ISES 1999 Solar World Congress Jerusalem, Israel
July 4-9, 1999
Editor: G. GROSSMAN
Conference Proceedings Volume 111
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The 1999 ISES Solar World Congress was held under the theme Solar is Renewable, adequately representing a Congress on the threshold of the 21 st Century in Israel- a pioneer in solar energy utilisation. We would like to thank our International team of Focal Point Editors and the many reviewers who helped make this event a success. We also wish to express our gratitude and thanks for their support to:
Ben-Gurion University of the Negev Israel Electric Company Israel Ministry of National Infrastructures Israel Ministry of Science Ormat Industries Ltd. Technion- Israel Institute of Technology Tel Aviv University World Energy Council Weizmann Institute of Science
Y. Zvirin - Congress Chair G. Grossman - Chair, Scientific Committee D. Dvorjetski- Executive Secretary H. Tabor-Chair, ISES Israel
vi
Congress Committee Y. Zvirin, Chair
D. Dvorjetski, Executive Secretary Organizing Committe D. Dvorjetski A. Elazari M. Epstein D. Faiman G. Grossman A. Kribus E. Shaviv H. Tabor D. Weiner Y. Zvirin Scientific Committee
G. Grossman, Chair J. Appelbaum D. Dvorjetski D. Faiman A. Kribus E. Shaviv D. Weiner Y. Zvirin Professional Tours Committee
D. Weiner, Chair U. Fisher Exhibition Committee
A. Elazari, Chair D. Dvorjetski Finance Committee
M. Epstein, Chair A. Shavit
Vll
International Advisory Committe D. Aitken, USA W. A. Beckmann, USA T. Book, UK A. Goetzberger, Germany Y. Goswami, USA O. Headley, Trinidad 1. G. Hestnes, Norway K. G. T. Hollands, Canada L. Imre, Hungary L.F. Jesch, UK H. S. Jeon, Korea D. Lorriman, Canada D. Mills, Australia
M. Nicklas, USA M. Oliphant, Australia E. de Oliveira Fernandes, Portugal D. Serghides, Cyprus L. Sherwood, USA 2. Silvi, Italy T. Tani, Japan M. Vazquez, Spain
viii
International R e v i e w B o a r d - Focal Point Editors
Solar Energy Systems for Buildings, Solar Architecture and Daylighting: A. G. Hestnes, NORWAY Flat Plate and Non-Concentrating Solar Collectors: W. A. Beckmann, USA Solar Thermal and Photovoltaic Concentrating Collectors: J. J. O'Gallagher, USA Photovoltaic Cells and Modules: M. Konagai, JAPAN Solar Collector Optical Materials: R. E. Collins, AUSTRALIA Solar Hot Water and Thermal Energy Supply B. D. Wood, USA Solar Thermal Electricity. A. Kribus, ISRAEL Photovoltaic Electricity:. J. Appelbaum, ISRAEL Active Cooling, Refrigeration and Dehumification: H.-M. Henning, GERMANY Space Applications: K. P. Bogus, THE NETHERLANDS Wind Power Systems and Solar-Wind Hybrids: M. Hirsch, ISRAEL Biomass Energy Conversion: R. P. Overend, USA Sustainable Hydroelectricity and Ocean Energy Conversion: D. Bharathan, USA Thermal Storage: J. Rheinlander, GERMANY Electrical Storage: D. Weiner, ISRAEL Hydrogen, Chemical Energy Storage, and Fuels: A. Heinzel, GERMANY Solar Radiation Measurement and Analysis: P. Ineichen, SWITZERLAND Indirect Solar Resource Evaluations: H. G. Beyer, GERMANY Education and Information Exchange: L. F. Jesch, UK Marketing and Commercialization: T. Book, UK Policy and Programs: A. Rabl, FRANCE Developing Countries: A. Ramachandran, INDIA Environmental and Social Impacts of Energy Systems: E. de Oliveira Fernandes, PORTUGAL Special Topics: C. Silvi, ITALY
ix
FOREWORD These volumes of Proceedings are the record of the 1999 ISES Solar World Congress, held in Jerusalem, Israel on the 45 th Anniversary of the International Solar Energy Society. The Congress was held under the theme Solar is Renewable, adequately representing a meeting on the threshold of the 21 st Century. The event also marks the 20 th anniversary of the Israeli Section of ISES, founded in 1979 - the year ISES celebrated its Silver Jubilee. The tradition of the biennial congress of ISES has been established since 1973. This Congress followed meetings in Paris, France (1973), Los Angeles, California (1975), New Delhi, India (1977), Atlanta, Georgia (1979), Brighton, UK (1981), Perth, Australia (1983), Montreal, Canada (1985), Hamburg, Germany (1987), Kobe, Japan (1989), Denver, Colorado (1991), Budapest, Hungary (1993), Harare, Zimbabwe (1995) and Taejon, Korea (1997). Israel- a pioneer in solar energy with the highest per capita utilization in the world - has for a long time expressed its interest in hosting the Solar World Congress. The Israeli Section of ISES is happy and proud to have had the opportunity to organize the Congress in Jerusalem this year. The Congress organizers have made great efforts to assure the quality of papers. The Congress Scientific Committee, in consultation with ISES, has developed a review procedure by which to accept papers for presentation at the Congress and publication in the General Proceedings. Due to time limitations, it was decided to base the review on extended abstracts of at least 400 words and up to one page. The abstracts submitted were screened by the Scientific Committee and then referred to Focal Point Editors, depending on their technical category. The responsibility of each Focal Point Editor was to handle the review of the abstract by referring it to three qualified reviewers in the respective area, receiving their comments, and making the final recommendations to the Scientific Committee regarding acceptance/rejection of the paper and required revisions. We have recruited an excellent Review Board consisting of 26 Focal Point Editors from around the globe, covering the full range of ISES topics. Following the Call for Papers, 464 abstracts have been received. Of those, 192 papers were accepted as submitted, 1 2 5 - with recommended changes, 1 1 3 - with mandatory revisions and 34 were rejected. The Congress was attended by over 520 participants, representing 47 countries. The Program included 207 oral presentations that ran in six parallel sessions during the five days of the Congress, and 149 poster presentations in three main sessions. In addition, 10 plenary lectures and 14 keynote lectures were presented. A business track under the title Solar Means Business included presentations and discussions on market implementation of solar technology. The Congress further included two panel discussions and two workshops, dealing with "WIRE" (World-wide Information System for Renewable Energy) and with IPMVP (International Performance Measurement and Verification Protocol). An exhibition presented the latest in solar products. We wish to express our sincere thanks to the international team of Focal point Editors, who have done a remarkable job in handling the review of the papers in an expedient manner, and to the many reviewers who helped make the Congress program a success. I personally wish to express my pleasure of working as a team, on all aspects of the Congress organization, with my two colleagues: Yoram Zvirin - t h e Congress Chairman and Dubi Dvorjetski- the Congress Executive Secretary. It is my hope that the Congress participants as well as those who were unable to attend, will find these Proceedings a useful reference and resource material, describing the state-of-the-art in solar energy. We look forward to the next Congress to be held in Adelaide, Australia in 2001. Gershon Grossman Editor Scientific Program Chairman
Table of C o n t e n t s - Volume III
Flat Plate and Non-Concentrating Solar Collectors Hybrid Solar Collectors for Microclimate Forming System G. J. Basler, D. Kwiecien ............................................................................................................................................................. 3
Testing of a Flat Plate Collector with Selective and Nonselective Absorbers That Are Otherwise Identical W. S. Duff, D. Hodgson ............................................................................................................................................................... 4
Comparison Between a Simple Solar Collector Accumulator and a Conventional Accumulator A. J. Fasulo, J. Follari ................................................................................................................................................................ Solar Air Collectors - Investigations on Several Series-Produced Collectors H. Fechner, O. Bucek ................................................................................................................................................................ An Empirical Heat Transfer Equation for the Transpired Solar Collectors, Including No-Wind Conditions K. G. T. Hollands, G. W. E. van Decker ..................................................................................................................................... A CFD Heat Transfer Analyses of the Transpired Solar Collector under No-Wind Conditions K. G. T. Hollands, S. J. Arulanandam, E. Brundrett .................................................................................................................... Analysis of Thermal Performance on an Air-Type Solar Collector with 2- Glass Using Carbon Fiber Sheet as Collecting Material
11 17 23 29
X. -rn. Jiang, H. Baba, K. Kanayama, N. Endoh ......................................................................................................................... 35
Research and Development of Solar Collectors Fabricated From Polymeric Material A. I. Kudish, E. G. Evseev, M. Romrnel, M. KOhl, G. Walter, T. Leukefeld .................................................................................. 40
Study of a Mixed (Water Or Air) Solar Collector S. Laiot ...................................................................................................................................................................................... 50
Uncertainty in Solar Collector Testing Results E. Mathioulakis, K. Voropoulos, V. Belessiotis ........................................................................................................................... 50
Optimized Finned Absorber Geometries for Solar Air Heating Collectors K. Pottier, C. M. Sippel, A. Back, J. Fricke ................................................................................................................................. 62
Inclination Dependency of Flat Plate Collector Heat Losses G. Rockendorf, B. Bartelsen, M. Kiermasch ............................................................................................................................... 72
PV-Hybrid and Thermo-Electric-Collectors G. Rockendorf, R. Sillmann, L. Podlowski, B. Litzenburger ........................................................................................................ 76
Elastomer-MetaI-Absorber - Development and Application G. Rockendorf, B. Bartelsen, N. Vennemann, R. Tepe, K. Lorenz, G. Purkarthofer ................................................................... 83
Solar Absorber System for Preheating Feeding Water District Heating Nets K. Vajen, M. Kr~mer, R. Orths, E. K. Boronbaev, A. Paizuldaeva .............................................................................................. 90
Statistical Analysis of Solar Collector Test Results in View of Future Certification K. Voropoulos, E. Mathioulakis, V. Balessiotis ........................................................................................................................... 92
Thermal and Electrical Yield of a Combipanel H. A. Zondag, D. W. de Vries, A. A. van Steenhoven, W. G. J. van Helden, R. J. C. van Zolingen ............................................. 96
A Comparative Investigation of Radiation Heat Transfer in Transparent Insulation with Differernt Reflection Models Y. Zvinn, B. Aronov .................................................................................................................................................................. 102
Solar Hot Water and Thermal Energy Supply Thermal Destratiflcation in Small Standard Solar Tanks Due to Mixing During Tapping E. Andersen, S. Furbo ............................................................................................................................................................. 111
Integrated Thermal Improvements for Greenhouse Cultivation in the Central Part of Argentina J. R. Banal, P. D. Galimberti, A. Barone, M. A. Lara ................................................................................................................ 120
In Situ Short -Term Test for Large Solar Thermal Systems N. Benz, T. Beikircher, M. Gut, P. Kronthaler, C. Oberdorf, W. Sch~lkopf, H. DrOck ................................................................ 126
•
Solar Process Heat with Non-Concentrating Collectors for Food Industry N. Benz, M. Gut, T. Beikircher, W. Ru/~ ...................................................................................................................................
131
Laboratory Testing of Integrated Collector Storage (ICS) Systems with Transparent Insulation Material M. Bosanac, J. E. Nielsen ........................................................................................................................................................
137
Uncertainty in Economical Analysis of Solar Water Heating and Photovoltaic Systems S. Co/le, S. L. de Abreu, R. R ~ h e r ..........................................................................................................................................
141
Solar Pond as a Power Source for Desalination U. Fisher. .................................................................................................................................................................................
150
Multistage Still J. Franco, L. R. Saravia, S. Esteban ........................................................................................................................................ 155
Development of a Smart Solar Tank S. Furbo, E. Andersen .............................................................................................................................................................
Thermal Modelling and Performance Prediction of Drying Processes under Open-Sun-Drying H. P. Garg, R. Kumar. ............................................................................................................................................................... Medium Scale Solar Crop Dryers for Agricultural Products 0. Headley, W. Hinds .............................................................................................................................................................. The Marstal Central Solar Heating Plant: Design and Evaluation A. Heller, J. Dahm ...................................................................................................................................................................
160 170 175 180
A Combined Ejector Cooling and Hot Water Supply System Using Solar and Waste Heat Energy B. J. Huang, V. A. Petrenko .....................................................................................................................................................
188
A Solar Still with Minimum Inclination and Coupled to an Outside Condenser D. Inan, A. El-Bahi ...................................................................................................................................................................
191
Modelling of a Thermosyphonally Driven Discharge Unit of a Storage Tank U. Jordan, K. Vajen, B. Knopf, A. Spieler, F. Hilmer. ................................................................................................................ 197
Performance of Transparently Insulated Solar Passive Hot Water Systems N. D. Kaushika, K. S. Reddy ....................................................................................................................................................
203
Thermodynamic Study of a Regenerative Water Distiller G. Koury Costa, N. Fraidenraich ..............................................................................................................................................
211
The Performance and Analysis of a Multiple - Effect Solar Still Utilizing Solar and/or Waste Thermal Energy A. I. Kudish, E. G. Evseev, L. Horvath, G. Mink ....................................................................................................................... 216
Performance and Analysis of a Multiple Effect Solar Still Utilizing an Intemal Multi - Tubular Heat Exchanger for Thermal Energy Recycle G. Mink, L. Horvarth, E. G. Evseev, A. L Kudish ...................................................................................................................... 226
Modelling Thermosyphon Solar Water Heaters with Mantle Collector-Loop Heat Exchangers G. L. Morfison, G. Rosengarten, M. Behnia ............................................................................................................................. 236
Bridging the Gap: Research and Validation of the DST Performance Test Method for CEN and ISO Standards - Project Results D. Naron, M. Rolloos, M. J. Carvalho .......................................................................................................................................
245
Research on a New Type of Heat Pipe Vacuum Tube Solar Water Heater N. Zhu, Ho Zinian .....................................................................................................................................................................
253
Solar Process Heat: Distillation, Drying, Agricultural and Industrial Uses B. Norton .................................................................................................................................................................................
256
Brackish Water Destillation with Plane Microporous Membranes Driven by Temperature Difference L Odicino, J. Marchese, D. A. Perelld, G. Lesino .................................................................................................................... 261
Effective Solar Energy Utilisation - More Dependent on System Design Than Solar Collector Efficiency J. Rekstad, L Henden, A. G. Imenes, F. Ingebretsen, M. Meir, B. Bjerke, M. Peter ................................................................. 265
Dynamical Model for Solar Still Validated From Optical and Thermal Parameters Measured Experimentally H. Romero-Paredes, E. Torijano, A. V~zquez, A. TorTes, J. J. Ambriz, E. Torijano Jr. ............................................................. 271
Characteristics of Vertical Mantle Heat Exchangers for Solar Water Heaters L. J. Shah, G. L. MorTison, M. Behnia ......................................................................................................................................
276
A System for Solar Process Heat for Decentralised Applications in Developing Countries F. Sp~te, B. Hafner, Ko Schwarzer ...........................................................................................................................................
286
Performance of a Cascade of Flat Plate Collectors T. Tomson ...............................................................................................................................................................................
292
A Solar Absorption Air-Conditioning Plant Using Heat-Pipe Evacuated Tubular Collectors H. Zinian, Z. Ning .....................................................................................................................................................................
297
~
Xll
Advanced Fuzzy Control of the Temperature in the Test Chamber B. Zupancic, I. Skrjanc, A. Krainer, B. Furlan ...........................................................................................................................
304
Solar, Thermal and Photovoltaic Concentrating Collectors Design and Construction of a Line-Focus Parabolic Trough Solar Concentrator for Electricity Generation G. C. Bakos, D. Adamopoulos, N. F. Tsagas, M. Soursos ....................................................................................................... The Duct Selective Volumetric Receiver: Potential for Different Selectivity Strategies and Stability Issues X. G. Casals, J. I. Ajona ........................................................................................................................................................... A Parabolic Dish Concentrator From a Telecommunication Antenna: Optical and Thermal Study of the Receiver C. A. Estrada, R. Dorantes, E. Rincon ..................................................................................................................................... Efficiency Improvement of Parabolic Trough Collectors by Means of Additional End Reflectors T. H. Fend, Jo Leon, P. Binner, R. Kemme, K. -J. Riffelmann, R. Pitz-Paal ............................................................................... Experimental Performance of a PV V-Trough System N. Fraidenraich, E. M. de Souza Barbosa ................................................................................................................................ Performance Analyses of a Combined Photovoltaic/Thermal (PV/T) Collector with Integrated CPC Throughs H. P. Garg, R. S. Adhikari ........................................................................................................................................................ An Astigmatic Corrected Target Aligned Solar Concentrator for Solid State Laser Pumping M. Lando, J. Kagan, B. Linyekin, L Sverdalov, G. Pecheny, Uo Achiam .................................................................................. Nonimaging Fresnel Lens Concentrators for Photovoltaic Applications R. Leutz, A. Suzuki, A. Akisawa, T. Kashiwagi ......................................................................................................................... Thermo-Mechanical Design of a Large Compound Parabolic Concentrator for 500 KWt Solar Central Receiver System G. Miron, S. Weis, I. Anteby, B. Ostteich, E. Taragan .............................................................................................................. Simulation and Analysis of the Performance of Low Concentration PV Modules M. Munschauer, K. Heumann .................................................................................................................................................. Practical Design Considerations for Secondary Concentrators at High Temperatures J. O'Gallagher, R. Winston ...................................................................................................................................................... Comparison of Predicted and Measured Performance of an Integrated Compound Parabolic Concentrator (ICPC) J. O'Gallagher, R. Winston, J. Muschaweck, A. R. Mahoney, V. Dudley .................................................................................. Double-Tailored Imaging Concentrators H. Ries, J. M. Gordon .............................................................................................................................................................. Development and Test of an Equipment to Replace Broken Glass Envelops of Receiver Tubes in Parabolic Trough Collectors K. -J. Riffelmann, M. B6hmer, T. Fend, R. Pitz-Paal, C. Spitta, J. Leon ................................................................................... Cooling of PV Modules Equipped with Low Concentrating CPC Reflectors M. ROnnelid, B. Karlsson, P. Krohn, B. Peters ......................................................................................................................... A Solar Bowl in India S. Rousseau, G. Guigan, J. Harper ......................................................................................................................................... The Development and Testing of Small Concentrating PV Systems G. R. Whiffield, R. W. Bentley, C. K. Weatherby, A. Hunt, H. -D. Mohring, F. H. Klotz, P. Keuber, J. C. Minano, E. Alarte-Garvi .........................................................................................................................................................................
315 324
333 337 342
349 354 358
367 370 377
382 388
394 400 405
409
Active Cooling, Refrigeration and Dehumidification Thermodynamic Design of a Solar Refrigerator to Conserve Sea Products H. D. Arias-Vatela, W. Soto Gomez, 0. Castillo-Lopez, R. Bast-Brown ................................................................................... 419 Demonstration of a New Type of ICPC in a Double Effect Absorption Cooling System W. S. Duff, R. Winston, J. J. O'Gallagher, T. Henkel, J. Muschaweck, R. Christiansen, J. Bargquam ...................................... 424
xiii
Indirect Evaporative Cooling through a Concrete Ceiling B. Givoni, S. Nutalaya ..............................................................................................................................................................
428
Experimental Studies on a Hybrid Dryer S. Kumar, G. A. Mastekbayeva, P. C. Bhatta, M. A. Leon ........................................................................................................ 434
Combined Solar Heating and Radiative Cooling System M. Meir, H. Storas, J. Rekstad .................................................................................................................................................
441
Hybrid Solar/Gas Cooling Ejector Unit for a Hospital in Mexico J. L. Wolpert, M. V. Nguyen, S. B. Riffat .................................................................................................................................. 447
Thermal Storage The Freezing Process of Water Inside a Vertical Cylinder with a Finned Tube Y. Changsoon, S. Taebeom, K. Jaeyoon ................................................................................................................................. 455
The Ciclops System: Optimised Management of Middle-Sized-Hybrid Wind-PV-Diesel Plants E. Uobet, J. Sold, J. Pitarch, J. Prats .......................................................................................................................................
462
Solar District Heating with a Combined Pit and Duct Storage in the Underground M. Reuss, J. Po Mueller. ...........................................................................................................................................................
468
Solar Heating with Heat Pump and Ice Storage A. B. Schaap, J. M. Warmerdam, E. E. Gramsbergen .............................................................................................................. 475
An Analysis of Phase Change Heat Transfer in a Solar Thermal Energy Store A. Trp, B. Frankovic, K. Lenic ..................................................................................................................................................
484
Modelling of Two - Layer Stratified Stores J. van Berkel, C. C. M. Rindt, A. A. van Steenhoven ................................................................................................................ 490
Full T a b l e o f C o n t e n t s ......................................................................................................................................... 499 I n d e x o f A u t h o r s ...................................................................................................................................................... 512 I n d e x o f P a p e r s ......................................................................................................................................................... 542
xiv
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ISES Solar World Congress 1999, Volume III
XVIII.
Flat Plate and Non-Concentrating Solar Collectors
ISES Solar World Congress 1999, Volume III
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ISES Solar World Congress 1999, Volume III
HYBRID SOLAR COLLECTORS FOR MICROCLIMATE FORMING SYSTEM
Dariusz Kwieciel, Gerard Jan Besler Department of Environmental Engineering, Wroc fiw University of Technology, ul. Norwida 4/6, 50-373 Wroc ~w, Poland, tel. 00487 /3226435, fax. 00487 /3203532,
[email protected]
Contemporary and modem residential buildings most often are characterised by well thermal isolated walls and very tight doors and windows, which results in a very small infiltration of the outside air. Then gravitational ventilation cannot serve its purpose satisfactorily. Mechanical ventilation is thus indispensable. It makes solar energy and energy from the shallow depth of the ground possible to be used in microclimate forming. It comes out not only possible but also effective. Research on such solutions heve been being carried for a few years in Air Conditioning And District Heating Chair in Wroc ~w University of Technology [ ,2]. Conventional energy for heating and cooling air purposes is replaced by natural renewable energy. In the paper an analysis of the possibilities of the natural solar energy gain in hybrid (liquid- air) solar collectors is presented. The collectors co-operate with thermoventilation and domestic hot water systems. On the experimental hybrid solar collectors station, which was made in technical scale, the measurements were made in natural climate conditions. The hybrid solar collectors experimental characteristics were estimated for water and air. Based on the known mathematical models which describe heat exchange in solar collectors, theoretical characteristics of hybrid solar collectors were made. For the average climate conditions in Wroc aw, the hybrid solar collectors operation efficiency in the heat - ventilation and domestic hot water systems were analysed (fig. ). The solar collectors co-operation with ground heat and mass exchanger were also included in the conducted analysis. Results of the analysis showed that the heat gain for ventilation and thermoventilation is more effective in the cold period of the year then for domestic hot water (fig. 2). Co-operation of the solar collectors and membraneless ground exchanger in the ventilation and thermoventilation systems gives very good result because energy gain from the ground is the most effective in the months with the lowest insolation level (XI + I). At that time solar energy input in solar hybrid collectors has low efficiency. Two natural energy sources: solar energy from solar collectors and ground energy from the ground exchanger supplement each other and thus the total quantity of gained heat energy is kept at the same level during almost whole of the heat season, it means in months: XI -III. System with ground exchanger serves not only for heating and domestic hot water but also allows for good ventilation with
air cooling in the summer and in some cases is able to replace expensive conventional air-conditioning system.
F i g . . The system discussed: a,b-solar collectors; 2-grave heat container; 3-heater; 4-ventilator; 5-throttles; 6-outside air intake device; 7-suspension ceiling; 8-intake ventilating device; 9uptake ventilating device; 0-heat container for domestic hot water system; - ground heat and mass exchanger.
Fig. 2. Total energy gained from natural sources during whole heating period (X+IV) in various solar systems. [ ] Kwieciel D., Besler G.J.: Thermoventilationsystem for accommodations with use of solar energy. Report K070 , Wroc fiw University of Technoligy, 997. [2] Kwieciel D.: Hybrid solar collectors for microclimate forming system. Dissertation for Doctor of Engineering Degree, Wroc hw University of Technoligy, 998.
Keywords: hybrid collectors, microclimate, natural energy, solar energy, thermoventilation.
ISES Solar World Congress 1999, Volume III
TESTING OF A FIAT PLATE COLLECTOR WITH SELECTIVE AND NONSELECTIVE ABSORBERS THAT ARE OTHERWISE IDENTICAL
William S. Duff and David Hodgson Department of Mechanical Engineering, Colorado State University, Fort Collins CO 80523, USA Telephone: 1-970-493-1321, FAX: 1-970-495-0657, E-mail:
[email protected] Abstract- We tested a flat plate collector using two absorbers that were physically identical, except for the fact that one was nonselective and the other had a good selective coating. We experimented over a range of input temperatures, keeping the input temperature to the collector constant throughout any one day. During a full day test, data was recorded every five seconds. We used several measures to insure steady state operation had been achieved before we selected two half-hour reporting periods symmetrical about solar noon. We found 1) that there was no difference in the efficiency of the collector with the selective absorber before and after stagnation and 2) that the efficiency of the collector with the nonselective absorber was substantially lower than the collector with the selective absorber.
1. INTRODUCTION We tested a flat plate collector using two absorbers that were physically identical, except for the fact that one was nonselective and the other had a good selective coating. A selective surface will exhibit a high absorptance for radiation in the solar spectnm~ and a low emittance for the longer wavelengths produced at the collector operating temperatures. Radiative losses are significantly reduced while high fractions of solar radiation absorbed are attained by the choice of a good selective surface for the absorber of a solar collector. The better the selective absorber coating, the more efficiently a given solar collector will operate at high temperatures and under low solar radiation conditions. One often wants to know the actual performance of a selective surface in an application. Though direct measurement of the properties of a surface can be made by various optical means, use of such measurements to predict actual losses due to radiation from the surface in the application introduces a variety of errors. These errors include inaccuracies in the optical measurement process itself and in extrapolating it to the conditions of the application. In non-evacuated collectors, conductive and convective losses are a significant fraction of the total losses. Conductive and convective losses can be modeled, but the complex geometry of flat plate collectors usually require a finite element approach to get good predictions. Even so, errors are introduced in the approximations and assumptions and one can never be sure that all factors have be taken into account. In-situ calorimetric measurements along with accurate measurements of solar radiation and other environmental factors can provide accurate estimates of the performance of a collector. However, because there are still uncontrolled, unmeasured or unknown processes, separating out the performance of a component of the collector, such as the absorber, from the performance of the collector itself can be a significant, often daunting, task. One way to approach this issue is to eliminate as many uncontrolled, unmeasured and unknown processes as possible and look at relative differences in the performance of the collector when one component property is changed. In our experiment, we used the same collector to evaluate the performance of two physically identical absorbers, one
nonselective and the other with a good selective surface. In addition, we evaluated the performance of the selective absorber before and after stagnation. 2. T E S T I N G SET-UP The collector was oriented south with a slope equal to the latitude of 40 degrees. The water flow rate was maintained as close to 2.2 kg/min as possible. The experimental equipment consisted of 9 Data acquistion software: Labtech by Keithly 9 Data acquisition hardware: HP 75000 Series B 9 Flow meter: Micro Motion model d-40 9 Boiler: Argo Industries model AI20-240 9 Power variac: Powerstat 1296d 9 Thermocouples: t-type (tested before installation) 9 Pyranometer: Epley Precision (calibrated before test) 3. E X P E R I M E N T A L APPROACH Efficiencies over a representative range of collector temperatures were experimentally determined and a collector efficiency curve was created for each of the three test sets using ANSI/ASHRAE standard 93-1986 as a guide. 9 The first test set was performed by testing the collector with the selective absorber before the absorber had been allowed to stagnate. To ensure that the absorber did not reach high temperatures the collector was covered at all times when it was not being tested. 9 The second test set was performed by testing the collector with the selective absorber after it had been allowed to stagnate for thirty days. 9 The final test was performed by testing the collector with the non-selective absorber. To create the curves, five or six of the most stable full day tests were selected from a much greater number of full day tests for each of the three test sets. During a full day test, data was collected every five seconds. Five parameters were recorded: 9 ambient temperature 9 temperature of the water entering the collector 9 temperature of the water leaving the collector 9 the water flow rate
ISES Solar World Congress 1999, Volume III
9 radiation incident on the plane of the collector. MATLAB was used to analyze the data and calculate a representative instantaneous efficiency. To minimize the effects of the thermal mass of the collector, the instantaneous efficiency was calculated by averaging a morning and an afternoon efficiency measurement. The morning and afternoon efficiencies were determined by first finding two symmetric stable fifteen-minute intervals of data on either side of solar noon. The data for these fit~een-minute periods were then averaged and the efficiency was calculated by mass flow rate * specific heat * temperature rise across collector radiation level * aperture area of collector These stable periods were selected by searching the calculated instantaneous efficiencies in the four hour period that had solar noon as its midpoint, excluding 15 minutes on either side of noon, and selecting the symmetrically placed set of fifteenminute periods having the smallest combined standard deviation of efficiencies. Once the fifteen-minute periods were selected, they were checked to make sure they conformed to Section 8.3 of the ANSI/ASHRAE 93-1986 standard. If not, the day was eliminated.
4. RESULTS Table 1 shows the results of all of the selected test days. The entries are the average of the selected fifteen-minute morning and afternoon test periods. Figure 2 shows the average efficiency for each day as a function of the difference between the collector temperature and the ambient temperature divided by the radiation level. The collector temperature used was the average of the incoming and outgoing water temperatures. From Table 1 and Figure 2, it is clear that the selective surface outperformed the non-selective surface. The selective surface's efficiency was about 7 percentage points higher than that of the non-selective surface when the collector temperature was close to the ambient temperature. At higher operating temperatures, the difference was closer to 12 percentage points. The collector with the selective absorber installed was stagnated for a thirty-day period at collector temperatures ranging up to and above 170C. As can be seen in Figure 2, the testing indicates that there was no significant degradation in the performance of the selective surface due to stagnation. 5. CONCLUSIONS
Figure 1 shows a sample plot of data collected. The top graph shows the measured temperatures and flow rates. The bottom graph shows the incident sunlight and a calculated instantaneous efficiency. The vertical lines on the top graph indicate the two fifteen-minute periods symmetric on either side of solar noon which were used to calculate the average instantaneous efficiency.
There was little or no difference in the efficiency of the collector with the selective absorber before and aider stagnation. However, the efficiency of the collector with the nonselective absorber was substantially lower than that of the collector with the selective absorber.
ISES Solar World Congress 1999, Volume III
Figure I" Sample Daffy Data Collection.
ISES Solar World Congress 1999, Volume III
0 = Selective A b s o r b e r b e f o r e S t a g n a t i o n x = Selective A b s o r b e r after S t a g n a t i o n * = Non-Selective Absorber
Figure 2: Efficiency Curves for Different Absorbers.
ISES Solar World Congress 1999, Volume III
Table 1: Testing Results SET 1" Selective Absorber Before Stagnation
4-Jun
Radiation Tamb Flow Tin (W/m^2) (C) (kg/min) (c) AM 959 28.2 2.2 70.2 992 30.2 2.1 71.8 PM
Tout
Temp.Rise
(O
(O
Efficiency
81.0 84.1
10.8 12.3
0.46 0.5
12-Jun
AM PM
928 926
27.4 31.8
2.3 2.3
14.8 15.0
31.2 32.1
16.4 17.1
0.78 0.81
13-Jun
AM PM
932 976
31.4 32.1
2.2 2.2
41.9 42.4
56.8 58.0
15.0 15.6
0.67 0.67
29-Jun
AM PM
935 961
32.6 33.1
2.3 2.2
83.3 83.2
93.4 93.5
10.1 10.3
0.46 0.45
6-Jul
AM PM
915 925
32.4 32.9
2.3 2.3
59.6 59.8
72.3 72.8
12.7 13.0
0.59 0.61
7-Jul
AM PM
928 922
37.5 38.0
2.3 2.3
50.2 50.3
64.8 65.1
14.6 14.7
0.67 0.68
ISES Solar World Congress 1999, Volume III
Table 1 (Continued)" Testing Results SET 2: Selective Absorber After Stagnation Flow Tamb (C) (k~min) 32.4 2.3 2.3 33.2
8-Aug
AM PM
Radiation (W/mA2) 980 997
12-Aug
AM PM
1005 1026
33.7 34.3
23-Aug
AM PM
1032 1019
24-Aug
AM PM
3-Sep
AM PM
Tin
Tout
Temp.Rise
(c)
(c)
(c)
Efficiency
18.2 18.4
36.3 36.8
18.1 18.4
0.81 0.81
2.3 2.2
30.9 31.2
48.6 49.5
17.7 18.3
0.75 0.75
31.7 36.8
2.2 2.1
72.3 75.0
85.0 88.4
12.8 13.4
0.51 0.52
1083 1071
33.9 35.2
2.1 2.1
58.9 59.1
74.2 75.0
15.3 15.8
0.57 0.60
1036 1014
33.6 35.3
2.2 2.1
31.1 30.8
49.2 49.2
18.1 18.3
0.71 0.73
ISES Solar World Congress 1999, Volume Ill
10
Table 1 (Continued): Testing Results SET 3: Non-Selective Absorber Radiation Tamb Flow (W/mA2) (c) (kg/min) 1051 29.7 2.3 1010 30.9 2.2
19-Sep
AM PM
25-Sep
AM PM
1072 1061
29.2 30.7
27-Sep
AM PM
1074 1074
8-Oct
AM PM
10-Oct
AM PM
Tin
Tout
Temp.Rise
(c)
(c)
(c)
Efficiency
19.1 19.4
37.0 37.1
17.9 17.6
0.72 0.73
2.0 2.0
77.3 77.8
88.2 88.8
10.9 11.0
0.39 0.40
25.7 26.6
2.3 2.3
40.1 40.1
54.6 54.6
14.5 14.4
0.59 0.58
1065 1061
28.4 24.6
2.3 2.3
25.8 25.7
41.6 40.9
15.9 15.2
0.66 0.63
1049 1055
27.1 27.7
2.3 2.3
58.9 59.0
69.9 70.6
11.0 11.6
0.46 0.47
ISES Solar World Congress 1999, Volume Ill
11
COMPARISON BETWEEN A SIMPLE SOLAR COLLECTOR ACCUMULATOR AND A CONVENTIONAL ACCUMULATOR Amflcar Fasulo and Jorge Follari Universidad Nacional de San Luis Chacabuco y Pedernera ? 5700 San Luis ? Argentina Fax 054 2652 430224 - e-Mail
[email protected]
Abstract- We have shown that, in dry regions with abundant solar radiation at a latitude lower than 40~ as the central-western part of Argentina it is possible to obtain domestic hot water by means of very simple collector accumulators less expensive than the current ones. The experimental assessment of a solar accumulator collector yielding daily 3001 of hot water is reported in this work. Therefore, the diurnal and daily global efficiencies and the nocturnal thermal losses have been systematically determined over a six-month period, from austral summer to austral winter. The results are compared with those obtained from two other systems tested at the same time. These systems are also designed to yield daily 3001 hot water. They are: A high quality solar system composed of a 4 m2 plane collector and an accumulator storage insulated by conventional material; an integrated plane and accumulator collector, IPAC, whose semitransparent thermal insulation has been reinforced. The new systems themselves provide hot water over 40~ during six months and reduce energy expenses the remaining six months, when installed in series with systems using conventional energy sources. Graphs and tables show the results obtained, such as diurnal and daily global efficiencies and nocturnal thermal losses of the systems.
1. INTRODUCTION The city of San Luis (Argentina), situated at 33.27 ~ South and 66.2 ~ West, with a temperate and dry climate, posseses an abundant wealth of sunshine over most of the year. A similar situation is there for the whole western region of the country from 40~ latitude northward. This favourable situation for development of solar energy esploitation allowed it to constitute itself as pioneer in the development and use of solar water heaters, Follari et al. (1998). The relatively high cost of solar devices relative to conventional ones however has limited their diffusion, and on arrival of the natural gas networks of low cost, use of them almost has disappeared, Fasulo et al. (1999). In spite of these circumstances, in isolated form there still subsists the use of solar water heaters, in most cases connected in series with conventional devices in order to assure provision with warm water at any hour and during all days of the year. On the other hand in the poor districts of the town we observe painting with black colouring of the domestic water reserve tanks. The residents argue that by this means they dispose of hot water during the summer period; what for sure they don't realise is that during the winter period energy consumption for warming water increases. Anyhow, this attitude gives evidence of the existence of a clear concience on the possibilities to
take advantage of by use of solar energy, and the wish to participate in their use. While new materials arrive on the market it is possible to introduce innovations in the solar devices, in the continuous search for reducing their cost, Torres et A1 (1997). Thus, a simple envelope of alveolar polycarbonate would not just allow improving the efficiency of those rudimentary collectors, but also would reduce the negative effects of the black paint during nighttime, particularly in the winter months. With these ideas in mind, in 1997 we began experimenting with an integrated collector storage, ICS, consistent in a tank of stainless steel of circular cross section 1 meter high and with 384 liters capacity. The literature schows the development of ICS, that are completely different, Schmidt et A1 (1988). These ICS" are similar to the plane collectos in that they have an inclined front surface facing the sun, with sides and back surface protected by opaque insulation. The cold water, in the ICS, enters by pressure from another reservoir situated at a higher altitude, and the warm water flows out through a pipe situated at the center of the tank's lid. In its interior and parallel to the base, at some 2 cm distance, there is a plate of the same material with perforations far from its center. This plate has the function of avoiding that the cold water current entering the tank at the center of the base might destroy the established thermal stratification. The tank is covered with matte black colour and enclosed in a
12
ISES Solar World. Congress 1999, Volume III
box of alveolar polycarbonate of 4 mm thickness. This device was compared to two solar water heaters designed to provide some 150 liters of hot water daily; one of them of low cost and craftmanshipmanufactured, T(100). The other one industrially produced, the one of best quality obtainable on the local market, T(160). The achieved results show us that this ICS is capable of producing - at least during six months of the year, from mid-spring till midautumn - 150 liters of warm water per day with temperatures above 40~ Fasulo et al. (1998). The efficiencies of the three compared devices measured by: 1)Extracted water, T(av) is 43~ for ICS Vs 48.5~ for the T(100), in the summer; 36~ for ICS Vs 44~ for T(100)and 51~ for T(160), in the fall (04/11 to 04/30). 22~ for ICS Vs 29~ for T(100) and 37~ for T(160), in the winter (06/02 to 06/19) 2) Efficiency of the ICS as solar collector, 0 in(av), determined by measuring the temperature increment of the water of the interior of the tank, relative to the periods of each of the extractions morningtime, evening, and nighttime; this last one permits us to determine 3) L(av), the nocturnal thermic losses, around of 7 MJ for ICS Vs 1.5 MJ for T(160). 4) The daily net efficiency of the devices, 0in-net(av), including in the former the nocturnal thermic losses. The table gives us account of the good prospects that ICS offers, as 0in(av) in all cases results superior to that of the other devices; at the same time it reveals the main defect it has: This lies in the high nocturnal thermic losses, by this having the effect that 0in-net(av) is inferior to that of the two devices equipped with solar collector with plane plaques. Finally, we show the estimated amortization time of the ICS for different conditions where the ICS is put in series with a heater functioning on some kind of conventional energy, as there are: Gas in tubes, electricity, or natural gas from provision network. We found that the amortization periods of the ICS were of 6 years, 3.5 and a half years, and 16 years, repectively. 2. INTEGRATED PLANE AND ACCUMULATING SOLAR COLLECTOR (IPAC). In a second stage, developed between end 1997 and 1998, we modified the device demanding it a higher rendering. For this purpose we combined it with a plane solar collector of 2 r~ surface, and positioned above it we put the accumulating solar collector that had been used in the former experiment. In the present case we reinforced the semitransparent coveting of the ICS adding a second layer of polycarbonate, separated some 3 cm from the original layer. We doubled the demand on the device as for the volume of warm water to be produced, so changing to 300 liters of water daily obtained in three extractions: 100 1 in the morning,
before sunrise, 100 1 at mid-day, and 100 1 in the evening, immediately after sunset. In the experiment the device was compared to a high quality commercial water heater designed por producing 300 1 of hot water daily, composed of two collector plaques of 2 ~ each, connected to a reserve tank of 270 1 protected by opaque thermic insulation, conventional system, CS. The results showed that the device works quite satisfactorily during the 6 months around summer, with ouput temperatures above or very near 40~ A. Fasulo et al. (1999). The background part of figure 1 shows us both devices on the test bench. The thermal losses in this last case are slightly higher in the winter period. These are consequences partly of the diverse climatic conditions, but mainly due to the circumstance that the IPAC operates at higher temperatures than the ICS. In this experiment, consisting in the systematic daily extraction of 300 1 of warm water the way indicated before, during four periods of no less than 15 days each and extending these periods until each of them included days of full sun, days partly clouded, and cloud-covered, we determined: 1) An average temperature of the IPAC compared to the CS, of: A) 46EC vs. 49EC for the period 13th to 31 st of January. B) 38EC vs. 43EC for the period 3ra to 27th of March. C) 34EC vs. 40EC for the period 29th of April to 22~a of May, and D) 31 ~ vs. 38E C for the period 16th to 30th of June. Thermal losses during each of the indicated periods and for each of the two devices were of: 4.35MJ vs. 0.26MJ, 5.38MJ vs. 0.95MJ, 7.46 MJ vs. 1.28 MJ y 8.37MJ vs. 1.33MJ, respectively.
3. A NEW ICS Keeping in mind from the former experiments: That one of the major advantages of those devices is their high accumulative capacity and their greater exploitation of diffuse solar radiation, allowing them to overcome cloudy days in the provision with warm water, as well as the major defect they reveal (high nocturnal thermal losses), a new ICS was designed meant for comparison with the IPAC and the conventional system, CS, of 4 n~ collecting area and thus produce 300 1 of warm water daily. This new ICS was constructed with a metallic cylinder made of stainless steel and of 768 1 capacity, 2 m high, covered by three envelopes of alveolar palycarbonate of 4 mm thickness. Thermic control of the device is done by means of five thermocouples installed in the length of the central axis of the cylinder and positioned at: 5 cm, 50 cm, 1 m, 1.5 m, and 1.98 m from the tank's base. On the exterior surface of the tank and on the inner sides of each of the polycarbonate coverings, as well as on the lid, we put thermocouples at the same altitudes as those positioned in the interior of the device.
ISES Solar World Congress 1999, Volume III
13
Fig. 1. View of the system when being built. At the back the two other systems for comparison.
4. THE IMPROVED IPAC In this third stage, we introduced the following improvement to the IPAC: Given that its worst defect consists in the thermal losses, their reduction was sought introducing a third covering of alveolar polycarbonate, making the calculus that by this means we would increase the thermal gradient by some 5~ resulting a total difference of about 25~ between the interior surface of the ICS an the environment. Figure 1 shows us a photo of the three solar heaters installed on a test bench. As in the previous cases, the experiments cover a minimum of 15 consecutive days that must be extended according to the requirement of comprising sequences of completely sunny and completely clouded days. Each of those days water extactions are done: One in the morning before sunrise, one at mid-day, and one in the afternoon immediately after sunset, determining the entrance and exit temperatures of the water, as well as all the other environment variables, i.e. maximum and minimum temperatures of the day, humidity, velocity and direction of the wind. Thermal control of the device is complemented determining all the temperatures in its interior and on its coverings immediately before and after each water extraction. 5. DATA ANALYSIS Data analysis in the first line presents the difficulty of comparing completely different solar collectors. On one side we have the CS that has a collecting area of 4 m2 inclined 45 ~ and north-oriented. The ICS has a net area of 2.17 m2 , in its major proportion composed of a vertical cylindric surface, and a horizontal surface - the lid. As the last we have the IPAC that is composed of a combination of those two. In table 1 we show the
dissimilarity of the sun-exposed surfaces, as well as the quantities of total radiation that reaches each of the devices at two extreme seasons of the year. We also find there the water quantity we should extract from each device if this would be performed proportional to the surface, starting from the condition that we can extract 300 1 daily from the CS. The volume we shall extract from the second experiment will be proportional as well to the areas of the devices as well as to the radiation that reaches them. Collector Surface m2 CS 4 ICS 2.17 IPAC 3.35
january n (MJ) 115 83 108
june propr extracted H (MJ) liters 94 100 100 52 60 70 76 86 90
Table 1. Areas of each device, radiation reaching them in each of the two extreme seasons of the year (summer and winter) on clear days, water volumes proportional to both and the extracted. Figure 2 shows us the exit temperatures of the IPAC for the 180 extractions performed, between Julian days 16 and 131; here we can appreciate the vast dispersion these data exhibit; owing to the changing climatic conditions, lower value data almost every time correspond to the extractions done in the morning, before sunrise. This pattern of dispersion is present in all three devices, with a distribution slightly above the shown for the CS and slightly below for the ICS, so that the data will be presented as mean values for each of the different experiments we performed. In table 2 we present the obtained results: In each columns the significant data of one experiment. The
ISES Solar World
14
Congress 1999, Volume III
files agrouped to show the types of data for each of the experiments. Where: Tair(av) is the average daily temperature of air. Tout is the temperature (average) of the extracted water. )T = Tout - Tin, where Tin is the temperature (average) of the cold water that had entered the device during the preceding extraction. Qout is the amount of energy in Mj gained at each extraction. Qint is the amount of energy stored in the tank during each of the diurnal periods between extractions of water (between 5 and 6 hours). L(av) is the nocturnal thermal loss of the water quantity the tank contained. 0r-out is the relative efficiency of the device calculated departing from Qout, in Mj, and the horizontal solar radiation, in Mj, summed up for the hours passed between the subsequent extractions. 0r-int is the relative efficiency calculated on the base of Qin (MJ) and the horizontal radiation. I
Experience/da
Fig. 3. Tout vs. secuence of water extractions of IPAC.
HI
n
t
iTai r (oc) F I (MY) Tout
,i
,
( ~C ) 9
)Tout = Tout-Tin 9 ( ~C ) Qout = miCTout 9 (MJ) Qint = CMiTint ,, ( M J ) L = CMiTint 9 (MJ) Or out = Qout/SiFI '0rint= Qint/St"I ,
1
1/2 days ICS IPAC CS 9
ICS IPAC . C ICS IPAC . CS ICS IPAC . CS ICS IPAC . CS ICS IPAC CS ' ICS IPAC i CS
23.9 ,,12"04
41.3 44.2 47.1
24.6 11.82 45.5 45.5 46.3
ml
i48
I
,+
10.99 3i.3 38.9 44.4
IV 17.5 8.39 34.7 43.5 46.7
V
9
16.4 8.81 31.5 38.7 42.7
VI 11.4 4.48 24.8 31.6 34.8
9
,
,
12.i 18.8 22 3.55 7.10 9.21 ---
'
9
20.2 23.0 . 26.0 8.44 9.63 . 10.87 16.67 14.55 . 16.34 -7.04 -5.02 .-0.41 0.34 0.29 0.24 ' 0.66 0.23 . 0.24
24.3 24.4 25.2 7120 9.19 10.55 31.43 14.54 16.48 -7.29 -5.40 -0.71 0.34 0.29 0.24 ' 0.62 0.22 . 0.24
12.9 20.3 26.0
S.40 8.48 10.86 11.65 14.73 16.75 -6.65 -4.87 -1.03 0.32 0.29 0.27 0.53 0.44 0.27
'
k
.
17.1 25.8 29.1 5.0 9.7 12.7 12.30 14.65 17.58 -4.73 -4.83 -0.28 0.40 0.47 0.39 0.72 0.58 0.39
'
14.2 21.4 25.5 " 4.16 8.08 10.66 9.66 11.00 . 17.42 -5.99 -5.34 -0.85 0.30 0.36 ~ 0.35 0.49 0.30 0.38
"
"
9
9
,
'
,
,
i
9 ---
,
,
ii .
0.36 0.47 0.51 ---
Table 2 Experimental results. I :16 to 34 Julian Days, extraction mi =100 1. Each device; II: 35 to 43 J.d. extraction ml=100 1. of CS, m2 = 90 1. of IPAC and m3 = 70 1. of ICS; III: 95 to 109 J.d., mi = 100 1. of each device; IV: 110 to 118 J.d., ml=100 1. of CS, m2 = 90 1. of IPAC and m3 = 70 1. of ICS; V: Idem IV 121 to 131 J.d.; VI: 165 to 178 J.d., ml=100 1., m2 = 901. and m3 = 70 1. ; M I = 270 1. (CS); M2 = 384 1. (IPAC) and M3 = 768 1.(ICS).
ISES Solar World Congress 1999, Volume III
Fig. 4. Gradual development of Tout and H for the three devices, having extractions of 100 1 of water in each operation.
6. RESULTS In table 2 first of all we observe that with the ICS we obtained values of Tout above 40~ for the summer period, extracting 3001 of water per day. In the column Qint we f'md a poor energy accumulation. This is an important aspect for this type of collector having in mind its accumulation capacity, as it gives account of one of the most important advantages compared to the CS, as it allows overcoming isolated clouded days. In figure 4 we can see this characteristic of the ICS when the exit temperatures of the three devices for three summer days are shown and after a sequence two clear days follow two days of low radiation. The ICS overcomes the first clouded day satisfactorily, begins declining at the second, and later when full radiation returns shows the effect of working at the limit of its reserves presenting lower temperatures of the mornig extractions. In consequence a second experiment was designed, this time extracting water in quantities proportional to a combination of the net collecting surfaces of each device and the amount of radiation they receive. The second line of table 2 shows us the obtained results. There first of all we can see that now the exit temperatures of all three collectors are similar, Qout of the ICS drops and slighty also that of the IPAC. A strong increase of the Qint of the ICS can be observed, as we anticipated, while the other two exhibit little variation with respect to the former experiment. Figure 5 shows us the Tout of the three heaters for a sequence of 11 extractions, the first 6 corresponding to two days of plain sun, following one day of low radiation (the 4 subsequent events), and finalizing with a fourth day with full sunshine. From comparison of the two graphs we can deduce some conclusions about the limits of warm water production each of those devices impose.
55
TO oC
50
50
45
45
40
40
35 I .3~
3o
25
cs (o~))
2o
20 15 10 5 0
9
~
i
~
9
i
62
9
I
6,
9
15
6'~
9
6'~
Secuence of estractions bewteen 35 and 38 Julian Days
9
70
55
55
50
50
45
45
4O
40
T o 35
-m-i
ut 30' ~ 252 2o-~ 15-' lO.~ 52 oSr162
(MJ) ISC
-A-
IPAC (~
I
i
60
62
35 1 3O
-0--
(~
- V - cc
9
,
MJ
(oc)
9
i
64
"
66
"
i
"
68
70
o f e s t r a c t i o n s b c w t e e n 35 a n d 38 J u l i a n D a y s
Fig. 5. Gradual development of Tout and H for the three devices, having water extractions of: 1001 for the CS, 90 1 for the IPAC, and 701 for the ICS. The third line of table 2 shows the results of the first experiment of the southern auamm period, performed between the 5~ of April and the 10~ of May. Here it is intended to work in a way similar to the summer experiment, i.e. one sequence of measurements extracting 3001 of water from each of the devices and a second sequence applying different extraction values to each heater according to the considerations before mentioned. During the fast part environment temperatures very below those normal for the season were registered, reaching minima of-4~ typical for the winter seasonl. This results in the noteworthy decreases of the Tout of both accumulator collector devices, in consequence only the CC hardly surpasses the working temperature of 40~ In the second experiment of this period, fourth line of table 2, with the mean environment temperature returning to its normal condition of +2.5~ compared to the former, but with a stronger cloudiness having the effect of a reduction of 5 MJ in the average daily H, we performed the second designed part, i.e. extracting 70 1 from the ICS, 90 1 from the IPAC, and 100 1 from CC, respectively. Compared to the previous series we see that the conventional device following the environment temperature increases Tout by 2.2~ whereas the ICS increases its Tout by 3.4~ and the IPAC by 4.5~ the increment in the )T is of 3~ 5~ and 5~ respectively. The autumn experiment is completed with
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a series of 10 days, measured beginning with the 11th of May, fitth line of table 2. We can observe that 0 r - o u t as well as 0r-int of the CS, are lower in summer. This is a direct consequence of the fact that the plane collectors, inclined 45 ~ directed to the north, miss the first and the last hours of solar radiation, whereas the ICS receives radiation all the time. In the sixth line we show the results of the winter period, measured between the 14~ and 26th of June.
7. CONCLUSIONS 1) 2)
3)
The ICS takes better advantage of the solar radiation than the CS, in summer. The ICS can replace the CS profitly in technical and economic respect during the summer months, most of spring, and beginnings of autumn in regions temperate, dry and with abundant solar radiation. Search for improvements of the semitransparent thermal isolations must go ahead, in order to
succeed in making the ICS competitive for the complete annual period. REFERENCES Follari J y Fasulo A. (1998) Veinte afios con los calefones solares Argentinos. Energias Renovables y Medio Ambiente. 5, 1 - 6 Fasulo A., Perello D. And Follari J. (1998) World Renewable Energy Congress V. 4, 2307 - 2310 Torres M., Follari J. and Fasulo A. (1996) An/disis t6rmico comparativo entre colectores pianos con cubierta de vidrio y policarbonato. ASADES I , 5.17 Schmids Ch., Goetzberger A. and Schmid J. (1998) Test Results and Evaluation of integrated collector storage systems with transparents insulation. Solar Energy 41, 5, 487 Fasulo A., Perello D. And Follari J. Comparison Against collector accumulator with semi-transparent insulationand and conventional. EuroSun 98.2, III, 3, 4-1
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SOLAR AIR COLLECTORS INVESTIGATIONS ON SEVERAL SERIES-PRODUCED COLLECTORS Hubert Fechner, Otto Bucek Dept. of Renewable Energies, Arsenal Research - Austrian Centre for Research and Testing, Faradaygasse 3, A-1030 Vienna, Austria, Phone: +43-1-79747-299 Fax: +43-1-79747-390, email: fechner.h@_,arsenal.ac.at
A b s t r a c t - Testing of solar liquid collectors is described in international standards (ISO 9806, prEN 12975). A standardized procedure for testing air collectors does not exist so far. Within Task 19 of the IEA-"Solar Heating and Cooling Program" tests of most of the few worldwide available types of solar air collectors were carried out. Collectors from Australia, Canada and several from Europe were tested at the Austrian Research Centre Arsenal. Development of testing conditions and appropriate presentation of results, as well as tests on efficiency, leakage, pressure drop, dependence on internal mass flow and wind effects was the aim of this project. Long-time proven products as well as promising prototypes have been tested. Various types of collectors were investigated: glazed modules with air flow below or on both sides of the absorber, different colored absorber and corrugated profiles, textile absorbers but also cheap site built collectors and uncovered perforated collectors. The choice of air flow pattern depends on the application. Generally, four distinct air flow patterns exist: A very simple air collector can be constructed with the air flow between the absorber and the glazing. Due to the high heat loss induced by the convective heat transfer to the glazing the efficiency will be quite low, especially, if a relative high temperature increase is needed. The air flow behind the absorber is probably the most common solution. Air flow on both sides of the absorber is used in medium temperature applications to increase the effective heat transfer area. The fourth air flow pattern - air flow passes through the (porous) absorber - offers the possibility of a cost effective solution for medium temperature collectors.
I. INTRODUCTION Active Solar Air Systems have been known for many decades. Although they are not very wide spread so far, solar air applications are a promising way of meeting the heating demand on ecological basis.
-
-
Starting with simple constructions at the end of the 19th century in the U.S.A., solar air systems are now in use for space heating, preconditioning of air as well as for cooling applications, for hay drying, for drying of tobacco, crops, fruits and timber. Air heating is tightly connected to architectural matters, but contrary to passive solar design, active air systems provide better heat distribution and regulation, which results in improved heat gains and finally more comfort. In order to pool the experience in designing air systems for space heating, the International Energy Agency (IEA) initiated a five year project: Within Task 19 "Solar Air Systems" of the "Solar Heating and Cooling Programme" more than twenty experts from nine countries, coordinated by the Operating Agent Arch.Robert S. Hastings, have worked out:
A book illustrating 33 exemplary buildings with diverse solar air systems. A catalog of manufactured components and guidelines for selecting them, A PC-based, easily used program to predict energy performance and comfort, A handbook for designing Air Heating systems.
2. GENERAL SPECIFICATIONS The central part of each solar system is the collector, where the energy radiated from the sun is collected and converted such that it can be easily distributed atterwards. Solar air collectors as active part of air heating systems are not really "common" products so far. Testing of various series produced air collectors in a reproducable and comparable way have not been done so far. Worldwide not even a dozen of companies are manufacturing air collectors, scarcely half a dozen have a relevant output. Beside manufactured collectors, site built collectors are important as well. Founded by the Austrian Ministry of Science the Austrian Centre for Research and Testing "Arsenal Research" has invited
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manufacturers of solar air collectors to bring their products to Arsenal for being tested. Seven manufacturers from seven different countries, mainly from Europe but also from Canada and Australia, had taken up the offer to have their products tested. Long-time proven products have been tested as well as promising prototypes. The main topics of development, investigation and research during this project have been: -
-
-
-
-
Development of a steady state testing procedure for solar air collectors, suited for all types Discussion on physieaUy suitable efficiency presentations Development of different performance descriptions adequate for all common operation modes A comparison of available products Investigation of the technical behaviour of different types of air collectors Recommendations for an optimised utilisation of solar air collectors Recommendations for improvements of tested products Recommendations for a standardised testing procedure of air collectors, which can finally be integrated into an internationally starting standardisation process on ISO or CEN level. Adaptation of the existing solar-laboratory-facilities for testing solar air collectors
Testing requirements for solar-liquid collectors have been intensively developed since about 1980, now, beside an ISO (International Standardisation Organisation) standard there is also a new CEN (Comit6 Europden de Normalisation) testing standard under development. For solar-aircollector-testing there exists no standard. Compared to water collectors, the measuring procedure for solar aircollectors needs even more expenditure. Generally, measuring of air-temperatures and air-mass-flows requires higher effort for gaining comparable a c c ~ e s . Moreover, leakage, the air flow distribution inside the collector and the much lower heat transfer from the absorber to working fluid are further complex affects. Opposite liquid solar collectors the efficiency of solar air collectors is strongly influenced by the actual mass flow rate inside the collector due to the often rather low heat transfer coefficient between absorber and air. This heat transfer coefficient is highly dependent on the air speed. It is, therefore, often difficult or even impossible to extrapolate from tests of small modules of solar air collectors in test rigs to larger solar air collector arrays as the air flow pattern might be different.
The physical behavior of air collectors differs from liquid collectors mainly due to the much lower heat transfer and the lower heat capacity of air. Effects of uneven air flow pattern due to unproper installation and connection of the collectors with the manifolds in collector arrays makes the prediction of the solar gain difficult. The result of this work is a comparison of different collector types, further collector-improvement in collaboration with the manufacturers but also a detailed description on how to deal with the presentations of air collector efficiency.
3. EFFICIENCY PRESENTATIONS Optical features (absorptance and emittance of the absorber, transmittance of the cover), materials used (absorber material, cover material, frame, insulation) and constructing characteristics (mainly the airflow-principle and the effective heat transfer area) of the collector are of basic importance for the efficiency. However, the respective operation condition of the collector is decisive as well and the efficiency decreases with increasing temperatures within the collector because of the increasing heat losses. The efficiency of a solar-(air)-collector is defined as the ratio of useful gain of the collector (Qu) to the respective insolation Gr at the collector reference area A~.
111= Q u _
Qu Q sot AcGT
The useful gain is described by the massflow, the heat capacity and the temperature rise:
Qu =rn*c p * (To - Ti) As collector reference area can be considered: aperture area, absorber area or gross area. A general equation for solar collector performance is based on Hottel, Whillier and Bliss:
110= Fo['C(X-UL(To-T,)IQ,ol] where Ta is the ambient temperature and Fo is the collector heat removal factor in relation to To - the collector outlet temperature and TI0 is the efficiency when the outlet temperature is taken as reference. Fo accounts for the fact that the absorber temperature is not the same as the outlet air collector temperature neither in the horizontal, direction nor vertical.
ISES Solar World Congress 1999, Volume III
It appears that the efficiency rl depends on the operation conditions of the collector. It decreases with increasing temperatures, because of the increasing heat losses. Important is, how to define the operation conditions of the collector given by the temperature difference between the "overall collector temperature" Tk and ambient Ta. An efficiency curve can be drawn in dependency of a certain reference temperature which corresponds to the collector temperature Tk. In a physically correct way one has to take a weighted mean temperature Tk of the whole collector box, but in the measuring practice this is not practicable. That is why three temperatures are for choice: The inlet temperature (Ti), the outlet temperature (To) and a so called ,,mean" collector temperature (Tm) which can be calculated as the arithmetic mean value between inlet and outlet temperature. Efficiency curves of a solar collector corresponding to the three possible reference temperatures for a constant mass-flowrate are shown below:
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Measurements indicate that the physical mean temperature (Tk) of the collector is often much closer to the outlet temperature (To) than to the arithmetical mean temperature (Tm). Therefore the presentation of the collector efficiency curve using the outlet temperature (To) often seems to be the best solution. A proposal for doing this is also given by DuffleBeckmanr~ (Solar engineering of thermal processes, J.Wiley&Sons Inter-science, New York 1991)
4. AIR C O L I ~ C T O R TYPES For different applications of air heating systems (space heating, preconditioning, drying-processes.... ) different collector types are the best choice each. Moreover, the result of this project was a contribution to the handbook for designing solar air systems which is one main output of the IEA-Task 19 collaboration, as well as recommendations for standardising of air collector-testing procedures and an input for further discussions on evaluating the performance of solar air collectors. In principle there are 4 different construction modes of solar air collectors: -
-
Flow below absorber, "Underflow'' (mass-flow behind the absorber, the air gap between absorber and cover operates as insulation) Flow above absorber, "Overflow" (mass flow only between absorber and cover) Flow on both sides of the absorber Perforated absorber, the air flow penetrates through the absorber (black felt, porous metal .... )
Fig. 1 Efficiency related to different Reference Temperatures For solar liquid collector it is the custom to present the efficiency related to the "mean collector temperature" (Tm) representative for the heat losses of the collector. For liquid collectors, where the temperature difference between inlet and outlet is very small (normally less than 10 K) and the heat transmission from the absorber to the fluid is high, the arithmetical mean value (Tm) is in fact very close to the physical mean temperature (Tk) of the collector.
For air collectors the difference between inlet and outlet can be up to 30K or 40K dependent on the mass flow. Also important is the amount of heat transmission from the absorber to the fluid, which is for air collectors usually not that high. Due to these effects there will be no longer a linear increase of the fluid temperature along the absorber plate and the arithmetical mean value (Tm) is often not representative for the heat losses of the collector.
Fig. 2 Air flow principles in Solar Air Collectors
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ISES Solar World Congress 1999, Volume III
Influence of different air flow principles:
Advantages and disadvantages: Flow above absorber: + Simple construction - High losses, especially at a high difference between absorber and ambient temperature - Highly decreasing efficiency at high surrounding air velocities (mainly depending on the cover) - Only one surface of the absorber is used as effective heat transfer area - Double glazing reduces the losses but decreases the solar input Flow below absorber: + Air in the gap between absorber and glazing operates as insulation (few losses at high differences between absorber and ambient temperature) - Only one surface of the absorber is used as effective heat transfer area Flow on both sides of the absorber: + Double effective heat transfer area - At high differences between absorber and ambient temperature the heat losses due to the hot air directly under the cover increase and the dependency of surrounding air speed decrease Porous absorber: + High heat transfer-coefficient - High pressure drop - Depending on ambient air conditions (dust, pollution...) the absorber is often under a high technical stress.
3...unglazed perforated trapezoid absorber panel, aluminimn, anthrazit, strongly dependent on wind 4...glazed plane absorber, facade element, underflow 5...glazed, rippled absorber, air flow on both sides 6...glazed plane absorber, facade element, air flow on both sides 7...glazed site built collector, selective absorber, trapezoid profile, underflow 8...glazed plane absorber, facade element, underflow From liquid systems we know that judging a solar system is often to much concentrated on the thermal performance of the collector; other features of the system like control strategy, mounting of temperature sensors, connecting the modules, storages, insulation matters and many other questions should also be considered carefully. But this project focused on the assessment of the features of air collectors. As a result of these investigations a lot of hints can be derived how to built a solar air collector for a certain application. 5. OPERATING SOLAR AIR COLLECOTRS If nmning a solar aircollector usually 4 multiple combined effects must be considered: A) The higher the mass-flowrates the higher the efficiency (reason: at higher mass-flowrates two combined effects occur: the heat transfer from the absorber to air increases and the Outlet temperature- and therefore the heat losses decrease) B) The effect of air flow leakage increase with the air flow-rate C) The electrical power requirement for the fan increase with the mass-flowrate D) For heating purposes a certain temperature level is often needed, which further restricts the possible mass-flowrates
Fig. 3 Efficiency of Solar Air Collectors 1...glazed collector (low iron), aluminium absorber with uprofiles, selective coating, underfiow 2...glazed collector, black textile absorber
Fig.4 Temperature Rise and Efficiency Large effective heat transfer-areas are advantageous but constructions where the air is forced to flow in tight profiles
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causes high pressure drops. It is a challenge for the constructing engineer to find a compromise in high heat transfer and low pressure drop.
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conditions. The leakage rate depending on the mean static pressure of the collector should be determined generally.
Air-flow pattern: Pressure drop:
The air flow pattern inside the collector is very important for a correct assessment of the performance. Normally, if you work with one collector only, the air flow pattern near inlet and outlet is often not satisfactory. To reach an even air flow pattern from inlet to outlet for each tested collector a special connection box was built.
The Pressure drop is important for the number of collectors in series and the electrical power of the fan. Pressure drop is about a square function of air velocity. Pressure drop increases about linear with air density.
Temperature measurements: During an engineering process for solar air systems, choosing an air collector should be based considering the following aspects: -
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The desired temperature rise A low temperature rise (f.e. in stores, factory halls, sport halls, drying and preconditioning processes...) otten favours cheap collector constructions (simple absorber profile, cheap or even no glazing, minor insulation) A high temperature rise (for heating offices, living rooms .... ) mainly need high performing constructions The design and optical features Location, climate, orientation The costs Reasonable pressure drops; (Optimisation according to the temperature rise needed)
While the measurement of the inlet temperature is rather easy, to achieve a precise determination of the outlet temperature is difficult. Several layers of different air temperatures are often close adjacent, a ~ i f i c mixing device- optimized according to fluid dynamic experiences - at the outgoing duct just in front of the sensors and a sophisticated arrangement of temperaturesensors are necessary. Effects of condensation must be considered carefully.
Conditioning: For testing the collector with different air temperatures, the preparation of differently conditioned air is needed. An enclosed climatic chamber with 150 m 3 met these require-ments. Temperatures between some degrees below zero and up to 60 ~ could be reached at the collector inlet. Control devices and various control heating devices cared for stability.
7. C O N C L U S I O N 6. EXPERIENCES MADE CONCERNING THE TESTING PROCEDURE AIR COLLECTOR TESTING -
R
E
C
O
M
M
E
N
D
A
T
I
O
N
S
F O R
No standardised testing procedure exist for Solar Aircollectortesting so far. Starting a standardisation process for testing solar aircollectors has been already discussed in the Technical Committee 180 of the International Standardisation Organisation (ISO), but work is still resting. (Reuss et al, 1993) Generally, measuring of air-temperatures and air-mass-flows requires higher effort for gaining satisfactory accuracies. Moreover, leakage, the air flow pattern inside the collector and the much lower heat transfer from the absorber to the heattransfer-medium are further complex affects.
Leakage: For an accurate measuring process 2 fans are needed, one at the inlet and one at the outlet. For testing reasons the fans had to work such, that the mean static pressure at the collector is equal atmospheric pressure. Only by that you can manage the leakage rate to be minimised. It was also possible to carry out tests with only one fan near inlet or outlet the collector to simulate realistic
Solar air collectors are not wide spread so far. As main obstacle for a wide dissemination appears lacking information as well as lack of confidence on how these systems will work. Testing of the respective components is therefore essential. Test results from independent test institutes can serve as efficiency proof. Testing within this project showed a wide range of effectivity among the different products. Experiences how to optimise a collector as well as precise recommendations for a testing procedure of air collectors are result of this project. Further Investigations seem to be necessary in the general issue of presenting the thermal energy output of air collectors. The problem with reference a r e a - well known from liquid collectors as well as the problem of the reference temperature are open for further discussions.
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REFERENCES 1. Duffle J. & W. Beckman. ,,Solar Engineering of Thermal Processes "' J . Wiley & Sons Interscience, New York 1991) 2. Morhenne J. & M. Fiebig. Entwicldung und Erprobung einer Baureihe yon optimierten, modularen Solarlufierhitzern .f'dr Heizung und Trocknung, Ruhr-Universit~t Bochum, 1990. 3. Dai, Hui and Li. Fully developed laminar flow and heat transfer in the passages of V-corrugated solar air heater, ISES "91, Denver, proceedings. 4. Lo, S. N. G., Deal, C. R. & B. Norton..4 School Building Reclad with Thermosyphoning air panels, Solar Energy Vol. 52, No. 1, pp. 49-58, 1994. 5. Biondi, P., Cicala L. & G. Farina. Performance analysis of solar air heaters of conventional design, Solar Energy, Vol. 41, No. 1, pp. 101-107, 1988. 6. ,41tfeld, K. Leiner, W. & M. Fiebig. Second Law optimisation of flat-plate solar air heaters, Part 1: The concept of net exergy flow and the modelling of solar air heaters. Solar Energy 41, 127-132, 1998. 7. ,41tfeld, K~ Leiner, W. & M. Fiebig. Second Law optimisation of flat-plate solar air heaters, Part 2: Results of optimisation of and analysis of sensibility to variations of operating conditions. Solar Energy 41, 309-317, 1998. 8. Reuss, M. Recommendations for standard procedures for testing of air heating solar collectors, Bayrische Landesanstalt fLandtechnik, ,4ugust 1993 (ISO TC 180/SC5/N53) 9. Corazza ,4., et. al. Design, development and performance studies of a large sized solar air heater in nonconventional mode of operation. Int. Conf. ,41ternative Energy Sources Today and for the 21st century. Brioni, oct. 5-8, 1988. 10. CE-Standard of solar collectors, Thermal solar systems and components - Collector- General requirements, CEN TC 312 N164, a draft paper by CE TC 312-PT1. 11. Gupta D., Solanki, S.C. and J.S. Saini, Thermohydraulic performance of solar air heaters with roughened absorber plates. Solar Energy Vol. 61, No. 1, pp. 33-42, 1997. 12. Keller, J., V. Kyburz and `4. K6lliker, Untersuchungen an Lufikollektoren zu Heiz- und Trocknungszwecken, 1988. Schlussbereicht des KWF-Projektes Nr. 1296. Paul Scherrer Institut, W'renlingen und Villingen, CH-5232 Vilh'ngen PSI. 13. ,4bbud, I..4.., G.O.G. L6f and D.C. Hittle, Simulation of solar air heating at constant temperature. Solar Energy Vol. 54, No. 2, pp. 75-83, 1995. 14. Matrawy, K.K., Theoretical analysis for an air heater with a box-type absorber. Solar Energy Vol. 63, No. 3, pp. 191-198, 1998. 15. Jensen, S.O., Roof Space Collector, Validations and simulations with EMGP2. 1987a. Institute for Energy and Building, Technical University of Denmark, Report No. 87-15. 16. Jensen, S.0., O. Olesen and F. Kristiansen. Lufi/vceskesolfangee. 1987b. Solar Energy Centre Denmark, DTI Energy. ISBN: 87-7756-470- 7.
17. Morck, O. and P. Kofod, Udvikling af luftsolfanger. 1993, Cenergia Energy Consultants, Denmark. ISBN 8 7-90314-02-6. 18. Muff, Christoph; Solarluft~steme- Vortrag Trisolar 98 Bregenz/,4 ustria 19. IEA Solar Heating and Cooling Programme, 1998 Annual report with a feature on Solar ,4ir Heating, Morse Ass. Inc. 1808 Corcoran Street, N.W. Washington, DC 20009, USA, March 1999.
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AN EMPIRICAL HEAT TRANSFER EQUATION FOR THE TRANSPIRED SOLAR COLLECTORS, INCLUDING NO-WIND CONDITIONS Gerald W. E. Van Decker Active Solar Energy Technologies, Natural Resources Canada, 580 Booth St. Ottawa, Ont. Canada, K1A 0E6, 613 996-3648, 613 996-9416,
[email protected] K. G. Terry. Hollands Department of Mechanical Engineering, University of Waterloo, Waterloo, Ont., Canada N2L 3G1, 519 888-4053, 519-746-0852, kholland @ solar 1.uwatedoo.ca
Abstract - The unglazed transpired solar collector is now a well-recognized solar air heater for heating outside air directly. Example applications include pre-heating ventilation air and heating air for crop drying. The outside air in question is drawn straight from ambient, uniformly through the whole surface of a perforated blackened plate (the absorber plate) exposed to the sun. An important parameter fixing the collector's efficiency is the heat exchange effectiveness, ~. Once e is known, finding the collector efficiency is straightforward. Recently, Van Decker et al. presented measurements of e under various wind speeds and suction velocities plates perforated with circular holes of various diameters and spacings, laid out on either a square or triangular layout. They also developed a predictive equations for ~, which contained various parameters adjusted to fit their measurements, but their equation did not cover wind speeds down to zero (still air). The present paper extends that earlier work so as to cover the zero wind speed case. This new model predicts the measured data of Van Decker et al. andKutscher with a root mean square error of about 5.5 %. (This degree of uncertainty would lead to a contribution to the error in the predicted efficiency of roughly 2.5%.) The model also gives the breakdown of the contribution to the total heat transfer from each of the plate regions: the front, the hole and the back. 1. INTRODUCTION Unglazed, transpired solar collectors (HoUick and Peter, 1990; Kutscher et al., 1991, 1993, 1997) have been the subject of a number of investigations. They are effective devices for applications where outside air is to be heated directly, such as in heating ventilation air for buildings and crop drying. The outside air in question is drawn straight from ambient, through the whole surface of a transpired, darkened plate (the absorber plate). Tests conducted on several installations indicate that the unglazed transpired collector (UTC) gives annual solar collection efficiencies reaching 72% (Carpenter and Kokko, 1991). Typical installations have simple paypack periods of 2-8 years, making the UTC an attractive investment. Over 70 large systems each having collector areas between 500 to 10,000 m 2 have been installed and are successfully operating for fresh-air heating in Canada, the United States, Germany, and Japan and heating process air for crop drying in countries throughout the world.
1.1 Prior Work on Heat Transfer Modeling A recent review of the heat transfer principles of the UTC has been given by Hollands (1998). Kutscher et al. (1991, 1993) determined that, as the air travels across the face of the collector (driven by the wind), the thermal and velocity boundary layer thicknesses reach an asymptotic value at a very short distance from the edge of the plate (about 0.1 m), so that almost all the plate is in the asymptotic region. This fact laid the foundation for their performance model. When one approximates the radiation loss by a linear equation (HoUands, 1998), the Kutscher et al. (1991) model for the collector efficiency r/reduces to
s/(1
h /(
CpV s ) )
r
where txs is the solar absorptivity of the plate, hr is the radiative heat loss coefficient, Vs is the superficial suction velocity (volumetric rate at which air is sucked through plate, per unit area of plate), eis a "heat exchange effectiveness" (see below), and p and Cp are the density and specific heat of the air, respectively.
Values of r/ranging from 50 to 80% are common in practice, and typically Vs ranges from 0.03 to 0.08 m/s. A key item in the efficiency equation is the quantity e, defined by -
ro-r. L-T
,
(1)
where To is the mean air temperature leaving the plate at the backside and p is the plate temperature, and T, is the ambient air tempearture. Effectiveness e has to be evaluated, say from experimental data, or from analysis, but once it has been evaluated, determining r/is straight forward. Based on his extensive measurements on relatively thin plates, Kutscher (1994) presented a predictive model for e, for UTCs with circular holes on a triangular layout. Cao et al. (1993) and Golneshan and Hollands (1998) reported numerical and experimental correlation equations for e for a plate with perforations consisting of an array of slits, with the wind flow assumed to be transverse to the slits. Using computational fluid dynamics (CFD), Arulanandam (1995) (see also Arulanandam et al. (1999)) obtained a correlation equation for e, applying to a plate with circular holes on a square layout, but only under no wind. Van Decker et al (1996) reported an extensive set of measurements on thin and thick plates with circular holes on a square or triangular layou, over a range of typical wind speeds, as well as a correlation equation for ~, fitted to their measured data. 1.2 Present Study The present paper extends the treatment of Van Decker et al, to cover the no-wind conditions. It presents a new model for e that is shown to predict both Kutscher's data and that of Van Decker et al., within an rms error of roughly 5.5 %. Like the Van Decker et al model, the new model gives a breakdown of the contribution to the heat transfer on each of the parts of the plate: the outside face, the hole, and the back of the plate. The breakdown given in the present paper is expected to be more accurate than that given
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Table 1: Characteristics of the Plates Tested by Van Decker et al. (1996) Plate No. 1 2 3 4 5 6 7 8 9
Plate Material
Aluminum Aluminum Polyvinyl Chloride Stainless Steel Polyvinyl Chloride Polyvinyl Chloride Polyvinyl Chloride Stainless Steel Polyvinyl Chloride
Hole Pitch* P (ram) 16.89 16.89 16.89 13.33 13.33 8.00 24.00 24.00 6.67
Diam. D (ram)
Thickness t (mm) 1.60 1.60 1.60 1.60 0.79 1.20 3.60 3.60 0.93
Plate Thermal Conductivity, k, W/(mK) 0.86 0.86 1.69 1.57 3.11 6.51 1.60 0.57 1.97
186 186 0.149 15.12 0.149 0.149 0.149 15.12 0.149
*Shortest distance between two holes.
by Van Decker et al.
properties with one from Kutscher' s plates, good agreement was observed with the data of Kutscher.
2. REVIEW OF STUDY OF VAN DECKER ET AL.
2.1 Apparatus and Method A description of the apparatus and method used by Van Decker et al. is given in their paper (Van Decker et al., 1996).We give a only brief review of it here. Aside from the properties of air, which are more or less constant, the heat exchange effectiveness in the asymptotic range depends on five plate parameters: minimum distance between holes (called the pitch), hole diameter, plate thickness, and thermal conductivity, and two velocities: the suction velocity and wind velocity. The geometric properties of the nine 60cm by 60 cm plates they tested are listed in Table 1. The properties of Plate 1 were made identical to a plate in Kutscher' s study, in order to permit a direct comparison. The remaining plates had holes with the square geometry, both to provide experimental data on a layout other than the triangular hole layout and so that the CFD model of Arulanandam (1995) could be used in the interpretation. The asymptotic performance of each plate was measured on a test rig (Golneshan and Hollands, 1998) Primary components included: a solar simulator (or short wave radiant heat source), an air suction system, and a wind tunnel. Each plate was installed on a suction plenum, which divided the plate into 7 sections of equal area. By limiting the measurements to the downstream plenums, the asymptotic conditions were enforced. Temperatures of the plate, upstream air and outlet air were measured with thermocouples. Each plate (except Plate 2) was tested over
2.3 Model Development Preliminaries Van Decker et al. developed a model for their data, as follows. They assumed that the plate is isothermal, with a single temperature Tp. This was supported by direct measuremnts and by the fact that plate thermal conductivity was not found to be an important parameter. The air-heating by the plate takes place in three regions: the front-of-plate, the hole, and the back-of-plate. Each region was assigned an effectiveness (denoted ep eh, andeb, respectively), as follows:
to, -r Ef - T - T =
ro~ -fro, , ~h ~
ro -ro~ , % -= ~
T-Trol
L-To2
(2)
where (referring to Figure 1) To1 is the bulk mean temperature of the air as it enters the hole and To2 is the bulk mean temperature of the air as it exits the hole. They also defined a combined effectiveness ejh for the front of the plate plus the hole as
%-
to2 - r . T,,- r
"
(3)
the same set of wind and suction velocities: for the wind velocity this set ranged from 0.0 m/s (no wind) to 5.0 m/s; and for the suction velocity the set ranged from 0.028 m/s to 0.083 m/s. The error limits of e were estimated to be _+0.019.
2.2 Results The observed values of effectiveness e ranged from 0.32 to 0.91. The effectiveness was found to decrease with increasing Vs, P, and D, and to increase with increasing Uw and with t. Thermal conductivity, k, was found to have a very weak effect. For the plate (Plate 1) with the common
Figure 1. Sketch of the plate, showing the three temperatures, To1, I"o2,and To~.
ISES Solar World Congress 1999, Volume III
geometry, as will be explained below. From Eqns (1), (2), and (3) it was easy to show that
% = 1 - (1 - ~ ) / ( 1
and
-c#).
e = 1 - (1 - ef) (1 - e h) (1 - %)
(4)
(5)
2.4 Back-of-Plate Model Arulanandam (1995) (see also Aralandam et al. (1999))had used a computational fluid dynamics (CFD) code to model the flow and heat transfer on the front face and in the hole for the no-wind situation, and correlated the results by an single equation for ejh. For the same conditions, the resulting model was found to predict values efh that are consistently less than the corresponding measured c, this was expected since the model did not include the additional temperature rise (To - I"2) associated with the back-of-plate heat transfer. The difference between the measured e and the computed elh was roughly constant, at about 0.18. For each of their no-wind data-point, Van Decker et al. determined the e~ as predicted by Arulanandam, and substituted this and their measured e into Eqn (5) to get a corresponding value for eb. These results were then correlated by a model for the back-of-plate heat transfer giving the following equation for eb
% - ~
1 + dRe~
'
(6)
where Reynolds number Reb is equal tO VhP/V where Vh is the velocity in the hole, (=Vs4p2/nrD2), and coefficients d and e being found to be equal to 0.144 and 1/3, respectively. The authors then assumed, for the purpose of model development, that eb is independent of U,. This meant that Eqn. (6) could be used for eb for all wind speeds. 2.5 Front-of-Plate Model For a plate perforated by long slots rather than circular holes, Golneshan (1994) has described a 2D momentumintegral analysis that predicts the heat transfer on the front of the plate under conditions of significant wind. He found that eI should depend upon only parameter, y, as defined by:
y =Re)lRe = V)P I Uwv ,
(7)
where the Reynolds numbers Rew and Res are given by Re,, =U,cP/v. and Res=V~P/v, respectively. In particular, he obtained eI = 1 - ( a + by-it2) -1 ,
(8)
where a and b are constants. Van Decker et al. adapted this model to the plate with circular holes, adjusting the values of the coefficients a and b, to values appropriate to this
2.6 Hole Model Since the hole Reynolds number Reh = VhD / v is found to be much less than 2000, Van Decker et al. assumed laminar flow in the hole. The flow in the hole is (hydrodynamically and thermally) developing flow in an isothermal circular tube, but with the fluid (air) entering the hole being non-isothermal: the temperature nearer the plate is hotter than that nearer the centre-line. The ample literature information on the heat transfer when the entering fluid is isothermal guided their model development. They assumed a linear fit for the Nusselt number as a function of the Graetz Gr, obtaining ~h = 1 - e x p { - 4 ( c ( P / D ) + 3 . 6 6 ( D ) Rehlpr-1)},(9)
in which c was a constant to be determined. 2.7 Correlation With appropriate values of a, b and c, Eqns (4), (9)constituted a model for e, which, was fit to all the "windy data" (i.e., all the data points for which the wind speed was not zero), the constants a, b and c being adjusted so that the root mean squared difference between the data and the model was minimized. The result gave a=0.8434, b--0.4867, and c=0.00665. This model was found to fit the windy data with an rms error of 5.2%. 3. N E W
MODEL
For the conditions of no-wind or low values of wind speed U, the above model of Van Decker et al. gives unreasonable, negative answers. To overcome this problem, Van Decker et al. suggested switching to the Aralanandam equations for the no-wind conditions, but such a strategy could not in itself make clear the range of U, where one uses one equation and the range where one would use the other. Also it is much more convenient to have just one equation to cover the full range of wind speeds of interest (as does the recommended equation of Kutscher). So we undertook to develop such an equation. 3.1 Model Development For the conditions of no wind, y is equal to infinity, and Eqn (8) gives negative values of %, which leads to unrealistic negative values for e.. Indeed, negative values for e at zero wind speed will always be predicted by the model whenever a is made to be less than unity. When a is put equal to unity, Eqn (8) yields ei= 0 for zero wind. In fitting coefficients a, b, and c to their data, Van Decker et al. had noted that forcing a =1 gives almost as small an rms error in the model as letting a be free to take on any value. On the other hand, letting a take on increasingly large values greater than unity gives greater and greater error. In subsequent developments, we have found that building a comprehensive model on one that predicts negative e under no-wind circumstances was unwise. So we fixed a at unity and proceeded with the no-wind model development from there. With a equal to unity Eqn. (8) reduces to
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ISES Solar World Congress 1999, Volume III
26
r = (1 + eylt2) -1 ,
(10)
where e = lib. It was felt that (with the possible exception of their respective parameters settings, which might need slight adjustments) the Van Decker et al. models for the hole and back-of-plate heat transfer are quite satisfactory. Thus the model that needs revision is the model for the front of plate heat transfer: that is, Eqn (10) for eI. Using the Van Decker et al. models for eh and eb, it is possible calculate what eI would have to be to get the values of e that was actually measured at no wind conditions, and we did this for every no wind data point. Then from dimensional arguments, we concluded that this eI should be mainly dependent on the suction velocity Reynolds number, Res=V~/v. So we correlated this eI against Re,. A strong correlation was found, and moreover it was one that could be fitted by an equation very similar to Eqn. (10): ey=(1 +fRe,) -1
(11)
with Res replacing yln, and a new coefficient f which was fitted to the no-wind data, giving f = 0.0654. When it is recalled that y = Re,2/Re, we see that we can express Eqn (10) as (~f =(1 + eRe,Rew-lt2) -1
Effectiveness Measured by Van Decker
Figure 2. Hot of both windy and no-wind data of Van Decker et al (1996) in the form of measured effectiveness vs. that predicted by the model of the present paper, namely Eq. (14). The Legend refers to the Plates listed in Table 1. values of 6. The result gave e=1.42, b=O.0400, c=0.00510, and d=0.294. So, in summary, the new model for effectiveness model for a UTC with a square hole geometry is given by:
(12) =1-
1-
max[Re -1/2 , 0.04]
1+1.412Re $
Eqn (12) applies to the windy situation and Eqn.(11) for the no-wind situation. We propose combining them as in the following equation
ex~_4(O.Oo51P + 3.66
D PrRehD el= 1 + eResmin[Re~lt2,b]
W
)HI_
1 ,] 1 +0.294Re;/~
(14)
(13)
where b is equal to fie, and the symbol "min[x,y]" means that one is supposed to take the minimum of x and y. Because of the nature ofthe min[x,y] function, one finds that this equation automatically shifts to Eqn. (11) under no wind (Rew = 0) conditions, while Eqn (12) is found to apply whenever Rew is greater than b 2, which was found to be about 700, and since almost all of the data points under windy conditions had Re, greater than 700, the model reduces to the V an Decker et al. (1996) model for the windy cases of their experiments. 3.2 Fitting the Model to the Square-pitched Data The entire model for e , as given by Eqns (5), (6), (9), and (13), was fitted to the full data-set containing measurements on plates with holes on a square layout, both windy and not, to obtain the values of the parameters e, b, c, and d that minimized the sum of the square of the deviations between the model for e and the measured
This model was found to fit all the data with a root mean square error of 4.3%. Figure 2 compares the data to the model, by plotting the prediction on one axis and the measured on the other. 3.3 Testing the Model on Triangular-pitched Plates Van Decker et al' s Hate 1 had the same values of t, D, k, and P (P being the shortest distance between two holes) as Plate 2, but the holes were laid out in a triangular arrangement, rather than square. For a given Vs and U~, e for Plate 1 was found to be about 0.05 greater than for Plate 2. That is, other things being equal, the triangular layout performs slightly better. Van Decker et al (1996) found that the same model could be used for both plates if one uses an appropriatelyadjusted value for the pitch. That is a triangular-plate model (like that developed by Kutscher) can be used for a square pitched plate if one uses for the pitch P a value that is ~" times the square layout pitch, where scaling factor ~ is equal to 1.6. Conversely, a square-plate model (like Eqn
ISES Solar World Congress 1999, Volume III
placements It captures the effect of a range of variables: suction velocity Vs, wind velocity U~, hole pitch P, hole diameter D, and plate thickness t, having been tested over the following ranges of these variables: 0.028 m/s _< Vs -< 0.083 m/s; 0 m/s _< U| _<5.0 re~s; 7 mm <_P <_24 mm; 0.8 mm <__D <_3.6 mm; 0.6 mm <_t <_6.5 mm, as well as on plates having thermal conductivity k ranging from 0.15 W/mK to 200 W/mK. Under typical operating conditions, about 62% of the ultimate temperature rise of the air is predicted to occur on the front surface, 28% in the hole, and 10% on the back of the plate. REFERENCES Arulanandam S.J. (1995) A Numerical Investigation of Unglazed Transpired-Plate Solar Collectors under Zero-Wind Conditions, M.A.Sc. Thesis, University of Waterloo, Waterloo, Canada. Effectiveness Measured by Kutscher
Figure 3. Plot of both windy and no-wind data of Kutscher (1994) in the form of measured effectiveness vs. that predicted by the model of the present paper, namely Eq. (14). (14)) can be used for a triangular pitched plate if one uses for the pitch P a value that is l/~'times the triangular layout pitch, where, again, scaling factor ~" is equal to 1.6. This means that Eqn (14), with this proviso, predicts results for triangular plates, so we tested it on Kutscher's data, and obtained the comparisons shown in Fig 3. Excellent agreement was found. The root mean square deviation between the predictions and the measured data were found to be 6.6%, when expressed as a per cent. Also the mean bias error in the model was found to be essentially zero. Thus Eqn (14) predicts Kutscher's data with essentially the same fidelity as it fits the data of Van Decker et al. 3.4 Relative Contributions of Regions A useful feature of the model is that it provides information on the contribution to the overall heat transfer provided by each of the three regions: the front-of-plate, the hole, and the back-of-plate. The important contribution of the back-of-plate and the holes is to be particularly noted. As an example, suppose we consider the following settings, which are representative of a commercial solar collector plate operating under average suction and wind speed conditions: P = 16.9 mm, D = 1.60 mm~ t = 0.8 mm, Vs = 0.04 m/s, and U**= 2.4 m/s. Under these conditions, the model predicts es = 0.405, eh = 0.306, eb = 0.149, e~ = 0.588, and e = 0.649. This means that 62% of the ultimate temperature rise of the air occurs on the front surface, 28% occurs in the hole, and 10% occurs on the back. This information constitutes useful knowledge for those trying to optimize plate or to develop improved plates.
4. CONCLUSIONS Equation (14) provides a suitable model for the thermal effectiveness e for UTCs, including the case of no-wind. With suitable interpretation of the measure of pitch, it is useful for plates laid out with square and triangular hole
Arulanandam S.J. Hollands, K.G.T., and Brundrett, E. (1999) A CFD Heat Transfer Analysis of the Transpired Solar Collector under No-Wind Conditions, Proceedings, Solar World Congress ,1999: Biennial Meeting of the International Solar Energy Society, Jerusalem Israel, Carpenter S.C. and Kokko J.P. (1991) Performance of Solar Preheated Ventilation Air Systems, Proceedings Annual Conference of Solar Energy Society of Canada Inc., Toronto, Canada, Solar Energy Society of Canada Inc., Ottawa, pp. 261-265. Cao S., Hollands K.G.T. and Brundrett E. (1993) Heat Exchange Effectiveness of Unglazed Transpired-Plate Solar Collector in 2D Flow. Proceedings of ISES Solar World Congress 1993, Budapest, Hungary, 5, pp. 351-366. Golneshan A.A. (1994) Forced Convection Heat Transfer from Low Porosity Slotted Transpired Plates, Ph.D. Thesis, University of Waterloo, Waterloo, Canada. Golneshan, A.A. and Hollands, K.G.T. (1998)Experiments on Forced Convection Heat Transfer from Slotted Transpired Plates, Proceedings CSME Forum 1998, Toronto Canada, Canadian Society for Mechanical Engineering, Hamilton, Canada Volume 1, pp78-88. Hollands, K.G.T. (1998) Principles of the Transpired-plate Air Heating Collector: the Solarwall, Renewable Energy Technologies in Cold Climates'98 (Incorporating the 1998 Annual Conference of the Solar Energy Society of Canada Inc.) The Solar Energy Society of Canada Inc., Ottawa, pp 139-144. Hollick J. C. and Peter R. W. (1990) United States Patent No. 4,934,338. Kutscher C. F. (1994) Heat Exchanger Effectiveness and Pressure Drop for Air Flow Through Perforated Plates With and Without Crosswind, J. of Heat Transfer, 116, 391-399. Kutscher C. F., Christensen C., and Barker, G. (1991) Unglazed transpired solar collectors: an analytic model and test results. Proceedings of ISES Solar Worm Congress 1991, Pergamon Press, Vol. 2, Part 1, pp. 1245-1250.
27
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ISES Solar World Congress 1999, Volume III
Kutscher C. F., Christensen C., and Barker, G. (1993) Unglazed transpired solar collectors: heat loss theory. ASME J. ofSolar Eng, 115, No. 3, 182-188.
ISES Solar World Congress 1999, Volume III
A CFD HEAT TRANSFER ANALYSIS OF THE TRANSPIRED SOLAR COLLECTOR UNDER NO-WIND CONDITIONS
S. J. Arulanandam Department of Mechanical Engineering, University of Alberta, Edmonton, AB., Canada, T6G 2E5, 780 492 3598
[email protected] K' G. Terry, Hollands and E. Brundrett Department of Mechanical Engineering, University of Waterloo, Waterloo, ON, Canada, N2L 3G1, 519 888 4053, 519 746 0852, kholland@solarl .uwaterl00.ca,
[email protected]
Abstract - The unglazed transpired solar collector is now a well-recognised solar air heater for heating outside air directly. Example applications include pre-heating ventilation air and heating air for crop drying. The outside air in question is sucked straight from ambient, uniformly through the whole surface of a perforated blackened plate (the absorber plate) exposed to the sun. An important parameter fixing the collector's efficiency is the heat exchange effectiveness, e. Once e is known, finding the collector efficiency is straightforward. The effectiveness depends on the wind speed, the suction velocity, and the plate geometry. This paper is about determining this effectiveness by computational fluid mechanics (CFD), for conditions of no wind. The computational domain extended over a representative element including one hole and the region immediately adjacent it, extending to half the distance between the holes. Simulations were carried out over a wide range of conditions, and the results are incorporated into a correlation model, which will be reported in the full paper. Because of the no-wind assumptions, the model is of limited direct use, but when combined with experimental data, the model has permitted a wideranging correlation equation to be obtained by other workers
1.0 INTRODUCTION Unglazed transpired-plate solar air-heating collectors have been the subject of a number of recent investigations. They are effective devices for cases where outside air is to be heated directly, such as in heating ventilation air for buildings and for drying applications. The outside air in question is drawn straight from ambient, through the entire surface of a perforated blackened plate. The glazing, traditionally used for reducing the plate's radiant and convective losses, is dispensed with in this collector. It is not required, because the convective boundary layer is continually sucked off, thus virtually eliminating the convective loss. The intimate heat transfer between the plate and the sucked air keeps the plate temperature low, minimizing the radiant loss. In commercial production, these collectors have been built to cover areas on the sides of buildings of the order of thousands of square metres. The available heat transfer theory (Kutscher et al, 1993; Hollands, 1998) allows one to predict the collector efficiency, but only once a quantity called the heat exchange effectiveness, e, has been specified. Defined as the actual temperature rise of the air as it passes through the plate divided by the maximum possible temperature rise, e is a measure of the intimacy of the convective heat transfer between the plate and the air, and can be determined by purely convective considerations and/or analyses. Most published studies have determined e experimentally, (Kutscher, 1994; Van Decker et al., 1996, 1999.) This paper is about determining this effectiveness by computational fluid mechanics (CFD). CFD studies of the
problem have been carried out before. Kutscher (1992) carded out some CFD studies, but the range of parametric settings did not extend over the full range of interest. The CFD studies of Cao et al. (1993) restricted their work to the case where the perforations are in the form of long slits rather than circular holes. Because of the limitations of CFD simulations in the present study, it was necessary to put restrictive assumptions on the problem statement: the assumption of no-wind conditions and the exclusion of the effect of the heat transfer on the back of the plate. Simulations were carried out over a wide range of conditions, and the results are incorporated into a correlation model. Because of the assumptions, the model is of little direct use, but when combined with experimental data, it has permitted a wide-ranging correlation equation to be obtained (Van Decker et al., 1996, 1999). 2.
MODEL DEVELOPMENT
2.1 Defining the Domain
Figure l(a) is a sketch of the transpired plate with circular holes on a square pitch arrangement in zero win& The spacing between the holes centres is the hole pitch P. Since every hole should behave in a similar manner, one need only study a representative element, like the one shown in a dotted line in Figure l(a). Several planes of symmetry, introduced by the fact that there is no wind, imply that one need only study the reduced domain sketched in Figure l(b) with symmetry boundary conditions applying at the planes z = 0, z = P/2, y = O, and y = P/2. The computational domain is also bounded by the planes x = 0 and x = x,, In the drawing, Vs represents the suction velocity, or the volumetric rate at
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ISES Solar World Congress 1999, Volume III
30
~ . . . . . .
/
The other part of the x = 0 plane is the flow outlet plane, which was given the outflow boundary condition (Patankar, 1980):
el~
~/T~platv
OT = O, u > O f o r at x = x o, -~x
,OO OooO or176 i1 o~ /
//
/lff
/
(
,/,
| ~i
',
I
.~y
/
The boundary conditions symmetry are as follows:
// /
X
is
;
at y = O o r
i11 III
y2 + z 2 < D 2 / 4
along
y=Pl2,_~=O,v=w=O
~_~=O,u>O
',/
vs
the planes
(3) of
(4)
~T
at z = O or z = P/2,-~z = O, v = w = O -.~z = O, u > O (5)
For the front surface of the plate, we have t/'
,o,o,o, I/'i ,,,11
at x = t and y2 + z 2 > D 2 / 4 ,
,,,ll,"
/// i'N ,'/~ii'___
f /
k 0_~
V = (O,O,O), a G = - ~ ~ lt_
which air is sucked through the plate per unit plate area. At distance x**from the plate this air stream is assumed to be in uniform motion. 2.2 Governing Equations and Boundary Conditions The relevant governing equations for the air velocity V(u,v,w) and the temperature T are the equations for the conservation of mass, momentum, and energy, assuming laminar flow, steady conditions, constant properties, no external forces and no viscous heat dissipation. These are given in textbooks (e.g., Arpaei and Larsen, 1984) and will not be repeated here. For the region containing the solid, (i.e., the plate), the governing equation is simply the steady heat diffusion equation. The inlet boundary condition is the plane x = x., Ideally, this plane would be located an infinite distance from the plate, but any CFD representation has to place it at a finite distance of magnitude sufficiently large to realistically represent the infinite distance. The following boundary conditions were assumed to apply there:
P=Poo
(1)
where T** and P,. are the ambient air temperature and pressure, respectively. We modelled the back of the plate as adiabatic, so that ~T = 0 at x = x o, "~x
for
y2 + z 2 > D 2 / 4
(6)
where G is the solar irradiance, o~is the solar absorptivity of the plate, ks and k are the thermal conductivity of the plate and air, respectively, and h, is the radiative heat transfer from the plate to the radiant surrounds, which are assumed to be at the ambient air temperature, T.. At the interface inside the hole, there must be a balance of heat fluxes, leading to the boundary condition:
Fig.1 (a): Sketch of the representative elements of the absorber plate and (b) definition of the computational domain
V=(Vs,O,O ) T=T**
Ox lt+ + hr (T - T**)
~y
X
atx=x**,
_kb_~
(2)
at O < x < t and y2 + z2 = D 2 / 4 ,
v (o,o,o),o~-k,
a,- I(~,/~)+
(~/~)_
where r 2 = y2 + z 2" 2.3 Dimensional Analysis The governing equations and boundary conditions were transformed into non-dimensional equations by introducing certain dimensionless variables (Aralanandam et al., 1995). The following dimensionless groups arose: The plate porosity r = ~292/4P 2 , a Reynolds number defined by Re D = VhD[V
where
Vh = V~/cr
(7)
a non-dimensional approach distance x** = x**/D and a non-dimensional plate thickness t* = t / D , the "plate admittance" Ad, and radiative Nusselt number Nu, defined by Ad = k ~ t / ~
and
Nur = h,D/k
(8)
respectively and the Prandtl number of air, which is fixed at about 0. 7. The heat exchange effectiveness e is defined
by (r0 - r.)/(r, - ~.) where ~0 i~ tho b u ~ outlet temperature at x = 0 and Tp is the average plate temperature. When it is expressed in terms of dimensionless quantities, there results e=e
9
eD, G , t , A d , N u r
)
(9)
ISES Solar World Congress 1999, Volume III
The heat transfer can also be expressed in terms of a Nusselt number, defined by Nu = ReD Pr ln(1 - 6) t7
(10)
From Equations (9) and (10) it follows that Nu=Nu(ReD, tr, t* ,Ad, Nur)
3.
COMPUTATIONAL MODEL
FLUID
(11)
MECHANICS
The governing equations were solved with the appropriate boundary conditions using TASCflow, a finite volume based CFD code. Using TASCflow, the computational domain shown in Figure 1(b) was divided into a finite set of control volumes. The TASCflow solver was used to solve the algebraic equations that result from integrating the governing equations over each control volume. The control volume formulation method is fully conservative, with the formulation guaranteeing conservation of mass, momentum, and energy over each control volume.
Fig. 2:
Plot of the grid lines used, at a constant x-plane
3.1 GridDesign The domain was broken down into a set of control volumes, with a node at the centre of each volume. The total number of nodes N in the resulting grid is limited by constraints on the available computer memory and also by CPU time. In breaking down the domain, a rectilinear grid with uniformly spaced nodes was tried, but there were several problems with this approach, the most significant being the large number of irregular control volumes located at the edge of the circular hole. Later a series of straight lines approximating circular arcs centred along the x-axis was used to create the grid in the region within D/2 of the centreline. A different series of straight lines was used to approximate the square cross-section of the domain at the edges of the domain. This choice of grid lines improved the grid in the vicinity of the hole, but it led to having several nodes co-located at the origin. So an additional grid was created for the region around the x-axis, and was then attached to the main grid. Thus, the final grid actually consisted of two grids: the main grid and the sub-
grid. A view of the final grid design is given in Figure 2. Preliminary investigations indicated that convergence was improved by increasing the number of nodes near the solid and inside the hole in order to fully capture the characteristics of the flow as it approaches the plate and enters the hole. Therefore, the nodes were distributed using an expansion ratio R, where R is the ratio (in a given direction) of the length of the last flux element to that of the first flux element. The expansion ratio used in the grid construction near the walls was 25, resulting in a higher density of grid points in regions near the walls, where they were most needed. 3.2 Grid Refinement and Other Preliminary Studies The approximations introduced in the discretization process become more precise as the gird is refined, i.e. as the number of nodal points N is increased. Grid refinement studies are used to study the sensitivity of the numerical solution size of the grid and then decide what value of N will be used for the bulk of the simulations. A grid refinement study is usually carried out by systematically increasing the number of nodes and comparing the results --in terms of global parameters, such as the effectiveness. For a three-dimensional problem, the number of flux elements is successively doubled in all of the three directions. Thus, if the first grid has NI nodes in total, the second will have 8N1 and the third 64N1 and so on. The solution as N approaches infinity can be extrapolated from the finite N grid solutions by using an extrapolation method such as Repeated Richardson (RR) Extrapolation (Zwillinger 1992). Comparing the global parameters from each finite grid with the extrapolated results can then be used to estimate the error induced by having a finite value of N. In the present study, the grid refinement study was completed using the following combination of parameters: ReD = 1375.5, tr = 0.0005, t* = 0.67, Ad = 4.93, and Nut = 0.26. This combination of parameters, having a high suction velocity, low conductivity and low porosity was the most difficult computationally, because it has high velocity gradients inside the hole and a large variation in temperature across the plate. Ideally the x***plane is located an infinite distance in front of the plate, but in practice a suitably large value must be used. Several runs were made in which x** was reduced in steps of about 10 from 168 to 42. Within round-off error, the same results were obtained at each setting. So, at x*** = 42, the inlet boundary face is far enough away from the plate to not affect the results. Thus, a value of x* > 42 was used for the grid refinement study and all subsequent simulation runs. An initial coarse grid, Grid A, was first constructed using a 6 x 12 x 12 grid with 6 flux elements in the xdirection, 12 flux elements in the y-direction, and 12 flux elements in the z-direction. (A flux element is a hexahedron defined by eight nodes, one at each vertex.) The number of flux elements in Grid A was doubled in the three directions to produce Grid B, a 12 x 12 x 24 grid. Similarly, Grid B was doubled to produce Grid D, a 24 x 48 x 48 grid. Computer memory limitations eliminated the possibility of running a fourth gird by
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doubling Grid D. Instead, a fourth grid, between Grid B and Grid D, was constructed by using a refinement factor of 1.5; thus Grid C was an 18 x 36 x 36 grid. The effectiveness was calculated for the four different grids and plotted against the inverse of N. The Repeated Richardson (RR) Extrapolation method was used to estimate the value of the effectiveness and the efficiency as the number of nodes approached efficiency, and the results are presented in Table 1.
Grid Size Grid A (6 x 12 x 12) Grid a (12 x 24 x 24) Grid C (18 x 36 x 36) Grid D (24 x 48 x 48) RR Extrapolation
"Global Parameters E T1 0.1745 81.98 % ' 0.1473 66.78 % 0.1365 66.60 % 0.1362 66.45 % 0.1362 66.45 %
Table 1- Grid refinement study results
captured by the variables in Equation (9). 3.4 Comparison to Similar Studies Kutscher (1992) completed numerical studies for a transpired plate absorber with a hexagonal pitch under nowind conditions. While not identical to the square-pitch configuration investigated in the current study, the configurations were similar enough to merit a detailed comparison of the results. Kutscher's work covered high porosity plates and include heat transfer from the back of the plate, but in the current numerical model the back surface of the plate has been modelled as an adiabatic surface. This significant difference in the models was easily accounted for in the comparison of Kutscher's numerical results, since he broke down the total heat transfer into three sequential heat exchange processes: that occurring along the front surface of the plate, inside the hole, and along the back surface of the plate. To compare the present results, Kutscher's results for the front and hole regions were combined. The results of the two codes were found to agree to within about 1%.
Reo = 1375.5, G = 0.005, t* = 0 . 6 7 , A d = 4.93,Nu~= 0.26
4. Based on the results of this study, and considering the amount of CPU time available, Grid C was chosen as a suitable grid with acceptably small error and convergence time. Figure 3 gives a plot of the velocity vectors obtained using Grid C.
SIMULATION RESULTS
4.1 R a n g e o f P a r a m e t e r s The important ranges for the non-dimensional parameters were chosen based on a number of factors, including manufacturing constraints and solver limitations. Thus, the following ranges were specified for the dimensionless parameters: 150 <_ReD <- 1350, 0.005 <_ G<_ 0.02, 0.67 <_t* <_2.0, 5 <_Ad < 1150, and 0.13 <_Nur < _ 0.52. The physical properties of air were kept constant for all simulations, evaluated at 300 K and atmospheric pressure. The ambient air temperature, T,,, is 300 K. A constant value of aG = 800 W/m 2 is used for all
simulations. Simulations were carried out over a wide range of combinations of specific values of the dimensional parameters inside their ranges: a total of 216 simulations were carried out in all.
Fig. 3" Fluid velocity vector plot obtained in the set of conditions applying in Grid C of the grid refinement study 3.3 Dimensionless Parameter Validation Study A study was undertaken with two different sets (set 1 and set 2) of dimensional variables chosen such that the dimensionless parameters were the same (namely Ren = 540, o'= 0.0111, t* = 2.0, A d = 1150, and ?Cur = 0.35) but the dimensional parameters were different, as follows: DI = 0.001588 m, PI = 0.0134 m, h = 0.003175 m, Vs~ = 0.06 m/s, ks1 = 15.12 W/mK, zpl = 0.45; and 1)2 = 0.001191 m, P2 = 0.01 m, t2 = 0.002382 m, Vs2 = 0.08 m/s, ks2 = 15.12 W/mK, zp2 = O.60. (In the proceeding,
subscripts 1 and 2 refer to the two sets, and zp is the emissivity of the plate.) The resulting values for e were el = 0.367 and e2 = 0.366, which are within 0.3% of each other, verifying the integrity of the computational code and demonstrating that the behaviour of e is indeed
4.2 Simulation Results To illustrate the effect of the five dimensionless parameters on the collector performance, plots were prepared in which each parameter was varied while the others were held constant. The strongest effect on e was observed for ReD; the next strongest was observed for or, then t*, and then Ad. No effect of Nut was observed. Figure 4 shows a typical plot of this set. Further plots are given by Arulanandam (1995). Figure 4 demonstrates the effect of the admittance. The admittance, Ad, represents the ability of the plate to conduct heat. Because of the high conductive coefficient near the hole associated with the high velocities there, the region of the plate surrounding the hole takes up a lower temperature that the rest of the plate. Heat from the outer regions of the plate is therefore conducted towards the hole. For high A d plates, more heat is conducted towards the hole, increasing e. In the high A d cases, corresponding for example to plates made of stainless steel (ks = 15.121 W/mK) or aluminium (ks = 131 W/mK), the value of e approaches the asymptotic value; the asymptote is approached as the plate approaches isothermal conditions with the plate temperature almost uniform. At low values o f Ad, the non-uniform
ISES Solar World Congress 1999, Volume III
33
Using SAS, the values of the coefficients were found, and the best-fit correlation followed. This fit had a coefficient of determination, Re, of 0.9915. (In the following, the correlation data and the TASCflow generated data will be distinguished by subscripts: C being used for correlation data and T for TASCflow data.) Following this, to check the limits, additional runs were completed for plate conductivity approaching zero and infinity. The additional points were added to the original simulation study data points, resulting in the following equation:
NUD =
Fig. 4: Plot of simulation results in the form of effectiveness vs. Reynolds number at various values of the admittance Ad. Values of tr, Nu,, and t* are as noted.
5.25Re~ 6 t7~ +0.15t*) 7.89 1-1-~ 13+Ad
(13)
Figure 5 is a plot of all of the data points used to generate the coefficients in Equation (13), with Nuo,r plotted against NuD,c calculated with Equation (13). The coefficient of determination for this fit is R e = 0.9887. Although this R 2 value is lower than that obtained without the additional points the result is a better fit for the data limits of Ad, i.e. as Ad approaches zero and as Ad
temperature distribution results in higher temperatures away from the hole, leading to higher radiative losses in this region and thus lowering ~. 4.3 Development of Correlation Equation One of the objectives in this study was to relate the collector performance parameter e to the non-dimensional parameters ReD, tr, t*, Ad, and Nu~. One way to express the dependence of e on these parameters is through the Nusselt number. That is, from the analysis presented in Section 2, the heat exchange effectiveness e can be expressed as a function of NUD (Equation (10)), and then NUD can be correlated as a function of these five dimensionless parameters, once Nuo(Reo, ~, t*, Ad, Nur) is known, e can be easily found using equations presented in Section 2. This is the strategy adopted here. The statistical sottware package SAS was used to perform a non-linear multi-variable regression analysis of the data points generated using TASCflow. The curve fit program minimises the sum of the squares of the differences between the correlation equation values and the data points. That Nut does not have a significant effect was confirmed by the SAS results. As a result, the correlation equation was developed for only four parameters: Re, tr,
t*, and Ad. A simple power law form was used to capture the effects of ReD and tr on NUD. The form used to capture the other variables took into account the behaviour of Nuo as t* and Ad approached the limits of 0 and oo. Several forms were tried, but the best results were obtained using the equation of the form:
NUD = flo Re#o'cT'~(l + fl3t*) 1 + f14 ~
(12)
Fig. 5: Comparison of predictions of Nuo given by Equation (13) (denoted Nuoc) with the corresponding NUD given by the CFD simulation (denoted NUDT) Two additional forms of the correlation equation were tested in an attempt to improve the data fit. First, a higher order term was added to the t* expression, then a higher order term was added to the Ad expression in the denominator of Equation (13). The improvement with the additional terms was negligible. Thus, Equation (13) was chosen as the most suitable fit for the data. 5.
CONCLUSIONS
The heat transfer and flow characteristics have been investigated for a representative element of an unglazed transpired-plate absorber with holes in a square pitch arrangement under no-wind conditions. The study was completed using TASCflow, a state of the art CFD code.
34
ISES Solar World Congress 1999, Volume Ill
A mathematical model was developed with the relevant boundary conditions and interfacial (solid-fluid) conditions specified. A dimensional analysis of these equations established that e depends on five nondimensional parameters: Reo, tr, t, and Ad using Equation (13). The computational grid was designed to handle a wide range of the dependent parameters for the simulation study. Validation of this grid was established by completing a multi-step grid refinement study, a dimensionless parameter dependence study, and a comparison with published data from Kutscher (1992). The simulation study was conducted by systematically varying each of the five dimensionless parameters for a total of 216 runs and analysing the results. Using SAS, a statistical software package, a correlation equation was derived for Nuo (R" = 0.9887). The Nusselt number is shown to be almost independent of Nu~ and can be estimated as a function of ReD, a, t*, and Ad using Equation (13). The results demonstrate the potential for low conductivity absorbers in low porosity, low flow situations. For the same plate geometry, changing the plate conductivity from ks = 0.196 W/mK to ks = 15.121 W/mK resulted in a 10-20% drop in the effectiveness, with the percent reduction increasing as ReD was increased. The effect on the thermal efficiency, r/, would be less pronounced, with approximately a 5% reduction in 17expected. The results of this study demonstrate that if transparent-plate absorbers were to be made from lower conductivity materials, acceptable efficiencies could be achieved. REFERENCES
Advanced Scientific Computing Ltd. (1994). TASCflowVersion 2.3.1: Theory Documentation, Advanced Scientific Computing Ltd, Waterloo, Canada. Arpaci, V.S. and Larsen, P.S. (1984). Convection Heat Transfer, Prentance-Hall Inc., New Jersey. Cao S., Hollands K.G.T. and Brundrett E. (1993). Heat Exchange Effectiveness of Unglazed Transpired-Plate Solar Collector in 2D Flow. Proceedings oflSES Solar World Congress 1993. Budapest, Hungary, 5, pp. 351-366. Hollands, K.G.T., (1998). Principles of the TranspiredPlateAir Heating Collector: the SOLARWALL. Renewable Energy Technologies in Cold Climates, 1998 Annual Meeting of the Solar Energy Society of Canada Inc. SESCI, Ottawa, pp. 139-144. Kutscher C.F. (1992). An Investigation of Heat Transfer for Air Flow through Low Porosity Perforated Plates, PhD Thesis, University of Colorado, Dept. Mech. Engineering Kutscher C.F. (1994). Heat Exchanger Effectiveness and Pressure Drop for Air Flow Through Perforated Plates, With and Without Crosswind, Journal of Heat Transfer, 116, 391-399.
Kutscher C.F., Christensen C., and Barker, G. (1993). Unglazed transpired solar collectors: heat loss theory. ASME Journal of Solar Eng., Vol 115, No. 3, pp. 182188. Van Decker, G.W.E., Hollands, K.G.T., and Brunger, A.P. (1996). Heat Exchange Effectiveness of Unglazed Transpired-Plate Solar Collector in 3D Flow, Proceedings of EuroSun "96, Freiburg, Germany, A. Goietzburger, J. Luther Editors, DGS - Sonnen energie Verlags GmbH, Munchen, Germany, pp.130-135. Van Decker, G.W.E., and Hollands, K.G.T. (1999) An Empirical Heat Transfer Equation for the Transpired Solar Collectors, Including No-Wind Conditions, Proceedings, Solar World Congress, 1999: Biennial Meeting of the International Solar Energy Society, Jerusalem, Israel Zwillinger, D. (1992). Handbook of Integration, John Bartlett Publishers, Boston.
ISES Solar World Congress 1999, Volume III
ANALYSIS OF THERMAL
PERFORMANCE
35
ON AN AIR-TYPE SOLAR COLLECTOR
WITH 2-GLASS USING CARBON FIBER SHEET AS COLLECTING
MATERIAL
Xi-meng JIANG Student at Graduate Course, Kitami Institute of Technology, 165 Koen-cho Kitami, Hokkaido, 090-8507 JAPAN, Tel. +81-157-269231 Fax. +81-157-239375 E-mail: dse97802/
[email protected]
H i r o m u BABA, Kimio KANAYAMA, Noboru ENDOH Dept. of Mech. Engng., Kitami Institute of Technology, 165 Koen-cho Kitami, Hokkaido, 090-8507 JAPAN, Tel. +81-157269209 Fax. +81-157-239375 ABSTRACT- In this paper, the analysis of thermal performance on an air-type 2-glass solar collector using carbon fiber sheet as collecting material (air-type CF-sheet solar collector) is described. A model collector (2m•215 was made with 5 steps arranging CF-sheet. Based on this collector, the heat balance equations were designated and the temperatures, thermal efficiency and the overall heat loss coefficients were calculated. The operating parameters consist of the solar radiation intensity, flow rate of air, inlet temperature and ambient temperature. It was confirmed that the thermal efficiency became higher along with the higher ambient temperature, higher radiation intensity and higher flow rate of air and decreased with high inlet temperature. The 2-glass solar collector was superior to the 1-glass one and the overall heat loss coefficient became higher along with increase in ambient temperature. 1. INTRODUCTION Because the heat capacity of the air is small and heat transfer coefficient from absorber to air is low, it can be thought the performance of the air-type solar collector is inferior to the liquid-based collectors (L~ et al., 1975). However, there are advantages of simple construction, no freezing, and hot air can be used for space heating directly. Therefore, the performance improvement of airtype solar collector has been desired. Air-type collectors have been constructed in several configurations: for example, flow over or under the absorber, and flow on both sides of absorber (Parker et al. 1993). There are many different designs such as, Vgroove solar collector (Gama, et al., 1986) and honeycomb structure collector (HoUands, 1973), except for flat-plate absorber. The CF-sheet solar collector is special one of the air-type collector using two pieces of CF-sheets as collecting material. For a single-glazing (1-glass) air-type CF-sheet solar collector, based on the structure, given heat balance equations were calculated by assuming a lot of conditions. And thus, it's thermal performance was made clear by a numerical analysis in the previous report by the authors (Jiang, et al., 1997; Jiang, et al., 1998a). At the same time, using a large-scale solar simulator the measurement of the thermal performance was done under the stabilized conditions of the radiation intensity indoors (Jiang, et al., 1998; Jiang, et al., 1998b). As a result, there is a little difference
between the experimental values and calculated values for the collecting efficiency, but the ranges of variable data were agreed. Therefore it will be thought that the analytical method for the performance of the 1-glass CFsheet collector was verified. In this paper, the thermal performance on a doubleglazing (2-glass) air-type CF-sheet collector is analyzed applying the same method as above of 1-glass CF-sheet collector. The expressions of energy balance on each component of two glass plates and two CF-sheets could be yielded and were solved by using a repeat calculation (step-by-step) method to analyze the collector efficiency under the steady state conditions. NOMENCLATURE A effective area per one step of CF-sheet (= 0.327 ) Cc conductance of air flow, W- (7 9? ) 1 Cd coeff, of heat transmission to bottom, W. (7 9? ) 1 D height of the channel for air flow (= 0.13m) E.~m) solarradiation intensity of air ma.~ m at wavelength ?, W- (? 9~m) ~ F.~(T) radiation intensity of black body of temperature T at wavelength ?, W. (? 9~m) 1 G1, G2 glass 1, glass 2 Gr Grashof number J solar radiation intensity on the collector surface V~. m- 2 ka conductivity of air W. ( m- ? ) " 1 L length of channel for air flow per one step of CFsheet( = 0.328m)
ISES Solar World Congress 1999, Volume III
36
m air mass Q thermal energy per effective area of the collector V~r. m- 2
$ 1 , $ 2 CF-sheet 1, CF-sheet 2 (area of 1 step of CFsheet is 0.35mx0.9m) T * , T temperature IL ? t time s UL heat loss coefficient of the collector W. (m. ? ) ~ V flow rate of the air m 3. min" ~ Greek
those of other fibers. The top cover is composed of two glass plates of 3 mm thick each, between which there is a space of 15 mm thick to prevent heat loss from the collector by convection and radiation.
letters
a absorptance of CF-sheet t transmittance of glass plate t*s fraction of ray-trapping of CF-sheet (= 0.9) e emittance ? collector efficiency ? tilt angle of collector ? ? wavelength ~m ? T difference between average and ambient temperatures: ( Tin+ Tex)/2- Ta = TA- Ta ? Subscripts
a ambient A averaged G glass S sheet ? wavelength
Fig.1 Structure and dimension of the 2-glass air-type CF-sheet solar collector
2. N U M E R I C A L ANALYSIS 2.1 Construction of CF-sheet collector The structure and dimension of a 2-glass air-type CFsheet model collector are shown in Fig.1. The inner space of the collector casing made by the stainless is divided into three channels by two partition plates, and the inside of the case is well insulated. One piece CFsheet is bound on a layer at the bottom insulator, and another one are obliquely inserted in each flow channel of the space in five steps to be crossed along air flowing. The physical properties of the pitch-series carbon fiber used for the absorbing material of this solar collector are shown in Table? compared with
2.2 Mathematical model and basic equations In order to analyze the thermal performance of the 2glass air-type CF-sheet collector, one step of collector structure is shown in Fig.2, by extracting from the model collector circled with K in Fig.1. From this figure on the mathematical model, a mechanism of heat balance of the CF-sheet collector under the steady state conditions can be seen. From Fig.2, the collector is composed of four elemental components of two CF-sheets and two glasses. The incident rays are absorbed or transmitted through these components, and heat transfer due to the radiation and convection is occurred among them and surroundings. And the equations of energy balance (1)~ (4) can be held on each component as follows:
i
- ~ ~ . ~ i I ~ ! 9
!:~i ......~...~.._ , i l
.....:
i:i~~:fi~m
i
iP A N
ii
IJnr
~. . . . . . . . . . . . . . . . . . .
::':::: ~ ~..~=~~:~ i
~ :
~'~ ' : ~
......................................... L...............:;~::.~:..........~ . ..........~:::!~:-.;~._....i ............~ ~ : ; ,
~:=._
................................... ...i_....,.~~.~.~
~'~:
m
~~
~•
::~._~;~i~A.
!
~
ii!i~]
:,::~r
~~i~.
: ~ ' -
................._ ~ I ] : ~ I L
i i!!i.~s
i
..2~:I~
~
I
i
:~:
:~s
~i
..........L.L~-k!::~,..:.!L~ ...... : : ~ i :
! ii ~ ] : ~ :
.2:7(X)
:~:
i ::~::~~:
i
--
. . . . . .
~~...:~
i~~~z~:~
i.S~
~:
"
~
::~
~~.
.... , : ~ : ~ L : , . L
............................
...................!
...:~~%~.L ......
~
~ ~
6~i
~..............................
i ~~::
:~i
~,~~::: .....i........~,::~::~ ...... .........!~:~:~'~:~)~i!~!'!~:i::~::~i:~i:::~:~:;'~ ....... ....!....... ..i)i:!~:~:iV~' ......~.....:::~i~:.-illi~! .. ~.....:i .............=......................... =========================================================================================================================================================================================================================== i~...~~
.........i.....::.::.I:o:; ... ~.:~]i..i...i.....ii......:1,~.:;i]]...i...]iii.[.!1.~s ~]]II[S]]iii~::~.:...i]]iil]][.i..ilEiii::~..]] ....... i[.......]i=...[] ...... ....
i:~~.!~:~.~~:~....i. ......:...:.....~ m ~ io~i~l:~:: ................ "
~
! ..........J~;~: ........ ............i ........]..~...~... ! c~ood i .(o~,
........:i ..........=:~..s.. ........ ...............!...........~.~i....~.~.~. ......... ..... L = : .....:~i ...... ..........:..... .............. " A:.l~e.~or: i::, F ~ r ~ r ~ I I ~ i i i.)~t~br..e,cid
ISES Solar World Congress 1999, Volume Ill
For sheet1; (mscs / A" ) dTs~ T -- QOSI--(Qs1G2- QG2S,)- (Qs1s2- QS2S1)-- Q$10
-- (Qs1G1-- QG1S1)-- QFAG2"+"QFAS2-- QFAE
(1)
For glass 2; (~/.,
d%2 = QoG2-(QG2G1-QG1G2)--(QG2s1-QS1G2)--QG20
For glass 1; ( n r ~ / " "dTm A ) T
-
= I ~ 3 1 - (QGIG2 -- QG2G1) -- (QG1S1 -- QS1G1) -- QGIO
,G0-
(3)
For sheet 2; (~/
" dTs2 = Qos2- (Qszsl-Qsm)- (Qsm2- QGzs2)-Qs20 - (Qsm,- Qms2)- Q~As2-QcBo
(4)
where, ms and cs are mass and specific heat of CFsheet, and mG and CG are mass and specific heat of glass plate repectively. The differential terms on the left hand in Eqs. (1) ~ (4) are the variation of thermal capacity of energy stored in each component due to the temperature change. When each component attained in equilibrium, they become to be 0. For Eqs. (1)~ (4), we will put the useful heat gain as QFAE= D1, heat loss by covection from glass 2 to surrondings as QFAG2= - D2 and heat loss by covection from sheet2 as QFAS2=D4. By assuming D3=0 and substituting them into the above, the next equations (5)~ (8) can be yielded.
Dl=Qosl- (Qs1G2- QG2Sl)- (QsIs2- Qs2s1) - Qsm- (QslG1- QGm0+ D2+ D4
(5)
D2-QoG2- (QG2G1- QG1G2)- (QG2s1- QS1G2)- QG20 - (QG2s2- QS2G2) - (QCG2G1- QCG1G2)
(6)
D3-QoG1- (QGIG2- QG2G1)- (QG1s1- QS1G1)- QG10 - (QG2s2- QS2G2) - (QCG2G1- QCG1G2)- QCGlO
(7)
D4-Qos2- (Qsgm- Qms2)- (Qs2G2- QG2S2) - (Qs2G1- QG1S2)- Qs20- Qcso
(8)
where Qom, QOG2, QOG1and Qos2 are thermal energies of the incident radiation which are absorbed by sheet 1, glass 2, glass 1 and sheet 2 respectively. QS1G2, QSlG1, Qslo and Qsm2 are thermal energies by radiation from sheet 1 of temperature Tsl to glass 2, glass 1, surroundings and sheet 2 respectively. QG20, QG2G1, QGS1 and QGS2 are thermal energies by radiation from glass 2 of temperature TG2 to
37
surroundings, glass 1, sheet 1 and sheet 2 respectively. QGlO, QG1G2,Qmsl and Qms2 are heat thermal energies by radiation from glass 1 of temperature TG1 to surroundings, glass 2, sheet I and sheet 2 respectively. Qs2m, QS2G2, QS2G1 and Qs20 are thermal energy by radiation from sheet 2 of temperature Ws2 to sheet 1, glass 2, glass 1 and surrounding respectively. QcGlo, QCG2G1, QCG1G2, QFAS2 (-- D3) and QFAG2 (--- D2) are thermal energy by convection from glass 1 to surroundings, from glass 2 to glass 1, from glass 1 to glass 2, from air flow to sheet 2 and glass 2, respectively. QcBo is heat loss downward from sheet 2 through the bottom insulation to outside by heat transmission due to conduction. QFAE (= D1) is useful energy gain per effective area of the collector. On each component of thermal energy in the above equation: (1) Absorbed energy of solar radiation For sheet 1; 2 *
Qosl = Jxl a:Gx~: sotsEx(am)d~
(9)
For glass 2; qoG2 = fx~2XGxaGEx(am~b~
(10)
For glass 1; QoGI = ~ omEx(am)d~
(11)
For sheet 2; Qos2 = J'x~,l:Gx~(1- ,l:*s)1:*sasEx(a~)dk
(12)
where ET(m) is Moon's spectral radiation intensity (Moon and Franklin, 1940) of incident rays with air mass m. The E?(m) with m=l.5 is usually used in this calculation.
Fig.2 Heat balance in a model of the 2-glass collector( Extracted from K in Fig. 1)
ISES Solar World Congress 1999, Volume III
38
(2) Thermal energy by radiation from sheet 1 of temperature Tsl From sheet 1 to glass2 ; QSlG2 - I~.)'~0~s6'l;*sEsWx(T*sl)d~
(13)
Qslo= iX1
/
.
x
r
* x
S~s[FxtT sI)-Fx(T a)]d~,
(15)
From sheet 1 to sheet 2; Qs,s2 = Ii~2z*s~~ s F x ( T * s l ) ~
(16)
(3) Thermal energy by radiation from glass 2 of temperature T62 From glass 2 to ambient; QG20
= IX~'; T ~ 3 ~ . x } [ F x ( T * G 2 ~ F x ( T * a ~ d ~
G2)d~
QG2s2= I~}x*smseGF~(T'o2)~
F4W*)=C1/(?5(eC~Z'- 1))
(29)
(18)
(5) Thermal energy by convection From glass 1 to the ambient;
(19)
From glass 2 to sheet 2;
Qcmo=hc(Tm
- Ta)
(30)
Coefficient of heat convection between the top cover glass and the ambient is given by Eq. (31).
, =
(4) Thermal energy by radiation from glass 1 of temperature TG1 From glassl to ambient; ~,2
In these equations abovet~is ~ transmittance of the glass cover (Kanayama, et al., 1981). The absorptance of the CF sheet as=0.85 (Jiang, et al., 1998b), emittance of CF-sheet es=0.85, a fraction of ray-trapping of CFsheet t's=0.9, absorptance of glass aG=0.05 and emittance of glass ~=0.88 are assumed. The thermal energy by radiation from a black body of temperature T* is expressed by Planck's law.
where C1= 3.7417x108 W-~n 4/? and C2= 1.4388x104 ~m-K
From glass 2 to sheet 1;
Qo,
(27)
(17)
From glass 2 to glass 1; QG2m = I:~ s
Qs2ol - Ix~2T,6x(Ic,~.0- ,i;*s ~* sEsFx(T*s2)d~
(14)
From sheet 1 to ambient; "
From sheet 2 to glass 1;
Qs:o= I~:'cox~'C's(1-'Cs~[Fx(T's,)--Fx(T',~d~(28)
Qs,ol = Ix~:'r~x0~'l;*sesF~.(T*Sl)~ 2 * 'T,GX "C
(26)
From sheet 2 to ambient;
From sheet 1 to glass 1;
p~,2
Qs2G2 = I~20Co(1- X's)~*sesF~(T*s2~I~
.
.
(21)
Qo,o =Ix I ~[Fx(T Ol)-Fx(T a)]d~
From glass i to glass 2;
QGlO2= I~; ~GF~(T'GI)~
(22)
From glass 1 to sheet 1: QGIS1-- I~,~2~,~'1~* ss
s1)d/~
(23)
From glass 1 to sheet 2;
Qols2= I;~21:ox(l1:*s)~*saseoFx(T*ol)dZ
(24)
hc=4.8+3.4v (v? 5m/s) hc=6.12xv ~
(v> 5m/s)
(3~)
where, v is wind velocity and the calculation was done under a breeze of v=0.5m]s. From glass 2 to glass 1; QCG2GI=Cc(TG2- TG1) (TG2>Tin) =0 (TG2? TG1)
(32)
From glass 1 to glass 2; QcG1G2=Cc(TG1- TG2) (Tin>TG~) =0 (TGI?TG9
(33)
From air medium to glass 2; (5) Thermal energy by radiation from sheet 2 of temperature Ts2 From sheet 2 to sheet 1; fZ2
* 2
- TG2)
(34)
From air medium to sheet 2; /
*
\
-
Qs2s, = ix1 "l;s ~sesFztT s2)dX From sheet 2 to glass 2;
QFAG2=CFIxCc(TA
(25)
Q F A s 2 = C F 2 x C c ( T A - Ts~)
(35)
ISES Solar World Congress 1999, Volume !11
where TA is average temperature of air flow. CF1 and CF2 are amendment coefficients which are very difficult to find directly out. Therefor QZAG2 and Qz~2 are calculated by giving empirical values in Eqs.(34) and ( 3 5 ) . Cc is thermal conductance by convection and conduction (Jiang, et al., 1998a; Kanayama and Baba, 1985) From sheet 2 to the ambient downwards; QCBO=Cd(Ts2- TD
(36)
where a coefficient of heat transmission between the sheet 2 and the surrounding, Ca, is given by Eq. (37). Cd= I. 01(1.5+ 1.O/hc)
(3 7)
Ultimately, the total useful energy collected by a collector equals QFAEx5A~ so that the collector efficiency ? is given by Eq. (38). ? =QF~LFJJ
(38)
Fig.3 Flow chart of the calculation V--1 I
i ~
i
TO2=Ts2=T~=Tin Tsl=Tin+5
ICal.of ol .,~ I
1 I-
JCal.of DI''D41 YES I Tsz=Ts~-}fr
I~-~
~
39
It is seen from Fig.2 that in the calculation model, a semi-infinite width over the lateral direction is assumed for this calculation, but heat loss from the side wall of the collector must be considere& Therefor Cd was increased by 1.57 times for this calculation, according to fraction of side wall to all the surfaces of the collector. 2.3 Procedure of calculation By the above procedure, according to Eqs. (4) ~ (8), a flow chat of calculation for the collector efficiency is shown in Fig.3. First of all, initial values of the ambient and inlet temperatures, radiation intensity and flow rate should be given. We make the temperature of the glass 1 to be equal to the ambient temperature, make the temperatures of glass 2, sheet 2 and the average temperature of air flow to be equal to the inlet temperature, and we make also the temperature of sheet 1 to be equal to the inlet temperature plus 5 ?. And then D~, 1)2, D3 and D4 are calculated repeatedly, until Eqs. (2), (3) and (4) reach to be in equilibrium, when the temperatures of glass 1, glass 2, sheet 1 and sheet 2 are increased or degreased differently. If the Eqs. (2), (3) and (4) were in equilibrium, the temperature of each component of the collector is determined. And collector efficiency is obtained as a ratio of DI(=QF~m) to the incident radiation intensity. The results of temperatures and thermal gains for various parts of collector are given in Fig. 4, when the intensity is 800W/m 2, flow rate of the air is 3m3/min and inlet temperature is 60?. The heat loss from surface of the collector by radiation and convection is:
JTG2=Tc2~ }fCT
J
!~176 --1 9 D2=QI,aG2 ~
I
NO
QoGlo+QGlo+QG~.o+Qslo+Qsgo =256.8W/m 2
(39)
NO
The heat loss from bottom of the collector by conduction QcBo=67.1W/m9, then heat loss from coUector QL:
Cal. of D1
Tsl:Tsl -} fr
1000
al .of D1 .'D4
>..,
,--
~
(t3 c-" I,I
.--2 D4=Q,,as2
--3 9 D2 =QFAG2
'
N3~--~ JTsI=TA+fr
.,k.~o~.~,o,, .... ~.,.~o ..... ~ - - , - - . ~ ~ "";bD2 (TG2 ), .b .. -bQFaG2-b ~
...........~'EV+Tin
,, ~ b D 4 (Ts2),-1-b -f -bQi,As2 .b ~ bD4 9 (Ts2)n"b'. ~ -b ~ ~ Tsl -,TA
~
L__ n+ = ,x /2 NO
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E) of) c(].3
--~ !
I DI:Q,~,
800
200
0
l 800 d
_
I
I
I
I
i
V =3m3/min .--a ~/d =0.0691 a .=0.442J J
9
504.9 _ Qosl
350.: OG~s ~91.5 - 2 448.211 sL~ 56.1 Qos2
414.5 Qc2GI - 297.7
Qsl
QoG2 21.8
QFA02 ., qclsl 0.2
i
353.4 QFAr
Q.s2GI )cGto
0.4
9 9II ~176176
i
T i, = 60.0Ci -Tex =70.4Ci ~, TGI =23.9(i ~, T G2 - 47.7(J ~, Ts~ = 67.9(J ~, Ts2 = 63.40 ~, TA = 65.35 A
.
II
CC2G~ 0 . Qcls
Q~ ~
(i J~F^S2 Qs2G2 QG2oQcGIG2 I..5 47.1 0.7 0
(,BO
111 | as2o 0
I
I WRITE T, Q, " f * l
Fig.4 Partial thermal energies and temperatures in various parts of CF-sheet collector
40
ISES Solar World Congress 1999, Volume III
R E S E A R C H AND D E V E L O P M E N T OF SOLAR C O L L E C T O R S FABRICATED F R O M POLYMERIC MATERIAL AVRAHAM I. KUDISH*, EFIM G. EVSEEV*, MATTHIAS ROMMEL**, MICHAEL Kt3HL**, GERD WALTER*** and TIMO LEUKEFELD*** *Solar Energy Laboratory, Institutes for Applied Research, Ben-Gurion University of the Negev, Beer Sheva 84105, ISRAEL.; Tel: +972 7 6461488, Fax: +972 7 6472916, E-mail:
[email protected] ** Fratmhofer Institute for Solar Energy Systems ISE, Oltmannsstrage 5, D-79100 Freiburg, GERMANY; Tel: +49 761 4588141, Fax: +49 761 4588100, E-mail:
[email protected] ***Institut flit W~metechnik und Thermodynamic, Technische Universit~it Bergakademie Freiberg, Gustav-Zeuner Str. 7, D-09596 Freiberg, GERMANY; Tel: +49 3731 393494, Fax: +49 3731 393655, E-mail:
[email protected] Abstract - The successful development and utilization of solar collectors fabricated from polymeric materials has many fabrication by techniques, such as extrusion and molding, which can, a priori, result in significant reduction in production costs; Co.) eliminate corrosion problems, especially with regard to their application in sea water desalination systems; (c.) reduced weight per unit area of solar collector which results in significant reductions in both shipping and labor/installation costs. These advantages are contingent to overcoming the distinct disadvantage solar collectors fabricated from polymeric materials have vis-a-vis those fabricated from metals, viz., their inherently low thermal conductivity coefficients. This disadvantage can be compensated for by a proper design of the solar collector, which is the crux of this investigation. We have attempted to overcome this disadvantage by two very different approaches in the design of solar collector prototypes fabricated from polymeric materials. 1. Coaxial tubular solar collector: design concept that entails the use of an inner black tube as a solar absorber in intimate contact with an outer transparent tube as an insulator. In theory, the inner black tube should be of minimum wall thickness to compensate for its low conductivity and the outer transparent tube should have a wall thickness that optimizes the trade-off between the material insulation and its transmittance of the solar radiation. 2. Selectively coated polymeric absorbers fabricated from double-walled polymer sheets which function as the solar collector absorber plate. Their upper surface serves as the substrate for the selective coating and the heat exchange flows through the channels between the two walls. The selective surface was applied via a sputtering technique, utilizing the equipment available at ISE. The results and analysis of the performance testing on these two solar collector prototypes will be reported upon in this paper. 1. INTRODUCTION There is a relatively long history detailing the investigation of the utilization of polymeric materials in the design of solar to thermal energy conversion systems. One of the earliest reports on the use of polymer materials in solar collector systems was by Tabor and Zeimer (1962). They used an inflated polymer film as a cylindrical concentrator. Gerich (1977), also reported on the utilization of an inflated cylindrical concentrator, fabricated from a polymeric film, for the production of industrial process heat. Dickinson, et al. (1976) and Kudish (1980) both reported on the utilization of shallow solar ponds (SSP) fabricated from polymeric materials for the production of low grade thermal energy for both domestic and niche industrial process heat systems. The shallow solar pond is essentially a large water bag or pillow placed within an enclosure with a transparent upper glazing. Harris, et al. (1965) tested a solar water heater similar in design to that described by Dickinson, et al., the major difference being that a black butyl-rubber tube is substituted for the water pillow. More recently, Tsilingiris (1997) has reported upon the performance of such a system with the major design change being the use of glass as an upper glazing to resolve the problems associated with the UV degradation of most polymer glazings.
The idea of using such a simple device, viz., a water pillow, for solar to thermal energy conversion was not new. The Japanese had been using numerous variations of this idea to heat domestic hot water since the 1930's. In fact, Meinel and Meinel (1976) reported that 39 patents were issued for solar water heaters in Japan during the 1930's and another 20 were issued in 1940's, the majority of these being for the water pillow type. A study on such a commercially produced Japanese pillow-type water heater had been reported by Khanna (1973). Gopffarth, et al. (1968) tested a plastic solar water heater in which the water pillow was formed by heat sealing one black layer and one clear layer of polypropylene. They also investigated the use of tedlar as an upper glazing for the enclosure. The utilization of polymeric materials as upper glazings in solar collectors has been reported in the literature numerous times, e.g., Whillier (1963) and Grimmer and Moore (1975). Kudish and Wolf (1978) did extensive testing on a compact SSP designed for both recreation and military use. Their prototype design consisted of a water pillow placed within an insulated container to maintain the water temperature for overnight storage. The inner surface of insulated container cover was fitted with a
ISES Solar World Congress 1999, Volume III
mirrored surface which functioned as a reflector when open and operating during sunshine hours. Another approach to the incorporation of polymer materials in the construction of solar collectors is the fabrication of selectively coated absorber plates using polymeric substrates. Such corrosion-free absorber plates would be ideal for preheating the feedstock for the desalination of sea water by reverse osmosis. Rommel, et al. (1997) have done extensive work on the design, coating and testing of such absorber plates. In the following, we will describe the solar collectors fabricated from polymeric materials under investigation, their performance testing and a simulation model for the tubular collector design prototypes. 2. SOLAR COLLECTOR MODULES We have attempted to overcome the major disadvantage that solar collector fabricated from polymeric materials have vis-a-vis those fabricated from metals, viz., their inherently low thermal conductivity coefficients, by two very different approaches to their design.
41
decided that prior to attempting the fabrication of such a solar collector design, i.e., by an extrusion technique, it was best to determine if such a design was able to convert solar to thermal energy at high efficiencies. Therefore, it was decided to first fabricate a prototype coaxial tubular solar collector from off-the shelf stock. This prototype was fabricated from black natural rubber and transparent PVC tubing by inserting the black tube within the transparent tube. The prototype solar collector was of a riser and manifold design, i.e., the coaxial tubes functioned as risers and were connected to upper and lower manifolds fabricated from the same black natural rubber tubes. The riser tubes were all a nominal 1 m in length and the solar collector consisted of eight risers. Prototype 1, consisting only of black natural rubber tube risers, served as a reference standard for comparison of the performance of the coaxial solar collector, prototype 2. Theprototype 2 risers were constructed by inserting the black natural rubber tube into the transparent PVC tube. The prototype solar collectors were each placed on a support constructed from wood, approximately 1.2 x 1.2 m, which was painted black and were attached to the surface by plastic ties, such as those used in electronic equipment. The spacing between risers was 15 cm. The solar collectors were positioned at a tilt angle of 30 ~ towards the south. The heat exchange fluid, water, was pumped from a storage tank through the collectors by a peristaltic pump and Type T, copper-constantan, thermocouples were inserted at the entry and exit points to the solar collector manifolds. The dimensions of the tubing are listed in Table 1.
Coaxial tubular solar collector This design concept entails the use of an inner black tube as a solar absorber in intimate contact with an outer transparent tube as an insulator. In theory, the inner black tube should be of minimum wall thickness to compensate for its low conductivity and the outer transparent tube should have a wall thickness that optimizes the trade-off between the material insulation and its transmittance of the solar radiation. In essence, the outer transparent tube functions as a glazing. It was Table 1. Dimensions of tubing used in fabricating coaxial tubular solar collectors. Prototype No. 1. reference standard 2. coaxial tubes
Black tube ID(mm) OD(mm) 9 15 9 15
The conductivity of the polymeric materials being utilized in the fabrication of the coaxial solar collector prototypes have been measured at TUB and are as follows: 0.148 and 0.090 Wm-IK1 for the black natural rubber and transparent PVC, respectively. Selectively coated polymeric absorbers Double-walled polymer sheets were chosen to function as the solar collector absorber plates. Their upper surface serves as the substrate for the selective coating and the heat exchange fluid, water, flows through the channels between the two walls. The selective surface was applied via a sputtering technique, utilizing the equipment available at ISE. The procedure involves first applying a metal,
PVC tube
ID(mm) OD(mm) 16
20
molybdenum, layer on the polymer substrate that functions as an IR mirror and then a cermet layer, consisting of metallic chromium and its oxide and nitrate salts, which functions as the absorbing layer. The polymeric materials presently under study are polypropylene and polycarbonate. The optical properties of the selectively coated absorber plates have been measured at the ISE and are as follows: gold-green hue: {x= 90-92%, e = 9% blue hue: o~= 93-94%, e = 9%. where o~ and e are the absorbance and emittance, respectively. The different hues are a function of the coating composition. The dimensions of the
ISES Solar World Congress 1999, Volume III
42
selectively coated polymeric absorber plates and their hue are listed below in Table 2. The selectively coated polymeric absorber plates were outfitted with perspex manifolds machined on a lathe. The absorber plate being inserted into a slot machined into the perspex, whose width was of essentially the same dimensions as the corresponding plate thickness; producing a very fight fit. An epoxy cement was applied between the absorber plate and the perspex manifold. In addition, a ribbon of Si RTV was
applied along both sides of the slot as a further measure to assure that the joint was leakproof. The absorber plates were then inserted into a wooden casing and covered with a 4 mm glass glazing. The heat exchange fluid, water, was pumped from a storage tank through the collectors by a peristaltic pump a n d Type T, copper-constantan, thermocouples were inserted at the entry and exit points to the solar collector manifolds. The solar collectors were positioned at a tilt angle of 40 ~ towards the south during performance testing.
Table 2. Description of selectively coated polymeric absorber plates. Polymer substrate
Thickness of double-walled sheet (mm) 6.5 10 10 10 10 16
Polypropylene- 1 Polypropylene-2 Polycarbonate- 1 Polycarbonate-2 Polycarbonate-3 Polycarbonate
3. MEASUREMENTS Experimental setup The performance testing of the solar collectors was done on the roof of the building housing the Solar Energy Laboratory (BGU). The test loops consisted of a solar collector, peristaltic pump (both single and multi-head) and storage tank. The solar radiation intensities and ambient temperature were monitored by the laboratory's meteorological station located on the same roof. The meteorological station measures the global radiation, both on a horizontal surface and on a surface tilted at 40 ~ towards the south, and the normal incidence radiation. The global radiation being measured by Eppley PSP pyranometers and the normal incidence radiation by an Eppley NIP. The inlet and outlet temperatures of the heat exchange fluid, water, for each solar collector were measured by means of Type T thermocouples, which were scanned at 30 minute intervals. The peristaltic pmr~s, viz., the individual pump heads, were calibrated in the laboratory and the flow rates were checked periodically as part of the performance testing procedure. The average flow rates of the heat exchange fluid through the solar collectors were for the
Wall thickness
Hue
(mm) 0.75 0.75 0.75 0.75 0.75 1.5
(ram) gold-green blue gold-green gold-green blue gold-green
0.009 0.015 0.016 0.020
m m m m
500x1000 500xl000 400x600 400x600 400x600 500xl000
i. tubular solar collectors: reference standard- 5.7, 10.9 kghl; coaxialtubes- 7.2, 15.7 kgh "1 ii. selectively coated polymeric absorbers: 13.8 kgh1
As mentioned previously, the tubular solar collectors were positioned at a tilt angle of 30 ~ towards the south, whereas the selectively coated solar collectors were tired at an angle of 40 ~ towards the south. 4. SIMULATION MODEL The tubular solar collector consists of N risers which are mounted in a parallel array between two manifolds. The tubular solar collectors are described in Table 1. The storage tank was constructed from polypropylene and was tminsulated. The values for the relevant design parameters used in this analysis are listed below in Table 3, whereas the Nomenclature is listed at the end of the manuscript. The simulation model is simplified by assuming that the storage tank can be characterized by a single average temperature with regard to the thermal energy losses through its walls to the ambient and the inlet water temperature to the collector is the same as the outlet water temperature from the storage tank.
Table 3. Values of design parameters utilized in this analysis. Da,i = D~o = Ol,i = Dl,o = N=8
Dimensions
ma = 0.13 kg/riser m l = 0.12 kg/riser rest= 50 kg Mf = 10.9, 15.7 kgh "1 L = 0.98 m/riser
Aref = 0.50 m "2
A = 0.60m -2 A~t = 0.96 m 2 8~t- 0.02 m
ISES Solar World Congress 1999, Volume III
Thermal energy analysis The solar radiation incident on the tubular solar collector was calculated from the corresponding global and normal incidence radiation values measured by the BGU Solar Energy Laboratory's Meteorological Station. The measured values were converted to those incident on the tubular collectors by means of the following equations: Gc = GbeamCOS0+ Gdin(1+COSl3)/2, (1) where Gear = Gg~ob - GbeamSina. The incident angle 0 for south facing tubular collector along a north-south axis and inclined at a tilt angle 13 from horizontal is given by (Duffle and Beckman (1980)), cos0 = cosiS{sin2t.o + [cos(O- ~)costo + taniSsin(~ ~)]2} 1/2. (2) The solar altitude angle a is given by sina = cost~cos~costo + sin~sinS. (3) The overall thermal energy balance on the tubular solar collector modules is as follows: mlcldT1/dt = (xo01Gc(t) Al,efr + Al,efrUamb(Tamb-T1) + AaUI(T~-T1), (4) macadTa/dt = (Xa)aGc(t)Aa,efr+ AaUI(T1-Ta) + Qdt), (5) rn~c,~dTst/dt = A~tU~t(W~b-Tst) + Qst(t). (6) The notation, Ai,efr, refers to the effective area of the tubes, i.e., that upon which the solar radiation is incident. The overall heat transfer from the outermost surface to the ambient is given by, Rabl (1985) as Uamb =0-5E2G(I+cos~)(T14- Tsky4)/(T1 - Tamb) + heonv,
(7) where Tsky = 0.0552(Tamb)1"5 (8) cf., Swinbank (1963), and heonv = NUambK["a~l,o . (9) The Nusselt number, as a function of the Reynolds number, is estimated from the following correlation (Eckert (1972)), NUmb = 0.43 + 0.48(Re~mb)~ (10) in the case of laminar flow. The overall heat transfer coefficient U1 for a two long concentric tubes is then given, Oszisik (1977), as: UI = [l/Ca + (Da/D1)(1/e1 -1)]6(Ta 2 + T12)(Ta + T1) + hi. (11) The overall Nusselt number NUl for two horizontal concentric cylinders, based upon the inner cylinder is given (cf., Kuehn and Goldstein (1978))as: NuI _. [ (NUl,~o~d )15 + (NUl,~onv)1511/15 , (12) NUl,cond = 2/ln(Dl,i/Da, o), (13) NUl,~o,v= [ 1/Nui + 1/Nuo]l, (14) Nui = 2/111{ 1+2/[(0.5Rai1/4)15+(0.12Rail/3) 15 ]1/15}, (15) N u o - -2/111{ 1-2/[(Raol/4)lS+(0.12Raol/3) 15 ]1/15}, (16) R a i - 2g(Ta- Tb)Da,oPr/(Ta + Tb)rE, (17) Rao = 2g(Tb- TI)DI,iPr/(Tb + TI)V 2, (18)
43
The average bulk temperature, Tb, in Eqs. (18) and (19) is calculated from the following equation, (Tb - TI)/(Ta - Tb) = Nui/(Nui + Nuo)"1. (19) The left-hand-side of the above equation is the average dimensionless enclosure temperature between the inner absorber and outer transparent envelope boundary layers, cf., Kuehn and Goldstein (1976, 1978). It has been assumed that the convective heat transfer in the annular space between the two concentric tubes is totally suppressed, since the annular space is relatively small (~ 1 mm). Therefore, the heat transfer is via simple conduction, i.e., hi = 2~'ai~l,oln(Dl,i/Da, o) .(20) In order to develop an expression for Qa(t) in Eq. (5), the useful thermal energy removed by the heat exchange fluid as it flows through the absorber tube, we consider a small section of thickness Ax in the fluid flow direction (longitudinal conduction in the absorber is neglected). The energy balance on this infinitesimal element is given as Mfcf(~iTf/~ix)LaAx = AaUf(Ta-Tf)Ax (21) The overall heat transfer coefficient from the absorber tube to the fluid is given by Uf = [(Da, o/2K~a)hl(Da, o/Da, i) + 1/hf] -1, (22) where the heat transfer coefficient between the fluid and the tube, he, is given for the laminar flow regime (Ref< 2100) by Hedderich (1982), as Nuf = 1.86(Ref Prf Da.i/La)l/3(l.L/~.tf)~ , (23) where Re f 4Mig(l.t/1;Da,i). (24) This above equation is the Sider & Tate version of the Dittus-Boelter equation for viscous liquids, but neglects the factor (B/l.tf)~ which is close to unity for the heat exchange fluid under consideration, viz., water. The relationship for the fluid temperature at the outlet from the solar collector is obtained by integrating Eq. (21) under the following boundary conditions: at x=0, Tf--- Tf, i, where Tf, i = Tst,o for our closed loop system and at x = Lx (La -> Lx), Tf = Tf, o = Tst,i), and is given as Tf, o = Ta + (Tf, i- Ta)exp[-Uf-Aa/(Mfcf)Lx/La]. (25) The useful thermal energy Q~(t) is given as Qdt) = UaA~(Tf,i - Tdt)), (26) where Ida = Mfcg'Adl - exp{-Uf-Aa/(Mfcf)Lx/La}]. (27) The physical properties of the heat exchange fluid, of., Eqs.(23) and (24), are determined at its outlet temperature. The thermal energy introduced into the storage tank Qst(t), of., Eq.(6), is given by Qst(t) = Mfcw(T~o - Tst). (28) and the overall thermal energy loss from the storage tank, through its walls, to the ambient is given by =
ISES Solar World Congress 1999, Volume III
44
solution of Eqs. (4) - (6). The values for the parameters used in the simulation are listed in Table 5.
U~t = 5st/V~t. (29) We have utilized a forward time-step marching (an explicit) finite-difference scheme for the numerical
Table 5. Values of parameters used in the simulation model.
(xa)~= 0.85 el = 0.92 ea = 0.9 C 1 "- 2740 Jkg-lK1 ca- 1900 Jkg-lKl
solar collectors were 9.6 and 27.7%, respectively. The maximum hourly efficiencies were 11.0% for the reference and 35.7% for the coaxial tubular solar collectors. Though these results are promising, viz., significant performance enhancement, the absolute performance values, outlet temperatures and efficiency, are still relatively low and thereby limit such a design to applications such as the heating of swimming pools. We intend to continue this study to determine both the optimum and practical dimensions, i.e., wall thicknesses, for such a coaxial tubular solar collector design. The simulation model utilizes the measured average hourly solar radiation, both global and normal incidence (cf., Eq. (1)), and ambient temperature values together with the initial water temperature and then calculates the inlet and outlet water temperatures at the end of a predetermined time interval. The model was validated by comparing the calculated to the measured inlet and outlet water temperature values. The results of such an analysis are shown in Fig. (2) for the reference tubular collector together with the average hourly efficiency values.
5. RESULTS AND DISCUSSION Tubular solar collectors The performance testing of the reference and coaxial tubular collectors were done in parallel in order to determine the performance enhancement, which can then be attributed to the coaxial design. The side-by side performance testing procedure assures that both prototype collectors are subjected to the same ambient conditions. A typical set of experimental data for these tubular solar collectors is shown in Fig. 1 together with the global solar radiation intensity incident on the collector surface. The latter being calculated from Eq. (1), utilizing the measured horizontal global and normal incident solar radiation. It is obvious that there is a significant enhancement in collector performance as a result of the coaxial design, viz., the addition of the transparent envelop for thermal insulation. The coaxial tubular solar collector achieved a maximum outlet temperature > 50~ whereas the reference collector exhibited a maximum temperature < 40~ These results are also reflected in the values for the collector efficiencies, cf. Figs. (2) and (3). The average daily efficiencies for the reference and coaxial tubular
900
60 x o
X
50 x
e~
E
X
X
0
0
0
800 700
@ m
L_
~- 30
X
o
o~ 4 0
I-
KI = 0.148 W m l K "1 ~a = 0.090 W m l K "1 r~t = 0.024 Wm-lK"1 Kf = 0.63 W m l K "1
Cw= 4190 Jkg-lK"1 Mf = 10.9, 16.9 kg.h"1 p ~ = 1.165 kgm"3 pw = 958 kgm-3 Pa- 1100 kgm"3 ~:m - 0.027 Wm]K 1
('171~)1-" 0 . 0 4
O
D
o
m
D
m m
|
I
B m X Global radiatio
8
I
I
9
10
11
I
Houm 12
"O
- 400
-200
mTi (ref) DTo (ref) I
5 0 0 .__.
-300
OTo
10
'O
.,,,.
m m
eTi
20
600
a m
m .D O
- 100
I
I
13
14
15
Fig. 1 Inlet and outlet water temperatures for reference and coaxial tubular collectors and corresponding global radiation
ISES Solar World Congress 1999, Volume III
45
data, May 26, 1999. The flow rates are 5.7 and 7.2 kgh-1 for the reference and coaxial tubular solar collectors, respectively.
50
12 -10
40
-
= 30
_0 ..... J
L_
i1) l:k
0-
_-0 .....
9 ...... n
E 20 I--.
=
10 .........................
8
J_ . . . . . . . . . . . . . . . . . . . . . . . . . . .
9
t ........
10
1
11
1
Houl
12
I
13
8
Ti (ref) expt'/ I Ti (ref) mod~ TO (ref) expt 1 To (ref) mod t Efficiency (%~
i
14
,I
15
v
u t--
1
-- 6 u --
4
--
2
uJ
,,I
16
17
Fig. 2 Comparison of experimental and simulation model values for inlet and outlet water temperatures for the reference tubular collector together with average hourly efficiency (%), May 26, 1999. Flow rate = 5.7 kgh"1.
A similar comparative analysis for the coaxial tubular solar collector is presented in Fig. (3) together with corresponding average hourly efficiency values. It is apparent that there is very good agreement between the values calculated by the simulation model and the experimental data. This simulation model will be utilized to help determine the optimum wall thickness dimensions for such a coaxial tubular solar collector design.
Selectively coated pol~'neric absorbers The selectively coated polymeric absorber plates all exhibited relatively high efficiencies aoutlet temperature during performance testing. In fact, the heat exchange fluid flow rates had to be increased in order to prevent damage to the prototype selectively coated absorber plates due to thermal stress. The flow rate during performance testing was 13.2 kgh1 for all selectively coated absorber plates. Results, typical for such performance testing, are shown in Fig. (4) and (5) for the 10 mm thick polypropylene-2 and 15 mm thick polycarbonate double-walled absorber plates, respectively.
ISES Solar World Congress 1999, Volume III
46
40
60
35
50 -
30 IJ
30
= r
-20
Q.
_---
E I- 20
qt
9
...... El
ql
10
r
O
'"
8
I
9
,
, I
10
I ........
I
11 H o u r 1 2
Ti expt'l
,
U m
- 1 5 ~ 111
Ti model To expt'l To model Efficiency
-10 -5
'
i
,
i
13
14
15
16
O
17
Fig. 3 Comparison of experimental and simulation model values for inlet and outlet water temperatures for the coaxial tubular collector together with average hourly efficiency (%), May 26, 1999. Flow rate = 7.2 kgh 1.
In both cases the observed average daily absorber plate efficiencies were relatively high, viz., 58.6 and 67.4% for polypropylene and polycarbonate, respectively. The maximum outlet water temperatures were in the range of 60~ and the average daily temperature gradient between the inlet and outlet streams were ~ 14~ in both cases. The average daily efficiencies, based upon a series of tests performed during April and May 1999, on the selectively coated
absorber plates were as follows: 1] (polycarbonate-1) = 58.6%, 1] (polycarbonate-2) = 49.3%, I] (polycarbonate) = 58.6%, rl (polypropylene-1) = 64.3% and (polypropylene-2) = 60.3% (of., Table 2 for details on the absorber plates). The maximum outlet water temperatures and average daily temperature gradients between the inlet and outlet were similar to those reported in Figs. (4) and (5).
70
1100
v
-
>" t,J 60
1000
900 800 700 - 600 - 500 - 400 -300
t,"
~9 50
-
..,,,.
"' 4 0
30 ~- 20 ~-9
EIO i-0
To
10
11
12
13
t~
-~ .o O
(%) -
9
._~
-200
Efficiency
8
r
o
o.,.., ,i-,a
14
15
100
16
Houl
Fig.4 Inlet and outlet water temperatures and average hourly efficiency (%) for polypropylene -2 solar collector and corresponding global radiation data, May 13, 1999. Flow rate = 13.2 kgh q.
ISES Solar World Congress 1999, Volume III
47
80
1100
>. 7 0
1 000 900
v
t,j t--
u
60
"I
El
800 - 700
4o = 30
r
o
...,.
- 600
"0
- 500
._~
er
- 400
L_
~
20
-300
.,-!. To
E ~10 I-0
"~ 0
200
Efficiency (%)
100 8
9
10
11
12
13
14
15
16
Hou=
Fig. 5 Inlet and outlet water temperatures and average hourly efficiency (%) for polycarbonate solar collector and corresponding global radiation data, May 19, 1999. Flow rate = 13.2 kgh1.
The results form the performance testing of the selectively coated absorber plates, to date, are very encouraging with respect to both efficiency and maximum outlet water temperature. The major design problem is to find a polymeric material that can be (a.) fabricated as a double-walled sheet, (b.) serve as a substrate for the selectively coated surface, and (c.) is stable under the thermal stresses to which a solar collector is subjected. Obviously, this is not a simple task; especially since the economics of the thermal energy conversion process is of utmost importance. We are also in the process of developing a simulation model for the selectively coated absorber plate solar collectors in order to optimize the design with regard to the double-walled sheet dimensions. 6. CONCLUSIONS The design and performance testing of two different solar collector design prototypes, fabricated from polymeric materials, has been presented. Such solar collectors have a number of inherent advantages vis-avis those fabricated from metals. These include the following: (a.) fabrication by techniques, such as extrusion and molding, which can, a priori, result in significant reduction in production costs; (b.) eliminate corrosion problems, especially with regard to their application in sea water desalination systems; (c.) reduced weight per unit area of solar collector which results in significant reductions in both shipping and labor/installation costs. These advantages are offset significantly by their inherently low thermal conductivity coefficients.
We have attempted to overcome this disadvantage by two very different approaches in the design. 1. Coaxial tubular solar collector: design concept that entails the use of an inner black tube as a solar absorber in intimate contact with an outer transparent tube as an insulator. In theory, the inner black tube should be of minimum wall thickness to compensate for its low conductivity and the outer transparent tube should have a wall thickness that optimizes the trade-off between the material insulation and its transmittance of the solar radiation. 2. Selectively coated polymeric absorbers fabricated from double-walled polymer sheets which function as the solar collector absorber plate. Their upper surface serves as the substrate for the selective coating and the heat exchange flows through the channels between the two walls. The coaxial tubular solar collector exhibited a significant performance enhancement relative to the reference prototype, a simple black tubular collector. The side-by-side performance testing of the tubular collectors gave an average daily efficiency of 27.7 and 9.6% for the coaxial and reference solar collectors, respectively. In addition, the coaxial collector achieved maximum outlet water temperatures > 50~ whereas the reference collector achieved a maximum outlet water temperatures < 40~ Though significant performance enhancement has been achieved, the absolute performance values, outlet temperatures and efficiency, are still relatively low and thereby limit such a design to applications such as the heating of swimming pools. We intend to continue this study to determine both the optimum and practical dimensions,
ISES Solar World Congress 1999, Volume III
48
i.e.,wall thicknesses, for such a coaxial tubular solar collectordesign. A simulation model, which utilizes the measured average hourly solar radiation and ambient temperature values together with the initialwater temperature and calculates the inlet and outlet water temperatures at the end of each hour has been developed. It has been validated by comparing calculated to the measured inlet and outlet water temperature values. This simulation model will be utilizedto help determine the optimum wall thickness dimensions for such a coaxial tubular solar collectordesign The selectively coated polymeric absorber plates all exhibited relatively high efficiencies and outlet temperature during performance testing. The average daily absorber plate efficicncies for all those tested were relativelyhigh, viz.,in the range of 50 - 60. The m a x i m u m outlet water temperatures in the range of 60~ and the average daily temperature gradient between the inletand outlet water streams ~ 14~ were measured for all selectively coated absorber plates tested. The results from the performance testing of the selectively coated absorber plates, to date, are very encouraging with respect to both efficiency and maximum outlet water temperature. The major design problem being to find a polymeric material that can be (a.) fabricated as a double-walled sheet, (b.) serve as a substrate for the selectively coated surface, and (c.) is stable under the thermal stresses to which a solar collector is subjected. Obviously, this is not a simple task; especially since the economics of the thermal energy conversion process is of utmost importance. We are also in the process of developing a simulation model for the selectively coated absorber plate solar collectors in order to optimize the design with regard to the double-walled sheet dimensions.
Subscripts transparent tube 1 a absorber tube ambient amb bulk b c collector cax coaxial tubular collector conduction cond convection conv f fluid i inside/initial/inlet outside/output/outlet 0 ref refetubular collector
sky
sky
st
storage tank water
w
NOMENCLATURE A c
D G g h L M m N Nu Pr Q Ra Re T t
U
area (m2) heat capacity (JkglK ~) tube diameter (m) solar radiation ( W m "2) gravitationalconstant (ms 2) heat transfer c o e f f i c i e n t (Wm2K "1) length (m) mass flow rate (kgs "1) mass (kg) number of risers Nusselt number Prandtl number energy (W) Rayleigh number Reynolds number temperature (K) time (s) overall heat transfer coefficient (Wm2K 1)
Greek O~
B 8 s K
0
g P 0
solar altitude angle slope hour angle thickness (m) emissivity thermal conductivity (WmqK "1) latitude incident angle dynamic viscosity (kgm-ls"1) density (kgm"3) Stefan-Boltzmann constant (Wm2K -4) transmittance-absorptance product
ISES Solar World Congress 1999, Volume III
Acknowledgment- This research was supported under Project No. GR.01463 E 1071, Joint German - Israel Research Program, Israel Ministry of Science Bundesministerium Rir Bildung Wissenschaft, Forschung und Technologie. REFERENCES Dickinson, W.C., Clark, A.F., Day, J.A. and Wouters, L.F. (1976) The shallow solar pond energy conversion system. Solar Energy 18, 3-10. Duffle, J.A. and Beckman, W.A. (1980) Solar Energy of Thermal Processes, 762 pp., Wiley Interscience, New York. Eckert, E.R.G. and Drake, R.M. (1972) Analysis of Heat and Mass Transfer, 675 pp., McGraw-Hill Book Co, New York. Garg, H.P., Chakravertty, S., Shukla, A.R., Agnihotri, R.C. and Indrajit (1983) Advanced tubular solar energy collector: A state of the art. Energy Convers. & Mgmt 23, 157-169. Gerich, J. (1977) An inflated cylindrical concentrator for producing industrial processing heat. J.Proc. FRDA Conf. Concentrating Solar Collectors, Atlanta, Georgia 2, 103-115. Grimmer, D.P. and Moore, S.W. (1975) Practical aspects of solar heating: A review of materials used in solar heating applications. Los Alamos Scientific Laboratory, University of California, LA-UR-1752, Los Alamos, New Mexico. Gopffarth, W.H., Davison, R.R., Harris, W.B. and Baird, M.J. (1968) Performance correlation of horizontal plastic solar water heaters. Solar Energy 12, 183-196. Harris, W.B., Davison, R.R. and Hood, D.W. (1965) An experimental solar water heater. Solar Energy 9, 193-196. Hedderich, C.P. (1982) Design and optimization of aircooled heat exchangers, ASME J. Heat Transfer 104, 683-690. Khanna, M.L. (1973) A potable-type solar water heater. Solar Energy 15, 269-272. Kudish, A.I. and Wolf, D. (1978) A compact shallow solar pond hot water heater. Solar Energy 21,317-322. Kudish, A. (1981) Sede Boqer shallow pond project. Energy 6, 277-292. Kuehn, T.H. and Goldstein, R.J. (1976) Correlating equations for natural convection heat transfer between horizontal circular cylinders, Int. J. Heat Mass Transfer 19, 1127-1134. Kuehn, T.H. and Goldstein, R.J. (1978) An experimental study of natural convection heat transfer in concentric and eccentric horizontal cylindrical annuli, ASME J. Heat Transfer 100, 635-640. Meinel, A.B. and Meinel, M.P. (1976) Applied Solar Energy, p. 13, Addison-Wesley, Reading, MA. (3zisik, M.N (1977) Basic Heat Transfer, 572 pp., McGraw-Hill Book Co, New York.
49
Rabl, A. (1985) Active Solar Collectors and Their Applications, 503 pp., Oxford University Press, New York. Rommel, M., K6hl, M., Graf, W., Wellens, C., Brucker, F., Lustig, K. and Bahr, P. (1997) Corrosionfree collectors with selectively coated plastic absorbers. Desalination 109, 149-155. Swinbank, W.C. (1963) Long-wave radiation for clear skies. J. Roy. Meteoro. Soc. 89. Tabor, H. and Zeimer, H. (1962) Low cost focusing collector for solar power units. Solar Energy 6, 55-59. Tsilingiris, P.T. (1997) Design, analysis and performance of low-cost plastic film large solar water heating systems. Solar Energy 60, 245-256. Whillier, A. (1963) Plastic covers for solar collectors. Solar Energy 7, 148-151.
ISES Solar World Congress 1999, Volume Ill
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STUDY OF A MIXED (WATER OR AIR) SOLAR COLLECTOR Sylvain LALOT M.E.T.I.E.R., E.I.P.C., Campus de la Malassise BP 39, 62967 Longuenesse Cedex, FRANCE, Telephone : +33 3 21 38 85 10, Fax : +33 3 21 38 85 05, E-mail :
[email protected]
Abstract
-
paper presents a new concept of solar collectors. Usually, solar collectors are designed as water collectors or as air collectors. Here is presented a collector that can be used as a water collector as well as an air collector. First, the governing equations are given for steady state, explicitly including the ratio of the actual convective surface to the collector surface. Then it is shown that the product of the actual convection coefficient with the surface ratio plays a main role in the collector efficiency. Then it is shown that is possible to compensate the low value of the convection coefficient for air by a high convective surface. This can be achieved using standard heat exchangers components : finned tubes blocks. The idea is then to use the air side to build an air collector and the tube side to build a water collector. The feasibility of such a collector is shown and the performances of the proposed collector are given trough the efficiency curves. It is shown that the mixed collector has performances comparable to those of standard collectors; the performances of the air collector being higher than the performances of the water collector. Improvements are proposed such as the use of selective coatings and technical cover. Finally the dynamic behavior of the collector is briefly studied. It allows the determination of the thermal capacity of the collector. This
1. INTRODUCTION The scientific study of solar collectors has begun many years ago (Desautel, 1978) and many geometries has already been proposed : flat plate non-concentrating collectors, parabolic collectors... The fluid which is used may be water, air, oil, molten salt . . . . To increase the efficiency of the collectors, progress has been made on the coating of the absorber, on the geometry of the absorber (mainly for air collectors) , on the quality of the cover that acts as an infra-red barrier. Many devices has been introduced in the manufacturing of the collector to decrease thermal losses : anti-convection cells, transparent insulation (Aronov and Zvirin, 1997) . . . . To characterize accurately a collector and to manage efficiently a whole installation, the resource itself and its measurement are widely studied (Perrin de Brichambaut and Lamboley, 1974) : direct solar radiation and indirect solar radiation. In the last few years some new concepts have been developed. Hybrid collectors have been proposed : one part of the radiation is transformed by photovoltaic cells, another part is absorbed by a cooling medium (air for instance). But in the large majority of the cases, the solar collector is designed to heat one fluid. Here is presented a collector that can be used as a water collector as well as an air collector. In this case, the energy is used either by water or by air. It could be used by both at the same time, but this has not been already tested. First, the governing equations are presented for steady state, introducing explicitly the ratio of the actual convective surface to the eaptation surface. Then a description of the proposed collector is derived from the equation giving the efficiency of a
collector. Then, it is shown that the mixed collector has a good efficiency for water as well as for air. Then the governing equations of transient states are given. Experimental results allow the computation of the thermal capacity of the collector. Finally, the determination of the time response of the collector is achieved for both fluids, air and water. It should be noticed that the present work has been carried out for a company some times ago and that confidentiality has delayed the publication of the results
2. GOVERNING EQUATIONS In general, a water solar collector is built from "tubes" where the fluid flows, and from solid parts that receive the solar energy and transmit it to the fluid. For air solar collectors, it has been found that it is interesting to increase the convective surface. Figure 1 shows a schematic of a typical absorber. In this representation, the cover is characterized by its ability 2" to transmit the energy; the thermal losses are represented by a conductance g . In a first analysis, it is possible to group the characteristics of the cover and of the absorber in a single characteristic/70 = T E
E.
This study focuses on collectors in which the fluid has only one pass in the absorber. In this case the energy balance may be written as follows : E
_K
(1)
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51
Fig. 1. Schematic of a solar collector and"
r: (x) = r:, +
dr: :x) = kt Aa (to : x ) - r: :x))
(2)
Ac
K,
(or+ K, )r~ c z L
(4) Introducing the reduced mass flow rate n ~ - - ~
r&
, the
wL global convection coefficient ~ = k t
loss conductance
K,
K -~, wL
Aa
Ac
, and the reduced
Equations (1) and (2) lead to
the governing differential equation of the evolution of the fluid temperature 9
dr:(x) ~o I~ - K , ( r : ( x ) - r . ) -
a (a+K,)~
c:
It is then easy to find the evolution of the fluid temperature"
From Equation (4) it is possible to calculate the efficiency of a collector. The efficiency is defined as the ratio of the actual energy taken by the fluid n~c: (T/o - T/,)to the normal incident radiation A c I s . It may be written as follows"
2 r ~ c / thF _ct_K~_ K, L2 (a + K, )n~ :
](
rio-
gr (Tf o + Tf i )/ 2 - Too"~ IS
(5) For given operating conditions, the efficiency of the collector depends on the quality of the heat exchange between the absorber and the fluid. This is shown in figure 2"
ISES Solar World Congress 1999, Volume III
52
_
0.90.80.70.60.50.4-
0.5
0.3-
ot
0.2-
oJ + K r
0.7 A
0.9
..-o--1
0.1I
0
Kr
I
0.5
1
I
I
1.5
2
Fig 2" Influence of the convection coefficient on the efficiency of a solar collector For a typical water solar collector, it is possible to suppose 9 let = 2 0 0 0 W / m 2 K , K r = 1 0 W / m 2 K , ~ Ac
to ~ a+K,.
= 0.99"
0.5" This leads
The absorber surface is large enough to
This can be achieved using a well known technology used in the manufacturing of heat exchangers : the use of finned tubes. Figure 3 shows a detail of an uncoated finned tubes block. The geometrical characteristics are : - fin pitch : 1.7 ram, tube pitch : 35 ram, - fin height: 35 mm. -
assure a good efficiency. But the convection coefficient between air and an absorber is much lower, it can be considered that kt = 1 0 0 W / m 2 K is a large value. So, to get a good efficiency, the area ratio has to be at least 10.
Fig. 3. Detail of an uncoated finned tubes block
53
ISES Solar World Congress 1999, Volume III
So, the area ratio is large"
-
A,, .
-
20.17
.
A~ The use of fins allows the use of both convective surfaces, as usual in a heat exchanger. So, this leads to the fact that the absorber could be used to heat air (along the fins) or water (in the tubes). To prove the feasibility of such an solar collector, a prototype has been manufactured. It has been built using standard finned tubes blocks (painted with a standard black paint, E = 0. 9), a polycarbonate cover ~" = 0. 83, and rock wool as bottom insulation. The fins are made of aluminum and the tubes are made of copper.
3.
PERFORMANCES COLLECTOR
OF
THE
From these results it is possible to calculate the characteristics of the water collector"
7"1o = O. 7 9 3 K r =8.55W/meK This shows that the efficiency of the water collector is close to the efficiency of standard collectors. The performances could be improved using a selective coating instead (Tal-tarlo I. and Zvirin Y., 1988) of standard black paint, and using a technical cover (Zvirin Y. and Avichai Y.,1989).
NEW It can be seen that the efficiency of the air collector is higher than the efficiency of the water collector in the first part of the efficiency curve. This can be explained by three facts. First, there is no need to take the efficiency of the fins into account :
The prototype has been tested at the French technical center "CETIAT" (CETIAT, 1984). For steady state, the results can summarized by the efficiency curves (Figure 4). For water, the variation of the mean temperature is obtained by the variation of the inlet temperature (using an external heater). So, equation (5) shows that the efficiency is a straight line. For air, as the inlet temperature is the ambient temperature, the variation of the mean temperature is obtained by the variation of the mass flow rate. In this case, to increase the mean temperature, one has to decrease the mass flow rate. This induces a decrease of the convection coefficient. Then equation (5) and figure 2 show that this leads to a decrease of the efficiency; then the efficiency curve is no longer a straight line.
/7o is higher. Secondly, the area ratio is very high and the first term i n equation (5) is higher for the air collector. Thirdly, the heat losses due to the convection over the absorber are null : the heat is used by the flowing air, so
K,
is reduced. Here again the performances could be
improved using a selective coating instead of standard black paint, and using a technical cover.
_
0.9
-
0.8
-
0.7
-
0.6
-
0.5
-
0.4
-
0.3 0.01
air .." w a t e r
r
00 s
i
002
1
(r o +
O Fig. 4. Efficiency curves of the mixed collector
ISES Solar World Congress 1999, Volume III
54
4.
DYNAMIC THE NEW
CHARACTERISTICS
OF
COLLECTOR
The dynamic behavior of a solar collector is governed by the following equation :
rlo ls = K, (T/ - T= )+ a + K , Ln~ c/
+
~)x TCA + +
o~+K, cz
(6)
P.r c/V,
TCA n~ c: L ~ ~ T: ~
It can be seen that the time response of the water collector is about 5 minutes. As there is no analytical solution to equation (6), it is necessary to use a numerical method to deduce l~om the experimental data the value of the thermal capacity of the absorber. A standard finite-difference technique has been used, and it has been found that the value of the thermal capacity of the absorber is TCA=12 kJ/m2K. This value can be compared with the known values; the weight of aluminum in the absorber is about 5.2 kg/m2, the heat capacity of aluminum is about 880 J/kg K; so the thermal capacity of aluminum is about 4900 J/m2 K. The weight of copper in the absorber is about 12 kg/m2; the heat capacity of copper is about 400 J/kg K. So the thermal capacity of the copper tubes is about 4800 J/m2 K. This makes about 9700 J/m2 K for the absorber. We can deduce that the thermal capacity of the insulation is about 2300 J/m2 K.
+
a ~x~t p: c: V r TCA ~ 2 T/
Using the value of the thermal capacity of the absorber, it is possible to numerically calculate the time response of the air heater. It is found that the heater also needs 5 minutes to reach the stabilization.
~t 2
To determine the thermal capacity of the absorber, T C A , it is possible to use the response to a step of energy. This has been done for the water collector. The results given by the CETIAT are given in figure 5.
T:o (t ) - T: ,
_
0 0.9-
~
X X
.
T:o ( +Oo)- T:,
X
X
X
x
0.8-
0.70.60.50.40.30.2Time (minutes)
0.1I
0
1
I'
I
I
I
I
2
3
4
5
6
Fig. 5. Response of the water collector to step of irradiation
ISES Solar World Congress 1999, Volume III
55
Greek symbols
5. C O N C L U S I O N S
O~
It has been shown that the fine study of the governing equation of a solar collector can lead to a new concept for a solar collector. Based on the geometry of finned tubes heat exchangers, the proposed collector is able to heat water or air and the experiments have shown that in both cases the collector has a good efficiency. It has also be shown that the time response is quite short and independent of the fluid used.
ACKNOWLEDGMENTS
global convection coefficient
= kt A a /A c
W/m2K
E T]
emissivity collector efficiency
dimensionless dimensionless
/7o
maximum efficiency =E~'E
dimensionless
pf
density of the fluid
kg/m 3
"/"
transmittance of the infrared barrier
dimensionless
REFERENCES
The author would like to thank the French subsidiary of GEA for its technical and financial support.
Desautel J. (1978). EDISUD, Paris
Les
capteurs h61iothermiques.
Aronov B. and Zvirin Y. (1997). Theoretical investigation of solar collectors with transparent insulation covers by a novel calculation algorithm. In Proceedings o f lSES World Congress, Taejon, Korea
NOMENCLATURE Aa
absorber convective area
m2
Ac
collector aperture area
me
cf
specific heat of the fluid
J / kg K
E
fm efficiency
dimensionless
Is
normal solar irradiance
W / me
CETIAT (1984). Proc6s-verbal d'essais n~
K
thermal losses conductance
W/ K
Kr
reduced loss conductance
Tal-tarlo I. and Zvirin Y. (1988). The effects of radiation properties of surfaces and coatings on the performance of solar collectors. In J. Solar Energy Eng., vol. 110, pp. 217-225
= K/(wL) L
length of the collector
n~
total mass flow rate
kg / s
n~
reduced mass flow rate
convection coefficient
= n~/(wL)
kg/s m2
t
time
s
Ta
absorber temperature
K
Tf i
inlet fluid temperature
K
Tfo
outlet fluid temperature
K
T**
ambient temperature
K
T C A thermal capacity of the absorber
Vr
reduced volume of fluid
J / K m2 3 m / me
w X
width of the collector abscissa
m m
per square meter
CETIAT (1984). Proc6s-verbal d' essais n~
W/m2 K W / me K m
kt
Perrin de Brichambaut C. and Lamboley G. (1974). Le rayonnement solaire au sol et ses mesures. Editions Europ6ennes Thermique et Industrie, Paris
Zvirin Y. and Avichai Y. (1989). Improving the efficiency of solar collectors by glass coatings. In Proc. ISES Solar World Congress, Kobe, Japan
ISES Solar World Congress 1999, Volume III
56
UNCERTAINTY IN SOLAR COLLECTOR TESTING RESULTS Emmanouil Mathioulakis, Kostantinos Voropoulos and Vasilis Belessiotis Solar & other Energy Systems Laboratory, NCSR
>, 15310 Ag. Paraskevi Attikis, Greece Tel. +301 6544592, Fax +301 6544592, E-mail: [email protected] Abstract - A systematic assessment of all experimental error is presented leading to the determination of the uncertainty in the solar collectors testing results. The use of specific statistical tools allows not only the evaluation of the reliability of the testing procedure itself, but also the quantification of the goodness of fit and the prediction of the uncertainty in the collector instantaneous efficiency.
1. INTRODUCTION The basic scope of solar collector testing is the determination of the collector efficiency by conducting measurements under specific conditions defined by international Standards. The experimental results of testing lead to determination of the parameters of a more or less complex model capable of satisfactorily describing the energy behavior of the collector. The equation derived is considered to express the specific collector and can subsequently be used to predict its output under any conditions. Although several more elaborated models and testing methods have been proposed by various authors (Perers, 1997), in the present study, the Standard ISO 9806-1 is examined, mainly due to its extensive use and international application (ISO, 1994). More specifically, it is assumed (Duffle and Beckman, 1991) that the behavior of the collector can be described by a 2 or 3-parameter single node, steady state model n=/( T~): n= no-U0 Ti* n=no-U1
Ti'- u2 G (T i• )2
(la) (lb)
The above equations 1a and lb as well as the whole analysis presented in this paper are also valid for reduced temperature difference T'm calculated with respect to the mean collector fluid temperature. In this case the variable ~, where it appears, must be replaced by T'm. During the experimental phase, the output, solar energy and the basic climatic quantities are measured. During analyzing the data, a least square fitting of the model equation is performed on the measured data, in order to the determine parameters no and U0 or no, U1 and U2. In practice, this procedure determines only the equation of the collector behavior without calculating the uncertainties in the determined parameters, and thus the suitability of the concerned model is not evaluated with statistical criteria. Despite the widespread use of testing and the great importance of the testing results, an objective and standardized method for the determination of uncertainty in test results is still lacking. The question of uncertainty is crucial if one wishes to investigate the efficiency of each model, to consider the experimental uncertainties and to determine the uncertainty in the model parameters. It is noted here that only a limited number of publications deal with the accuracy of test results of solar thermal devices, mainly in the context of test
procedures of solar water heaters (Burges et al. 1991a, 1991b). A corresponding analysis for solar collector testing methods has been proposed by Proctor (1984a, 1984b, 1984c). In this analysis only the uncertainties related to measuring device errors and the standard least square technique have been considered. However, as will be discussed later on, this approach is equivalent to assuming a good fit and prohibits an independent assessment of goodness-of-fit In this publication we develop the general rules of uncertainty analysis and their application in a typical case of a commercial collector tested according to ISO 9806-1. The test were conducted in the Solar & other Energy System Lab which operates under the EN45001 Quality Assurance System, with strict respect to the requirements of the testing standard. 2. UNCERTAINTIES EXPERIMENTAL DATA
ASSOCIATED
WITH
The terminology used in uncertainty calculation is often confused, leading to different interpretations. In area of testing performed according to commonly accepted standards, uncertainties in experimental data should be determined according to the recommendations of ISO VIM (1995), by taking into account Type A and Type B uncertainties. Type A uncertainties are those determined by statistical means while Type B uncertainties are determined by other means (ISO, 1995a; ISO, 1995b). The uncertainty which is associated with each measurement, is the accumulated result of the uncertainty of the measuring insmmaent (Type B uncertainty), the uncertainty which represents the deviation of the measured value during sampling of data (Type A uncertainty) and the uncertainty which derives from the fact that the measurement may not represent the true value of quantity. Generally, in cases where an attempt is made to describe the behavior of a certain system with an approximate model, a distinction on the following should be made: 9 On the one hand, the Type A and Type B uncertainties which characterize every measurement itself and which are related to the quality of the measuring instrument and the stability of the measurement. These uncertainties can be determined quantitatively. 9 On the other hand, the uncertainties which are related to the degree to which the measurement or the model is representative, and which characterize the quality of the methodology followed. These uncertainties cannot be
ISES Solar World Congress 1999, Volume III
determined quantitatively and, after all, their determination has no meaning. Their influence is reflected on the ability of the methodology used (model and testing method) to describe the phenomenon. If, for example, it is proved that the certain experimental results are not represented satisfactorily by eqn (1 a) or (1 b), the whole methodology, or the suitability of the specific equation is in question. In our case, Type A uncertainties derive from the statistical analysis of the repeated measurements at each point of the steady-state operation of the collector. It should be brought in mind that, according to the Standard, N measurements are taken for 15 minutes (about 30 measurements), and the average value for each measured quantity is found. For every operation point of the collector, the best estimate of a quantity X is the arithmetic mean X of the N observations xj and its Type A uncertainty is the standard deviation of the mean (Fuller, 1987): N
(xj -~)~
57
provided by calibration certificates of sensors used for the measurements for this study, leads to the values of Table I. In most cases a measurand Y is determined indirectly from N other quantities X1, X2. . . . XN through a functional relationship Y=/(X1, X2.... XN). The standard uncertainty in the estimate y is given by the law of error propagation (ISO, 1995a; Fuller, 1987):
%=
dfl
2+2
~~u(xi,xj)l
~
(4)
where u(xi,xj) is the covariance associated to xi and xj. In our case, eqn (4) is used for the evaluation of combined uncertainty in the efficiency values n and of the reduced temperature difference Ti', which are calculated as a function of Tin, Tout, AT, Ta, m, G and Ar The calculation is conducted following the steps described in the flow chart of Figure 1.
0.5
(2)
j=l
N(N-~)
O-&x -
By nature, Type A uncertainties depend on the specific conditions of the test. Thus, they include the fluctuations in the measured quantities during the test which lie within the limits imposed by the Standard, and also the fluctuations in the testing conditions not considered by the model. Such fluctuations concern, for example, the air speed or the percentage of diffuse irradiance in global. Type B uncertainties derive from the calculation of uncertainties over the whole measurement, taking into account all available data, such as sensor uncertainty, data logger uncertainty etc. Although the Standard defines the upper limits of the accuracy of the measurements, the uncertainties that have to be taken account are the ones associated with the specific sensors used in the test. If there are more than one independent sources of uncertainty, (Type B or type A) ui, the final uncertainty is calculated according to the general law of uncertainties combination (Dietrich, 1991): (3) i
Figure 1: Propagation of uncertainties and synopsis of fitting procedure
Table I: Type B uncertainties in measurements The application of the above methodology for the calculation of Type B uncertainty based on the information
Figure 2 shows the expanded standard uncertainties ~n and ~T* of n and T: respectively, as calculated for a specific black-painted collector for each measurement point. The horizontal bars refer to ~T* and the vertical ones to an. In order to show the figure more clearly, only some of the 32 points (measured in the laboratory) are presented. The values of the uncertainties are presented in figure 2 as expanded uncertainty ax, as is the usual practice. The expanded uncertainty in an estimate x is obtained by multiplying the
ISES Solar World Congress 1999, Volume III
58
combined standard uncertainty Ux by a coverage factor k=2, corresponding to a level of confidence of 95% (ISO, 1995b).
(6)
[Yi - y(xi;al, a2...a M)] 2 i=1
0.9
o.8
The problem with this approach is that, in reality, the typical deviation o is almost never constant and the same for all points, but that each data point (xi, Yt) has its own standard deviation 6i. Another very interesting alternative is the use of the weighted least square (g'ZS) method, which calculates, on the base of the measured values and their uncertainties, not only the model parameters but also their uncertainty. By this way a qualitative evaluation of fitting can be performed. In the case of WLS, the maximum likelihood estimate of the model parameters is obtained by minimising the chi-square function (Press et al., 1996):
,~
0.7
0.6 0.5 0.4
++
0.3 0.2
#
0.1
Z ~'= N ,yi-( v( xi;a~ a2...aM,/2 3~ 0 -0.01
i
/
0.01
i
i
0.03
i
i
i
0.05
(7)
i
0.07
i-~
u~
where 11/-is the variance of the difference yi-y(xi; al, a2 ...as): Figure 2: Values of n, T~ and combined standard uncertainties in n, In the case of the 3-parameter model the quantity G( T~')2 is treated as an independent variable, thus its uncertainty is calculated separately, by applying the law of propagation of errors on equation G(~ )2=(Ti-Ta)2/G. In fact a 2-dimensional linear fit is required, since a single variable n is modelled as a function of two variables ~" and G ( ~ )2. 3. THE FITTING PROCEDURE The general problem of fitting is to find a model with M parameters aj to represent a series of N observations (x. YO with the greatest accuracy: y(x)----y(x; al...a~
(5)
In the above equation a single variable y can be a function of either a single variable x or a vector x of more than one variables, in the case of a multidimensional model. The basic methodology is always the same (Press et al., 1996; Dietrich, 1991): a figure-of-merit function is selected, to give an indication of the difference between the real data and the model. After this, the model parameters are selected so that the value of the function is minimized. The deviations of the model from the real data can be attributed to experimental errors but also to model weaknesses. The least square (LS) method tries to give an answer to this question: given a set of parameters a:, a2...a~, what is the probability that this set is the desired one?. Assuming that every point of our data is associated to an error which follows a normal distribution around the "true" value with standard deviation o which is the same for all points, the maximalization of the probability that this is the correct set of parameters leads to the minimization of the ftmction:
u~ =Val~yi - y ( x i ;al,a2 ...aM) )
(8)
Since the parameters al...aM are to be calculated, not all the terms that appear in eqn (7) are statistically independent, for this the degrees of freedom are v=N-M. It emerges from eqn (8) that the quantity ~ depends on the experimental uncertainties ~ and ,~. With this consideration in mind, the chi-square merit function actually gives an idea about the relation between the model deviation from the experimental data and the uncertainties in the measurements. A relatively good model will be able to explain the deviations observed on the base of the experimental errors and the and the corresponding X2 function will have a value close to v. Among the advantages of the use of the weighted least square is the fact that the real experimental uncertainties are taken into account in determining the model parameters, the fact that it allows the calculation of the uncertainties in these parameters, and also that it gives a realistic estimation of goodness-of-fit. However, even in the case that a least square fitting is selected by neglecting the uncertainties ui in the phase of the calculation of parameters al...aM, the chi-square function and the goodness-of-fit can still be determined afterwards using eqn (7). From the values of ~ and v the probability Q(0.Sv, 0.5 Z~) that the data do not fit the model by chance can be calculated (Press et al., 1996, Bajpai et al., 1977):
e-tt~-~dt' a>O, r(a)=
Q(a,x)= F(a) x
t'-~e-tdt
(9)
0
The probability Q can be explained as a quantitative indication of goodness-of-fit for the specific model. Generally speaking, if Q is larger than 0.1, then the goodness-of-fit is believable. If it is larger than 0.001, then the fit may be acceptable, under certain conditions. If Q is less than 0.001, then the model (or the estimation procedure) can be called into question.
ISES Solar World Congress 1999, Volume III
In the case of solar collectors, where a 2 or 3-parameter model is concerned, the denominator in eqn (7) is written as follows: 2-parameter model Y=a+bX: t~
--'U~i+ b 2 u,q2
(10a)
2 +c2 Ux2i 2 (10b) 3-param. Y=a+bXI+cX2: u~=uy2i + b 2 Ux~ So, the purpose is to minimize eqn (7) with respect to al...aM. Unfortunately, as can be seen from eqns (7), (10a) and (10b), the occurrence of b and c in the denominator makes the eqn (7) non linear. Its solution by analytical methods is possible only if the uncertainty in xi can be considered negligible (Press et al., 1996). Otherwise, the solution is possible by using numerical methods for minimisation of non-linear functions. Generally, the requirements for the acceptance of a good fitting can be reported as follows (ISO, 1995a; Press et al., 1996, Bajpai et al., 1977): I. The goodness-of-fit, i.e. the probability Q(0.5v, 0.5Z2) that the data do not fit the model by chance, should be high or, equivalently, the chi-square statistic should be about the number of degrees of freedom. II. The determined parameters al...aM should be independent, i.e. Covar(ai, aj)<
Uu~ and
Uu2 in parameters no, U0, U1 and U2 is more complicated, because of the non-linearity present in eqn (7). Our strategy is therefore to find these uncertainties numerically. The method for the case of a 3-parameter model is presented below; for a 2-parameter model the same
59
methodology is followed. For a more detailed review of the mathematics of the method, see Press et al. (1996). Let K be a matrix whose NxM components k0 are constructed from M basic functions evaluated at the N experimental values of T~ and Gi( T~ )2 weighted by the uncertainty ui (M=2 or M=3 for a 2 or 3-parameter model respectively). Let also L be a vector of length N whose components li are constructed from values to be fitted, weighted by the uncertainty ui
ki,1 = 1., ki,2= Ti* ' ki,3 = Ui
, li, j -
Ui
Ui
(11) Ui
The normal equation of the least square problem can
be
written: (Kr 9K) * INV(C) = K r - L
(12)
where C is a matrix whose diagonal elements Ci,i a r e the variances (squared uncertainties) of the fitted parameters and the off-diagonal elements cki, i~j, are the covarianees between these parameters. Eqn (12) can be solved by a standard method, for example, by Gauss-Jordan elimination. It should be noted that the calculation of covariances between the fitted parameters is necessary to estimate the acceptance criteria of fitting and to calculate the uncertainty in the predicted values of efficiency n given. The calculation of collector efficiency for given values of irradiance and inlet water temperature can be easy done by entering the calculated parameters in eqn (1). The uncertainty in predicted values of n is calculated by eqns (13) and (14) for the 2 or 3 parameters model respectively. Eqns (13) and (14) derive from eqn (4), where the values of irradiance and inlet water temperature are supposed to be known without uncertainty. Similar relations can also be derived from eqn (4) in the case that the values of irradianee and inlet water temperature are accompanied by known uncertainties.
U~o + 2 T ;
Cov(no,Uo)[
(13)
2
+
+
+2 i"
2 GT;3 CovCU1,U2)+2 GT; 2Cov(no,U2)]~
+
(14)
5. RESULTS The results presented here concern the typical black-painted collector, mentioned in a previous stage. The efficiency and the reduced temperature difference T: were calculated for 32 steady-state operation points, as well their respective uncertainties. The experimental data, together with the 2-parameter and 3parameter WLS fittings, are shown in Figures (4). The graphical representation of the 3-parameter model is given only for indicative purposes, since a 3-dimensional graph is
ISES Solar World Congress 1999, Volume Ill
60
normally required for the complete representation of n as a function of ~" and G( T~)2.
.9
,
,
,
,
,
,
,
known coefficient of regression R 2 in the table of results. This coefficient is often used as a criterion of the suitability of fitting. In some cases a 3-parameter model describes the collector behavior better than the 2parameter model, especially for a collector with a black painted absorber. It is noteworthy that their difference is indicated not by coefficient of regression R 2, but by goodness-of-fit Q, which, in the case of the 2-parameter model was often below, or very close to, the acceptability limit. After all, R 2 indicates nothing about the quality of the model.
,
0.7
0.5
n= no-U1 "/i - U2 G('I i )2
Model: """.
Quantity
0
method o ffim'ng
0.3
LS
~ZS
0.748 6.67 KIWm "2 0.026 K2W2m"4
ou 1
0.6768 6.527 KqWrn "2 0.02418 K'2W2m"4 0.00493 0.389 K'lWm "2
ok
0.0068 K2Wnl "2
no
0.1 -0.01
m
0 01
0 03
0 05
0
07
Figure 4: Experimental data (o), 2-parameter model (dashed line) and 3-parameter model (solid line)
on 0
m
.
The results of least-square and weighted least square fitting for a 2-parameter model are presented briefly in Table (II), while Table (HI) contains the results concerning a 3parameter model. Despite the fact that the above results concern the specific collectorg, and that the results differ from one collector to another, the following generally applicable remarks can be made, based on the analysis of a large number of test results conducted in our laboratory:
9
Model: Quantity
.
nq
.
.
Uo
.
O~
m
OUo
9
. C o v ( % Uo)+.
.
23
v-N-2 . Q /~ Remarks
,
n= ng-U9 I i method o ffitting WLS LS 0.76 0.75 8.45 K'lWm "2 8.37 KqWm 2 0.03 0.077 K ' I W m "2 -2.1e-4 95 97 30 30 8e-9 5e-9 0.99 0.99 goodness of fit: goodness-of-fit: v e ~ low v e ~ low
Table II: Results obtained using Least Square (LS) and Weighted Least Square (WLS) for a 2-parameter model I.
The problem of the evaluation of the candidate models is emphasized more clearly with the use of the well-
u1 U2
39 29 0.09 0.997
39 29 0.1 0.997
goodness of fit: acceptable
goodness of fit: acceptable
x
v-N-3 .
P.
lg Remarks
Table III: Results obtained using Least Square (LS) and Weighted Least Square (WLS) for a 3-parameter model II. The acceptance criteria of the model parameters are almost always satisfied with the exception of U2, the expanded uncertainty of which sometimes exceeds its own value. If this happens, the fitting should be reconsidered, otherwise there is a possibility to accept negative values of parameter U2. In this case it is preferable to repeat the fitting by considering the 2parameters model. HI. The deviation of the model parameters values between LS and WLS is not particularly large in most cases. In spite of this, the quality of the fitting is improved significantly in the case of WLS. 5. CONCLUSIONS We have developed a methodology for the evaluation of uncertainties in the testing results of solar collectors according to ISO 9806-1 and for the assessment of the performance of the models in use. This methodology is based on estimation of the experimental uncertainties and on the implementation of the weighted least squares fitting.
ISES Solar World Congress 1999, Volume III
The basic conclusion is that a realistic assessment of the quality of the test results is not possible unless a check of the model suitability and of the whole testing procedure has previously been conducted. Implementation of the weighted least square permits this check and at the same time enables the assessment of the possible expected error when the collector characteristic equation is used. It must also be stressed out that if the goodness-of-fit is not acceptable, the combined standard uncertainties in model parameters will certainly not be reliable and the use of the model is hazardous. Finally, it is noted that the proposed methodology is not limited to testing but it can be used also for the evaluation of a solar collector model of any kind.
Dietrich C.F. (1991), Uncertainty, calibration probability, 2nd edn, Adam-Hilger, Bristol.
61
and
Fuller W.A. (1987), Measurement error models, John Wiley, New York. ISO (1995), Guide to the expression of uncertainty in measurements, ISO ed., Switzerland. ISO (1994), Standard 9806-1. Test methods for solar collectors - Part 1: Thermal performanceof liquid heating collectors including pressure drop, ISO ed., Switzerland. ISO VIM (1995), International vocabulary of basic and general terms in metrology, 2~a ed., ISO ed., Switzerland.
NOMECLATURE
Ac C
collector aperture, m 2 specific heat at mean temperature of water, J kg-1 K
G m
global incident solar irradiance, W m -2 mass flow rate through the collector, kg s-1
n
Ta
aTi
1
collector efficiency n=,~c(T~-Ti~) AcG ambient air temperature, ~ reduced temperature difference T"i =( Tin-Ta)/G, K W-lm2
Y Tm Tm To~t v
UA,q UB,q
Uq AT P
%
reduced temperature difference T=(Tm-Ta)/G, KW-1 m2 temperature of water in collector inlet, ~ mean temperature of water inside collector Tin-'--(Tin + To~t)/2, ~ temperature of water in collector outlet, ~ volume flow rate through the collector, m3 s1 Type A standard uncertainties in an estimate q Type B standard uncertainties in an estimate q standard uncertainties in an estimate q temperature difference AT=Tout-Tin, K density of water, kg m 3 expanded uncertainty at a level of confidence of 95% in an estimate q
REFERENCIES
Bajpai A.C., Mustoe L.R. and Walker D. (1977), Advanced engineering mathematics, John Wiley, New York. Bourges B., Rabl A.., Carvalho M.J. and Collares-Pereira M. (1991), Accuracy of the European solar water heater test procedure. Part 1: Measurement errors and parameter estimation, Solar Energy 47(1), 1-16 Bourges B., Rabl A.., Carvalho M.J. and Collares-Pereira M. (1991), Accuracy of the European solar water heater test procedure. Part 2: Long-term performance prediction, Solar Energy 47(1 ), 17-25 Daffie J.A. and Beckman W.A. (1991), Solar engineering of thermalprocesses, 2nd edn, Willey, New York.
Perers B. (1997), An improved dynamic solar collector test method for determination of non linear optical and thermal characteristics with multiple regression, Solar Energy, 59(46), 163-178 Press W., Teukolsky S.A., Vetterling W.T. and Flannery B.P. (1996), Numerical recipes, 2nd edn., Cambridge University Press, Oxford. Proctor D. (1984), A generalized method for testing all classes of solar collectors-I: Attainable accuracy, Solar Energy, 32(3), 377-386 Proctor D. (1984), A generalized method for testing all classes of solar collectors-II: Evaluation of collector thermal constants, Solar Energy 32(3), 385-394 Proctor D. (1984), A generalized method for testing all classes of solar collectors-l: Linearized efficiency equations, Solar Energy 32(3), 395-399 (1984)
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62
OPTIMIZED FINNED A B S O R B E R GEOMETRIES FOR SOLAR AIR HEATING COLLECTORS Klaus Pottier, Carl Martell Sippel, Andreas Beck and Jochen Fricke Bavarian Center for Applied Energy Research (ZAE BAYERN), Am Hubland, D - 97074 Wiirzburg, Germany Tel: ++49 931 70 564 44, Fax: ++49 931 70 564 60, E-mail: [email protected] Abstract - Solar air heaters are limited in their thermal performance due to the low density, the small volumetric heat capacity and the small heat conductivity of air. For high solar gains, an efficient thermal coupling between absorber and fluid is required, while the electrical power for the fan operation ought to be as small as possible. As heat transfer augmentation usually increases friction losses, an optimum fin geometry has to be found. For a solar ventilation air preheater, mounted on a south facing facade in Wiirzburg/Germany, optimized geometries have been derived. For a specific air mass flow rate of 70 kg/(m2h) the air gap for smooth absorbers without fins should be about 7 to 8 mm to get a maximized net energy output of 190 kWh/(m2a) during the heating season from October through April. Finned absorbers perform much better. Continuous aluminum-fins, 0.1 mm thick and spaced about 6 mm apart in a 30 mm wide air gap yield about 245 kWh/(m2a) net output. Offset strip fins do not show an improved performance compared to optimally spaced continuous fins, due to the larger electrical power for this geometry. However, offset strip fins yield high net energy gains for large fin spacings.
1 INTRODUCTION Solar air heaters offer some advantages over solar water heaters: Freezing or boiling of the fluid does not occur, they run at small fluid pressures and leaks decrease the thermal performante somewhat but do not disable the whole system. However, solar air heaters are limited in their thermal performance due to the low density, the small volumetric heat capacity and the small heat conductivity of air. Generally a solar air heater consists of a casing, which holds a solar absorber, a transparent cover and a back insulation. An air gap between cover and absorber reduces heat losses towards the front. The air flows in the gap between the absorber plate and the thermal insulation. To accomplish high thermal efficiencies heat has to be transferred efficiently from the absorber to the flowing air. This measure decreases the temperature of the absorber plate and thus reduces the heat losses through the glazing. In order to reduce the electrical power necessary to pump the air through the collector the pressure drop inside the heater has to be minimized. Measures for heat transfer augmentation generally increase the pressure drop. Since, according to Morhenne et al. (1990), the ventilation power may be as high as 20 % of the thermal output of the collector, optimization should address heat transfer and pressure drops. Heat transfer can be augmented e.g. via rough surfaces, displacement devices and/or fins (Fig. 1). Offset strip fins (Manglik and Bergles 1995) are oriented in flow direction. The offset is commonly uniform and equals half the fin spacing. Because boundary layers vanish and redevelop periodically in such channels, the heat transfer coefficient is high. The solar air heater of Kuzay et al. (1975) had 1 mm thick and 19 mm long aluminum fins, forming channels 8 mm wide and 25 mm high. The thermal efficiency reached 74 %, when operated with a specific air volume flow of 82 mV(m2h). Mattox (1979) discussed the utilization of offset strip fins for solar air heaters, too. This paper presents optimal fin geometries for a solar ventilation air preheater, mounted vertically on a south facing wall in central Germany. For solar collectors used to preheat ventilation air, the mass flow is given by the fresh air demand of the building. Since the flow rate and therefore the pumping power is independent of the solar insolation the entire
thermal energy gain through the full heating period has to be taken into account to optimize the heat transfer geometry of the collector (Heibel and Hauser 1996).
Fig. 1: Solar air heaters with heat transfer augmentation with single pane cover and air flow between absorber and insulation. a) and b): sectional drawings in flow direction with roughness inducing wires, according to Prasad and Saini (1988) and wavy passage as discussed by Piao et al. (1994). c) and d): sectional drawings perpendicular to flow direction with continuous fins as rectangular or triangular air flow channels, as studied by Diab (1981) or Kabeel and Mec~'ik (1998), respectively. 2 INVESTIGATED SOLAR AIR PREHEATER SYSTEM The investigated air preheater (Fig. 2) consists of three modules, each with about 3 m2 of absorber area. For architectural reasons the concept of an air heater with flow behind the absorber with single glass cover was chosen. This system is more complicated than unglazed transpired solar collector-systems, as described by Kutscher and Christensen (1992), but it is well adaptable to buildings with standard glass-facades. Since the air gap between the glazing and the absorber plate reduces thermal losses, this system can also be used to heat circulating air. The fresh air enters the module from the bottom and is warmed up while flowing to the top. During the heating season a fan forces the preheated air into the building. In summer, the air outlet inside the room is closed and an overheating protection flap opens passively. In this case buoyancy forces create an air flow through the collector which limits the absorber temperature.
ISES Solar World Congress 1999, Volume III
63
results of Ebadian and Dong (1998), who gave critical Reynolds numbers for sharp-edged rectangular channels with various aspect ratios, one gets an equations for all aspect ratios from 0 (fiat plate) to 1 (square channel): Re c = 2200 + 900 exp(- 4.754. fl).
(2)
The aspect-ratio fl of the channel is defined by
fl=
we~he for wo h~ '
(3)
with the channel width w~ (fin spacing) and the channel height h~. To simplify the heat transfer equations we use normalized lengths with respect to heat transfer and hydrodynamic flow, y* and y+, respectively: y* =
Fig. 2: Investigated solar ventilation air preheating system.
Y
D h Re Pr
(4)
Y Dh R e '
(5)
and The three modules were mounted onto the south facing wall of the ZAE research building (Fig. 3). The temperatures of the air stream, the glazing, the absorber plate and the outer wall surface were measured on various locations with platinum resistors. A flow sensor, as used for flow volume control in ventilation systems, in combination with a differential pressure gauge was used to determine the air flow rate. The system has also been used to validate the numerical model described in chapter 4, see Pottier et al. (1998a).
Y§ =
with y being the location of the fluid in the channel and Pr the Prandl number of the fluid. The pressure drop zip inside the channel is calculated via the apparent Fanning friction factorfavp which incorporates both the skin friction and the change in momentum rate in the hydrodynamic entrance region:
4L p u 2 AP= f a v v - ~ h - " ~ - .
(6)
L is the length of the channel, p the density of the fluid and u the average velocity of the fluid in the channel. For Re between Rc and 10000 we use the interpolation formula from Gnielinsky (1995) to calculate the transition Nusselt numbers:
Nut,. - (1 - 7') Nul(R%) + 7"Nut~(Re = 104),
(7)
with taking the average laminar Nusselt number at the value of Rec and taking the average turbulent Nusselt number at Re = 10000. The coefficient 7'is calculated by
Fig. 3: Test modules mounted on the south facing wall of the ZAE building in Wiirzburg/Germany with air inlet meshes on the bottom and overheating protection flaps on the top. 3 HEAT TRANSFER AND F R I C T I O N
R e - Rec 7' = R104 e -- - - ~ "
(8)
For the calculation of the transition friction factors Nu in Eq. (7) is replaced by favv- The heat flow for finned absorbers is calculated by taking the fin efficiency into account.
3.2 Absorber with continuous fins 3.1 Preliminary remarks The convective heat transfer coefficient in enclosed spaces is calculated by the Nusselt number, which is defined by Nu=
hDh 2
'
(1)
h is the convective heat transfer coefficient, 2 the thermal conduetivity of the fluid and Dh the hydraulic diameter. The flow is laminar for Re < Rec and turbulent for Re > 10000. Taking the
3.2.1 Heat transfer for laminar flow The local Nusselt number for thermodynamically developed flow (y* >> 1) is given by Shah and London (1978): Nu**,l= 8.235 9(1- 2.0421fl + 3.0853fl 2 -
2.4765fl 3 + 1.0578fl 4 - 0.1861fl5).
For short path lengths (y* << 1) VDI (1997) states:
(9)
64
ISES Solar World Congress 1999, Volume III
B
Nuo, l
~
1
(10)
3.44
Af + B f / ( 4 y + ) - 3 . 4 4 / ~
~
l + C f .(y§ -2
with B = 0.4849. The intermediate values were taken from Wilbulswas (1966), as reviewed by Shah and London (1978). With these values and Eq. (9) and (10) we get Nuo~ Nut = Nu**'l + ~]1 + C y *
(11)
C (1/4 < fl < 1) is given by
To get average Nusselt numbers between Yl* and Ye* the following integration has to be performed: 1 Y; , , Nu(y*) dy*. Y2 - Yl y~
-
The parameters Af, Bf and Cf are given by Shah (1978) for discrete values of the aspect ratio only. From fit curves the following relationships are derived: A f = 24.00-30.88 fl + 34.56 f12 -13.52 fl 3 ,
(19)
Bf = 0.6740+l.123fl+l.442f12-3.294fl 3 +1.485/$ 4,
(20)
(12)
C = 64.52 +434.2- (fl-1) 4 .
~uu
(13)
This gives
Cy=
(0.2900- 0.1312 fl + 16.86 fl 2 - 23.57 f13 + 9.460 f14). 10-4 "
=
Nu**,I +
, (y2 _ yl) ~V~"
(14)
'
with D(y')= In[ 2.4Cy'.O+Cy*)+2Cy'+ll.
(15)
In contrast to CFD-simulations shown by Merker (1987) Nu diverges for small flow lengths at the beginning of the channel when computed with Eq. (10). Therefore we calculate the heat transfer coefficient for the first knot by using the local Nu number and use the average Nu for the other knots. For the optimization procedure we calculate the average Nu by omitting the first 10 % of the flow path length.
(21)
3.2.4 Friction factor for mrbulent flow For turbulent flow the relationship from Colebrook, as given by Ebedian and Dong (1998), will be used:
f = 0.4091. [In(Re/7)] -2 . Nil I
(18)
(22)
To take the developing flow in the first sections of the channel into account an additional term, given by Altfeld (1985), is added to get the apparent friction factor: f~p = 0.4091-[In(Re/7)]-2 + 0.01625 D ~ . Y
(23)
3.3 Parallel plates 3.3.1 Heat transfer for laminar flow If the back wall of the solar air heater is adiabatic, the developed Nu is given by Shah and London (1978): Nu**,l = 5.385.
(24)
3.2.2 Heat transfer for turbulent flow The calculation of heat transfer and friction factor for turbulent flow in rectangular channels can be based on the equations for circular ducts, when the hydraulic diameter is corrected. By using the relationship from Jones (1976) we get a simple formula for the corrected hydraulic diameter Dh~:
For short path lengths we use Eq. (10) as described above. The intermediate values for Nu were taken from Heaton et al. (1964) to get a simple equation for all flow path lengths:
Dh~ = D h 9(0.6081 + 1.812 fl - 1.292 f12).
(16)
with E = 141. The average Nusselt number between Yl* and Y2* is given by
(17)
-~uut = Nu..,t "~
Nuo~ Nut = Nu**,l + ~ , l+Ey*
(25)
The average Nu for developed turbulent flow is given by 2B (arctan4Ey: - arctan E E ~ ~ ~u~, = (f/2)(Re-1000)Pr[l+ 2"4254(D~/Y)~
,
1+ 12.7 ~/(f / 2)(Pr 2/3-1) based on the results of Gnielinski (1976), Bhatti and Shah (1987) and Mills (1962). The friction factor f u s e d in Eq. (17) is given in Eq. (22). 3.2.3 Friction factor for laminar flow Shah (1978) gives the apparent friction factor for parallel plates and rectangular channels:
( y ; - y ; ) ~'ff
(26)
As described above we don't use )'1" = 0 in Eq. (26) to avoid computation errors for short path lengths. 3.3.2 Heat transfer for turbulent flow Here we use equations given by Altfeld (1985): Nut, = Nu..,, + F. exp(-G y*),
with
(27)
ISES Solar World Congress 1999, Volume III
Nu.~,t~ = 0.0158 Re ~ ,
(28)
j = 0.6522Re -~176 ~--0.1541~'0.1499r 9[1+ 5.269.10 -5 Re 1"34~fl0.504~-0.456~:-1.055]0.1
F =0.00181 Re+ 2.92,
(29) f =
(30)
G = 0.03795 Re Pr.
65
(33)
9.6243 Re -0"7422fl--0.1856~,-0.3053~:-.0.2659 9[1+ 7.669.10 -s Re 4"429/~~176162162176 ] 0"1
(34)
The average Nu is given by integration of Eq. (27):
Nu ~ = Nu.,~ 4
F (exp(-G y; )-exp(-G y; )) G(y;-y;)
4 NUMERICAL MODEL
(31)
3.3.3 Friction factor for laminar flow The friction factor for laminar flow is calculated with Eqs. (18) - (21).
3.3.4 Friction factor for turbulentflow The friction factor for turbulent flow is calculated with the equation from Beavers et al. (1971), extended by the developing flow expression, as done for Eq. (23) to getfapp: fapp =0.1268 Re-~
Dh .
(32)
Y
3.4 Absorber with offset strip fins Empirical equations to estimate heat transfer and pressure drop properties of small offset strip fin surfaces are available from Manglik and Bergles (1995) for compact heat exchangers only. Here the channel widths and lengths usually are in the order of some mm and smaller. A typical geometry is shown in Fig. 4.
4.1 Structure of the model To calculate solar heat gains of fiat plate solar air heaters the numerical simulation program "SoLuTion" S(S_.qlar-LuttTransiente-Simulatio._qn_) was developed. It is based on the method of finite differences, as shown by Kreith and Bohn (1986) and used by Blomberg (1996). Energy balances are computed for every knot of the modeled system. To calculate covers and absorbers with low heat conductivity (e.g. polymer materials), these parts can be modeled by two knots each. The wall behind the collector is designed to consist of two different layers (e.g. an insulation layer and a brick layer). Each of these layers can be modeled by several knots perpendicular to the air stream direction. The PC-based program allows for transient simulations of the heat gains in dependence of geometry, optical and material properties of the collector, air mass flow rate and time dependent weather data. Apart from conduction heat flows the model accounts for all longwave radiative and convective heat exchanges inside and outside the system as well as for insolation. The model was described in detail by Pottier et al. (1998b). Fig. 5 shows the graphical representation of the model.
Fig. 4: Offset strip fin heat exchanger, from Manglik and Bergles (1995). For solar air heaters, larger fins are necessary. These geometries differ not only in absolute size (which could be easily taken into account by using the theory of self similarity and dimensionless numbers for heat transfer) but also in relative size, e.g. the ratio of fin wall thickness to fin length and to fin spacing. The reliability of the available equations is not yet established. For this reason an apparatus has been built and heat transfer properties of offset strip fin geometries suitable for solar air heating collectors were investigated. These measurements are described in detail by Pottier et al. (1999). For the used geometries the heat transfer values are to some extent lower, however show the same tendency as the calculated values and lie within the error range of the given equation from Manglik and Bergles (1995). Therefore we use these equations for calculating offset strip fins for all Re numbers. The definitions o f t , ~ and ~ are shown in Fig. 4.
Fig. 5: Model with multiple knots in the x- and y-direction.
4.2 Validation of the model The following figures show the validation of the model for an absorber with 200 mm long offset strip fins, about 27 mm high and with an average spacing of 12.5 mm. The measurement was performed in December 1998 with solar insolation and air mass flow data as shown in Fig. 6. In Fig. 7 the measured and the calculated outlet air and cover temperatures are depicted. The temperatures coincide to within 2 K. Fig. 8 shows the absorber temperatures at three positions (bottom, center and top). With the accuracy of the insolation measurements of about 3 % and the accuracy of the air mass flow rate measurements of about 5 % the model predicts the temperatures and hereby the efficiency
66
ISES Solar World Congress 1999, Volume III
of the collector well. As shown by Pottier et al. (1998b) the model is accurate for a collector with a smooth absorber, too.
Germany. The solar and the longwave radiation data were corrected as described below.
5.1 Correction of insolation data Pottier et al. (1996) showed deviations from the insolation data given by the TRY to the 35 year-average data measured by the German meteorological service. To correct the insolation data both the beam and diffuse radiation values in the TRY were multiplied by the following factors given in Table 1: Table 1: Correction factors for the solar insolation in the TRY. Oct. Nov. Dec. Jan. Feb. Mar. Apr. 1.046 0 . 9 6 1 1 . 1 8 8 1.029 1.229 1.304 0.858
Fig. 6: Insolation and air mass flow rate through the collector. The rapidly changing insolation is welcome to test the transient simulation. The air mass flow rate was actively varied.
5.2 Longwave radiation data In the TRY the longwave radiation data, based on the effective sky temperature, is given for horizontal surfaces only. A vertical oriented surface exchanges longwave radiation with both, the ground and the sky. The simplest model to calculate the longwave radiation exchange would be to assume that the ground is at ambient temperature and the nearby sky seen by the collector is at the temperature calculated from the horizontal longwave radiation data given by the TRY. However, to be more accurate, we use ground temperatures computed by a transient model in dependence of climatic data and average soil physical properties as described by Weinl~er et al. (1999). The sky radiation depends on the effective air temperature and is higher for air layers close to the ground as for air layers high in the atmosphere. The sky radiation was calculated by the model of Unsworth and Monteith (1975), which has been adopted to the local sky conditions also described by Weinlader et al. (1999). 6 OPTIMIZATION PROCEDURE
Fig. 7: Measured and calculated temperatures for the outlet air and the cover as well as the temperature of the ambient air.
Fig. 8: Measured and calculated temperatures of the absorber at three locations (bottom, center, top). 5 WEATHER DATA Accurate weather data are needed in order to calculate yearly solar energy gains. All values for ambient temperature, humidity, wind speed, beam and diffuse solar radiation as well as for longwave radiation were taken on a hourly basis from the German test reference year (TRY), type 5, which is valid for central
In the previous chapters it has been shown, that we have a reliable theory, weather data and computing aids to calculate the long term thermal energy gain for solar ventilation air preheaters, mounted on vertical surfaces in central Germany. For a system optimization the following assumptions are made: a) The solar air heater preheats fresh air for ventilation purposes. The system is in use from October through April, including the nights. This timespan m o u n t s to 5088 hours. b) All heat gained by the collector can be used to decrease the building heating demand. This assumption holds for low insulated, heavy structured buildings. In well insulated, lightweight buildings, some of the heat gains cannot be used due to the smaller heating requirements. c) We assume an adiabatic building wall behind the collector, so there is no heat conduction through the wall into the interior. This assumption is good for well insulated walls, but, interestingly, with small errors it describes low insulated walls, too. In the latter ease there is a large heat flux through the wall to the outside, but most of this heat is recovered by the air in the collector. In case of a south facing wall with a U-value of 1.4 W/(m2K) equipped with a collector and with an air mass flow rate of 75 kg/(m2h) and a heat transfer coefficient from the absorber to the air of 80 W/(m2K) about 87 % of the wall transmission losses are regained. Throughout the heating season 241 kWh/m2 are
ISES Solar World Congress 1999, Volume III
d)
gained by the collector whereas with the adiabatic wall assumption a gain of 254 kWh/m s is calculated. As a rule of thumb, the whole thermal energy savings of a ventilation air preheating collector mounted on a south facing fagade amounts to the yearly heat gain of the collector calculated with adiabatic back plus the thermal energy transmission losses which the wall would have without the collector. The climatic data are given by the corrected TRY. The average values for the operation time are: solar insolation: 76.1 W/ms, ambient air temperature: 4.0 ~ effective longwave radiation temperature (ground and sky): 0.8 ~ wind velocity: 3.4 m/s.
300[
J
.~ 250-]
]pr-
i
(35)
is the heat-removal-factor. For an adiabatic back wall the collector-efficiency-factor F" is h
(37)
In these equations ('cOt)eft is the effective transmissionabsorption product of cover and absorber, h the heat Wansfer coefficient between absorber plate and air, U the overall heat loss coefficient of the collector, Ti the inlet air temperature, Ta the ambient air temperature, G the global insolation, A the absorber area, n~ the air mass flow rate and c the specific heat capacity of the air. For solar air preheating the inlet air is at ambient temperature and Eq. (35) is reduced to r~ = FR . ( r a ),g .
=
I
[]
o v A []
50
5'0
100
150
200
125 kg/(n~h) 100 kg/(n~h) 75 kg/(rn=h) 50 kg/(rnah) 25 kg/(m=h) 12.5 kg/(rn=h)
2~)
360
heat transfer coefficient [WI(m=K)]
With these assumptions we calculated the solar heat gain for a solar air heater, mounted on a south facing wall. The heater is single glazed, the air flows behind the 1 m wide and 2 m long absorber plate. To reduce the computation effort we first calculate the yearly heat gain of this collector in dependence of the heat transfer coefficient between the absorber plate and the air stream, with the air mass flow as a parameter. The heat transfer coefficient between the back wall and the air-stream was set to zero. As given by Duffle and Beckmann (1991), the thermal efficiency 7/of solar air heaters can be calculated analytically by
h+U
I
lOO
0
F'=~.
I
....
6.1 Yearly thermal energy gain
r/= FR. [(za),g- U(~-G To)] 9
I
67
Fig. 9: Yearly solar heat gains for ventilation air preheating on a south facing fagade in central Germany versus heat transfer coefficient, with the specific air mass flow rates as parameter. The data points are calculated via the numerical model, the curves are least squares fits. The average effective wansmission-absorption product for the heating period and a south-facing collector is (~:tz)eff = 0.772,
(40)
which differs substantially from (*'a)eff = 0.854 for perpendicular incidence. For stationary calculation taking all angles of incidence into consideration, one gets ( z a ) w = 0.825. For nonsouth collector orientations we get the following coefficients: Table 2: Effective transmission-absorption-product for nonsouth orientated single glazed air ~ reheating collectors. wall orientation north [ west east (,/'a)ar 0.673 [ 0.725 0.733 As the dependence of the average absorber temperature on the air mass flow rate is smaller for non-south oriented collectors, U =- 4.8 W/(mSK) is sufficiently accurate. Fig. 10 shows the yearly solar heat gains for a specific air mass flow rate of 75 kg/(m% and a heat transfer coefficient of 80 W/(mSK) between absorber plate and air for different orientations.
(38)
With Eq. (36), (37) and (38) and the results of the numerical calculations shown in Fig. 9, U and (z'a)~ff- averaged over the whole heating season - can be derived via fit curves, also shown in Fig. 9. The average heat loss coefficient amounts to O'=(4.8+0.88.exp(-83M(~g]] "~.m2K Vr
(39)
U is the higher the lower the air mass flow rate is chosen. This is due to the higher average absorber temperature which is accompanied by an increased radiative heat transfer coefficient in the case of a low mass flow rate.
Fig. 1O: Comparison of yearly solar heat gains for single glazed air preheating collectors mounted on differently oriented fa(}ades. Ts~ and Tgro=~ are the effective radiation temperatures as described in chapter 5.2.
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68
6.2 Yearly net energy gain
300
For the optimization of the absorber geometry the yearly solar heat gain Qsota,, the yearly electrical energy consumption of the fan Qa, the energy for manufacturing the fins Qa. and the efficiency for electricity generation fief are necessary. Knowing the heat transfer coefficient from the absorber plate to the fluid for smooth and finned absorbers based on the absorber area the yearly solar heat gain is:
solar heat gain
=--i
r 9 200_=
maximum net energy gain
m
>, 10oL 0 %
",,heat gain through fan
0
~--,....._,
Q,I,
= ~'G
A At ,
(41)
with the winter-averaged thermal collector efficiency ~', calculated by Eqs. (36) - (40), the winter-averaged insolation for the south facing facade G (76.1 W/m'), the absorber area A and the fan operation time At (5088 h). The electrical energy consumption of the fan is:
6pn~m
Qa = ~
,
(42)
P ~Tfan
with the pressure drop zip, derived from Eq. (6), the air mass flow rate n~, the average air density p and the efficiency of the fan r/f=,, which was supposed to be 63 % (70 % mechanical, 90 % electrical). The primary energy for manufacturing of the aluminum-fins q/~. is 172 MJ/kg (Corradini 1997). If the lifetime of the collector system is assumed to be 20 years, 1/20 of this energy will be lost in one year. Therefore the energy necessary for manufacturing of the fins, on a yearly basis, is
cO
0-
-100
....-
,,-~primary // pumping ,/ energy 4 ;
smooth absorber mass flow rate: 70 kg/(rrFh) 12 1'6 20 channel width [mm]
0
Fig. 11" Contributions to the net energy gain for a south facing collector with smooth absorber versus channel width. Fig. 12 shows the results for different orientations and different specific air mass flow rates. The optimum channel width is in the range from 5 to 10 mm. It ought to be chosen larger for higher specific air mass flow rates and for smaller average insolation values. 200
I
I
I
I
I
I
I
I
~150,....= .=_ m
Qf~ = q~mf~ 20
(43) '
with m/~. the mass of the fins used in the system. The efficiency for the electricity generation r/etis 31.4 % and was taken from FIE (1996). With these values the yearly net energy output Q,~ of the collector is:
P 1000 I= 0
/ /'/
=
I /
e~
(44)
Because the fan is located in the air stream, the electrical energy Qa, necessary for running the fan during 5088 hours, will increase the air temperature. Therefore it is an additional heat gain besides the solar heat gain Q,ola,.. Our goal is to maximize Q.~t with respect to absorber geometry, e.g. gap width, fin thickness, fin spacing and fin length. It is not necessary to take the pressure drops of the in- and outlet and the manufacturing energy of other collector-parts into account. Q,t would be smaller in this case, but the optimal geometry would still be the sanle.
7 OPTIMIZATION RESULTS
7.1 Smooth absorber without fins Fig. l l shows the contributions to the net energy gain for a smooth absorber. Because the absorber is finless the only energy loss is due to the primary energy consumption of the fan. The yearly net energy gain amounts to about 190 kWh/m,` for a channel width of 7 mm.
./" /
-
............... _?----
__ 6.5 mm
,n ~ ~ - - - e a s t 70 kg/(rrF~" ,,.,.z_ ..... - - - e a s t 35 kg/(rrFh)
.~ 50-
8
I [ I:' / I/ , [ / / 0 0
O., =Qsota,.+Qa(1-1--'~-Q.a,, L rl,1
/ ,/-/-
;
...... north 70 kg/(m=h) ...... north 35kg/(rrF.)
~
~
;0
1'2
;4
1'6
18
channel width [mm]
Fig. 12: Yearly net energy gains for south, east and north oriented single glazed solar air preheaters with smooth absorber in dependence of channel width for two specific air mass flow rates. Optimum channel widths for each case are indicated.
7.2 Absorber with continuous fins The optimization of an absorber with continuous fins is more complicated as for a smooth absorber, as three geometrical properties can be altered simultaneously. Therefore an numerical optimization method, the mutation-selection method, given by Kinnebrock (1994), was used. All results refer to a south facing single glazed collector. Fig. 13 shows the contributions to the net energy gain for the finned absorber with 0.1 mm thick aluminum fins and a gap width of 30 mm. In comparison to Fig. 11, there is a small additional energy loss caused by the energy content of the fins. The yearly net energy gain for the finned absorber is about 250 kWh/ms and therefore about 30 % higher than for the smooth absorber. Fig. 14 shows the optimized values for the channel with constant fin thickness of 0.1 mm and the optimized values for the fin thickness with constant channel width of 30 mrn.
ISES Solar World Congress 1999, Volume III
300
~'
.~ 200-
solahr ~
Fig. 16 shows the yearly net energy gain for absorbers with continuous fins for different optimization constraints, given in the legend.
~et --'-~~
100-,\/
"-~r' -r
69
ma,~mum energygain 240.r
/ " " - . . h ~ t gain " through fan r "---2_ -
= I= 9
~'~ 210r m ~ 180-
0-
tO
~ fin s p a c i n g [mm]
60-
I
I
I
I
- o - optimized channel width (fin thickness O.1 mm)
/
......
OI
o
2'0
-0.12
~
o_,
6'0
~
--.1~
do
~o
e
=ll W
-5 ,-., -4 -3 -2 -1 0
specific m a s s f l o w rate [kg/(m:h)]
I
-0.10 r
E 50" =.d Jr -o 403= Q re- 30ra ,J= o 20-
Fig. 16: Yearly net energy gains on let~ axis and masses of the used fins on the right axis in dependence on the specific air mass flow rate for absorbers with continuous fins.
Pi,
-0.08 o~ o -0.06 =,,,-! 3 -0.04 .3.
10-
~
150--
L_ m 120--
Fig. 13: Contributions to the net energy gain for a finned absorber versus fin spacing. 70-
--o--fin thickness oplimized, channel width 30 mm --~--fin thickness O.1 mm, channel width 30 mm
L.. @
energy pumping "" "'"-"" : " -. .".-. -. ". ~ ~ ' ~ " ~ fin gap continuous mass thickness: width: flow rate: . fins 30 . . . 0.1 .mm 70ram -kg/( - - -m=h : -) "~ o~>, "- -100 t .......................................................... of energy content fins/ / . / / ""primary
l m
--o--fin thickness O. 1 mm, channel width optimized
--or-optimized fin thickness (channel width 30 mm)
2'0
4'0
8'o
do
-0.02
~6o o,oo
specific mass flow rate [kg/(m:h)]
Fig. 14: Optimized values for channel width and for fin thickness in dependence of specific mass flow rate. In both cases the fin spacing was not constrained in the optimization procedure. With increasing specific mass flow rate both the channel width and the fin thickness increase. In Fig. 15 the optimized fin spacing is given. The fin spacing for the higher flow rates shows only a small dependence on the two parameters, channel width and fin thickness.
Without fin spacing, fin thickness and channel width constraints we get the highest yearly net energy gain. This is not shown on the graph. If only the fin thickness is held constant, the gap width increases to large values as given in Fig. 14. In this case the heat transfer equations for the rectangular continuous fins are no longer valid and the equations for the smooth absorber are taken to compute the heat gain. This is the reason for the strong increase in energy gain and fin mass for flow rates over 60 kg/(m2h) in Fig. 16. Practical limitations such as constraining the channel width to the maximum value of 30 mm and the fin thickness to the minimum value of 0.1 mm do not lower the yearly net energy gain significantly. In this case an optimal fin spacing for a specific air mass flow rate of 70 kg/(m'h) would be about 6 mm.
Z 3 Absorber with offset stn'pfins Fig. 17 shows the energy gain for various offset strip fins in comparison to continuous fins. The specific air mass flow rate is 70 kg/(m2h). 250/
=
,
=
,
=
/
IllB1
,,o] |oo]
"=
,_..,
I= "64r
tn_ 111 rr
[
210
2--D-fin thickness O.1 mm, channel width 30 mm - o - fin thickness 0.1 mm, channel width optimized --o--fin thickness optimized, channel width 30 mm
1 o
o
2'0
io
6'0
8'o
specificmassflowrate[kg/ln~hl]
1;o
Fig. 15: Optimized fin spacing in dependence of specific mass flow rate, with fin thickness and channel width as parameters.
~'~ 190],
.... ~ =
>' 180 0T
- - - 100 mm long offset fins ~ continuous fins T
o
~
;o
1's
2'0
long offset fins
2'5
30
fin s p a c i n g [mm]
Fig. 17: Yearly net energy gain in dependence on fin spacing for continuous and offset strip fins.
ISES Solar World Congress 1999, Volume III
70
The offset strip fins do not perform better than optimally spaced continuous fins. However, if the fin spacing is restricted to non-optimal values, offset strip fins will give better results as continuous fins. Fig. 18 explains why the performance of continuous fins is superior to that of offset strip firm. For a gap width of 30 mm and a fin thickness of 0.1 mm curves for the heat transfer coefficient based on the absorber area in dependence on the pressure drop in the air flow gap are depicted. Since in the interesting region (h > 100 W/(m2K)) the continuous fins give the highest heat transfer for any pressure drop, their performance is superior to all the offset strip fins. 300
i
~'
i
i
i
i
i
i
i
i
optimum for preheating ventilation air (fin spacing 5.9 mm)
250-
--'"
To show the efficiency potential of thin, closely spaced continuous fins for solar air heater applications Fig. 20 was derived. The two selected solar air heaters differ only by the geometry of their fins. One has 1 mm thick fins, spaced 27 mm apart. This (standard) fin geometry is found in a commercially available solar air heater. The other has 0.1 mm thick fins, spaced 5.9 mm apart (optimized geometry). The absorber is 1 m wide and 2.5 m long and has the optical and radiative performance of the selective coating "Black Crystal IF' ( a = 0.937, e = 0.065), see (Brunold 1999). The ambient temperature is 20 ~ the air inlet temperature is 80 ~ the specific air mass flow rate is 72 kg/(m2h) and the channel width is 28 mm. The back wall is made of a 60 mm thick insulation with a thermal conductivity of 0.O40 W/(mK). I
I
I
80.t
/Cf-'~
:
gB
q = 0.77 - 2.65 W/(m=K) - AT / G
~40-
~ continuous fins - - - offset strip fins - - - - smoolh absorber
=9
I
optimized geometry standard geometry
.5r~ 60-
50-
I
o o
0
o
o
~
1'0
1'5
2'o
2'5
3'o
3'5
,~
4'5
50
20-
q = 0.69 - 2.37 W/(rrFK) - AT / G
" " D ~ ' ~ ~
pressure drop [Pa]
Fig. 19: Heat transfer coefficient (based on the absorber area) in dependence on the pressure drop for a smooth absorber and for absorbers with continuous and offset strip fins. The specific air mass flow rate is 70 kg/(m2h). 8 CONCLUSIONS In this paper an optimization method for solar air heaters with flow behind the absorber plate has been introduced. It maximizes the net energy output of the system and takes the long term solar heat gain, the electrical pumping energy and the energy for manufacturing the fins into account. The method was applied to solar ventilation air preheating collectors for which important conclusions can be derived: a) Continuous fins provide the highest net energy gain if they are spaced close to each other. The optimal distance between the fins is about 5 to 10 mm. In the case of a highperformance collector nmning at much higher average insolation values the optimal spacing is generally smaller. b) Due to higher pressure losses offset strip fins show reduced net energy gains compared to optimally spaced continuous fins. However, they show good results generally for large fin spacings. c) The optimum flow regime is laminar, accompanied with low Nusselt numbers and large heat exchange areas. d) In contrast to the second law optimization which considers exergy instead of energy (Altfeld 1985), the obtained fin spacing in this work is much smaller and predicts higher thermal heat gains. To our knowledge there is no commercial solar air heater which incorporates the optimized geometry. They all seem to have notideal fin spacings. On the other hand most of the water to air heat exchangers used for air-condition purposes use thin and closely spaced fins, as recommended in our study.
0
o.oo
o.b~
o.~o (Tin -
o.:~
o.~,o
o.~,5
.-.
0.30
T : m b ) / G [Km=/W]
Fig. 20: Thermal efficiencies of two solar air heaters with different fin geometries for an ambient temperature of 20 ~ and an air inlet temperature of 80 ~ The solar air heater with the optimized fin geometry is about 13 % more efficient as the heater with standard geometry and, not seen in this graph, it has a much shorter thermal response time. Therefore, apart from the higher stationary efficiency, the collector utilizes short time insolation even better. The optimal continuous fin geometries derived in this paper are based on empirical equations and have not yet been validated. Therefore it is necessary to build and test collectors with the proposed optimal geometry. Because the pressure drop increases fast with decreasing fin spacing, the optimal fin spacing for real solar air heaters may be somewhat larger than calculated. ACKNOWLEDGEMENTS This work was supported by the Bavarian Research Foundation (BFS), Munich, within the project "SOLEG" which deals with solar assisted energy supply of buildings. We thank our industrial partners GlasKeil/Wfirzburg, Gebrtider Schneider/Stimpfach and Grammer Solar-Luft-Technik/Amberg, all in Germany as well as SIT EuropeNienna, Austria. REFERENCES
Altfeld K. (1985). Exergetische Optimierung flacher solarer Lufierhirzer. VDI-Fortschrittsberichte. Series 6, No. 175. VDIVerlag, Dfisseldorf.
ISES Solar World Congress 1999, Volume III
Beavers G. S., Sparrow E. M., Lloyd J. R. (1971). Low Reynolds Number in Round Pipes and Infinite Channels and Heat Transfer in Transition Regions. J. Basic Eng. 93,296-299. Bhatti M. S., Shah R. K. (1987). Turbulent and Transition Flow Convective Heat Transfer in Ducts. In Handbook of Single Phase Convective Heat Transfer., Kakac S., Shah R. K., Aung W. (eds), Wiley-Interscience, New York. Brunold S. (1999) Qualification Tests of Thermafin Manufacturing, LLC (TML) "'Black Crystal 2 ""Solar Collector Absorber Coating with Respect to Thermal Stability and Resistance to Humidity Involving Condensation. Report, SPF, Hochschule Rapperswil. 5.5.1999. Corradini R. (1997) Ganzheitliche Bilanzierung von Metallen. Thesis, Lehrstuhl ftir Energiewirtschatt und Kraftwerkstechnik, TU-Miinchen, Germany. Duffle J. A., Beckman W. A. (1991) Solar Engineering of Thermal Processes. 2~dedn. Wiley-Interscience, New York. Blomberg T. (1996) Heat Conduct'on in Two and Three Dimensions. Report TVBH-1008, Lund University, Lund, Sweden. Diab M. R. (1981) Experimental and Analytical Study of Heat Transfer Characteristics of Solar Air Heater Incorporating a Finned Absorber. PhD-Thesis, Purdue University, West Lafayette, Indiana. Ebadian M. A., Dong Z. F. (1998) Forced Convection, Internal Flow in Ducts,. In Handbook of Heat Transfer. Rohsenow W. M., Hartnett J. P., Cho Y. I. (eds). 3ra edn. MacGraw-Hill, New York. FIE (1998) Die Bereitstellung von elektn'scher Energie in Deutschland (1996). Forschungsstelle ftir Energiewirtschaft, Am Bliitenanger 71, D-80995 Miinchen, Germany. Gnielinski V. (1976) New Equations for Heat and Mass Transfer in Turbulent Pipe and Channel Flow. Int. Chem. Eng. 16, 359-368. Gnielinski V. (1995). Forsch. im Ing.-Wes. 61, No 9, 240-248 Heaton H. S., Reynolds W. C., Kays W. M. (1964). Heat transfer in annular passages. Simultaneous development of velocity and temperature fields in laminar flow. Int. J. Heat Mass Transfer 7, 763-781. Heibel B., Hauser, G. (1996). Durchstr~mte Vorhangfassaden zur gorwiirmung der Zuluft mechanischer Liiftungsanlagen. AbschluBbericht, DFG-Forschungsvorhaben HA 1896/1. Mai 1996. Universit[it Gesamthochschule Kassel Jones O. C. (1976). An Improvement in the Calculation of Turbulent Friction in Rectangular Ducts. J. Fluid Eng. 98, 173181. Kabeel E., Mec~'ik K. (1998). Shape optimization for absorber plates of solar air collectors. Renewable Energy 13, 121-131. Kreith F., Bohn M. S. (1986)Principles of Heat Transfer. 4th edn. Harper & Row, New York. Kutscher C. F., Christensen C. B. (1992). Unglazed Transpired Solar Collectors. In Advances in Solar Energy, An Annual Review of Research and Development. Boer K. W. (ed). Vol 7 pp. 283-307. ASES. Kinnebrock W. (1994). Optimierung mit genetischen und selelm'ven Algorithmen. R. Oldenbourg Verlag, Miinchen. Kuzay T. M., Malik M. A. S., B6er K. W. (1975). Solar Collectors of Solar One. In Proceedings of the Workshop on Solar
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Collectors for Heating and Cooling of Buildings. Sargent S. L. (ed). May 1975, 99-108. Maryland University, College Park, New York City, Manglik R. M., Bergles A. E. (1995). Heat Transfer and Pressure Drop Correlations for the Rectangular Offset Strip Fin Compact Heat Exchanger. Exp. Therm. Fluid Sci. 10, 171-180 Mattox D. L. (1979) Evaluation of Heat Transfer Enhancement in Air-Heating Collectors. DOE-No. ALO-5352-T1, Northrop Services, Inc., Huntsville, Alabama. Merker G. P. (1987). Konvelm've Warmeiibertragung. Springer, Berlin, Heidelberg, New York. Mills A. F. (1962). Experimental Investigation of Turbulent Heat Transfer in the Entrance Region of a Circular Conduit. J. Mech. Eng. Sci. 4, 63-77. Morhenne J., Fiebig M., Barthel H. (1990). Entwicklung und Erprobung einer Baureihe von optimierten, modularen Solarlufterhitzern riD, Heizung und Trocknung. BMFT-Report-No. 0335003E6, Ruhr-Universit~it, Bochum. Piao Y., Hauptmann E. G., Iqbal M. (1994). Forced Convective Heat Transfer in Cross-Corrugated Solar Air Heaters. J. Sol. Energy Eng. 116, 212-214. Pottier K., Beck, A., Benz N. (1996). TestreferenzjahrUnstimmigkeiten in der Globalstrahlung. Sormenenergie 4/96, 22-23. Pottier K., Beck A., Fricke J., (1998a). Solarfassade zur Frischlutt-VorwErmung. In Proceedings of 11. Internationales Sonnenforum. 26.-30.07.1998.510-517. DGS, K61n, Germany. Pottier K., Beck A., Fricke J. (1998b). Dynamische Simulation und Optimierung einer Solarfassade. In Proceedings of 11. Internationales Sonnenforum. 26.-30.07.1998. 791-798. DGS, K61n, Germany. Pottier K., Sippel C. M., Beck A., Fricke J. (1999). Heat transfer and pressure drop correlations for offset strip fins usable for solar air heating collectors. In Proceedings of 15th European Conference of Thermophysical Properties (ECTP). 5.-9. September 1999, Wiirzburg, Germany. In Press. Prasad B. N., Saini J. S. (1988). Effect of arnficial Roughness on Heat Transfer and Friction Factor in a Solar Air Heater. Solar Energy 41,555-560. Shah R. K. (1978). A correlation for laminar hydrodynamic entry length solutions for circular and noncircular ducts. J. F1. Eng. 100, 177-179. Shah R. K., London A. L. (1978) Laminar flow forced convection in ducts. Academic Press, New York. Unsworth M. J., Montheith J. L. (1975). Long-wave radiation at the ground ~+II). Quart. J. R. Met. Soc. 101, 13-34. VDI (1997). VDI-Wiirmeatlas: Berechnungsbliitter fiir den Wiirmeiibergang. 8. edn. Springer, Berlin, Heidelberg. Weinl~ider H., Pottier K., Beck A., Fricke J. (1999). Angular dependent measurements of the thermal radiation of the sky. In Proceedings of 15th European Conference of Thermophysical Properties (ECTP). 5.-9. September 1999, Wiirzburg, Germany. In Press.
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INCLINATION DEPENDENCY OF FLAT PLATE COLLECTOR HEAT LOSSES Bernd Bartelsen, Markus Kiermasch, Gunter Rockendorf Institut fiir Solarenergieforschung GmbH, Hameln/Emmerthal (ISFH), Am Ohrberg 1, D- 31860 Emmerthal, Germany, Tel. +49 5151/999-522, Fax +49 5151/999-500, e-mail [email protected] Abstract - For flat plate collectors the natural convection in the air gap between absorber and transparent cover is of major importance regarding the collector heat losses. The collector inclination angle affects the natural convection phenomenon and thus influences the collector heat loss coefficient. This has been investigated by experiments on four single glazed selective flat plate collectors, including an additional variation of the ambient air speed. For these collectors a reduction of the effective collector heat loss coefficient of around 0.1 W/m2K per 15~ increase of the inclination angle has been found in the angular range between 15 ~ and 90 ~ As a practical conclusion, this effect has to be taken into account during collector tests carried out on tracking devices. Furthermore, the higher efficiency at 90 ~ inclination angle is an advantage for fafade collectors.
1. I N T R O D U C T I O N For typical selective flat plate collectors the heat transfer by convection in the air gap between absorber and transparent cover is the dominant part of the collector heat losses. The effect of natural convection in inclined rectangular enclosures has been studied by different groups. A well known correlation describing the free convective heat transfer across an inclined air layer of large aspect ratio is given by the following equation (1) (Hollands et al., 1976). 1708 "]*. [1 _ sin(1.8 9T)1"6 91708] Nu = 1 + 1.44-[1 - R a - c"osyA Ra: co's~, +{FRa-eos~'] 1/3 }* / 5830 J - 1 The brackets with asterisk stand for:
(1) [~ *=(IXI + X ) / 2
Equation (1) quantifies the heat transfer by natural convection between two parallel plates, where the inferior one is heated. It is based on measurements with an isothermal temperature distribution on the heated plate and an even absorber surface. This equation is given a high accuracy in a range from 0 ~ (horizontal) to 60 ~ an extension up to 75 ~ is possible with a higher uncertainty. Experimental investigations carried out at the ISFH (Institut fiir Solarenergieforschung GmbH, Hameln/Emmerthal) in 1993 (Bartelsen et al., 1993) showed, that in a real flat plate collector, with a distinct temperature profile and an unevenness of up to 10 mm, the convective heat transfer increases significantly if compared to ideal surfaces. In particular these investigations on a test collector with variable gap size led to the following results: 9 higher heat transfer coefficient for real flat plate collectors than equation (1) gives, 9 only small reduction of the heat transfer coefficient with increasing gap sizes and no maximum resp. minimum at small gap distances, 9 similar curves of the heat transfer coefficient at different inclination angles (30 ~ - 60~ 9 no influence of the fluid mass flow rate and the orientation (horizontal or vertical) found, 9 higher heat transfer coefficient for an inverted flow direction with vertical orientation and 9 lower heat transfer coefficient at low (or zero) irradiance levels.
In numerous efficiency measurements on commercial collectors, carried out in the ISFH solar simulator test facility, the experience was made, that absorbers with even surfaces made out of one metal sheet in most cases lead to smaller heat loss coefficients than absorbers out of single stripes, especially if the stripe absorbers show a higher unevenness. Collectors with thick absorber sheets often showed smaller heat loss coefficients if compared to collectors with thin absorbers. These effects underline the main difference between equation (1) and real collectors: real absorbers in solar collectors always have a well marked temperature distribution, especially in vertical direction to the fluid pipe, and, if they are made out of stripes, they do not represent an even plate. Within the investigation discussed before the influence of the collector inclination on the heat loss coefficient has been considered only for a small angular variation (between 30 ~ to 60~ Especially the question of vertical collectors (90 ~) has not been considered there. With the increasing interest in fafade collectors, the question about the collector heat losses in vertical position arises. 2. E X P E R I M E N T A L INVESTIGATIONS In the following, the effect of the collector inclination on the natural convection in the air gap and thus on the collector performanee will be discussed by the presentation of experimentally determined heat loss coefficients for different flat plate collectors as a function of the inclination angle. For the measurements, two versions of a flat plate collector prototype with typical absorber and glass construction, one with normal opaque insulation, the second with an unusual high thickness of back and side insulation have been produced and investigated. The normally insulated prototype (A) has a thermal insulation of 5 cm thickness (back) resp. 2 em (side) and the highly insulated one (prototype B) a mineral wool thickness of 25 em resp. 5 cm. The insulation thickness of the second prototype is a typical value for the integration of collectors into the facade of advanced low energy houses, which is performed without thermal decoupling by a ventilation layer. In addition to these investigations, one flat plate collectors of commercial production has been measured at different inclination angles between 0 ~ and 90 ~
ISES Solar World Congress 1999, Volume Ill
The tests have been performed in agreement with ISO 9806-1 with highly reproducible test conditions in the solar simulator test facility. The high reproducibility of the test facility is a necessary condition, as only small efficiency differences have to be identified. The collector parameters are referring to mean temperature of the collector heat transfer fluid Tm and the aperture as reference area. The measurements were carried out at an irradiance level of about 820 W/m 2, an ambient air temperature of around 22 ~ and an air speed of 3 m/s. Some efficiency curves have in addition been recorded without forced convection of the ambient air (air speed below 0.5 m/s). The results of the collector tests with the highly insulated prototype are shown in table 1. prototype B
1"10[-]
a 1 [W/Kmz] a2 tw/K~n~]
0 ~ inclined
0.764
3.86
0.013
15~ inclined
0.765
3.85
0.013
30 ~ inclined
0.767
3.75
0.013
45 ~ inclined
0.769
3.62
0.013
60 ~ inclined
0.770
3.52
0.013
90 ~ inclined
0.773
3.35
0.0i3
Tab. 1. Efficiency parameters for the highly insulated collector prototype B at various inclination angles Figure 1 displays graphically the efficiency curves of the highly insulated collector prototype B versus AT/G, where AT is the difference between the mean fluid temperature and the ambient air temperature Ta, with the inclination angle as parameter (numerical data from Tab. 1).
The test results of a fiat plate collector from series production with a selective absorber plate (single sheet, meandering pipe connection) are shown in table 2. commercial collector
110 [-]
,--., 0,7
t="
/3 0 ~ 15.
0,6
0,02
0,04
AT/G [Km'/W] Fig. 1: Efficiency curves of the highly insulated prototype at vari-
a 1 [W/Kmz] a2 [w/Ir
0 ~ inclined
0.778
3.86
0.011
45 ~ inclined
0.780
3.65
0.011
90 ~ reclined
0.783
3.28
0.011
~]
Tab. 2: Efficiency parameters for a commercial collector at various inclination angles These results show, that the heat loss coefficient decreases with increasing inclination angles. This effect is caused by the convective heat transfer in the air gap between absorber and glass pane. From table 1 and table 2 it may be derived, that a reduction of the linear heat loss coefficient a 1 of about 10% at an inclination of 90 ~ results if compared to the parameters determined at an inclination angle of 45 ~ (standard test conditions according to different standards like prEN 12975-2 or DIN 4757-4) 1). At horizontal installations or small inclination angles the heat loss coefficient raises for about 5%. The temperature dependent heat loss coefficient a2 showed a far-reaching independence of the inclination angle. In order to make the changes of the flat plate collector heat losses clear, figure 2 presents the measured effective heat loss coefficient as a function of the collector inclination angle. The effective heat loss coefficient is calculated for a temperature difference of 40 K between the ambient air temperature and the mean temperature of the heat transfer fluid Uloss,40K
0,8
73
=
a 1 + a 2 940K
(2)
For the presentation of figure 2, results from different collector tests have been taken into account: 9 standard flat plate collector from series production (tab.2) 9 normally insulated collector prototype A 9 normally insulated collector prototype A, measured with an ambient air speed below 0,5 m/s (free convection) 9 highly insulated collector prototype B (tab. 1) 9 highly insulated collector prototype B, measured with an ambient air speed below 0,5 m/s (free convection) 9 test collector with variable gap size from former investigations (1993) Figure 2 shows, that the effective heat loss coefficient of all collectors decreases with increasing inclination angles in a similar way, independently of the construction and the air speed conditions above the transparent cover. This can easily be recognized by the parallel shape of the curves.
ous inclination angles (irradiance level 800 W/m 2, air speed 3 m/s) Figure 1 shows, that the collector efficiency increases with increasing inclination angles. This leads especially at high temperatures to a significantly higher efficiency for a vertical collector if compared to a horizontally mounted collector. Due to the reduction of the heat loss coefficient, the collector efficiency factor and thus the conversion factor 110 of the collector increases slightly.
1.The standard ISO 9806-1 requires that the collector shall be mounted at an inclination angle equal to the latitude (+ 5% but not less than 30~
ISES Solar World Congress 1999, Volume III
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3. PRACTICAL RELEVANCE 4,75
I
3
r
n
prototype A
2,75
I
~
A
prototype B
prototype B
rut ~lm:lDr H
Scmmu~on~,econvect~n =Scrnmu=aar~freeoo.voceon re=r=19931 I -_. o r 0 r II
: ,
I
0
=
15
1
~
30
I
=
45
I
~
60
I 75
,
I ' I 90
Collector inclination angle [~
Fig. 2: Effective heat loss coefficient of different collectors at a temperature difference of 40 K for various inclination angles. The effect is very small below 15~ for angles between 15~ and 90 ~ it is significantly higher, whereas an approximately constant gradient may be stated. As a rule of thumb, it can be derived that the effective heat loss coefficient will be reduced by about 0.1 W/m2K (i-0.02) per each 15~ tilt angle increase in the tilt range between 15~ to 90 ~ For the different collectors presented in figure 2, simulation calculations with a theoretical collector model have been performed. In this model the natural convection in the air gap is described with the correlation from Hollands, equation (1). 4,75 4,5
J3,75
pnmtypeB 3,5
3,~
9
"
-~-
~ i cm i n ~ a ~ n
lamuhllJol~
I
0
prmotypeB
~ cm i n s u l i i o n
15
,
I
30
,
I
45
,
,
60
75
90
Collector inclination angle [*]
Fig. 3" Effective heat loss coefficient of the collector prototype B at a temperature difference of 40 K versus inclination angle, in comparison to model calculations. The results for an inclination angle below 60 ~ show a good correspondence of the curve shape. For an angle between 60 ~ and 75 ~ the simulation results indicate smaller heat loss coefficients of the calculated as compared to the measured coefficients. It is reminded, that the best accuracy of equation (1) has been specified in an angular range up to 60 ~.
The practical effect of the inclination dependency of flat plate collector heat losses will be discussed with special regard to the calculation of the yearly energy output of collector systems using simulation programmes and the measurement of collector efficiency parameters on outdoor tracking devices.
3.1 Energy output of collector systems Collector parameters, derived from efficiency measurements according with test standard ISO 9806-1 or other standards are normally valid for an inclination angle of 45 ~ The 45~ are used in general within simulation programmes to calculate the thermal collector output of solar systems and the solar fraction of the heat demand. For the following discussion, the collector parameters are used to determine the annual output of a typical domestic hot water system in Germany. If a collector is installed horizontally or with a small angle to the ground (0 ~ - 15~ the application of the efficiency parameters identified with the relevant inclination angle reduces the energy output of the collector by nearly 2% if compared to simulation results with collector parameters which have been identified at 45 ~. A vertically installed collector has a lower heat loss coefficient and therefore the energy output will increase by up to 4% if compared to the 45 ~ parameters. If the integration into a facade with thermal coupling (no backside ventilation) is considered, the higher rear insulation of the collector will additionally reduce the heat loss coefficient by about 0.3 W/m'K. This difference has been determined from the efficiency curves of prototype A and B. If this improvement is taken into account in addition to the influence of the vertical mounting, the energy output will increase by about 8%. Further positive effects of the fagade position like the lower wind velocity above the outer collector surface cannot be quantified and are therefore not considered in these discussion. With regard to the accuracy of collector efficiency measurements and comparability between different test laboratories, the parameters identified at an inclination angle of 45 ~ are sufficient for the calculation the annual energy output of roof mounted collectors. For vertical facade collectors the lower heat loss coefficient should be taken into account. 3.2 Discussion of collector tests on tracking devices For collector tests carried out on tracking devices the variable inclination angle during the tests will influence the measured efficiency parameters. The influence on the conversion factor % in the inclination range between 30 ~ and 70 ~ for typical fiat plate collectors may be ignored (less than 0.5 percentage points). On the other hand the heat loss coefficient can vary significantly. Depending on the test procedure of the performance measurements, the combination of collector temperature and inclination angle may lead to heat losses at a high collector temperature (80 ~ that differs by about ~- 6% from the value at constant 45 ~ In the regression analysis of the measured data this may lead to different impacts onto the two collector heat loss parameters a 1 and a2, thus creating a unrealistic modification of the curves shape. To avoid this error the efficiency analysis should be carried out only with data coming from a limited range of inclination angles.
ISES Solar World Congress 1999, Volume III
As an alternative a selection of the measured data with a balanced proportion of different inclination angles below and above 45 ~ especially for high temperatures is also possible to attain a suitable analysis of the performance measurements. If one of these recommendations is taken into account, the inclination dependency of flat plate collector heat losses has no critical influence on the results of collector tests carried out on tracking devices. 4. CONCLUSION The collector inclination angle influences the heat transfer by natural convection in the air gap between absorber and the single glass cover and therefore the collector heat loss coefficient is affected. For an inclination angle range between 15 ~ and 90 ~ a reduction of the effective heat loss coefficient by about 0.1 W/m2K per 15 ~ inclination angle increase could be identified for each set of collector efficiency curves, where different constructions and ambient air speed conditions have been investigated. This behaviour may also be found by simulations using the equation of Hollands et al., if only an angular range of up to 60 ~ is regarded. The consequences of the inclination dependency for flat plate collector heat losses are of minor importance for the calculation of the energy output of roof mounted collector systems as well as for the analysis of collector efficiency tests on tracking devices. But for vertically mounted or facade integrated flat plate collectors the benefit of the lower loss coefficient should be considered. NOMENCI~TURE a1
linear collector heat loss coefficient, referred to T m (W/m2K) a2 temperature dependent collector heat loss coefficient, referred to T m (W/m2K Y inclination angle, between collector surface and ground AT temperature difference between mean fluid temperature and ambient air temperature (K) G Solar irradiance (W/m 2) rl collector thermal efficiency, referred to Tm (-) rl0 conversion factor (11 at AT = 0), referred to T m (-) Nu Nusselt number (-) Ra Raleigh number (-) Tm mean temperature of heat transfer fluid (~ UL effective heat loss coefficient of collector, referred to T m (W/m2K) UL,40K effect, heat loss coefficient (UL at AT = 40 K) (W/m2K) REFERENCES Hollands K.G.T, et al. (1976) Free Convective Heat Transfer Across Inclined Air Layer. ASME Journal of Heat Transfer, 98, pp. 189-193. Bartelsen B., Jard]en S., Rockendorf G. (1993) Heat Transfer by Natural Convection in the Air Gap of Flat Plate Collectors. In Proceedings of the ISES Solar World Congress, 23-27 August, Budapest, Hungary, pp. 267-272, Pergamon Press, New York.
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PV-HYBRID AND THERMO- ELECTRIC- COLLECTORS Gunter Rockendorf and Roland Sillmann Institut ftir Solarenergieforschung GmbH, Hameln/Emmerthal (ISFI-I), Am Ohrberg 1, D- 31860 Emmerthal, Germany, Tel. +49 5151/999-521, Fax +49 5151/999-500, e-mail [email protected]
Lars Podlowski and Bernd Litzenburger SolarWerk GmbH, Iserstr. 8-10, D- 14153 Teltow, Germany, Tel. +49 3328/448-300, Fax +49 3328/448-301 Abstract Two different principles of thermoelectric cogeneration solar collectors have been realized and investigated. Concerning the first principle, the thermoelectric collector (TEC) delivers electricity indirectly by first producing heat and subsequently generating electricity by means of a thermoelectric generator. The second principle, the photovoltaic-hybrid collector (PVHC) uses photovoltaic cells, which are cooled by a liquid heat transfer medium. The characteristics of both collector types are described. Simulation modules have been developed and implemented in TRNSYS, in order to simulate the behaviour of typical domestic hot water systems. The discussion of the results shows, that the electric output of the PVhybrid-collector is significantly higher than that of the thermoelectric collector.
1. INTRODUCTION The aim of thermoelectric- hybrid- solar collectors is to cogenerate thermal and electric energy within the same module. In cooperation with the company SolarWerk, Teltow (Germany) two different types of thermoelectric- hybrid- collectors have been constructed and investigated at ISFH according to their corresponding physical principles. The first type is called thermoelectric collector (TEC). The principle is to combine a solar thermal collector with a thermoelectric generator (TEG), located between absorber and fluid pipe of the collector, delivering the electric energy. The second collector type is the photovoltaic-hybrid collector (PVHC). The idea of this collector is the combination of photovoltaic (PV) cells with a thermal collector. The PV cells are laminated on the surface of the solar absorber, which is cooled by a liquid heat transfer medium. For both collector types basic work has been carried out. Collector prototypes have been constructed and evaluated by experiments. Furthermore, mathematical collector models has been developed and validated, and thus system simulations could be carried out. Both collector constructions and the main results of the investigations will be described. 2. T H E R M O E L F . ~ C
TEG to its cold junction. This local concentration of heat may be obtained by a gravity assisted heat transfer processes like e.g. boiling-condensing process in heat pipes or thermosyphon cycles. For the following development a water filled heat pipe has been applied as appropriate solution. To generate a high amount of electric power, a high temperature difference at the TEG is necessary. This can only be achieved by a high thermal resistance of the TEG, which consequently leads to a high absorber temperature, if a significant amount of heat will be led over the TEG. The high absorber temperature however increases the thermal losses of the absorber and therefore reduces the solar heat production in the collector part. This results in a reduction of the thermal and electric gains. Therefore it is necessary to use high temperature collectors, what at least requires evacuated tubular collectors (ETC), or even better ETC with concentrating mirrors. Figure 1 shows a principle solution using an ETC with heat pipe, which has been investigated here.
COLLECTOR
2.1 Design principles and selected construction The thermoelectric collector (TEC) combines a solar thermal collector with a thermoelectric generator (TEG). The TEG, which delivers the electric energy, is located between absorber and fluid pipe of the collector. Peltier elements, which are normally taken for cooling purposes, were specially designed for electricity generation in order to use them as TEG. The thermal resistance of the TEG causes a temperature difference, which is proportional to the heat flux from the absorber to the fluid. Furthermore, this temperature difference is proportional to the electric power. Thus, for a high electric performance, all solar thermal heat has to be conducted over the TEG. Therefore a clear separation between the absorber and the fluid part of the collector is necessary, in order to concentrate the solar heat to one point, the hot junction of the TEG, and then to let it pass over the
Fig. 1: Scheme of thermoelectric collector Using a dry coupling, the condenser of the heat pipe heats the lower part of the heat exchanger to a high temperature, which will act as the hot junction of the TEG. The upper part of the heat exchanger is cooled by the heat transfer fluid of the solar loop, thus acting as the cold junction. The TEG is arranged between the hot and the cold junction. The heat passing the "lEG causes an electric power, which in its maximum power point (MPP) is proportional the temperature difference between the hot and cold junctions. The amount of heat transferred over the "lEG may be directed to
ISES Solar World Congress 1999, Volume I!!
an application like hot water preparation. 2.2 Investigations on thermoelectric generators A central objective of the development work is to investigate the behaviour of appropriate TEGs with regard to their electric and thermal properties. For this purpose, a heat exchanger test stand has been built up and the behaviour of different TEGs (area around 9 cm 9-, thickness 3 - 5 mm, manufacturers Kunze and TECOM) has been measured for varying boundary conditions (Giebel, 1997). The interactions of the electric and thermal properties depend on various parameters taking into account the different physical effects (mainly Seebeck- and Peltier- effect). E.g., the electric output is a function of mean TEG temperature, temperature difference and inner electric resistance of the element and is therefore coupled with the thermal resistance between hot and cold junction, which has been found is mainly depending on the mean temperature and the electric current generated by the dement. Thus, both the thermal and the electrical characteristics are depending on each other. Further complication is caused by the dependency of the inner resistance of the mechanical pressure, to which the element is exposed by the heat exchanger package. Finally, practical considerations like the heat transfer resistance between the TEG and the heat exchanger resp. the bypass heat flow caused by the clamping mechanism have also to be taken into account. For theses reasons, only simplified dependencies could be developed. The correlations, however, have been proved to be sufficiently precise for the description of the whole collector. A deviation between measured and calculated results of less than +/ - 3% concerning the electric output has been found during the collector investigations. For the behaviour of the electric output, the following simplified formula has been worked out:
Pel =
Rload Rl~
(Ri +
2 2 9[b I - ATTEG2 + b 2 9ATTE G 9Tavg (1)
2 + b 3 9ATTEG2- Tavg ]
77
Three vacuum-tubes with heat pipe (producer Thermomax, UK) with 0.l m 2 absorber area each and water as heat pipe medium have been connected via the specially designed heat exchanger to a fluid circuit. Figure 2 shows the construction of one heat exchanger element.
Fig. 2: Construction of the heat exchanger of the thermoelectric collector-prototype Special care was necessary in order to avoid additional thermal resistances between the heat exchanger and the hot resp. cold junction of the TEG. Furthermore, as only the heat passing via the TEG produces electricity, any bypass heat flow has to be minimized. This is important for the design and the selection of the clamp device and the surrounding insulation material. Finally, the temperature stability of the applied material has to be high enough to withstand the expected high temperature, especially in case of stagnation. The prototype collector has been tested in agreement with ISO 9806-1. To assess the influence of the TEG-integration, a modified collector without TEG has been investigated, too. Due to the small amount of electric output, the thermal and electric yield may be discussed separately. Figure 3 shows the thermal efficiency curves of both collector prototypes. i
The heat transfer capability between hot and cold junction may be described by
i
I . . . .
-r
"-'0.
UTE G- C1 9I -t- C2 9Tavg + c 3 9I. Tavg + c 4 9I. T2avg (2)
. . . .
I . . . . .
1 . . . . .
I
I
7
i
-.L . . . .
_1
I
I
I
I
~-_._,,,,_,,,~.. /
+ UTEGO It has been shown during the TEG experiments and within the collector tests (section 2.3), that both equations describe the measured behaviour with a high accuracy. The measured performance of the TEGs is at 60 W input heat and 20 ~ fluid temperature between 1.3 and 2.0 W, i.e. an efficiency of around 2.3 to 3.2% has been achieved, while the TEG is operated in MPP. The thermal conductivity of one TEG is around 0.4 W/K, which causes an overtemperature at the hot junction of around 150 ~ (in MPP-operation, irradiance level approx. 900 W/m2). 2.3 Construction and assessment of the thermoelectric collector A prototype of a thermoelectric collector has been constructed.
! ,
I
A
I I
I I
-'41"-,
I
,
"" I,,,.
I
"
--O I
" I
I
,
I
,
T/G in K/(W/ma)
Fig. 3" Thermal efficiency curves of thermoelectric collector, compared to same collector without TEG, irradiance level approx. 800 W/m 2, air speed 3 m/s, referring to aperture area and mean fluid temperature The installation of the TEG with its high thermal resistance leads to a drastic decrease of the collector efficiency factor and thus reduces the conversion factor rl0 by around 45%, if compared
78
ISES Solar World Congress 1999, Volume III
to the identical collector without TEG. The electrical efficiency came up to a maximum value of 1.1% of the incoming solar radiation, which is around 2.8% of the transferred heat. The integration of a TEG rises the absorber temperature and by this way the losses of the solar collector are increasing significantly, whereas the electric output remains rather small.
if the improved TEC would be operated with constant fluid inlet temperature (10 ~ over one year. With these improved elements, annual simulations of typical solar domestic hot water systems have been carded out, where the heat transfer capability of the TEG has been varied. Figure 4 shows, how the variation of the conductivity affects the output of electric and thermal energy.
2.4 Simulation of thermoelectric collector and system For the calculation of yearly energy gains, a dynamic simulation model was developed. As the thermal and electrical properties may not be isolated, an iterative calculation process is necessary. The model has been validated with the experimental results of the prototype collector tests (Sillmann, 1997). It has been transferred to a TRNSYS simulation tool, and thus, it could be implemented in a solar system simulation programme.
I
-9
I-
i. . . .
r',
-: . . . .
I
I
-r
r
I
~ . . . . . . . .
I
t
....
r----]
-
*, . . . .
er
l=
,
!
The simulations lead to the following results: 9 The thermal connection between condenser and TEG and between TEG and fluid must be good. A minimum heat transfer capability between absorber and fluid except for the TEG itself of 20 W/m2K should be achieved for the used vacuum tube and heat exchanger, referred to the absorber area. The bypass heat flow should be minimized in order to come up to higher electric gains, whereas a higher bypass heat transfer increases the thermal output. 9 If the collector is operated during the whole year with a constant inlet temperature of 10 ~ (if irradiance is above 10 W/ m2), a thermal output of 660 kWh/m 2 and an electric output of 14 kWh/m 2 may be expected at Hanover (Germany). At 90 ~ inlet temperature, the output is 260 resp. 6 kWh/m 2 per year. 9 If a vacuum tube with a significantly lower loss coefficient as compared to the prototype (stagnation temperature around 70 K higher) would be used, the thermal output would nearly not be affected and the electric output would increase by around 5% at 10 ~ and 65% at 90~ fluid inlet temperature. That means, that a better insulated collector first of all promotes the electric gains. 9 An increase of the thermal conductivity of the TEG by a factor 2 would increase the thermal output to 750 kWh/(m2a) (constant fluid inlet temperature of 10 ~ and decrease the electric gain to approx. 50%. On the other hand, a reduction by 75% would lower the thermal output down to 320 kWh/(m2a), whereas the electric gain would rise to approx. 40 kWh/(m2a). The prototype of the TEC has shown the technology inherent disadvantage, that the high thermal performance of evacuated tubular collectors will be significantly decreased and the returned electric energy only comes up to small values. Therefore the conversion efficiency of the TEG has to be improved. The best laboratory elements attain effiency values, which are about 3 times higher than the used elements, they come up to around 30% of the Camot efficiency (Rowe et al., 1995), which is assumed to be at the upper technical limit. This means an efficiency increase of the TEG by a factor 3 if compared to those TEGs used in the prototypes. This theoretical TEG improvement would nearly not influence the thermal output of the improved thermoelectric collector 1), but it would enlarge the electric gain by a factor 2.5, 1. The improved TEC is a theoretical collector combination, in which the far better TEGs with a 3 times higher conversion efficiency have been applied together with the vacuum tubes of the prototype.
o
#
#
S# # # conductivity factor of the TEG
S#
Fig. 4: Annual energy yield of a thermoelectric collector with efficiency-improved TEG in solar domestic hot water system (Hanover) vs. conductivity factor of the TEG. The conductivity factor is a multiple of the measured heat transfer capability of the prototype (collector area 5 m 2, energy demand 2600 kWh/a). If 5 m 2 of the constructed prototype collector with the improxrA.TF,G would be installed in a typical hot water system, a yearly gain of 1650 kWh thermal and of 50 kWh electric energy could be achieved. Figure 4 shows, that a lower TEG conductivity increases the electric gain, but strongly reduces the thermal gain. If the conductivity is halved, the yearly electric output is nearly doubled to 100 kWh/m 2, whereas the thermal gain is reduced by around 500 kWh/(m2a) (i.e. 30%). The discussion shows, that a thermoelectric collector with acceptable technical properties needs high efficient vacuum collectors and high efficient TEGs, both at the upper technical limit. But even with these components, it seems to be unrealistic to develop a thermoelectric collector with economically promising prospects.
3. PHOTOVOLTAIC- HYBRID- COIJ,ECTOR
3.1 Design principles of a PV-Hybrid-Collector The idea of the photovoltaic hybrid collector (PVHC) is the combination of photovoltaic (PV) cells with a thermal flat plate collector. The PV cells are laminated onto the surface of the alum i n i m solar absorber. On its back side, a copper fluid pipe is clamped on. This PVH-absorber is cooled by a liquid heat transfer medium. The absorber is integrated in a standard ahminium frame, usual for solar thermal collectors with normal solar glass, an air gap of 3 cm and a backside insulation (mineral wool) of 5 era. Figure 5 shows the explosion drawing of the PVHC.
ISES Solar World Congress 1999, Volume III
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This model has been developed interactively with the experiments at the first prototypes. The development process has given, that it is sufficient to introduce only one thermal capacitance into the model; it is located in the fluid node. This inaccuracy is acceptable, as the thermal resistance between the fluid and cell node is rather low (this has not been the case at the TEC, compare chapter 2). The validation procedure of the f'mal simulation programme shows a good agreement with the measured data, also in the dynamic parts of the time series.
3.3 Performance of the PV-hybrid-collector
Fig. 5: Construction of the photovoltaic-hybrid collector (PVHC) Important for the effectiveness of the PVHC is a small thermal resistance between the PV cells and the fluid. Both the thermal and the electric performance decrease with rising fluid temperature. As the absorber temperature is strongly affected by the fluid temperature, the design and operation of the hot water system, e.g. the solar fraction, is very important, too.
The general thermal behaviour is similar to that of a nonselective flat plate collector. The reason is the high thermal emissivity of the laminate-cell-package (e around 0.90), which leads to a high thermal loss coefficient. The conversion factor rio of the PVH-collector is only somewhat smaller than that of a nonselective flat plate collector. The difference is caused by the absorption coefficient a, which comes up to 0.915. This is a few percent points lower than with a black solar colour. The electrical operation mode has to be taken into account. If the PV-part has no efficiency (open resp. short circuit), the heat generation will be higher if compared to maximum power point (MPP) operation.
3.2 Collector model In order to support the development work, a simulation model is necessary. The model is used for the component optimization and the simulation of a whole solar system. To use the universal properties of TRNSYS, the PVH-collector module was written as TRNSYS- type and implemented into a TRNSYS configurations of a solar domestic hot water system. The PVH-collector model has to describe both the electric and the thermal behaviour, as well as the interaction between these characteristics.
"*/ ) 0..
J o
A T/G in K/(W/m a)
Fig. 7: Thermal efficiency curves of the PVH-collector, with zero or with maximum electric efficiency, (according to ISO 9806-1, referring to mean fluid temperature and aperture area)
\
U
, !o+
+-
-r---!
II )f
I
The efficiency curves have been measured at an irradiance level of about 820 W/m 2 and at an ambient air speed of 3 m/s. Figure 7 shows, that the curve in electric MPP-operation is moved almost in parallel downwards if compared to the open circuit curve by around 0.10. That means, that the extraction of electric energy directly affects the zero loss efficiency rio, but nearly not the heat loss coefficient. Consequently, the stagnation temperature of a PVH-collector decreases if the electric part is operated in the MPP- point (see below). thermal coefficients
. J
Fig. 6: Thermal and electrical model for the PVHC Figure 6 shows the node model as basis for the thermal considerations, including the integrated Two-Diode-Model for the description of the electric performance of the PV-cells. The interaction between both models is taken into account by the cell temperature, which on the one hand is affecting the cell performance and on the other hand is almost equivalent to the absorber temperature and therefore directly influencing the thermal gain and loss mechanisms.
electrical coefficients (STC)
rio [-]
0,726 (PV-opencircuit) 0,633 (PV-MPP)
Isc [A]
2,84
a1 [W/m2K]
5,88 (PV- opencircuit) 5,64 (PV- MPP)
Uoc [V]
107
a2 0,016 (pv- opencircuit) PMPP [W/m2K2] 0,015 (PV-MPP) [W] riel [-1
220 0,103 (aperturearea) 0,121 (cellarea)
Tab. 1: Performance parameters of the PVH-collector, prototype
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I S E S S o l a r W o r l d C o n g r e s s 1999, V o l u m e III
The thermal parameters are determined according to ISO 98061 and they are referring to the mean fluid temperature and the aperture area (2.1 m2). The electric performance indicators are the aperture area and cell area related efficiency data, the electric power of the module (cell area 1.81 m2), the open circuit voltage and the short circuit current, all referred to standard test conditions (STC, irradiance 1000 W/m 2, cell temperature 25 ~ Table 1 directly shows, that the difference between the conversion factors rio in either MPP or open circuit is nearly equivalent to the electrical MPP-efficiency, what is a nice approval for the first law of thermodynamics in this special case. During the efficiency measurements with low inlet temperature, the mean cell temperature is only 11 K above the fluid temperature. For the implementation of the simulation programme, more detailed parameters than the ISO- coefficients and STC- parameters are required. For this purpose, the parameters of the Two-DiodeModel have been identified as well as the single resistances between the cell node and the fluid resp. the environment, as displayed in figure 6.
3.4 Simulation of the PVHC system yield As basis for the system simulations, a domestic hot water system with single drinking-water storage and internal heat exchanger has been placed at Wtirzburg, Germany, with an inclination angle of 35 ~, south. The annual heat demand is 2600 kWh/a, at a demand temperature of 45 ~ The collector area was enlarged from 1 module (2.1 m 2) in steps up to 5 modules. Figure 8 shows the results of these simulations. 5O0 4#5
•E
400
_= _= m
S00 " " " , ~ 2
3
combinations. 2 PVH & 1 SFP 2 PVH & 2 SFP 4 PVH & 1 SFP PVHth [kWh/m2]
291
232
193
SFPth [kWh/m2]
399
310
352
PVHel [kWh/m2]
90,8
88,5
87,0
SF
0,53
0,60
0,60
Tab. 2: Annual output of PVH and selective flat plate (SFP) collectors and solar fraction, for different module combinations (2.1 m 2 collector area for each module) Table 2 shows, that the combination of PVH-collector and selective flat plate collector leads to higher solar fraction values. The electric output of the PVH-collector is somewhat lower than that of standard PV-modules, which would come up to an annual yield of about 100 to 110 kWh/m 2. This difference is mainly caused by the higher reflection losses at the glass pane and a relatively low ratio of cell and aperture area. At the PVHC this ratio is 0.85, whereas around 0,90 for standard PV modules are typical. If the annual yield is related to the electric power (instead of module area), the electric output of the PVH-collector is nearly identical to that of standard PV-modules. The influence of the module temperature is rather small, which will be discussed in the foliowing. As the efficiency of PV-cells rises with decreasing temperatures, low module temperatures are desired. One of the main questions of PV-hybrid systems is, whether the mean operation temperature is higher if compared to standard PV-modules. For this purpose, the irradiation weighed mean cell temperature is defined as follows:
4#5
~ 4 ~
200
[-1
Tcell:
I(Ee-Tcell)dt/IEedt t
(3)
t
0
r0 lOO ,
~
I
1
,
I
,
I
,
I
2 3 4 number of collector modules
;
I
5
, I
6 P ##
I . . . .
~- . . . .
I
I
I
I
I
I
I
I
I
I
4- . . . . . I
Fig. 8: Annual thermal and electric energy gain versus the number of PVHC-modules, with additional data of the solar fraction of the hot water demand (SF) While the thermal gain decreases from 432 to 177 kWh/(m2a), the electric gain only decreases from 92 to 86 kWh/(m2a). The solar fraction of the thermal demand increases from 22 (1 module) to 50% with 5 modules. As thermal part of the collector shows the performance of a nonselective collector with additionally reduced thermal gains, the system output in central Europe is restricted to a solar fraction of around 50%. That means, that significantly higher solar fractions, corresponding to a 100% covering of the demand during summer may hardly be achieved. On the other hand, this collector should ideally be used in preheating systems, where both the thermal and the electric gains can benefit from the low fluid inlet temperature. In order to attain a higher solar fraction, a series installation of PVH-collector modules and standard selective flat plate collector modules is possible. Table 2 shows the annual output of different
~# I #
I
. . . .
i
i
I
I
I
I
I
I
1- . . . . I
E S
i
-I- . . . . .
I . . . .
I
I
-t-
l" E6
I
i
I -
-J
. . . . .
I. . . . .
I
I I I
. . . .
I-
. . . .
I ,
I
I. . . .
3-
. . . .
I
I
I
I
I
I
I
--t . . . . . . . . . .
I ,
I I
t. . . . .
I ,
/
4,.
I ,
,
6
I
,
#
number of collector modules
Fig. 9: Irradiation weighed mean cell temperature versus number of modules, in comparison to standard PV-modules Figure 9 shows, that the mean cell temperature is increasing with the number of modules, but even with 5 modules (10.5 m 2) it is still in the range of standard PV-modules, for which a wide variety exists, depending on wind exposure and integration technique. In a warmer climate like in southern Europe, the mean cell temperature of the PVHC could even be lower than that of standard PV-modules.
ISES Solar World Congress 1999, Volume III
3.5 Reliability questions and possible potential of improvement The PVHC comes up to a stagnation temperature of 147 ~ at 1000 W/m 2, 30 ~ ambient temperature and calm wind conditions, if operated in open circuit conditions. This is a typical value for nonselective collectors. If the module is operated at the same time in its electrical maximum power point, the stagnation temperature decreases by around 12 ~ Special regard must be given to the laminate construction, the electric cables and the connecting boxes, which all have to withstand these extreme temperatures. The collector prototype already shows a good performance, which has only little potential of improvement for this construction type. The thermal contact between absorber and fluid may still be improved insignificantly, by which the thermal gain may be increased by about 2 to 4%. New high efficiency cells could lead to an enhancement of the electric gain.
4. CONCLUSIONS The principle of the TEC is to produce first heat, and then to transfer this heat over the thermal resistance of the TEG, where it will partly be transformed into electricity, the remaining heat has to be cooled away. It follows from this serial energy flow, that the absorber must be maintained at a high temperature as the electrical generator needs a high temperature difference. Therefore, even with high efficiency collectors, the thermal efficiency will decrease significantly. The first requirement is to use solar collectors with very low loss coefficients, e.g. by concentrating the irradiance. The further disadvantage is the low conversion factor of TEGs, where only a value of 30% of the Camot efficiency seems to be realistic. In contrary to the TEC, the principle of the PVH-collector is the direct electricity production, i.e. the efficient direct use of the high exergy content of the radiation, and only the remaining radiation energy will be transformed to heat. This heat will be used on a temperature level as requested by the solar system. Hence, the PVH-collector produces heat and electricity in parallel. The comparison of the solar system simulations between the existing PVH- collector prototype and the advanced extrapolated TEC shows the advantage of the PVHC-principle. The improved TEC (5 m 2 evacuated tubular collector) would lead to an electricity gain of only 50 kWh/a and meet the thermal demand with a solar fraction of 53%. The PVHC (10.5 m 2) delivers around 920 kWh/a electric energy and covers the thermal demand by 50%. Regarding the same collector area of 5 m 2, the PVH-collector comes up to an electricity production of 450 kWh/a, what at any rate is 9 times higher than the electricity production of the advanced extrapolated TEC. Precise cost statements or estimations are not available. It may however be assumed, that the production technology of the PVHcollector with its known processes from PV and thermal flat plate collector technique has a higher economic potential. It therefore may be concluded, that the TEC will only be of interest for special applications. In a direct comparison the PVHcollector technology shows many advantages. However, hybridcollectors often show a non optimal behaviour in comparison to the parallel operation of the basic technologies. But for specific applications and special purposes, the advantages of only one type of solar module for heat and electricity production may be so convincing, that to our opinion these collectors will occupy a place in future developments and a future market.
81
NOMENCLATURE A a1
area (m2) constant collector heat loss coefficient, referred to Tm (W/m2K) a2 temperature dependent collector heat loss coefficient, referred to T m (W/m2K2) b 1,b2,b3 coefficients to calculate the electric energy gain of the TEG (different dimensions) Cl,C2,C3,C4 coefficients to calculate the heat transfer capability of the TEG (different dimensions) Cfluid capacity of the PVHC referred to the fluid temperature
0d/K) D1,D2 G I Ise J Jph K Pel PlvtPP Quse Ri Rload Rs Rsh Tamb Tavg Teen Tfluid Tm AT
diodes of the Two-Diode-Model global solar irradiance (W/m2) current (A) short circuit current (A) current (A) photo-current (A) incident angle modifier coefficient (-) electric power (W) electric power in MPP operation (W) useable heat-flux (W) inner resistance (fl) load resistance (f~) serial resistance of the Two-Diode-Model (f2) shunt resistance of the Two-Diode-Model (f~) temperature of the ambient air (K) average temperature of the TEG (K) temperature of the PV- cells (K) temperature of the heat transfer fluid (K) mean temperature of heat transfer fluid (K) temperature difference between mean fluid temperature and ambient air temperature (K) ATTEG temperature difference between the hot junction and the cold junction of the TEG (K) U voltage (V) Uoe open circuit voltage UAlu heat transfer capability of the aluminium absorber (W/m2K) Ueonv fluid heat transfer capability between fluid pipe and fluid " (W/m2K) UEVA heat transfer capability of the laminate (W/m2K) Uloss effective heat loss coefficient between the absorber and the ambient (W/m2K) UTEG heat transfer capability of the TEG (W/K) UTEG0 heat transfer capability of the TEG, constant part (W/K) O~ absorption coefficient (-) conversion factor; i.e. thermal collector efficiency at AT rl0 = 0, referred to Tm (-) electrical efficiency (-) riel incident angle (o) O
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ISES Solar World Congress 1999, Volume III
REFERENCES
Giebel, Ulfert; Untersuchungen an thermoelektrischen Elementen zur Stromerzeugung in Sonnenkollektoren; diploma thesis at Institut fttr Solarenergieforschung; Emmerthal; 1997 Sillmann, Roland; Konstruktion, meBtechnische Bewertung und Simulation eines thermoelektrischen Kollektors; diploma thesis at Institut ftir Solarenergieforschung; Emmerthal; 1997 Rowe, D.M. et al.; CRC Handbook of Thermoelectrics; CRC Press, Inc.; 1995 TRNSYS 14.1; A transient system simulation program; Solar Energy Laboratory University of Winconsin; Madison, USA; January 1994 Litzenburger, B.; Podlowski, L.; Rockendorf, G.; Sillmann, R.; Entwickhmg eines PV- Hybrid- Kollektors; in Proceexlings of the 8. Symposium Thermische Sonnenenergie; Ostbayrisches Technologie Transfer Institut e.V., Regensburg; 1998; pp. 77-82
ISES Solar World Congress 1999, Volume III
E L A S T O M E R - METAL- A B S O R B E R
83
- DEVELOPMENT AND APPLICATION
Bemd Bartelsen, Guntet Roekandorf Institut ftir Solarenergieforschung GmbH, Am Ohrberg 1, D-31860 Emmerthal, Germany, Tel. +49 (0)5151/999-522, Fax -500
Nortmrt I/ennemann Fachhochschule Osnabrtick, Albrechtstr. 30, D-49076 Osnabrtick, Germany, Tel. +49 (0)541/969-2940, Fax -2999
Rainer T e l l , Klaua Lorenz Solar Energy Research Centre - Dalarna University College, S-78188 Borl/inge, Sweden, Tel. +46 (0)23/778-703, Fax -701
Gottffi~t PurkarthoIer Arbeitsgemeinschaft Erneuerbare Energie, A-8200 Gleisdorf, Austria, Tel. +43 (0)3112/5886-16, Fax -18 Abstract - A new principle of a solar collector, that consists in appropriately shaped metal form plates as absorber and clipped in elastomer fluid pipes, the so called elastomer-metal-absorber, will be presented. The advantages are its freeze resistance, the seawater suitability and new possibilities for cost reducing collector installation and system techniques. The design parameters including a detailed analysis of the thermal resistance between absorber and fluid will be discussed, where special regard is given to the development of an appropriate elastomer material with high thermal conductivity as one of the key items. The first development steps have shown, that absorbers with a high thermal performance may be constructed. Finally, the idea to apply the principle of the elastomer-metalabsorber to metal roofs and faqades will be presented. This idea is followed up within a development project. 1. INTRODUCTION The idea of a combined absorber with a metal absorber sheet for the absorption of the solar radiation and a flexible elastomer fluid pipe for the transport of the solar heat has been developed.
Figure 1: elastomer-metal-absorber construction. As shown in figure 1, a round shaped clip profile is integrated into a metal plate, which has an absorption layer for solar thermal conversion. In this profile an elastomer tube for the heat removal is clipped in. The application of this elastomer-metal-absorber in solar thermal collectors offers the following potential advantages and essential possibilities: 9 Due to its inherent freeze resistance, operation without an antifreeze additive is possible. 9 System installation without heat exchanger in the solar loop may be discussed. 9 Operation with a corrosive fluid is possible, e.g. direct flow with sea or brackish water. 9 New and simplified techniques for the collector and system installation can be developed. The most promising application results from the new installation possibilities for the collector and the system. It is intended to integrate this new collector concept into roofs and faqades made out of metal form sheet elements. This idea will be presented in the following. Furthermore the elastomer-metal-absorber concept seems to be an attractive collector for the solar desalination of brackish and sea water, as the collector may be operated directly
with corrosive liquids without cost intensive corrosion protected heat exchangers. The desalination process should be designed to operate on a low temperature level (e.g. around 70 ~ The use of elastomer tubes in collectors requires appropriate absorber constructions. Different constructions have been developed and investigated with regard to the internal heat transfer resistance in combination with freeze-thaw-cycles. These absorbers have been integrated in solar collector prototypes, and thus the thermal performance and the reliability have been examined. A high thermal efficiency may only be achieved with a low heat transfer resistance of the complete construction. In order to minimize this thermal resistance between absorber and fluid, the low thermal conductivity of standard elastomer material has to be improved. Therefore different elastomer mixtures with significantly higher thermal conductivities and acceptable mechanical properties have been developed and investigated. In the following, the results of the development and analysis work will be presented, future applications will be discussed.
2. COLI~CTOR DESIGN It is evident, that the low thermal conductivity of normal elastomer material results in a high thermal resistance between absorber and fluid, which lowers the thermal performance of a collector with this design. For this reason, a theoretical study of the absorber heat transfer had to be performed first. The results of appropriate numerical calculations have led to the following conclusions: 9 A direct contact of the metal absorber fin with the elastomer tube is necessary, no additional adhesive or contact material should be used. 9 The thermal conductivity of standard black elastomer material (around 0.25 W/mK) should be increased to a value of around 0.7 to 1.0 W/mK. The contact area between the absorber fin and elastomer tube must be large, the wall thickness of the tube should be small and finally, the diameter of the tube should be large.
84
ISES Solar World Congress 1999, Volume III
The last two requirements are in contradiction to the necessary strength of the elastomer tube at operation pressure. Of special importance for a high thermal performance is the contact between the metal absorber and the elastomer tube. Therefore, during the first development steps, different absorber stripe constructions as well as different collector prototypes have been investigated with special regard to the heat transfer characteristics, the thermal performance and the reliability. Figure 2 shows five different constructions of realized absorber shapes, which have been investigated up to now.
3. ANALYSIS O F I N F E R N A L T H E R M A L R F ~ I S T A N C E The efficiency of a solar collector mainly depends on the quality of the absorber. Beside the absorption and the emission of the coating, the capability to transfer the heat from the absorber to the fluid is important. Figure 3 shows the thermal resistance network of a typical absorber stripe.
Figure 3: Simplified thermal steady state model of absorber stripes.
Figure 2: Different design types of the elastomer-metal-absorber. Type "A" is a typically soldered or welded absorber construction which has been used for the first experiments. Type "B" is an absorber construction out of roll bended aluminium sheets. The clip profile, which embraces the elastomer tube, is integrated in the sheet, thus no welding or soldering is necessary. Type "C" and type "D" are aluminium roll shaped constructions, which are used as absorber in typical thermal collectors. Normally a copper fluid tube instead of the elastomer tube willbe used in the clip profile. Type "E" is a specially developed absorber construction out of roll bended aluminium sheets. This clip profile is an improvement of the types "A" to "D" and takes the capabilities of a roll-form machine for 1 mm thick aluminium sheets into account. These different constructions have been used as absorbers in collector prototypes for the measurement of the thermal performance and as single absorber stripes for the investigation of the internal heat transfer capability.
The heat has to pass four single resistances on its way to the fluid. These are the resistance of the absorber fin and of the base connection, the tube wall resistance and the convective resistance between tube and fluid. The serial connection of these single resistances is equivalent to the total resistance between the absorber and the fluid, (1/Uint). For the elastomer-metal-absorber the internal thermal resistance resp. the internal heat transfer capability depends on: 9 fin resistance - characterised by the tube distance W, the base diameter D, the fin thickness sf and the fin conductivity kf. 9 connection between fin and tube - characterized by the connection technique and its production quality. 9 tube resistance - characterised by the conductivity of the elastomer k t, the tube diameter d t, the wall thickness of the tube st and the contact area between the clip profile and the tube, i.e. the contact angle 9. 9 convection between the inner tube wall and the fluidcharacterized by the convective heat transfer coefficient txfluid, which is a function of the fluid, its flow velocity and its temperature and the inner tube diameter and surface. In metal absorbers, the resistance of the tube wall (1/Utube) is normally neglectable because of the high conductivity of the metal fluid tube. However, in the case of the elastomer-metalabsorber, the tube resistance is very important for the total resistance of the absorber construction. It may be summarized, that the internal heat transfer resistance depends on the construction parameters of the absorber sheet and the fluid tube, the connection technique between the absorber sheet and the fluid tube, its production quality and of the operation parameters. The dependency of the internal thermal resistance of these different construction parameters has been determined by calculations, for which the following base case parameters of the elastomer-metal-absorber have been used:
ISES Solar World Congress 1999, Volume III
fin W [mm]
D [ram]
100
7
tube sf kf dt,i [ram] LW/mK] [mm] 1
200
9
conv.
st [ram]
tp [o]
r d L-W/m2]
2
~',7~)
2000
Table 1: Base case parameters for the calculation of the internal thermal resistance of the elastomer-metal-absorber. Figure 4 presents results of Uint-calculations carded out for different tube distances W, in figure 5 the tube diameter to wall thickness ratio is varied. In both figures, the parameter is the thermal conductivity of the elastomer material.
85
importance is the tube distance. Furthermore, the tube diameter resp. the thickness of the tube as well as the contact area between the clip profile and the tube have a clear influence on the internal heat transfer resistance of the construction and therefore on the efficiency of the collector. As the collector efficiency factor F' and thus the conversion factor 11o depend on the ratio Uint F'= Ulos s + Uint ,
(1)
the Uint-value should be maximized by optimization of the complete absorber construction. For a nonselective single glazed collector Uint should be higher than 50 W/m2K (Uloss = 5.5 W/m2K, F ' = 0.90) and for an unglazed absorber a value of more than 70 W/m2K is desired (Uloss = 15 W/m2K, F' = 0.82). These relatively high Uint-Values may only be achieved with a thermal conductivity of the elastomer of at least 0.6 W/InK, if a realistic tube distance of more than 80 mm is assumed. The other alternative, to reduce the tube wall thickness, has clear boundaries: A long term reliability requires a wall thickness of at least 1.5 ram. Therefore the thermal conductivity of standard black elastomer material (kt = 0.25 W/mK) has to be increased significantly.
4. DEVELOPMENT OF THE EI~STOMER-MATERIAL
Figure 4: Internal heat transfer capability Uint versus tube distance, parameter is the thermal conductivity of the elastomer material.
The department of material technology of the University of Applied Science in Osnabrtick has developed an elastomer material based on ethylene-propylene-dien-terpolymer (EPDM) for the application in the elastomer-metal-absorber. Main part of the development was to increase the low thermal conductivity of typical elastomer material with a simultaneous improvement of the mechanical strength. In addition to typically used well conductive f'dling materials like carbon black, particles out of aluminium and graphite have been applied during the development steps. The EPDM mixture is varied with different types of carbon black with high electrical conductivity, two kinds of aluminium particles and various kinds of graphite powder. The measured thermal conductivity and the tensile strength of some of the elastomer mixtures are presented in figure 6.
Figure 5: Internal heat transfer capability Uint versus the ratio of mean tube diameter to the tube wall thickness, parameter is the thermal conductivity of the elastomer material. As figure 4 and 5 show, the internal heat transfer capability of the elastomer-metal-absorber is mainly determined by the low conductivity of the elastomer material, which leads to a high thermal resistance of the tube. This tube resistance becomes even more important, if the amount of collected heat transported over this resistance increases. Therefore the second parameter of major
Figure 6: Thermal conductivity and tensile strength of different elastomer mixtures.
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ISES Solar World Congress 1999, Volume Ill
The mixture indication starting with a "V" labels the first series with only one single additive, the indication "M" is for laboratory mixtures with two conductive filling materials and the indication "D" stands for elastomer mixtures produced by using an industrial mixing device. For the elastomer mixtures in figure 6 the following filling materials have been used: "V-0" pure polymer "V-I" addition of 100 phr 1) aluminium particles "V-2" addition of 40 phr carbon black "M-4" addition of 40 phr carbon and 80 phr aluminium "M-7" addition of 30 phr carbon and 80 phr graphite "M-13" addition of 10 phr carbon and 90 phr graphite "D-4" addition of carbon and graphite The pure polymer without filling materials shows a thermal conductivity of about 0.2 W/mK. Aluminium as single filling material like in mixture V-1 improves only the thermal conductivity. If a conductive carbon black is added to the mixture (V-2), the tensile strength is raised more than six times and the thermal conductivity is doubled. The mixtures M-4 and M-7 contain two filling materials. Beside the conductive carbon black an aluminium or a graphite powder is added to the polymer. The thermal conductivity is raised up to around 0.8 W/mK, four times the value of the pure polymer, and the tensile strength is at a high level, too. For the efficiency measurement of the first improved prototypes, the elastomer mixtures M-5 (similar to M-4) and M-7, with a good thermal conductivity and a good tensile strength, have been used. From these new elastomer mixtures, tubes have been extruded and integrated into the test collectors. Due to the high carbon black content, the materials M-5, M-7 and M-13 have a very high viscosity during the mixing process and the extruded tubes show a high hardness and a low flexibility. Furthermore, the tube surface has a significant roughness. The conclusion of this first elastomer development step is, that high thermal conductivity and tensile strength values have been achieved, but the material is not appropriate for an industrial production process and does not result in the desired properties of elastomer tubes. The second EPDM development step therefore focuses on the improvement of the production parameters and the final elastomer material data like hardness, stress relaxation, torsion pendelum and ageing resistance. For this purpose, the content of carbon black has been reduced and the other components have been adjusted with regard to the special requirements. First result is the mixture D-4, the first sample produced in an industrial mixer, which shows a clear progress and already meets some of the requirements. However, further efforts are necessary for the optimization of the elastomer material for the use as fluid tube in the elastomer-metal-absorber, especially with regard to the production parameters, costs and long-term reliability. This work is going on. One problem is inherent with the application of EPDM as tube material. The temperature resistance is restricted to a short term maximum temperature below 160 ~ as the elastomer presents a
1. "phr" means "per hundred rubber", i.e. the number of weight parts of the Idling materialwhich will be added to hundred weightpart of the basis polymer material.
clearly decreasing strength with increasing temperatures and an accelerated degradation at such high temperatures. This has two consequences: 9 The stagnation temperature has to be reduced to a value below 160 ~ Therefore the heat loss coefficient must be higher than of commercial high performance flat plate collectors, which come up to more than 200 ~ at 1000 W/m 2, 30 ~ air temperature and low air speed. The heat loss coefficient a 1 (according to ISO 9806-1, referred to mean fluid temperature) must be higher than or equal to 4.5 W/m2K. 9 The system design has to avoid the simultaneous occurrence of high pressure and high temperature, which is the case for typical closed loop solar systems. 5. EXPERIMENTS O N THERMAL PROPERTIES
During the development steps of the elastomer-metal-absorber, the internal thermal conductivity between solar absorber and fluid, the Uint-value, has been determined by numerical calculations, measurements at single absorber stripes and measurements at complete solar absorbers during the performance test procedure of test collectors. Table 2 presents some of the most important results. profile
type
elastomer tube clip profile Oint q~ st i calcul. m e a s u r . [~ [mini [W/mK] [mm] ~V/m~] tW/m2K] 285
12,0
0,25
12,0
2,0
19,3
20,7
285
12,0
0,78
11,7
1,5
58,1
39,7
250
13,2
0,78
13,2
1,5
55,5
47,3
255
13,0
0,78
13,2
1,5
57,4
52,0
255
13,0
0,75
13,2
1,5
55,1
52,5
290
11,0
0,78
11,7
1,5
57,7
55,3
260
13,0
0,75
13,2
1,5
61,5
50,2
270
12,2 0,7- 1
12,2
1,5
51-61
58,7 ,
275
12,2
12,8
2,0
61,5
1,0
Table 2: Heat transfer capability Uint of different elastomer-metalabsorber constructions, measured and calculated values, tube distance is constant (W = 115 ram, except second line from bottom: W = 135 ram). The five types of absorber profiles presented in figure 2 have been investigated with different construction and material parameters. The calculated and measured internal heat transfer capability of the construction depends, like discussed in chapter 3, on the conductivity of the elastomer k t, the tube diameter dt, the thickness of the tube st and on the contact angle of the clip profile. The tube distance is the same for each construction (W = 115 ram) and the base diameter D is varied only in a small range. With the fin and tube construction parameters, the internal heat transfer capability has been calculated. These theoretical values may be compared with the experimental results. Up to now, the internal heat transfer capability has been increased from 20 W/m2K to 60 W/m2K, resulting in a collector efficiency factor which raised from 0.78 up to 0.92 for typical nonselective collectors (Uloss = 5.5 W/m2K). That means, that the realistic aim of 60 W/m2K has already been achieved, an objective for the future is 75 W/m2K.
ISES Solar World Congress 1999, Volume III
If the construction type C (see table 2, Uint = 57.7 W/m2K) would have been equipped with a metal fluid tube instead of the elastomer tube, the collector efficiency factor would be 0.95 instead of 0.91. The difference of 0.04 is the price for an absorber construction with elastomer fluid tubes. Also with future optimized constructions (Uint = 75 W/m:K) this difference will be around 0.03. Table 2 shows, that some calculated values fit rather well to the measured ones, others show significantly lower measured values. The main reason is the thermal contact between the tube and the metal profile. As some of the tubes showed a low flexibility, the contact to the absorber has been reduced, as the uneven and hard tube wall does not touch the whole embracing metal area. Therefore, the flexibility is an important quantity. For this reason, the fluid pressure normally has a positive influence on Uint and it could furthermore be remarked, that a heating-up under pressure also improves the thermal contact. Another important influence may also be derived from table 2. If the outer diameter of the tube is too small in comparison to the clip profile, the measured value of Uint are significantly lower than the calculated ones. Therefore, the outer diameter of the tube should be around 0.5 mm larger than the profile circle. Here the production tolerance has to be taken into account. Up to now, five different test collectors with an integrated elastomer-metal-absorbers have been constructed and investigated. For the collector frame, insulation and cover components of a standard flat plate collector have been used. Figure 7 shows two of the test collectors in front of the institute's building.
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The first test collector had a copper absorber plate with soldered clip profiles in form of the type A construction. With this collector, the base case investigations and the first measurements with the improved elastomer tubes have been performed. For the base case investigations, the absorber was equipped with a conventional rubber tube (low thermal conductivity of around 0.25 W/InK) and a black painted surface (EMA-1). For the second base case collector (EMA-2) an adhesive selective foil has been used instead of the black painted surface. The first improved test collector (EMA-3) contains the same absorber construction, but the rubber tube has been replaced by a tube made out of the improved elastomer (similar to mixture M-4, figure 5). Again an adhesive selective foil has been applied. The second improved test collector (EMA-4) was produced with a type B absorber construction (roll bended absorber sheets) and an improved elastomer tube with a larger diameter. The absorber coating is again the adhesive selective foil. The diagram in figure 8 presents the measured efficiency curves of the improved test collectors EMA-3 and EMA-4 in comparison to a typical selective flat plate collector as well as to the selectively coated base case test collector.
Figure 8: Efficiency curves of different test collectors compared with a typical selective flat plate collector, test conditions: irradiance level 800 W/m 2, ambient temperature 20 ~ air speed 3 m/s, according ISO 9806-1, referring to mean fluid temperature and aperture area. The resulting low conversion factor of 0.67 of the base case collector with a selective surface (EMA-2) is caused by the high thermal resistance of the rubber hose. For the first test collector with an improved elastomer tube (elastomer mixture M-5) and a selective absorber coating the conversion factor was raised up to 0.78. The conversion factor of the second improved test collector reached 0.81. The differences of this prototype EMA-4 is the use of a more flexible elastomer tube made out of mixture M-7 and the use of an other absorber construction, profile "B". These improvements during the first development steps have shown, that the proposed elastomer-metal-absorber construction gives a thermal performance close to that of typical flat plate collectors with selective metal absorbers. Figure 7: Test collectors in front of the institute's building.
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However, the conversion factor of an absorber construction with an elastomer tube will remain at least 3 % smaller in comparison to the same construction with a metal fluid pipe. As the reliability of the absorber is the most important condition for any future applications, first reliability investigations have been carded out on the elastomer tube, the absorber construction and the collector prototypes. Burst pressure, long term stability at high temperature and pressure and the torsion vibration properties have been investigated on the tubes, freezethaw-cycles have been performed with different types of absorber stripes and exposition tests on a complete collector prototype have been carded out. The results showed, that the existing problems should be solvable.
6. APPLICATION AS FACADE AND ROOF ELEMENT Industrially produced roofs and faqades often consist of corrugated metal form sheets made out of steel or aluminium. These roof or faqade constructions are widely used for industrial, public or residential buildings. The elastomer-metal-absorber concept will transform these metal form sheets into uncovered or transparently covered roof and faqade absorbers by integrating an appropriate clip profile into the form sheets during the production process. The elastomer tube can then easily be clipped into these profiles after the installation of the roof or faqade. Figure 9 shows the conversion of a typical metal form sheet (presented here as insulated sandwich plate) into an unglazed or transparently covered solar collector.
" (
"
( 8
2( f'
( 8
2 ( "4
(
2(
" I
Figure 9: Steps from a metal roof and fafade element to a solar collector. The first step of the conversion is the integration of the clip profile into the metal form sheet during the roll form process. The form sheet is covered with a paint of high solar absorptivity, with or without selective properties. The sheet will be mounted on the roof or faqade by normal roofing or metal processing companies. The optical and technical properties will be the same like for normal metal roofs. The second step is the integration of the elastomer tubes into the form sheets. The elastomer tubes will be connected via the manifold tubes to the solar system. Thus, an uncovered absorber results with only little extra costs, where the technical properties of the metal roof or facade remain unchanged. As an additional option for systems with higher demand temperatures a transparent cover may be added, using single glass panes or transparent plastic covers. By this way also low cost glazed collectors may be produced, which are specially suited for large systems.
The idea of this building integrated collector type has the following advantages: 9 Metal form sheets are a common and well proved technology. 9 The transformation into the elastomer-metal-absorber does not affect the reliability of the original roof resp. faqade. 9 The additional effort to transform metal roofs into unglazed absorbers seems to be very low, on the other hand, the metal roof and faqade elements gain by their new property as active solar absorber further attractivity. 9 The extension to glazed collector roofs for higher demand temperature is possible. 9 The integration may be performed with a high aesthetical quality and architectural acceptance. Typical examples for a future application of this concept are buildings with a high demand of low temperature heat, e.g. swimming-bath and sports halls, hospitals etc. for glazed collector constructions and outdoor swimming-pools and heat-pump systems for the unglazed absorber type. Domestic hot water and residential room heating purposes may also be taken into account. Due to the very low additional costs expected for the transformation of the metal building envelope into a glazed or unglazed solar collector, this concept has the potential to result in new solar applications with a high economic benefit.
7. DEVELOPMENT PROJECT- STATF.,-OF-THE-ART Despite of the encouraging results of the first development steps, this absorber type is not available up to now. Open questions are mainly the production and installation technology, the long term reliability and the long term thermal performance. A research and development project, funded by the European Commission, has started to develop and investigate the integrated elastomer-metal-absorber in roof and facade metal form sheet elements and possible heat use applications. The main tasks within this cooperation between industrial partners coming from various activity fields and research institutions are: 9 further improvement of the elastomer material with special regard to heat conductivity, mechanical strength and durability, as well as the production of appropriate elastomer tubes, 9 development of absorber constructions with focus on production parameters, thermal performance and reliability, 9 construction and assessment of test collectors, determination of thermal performance and reliability characteristics, 9 development of different solar system concepts, assessment of collectors in test systems, comparison and extrapolation. The first results of this project are encouraging: 9 An improved elastomer mixture with special regard to the industrial produceability has been developed, from which 9 first prototypes of an appropriate elastomer tube have been produced in an industrial extrusion machine. 9 The construction of the form sheet elements for the integration into metal-roofs has been performed, the tools for the roll form machine are ready and the first absorber form sheet elements are in production. 9 A simulation tool for the elastomer-metal-absorber has been developed, first different heat use concepts have been worked out and simulated. However, still a couple of problems exist which have to be solved on the way to an industrial product with high performance and reliability.
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Some items to be worked out are in the fields of: 9 assessment of the absorptive surface and its stability, 9 design of the hydraulic system and the manifolds, 9 development of the connection technique between tubes and manifold, 9 and system operation and security technique, with special regard to the fact, that water will be used as heat transfer fluid. It has to be pointed out here, that at the moment it is planned to transform the whole roof area into an elastomer-metal-absorber, i.e. the normal application are large collector areas. For the specific problems arising from this aspect, the development has to go on over intermediate stages like medium sized pilot and demonstration plants. 8. C O N C L U S I O N AND O U T L O O K The principle of the elastomer-metal-absorber with its clip prof'lle contact opens up new possibilities with regard to the heat transfer fluid, the collector and system design and the architectural integration. The development steps have shown that the proposed elastomer-metal-absorber construction already has a thermal performance close to that of typical flat plate collectors, with only a slightly lower conversion factor. The essential results up to now are the increase of the thermal conductivity of the elastomer material from 0.25 W/mK up to 1.0 W/mK, which in combination with an optimized absorber construction leads to an internal heat transfer coefficient of at least 60 W/m2K, a value comparable to standard flat plate collectors. The existing reliability problems seem to be soluble, the first results of the actual development project are encouraging. The special attraction of this building integrated design is given for the following reasons: 9 high expected cost reduction for unglazed absorbers or glazed collectors, 9 significant reduction of energetic amortisation periods, 9 well suited solution for repair or recycling, 9 enlargement of the solar market by new manufacturers and solar systems, especially in large commercial and public buildings, as well as in residential buildings, 9 and improvement of architectural acceptance by the high degree of building integration.
Acknowledgements-The work is funded partially by the European Commission, within the project ,,Faqade and Roof Integrated Solar Collectors with a Combination of Elastomer Tubes and Metal Form Sheet Elements", contract no. JOE3-CT98-0236, organized in the framework of the Non-Nuclear Energy Research and Technological Development Programme JOULE Ill.
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NOMENCIJ~TURE a1
linear collector heat loss coefficient, referred to Tm (W/m2K) 0t~luid convective heat transfer coefficient between the inner tube wall and the fluid (W/m 2) D base diameter (projection width of visible tube surface)
(mm) df dt dt,i dt,m F' G 11 kf kt tp sf st Tabs Tbase Tm Tt,out Tt,in Ufm Ubase Utube Ueonv
diameter of the clip profile (mm) tube diameter (ram) inner tube diameter (mm) arithmetic mean of outer and inner tube diameter (ram) collector efficiency factor (-) Solar irradiance (3br/m2) collector thermal efficiency, referred to T m (-) thermal conductivity of the fin material (W/mK) thermal conductivity of elastomer tube material (W/mK) contact angle of the clip profile (o) thickness of the fin (mm) wall thickness of the tube (ram) mean temperature on the absorber fin (~ temperature on the absorber base (~ mean temperature of heat transfer fluid (~ temperature on the outer surface of the tube (~ temperature on the inner surface of the tube (~ internal heat transfer conductivity of the fin (W/m2K) internal heat transfer conductivity of the base (W/m2K) internal heat transfer conductivity of the tube (W/m2K) convective heat transfer conductivity between tube wall and fluid (W/m2K) Uint internal heat transfer conductivity of absorber construction (W/m2K) Uloss overall heat loss coefficient of the collector, referred to T m (W/m2K) W tube distance (ram) REFERENCES Bartelsen B., Rockendorf G. and Vennemann N. (1996) Development of an Elastomer-Metal-Absorber for Thermal Solar Collectors. In Proceedings of the EuroSun '96, 16-20 September, Freiburg, Germany, pp. 495-499, DGS-Sonnenenergie Verlag GmbH, Mtinchen. Rockendorf G., Falk S.,Wetzel W. (1996) Bedeutung und Bestimmung des Kollektorwirkungsgradfaktors bei Sonnenkollektoren; 6. Symposium thermische Solarenergie, 08-10 May, Staffelstein, Germany, pp. 196-201, O T H e.V., Regensburg. Duffle J.A. and Beckmann W.A. (1991) Solar Engineering of Thermal Processes; 2n d edn, pp. 268-276, Wiley-interscience Publication; New York. B6kamp K., Vennemann N., Wallach J., Bartelsen B. and Rockendorf G. (1997) EPDM Compounds with Improved Thermal Conductivity for Thermal Solar Collectors. In Proceedings of the International Rubber Conference, 30 June - 3 July, Ntirnberg, Germany.
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SOLAR ABSORBER SYSTEM FOR PREHEATING FEEDING WATER FOR DISTRICT HEATING NETS Klaus Vajen, Marcel Kdimer FBPhysik, Universit~itMarburg, D-35032Marburg, Germany phone -H-49/6421/28-4148, fax ++49/6421/28-6535, [email protected] Ralf Orths Wagner & Co Solartechnik, Ringstr. 14, D-35091 C61be, Germany
Erkin K. Boronbaev, Astra Paizuldaeva Kyrgyz State University of Construction, Transport and Architecture, 34 b Maldybaevstr., KS-720023 Bishkek, Kyrgyzstan
A b s t r a c t - EPDM-absorbers, made of artificial rubber and well-known in Central Europe for heating swimming pools, have been installed to preheat domestic water in a heat and power plant in Bishkek (Kyrgyzstan). Measurements were carried out during the summer 1998. The special construction of the district heating net and the climatic conditions of Central Asia lead to a favourable environment for the utilisation of solar thermal energy. Fluid temperatures nearly always far below ambient temperature result in convective heat gains instead of losses. Collector "efficiencies" far above 1 as well as nightly heat gains were measured. Calculations of solar energy prices lead to about 6 Euro/MWh useful energy.
1. INTRODUCTION The heat supply of cities in the former Soviet Union usually is provided by one or more district heating (and power) plants. So it is in Bishkek, the capital of Kyrgyzstan. 350.000 inhabitants receive domestic hot water and energy for room heating from the central Heat and Power Plant of Bishkek City. The district heating net (fig.l), however, shows some differences to common Central European technology. In Bishkek (as in many other cities of the CIS) one finds an open circle system: domestic hot water is taken by the consumers directly out of the net without any heat-exchanger coupling. Thus in Bishkek the amount of 3000..4000 ma/h water has to be refilled into the net. This is carried out at one central place. Cold water is taken from the ground and artesian sources and led to boilers which heat it to the required temperature of 60~ Due to Kyrgyzstan's climatic conditions (Central Asia, latitude 43 ~ north, comparable with Rome), altogether these are nearly ideal conditions for the implementation of solar thermal systems. So it stood to reason to preheat the cold water directly by uncovered solar collectors. In Central Europe they are wellknown for swimming-pool heating. 2. MEASUREMENTS In June 98 a test plant of a 50 m2 EPDM absorber field was installed on one roof of the District Heating Plant of Bishkek City. The absorber had an inclination of 4 ~ to west. The water was taken from the pressure pipe (behind the pumps) and led back to the pressure-less pipe, so no extra pump was necessary to force water circulation. The collector flow rate was varied by a hand valve and also unspecified altered due to pressure
changes in the net. The variation width was 10 l/m2h up to 1201/m2h. Measurements took place from June to October 1998. Apart from the flow rate, the global, diffuse and long wave radiation, Tin, Tout, Tamb, humidity and the wind-speed were measured, automated by a computer system Mean values of up to 4000 single measurements were stored on the harddisk minutely. The system worked nearly without problems during the summer.
Fig. 1. Simplified scheme o f the distn'ct heating net in mar absorber
,
~oreoe
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m II
c~a w ~ r 0 ~ ' Q
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:l
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(60~ C in summer)
an U
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| ~
l
=
~ m
=
=
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=
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= u
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=
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= , m =
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(only in winter)
Bishkek. Huge cold- (not included in the figure) and hotwater storages lead to nearly constant cold water and heat demands. In order to use an existing pump, the back flow from the collector was connected before the turn-off o f the collector forward flow. The flow rates were about 3000 m3/h through the pump and 5 mS/h through the absorber. 3. RESULTS Data from June, 13 to August, 10 were taken for the following evaluation. The cold water inlet temperature was always about 12 to 13~ The ambient temperature, however, was nearly always higher, even at night. This leads to the
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unusual behaviour, that the net energy balance of the absorber shows profits from the surroundings instead of losses, see fig. 2.
Fig. 2. Example o f the measured temperature courses. During the selected days the ambient temperature was always even higher than the outlet temperature. O f course, this temperature difference depends on the collector flow rate. I f the dew-point is above the inlet temperature, at least on a part o f the absorber condensation occurs.
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independent of the flow rate. The average heat gains were 0.7 kWh/m2night or 80 W/m2, respectively (see fig. 4). With a further developed absorber model, taking into account also condensative heat gains, the measured results could be reproduced. More detailed results of the measurement and the modelling will be published later.
Fig. 4. Measured hourly mean values o f the nigh@ useful power gains (June to August 1998). 4. CONCLUSIONS
Fig. 3. Measured hourly mean values o f the collector efficiency during the day, which is nearly always > 1 at high flow reates (June to August 1998). Since the efficiency of an "'ordinary collector" depends on the wind speed an "'efficiency field'" is sketched in. In fig. 3 the collector efficiencies during the day (this means 5.30 to 20.30 h local time) are shown. Best results could be achieved at flow rates higher than 40 l/m2h. The high ambient temperatures lead to collector "efficiencies" more than 1. The highest values could be observed in the early morning and the late afternoon. Note, that unlike usual collector characteristics nearly all dots can be found in the 2~a quadrant of the coordinate system. In contrast to the figure in the "Book of Abstracts" fig. 3 does consider partly shading of the absorber field. In contrast to the collector efficiency at day, during the night (20.30 to 5.30 h local time) the power gains were nearly
The yearly heat gains of an uncovered collector connected with the district heating net in Bishkek can be estimated to be higher than 1100 kWh/m2. So the absorber heat gains can be expected to be more than twice as high as common for collector systems in Central Europe, furthermore the installation costs of the absorbers are very low. With the results measured at the test-plant solar energy prices of about 6 Euro/MWh useful energy can be expected for an absorber area > 1000 m 2. This is below the today's prices of fossil fuels on the world market. So the absorber system in Bishkek could be a solar thermal installation able to compete economically with all conventional energy sources. The estimated technical potential only in Bishkek is higher than 40000 m 2 absorber field. An installation (abt. 1000m 2) in Bishkek is under consideration. ACKNOWLEDGEMENTS The authors would like to express their sincere thanks to the following persons and institutions for financial and logistic support: Unversiti4t Marburg, Wagner Solarteclmik in Crlbe (Germany), International Bureau of the German Ministry of Education and Research, Heat and Power Plant of Bishkek City and to the Embassy of the Federal Republic of Germany in Bishkek.
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STATISTICAL ANALYSIS OF SOLAR COLLECTOR TEST RESULTS IN VIEW OF FUTURE CERTIFICATION Emmanouil Mathioulakis, Kostantinos Voropoulos and Vasilis Belessiotis Solar & other Energy Systems Laboratory, NCSR >, 15310 Ag. Paraskevi Attikis, Greece Tel. +301 6544592, Fax +301 6544592, E-mail: [email protected] Abstract - This paper deals with the need to develop certification schemes for solar products as a means for a further promotion of solar energy applications. It also examines the ways in which these schemes could be implemented. More specifically, an analysis of the results of solar collector efficiency testing is presented, aimed at depicting the existing situation. A methodology for the exploitation of these results is proposed, leading to a realistic approach to the criteria that could be used in future certification schemes
1. INTRODUCTION Quality is well known to be the most essential factor for the survival of every commodity produced and offered for sale in today's strong competitive market. The quality level of a product can be proved with its certification and marking, based upon specific evaluation criteria and generally accepted procedures. Moreover, it is commonly accepted that the absence of objective and undisputed mechanisms for the assessment of solar collectors and the qualitative characteristics of the systems is one of the main obstacles to the further promotion of these products. These mechanisms should be based on results from tests undertaken in accordance with existing national, European or international testing standards, and should permit, within the framework of a certification scheme, the classification and marking of products depending on their performance. A certification scheme should not be based on abstract concepts. It can be realistic only if it takes into account the actual situation of the productive sector. In this regard, the exploitation of the existing test results could be an essential aid in the development of a certification scheme. In this paper, an attempt is made to relate an evaluation mechanism that could be used in a solar collector certification scheme with the experience gained from their testing to date, particularly concerning efficiency. 2. CERTIFICATION OF SOLAR ENERGY PRODUCTS A survey of not only the Greek but also the international experience from the existing applications of solar thermal energy systems leads to two, at a first sight, contradictory conclusions (EU, 1996). On one hand, the technologies used nowadays are mature, which means that there are no critical technological matters that could obstruct the dissemination of such systems. On the other hand, practical applications still present important efficiency and reliability problems, resulting in the fact that that there has been a reduction in the faith which the potential users have in these products (Mathioulakis and Belessiotis). Several explanations can be given to this situation, one being, without doubt, the quality of products. There is a distance between the technological know-how and the products available in the market. This is due, amongst others,
to the new reality of flee movement of goods and to the inadequate quality control. Because of this, the issues that concern the evaluation of performance and the marking for quality of solar thermal collectors are of great importance. Certification of solar energy products can become an essential tool for dealing with the problems mentioned above. It should not be forgotten that the quality and reliability of the proposed solutions are of key importance when trying to formulate a favorable legal and financial framework for the promotion of solar energy usage. The certified quality of the products is a motive for consumers who are now able to make their choices based on objective criteria. It also constitutes a motive for manufacturers to improve their products. It facilitates a more realistic approach to the issues of financing and economical effectiveness of solar energy exploitation systems. Finally it contributes to market control since it introduces more transparency in the evaluation of existing technologies. Confrontation of this problem varies from country to country. In some cases, certification schemes are applied in connection with support actions, such as subsidies. However, market internationalization within the last few years is leading in the direction of a common international practice, which is facilitated by the gradual development of common standards concerning testing methods. A typical case is the that of the new European Standards in the field of solar energy products. Their basic target is the technical harmonization of testing methods in all European countries and their use as a base for certification. The new European Standards adhere to the testing methods of ISO, with minor changes, and add new Standards of <> (CEN, 1998). The certification procedure that could be developed on the basis of the new European Standards is shown in figure 1. All certification schemes concerning solar energy products are optional. But in some cases they turn out to be compulsory in praxis when they are used as a criterion for subsidizing the installations. The gradual harmonization of the certification schemes used in several countries is now a realistic target. However it requires the harmonization of testing methods, which is still in progress, and of the way in which the products are classified with respect to their performance.
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where no is the maximum collector efficiency and U0 the collector heat loss coefficient.. In some cases, a three-parameter model expresses the collector behavior more clearly, especially for collectors with non-selective absorbers. However, for the scope of the present paper and the homogeneity of the results, the model of equation 1 is used. : I n figure 2 the determined factors no and U0 of the instantaneous efficiency curve are plotted. On this figure the <> are distinguished, showing the separation between the technologies used and the direction of improved collector efficiency. It is shown from the graph that the collectors with selective absorbers are clearly separated from the rest with respect to their energy characteristics and especially the heat loss coefficient. Collectors with semiselective absorbers have improved performances compared with those which are black-painted but are till worse than the selective ones. Figure 1: General layout of certification procedure
The efficiency of the collector, which is a basic characteristic of quality, is examined in the next paragraph. The manner in which collector efficiency is introduced into the general evaluation scheme for certification is therefore of great importance. Furthermore the same methodology can be applied to the remaining qualitative characteristics of the collector or of the solar system. 3. COLLECTOR EFFICIENCY TEST RESULTS Evaluation can be reliable if the real situation of the products available in the market is taken into account. This situation is depicted in test results. Results from the testing of about 100 collectors, tested over the last 7 years in the Solar & other Energy Systems of NCSR "Demokritos', form the basis of the analysis presented. Tests were carried out according to the ISO 9806-1 Standard [ISO, 1994]. These collectors constitute a rather broad sample of the product, with regard to construction type, absorber surface treatment and year of production. Thus, they are representative of the majority of the products produced and used over the last ten years in the market. Through a statistical analysis of the determined factors of the instantaneous efficiency curve of the solar collectors, and also of the instantaneous values of efficiency in specific ordinary operating conditions, several important results can be derived. Three different types of collectors were separated, concerning the absorber surface treatment: black-painted, semi-selective and selective collectors. We should keep in mind that according to the ISO 9806-1 Standard, the energy characterization of the collector is achieved by determining, from testing data, the coefficients no ~:at U0 of the collector efficiency equation (Duffle and Beckman, 1991): n = no- U0 I m
(1)
Figure 2 - Maximum efficiency nO and heat loss coefficient U0 of tested collectors +: Black paint, x: semi-selective, o: selective It should be noted however, as shown in figure 2, that the separation between the three collector types is not complete, given the fact that the respective <>overlap. This conclusion is very important since it demonstrates not only that the final result surely depends on the technologies used, but also that the integration and the correct design of the collector play an important role as well. Moreover it shows that the only reliable method for collector energy characterization is testing, through which the whole technological, design and construction particularities of the specific product are incorporated in the final result. The same conclusion, i.e. that a product should not be seen as a <<sum of technologies>> but as a whole (seen only by testing), is also depicted in figure 3. The results given in this figure concern systems of the same type (thermosiphon-type) and contain the values of their energy output calculated from testing according to ISO 9459-2 (ISO, 1995b), for the same climatic conditions. It is observed that the type of the collector used in the system is important, but it is not the only factor that plays important role. The advantages of the selective absorbers can be easily lost due to wrong choices
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when integrating a collector with such an absorber into the system.
Figure 5: +: determined values of n, solid line: normal distribution
Figure 3 - Energy output for different systems as a function of the type of collector surface From the above, it can be concluded that the energy characterization of the collector, within the framework of a certification scheme, must be referred to test results, i.e. the coefficients no and U0. However, it may be better, for reasons of simplicity, for the evaluation to be based on the instantaneous efficiency n calculated under specific climatic and operating conditions. These conditions could be, for example, a global solar irradianee of 800 W/m2 and a temperature difference AT = Tm-Ta of 30 K. Figure 4 gives the results for the three types of collectors concerned.
From the graph of figure 5, the percentage of collectors whose performance lies above a certain range of values can easily be determined. Thus, it facilitates the correlation of the efficiency evaluation criteria in the framework of a certification scheme with the quality of products that are actually available in the market. It must be stressed that for a certification procedure to be objective, the physical quantities used for product evaluation should be accompanied by the confidence interval, within which the values of the characteristic quantities lie. Having this in mind, the introduction of procedures for calculating the uncertainties of results in testing standards, and especially those concerning efficiency, would be very useful (ISO, 1995a). 4. Conclusions
Figure 4: Instantaneous efficiency n for standard conditions (G=800 wm2, AT=30 K) Presentation of the instantaneous efficiency values of all collectors in a probability plot shows the distribution of the determined values over the range which they appeared (figure 5). This curve, with a tendency towards normal distribution, contrary to the similar curves of no and U0, results in the fact that a kind of collector rating could be based on the instantaneous efficiency. The classification criteria can be determined by the probability distribution curve and the quality level that is desired to be fulfilled.
Quality is well known to be the most essential factor for the survival of every commodity produced and offered for sale in the today's strong competitive market. The level of the quality of a product can be assessed by its certification and marking, by implementing specific evaluation criteria and commonly accepted procedures. This also applies to solar collectors, a product which uses the flee and abundant solar energy for a large range of applications, while simultaneously protecting the environment. A methodology according to which the marking of solar collectors will be made, based on real data, is the proper tool for the upkeep of a high quality level of solar collectors and also for the their continuous improvement. For example, the analysis of the determined factors of the instantaneous efficiency curve of the solar collectors, and also of the instantaneous values of efficiency in specific ordinary operating conditions, should be used for the evaluation of their performance. As a result, this could be the start, together with other quality indices, of the development of a certification scheme, in view of the subsequent marking of solar thermal products.
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NOMECLATURE G m n
Ta Im
Tm Tm Tout AT
global incident solar irradiance, W m -2 mass flow rate through the collector, kg s1 collector instantaneous efficiency ambient air temperature, ~ reduced temperature difference I m =(Tm-Ta)/G, KW-1 m2 temperature of water in collector inlet, ~ mean temperature of water inside collector Tm---( Tin + Tout)/2, ~ temperature of water in collector outlet, ~ temperature difference AT=To~t-Tm,K
REFERENCUES CEN (1998), <9~rEN12975-1: Thermal solar systems and components - Collectors - Part 1: General requirements>>, CEN ed., Brussels Duffle J.A. and Beckman W.A. (1991), Solar engineering of thermalprocesses, 2nd edn, Willey, New York. EU (1996), Sun in Action, Office for Official Publication of the European Communities, Luxembourg ISO (1994), Standard 9806-1. Test methods for solar collectors - Part 1: Thermal performance of liquid heating collectors including pressure drop, ISO ed., Switzerland. ISO (1995), Guide to the expression of uncertainty in measurements, ISO ed., Switzerland. ISO (1995), Standard 9459-2. Solar Heating- Domestic Water Heating Systems- Part 2: Outdoor Test Methods for System Performance Characterization, ISO ed., Switzerland. Mathioulakis E. and Belessiotis V., Active solar systems Review of technologies and applications in Greece, In Proceedings of NTUA National Congress <<Application of renewable energy sources>>, 30 November- 2 December 1998, Athens, Greece (in Greek)
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THERMAL AND ELECTRICAL YIELD OF A COMBI-PANEL Herman A. Zondag~ Douwe W. de Vries and Anton A. van Steenhoven Department of Mechanical Engineering, Eindhoven University of Technology, P.O.Box 513, 5600 MB Eindhoven, The Netherlands, Tel.: + 31-40-2472726, fax: + 31-40-2433445, [email protected] Wim G.J. van Helden, Netherlands Energy Research Foundation ECN, P.O. Box 1, 1755 ZG Petten, the Netherlands Ronald J.C. van Zolingen Shell Solar Energy BV, P.O. Box 849, 5700 AV Helmond, the Netherlands Abstract - A first, non-optimised prototype of a combi-panel was built of a PV-laminate and a sheet-and-tube absorber. The thermal efficiency at zero reduced temperature was found to be 54%, along with 8.5% electrical efficiency. The results of the measurements were used to verify the results of the simulations. It was concluded that the simulations and the measurements corresponded sufficiently well. Then, the simulations were used to find the annual efficiency of a PV/T-system that was used for hot-water production in a Dutch household, for which 33% thermal and 6.7% electrical efficiency was found. Finally, the simulations were used to quantify the contribution of the various loss terms to the reduction in thermal efficiency of a PV/T-system with respect to a thermal collector.
1
INTRODUCTION
A combi-panel consists of a PV-laminate that functions as the absorber of a thermal collector. In this way a device is created that converts solar energy into both electrical and thermal energy. The main advantages of combi-panels are: 1. An area covered with combi-panels produces more electrical and thermal energy than a corresponding area partially covered with conventional PV systems and partially filled with conventional thermal collectors. This is particularly useful when the amount of space on a roof is limited. In addition, installation costs are reduced. This will become increasingly important in the future when the price of PV will be reduced. 2. Combi-collectors provide architectural uniformity on a roof, in contrast to a combination of separate PV- and thermal systems. 2
The efficiency of the combi panel was measured and compared to the efficiencies of a conventional sheet-and-tube=type thermal collector and a multi=crystalline silicon PV=panel of the same length and width, which were positioned next to it in the test rig. A photograph of the test rig is shown in figure 2. The original collector surfaces were somewhat larger than the PV= laminate. In order to create similar areas for the PV laminate, the thermal collector and the combi=panel, the absorbing surfaces of the latter two were partly covered with insulation that had a reflective aluminium top layer. In figure 2 these covered parts appear as the white areas around the collector and the combi-panel. The uncovered parts have an area of 0.94 m 2 each.
SYSTEM
In order to quantify the efficiency of a PV/T-collector, an experimental prototype was built at the Eindhoven University of Technology. This was a non-optimised first prototype, that was built in order to be able to validate the simulations. The prototype was constructed by connecting a conventional PVlaminate, containing multi-crystalline silicon cells, to the absorber plate of a conventional glass-covered sheet-and-tube collector, as shown in figure 1. The panel was then integrated into a test rig on the roof of the department of Mechanical Engineering at the Eindhoven University of Technology.
Fig. 1. Cross-section of the combi panel
Fig. 2. The test rig. Left to right: a conventional thermal collector, the combi panel and a conventional PV-laminate. The system consisted further of a water tank of 130 litres. The water was drawn fi'om the tank into the thermal collector and the combi-collector by a NKF Verder ND 300 KT 18 diaphragm pump. The construction was such that the water that was heated by the system could either be returned to the tank or could be discharged on the sewage system in order to keep the water temperature in the tank at a constant value. In the latter case the water level in the tank was kept constant through a tap that was connected to the water supply of the building. The water flow through the combi-panel and the conventional thermal collector were measured independently with two rotary
ISES Solar World Congress 1999, Volume Ill
piston KENT PSM-LT PL 10 water volume meters. The volume flow was measured by dividing the counted amount of litres by the measuring time. The wind speed was measured with an EKOPOWER MAXIMUM cup anemometer. The irradiation was measured with a Kipp & Zonen CM 11 pyranometer. The temperatures of the PV-laminate, the combi-panel laminate and the collector absorber as well as the in- and outflow temperatures of the collector and the combi-panel were measured with thermocouples type K which were calibrated to an accuracy of 0.2 K. The thermocouples, the pyranometer, the anemometer, the two water meters and the electrical output of the combi-laminate and the PV-laminate were read out by a DORIC digitrend 220 datalogger. The time between two measurements was typically 11 seconds. The PV laminate was a standard Shell Solar PV-laminate consisting of 72 10xl0 cm2 EVA encapsulated square multi-crystalline silicon cells with a low-iron glass front and an A1/tedlar film at the back. The cell efficiency under STC is typically 13%. The laminate efficiency at 25 ~ is 9.7%. 3
97
Both the electrical yield and the thermal yield are lower than found for the conventional collectors, as expected. However, the results show that two combi-panels together produce more energy per unit area than one PV-laminate and one thermal collector next to each other, which makes them interesting for solar energy production.
C O L L E C T O R EFFICIENCY CURVES
3.1 Measurements The thermal efficiency was measured as a function of reduced temperature. For these measurements, a mass flow of 76 kg/(m2 hr) was used. The conditions for the measurements were: 1. During a time span of 15 minutes the radiation is at least 750 W/m 2 and its value does not vary more than 100 W/m 2. 2. The fluid inlet temperature and outlet temperature do not vary more than 0.2 K during the measurement. 3. The flow rate for the collector and combi-panel is around 20xl 0 ~ m3/s and varies not more than 1.4xl 0 "~ m3/s. 4. The wind speed during the measurements does not exceed 2 m/s. In order to check if the restrictions mentioned above are sufficiently strict a set conditions of a somewhat less restrictive nature was applied. It was found that the results did not change significantly. The thermal efficiency is calculated from quantities averaged over 15 minutes. The thermal efficiency of the combi panel and the thermal efficiency of the conventional collector are presented in figure 3. The electrical efficiency of the combi panel and the PV panel are presented in figure 4.
Fig. 4. The electrical efficiency of the PV-laminate (,) and the combi-laminate (o).
3.2 Simulations The thermal efficiency is simulated assuming thermal equilibrium in the various layers of the combi-panel and the thermal collector. The relations describing the radiation and convection heat transfer are put in local heat balances for all layers in the combi panel. The sheet and tube type combi-panel is considered as one long straight tube, which implies that symmetry is assumed with respect to the centreline between two successive pipes. The heat flow in the direction perpendicular to the flow direction is calculated with a relation obtained from the well known Hottel-Whillier model for thermal collectors, that is based upon this same assumption of symmetry. It is described extensively by e.g. Duffle and Beckman (1991 ). The Hottel-Whillier model leads to a temperature distribution between two tubes in a sheet-and-tube collector that is given by
T ( x ) - Tarot, - V,~I / h / = T b -Ta,,a, - Z a I / h
I
cosh(mx)
(1)
cosh(m(W-D)/2)
in which x,~ is the transmission-absorption coefficient, I is the irradiation, hi is the heat loss coefficient, T ~ , is the ambient temperature and Tb is the temperature at the absorber surface directly above the tube, W is the distance between the tubes and D is the tube diameter. A typical temperature distribution between the collector tubes is shown by figure 5.
Fig. 3. The thermal efficiency of the thermal collector (x) and the combi-laminate (both with (+) and without (o) electricity production).
Fig. 5. The temperature profiles of the laminate and the absorber between the tubes of a sheet-and-tube collector.
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The coefficient m in formula 1 determines the flatness of the temperature profile and is given by
m
2
= ~
F = tanh(m(W-D)/2) m ( W - D ) /2
(3)
For a conventional thermal collector, the heat collection efficiency factor is then defined as
F" =
l/hi
1
(4)
~bo.d
1
Lhl (D + (W - D)F) + Abo.db+ff.Dhf
q - WF" ('~al - hl (Zw - Za ) )
(5)
"~a,eff"="~ T-~elO-O.OO5(Tlam-25~
3.
4.
hi
1 "~
1
Whca
0.9 0.8 9tD -eo 0.7 tD
"•
0.6
"~ 0.5 I
0.4 0
I
I
I
0.005 0.01 0.015 0.02 reduced temperature in KmVW
I
0.025
(6)
For xa a value of 0.744 was found from a simulation of the optical characteristics of the PV laminate. For the transmission of the glass cover x a value of 0.92 was applied, based on general low-iron glass transmission data. In the equation for m an additional term appears, due to the fact that the silicon cells provide a parallel channel for heat conduction, along with the conduction sideways through the copper absorber plate.
m2 =
1
F is the fin efficiency factor and hca is the heat transfer coefficient between the cells and the absorber If hca is small, the temperature gradient between the laminate and the copper absorber will be large and a large heat loss to the ambient will occur, which will reduce the thermal efficiency. 5. Finally, a PV-laminate is not spectrally selective, so the emission coefficient was changed from 0.12, which is a typical value for a spectrally selective surface (e.g. Duffle and Beckman (1991)), to 0.9. The full set of equations provides a matrix, which was solved by a matrix solving procedure of MATLAB. This results in a set of efficiency curves. In figure 6 the calculated efficiency curves are presented together with a least-squares fit of the measurements that were presented previously in figure 3.
in which Ill is the heat loss at the top of the laminate and hf is the heat transfer coefficient to the water in the tubes. The useful energy gain per unit tube length is given by formula 5.
Tw is the temperature of the water. In order to account for the special characteristics of the combipanel, three equations had to be modified with respect to the equations for a conventional thermal collector. 1. Due to conservation of energy, the solar energy that is converted to electricity cannot be converted to thermal energy anymore. Therefore, in the heat equations, xa should be replaced by its effective value.
(8)
-i-~ Lh l ( D + ( W - 9 ) 1 7 ) + ~ff.Dhf
(2)
given by
2.
l/hi W(
h~
The temperature gradient across the absorber drives the conductive heat transfer to the collector tubes. At the same time, the high temperature between neighbouring tubes causes additional losses, which means that not all the heat can be collected. This is expressed in the fin efficiency factor F that is
W(
F" =
(7)
Only the effects of the silicon and the copper are expressed in the equation, since the heat conduction through the EVA and the glass are much smaller than these. If the heat conduction through the silicon is large, the temperature profile across the combi laminate will be flatter than if the heat conduction is small. In the equation for the heat collection efficiency, an additional term appears due to the heat resistance hca between cells and absorber. The bond conductance can be neglected due to the high silver content of the bond
Fig. 6. The simulations of the thermal efficiency (dashed) compared to the least squares fit of the measurements (solid) for a conventional thermal collector and a combi-panel either or not producing electricity. The figure shows a reasonably good agreement between the simulations and the measurements, although the difference between the curves is in the range 0%-4%, which is somewhat larger than the experimental inaccuracy, which was found to be around 1%. The differences still present might be due to a slight overestimation of the optical efficiency or to heat loss to the sides of the copper absorber (the parts which are covered by insulation in order to keep the area of the PV laminate and the eombi absorber-plate equal; see figure 2). In addition, the sky temperature was not measured. In the simulations, it was assumed to be equal to the value for a clear sky. This could also account for a part of the difference. The clear-sky temperature is calculated from the formula
T,ky = 0.0552Ta~
(9)
3.3 Estimating the loss terms Next, the simulations were used to obtain information about the loss mechanisms in a eombi-panel. Figure 7 shows the magnitude of the radiation loss, the convection loss and the back loss. Together with these losses, the thermal and the
ISES Solar World Congress 1999, Volume III
electrical efficiencies are indicated. Finally, the straight line on top is the sum of all these terms, which is equal to the transmission-absorption coefficient of the combi-panel, as expected. The calculation was done by setting Ta~ = 20 ~ and I = 800 W/mEand increasing the inflow temperature, which is of some importance because these settings determine the PVlaminate temperature, which determines the electrical efficiency and which is by itself not a function of reduced temperature. The relative magnitude of the radiation and convection losses depends on the sky temperature. The calculation was done for a clear sky using formula 9.
99
seems to have a substantial effect on the slope of the thermal efficiency curve.
Fig. 8. The efficiency curves versus reduced temperature, successively removing the special features of the combi-panel. From low to high: (1) combi-panel, (2) optical efficiency enhanced, (3) heat transfer enhanced, (4) spectral selectivity enhanced, (5) heat conduction sideways through silicon removed, (6) no electricity produced. 4 Fig. 7. The loss mechanisms in the combi-panel as a function of reduced temperature (solid lines); (1) back loss, (3) convection loss and (4) radiation loss. The dashed lines indicate the electrical efficiency (2) and the thermal efficiency (5). The dash-dot line (6) represents the sum of all these terms, and is therefore equal to xa. With respect to the reduction in the thermal efficiency of the PV/T-system in comparison to the conventional thermal absorber, simulations were performed in which the special features of the PV/T collector were successively left out. These features are 1. A lower optical efficiency of 0.744 for the PV laminate applied in the first prototype, instead of 0.89 for a conventional thermal absorber. This is particularly important for long-wavelength irradiation 2. A smaller heat transfer between the absorber (the PV laminate) and the water, as indicated by formula 8 (see above). The value of laea that was found from measuring the temperature difference across the combi-absorber, was approximately 45 W/m2K. Due to this heat resistance, the absorber surface is relatively hot and therefore thermal losses are enhanced. 3. The PV-laminate is not spectrally selective, since glass has a high emission factor in the inflated. This changes the emission of the absorber from 0.12 (for a spectrally selective absorber) to 0.9. This strongly increases radiation losses from the absorbing surface. 4. Due to the additional heat transfer sideways through the silicon (which provides a thermal path parallel to the copper absorber), the heat loss from the collector surface was slightly reduced. However, this effect is very small. 5. Due to conservation of energy, electrical energy can only be produced at the expense of thermal energy. The effect of these features on the efficiency curves is indicated in figure 8. Particularly the spectrally selective layer
SYSTEM EFFICIENCY
4.1 Dailyyield The thermal yield was simulated as a function of reduced temperature by assuming that at each moment the panel is in thermal equilibrium. In these simulations, the top loss was calculated from the empirical formula found by Klein (Duffle and Beckman, 1991, p. 260). In order to test the software program, the daily yield was measured and subsequently simulated. The ambient conditions during the day were those presented in figure 9. The inlet temperature was kept constant.
Fig. 9. Ambient conditions on July 12, 1997, against the hour.
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100
800
.
,
,
.
.
4.2 Annual Yield
.
Next, simulations were performed to find the thermal and electrical yield of the prototype combi-panel for the Dutch KNMI test reference year. The program was used to model the
600
.
~. 4oo
.
.
.
.
.
.
6O
0
5O
20o 0
I
10 600
,
12 ,
I
I
14 16 hour of day ,
|
I
18 .
20
[ 30 o 20
,
0
10
0
500
, 8
|
10
12
~: 400 0
.
,
.
14 16 hour of day ,
.
18 .
20 .
o 300 0
200
60
,~ 100
0
~ 40
I
8
I
10
12
14 16 hour of day
I
18
0
20
Fig. 10. Calculated (dashed) and measured (solid) thermal power for the conventional thermal collector (above) and the combi-panel (below), on July 12, 1997, against the hour. The results of the simulation are presented in figures 10-12. Figure 10 shows a good correspondence between the measurements and the simulations, although the calculated values tend to be slightly larger than the measured values, as found before. In addition, it can be observed that the simulations somewhat over predict the measured thermal efficiency in the morning and slightly under predict the measured thermal efficiency in the evening. It was found that this was due to the effect of the roof tiles, which effectively increased the heat capacity of the system. These results indicate that the assumption of thermal equilibrium in the simulations works quite well. On the basis of these results, it is concluded that hourly data are sufficiently accurate to give a good estimate of the annual thermal yield of the system, especially since the effect of the roof tiles largely cancels over a day. Figure 11 shows the electrical power of the system and figure 12 shows the temperature difference between the PV-panel and the combi-laminate. Figure 11 indicates that the electrical efficiency of the PV-laminate and the PV/T collector are almost the same. At the other hand, figure 12 indicates that in the PV/T-unit the temperature of the PV is much lower than the temperature of the conventional PV unit for the present case in which the inlet temperature was kept constant at approximately 18 ~ This implies that the electrical gains due to cooling of the PV by the water are of the same order as the optical loss of the PV/T-collector, that is due to the reflection at the glass cover.
0
0 i
I
8
10
12
14 16 hour of clay
18
20
Fig. 11. Calculated (dashed) and measured (solid) electrical power for the PV-panel (above) and for the combi panel (below), on July 12, 1997, against the hour.
60
50
L___
3o 20
I
8
10
12 14 hour of day
16
18
20
Fig. 12. Measured temperature of the PV panel (dashed) and the PV/T-laminate (solid), on July 12, 1997, against the hour. case in which two similar combi-panels with a joined area of 3.5 m 2 and a mass flow of 50 kg/(m2 hr) were used to heat a container of 175 litres of water l~om 10 ~ up to 60 ~ A boiler unit was assumed to do the remainder of the heating required if a temperature level of 60 ~ could not be reached by the PV/T unit. The tapping pattern was modelled atter the hot water withdrawal schedule of the ISSO (Institute for Study and Stimulation of Research in the area of heating and air conditioning), which is presented in table 1.
ISES Solar World Congress 1999, Volume III
The thermal and electrical efficiencies were found to be 33% and 6.7% for the configuration used, as compared to 54% for the conventional thermal collector and 7.2% for the conventional PV-laminate under the same conditions. The electrical efficiency was calculated from an efficiency of the PV-laminate of 9.7% at 25 ~ corresponding to figure 4, and an inverter efficiency of typically 92%. Due to reflection of 8% of the incoming light at the glass cover, a thermal efficiency of 0.92 x 7.2% = 6.6% would be expected if the temperature effect on the PV could be ignored. This implies an increase in electrical efficiency of 0.1% of the yearly electrical efficiency due to the temperature effect. Clearly, for the yearly electrical efficiency, the effect of the glass cover is much more important than the effect of the temperature, which largely cancels out over the year. However, if the present collector would have been used for a low-temperature system instead of for the production of hot tap water, the increase in electrical efficiency due to the temperature effect would have been larger, as indicated by the results presented in figures 11 and 12. Finally, the annual thermal and electrical efficiencies were calculated when the special features of the combi-panel were successively removed. The results are summarised in table 2.
Annual thermal efficiency
Configuration
Annual electrical efficiency
Annual thermal efficiency 33.4% 6.7% of the combi-collector Optical efficiency increased 41.0% 6.5% Heat resistance removed 44.7% 6.7% Emission factor reduced 49.9% 6.6% Additional heat transfer sideways 49.6% 6.6% No electricity production 54.4% 0% TABLE 2. Contribution of the various loss mechanisms in the annual electrical efficiency of the combi panel. This table shows that the thermal loss due to production of energy is smaller than the electrical gain. This is due to the fact that the PV is effectively cooling the system by converting irradiation to electricity instead of heat. This implies a small reduction in thermal losses. The thermal loss due to production of electricity is only 5%, whereas the electrical energy produced amounts to 6.7% of the yearly irradiation.
Hour
1
2
3
4
5
6
7
8
9
5
101
CONCLUSIONS
A non-optimised first prototype of a combi-collector was built. From the measurements the thermal efficiency at zero reduced temperature in the absence of the production of electricity was found to be 59%, which is 25% less than the thermal efficiency found for the corresponding thermal collector. The electrical efficiency with electricity production was found to be 54% and the corresponding electrical yield is around 8.3%. From the KNMI test reference year and the ISSO tapping schedule, an annual efficiency of 33% thermal and 6.7% electrical was found if the collector was employed in a domestic water heating system. From the simulations, the magnitude of the factors limiting the performance of the combi-panel can be determined. The reduction in the annual electrical efficiency is mainly due to reflection at the insulating glass cover on top of the thermal collector (approximately 0.6% absolute). The reduction in thermal efficiency of the panel is mainly due to the fact that the glass on top of the PV-laminate is not spectrally selective which increases radiation losses (5% absolute) and the fact that the absorption of the PV-laminate is lower than the absorption of the thermal collector due to reflections in the PV-laminate (8% absolute). This model has proven to be an important tool for further optimisation of the eombi-panel. The results were used to build an improved prototype of the combi panel, which is presently under study.
REFERENCES Duffle J.A. and Beckman W.A. (1991) Solar Rngineering of Thermal Processes, 2 ~ edn, Wiley Interscience, New York. Vries D.W. de, Helden W.G.J. van, Smulders P.T., Steenhoven A.A. van, and Zolingen R.J.C. van (1997). Design of a Photovoltaic/Thermal combi panel momentary output model, outdoor experiment, ISES 1997 Solar World Congress, August
24-30 Taejon Korea. Vries D.W. de (1998), Design of a PV/Thermal Combi Panel,
PhD Thesis Eindhoven University of Technology. Vries D.W. de, Steenhoven A.A. van, Helden W.G.J. van, and Zolingen R.J.C. van, (1999) A panel-shaped, hybrid photovoltaic/thermal device, Dutch Patent 1006838.
1 1 1 1 1 1 1 1 1 1 2 2 2 2 0 1 2 3 4 5 6 7 8 9 0 1 2 3 Tapping . . . . . . . + . . . . + + - - - + + + - + + TABLE 1. ISSO warm water withdrawal schedule, (-) no withdrawal, (+) 175/8 litres withdrawal.
2 4 -
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A COMPARATIVE I N V E S T I G A T I O N OF RADIATION HEAT T R A N S F E R IN TRANSPARENT INSULATION W I T H D I F F E R E N T R E F L E C T I O N M O D E L S
B. Aronov and Y. Zvirin Faculty of Mechanical Engineering. Technion, Israel Institute of Technology Haifa 32000, Israel Tel.: 972-4-8292070
Fax: 972-4-8324533, Email: [email protected]
- The present paper describes a comparative theoretical study of radiation heat transfer in TI structures, with three different models representing the reflection of the TI channel walls: specular, diffuse and a new anisotropic one. The latter assumes that the heat flux which impinges on the wall is reflected uniformly (isotropically), but only in the quarter sphere surrounding the specular direction, with nil reflection in the quarter sphere surrounding the incident direction (zero "back reflection"). For the specular reflection model a 1D ray tracing method is used and for the diffuse one a conventional Discrete Transfer Method is employed. For the third model, the 3D DTM has been modified to accommodate the anisotropie reflection mode described above. The radiation considered here is gray and the intensity reaching the TI structure is taken to be isotropic. The heat flux leaving it on the other side is calculated, i.e. the total transmittance of the TI is obtained. It is assumed, for convenience of comparison between the models, that the channel walls do not re-emit radiation (cold, at 0 K). As expected, the differences between the heat fluxes obtained with the specular and diffuse reflection models are enormous. The values obtained with the anisotropic model are in between, and closer to the specular case. The anisotropic reflection model was used by Aronov & Zvirin (1999) in a simulation algorithm for a solar collector with a TI separator placed between the cover and the absorber plate. The simulation results agree quite well with the experimental data ofRommel & Wagner (1992), which is an indication of the validity of our new anisotropic reflection model. Abstract
transmittance, 2", of the actual wall, and whose effective 1.
emissivity, 13e , is equal to the emittance, E, of the actual
INTRODUCTION
Transparent insulation (TI) has been used in various solar energy applications and systems both passive and active. The TI is made of a multitude of parallel capillary channels of rectangular, round or honeycomb cross section with transparent (glass or plastic) walls, see Fig. 1. Due to the rising spread of TI use, e.g. Goetzberger (1992), it has become apparent that more accurate modeling of the heat transfer inside it is needed, mainly because of quite significant differences between theoretical predictions and experimental data. One reason for this is the reflection characteristics employed in the available models, either specular or diffuse. It is well known that neither is completely correct, and in reality the reflection has components of both. A bundle of radiated thermal energy, when impinges on a surface, is partially absorbed, reflected and transmitted. In case of periodic structure, e.g. honeycomb transparent insulation (TI) as in Fig. l a, the transmission can be taken into account by effective reflection of an equivalent opaque wall. Hollands et al. (1984) developed an approximate model connected with a symmetry of the honeycomb according to the mirror-image technique of Eckert & Sparrow (1961): an opaque wall, whose reflectivity,
Pe , is equal to the sum of the reflectance,
/3, and
wall. The direction at which the effectively reflected bundle will travel from the surface is governed by an associated probability distribution. Surfaces that emit or reflect diffusely have a hemispherically uniform directional probability distribution. Other surfaces reflect specularly and there is also back-scatter reflection, meaning that all radiation is reflected back in the direction from which it arrived. Most surfaces, however, do not reflect in any of these ideal modes, but have an angular dependent reflection distribution in the hemisphere, which is bidirectional. A diffuse approximation is applied in many of the accurate multi-dimensional methods for solving the long-wave radiation problem. Schornhorst & Viskanta (1968), and Herring & Smith (1970) investigated experimentally the accuracy of the purely diffuse model for most common surfaces. These experiments indicated that the reflection of many surfaces of engineering importance was closer to the limit of specular reflection than to that of a perfectly diffuse one. However, applying an accurate specular reflection model is not practical: it requires complicated and time-consuming numerical calculations. This is not justified because the surfaces do not reflect perfectly according to an ideal specular model. Therefore, Aronov & Zvirin (1999) developed a new model, quite similar to the three dimensional Discrete Transfer Method (DTM), e.g. Lockwood & Shah (1981). The new model considers anisotropic reflection (the reflectivity includes transmissivity) of the cell side walls, so that the radiative flux from below is reflected upward only and vice versa. In the
ISES Solar World Congress 1999, Volume III
"classical" DTM, the reflected heat flux is calculated only approximately, while in the method developed by Aronov & Zvirin (1999) with anisotropic reflection, it is computed accurately, by means of iterations until full relaxation is achieved. This model was applied to investigate heat transfer in solar collectors with TI, as a conjugate radiation - conduction convection problem. Comparisons with the experimental results of Rommel & Wagner (1992) showed good agreement. It was demonstrated that the results for the collector performance obtained by using the 'conventional' diffuse reflection (isotropic) model, tend to significantly overpredict the more realistic values obtained by the new anisotropic model.
~
qi
interior surface of the channel, see Fig. lb, while the rest of its surface is cold as before. The different orientations of the emitting surfaces lead to different results both quantitatively and qualitatively.
2. STATEMENT OF THE PROBLEM The transparent insulation (TI) structure is assumed to be of honeycomb type with the same cell geometry parameters in the transverse directions. The TI elements are square in cross section, uniform and equally spaced. This assumption allows to consider the TI element walls as opaque, with an effective reflectivity, ,Oe, given by the sum of the real reflectivity and transmissivity. Further assumptions are as follows. Direct absorption of heat radiation in the air within the TI structure is negligible. Convection in the channels is suppressed. The effective reflectivity of the TI channel walls is independent of the angle. The TI element walls are cold (at 0 K), i.e. not emitting. As mentioned above, three models of TI cell walls reflection are compared: specular, diffuse and anisotropic.
/1/1/1/1/1/ -'~
103
H
Output heat radiation fluxes, qo, transmitted through TI /
/ /
structures of different channel lengths, h, and with different wall absorptivities,
qo
Ew = I-De,
are calculated under
condition of a reference input heat radiation flux, q i, for the two cases mentioned above: aperture and hot thin rim radiation. For the former, q i is determined as an isotropic black radiation that corresponds to a black plate radiation at a given temperature
b"i ./
I "i~ I / e x i t p l a n e
Figure 1. Transparent insulation (TI) structure: a) general view, b) representative ray trajectory in a single channel, emerging from i on the hot rim and leaving at O, after multiple reflections at the walls.
In the present paper, the new model is used to compare pure radiation fluxes through TI structure calculated with three reflection models: for the two above-mentioned models (diffuse and anisotropic) and specular. The simplest way for the comparison is that under conditions of cold TI cell walls (absence of re-emitting). The comparison of radiation fluxes through the TI channel of the cold walls for the three reflection models is carried out for isotropic input heat radiation at one of the channel edges for two cases: 1) the radiation emerges from the whole channel aperture and 2) from a thin rim on the
Ti .
For the case of hot rim radiation, the
temperature of the rim at the upper channel edge is determined by the following condition. The downward rim radiation flux (into the channel) is set equal to the same value as in the previous case (on the cross section plane), and is isotropic too. Maximal absorbing plate temperatures have been calculated by the anisotropic and diffuse reflection models. These were made for a solar collector with a TI structure of glass capillaries, having the same parameters and the same climate conditions as in Rommel & Wagner (1992). In this case, the TI structure was considered approximately as a honeycomb of square cross section channels, with effective channel wall thickness taken based on "aperture conservation". As mentioned above, the results were obtained by means of conjugate heat transfer analysis within the collector. The radiation is considered as semi-gray, with two spectral ranges: short-wave solar radiation (beam and diffuse), and long-wave infrared radiation emitted diffusely by the collector elements. The TI channel walls reflect specularly the short-wave radiation, and diffusely, into a quarter-sphere (according to the anisotropic reflection model), the long-wave radiation: the upward one is reflected isotropically in the upward direction only, within the limits of the corresponding quarter-sphere, and vice versa for the downward radiation. The temperature varies only with the axial direction, z: it is assumed to be uniform across the thickness of the TI element wall and the enclosed air. The side walls of the TI structure are insulated.
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2
3. G O V E R N I N G E Q U A T I O N S
J ; - i b (T,. ); J+- ( N ) - -- pe (N)q~c ( N )
The complete system of equations, which govern the nonlinear stationary problem of the conjugate heat transfer within an element of the TI structure, including the radiation, is presented in Aronov & Zvirin (1999). Thus only the radiation transfer equations for the three reflection models will be presented here. These equations correspond to the following boundary conditions: 1) Cold TI cell walls at 0 K, with the exception of the hot rim area in the corresponding case; 2) The channel under consideration ends in black covers at both of its edges; 3) In the aperture radiation case, the upper black cover temperature is specified corresponding to an isotropic heat flux; in the rim radiation case, the rim temperature is determined by the condition of having the same flux on the cross section plane and the upper black cover is cold; 4) The lower channel cover is cold in both of these cases. For the diffuse reflection model, the radiation transfer equations under these boundary conditions are as follows: 2~
dip d{~cos[g(M), ~]J(N)
q~.c(M) = 0
0
(1) 1 Jw ( N ) - - - Peq,,~
J, - i b (T~ );
point
M
on
the
channel
wall
are incident fluxes at point M,
obtained by integrating the radiosities:
J +(~)
over the lower
hemi- or quarter-sphere for the upper plate or the TI channel walls, and J - ( ~ )
over the upper hemi- or quarter-sphere for
the exit plane or walls, respectively. When N belongs to the upper plate or exit plane, the second of equations (4) includes
q~r
Thus J~(~)
for the direction
~
are
calculated under conditions of isotropic emission at the plate and anisotropic reflection (at the side walls) of the infrared radiation. A specular reflection case, under the above boundary conditions, can be solved in a way similar to that of Platzer (1992) for the insolation, since the heat radiation within a TI channel does not depend on the temperatures in the case of cold channel walls. His approach was used by Aronov & Zvirin (1999). In the case of isotropic 'aperture' input heat radiation, which is transmitted within the channel of cold walls, the above-mentioned calculation must be performed for every
the
~r2,and then the output ||
is incident flux at or
q+~(M), qL (M)
element of a discretized solid angle,
(2) q~,,~( M )
Here
U
(N);
Jo = 0 , where i b ( T i) - oTi 4 / z ,
(4)
exit
plane,
ffi, ffw ( N ) and fro are radiosities on the top plate, the TI channel wall and on the exit plane, respectively, q) and
nux, qo, can be obtained by umming of qo (fi)" For cross section channels, the corresponding equations for
qo(~)
are as follows:
qo ( ~ ) = q,~ ( ~ ) P :
costo
(5)
/
(6)
with n = n x + g/y ,
are the azimuth and polar angles and N in eqs. (1,2) is a point L/
on the other end of the beam ~r2, at its intersection with another channel wall or plate. Similar equations, but for two opposite directions (upward, downward), describe the radiation heat transfer in the case of the anisotropic reflection model:
cosco = sin ~c~
\sintp/
for x and y-oriented walls, respectively,
2g
q ~ (M) =
0
nx -
dip n/2 d~cos[~(M),~]J +(N) (3a)
Int(H tan ~ cos 9),
ny - Int(H tan r sin {p) (7) and H is the TI channel spacing.
2n
qi.r (M) = o
n/2
0
d{~ cos[g(M), ~ ] J - (N)
(3b)
In the case of 'hot rim' radiation, the algorithm for the calculation of qo is the following. At first, for every hot rim control element, N, and a channel wall surface element, M, the fluxes q'+mc( M , N , ~ ) , are calculated as in the anisotropic
ISES Solar World Congress 1999, Volume III
reflection model, see the kernel of the integral in Eq. (3a). Then,
250
for every set of M, N, h , the output fluxes, qo ( M , N , h )
~x Specular reflection model [] Anisotropic model o Diffuse model Aperture input heat flux ----- Hot rim input heat flux
13=0.15
2001
are calculated as in the previous case, Eqs. (5 - 7), with distance z between M and the lower channel edge, and integrated over the angle space.
105
150 o 100
4. R E S U L T S AND D I S C U S S I O N 501[
The radiation heat flux emerging from the TI structure was computed with all three reflection models as a function of the channel aspect ratio, ~ - - h / H ,
and
the channel wall
0
absorptivity, E , for the two cases of emitting surfaces defined above: 'aperture' and thin hot rim. The results are presented in Figs. 2 - 4 and Tables 1 -3. As expected, the differences between the heat fluxes obtained with the specular and diffuse reflection models are enormous. The values obtained with the anisotropic model are in between, and closer to the specular case, especially for ~ < 8
and E ___0.3
for the 'aperture'
case. For the hot rim radiation case, the values associated with the diffuse reflection model are the lowest, again, and tend to those of the anisotropic model with increasing ~
Figure 2.
6
for
8
12
Exit radiation flux,
16
qo,
20
24
as a function of the TI
channel aspect ratio, ~, for the three reflection models and two input radiation cases: 'aperture' and 'rim'. Infrared emissivity of channel walls E -
0.1
5.
and E,
more moderately than in the previous case. In contrast with the aperture radiation case, the output fluxes for the anisotropic mode for the hot rim case exceed those of the specular one at
~: <
4
120
A Specular reflection model [] Anisotropic model o Diffuse model Aperture input heat flux ---Hot rim input heat flux
13= 0.3
100
E < 0.5 and at larger values of ~ for smaller
80
E . This result is explained by the following. For small ~ and small E, the contribution of rays emitted at small angles is obviously more significant in the hot rim case than in the aperture case. In addition, within the anisotropic model these rays are reflected at effectively larger angles, therefore the number of collisions with the channel walls is lower, resulting in lower absorption, especially at low E. Figure 2 shows that for small infrared emissivity, E = 0.1 the exit heat flux,
qo,
O
40q 20
5,
obtained by the anisotropic reflection
4
6
8
10
12
14
16
model, is almost the same for both cases of 'aperture' and 'rim' input heat fluxes. The diffuse reflection model yields close q o values for the two cases, while by using the specular model there is a large difference. This is expected because the orientation has a dominant effect in the specular model. When
Figure 3. Exit radiation flux, %, as a function of the TI channel aspect ratio, ~ for the three reflection models and two input
E is large (Fig. 4, with E = 0.5) the results for qo differ
channel walls E -- 0.3.
radiation cases: 'aperture' and 'rim'. Infrared emissivity of
substantially, for the two cases with each of the three reflection models. As can be seen from Figs. 2 - 4, the anisotropic reflection model yields results for q o which are closer to those obtained by the specular model than by the diffuse one, for small and intermediate values of the aspect ratio, ~. The effect is stronger for the 'aperture' case of input heat flux than for the 'rim' case. These results are mainly due to the smaller number of ray collisions, and the larger view factors at small
The anisotropic reflection model was used by Aronov & Zvirin (1999) in a simulation algorithm for a solar collector with a TI separator placed between the cover and the absorber plate. In that simulation, as mentioned above, the radiation was taken as semi- gray, the solar short- wave and the infrared spectra were calculated separately.
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106
Table 1. Comparison between results for heat flux (W/m2) leaving TI structure, for 'aperture' input heat flux. Impinging radiation intensity: 1000 W/m2 ster, heat flux q i = 500 W/m2; square TI channels, H = 5 mm
h mnl
13 0.15
0.30
Mod
spec anis diff spec anis diff spec
0.50
anis diff
20 ~ =4
30 6
217 204 60 114 94 38 53 43 27
40 8
161 133 27 72 45 15 29 17 11
124 83 12 49 20 7 18 8 6
60 12
100 20
80 16
80 33 4 27 5 3 9 2.8 2.5
55 13 2 17 2 1 5 1.5 1.4
40 5 1 11 1 0.9 3 0.9 0.9
120 24 31 2 0.6 8 0.7 0.6 2 0.6 0.6
Table 2. Comparison between results for heat flux (W'/m2) for the same TI channel and reflection models as in Table 1, for 'hot rim' input heat flux of 500 W/m2 from above in downward direction.
20
30
40
60
80
100
120
6
8
12
16
20
24
76 125 20 27 36 8. 9.5 9.5 4.
52 80 8. 15 15 3.0 4.7 3.2 1.6
27 33 1.6 6.2 3.2 0.6 1.6 0.5 0.4
16 14 0.4 3.1 0.8 0.2 0.7 0.16 0.15
10 5.7 0.15 1.8 0.25 0.10 0.4 0.07 0.07
6.6 2.3 0.07 1.1 0.06 0.04 0.23 0.04 0.04
iilln
e 0.15
0.30
0.50
mod
~ =4
spec anis diff spec anis diff spec anis diff
122 200 55 56 86 27 24 32 15
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107
Table 3. Maximal plate temperatures, Tp,m=, obtained with two reflection models for the collector with glass capillaries of 100 mm height, 6.7mm diameter, 0.1mm thickness, 20 mm air gap, 45 ~ tilt angle with two glass panes of 4rmn, backside heat losses of 0.24 W / m z, at T ~ , = 24.3~
ir Epl
diffuse insolation fraction
insolation of 968 W/m z, Epsol = 0 . 9 5 .
Vw (m/s)
anisotropic model (~
diffuse model (~
1
271
419
3
266 269
417 i416
1 3 1
i264 266
414 i 412
3 1
261 283
410 421
3 1
278 281
419 418
3 1
276 278
415 414
273
i 411
.(%) 10 0.15 15
20 10
0.10
15
20
i
!
3
The simulation results, with ~ -E = 0.5
4('
-
-
A Specular reflection model [] Anisotropic model o Diffuse model Aperture input heat flux Hot rim input heat flux
15 and Eir _ 0 . 1 5 , agree
quite well with the experimental data of Rommel & Wagner (1992), which is an indication of the validity of our new anisotropic reflection model. The maximal absorbing plate temperature measured in the solar collector of Rommel & Wagner (1992) was 261~ They did not specify values of wind
-
speed, V w , solar diffuse radiation fraction and did not take into account temperature dependence of the plate absorptivity, ir E p, in the infrared spectrum (selective coating on the plate). In
0
20
0
Aronov & Zvirin (1999), these were taken as 1 and 3 m/s, 10, ir 15 and 2 0 % of total solar radiation and 0.10, 0.15 for Ep. 4
6
8
10
12
14
16 The same simulation result, Tp,ma x =
261~ C was obtained
for 20% fraction of diffuse insolation, wind speed of Figure 4. Exit radiation flux, qo, as a function of the TI channel
V w = 3 m / s and E p
= 0.15.
A quite close result,
radiation cases: 'aperture' and 'rim'. Infrared emissivity of
Tp,max = 2 7 1 ~
corresponds to diffuse insolation
channel walls E -- 0 . 5 .
fraction of 10%, V w =
aspect ratio, ~ , for the three reflection models and two input
1 m/s
ir
and the same ep = 0.15.
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The difference between the experimental result and a value of
6. R E F E R E N C E S
Tp,max obtained with the diffuse reflection model is drastic: about 150~ Comparative maximal plate temperatures for these models calculated for the same collector and under the same climate conditions as in Rommel & Wagner (1992) are presented in Table 3. The heat transfer coefficient between the cover and the ambient was taken from the correlation ~ g l . a m b - 5 . 7 + 3 . 8 V w , and the 'sky' temperature was calculated by Tsky -- Tam b - 2 0 K . It was shown here that the anisotropic reflection model yields results which are close to those obtained by the specular one (except for large aspect ratios). Furthermore, the results with the former are in good agreement with the experimental data of Rommel & Wagner (1992). This corresponds to the above mentioned observations of Schornhorst & Viskanta (1968) and Herring & Smith (1970), indicating that the reflection of many surfaces of engineering importance is closer to the limit of specular reflection than to a perfectly diffuse one.
5. C O N C L U S I O N S A theoretical investigation was performed of radiation heat transfer in a transparent insulation structure, in order to compare between three reflection models: specular, diffuse and a new anisotropic one, where the impinging radiation is diffusely reflected into a quarter sphere in the direction of the ray. By using the latter model, good agreement was obtained with available experimental results for the maximal plate temperature of a solar collector. The results for the heat flux, emerging from the TI structure, obtained by this reflection model, are quite close to those by the specular one, for the practical cases of small and intermediate aspect ratio of the TI channels.
Aronov B. and Zvirin Y. (1999) A novel algorithm to investigate conjugate heat transfer in transparent insulation application to solar collectors. Numerical Heat Transfer, 7, 757-777. Eckert E.R.G. and Sparrow E.M. (1961) Radiative heat exchange between surfaces with specular reflection. Int. J. Heat Mass Transfer, 3, 42-54. Goetzberger A. (1992) Guest Editorial, Special Issue: Transparent Insulation. Solar Energy, 49, 331. Herring R.G. and Smith T.F. (1970) Surface roughness effects on radiant transfer between surfaces. Int. J. Heat Mass Transfer, 13, 725-739. Hollands K.G.T., Raithby G.D., Russel F.B. and Wilkinson R.G.. (1984) Coupled radiative and conductive heat transfer across honeycomb panels and through single cells. Int. J. Heat and Mass Transfer 27, 2119-2131. Lockwood F.C.. and Shah N.G. (1981) A new radiation solution method for incorporation in general combustionprediction procedures. 18th Int. Symposium on Combustion, the Combustion Institute, 1405-1414. Platzer W.J. (1992) Calculation procedure for collectors with a honeycomb cover of rectangular cross section. Solar Energy, 48, 381-393. Rommel M. and Wagner A. (1992) Application of transparent insulation materials in improved fiat - plate and integrated collector storages. Solar Energy, 49, 371 - 380. Schomhorst J.R. and Viskanta R. (1968) An experimental examination of the validity of the commonly used methods of radiant heat transfer analysis. J. Heat Transfer, 90, 429-436.
ACKNOWLEDGEMENT The Research work has been supported by the Center of Absorption in Science, Israel Ministry of Absorption, to whom the authors are grateful.
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XlX. Solar Hot Water and Thermal Energy Supply
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110
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THERMAL DESTRATIFICATION IN SMALL STANDARD SOLAR TANKS DUE TO MIXING DURING TAPPING
Elsa Andersen and Simon Furbo Department of Buildings and Energy, Technical University of Denmark, Building 118, DK-2800 Lyngby, Denmark.
Abstract-Most small solar domestic hot water systems, SDHW systems, are not equipped with circulation piping. In many systems the pipes, in which the hot water is transported from the solar tank to the draw-off locations, are relatively long. Hence, the waiting time for hot water during draw-off is relatively long. In order to reduce this waiting time to an acceptable level, the flow rate during draw-off is often very large - typically about 20 l/rain. - at least at the start of the draw-off. As long as the flow rate during draw-off is small, the mixing rate inside the tank is small. However, if the flow rate is large, as mentioned above, the mixing rate can be relatively large if the cold-water inlet design is poor. Mixing results in destratification in the solar tank and with that reduced thermal performance of the SDHW system. Investigations indicate that the decrease of the yearly thermal performance caused by mixing during draw-offs can be as high as 23% if a marketed cold-water inlet design is used. Other tested inlet designs result in a decrease of 2-3% of the yearly thermal performance caused by mixing. Based on the investigations recommendations on the design of the cold-water inlet and on a test method for solar tanks concerning mixing during draw-offs are given.
1. INTRODUCTION
During the last decade detailed research has been carried out in order to determine why some small SDHW systems perform better than others. The influence of a large number of design parameters has been analysed and the SDHW systems have been improved according to the findings. However, new investigations show that the thermal performances of small Danish SDHW systems are much lower than expected and that the thermal performances of systems in practice are lower than the thermal performances of similar systems tested in the laboratory, Andersen (1998). Investigations have shown that the hot-water tank is the most important component for small SDHW systems with regard to the thermal performance, Furbo and Shah (1997). Thus, there is a large need to improve the design of hot water-tanks. Almost all hot-water tanks used for small SDHW systems in Denmark are so-called combi hot-water tanks. The domestic water in the combi hot-water tank can be heated both by the solar collectors and by means of an auxiliary energy supply system. The water at the top of the tank is heated to a required temperature by means of the auxiliary energy supply system. In this way it is always possible to tap hot water from the tank, also in periods without sunshine.
For increasing thermal stratification in the solar tank the thermal performance of the SDHW system is increasing. It is therefore very important that the tank is designed in such a way that thermal stratification is built up in the best possible way during operation. This means among other things that cold water should enter the bottom of the tank without any mixing during draw-offs. Most small SDHW systems are not equipped with circulation piping. In many systems the pipes, in which the hot water is transported from the solar tank to the draw-off locations, are relatively long. Hence, the waiting time for hot water during draw-off is relatively long. In order to reduce this waiting time to an acceptable level, the flow rate during draw-off is often very large- typically about 20 l/min. - at least at the start of the draw-off. Consequently, it is important that mixing caused by cold water entering the tank during draw-offs is avoided both at low and at high flow rates. Mixing results in an increased temperature level in the lower part of the tank and with that a decreased efficiency of the solar collector and an increased heat loss from the bottom of the tank. In the higher part of the tank mixing results in a decreased temporamre level and with that an increased auxiliary energy consumption both due to an increased thermal conduction from the hot auxiliary top towards the mixing zone and due to direct
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heating by the auxiliary energy supply system of the water from the mixing zone when this water reaches the auxiliary zone. A marketed Danish solar tank has earlier been investigated, Andersen and Furbo (1998) and Vogelsanger and Frei (1998). The tank suffered from mixing in a relatively large part of the tank during draw-offs with a high flow rate due to a poor design of the cold-water inlet. The mixing caused by a flow rate in the range of 12-20 l/min, reduced the yearly thermal performance of typical SDHW systems based on the tank by 10-14%.
2. T E S T E D T A N K
The tested solar tank is the so-called Danlager 1000 marketed by Nilan A/S. The tank is a mantle tank with both an integrated electric heating element and a heat exchanger spiral as the auxiliary energy supply systems. The data of the tank is given in Table 1. A schematic illustration of the tank with dimensions is shown in Fig. 1.
Amdlla~ Im~ ~ n g m "
Table 1. Data of the tested solar tank. Hot-watertank volumeexclusiveof auxiliaryheat exchangerspiral Volumein hot-watertank belowthe levelof the top of the tappingpipe Watervolumein hot-watertank at a waterpressureof 5.2 bar Volumeabove electricheatingelement Volumeabovelowerlevelof auxiliaryheat exchangerspiral Volumeof mantle Water volumein auxiliaryheat exchangerspiral Tank material Thickness of hot-watertank wall sides top and bottom Thickness of mantlewall Insulationmaterial Thickness of insulationmaterial sides top bottom material Auxifiaryheat exchangerspiral length inner diameter outer diameter Electric heatingelementpower Tappingpipe material length inner diameter outer diameter
1831 1781 1821 801
741 6.81
3.31 St 37-2 3mm 3mm 2mm PUR foam 37-70 mm 75-100 mm 0-30 mm St 37-2 1000 mm 18.5 mm 22 mm 1200 W PEX 1020 mm 16ram
20 mm
Figs. 2 and 3 show schematic illustrations with dimensions of the inlets and photos.
E]echtlc he(dlng element lid
,
140
.~
~ 131 I
. I
p
Oullet
from
amdllary
from
mantle
h ~ ~.mh~mr q ~ i
I~ld
1o
""*'~
n
i . ~ to ~.xnk../ heat ~ d ~ l e r qdml
II
II \ \
- " ~ ' "m' ~. ~' ~ mr
~.,
Inlet
Fig. 1. Schematic illustration of the tested solar tank.
The mantle and the auxiliary heat exchanger spiral were filled with water during the tests. Tests were carded out with the tank with three differently designed cold-water inlets: A marketed PEX pipe with 12 holes in four different levels leading the entering water in three different directions, a horizontal baffle plate and a half ball baffle plate.
.
-
.
IN
/II-]
~
Cold water t
Fig. 2. Schematic illustrations of the three tested cold-water inlets. Note that the drawing deviates from stipulated dimensions.
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Fig. 4. Two temperature profiles at the start of the draw-offs.
4. TEST RESULTS
The measured temperature profiles in the tank aider the drawoffs are shown in Fig. 5. The figure shows the ratio: Fig. 3. Photos of the three tested cold-water inlets.
T - T~old
throughout the tank.
Tmax- - L o l d 3. TESTS
T is the temperature in the tank, ~ T~okt is the temperature of the cold water entering the tank during the draw-off in question, ~
Twelve draw-off tests of the tank were carried out for each cold-water inlet. Tests were carried out with two different volumes tapped from the tank of 20 1 and 50 1, two different temperature profiles in the tank at the start of the draw-off shown in Fig. 4 and three different volume flow rates during draw-off of 5 1/min, 15 l/min and 20 1/min. Temperatures were measured inside the tank in different levels. The cold-water temperature as well as the temperature of the tapped water was measured during the tests.
Tmaxis the maximum temperature in the tank at the start of the draw-off in question, ~ This way of presenting the temperature profiles eliminates small differences in the temperature profiles at the start of the draw-offs and in the cold-water temperatures from one test to another.
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Fig. 5. Measured temperature profiles in the tank after the draw-offs.
The figure shows that the extent of mixing in the lower part of the tank for all three inlet designs is increasing for increasing flow rate during draw-off and for increasing tapped volume. It is obvious that the PEX pipe inlet results in the greatest extent of mixing and that the half ball baffle inlet results in the smallest extent of mixing. Calculations of the temperature profiles after the draw-offs for all 3 times 12 tests were carried out with a detailed simulation
program, Shah and Furbo (1996), in order to determine the exact extent of mixing in the lower part of the tank during the draw-offs. The simulation program was fitted until good agreement between calculated and measured temperature profiles after the draw-offs was achieved. In this way the different draw-offs can be simulated for the three tested inlet designs.
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Fig. 6. shows examples of measured and calculated temperature profiles af~er 20 1 draw-offs with volume flow rates of 20 1/min. for the two start temperature profiles both for the PEX pipe inlet and for the half ball baffle inlet. The extent of mixing occurring in the lower part of the tank during the drawotis is given in the figure. The extent of mixing is much higher for the PEX pipe inlet than for the half ball baffle inlet. It is further observed that the extent of mixing is much higher for start temperature profile 1 than for start temperature profile 2. That is: The extent of mixing is increasing for decreasing start temperatures at the lower part of the tank. The reason is that the temperature difference, and by that the density difference between the incoming cold water and the water at the lower part of the tank, can be so small that even relatively small vertical velocities of incoming water will overcome the natural forces trying to establish thermal stratification in the tank.
5. CALCULATIONS
The extent of mixing in hot-water tanks during draw-offs for a certain cold-water inlet is a function of the flow rate during the draw-offs, the temperature level and the thermal stratification in the lower part of the tank, the cold-water temperature and to a certain extent the volume tapped. It is therefore extremely difficult to simulate the mixing in a correct way. Further, the hot-water consumption pattern as well as the flow rates during draw-offs are not known for SDHW systems in practice. Nevertheless, calculations of the yearly thermal performance of a SDHW system were carried out. The data of the system taken into calculation are given in Table 2. The calculations were carried out with the same extent of mixing for all draw-offs. The extent of mixing determined by the tests with start temperature profile 1 was used.
Table 2. Data for the SDHW system. Solar colletor
2m~
Area Efficiency for small incidence angles Incidence angle modifier Heat capacity Tilt
vI=0.75-5.4.(Tm-T.)/E Ira=l-tan4-2(0/2) 5000 J/K m2 45 ~
Orientation
South
Solar collector loop Solar collector fluid Volume flow rate in the solar collector loop
Pipe length Heat loss coefficient of pipes Power of circulation pmnp
40% propylene glycol/water mixture 0.30 l/rain. 10m 0.25 W/K 30 W
Control system Differential thermostat control with one sensor in the solar collector and one at the bottom of the mantle. Start/stop difference
4K/2K
Solar tank Tested Danlager l O00 with the PEX pipe inlet and half ball baffle inlet
A ~ Fig. 6. Measured and calculated temperature profiles in the tank after 20 1 draw-offs with a flow rate of 20 l/min, and the extent of mixing in the tank during draw-offs for different start temperature profiles and inlet designs.
energysupplysystem
Electric heating element
Set point for the electric heating element Storage ambient air tempemtm~
50.5 ~ 20 ~
The yearly thermal performance of the SDHW system was calculated with hot-water consumption of 50 l/day, 100 l/day and 150 l/day heated from 10~ to 50~ Two flow rates were assumed during all draw-offs: 5 1/min. and 20 l/min. Hot water is tapped either three or six times every day. During each drawoff the same volume of hot water is tapped. Six different drawoff patterns were investigated: Tapping occurred at 7 am, 8 am
ISES Solar World Congress 1999, Volume III
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and 9 am or 7 am, 12 am and 7 pm or 6 pm, 7 pm and 8 pm or 6.30 am, 7 am, 8 am, 8.30 am, 9 am and 9.30 am or 6.30 am, 7 am, 11.30 am, 12 am, 6.30 pm and 7 pm or 5.30 pm, 6 pro, 6.30 pm, 7 pm, 7.30 pm and 8 pm. Fig. 7 shows the calculated net utilized solar energy and the performance ratio of the systems. The net utilized solar energy is the tapped energy minus the energy supply from the auxiliary energy supply system. The performance ratio is the ratio between the net utilized solar energy for the system with the inlet and volume flow rate in question and the net utilized solar energy for the system with the half ball baffle inlet and a flow rate of 5 l/min. The thermal performance of the system is increasing for increasing hot-water consumption. The system with the half
ball baffle inlet performs better than the system with the PEX pipe inlet and the systems with a flow rate of 5 l/rain, perform better than the systems with a flow rate of 20 1/min. If the PEX pipe inlet is used the decrease of the yearly thermal performance caused by mixing is about 3-4% if the flow rate during draw-offs is 5 l/rain., while the decrease is about 23% if the flow rate is 20 l/rain. If the half ball baffle inlet is used the decrease of the yearly thermal performance caused by mixing is about 0% if the flow rate during draw-offs is 5 l/rain., while the decrease is about 2-3% if the flow rate is 20 1/min.
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Fig. 7. Yearly net utilized solar energy and performance ratio as a function of the hot-water consumption for different consumption patterns, inlets and flow rates during draw-off.
6. RECOMMENDATIONS FOR INLET DESIGNS AND SOLAR TANK TEST METHODS
A minimum of mixing during draw-offs can be secured if the cold water enters the bottom of the tank horizontally with a sufficiently low velocity. Further, the velocity of the entering water should be reduced in the tank as much as possible before it hits the side wall of the tank. This can for instance be achieved with an inlet based on a horizontal baffle plate with a large diameter or on a large half ball baffle inlet. Obviously it is easy to design a relatively inexpensive coldwater inlet which only results in a small mixing rate during draw-offs and by that in a small reduction in the yearly net utilized solar energy of SDHW systems caused by mixing, even
for high flow rates during draw-offs. Andersen and Furbo (1998) therefore suggested the following test method to determine if a hot-water tank is suitable as a solar tank: The solar tank is heated to a constant temperature of about 60~ with the solar collector fluid circulating through the solar tank with a constant inlet temperature and flow rate. When steady state has been reached, the flow rate is stopped, and hot water is tapped from the solar tank with a constant volume flow rate of about 20 l/min, and a constant cold-water inlet temperature of about 10~ The tapping is stopped when a new steady state has been reached. During the tapping, the coldwater temperature, the hot-water temperature and the volume flow rate are measured. Based on the measurements the energy tapped from the solar tank from the start of the draw-off until the volume of the domestic water in the solar tank has been tapped is determined. If this energy quantity is smaller than
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96% of the energy of the domestic water in the solar tank in the temperature interval from the cold-water temperature to the temperature of the solar tank at the start of the draw-off, the mixing during the draw-off is considered to be unacceptable and the tank is less suitable as a solar tank. For typical SDHW systems this limit will result in reductions, caused by mixing during draw-offs, of the yearly thermal performance of about 3%.
7. CONCLUSION
The extent of mixing in a marketed solar tank during drawoffs has been investigated experimentally. Three different designs of the cold-water inlet in the tank were tested: The marketed vertical PEX pipe with 12 holes for the entering water, a horiziontal baffle plate and a half ball baffle inlet. 12 draw-off tests were carried out for each inlet.
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Based on the investigations it is recommended to design coldwater inlets in such a way that cold-water enters the bottom of the tank horizontally with a sufficiently low velocity. Further, the velocity of the entering water should be reduced as much as possible in the tank before it hits the side wall of the tank. This can for instance be achieved with an inlet based on a horizontal baffle plate with a large diameter or on a large half ball baffle inlet. It is easy to design a relatively inexpensive cold-water inlet resulting in a small mixing rate and with that in a small reduction in the thermal performance of SDHW systems caused by mixing. Therefore a simple test method with the aim to elucidate if the mixing during draw-off is considered to be acceptable is proposed.
REFERENCES
The draw-off tests were carried out with different temperature profiles in the tank at the start of the draw-off, with different volumes tapped from the tank and different flow rates during the draw-offs.
Andersen E. (1998) Thermal performance of small solar domestic hot water systems in theory, in the laboratory and in practice. The second ISES Europe Solar Congress EuroSun 98 Book of Proceedings Volume 2,111.3.1-1 - 111.3.1-7.
The tests showed that the mixing in the tank during draw-off is strongly influenced by the inlet design. The marketed PEX pipe inlet results in a large mixing, while the mixing in the tank is reduced to a minimum if the half ball baffle inlet is used.
Andersen E. and Furbo S. (1998) Simple characterisation of solar DHW tanks. Status report August 1998. Report SR-9817. Department of Buildings and Energy, Technical University of Denmark.
Further, the extent of mixing in the tank during draw-offs is strongly influenced by the flow rate during draw-offs. For increasing flow rate the extent of mixing is increasing. The extent of mixing is also influenced by the temperature level, the thermal stratification in the lower part of the tank, the coldwater temperature and the volume tapped. It is therefore extremely difficult to simulate the mixing correctly by means of simple simulation models.
Andersen E. and Furbo S. (1998) Mixing during draw-off in tanks for small SDHW systems. Proposal for maximum acceptable mixing. Report SRo9824, Department of Buildings and Energy, Technical University of Denmark.
A detailed simulation model was fitted by means of the test results in such a way that the mixing for the 12 tests for each cold-water inlet was simulated correct. On the assumption that the extent of mixing is the same for all draw-offs during the year, calculations of the yearly thermal performance for a small SDHW system were carried out with the model. The thermal performance of the SDHW system was determined for the different cold-water inlets for different flow rates during drawoffs and for different hot-water consumption and consumption patterns. If the PEX pipe inlet is used the decrease of the yearly thermal performance caused by mixing is about 3-4% if the flow rate during draw-offs is 5 l/min., while the decrease is about 23% if the flow rate is 20 l/min. If the half ball baffle inlet is used there is no decrease of the yearly thermal performance caused by mixing if the flow rate is 5 1/min, while the decrease is about 23% if the flow rate is 201/min.
Furbo S. and Shah L.J. (1997) Laboratory tests of small SDHW systems. 7 th International Conference on Solar Energy at High Latitudes North Sun '97, Finland, Proceedings Volume 1,153-160. Shah L.J. and Furbo S. (1996) Optimisation of mantle tanks for low flow solar heating systems. EuroSun '96 10. Intemationales Sonnenforum Proceedings, Freiburg, Germany. Book 1,369-375. Vogelsanger P. and Frei U. (1998) Simple Chara~terisation of Solar DHW Tanks. Round Robin T e s t - SPF Report. Solartechnik Priifung Forschung, Ingenieurchule Rapperswil ITR, Switzerland.
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INTEGRATED THERMAL IMPROVEMENTS FOR GREENHOUSE CULTIVATION IN THE CENTRAL PART OF ARGENTINA Jorge R. Barral, Pablo D. Galimberti, Adrian Barone Facultad de Ingenieria, Universidad Nacional de Rio Cuarto, Ruta Naciona136 km 601, Rio Cuarto, C6rdoba, 5800, Argentina, Telephone 54 358 4676243, Fax 54 358 4676246, E-mail [email protected]
Miguel A. Lara Instituto de Fisica Rosario, Conicet-Universidad Nacional de Rosario, Blvd. 27 de febrero 210 bis, Rosario, Santa Fe, 2000, Argentina, Telephone 54 341 4495467, Fax 54 341 4853200, E-mail [email protected] A b s t r a c t - A system to improve winter horticultural cultivation inside greenhouses has been developed in the central part of Argentina. Previous experiences showed that it is possible to avoid the harmful freezing effects using a combination of thermal curtains and heating tubes in which warm water from a geothermal source flows. In spite of these improvements, in many winter nights the temperature tends to remain low inside the greenhouses, causing plant growth detention. This problem was prevented by concentrating the energy delivered by the tubes as near the plants as possible. This was achieved adding a transparent and light synthetic blanket to the previous improvements to decrease convective heat transfer effects. The results were evaluated measuring temperatures at different places of the prototype. Prototype data of winter nights were plotted and compared with temperatures of other greenhouses. The system showed to be highly cost efficient. It maintained satisfactory temperature levels in the surroundings of the plants, allowing continuous growth of the cultivation; the raw material necessary for these improvements is a cheap common market material; the arrangement of all the system does not take more than a few hours, and the operation of the system is simple and no time consuming.
1. INTRODUCTION Horticultural Greenhouses are used to grow plants in quantity, with high quality and at a good timing. In a greenhouse it is possible to use the soil intensively, to create favourable environmental conditions, and to maintain an effective sanitary plant control, which results in massive production of high quality. The extension of good thermal conditions for the cultivation allows to obtain crops before and after the normal seasons, which results in higher prices for the product. The adequate environmental conditions for each type of cultivation are achieved controlling a set of variables, from which the most important are light, humidity, and temperature (Albright, 1991). For the central part of Argentina, the photoperiod of many typical horticultural products is not a crucial variable in winter. Moreover, the climate in winter is generally dry and sunny, which causes no problems with the greenhouse ambient humidity, since the warm and dry daytimes allow the ventilation of the greenhouse to deliver the moisture accumulated during the night. Therefore, the critical variable to be controlled is the temperature, especially at night, and, for that reason, heating systems are used. The use of fossil fuel is not recommended because of its contaminant effects, high prices and fastidious handling. The normal use of fossil fuels is their burning during the cold nights, which contaminates the inner greenhouse ambient at night and the environment globally. Moreover, it obliges the farmers to take care of the burning process and to have some transport facilities to carry the fuel. Many attempts have been made to provide thermal energy to horticultural production greenhouses by means of renewable energy (Santamouris et al., 1994). Most of the works are based on the accumulation of energy during the day, delivering it at nights and on the energy conservation. The accumulation of energy in water is widely used, using plastic tanks or tubes placed on the greenhouse floor, which
absorb solar energy during the day and transfer heat at night to the greenhouse ambient. The main problem of this kind of systems is that better accumulation of energy is obtained when large volumes of water are used (Santamouris et al., 1994). On the other hand, it is possible to use plastic surfaces to transfer heat to the greenhouse ambient, taking the energy from a source (Saravia et al., 1992). To avoid energy losses, mainly at night, systems of thermal curtains are used, which are basically made of polyethylene of low quality (Chandra & Albright,1980) (Seginer & Albright, 1980). Taking into account these elements, a heating and energy conservation system was designed and tested in a farm in the central part of Argentina. The system took advantage of a natural artesian spring, a geothermal source of low entalphy (Dicksonand Fanelli, 1995), which provides warm water at constant temperature of 28 0(3 to the farm without pump requirements. First, black polyethylene tubes were arranged on the greenhouse floor, allowing the warm water from the geothermal source to flow during nights (Galimberti et al., 1997). This arrangement has the advantage, compared with the accumulation systems previously described, that small volumes of water are necessary because the energy is provided permanently by a source at constant temperature. Second, inner thermal curtains were added to the walls and roof to decrease the energy losses by convection heat transfer effects (Adaro et al., 1997). The system was tested during two years, showing to be very effective to prevent the freezing problems the greenhouses used to be. Moreover, it is used today in normal production greenhouses in the farm (Adaro et al., 1998). However, in many winter nights the temperature inside the greenhouse can be so low that plants of many typical cultivations delay or completely stop their growth. As a consequence, they occupy the greenhouse, need watering and some horticultural tasks to live, but they do not bloom or fructify, which is expensive and useless. But, on the other hand,
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if it would be possible to maintain the temperature above some values during all the nights, an out of season crop would be harvested, allowing to the farmers to get a high price for their production in the local market. The use of hermetic polyethylene low tunnels for small plants showed to be very effective to avoid energy losses at night, in greenhouses in the East-Central part of Argentina (Francescangeli et al., 1994). The arrangement of the tunnels is really simple: each of them is applied over the furrow covering a line of plants. Then, the convection heat transfer is strongly reduced and microclimatic conditions are created near to the plants. This work describes the design of an assemblage to achieve acceptable temperature levels near the plants in winter to allow them to grow, bloom, and fructify continuously. It is based on the system of polyethylene tubes and thermal curtains previously described, and the addition of thermal blankets over the plants resembling the operation principle of the low tunnels. The results were highly satisfactory and the response of the system for a typical winter night is shown through the comparison of temperatures measured in a prototype with thermal blankets and in another greenhouse with only tubes and curtains. First, a brief physical description and results of the old system is made; then, the addition of the thermal blankets is analysed and its results shown. 2. DESCRIPTION EXPERIENCES
AND
RESULTS
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experiences were developed on one of the main production greenhouses (1000 m2), on a prototype built by the UNRC (105 m2), and three production greenhouses (105 m 2 each one). The section of the greenhouse is shown in Fig. 1 a), where the mentioned polyethylene tubes (110 mm diameter) and the thermal curtains are represented. Cheap common market materials were selected for the heating system and a simple construction methodology was conceived to build it, in order to minimise labour hours and avoid specialised employees. The tubes were made of 200 lain thick black low density polyethylene and had a diameter of 10 cm. They were placed on the greenhouse ground, between furrows, allowing heat transmission to the air and the ground simultaneously. The tubes were connected in series with only one inlet and one outlet for each greenhouse, as shown in Fig. 1 b). The water flow was controlled manually: the valves were opened before sunset, allowing the water to flow through the greenhouse tubes, and closed the next morning, when the sun began to warm the greenhouse. The greenhouse exit water was used for watering farm open sectors, which is convenient in winter, when the rain is scarce.
PREVIOUS
2.1 Motivation The Solar Energy Group of the Engineering College at National University of Rio Cuarto (UNRC) began its work on greenhouses subject in 1994, when the owner of a farm exposed his freezing problems to the group. The farm, called SIQUEM, is situated 10 km away from the UNRC, 33.2 ~ S latitude and 64.3 ~ W longitude, and has a low temperature geothermal source (ASHRAE, 1995), which provides underground water at 28 ~ The water flows freely and no pump is necessary for the farm consumption requirements. SIQUEM and UNRC signed an agreement to work conjointly on the problem; the farm basically provides some materials and the standard greenhouse labour and the University performs research work using its measurement equipment, providing new materials to be tested, and labour hours to arrange and develop the experiences. The analysis of the problem began with the selection of some materials for the tubes and thermal curtains, which were tested in laboratory and in the greenhouses of the farm. Then, using the results of these first experiences, the regional climatic data 03arral et al., 1995), and geometrical and physical information, some energy balances were approximated, which made the researchers think it would be possible to prevent the freezing problem with the arrangement described in the next point. 2.2 Physical description and handling of thefirst system The greenhouses used in the experiences are called "chapel greenhouses" because of their appearance. They are typical constructions in the central part of Argentina and they are made of wood beams and polyethylene walls and roof. These first
Fig. 1. a) Section of the greenhouse prototype; b) Top view of the greenhouse prototype The thermal curtains were made of 50 larn thick transparent polyethylene with no special optical properties since their main function was to decrease heat losses by convection and infiltration and they were not exposed to the exterior climate adversities. The ceiling was not moved during all winter and the double walls were opened and closed for ventilation in the same operation made for the simple walls.
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2.3 Experimental tasks performed The temperatures in different places of the greenhouses were measured with data acquisition systems, every 15 minutes, during all the winter season. Other variables, like relative humidity and other temperatures, were measured by means of a portable weather station placed inside the greenhouse whereas water flow rate and water pressure were determined manually. A pyranometer was used to measure global solar radiation inside the greenhouses and outside. An agronomist of UNRC worked in the determination of productivity and other aspects related to the biological behaviour of the plants. The deterioration of the materials used in the system was checked visually. The economical aspects were directly evaluated by the farmers from the results of production selling in the local market. 2.4 Results and conclusions of these previous experiences The following results can be summarised aider the experiences (Energex): a) The system is cheap and simple to operate. b) The materials of tubes and curtains did not suffer major deterioration, which allowed them to be used for other winter seasons. (puncture of tubes, care of the farmers) c) Since the thermal curtain is transparent and very thin, it did not cause a significant decrease on the solar radiation absorption of greenhouses. d) The two improvements were necessary to prevent freezingeffects in hard winters. e) The night temperature is a decisive factor in the plant growth acceleration. 0 It is possible to obtain a good early production and extended winter production. g) It would be possible to obtain a great out of season production with a few more degrees.
3. ASSEMBLY AND OPERATION INTEGRATED IMPROVEMENTS
OF
THE
After these previous satisfactory experiences for freezing and noting that out of season production was possible, the Solar Energy Group decided to test a system of thermal light blankets to minimise the heat losses in the vicinity of the plants. Since the greenhouse floor works as the accumulator of energy, the placement of these blankets cut the convection currents in contact with the floor surface and reduce them to a small circuit near to the plants. Then, not only is the heat provided by the tubes concentrated, but also the energy absorbed by the greenhouse ground on sunny days is retained to some extent. Considering that the temperature inside the tubes is almost constant and the heat transfer would be increased if greater surface of transmission were used, bigger tubes were adopted for this experiment. In addition, tubes with greater diameters have thicker wall thicknesses, which diminishes the puncture problem during the normal f a m ~ g labour. Figure 2 shows the section of the greenhouse provided with the thermal blankets. These thermal blankets are made of transparent synthetic material, they are light (17 g/m2) and not impervious to gases. The mounting process of the blankets is easy and it is no time consuming. The light synthetic blankets are simply supported by the same wires used to guide the plants and are maintained in position by means of clothespegs, avoiding displacements during daytime ventilation operations. Also, the blankets are set in such a way as to allow the farmers to walk along the greenhouse during normal inspection tasks. Since the blankets are transparent, they do not need to be moved if no horticultural work is necessary, and if they must be moved, they can slide effortlessly along the supporting wires.
Fig. 2. Section of the greenhouse with tubes, thermal curtains and thermal blankets Figure 3 shows the arrangement of the thermal blankets in a top view, in this case for a greenhouse covered by six panels. It also shows how the tubes are connected at the ends by PVC pipes, which was a progress from the old system, taking into
account the pathways of the farmers, where the damage for the tubes was more likely. It is important to remark that the blankets are going up following the growth of the plants, and although the convection effects become greater with this growth, the coverage of the
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plants can be maintained in a good degree if the total width of the blankets is foreseen taking into account this change on plant highnesses.
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it can be seen that there are no more than two degrees between the inlet and the outlet, which supports the previously stated concept about constancy of the heat source The temperatures for the greenhouse without thermal blankets are plotted in Fig. 5 for the same winter night selected for the prototype. Comparing the two greenhouses it can be concluded that for this cold night, under the thermal blanket the temperature maintained values over 13 ~ for all the night, while for the other greenhouse, the temperature was under 11 ~ for more than 7 hours. The temperature for cultivation of tomatoes and peppers, two of the most common products in this part of the country, must be maintained over 12 or 13 ~ if continuous growth and fructification is desired. Then, although the differences may seem to be small, they are crucial to define a profitable or not economical production
3. CONCLUSIONS
Fig. 3. Top view of the blankets and tubes arrangement 4. RESULTS The results of the experiment were evaluated from the measurement of temperatures at different locations of the prototype and in another greenhouse furnished only with thermal curtains. The cultivation adopted for the experiment was green pepper. Figure 4 shows a plot of the temperatures in the greenhouse with thermal blankets for a typical winter night. The temperature of the water inside the tubes is also plotted and
The system proved to be very efficient to provide the temperature levels required for the continuous growth of some important typical horticultural cultivations. It is important to remark that this thermal improvement does not imply structural changes in the greenhouses. It is necessary to think only in some wires to support the blankets for those plants that do not require wires to be guided and supported, which is not the case for tomatoes and peppers. Although the duration of the blankets is not yet tested, they are cheap, and the time to perform the mounting is very short. It is also no time consuming to open the clothespegs and slide the blankets over the wires when horticultural tasks are required; specialised labour is demanded for this task. Since they are transparent and impervious, they cover the plants during the day without problems and in permanent form for many days. The system does not present heat regulation problems. The valves just have to be opened before night and closed the next morning. Since the temperature of the resource is low, there is no problem of overheating for the typical greenhouse cultivations. A natural artesian spring is required to arrange the complete thermal system which is a restriction for its application. However, in the region where the experiments have been done, there exist a large geothermal field of low enthalpy. This become the system to be high promising for the local market. The next steps in this work is to select different kinds of horticultural products, with different seedtimes, in order to determine which is the most profitable methodology for the out of season production using thermal blankets. To do that it must be taken into account that the behaviour of the plants is not the same, mostly in times of bloom and fructification, than those times in normal season production. The study will imply a degree day analysis to determine the energy levels required for each kind of plants
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Fig. 4. Temperature measurements in the experimental greenhouse
Fig. 5. Temperature measurements in the greenhouse without thermal blankets.
REFERENCES
Adaro, J., Galimberti, P., Lema, A., Barone, A. and Fasulo, A., (1997), Calefacci6n de invemaderos con energias renovables, International Journal of Environmentally Conscious Design & Manufacturing, Vol 6, No. 4, pp. 3-7.
A.Adaro, P. Galimberti, A. Lema, A. Fasulo, J. Barral, (1998), Geothermal Contribution in the Greenhouse Heating In Proceedings of 7th International Energy Conference and Exhibition (Energex'98), Manama, Bahrain Albright, L. D., (1991), Production Solar Greenhouses, In Solar Energy in Agriculture 03. F. Parker Eds.). New York: Elsevier.
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ASHRAE, (1995), Ashrae Handbook, HVAC Applications, American Society of Heating, Refrigerating and AirConditioning Engineering, Atlanta. Barral, J. R., Adaro, J. A., Lema, A. I. and Fasulo, A., (1995), Variables clim~iticas en la regi6n centro sur de C6rdoba. In Proceedings of ASADES 95 (1), pp. 04.67-04.72. Chandra, P. and L. D. Albright, (1980), Analytical Determination of the Effect on Greenhouse Heating Requirements of Using Night C-klrmins. Transactions of the ASAE 23(4), pp. 994-1000. Dickson M.H. and Fanelli M. (1995) Geothermal Energy, pp. 14-15. John Wiley & Sons, Inc., Chichester. Francescangeli, N., F crrato, J.,Levit, H., and Lara, M. A., (1994), Comportamiento de distintos materiales opaeos y transparentes en la cobertura de tdneles bajos, en el interior de invemaderos, durante el invierno. Rivista di Agrieoltura Subtropicale e Tropicale, vol 88, n ~ 3, pp. 529-538. Cralimberti, P., J. Adaro, A. Lema, A. Barone, L. Grosso, and A. Fasulo, (1997), Estudio Comparativo de Diferentes Mejoms en Invemaderos Avances en Energias Renovables y Medio Ambiente 1(1), pp. 21-24. Santamouris, M., C. A. Balaras, E. Dascalaki, and M. Vallindms, (1994), Passive Solar Agricultural Greenhouses: A Worldwide Classification and Evaluation of Technologies and Systems Used for Heating Purposes. Solar Energy 53(5), pp. 441-426.
Saravia, L., Echaz~ R., Cadcna, C., Cabanillas, C., (1992), Calentamicnto solar de invcmadcros cn la provincia de Salta, Proceedings of A S A D E S 92 (I),pp. 04.67-04.72. Seginer, I., and L. D. Albright, (1980). Rational Operation of Greenhouse Thermal- Ou'tains. Transactions of the ASAE 23(5), pp. 1240-1245.
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IN SITU SHORT TERM TEST PROCEDURE FOR LARGE SOLAR THERMAL SYSTEMS Th. Beikircher, N. Benz, M. Gut, P. Kronthaler, C.Oberdorf, W. Schflkopf Bavarian Centre for Applied Energy Research (ZAE Bayem), Division: Solar Thermal and Biomass, Domagkstr. 11, D-80807 Munich, Germany Phone: +49-89-356250-0, fax: +49-89-356250-23 email: [email protected]
H. Drfick Institut fiir Thermodynamik und W[innetechnik (ITW), Universit~t Stuttgart Pfaffenwaldring 6, D-70550 Stuttgart, Germany Phone: +49-711-685-3536, fax: +49-711-685-3503 email: [email protected] Abstract - A short-term test procedure for large installed solar thermal systems has been evolved and validated. The developed ISTr-method Cm-situ _short term test) makes it possible to determine the yearly energy output of a large installed solar thermal systems under arbitrary standard operation conditions especially for those underlying the guaranteed _solar results (GSR). A detailed description of the different steps of the ISTT-method is given. For data acquisition an autarkic wireless measuring station for the meteorological quantities has been developed. New surface temperature sensors have been constructed and ultrasonic volume and magnetic inductive gauges have been applied by special adapters to record the volume flow rate in a non invasive way. The method is applied and validated for the example of a standard solar thermal system with 110 m 2 flat-plate collector area and store with a nominal volume of 6 m 3. For the dynamic evaluation of the measuring data, the transient behaviour of the system is modelled on the basis of detailed component models (e.g. Trnsys). Using the data recorded during a system operation period of 4 to 6 weeks, parameters for the most important components, such as the collector and the store, are determined by means of parameter identification. On the basis of the parameters determined for the single components, a simulation model for the whole system was created, in order to predict the yearly energy output of the solar system. For the system investigated, the difference between the energy output as predicted by the ISTT-method and the measured energy output integrally was found to be lower than +_3% for the energy delivered from the collector loop to the buffer store (GSR1) as well as for the energy delivered from the buffer store to the load (GSR 2). Also the monthly predictions commonly showed errors below _ 5%.
1. INTRODUCTION To raise peoples confidence in solar thermal systems, a reliable and inexpensive method is needed to control the performance after the installation. A suitable figure for the thermal performance is the yearly energy output delivered by the system for standardised reference conditions. In Germany it recently has become common, that the planner and builder of a large solar system must warrant a certain yearly energy output (ASEW, 1998), the so called guaranteed solar result (GSR). Commonly, the GSR is determined either as the amount of energy delivered from the collector loop to the buffer store (GSR 1) or as the amount of energy delivered from the buffer store to the consumer loop (GSR 2). In order to check whether the system performs as projected by the planner or builder, the GSR is compared to the actual yearly energy output Q~p (Q1 and Q2 for the actual yearly energy output after the collector loop and the buffer store, respectively) of the installed system. Q~p depends on both the performance of the system itself and the boundary conditions such as the weather and the hot water demand during the measurement. Therefore the same boundary conditions have to be used when comparing the GSR with the experimental result Q ~ . The boundary conditions can either be the ones underlying the design of the system or the real ones measured during the operation of the installed system. So far, only the
second approach is applied leading to cost intensive long term conventional monitoring over at least one year. To objectively compare Qc~p and the GSR as guaranteed by the planner, in this approach the GSR is recalculated under the real operation conditions, which may cause problems because the simulation program is commonly not exact for the real system investigated. Moreover, from the juridical stand of view, the planner has only to gum-antee the GSR for the conditions assumed during the planning process, which value actually can be checked only by the new ISTr-procedure described below. Therefore these it is difficult to separate the influence of the operation conditions from the system efficiency and it is impossible to determine the system output for any other standard operation conditions, for example supposed by the planner. By comparison, the yearly solar energy yield can be determined in a drastically quicker and low-cost way, fitting a component based system simulation model to measured data over a short period (4-6 weeks) and calculating the long term energy yield with the help of the adapted simulation model for the standard operation conditions as adopted by the designer. At the ZAE Bayern in co-operation with the ITW Stuttgart such a short-term test procedure, the so called _in-situ _short t_erm t_est method (ISTT-method) for large installed systems, has been developed. It promises an accuracy of better than 5% for the yearly energy output delivered under reference conditions by
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the collector loop (GSR1) and by the solar buffer store (GSR2), respectively. The costs nearly can be halved, compared to the long term measurement method applied so far. 2. DESCRIPTION OF THE ISTT-METHOD The ISTT-method aims to determine the yearly energy output of an installed solar thermal system for the boundary conditions of the design case on the basis of short term measurement which are used to identify the system performance. Therefore, Qexpcan be directly compared with the GSR as calculated by the planner.
9
in-situ measurements (duration 4 - 6 weeks) I1
11
sensitivity analysis of yearly system gain (variation of component parameters)
11
11
11
determination and verification of component parameters
validation of complete simulation model for the whole system
prediction of yearly system gain for reference conditions Fig. 1: Structure of the ISTT-method
Since solar thermal systems show a large variety in their design, it is nearly impossible to give a detailed recipe for the treatment of each single system configuration. The ISTT-method represents more a general procedure that provides a guideline for the determination of the performance of installed systems. The ISTT-method can be subdivided into five steps which are shown in Figure 1.
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2.1 ln-situ measurements As the basis for the parameter identification in-situ measurements have to be carried out over a period of 4 to 6 weeks. Details about the measuring technique and the location of the sensors se chapter 3. During the main part of this time the system is operated as usual under normal conditions. However, in order to drive the system into some extreme operation conditions needed for a reliable parameter identification, the control strategy of the system is disabled for a short period. 2.2 Sensitivity analysis A sensitivity analysis has to be carried out in order to get an impression how the system output is influenced by the single component parameters. The sensitivity analysis should be carried out with a component based system simulation program having a modular structure, such as Trnsys (Tmsys, 1994). The results of the sensitivity analysis provide an important basis for the decision which parameters have to be determined by means of parameter identification. 2.3 Determination and validation of component parameters In the ISTT-method, each component of the solar thermal system is described by a detailed parametric model. The parameters required to describe the thermal behaviour of the component in an optimal way are determined by means of a numerical parameter identification procedure (fit). Conventional numerical procedures are basing on the Levenberg-Marquardt algorithm like the commercial program package DF by insitu Software (Insitu, 1996) or Tmspid by TransSolar (TransSolar, 1997). The reliability of the parameters and the best set of parameters are determined by means of cross predictions: The total data sequence is divided into subsequences and the parameters are identified for any subsequence. Subsequently the subsequences which were not used for the parameter identification are predicted. The relative difference between prediction and measurement is an indicator for the goodness of the parameters determined, the best set identified by the lowest deviation over all subsequences. The determination and validation of the best parameter set is finished, when a certain part of the subsequences (for example 68%) can be predicted with a deviation below -I-5%. In this case, the insitu-measurement can be stopped. 2.4. Creation and validation of system simulation model Having determined all essential parameters of the components, the next step of the ISTT-method is the numerical simulation of the thermal behaviour of the whole system. Here, it is essential that the same component models which were used for the determination of the component parameters are now used in the system simulation model. If the simulation program used for the parameter identification is identical with the program used during the design process of the system, the simulation of the whole system is easily performed replacing the parameter values used for the design process by the values identified from the in-situ measurements according to step 3 of the ISTT-method. If a different simulation program was used during the design process, a completely new numerical model for the whole system has to be created on the basis of the component models used in step 2 of the ISTT-method.
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The validation of the system model is carried out in analogy to the way as described for the component models by means of cross predictions. Ideally, the model is validated at sequences containing all possible states of the system occurring under the real operation. 2.5 Prediction of system performancefor reference conditions The prediction of the system performance is carried out using the validated system simulation model from the previous step. The yearly system energy output Qoq, is calculated using the boundary conditions of the design case (reference conditions) and the parameters determined for the installed system (real parameters). A comparison of Qoq, and the GSR (the system output predicted for the design boundary conditions and the design parameters) can be used as a basis to judge the fulfilment of contracts related to guaranteed solar results. 3. NON-INVASIVE MEASURING EQUIPMENT To meet the requirements of a cheap, short-term and not invasive measurement, an autarkic wireless measuring station for the meteorological quantities has been developed, new surface temperature sensors have been constructed and ultrasonic volume gauges have been applied to record the mass flows in a non invasive way. The mobile meteorology station with photovoltaic power supply records the total radiation and wind speed in the collector plane as well as the ambient temperature. The station applied for a patent is mounted directly on the collector field via a acrylic-glass carrier supplied with suction-cups evacuated by a small vacuum membrane-pump, see Figure 2. By this kind of fixing, it is ready for operation within a few minutes and the cumbersome adjustment of the radiation and wind speed measuring devices can be avoided. The accuracy is the same as for conventional meteorological measuring systems with fixed installation. The weather data is collected and stored by an integrated data logger and is sent wirelessly via GSM900 mobile net to the central evaluating computer at the ZAE Bayem.
10 s and 18 s in the case of a temperature jump of 20 K for a Cu-pipe DN 20 and a steel pipe 90 respectively) compared to conventional sensors immersed in the fluid. To non-invasively measure the fluid volume flow rate, ultra sonic sensors have been applied, where the delay of a sound wave by the fluid motion is a measure for the flow velocity in the pipe. In the laboratory we tested the ultrasonic device of the German manufacturer Flexim, Berlin, for pipes between DN 25 and DN 40, which is typical for large solar systems and gained sufficient results for copper (accuracy 3% against magnetic inductive (MID) flow meters for volume flows rates between 750 and 2000 l/h), while for black and zinced steel the method is still not applicable (deviations up to 8%). A considerable amount of development has still to be spent to make the ultra sonic principle applicable to all pipe diameters and materials used in solar systems especially for small volume flow rates down to 100 l/h. In the solar system investigated (see also chapter 4), the volume flow rates ranged between 1000 and 1500 l/h and the ultrasonic sensors showed good agreement (1,25%) to conventional mechanical sensors of the German manufacturer Aquametro used by the ZfS, Hilden even in the case of black and zinced steel pipes. As an alternative, we developed a coupling device, which easily can be mounted in one of the dirt pans, which normally exist in large solar thermal systems. By these devices, the fluid can be fed into a mobile MID-sensor to achieve highest accuracy without disturbing the normal system operation, see Figure 3a and 3b.
Fig. 3a Fig. 3b
Fig. 2: Mobile meteorological station. We developed also a new clamp-on surface temperature sensor, which consists of a steel armed Ptl00 soldered in a thin silver plate. The response time of the sensor was investigated in laboratory experiments and proved to be sufficiently quick (t99oa =
Fig. 3: The coupling device mounted in a conventional dirt pan (3a) with the connecting tubes and the MID-sensor (3b) for precisely and non invasively measuring the fluid mass flow.
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4. SYSTEM INVESTIGATED With the described experimental equipment, we carried out different periods of short term measurements at a large solar hot water system built up within the scope of the BMBF-program Solarthermie 2000 in Munich with 108 m2 collector area and a buffer store with a nominal volume of 6 m 3 , see Figure 4.
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The onset of the solar loop is controlled by a radiation limit (E>150 W/m2), while the pump of the store charging loop is operated depending on the temperature difference between the collector outlet and the lower part of the store (4 K upper and 2 K lower dead band for the design case). The system investigated showed strong oscillations in the on- and offset of the store charging pump with a period of 5 min and below. 5. REALISATION AND VALIDATION OF THE ISTTPROCEDURE In order to determine the collector loop parameters we carried out 2 short term measuring sequences: 9 29.10.97-8.11.97, 11 days under normal operation
Fig. 4: The system investigated with the measuring equipment. The demarcation line for the GSR is either the energy delivered from the collector loop to the store (GSR 1) or the energy delivered from the store to the consumer loop or load loop respectively (GSR 2). The collector field consists of normal, single-glazed flat-plate collectors. The store is charged and discharged via external heat exchangers in a direct way without using any special designed stratification devices. The ratio of store volume to the collector area amounts to 56 1/m. The collector field is normally operated at relatively low temperatures (below 60 K over ambient), which can be read from Figure 7 showing half hour mean values for the total radiation in the collector plane against the collector operation temperature referred to ambient that.
Fig. 5:
9 24.-25.3.98, 2 days with strongly reduced discharging of the store leading to high collector temperatures up to 90 K over ambient. For the store loop we measured the normal operation over ten days (21.4.-30.4.98). For the further dynamic evaluation procedure, we modelled the system in the simulation programme Trnsys using the multiport store model (IEA, 1997) and an extended matched flow collector model (Isakson, 1995) including pipes and a heat exchanger (Spirkl et al., 1997). Subsequently, a sensitivity analysis (changing the design values according to table 1 by 10%) was carried out yielding that the collector parameters are influencing the yearly solar output in a much stronger way than the store parameters, see Figure 5. Following the results of the sensitivity analysis, for the store only the parameters (UA)~,a,, the overall heat loss capacity rate [W/K], Zs,~ and Zs,o~t, the relative height of inlet and outlet position of the solar loop as well as ZL,~, the relative height of inlet position load loop have to be numerically identified. The other parameters (store volume Vs, relative height of the load loop outlet ZL, o,,t, vertical effective thermal conductivity in the store 2~n) show only a small influence on the system output and have been fixed to the values as given by the producers or from
Sensitivity analysis of the system investigated.
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Table 1 design- and insitu-parameters for the collector loop and the store. p arameter
(UA) Zs,in Zs,out W/K
rn
m
ZL.in
'rl o
rn
[.]
01
02
W/mZK W/mZK=
f [-]
Ccol
Lp
up
6
k.l/Km=
m
w/maK
F]
/
insitu
42,4
2,4
0,0
1,0
0,84
2,72
0,0146 0,475 7,01
74
1,66
0,81
design
21,7
1,05
0,0
0,0
0,78
3,67
0,013
30
,29
0,81
test results. For the collector loop the following parameters have to be identified: the collector capacity C~b the optical efficiency r/0, the coefficients describing the constant and the linearly temperature depending part of the thermal loss coefficient, UI, U2, Up, the loss coefficient of the pipes and the angel modifier f. For the remaining parameters (as the upper dead band controlling temperature difference for the solar loop, A Ton) the design values can be taken. Additionally the efficiency e of the heat exchanger and the length of the pipes Lp, were numerically identified from separate experiments not exceeding one day. The wind depending part of the loss coefficient has been set to zero, due to low yearly mean wind speed at the location the system is mounted. Table 1 shows the parameters identified from the insitu-measuring sequences as well as the design parameters, on which the sensitivity analysis was based. With the insitu-parameters according to table 1 which were identified from the short term measuring sequence, the energy delivered from the collector loop to the buffer store (GSR 1) as well the energy delivered from the buffer store to the load (GSR 2) was calculated in 5 min-steps for the real operation conditions between March and August 1998, which were monitored by the ZfS Hilden. The calculated results were compared to the experimental solar output as measured by the ZfS Hilden: Over all six months, the deviations for the GSR 1 as well as for the GSR 2 were below 3% (integrally) and below 4% (monthly), respectively. 6. GUARANTEED SOLAR RESULT With the help of the now validated ISTT-procedure, it is possible, 1. generally to calculate the yearly solar output for arbitrary operation conditions (weather and load) and 2. especially to check the GSR 1 and GSR 2-values guaranteed by the planner for the standard operation conditions as used during the planning process. For the system studied in this paper, the GRS-values are not accessible for the public. Therefore we exemplary calculated a corresponding GSR-value from available test results and from data supplied from the manufacturer for the TRY Wiirzburg and for a special load profile (draw off three times a day at 6 a.rn., 12 a.m., 6 p.m.), which was obtained to 328 kWh/m2a referred to the collector aperture area. The ISTT-procedure yielded 337 kWh/m2a, which means, that the solar system in the frame of the accuracy of the ISTF-proeedure (5%) is working correctly as guaranteed by the planner. 6. OUTLOOK The ISTT-procedure is to be applied to two further large solar systems from the German project "Solarthermie 2000" with 110 m2 flat plate and 100 m 2 evacuated tube collector area and
0,338
12,13
8 m 3 and 2 m 3 storage volume, respectively. The costs for the ISTT-procedure is running up to 5000-7000 Euro, which is drastically less than the long term monitoring procedures applied so far. A further advantage is the fact that with the ISTTprocedure, once having identified the parameters, the solar energy output can easily be recalculated and transformed to arbitrary operation conditions and is not only obtained for the in situ conditions during the measurement as for the methods used until now. Therefore the recently developed ISTT-proeedure promises an interesting an low-cost alternative to check the efficiency of installed solar thermal systems. ACKNOWLEDGEMENT This project was supported by the German Federal Ministry for Research and Technology (BMBF, grant number 032 97 28 A). The authors additionally want to thank the ZfS Hilden for supplying data from long term monitoring the solar system investigated in this paper.
REFERENCES (ASEW, 1998). ASEW. Garantierte Resultate von thermischen Solaranlagen. Schlul3berieht Projekt SE/475/93/DE/FR 19931997, 1998. (lEA, 1997) H. Driiek, E. Hahne, Thermal Testing of Stores for Solar Domestic Hot Water Systems IEA Task XIV Report no. T.14.DCST.1A, Pages 111-127, TNO Report 96-BBIR0876/526.6.3573, Delft, Netherlands, 1997. (Insitu, 1996), Insitu Scientific Software, Dynamic system testing program (version 2.6.), c/o W. SpirE, Kriegerstr. 23D, D82110 Germering, Germany. (Isakson, 1995), P. Isakson, Solar collector model for testing and simulation, Building Services and Engineering, Royal Institute of Technology, Sweden, 1995. (Spirkl et al., 1997). In Situ Characterisation of solar flat plate collectors under intermittent operation, Solar Energy, Vol. 61, Nr. 3, pp 147-152, 1997. (TransSolar, 1997), Trnspid, manual for parameter identification with Trnsys, release 1.3, Stuttgart, 1996. (Trnsys, 1994), Trnsys, a transient simulation program, Solar Energy Laboratory, Version 14.1., University of MadisonWisconsin, 1994.
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SOLAR PROCESS HEAT WITH NON-CONCENTRATING COLLECTORS FOR FOOD INDUSTRY
N. Benz, M. Gut and Th. Beikircher Bavarian Centre for Applied Energy Research (ZAE Bayern), Division: Solar Thermal and Biomass, Domaglcstr. 11, D-80807 Munich, Germany, Phone: +49-89-356250-0, fax: +49-89-356250-23 email: [email protected] W. RuB Technical University of Munich, Chair for Energy and Environmental Technology in the Food Industry, Prof. R. Meyer-Pittroff. D-85350 Freising-Weihenstephan, Germany, Phone: +49-8161-713362, fax: +49-8161-714415 email: [email protected]
ABSTRACT - We present the planning of four solar thermal systems producing process heat for a large and a small brewery, a malt factory and a dairy in Germany. In the breweries, the washing machines for the returnable bottles were chosen as a suitable process to be fed by solar energy, in the dairy the spray-dryers for milk- and whey powder production and in the malt factory the wither and kiln process. Design calculations were made on the basis of a detailed investigation of the load demands and on the architectural facts. In the study we used four high efficient collectors, an evacuated plate--in-tube, an evacuated flat-plate, an evacuated tube-in-tube with CPC reflector and a flatplate with transparent insulation. In all industrial processes the solar yields are comparable to the yields of solar systems for domestic solar water or space heating. Up to 400 kWh/m2a (related to collector gross area) heat delivered to the processes are attainable. The investments in the solar systems mainly depend on the cost for the collectors, the cost for providing mountings and for piping. At best, heat costs of 100 US$/MWh are feasible (including 20 years depreciation of investments, 6% interest, maintenance and running costs).
I. INTRODUCTION Today, the thermal utilisation of solar energy is usually confined to domestic hot water systems and space heating at temperatures up to 60~ Industrial process heat has a considerable potential for solar energy as well. In developing countries, industry needs up to 50% of the national power consumption, in industrial countries the amount is between 35 and 40% (Garg 1987). Up to 25 % of all industrial heat is directly used in processes at temperatures below 180~ Most of the process heat is used in food and textile industry for such diverse applications as drying, cooking, cleaning, extraction and many others. Energy can be provided from high efficient fiat-plate collectors or concentrating collectors of low concentration ratios (Duffle and Beckman 1991). Heat costs in small domestic solar hot water systems are between 200 and 300 US$/MWh, in central solar heating plants they vary from 75 to 150 US$/MWh for systems with short term storage and from 150 to 250 US$/MWh for systems with seasonal storage respectively (Fisch, Guigas et al. 1996; Hahne 1996). Large scale solar applications for process heat benefit from the effect of scale, from simple installations without store and from poor demands on architecture as well. Therefore the investment costs should be comparatively low, even if the costs for collectors are higher. In the project we planned of four solar thermal systems producing process heat for food industry. We investigated in detail a large and a small brewery, a malt factory and a milk processing company. In this study, several high efficient solar collectors for process heat production are considered: evacuated tube
collectors, with and without reflectors, evacuated flat-plate collectors and collectors with transparent insulation. To cause economically easy terms, the facilities are planned without heat storage, i.e. the solar heat is fed directly into suitable processes (fuel saver). Therefore the maximum rate at which the solar energy system delivers energy must not be appreciably larger than the rate at which the process uses energy. Favourable conditions exist in food industry, because food treatment and storage are processes with high energy consumption and high running time. In breweries, solar energy can be supplied to several energy consumers in beer production. A promising consumer is the bottle washer. In large factories this machine works in multi shift, while the main capacity is required in summer. The specific heat demand of a bottle washer with a capacity of 20000 bottles/h is about 1200 MJ/h with a temperature level of 9 0 100~ Maltings have a high power consumption as well. Green malt (barley) is dried from 40% water content to 4%. The drying kills have very high running times due to economical and technical reasons, an optimum is an all-year round working. For drying, air from heat recovery is heated up to temperatures of about 100~ Dairies are very interesting factories for solar energy, because they are often working seven days a week. Due to their high and constant energy demand, drying processes are promising. In the production, milk and whey are spray-dried in huge towers with air which is heated from 60~ (from heat recovery) to 180~ These drying processes have a running time up to 8000 h/a.
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2. SOLAR COLLECTORS
9 Flat-plate collector with transparent insulation made of glass
capillaries (TI) (Benz, Hasler et al. 1998). To produce process heat in the temperature range above 100~ with solar collectors, high efficient systems are indispensable. Heat losses are low when gas heat conduction is reduced or even eliminated and when highly selective absorbers are used. An additional concentration of radiation leads to a further improved efficiency. In the last years both, new collectors with highly selective absorbers and with reduced gas heat conduction and non-tracking collectors with concentration were developed.
9 Evacuated tube-in-tube collector with CPC reflector (CPC) (Muschaweck, Spirld et al. 1998). Figure 1 shows the efficiency curves of the collectors for an irradiation of 1000 W/mL The efficiencies are related to the collector gross area because the maximum amount of collector area is, in general, limited by the building roof area. 3. DESIGN OF THE SOLAR SYSTEMS
For the design calculations we used the commonly used simulation program TRNSYS. The collectors were implemented as non-standard models, especially for the CPC we had to develop a new model due to the unusual incidence angle modifier. For the design the solar systems have to meet the following demands: 9 Interfacing with conventional energy supplies must be done in a way that is compatible with the process. 9 Collectors must be mounted on the available and suitable roofs of the buildings. Shading has to be considered. 9 Storage of heat has to be avoided. That means that the collector area is limited by the maximum solar power which must not exceed the base load of the processes. The calculations were carded out with the local irradiation conditions and with the time dependent load demand of the processes described above. The piping corresponds to the real structural conditions. 4. EXAMINED INDUSTRIAL PROCESSES Fig. 1: Thermal efficiency curves for the four collector types. The values are referred to collector gross area. In the study we used four collectors, which were, apart from the ETC, investigated or developed at the ZAE Bayem: 9 Evacuated plate-in-tube collector (ETC) (SPF 1998). 9 Evacuated flat-plate collector (EFP) (Benz and Beikircher
1999).
4.1
Large Brewery
The production of beer requires a rather substantial amount of energy. A biochemical-thermal breakdown leads primarily to the extraction of sugars from the germinated and kilned grain (malt). These sugars are fermented to alcohol by yeasts. After a successful fermentation, the finished beer is filtered and filled into kegs or bottles. The heat demand depends on the size of the brewery: the larger the factory the lower the specific energy
Fig. 2: Coupling of the solar collector field to processes in the large brewery (bottle washer and hot water circuit).
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consumption. On average, the total heat demand is 1.8 MJfl sales beer (Heyse 1995). Between 40 and 60 % of the total heat is used during the extraction (in the brewhouse). Most of the heat is required at a temperature level between 40 and 100~ Roughly 3 0 - 40% of the total heat is used for the filling into bottles or kegs and for the cleaning of bottles. The bottles are cleaned in washing machines in a series of steps in cleaning solution and water baths. The temperatures of the baths are always below 80~ The large brewery produces approximately 130 million litres of beer. Unfortunately, since the hot water is made available by a high-pressure hot-water net (150~ the brewhouse cannot be equipped with a solar plant for reasons of too low efficiencies at such a high temperature. The washing machines for the returnable bottles were chosen as a suitable consumer of heat. The temperatures of the baths are always below 80~ The favoured newest machine is used in a two-shift cycle. At the start of the first shift, the machine is heated to its running temperature. This process requires approximately 147.5 MJ per degree Kelvin the main cleaningsolution bath cools down. During the running time about 900 MJ/h of heat are required. The start of the shift is at six every morning, the end depends on the amount of beer to be filled. The heat demand was calculated from the data supplied by the brewery and brought down on a 3-minute-basis for further design calculations. The solar plant, depending upon the power available, is able to cover the heat demand for the operation as well as for the pre-heating of the machine, since the main solution bath (32,65 m 3) can be used as a storage (overheating to 95 ~ on days when the machine is not in use. This overheating compensates the overnight losses due to cooling until filling begins again. The solution bath of the bottle washer is heated with an external heat exchanger by solar process heat, which has a temperature between 85~ and 100~ The surplus of heat, i.e. when the machine is stopped and the storage is full, will be supplied to the central hot water circuit which runs at a temperature level above 130~ (see fig.2). Figure 3 shows the yearly energy yield and the losses of the piping for the four collector types. The largest share of the yield
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is supplied to the bottle washer. Only in case of the CPC and the ETC collector, a relevant share of the yield is supplied to the hot water circuit due to the high efficiency of these collectors at high temperatures. Neglecting economical valuations the ETCcollector is best. The yearly efficiency of the system is 25% so that 270 MJ/a are delivered to the bottle washer and the hot water circuit. This corresponds to a CO2-saving of 93 t/a. Although the solar systems runs as fuel saver, the heat demand of the bottle washer is covered by 22%. The collector gross area is limited to 1047 m 2, because the ETC-collector field in this case delivers a maximum power which covers the needs of the bottle washer. For the other collectors, the area could be increased. The collector size could be increased in general, if more then one bottle washer is supplied by solar process heat. Then the roof area would be the limiting factor.
Fig. 3: Yearly energy yield and losses in the large brewery depending on the collector type. The line graph shows the solar fraction of the heat demand of the bottle washer.
Fig. 4: Coupling of the solar collector field to processes in the small brewery (bottle washer and space heating).
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4.2 Small Brewery The small brewery produces 18 million litres of beer a year which is only 14% of the large brewery. The capacity of the bottle washer is 70% of the favoured machine in the large brewcry. In comparison it works only in one shift, usually from 6.30 a.m. to 1.30 p.m. Even in summer months, the machine runs only 4 days a week and has frequent down-times due to changeovers to other types of beer and breaks for snack or dinner. This discontinuous operation leads to only 100 hours a year in which solar heat could directly be fed into the bottle washer. Therefore a buffer store with a volume of 50 m 3 for the solar heat was planned. If the washer is down and if the main solution bath (25 m 3) is overheated to 95~ as well, the buffer store is heated up to 130~ In winter months, also space heating (70~ is enabled if the temperature of the solar heat is not high enough to deliver the bottle washer or the store (see fig.4)
used for the withering process. On average, the heating of 150.000 mVh process air takes 1.900 kW. This corresponds to a total energy consumption of 600 kWh/t finished malt. Electric and thermal energy is generated by three heating and power stations (total electric power: 615 kW, total thermal power: 1.240 kW) and a peak load heating boiler (thermal power 1.200 kW). Unfortunately, the factory is a complex of buildings with different, comparatively small roofs. Therefore space for the collector field is limited and the field has to be distributed to the suitable roofs. For the design calculations, solar heat is supplied to the hot water circuit between the heating and power stations and the peak load heating boiler, where the temperature level is 77~ (see fig.6).
Fig. 6: Coupling of the solar collector field to processes in the malting.
Fig. 5: Yearly energy yield and losses in the small brewery depending on the collector type. The line graph shows the solar fraction of the heat demand of the bottle washer. The gains and losses of the solar system are shown in figure 5. Again the ETC-collector has the best results (250 kWh/m2a) and saves 107 t CO2 a year. Only 40% of the heat is directly supplied to the bottle washer, 35% are buffered in the store and 25% are used for space heating. Due to the store, the heat demand of the washer is covered by 51%. The designed collector field area of 1056 m2 is limited by the area of the fiat roof which is partially used as car parking area. 4.3 Malt Factory The malt factory, which is situated in southern Germany, processes 50.000 t barley to malt a year. The main heat consumption appears in the double-floored kiln in which malt is alternately withered and kilned. 80-85 t green malt are processed in every charge, which takes about 36 h. The water content decreases thereby from 41-43% to 4%. The double-kiln operates continuously, 24 hours a day and 365 days a year. The kiln-process runs at a temperature level of 85~ exhaust air is
Fig. 7: Yearly energy yield and losses in the malting depending on the collector type. The line graph shows the solar fraction of the heat demand of the double-floored kiln.
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Figure 7 shows again the yearly energy yield and the pipe heat losses. An ETC-collector field supplies 360 kWh/m2a to the process by which a solar fraction of 2.4% and a CO2-saving of 66 t are achieved. The solar fraction is low, because the roofs only allow a collector area of 790 m 2. Pipe heat losses are rather high because the collector field is distributed to several roofs and the distance to the heating stations is long. The best solution would be a direct air heating atter the heat recovery at a temperature level of 30~ That would increase the gains from 360 to 580 kWh/m2a (with the evacuated fiat-plate collector), but is not practicable due to missing space in the facility.
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heat using waste-air-recuperators. The two following heatexchangers are steam-powered. The most favourable interface for the use solar process heat is to supply it to the recuperation circuit, which has a starting temperature of about 60~ (see fig. 2). In the dairy, both, the ETC and the evacuated flat plate collector attain the highest yield (400 kWh/m2a) due to the lower operating temperatures between 65~ and 90~ (see fig.9). The yearly solar efficiency is 34% and the CO2-saving is 156 t/a. Here the collector size of roughly 1600 m2 is limited by the roof area. If more space for the installation of collectors would be available, the solar yield could be easily increased.
4.4 Dairy The dairy investigated, processes milk and whey to milkand whey-powder. The annual production is roughly 57000 tons of powder. Four parallel production lines of varying sizes are used for this. A falling-stream evaporation leads to a concentration of the liquids from 6 - 9% dry substance to 50 - 60% dry substance. The subsequent spray-drying increases the drysubstance content to 95%. The drying plants are running almost continuously. The fourth production line is the most interesting for a coupling with a solar-energy plant, since it is used most frequently and requires the most heat. On average the machine is down for only thirty days a year. Above all else, this is due to the maintenance which has to be carried out every 14 days, and which requires 16 hours. Additionally, the machine is cleaned every 48 hours for two hours.
Fig. 9: Yearly energy yield and losses in the dairy depending on the collector type.
6. ECONOMICS
Fig. 8: Coupling of the solar collector field to the spray-dryer in the dairy. Production line 4 heats 120 000 mVh of outside air to 180~ The heating of the air is done by four heat-exchangers connected in series. Two of the heat exchangers are supplied with
The investments in the solar systems mainly depend on the cost for the collectors, the cost of providing mountings, the cost for piping and the planning, whereas costs for heat exchangers, pumps and controllers are less important. Table 1 shows the calculated investment costs for all systems including 10% for the planning. In the large brewery as well as in the malt factory additional mountings are necessary due to requirements on statics. This raises the costs up to 27%. The ETC in the large brewery is exceptional, because there are lightweight modules available with flat installation and individual alignment of the pipes which avoid such mountings. The buffer store in the small brewery raises the investments by 14%. The difference in the investment costs between ETC, CPC and EFPC is not great. However, the evacuated flat plate system is the low-priced one, except for the large brewery without additional mountings.
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Table 1: Total investment costs for the solar facilities collector gross area in US$/m2 small large malting brewery brewery TWD 690 750 638 504 619 ETC 453-556 CPC 597 535 659 EFPC 505 443 567
related to
cations benefit from the effect of scale and from simple installations without store. In mediterranean climate, the gains of the systems could be doubled and heat costs halved respectively.
dairy
ACKNOWLEDGEMENT
566 433 463 370
Table 2: Heat costs for the supplied solar heat to the process (20 years depreciation of investments, 6% interest, 1.5% of investments a year for maintenance and running costs, no subsidy) in
US$/MWh.
TWD ETC CPC EFPC
The project was kindly supported by the Bavarian Ministry for
Trade, Traffic and Technology (Bayerisches Staatsministerium fiir Wirtschafi, Verkehr und Technologie). We thank the SpatenFranziskaner-Briiu KGaA, the Bayerische Milchindustrie eG, the Bamberger Miilzerei GmbH and the Bayerische Staatsbrauerei Weihenstephan for co-operation. REFERENCES
large brewery 363 1 6 8 - 206 225 204
small brewery 361 217 237 193
malting 320 185 219 177
dairy 196 115 134 98
Table 2 shows the heat costs the supplied solar heat to the process. They are calculated with 20 years depreciation of investments, an interest rate of 6% and an amount of annually 1.5% of the investments for maintenance and running costs. A subsidy is not considered. The lowest heat costs result in the dairy with the EFPC, where 100 US$/MWh are attainable.
Benz, N. and T. Beikircher (1999). High Efficient Evacuated Flat-Plate Solar Collector for Process Steam Production. Solar Energy 5(2). Benz, N., W. Hasler, et al. (1998). Flat-Plate Solar Collector with Glass TL EuroSun 98, Portoroz/Slovenia. Duffle, J. A. and W. A. Beckman (1991). Solar Engineering of Thermal Processes, 2 edition. New York, John Wiley & Sons. Fisch, M. N., M. Guigas, et al. (1996). Large-Scale Solar District Heating- Status and Future in Europe. EuroSun 96', Freiburg, DGS-Sonnenenergie Verlags GmbH. Garg, H. P. (1987). Advances in Solar Energy Technologie. Dordrecht, Holland, Reidel Publishing Company. Hahne, E. (1996). Solar Heating and Cooling. EuroSun 96', Freiburg, DGS-Sonnenenergie Verlags GmbH. Heyse, K. U. (1995). Handbuch der Brauerei-Praxis. Niirnberg, Hans Carl verlag. Muschaweck, J., W. Spirkl, et al. (1998). Optimized reflectors for nontracking Solar collectors with tubular absorbers. Solar
Energy, to be submitted. SPF (1998). LTS-Katalog. Rapperswil, SPF Solar Prfifung und Forschung.
Fig. 10: Dairy. The spray dryers are situated in the huge building, the large flat roof has an area of approximately 4000 m 2. 7. CONCLUSIONS The detailed investigation of four factories in food industry showed that producing process heat with suitable collectors is promising. Best conditions for a realisation of a solar facility exist in the dairy, because of its suitable process and the structural facts. The factory has a large fiat roof which supports the collector field, pipes are short and the interface to the existing heating installation is simple (see fig. 10). The solar yields are comparable to the yields of solar systems for domestic solar water heating or space heating. In German climate, heat costs of 100 US$/MWh are attainable, when no additional mountings and no store are necessary. The appli-
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LABORATORY TESTING OF INTEGRATED COLLECTOR STORAGE (ICS) SYSTEMS WITH TRANSPARENT INSULATION MATERIAL Miroslav Bosanac and Jan Erik Nielsen Solar Energy Center Denmark, Danish Technological Institute, P.O. Box 141, 2630 Taastrup, Denmark, Tel. (45) 4350 4569, Fax: (45) 4350 7222, Email: [email protected]
Abstract In this work we used modified multi-node collector model for characterization of the ICS system performance. Multi-node collector model has distributed thermal capacities with heat transfer from one node to the next. Model characterizes themperature dependence of heat loss coefficient and incident angle modifier. Two independent test sequences were carried out and characteristic parameters identified. Specific artificial draw-off conditions were applied ( mainly continuous draw-off). Repeatability of identified parameters is satisfactory and the difference in yearly energy yields predected by each test sequences differs within 5 %. Thus, the developed method has potential for both (i) accurate performance prediction tool and (ii) diagnosis tool primarely for optimization of system design. 1.
Introduction
Although the ICS systems presents small part of the world solar domestic hot water market, they have huge potential as they have are simple construction and low cost, there is no pump or any controll necessary what make them maintenance free. Therefore, they have large application potential, specialy in developing countries. An integrated collector storage system may be tested by several methods, e.g. by the Standard ISO9459 - Part 2. The ISO9459/2 encompasses measurements of daily energy balances with a single draw-off in the evening. The predicts yearly energy performance of the system under the test without identifying its characteristic parameters. .
by the Standard I S O 9 4 5 9 - Part5. The DIS9459/5 deals with system performance characterisation by means of system dynamic test and computer simulation.
The main disadvantage of the existing methods is that they are not able to identify incident angle modifier (IAM) of the ICS systems. In particular for the ICS systems with transparent insulation material (TIM) the IAM influence is considered important and a new test method for characterisation of these systems has been developed. A multi-node collector-storage model is used to characterise capacity distribution in draw-off direction. The following parameters fully characterise the presented model (i) the optical efficiency, (ii) the overall heat loss
coefficient, (iii) the total thermal capacity of the system, (iv) the incident angle modifier coefficient (v) the effective layers-conductivity characterising storage stratification. As the parameters obtained from the test have their physical meaning, the application of the test results is not restricted to performance prediction but it allows also diagnostics on system physical behaviour. The method is validated by a laboratory test on the ICS system with TIM. The reliability of the method is judged with respect to (i) the repeatability of the identified parameters and (ii) the ability of the identified parameters to predict correctly the ICS system performance, primarily its energy yield. The identification of ICS system parameters is carried out using the dynamic fitting algorithm, namely the measured variables, i.e. the ICS inlet and outlet temperature, draw-off flow rate, ambient temperature, hemispherical and diffuse irradiance are input variables to the theoretical model. The difference between measured system power and modelled power is minirnised by optimising the set of parameters. As a result of this procedure the estimated errors of the parameters are given. The same model is used for simulation so that the error of the predicted energy output has been estimated.
138
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-
A linear dependence of the heat loss coefficient on the surrounding air speed as well as on the temperature difference between collector and ambience is assumed. Incident angle modifiers for beam (as a function of incident angle) and diffuse irradiance are used. Here are briefly described the main features of the model. Each node is characterised by: C, dT~/dt = A. F'[(z{x)0 G~q- UL(T,- T,)] - q ~ where Geq is the equivalent normal irradiance taking into account ir?adiance components multiplied by respective incident angle modifiers: Geq = K ~ beamGbeam+ K~difr Gd~r+ K~alb Galb, UL is the overall heat loss coefficient:
UL = Uo + Uv v + UT(T.- T.); q ~ is the rate of energy gain by the ICS node: q ~ = nkcp(Tn -T..I); Fig 1. ICS system on the laboratory test stand In order to enable use of a general simulation tool for prediction of the ICS system performance, a TRNSYS module has been developed. The corresponding set of differential equations is solved numerically within the TRNSYS program.
3. The ICS System The ICS system under the test has tempered 4mm glass and 100mm TIM (polycarbonate honeycomb). Cylindrical storage (160 liters) serves as an absorber. It is coated by black chromium.On the back side of storage, it is situated reflector of polish and anodized aluminium.
4. ICS System Model
(za)0 is the transmittauce-absorbtance-product at normal incidence. An incidence angle modifier for beam irradiance is defined by the modified Ambrosetti (Ambrosetti 1983) equation: K ~ b~(0)
-
1- tanl/r(0/2).
The incident angle modifier for diffuse irradiance assuming isotropic distribution is used as derived in (Bosanac et al. 1993). The incident angle modifiers for diffuse irradiance and for albedo are assumed to be equal. They are both derived in (Bosanac et a1.1993) as a function of the parameter r. Hence, the following parameters fully characterise the presented model: 9 9
The optical efficiency of ICS, F'0:o0. The overall heat loss coefficient if Tn--Ta and v--0,
9
The coefficient characterising wind dependence of overall heat loss, Uv. The coefficient characterising temperature dependence of overall heat loss, UT. The total thermal capacity of ICS, C. The incident angle modifier coefficient.
Uo. A modified multi-node collector model (Bosanac and Nielsen, 1997) has been used for this analysis. The model has the following features: The collector is modelled with distributed capacities in flow direction.
9 9 9
ISES Solar World Congress 1999, Volume III
The coefficient characterising stratification in the storage during draw-off (it is represented by thermal conductivity between the nodes). The identification of collector parameters is carried out using the dynamic fitting algorithm developed by Spirkl [Spirkl, 1990]. The dynamic fitting algorithm procedure is based on the following principle: The measured variables, i.e. ICS system inlet and outlet temperature, mass flow rate, ambient temperature, hemispherical and diffuse irradiance are input variables to the theoretical model. The difference between measured system power and modelled power is minimised by optimising the set of parameters. As a result of this procedure the estimated errors of the parameters are given. If the same model is used for simulation, the error of the predicted energy output can be estimated.
139
No-draw-off regime during at least 25 MJ/rn2 hemispherical irradiation has been received. Continuous draw-off (app. 1-10 kg/min) until the ICS system parameters being accurately identified (e.g. standard deviation of optical efficiency should not exceed 5% and standard deviation of heat losses should not exceed 15% of their respective values. The ICS system was monitored in August and September 1998 and in May 1999.
6. Identification Results
In order to enable use of a general simtdation tool for prediction of ICS performance in scope of the system, a TRNSYS module TYPE 59 [Bosanac, 1992] has been developed. The corresponding set of differential equations is solved numerically within the TRNSYS program.
5. Monitoring We consider the ICS system a collector with high thermal capacity. As steady-state test methods does not apply here, it must be used dynamic test method.
The minimal set of three parameters charactirizing the ICS system performance has been identified after 4 days of testing. Naturaly standard deviations of respected parameters were reduced whit each additional test day. Fig 2. Hemispherical irradiance along wit draw-off capacity rate in [W/K] during the monitoring sequence Tab. 1 shows comparison of identified parameters by two independent test sequences. It is shown that repeatability of test results is satisfactory. The difference in yearly energy yield predicted by these two sets differs 4.6 % for Copenhagen TRY.
Fig 2. Hemispherical irradiance during the monitoring sequence Test sequence consists of three subsectional parts: 9 Conditioning (continuous draw-off for at least 6 hours during night time)
Sequence: Aug. 3 lth, - Sept. 10 th, 1998
Sequence: M a y - 11 th, 1999
11o = 0.52 • 0.03
11o = 0.53• 0.03
Cc = (660 + 36) k J / K
Cc = kJ/K
AUL = W/K
AUL = (4.8
(4.6 + 0.3)
(664
•
6 th
22)
0.5)
W/K
Table 1. The identified parameters for the independent sequences
two
In order to enable use of a general simulation tool for prediction of the ICS system performance, a TRNSYS
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140
module has been developed. The corresponding set of differential equations is solved numerically within the TRNSYS program.
(xa)0 - product of cover(s) transmittance and absorber absorptance for normal incident angle 0 - incident angle of radiation
7.
8.
Conclusion
The simplified test method leads to a set of reliable parameters enabling accurate prediction of yearly energy yield of the ICS system. The multi-node collector model is capable for characterisation of detailed performance of the ICS system including incident angle influence on the ICS performance. However for identification of complete set of parameters far more data are necessary. Another problem to be solved is that parameters are not constants if independent sequences being used for identification.
R
E
F
E
R
E
N
C
E
S
Ambrossetti, J P (1983). Das neue Bmttow~irmeertragsmodel fiir Sonnenkollektoren. Technical Report, EIR Wiirenlingen, ISBN-3-85677-012-7. Bosanac M, Nielsen J E, (1997). In-Situ Test of Solar Collector Array. J. of Solar Energy, Vol 59 Nos. 4-6, pp 135-142. Bosanac M, Brunotte A, Spirkl W and R. S i z m (1994). Use of Parameters Identification for Flat Plate Collector Testing under Non Stationary Conditions. J. Renewable Energy Sources, 4, pp 217-222. Bosanac M, (1993). TYPE 59, TRNSYS Module for Dynamic Simulation of Collectors and Collectors Arrays, Expert Meeting lEA Task XIV, Rome 1993.
9. A C
c~ %
Nomenclature total aperture area of the ICS - the ICS total thermal capacity, NCn the heat capacity of each ICS node - specific heat coefficient of the fluid in the ICS - ICS efficiency factor incident total radiation on a fiat surface per unit
-
Klein, SA, Duffle, JA, Beckman, WA (1974). Transient considerations of flat-plate solar collectors. Trans. ASME, J Eng for Power, 96A, p 109.
-
F' G area G ~ , Gain, Galb - incident direct, diffuse and ground reflected radiation Kbean~Kdiff, Kal b - incidence angle modifiers, beam, diffuse and ground reflected radiation draw-off flow rate Ilk - parameter for incident angle modifier r number of nodes N rate of energy gain by ICS node q~ - ambient air temperature in vicinity of the ICS T. - ICS mean fluid temperature Tm - ICS node temperature T. reduced temperature T* overall heat loss coefficient UL overall heat loss coefficient when T=Ta, and Uo v--0 - coefficient characterising wind dependence on Uv the heat loss coefficient UT coefficient characterising temperature dependence of the heat loss coefficient v - wind speed in the collector plane dt - time step dT - increment in node temperature over the time step -
-
-
Labtech (1996). Data acquisition and process control software, Labtech 400 Research Drive, Wilmln"gton, MA01887. Mather, GR, (1982). Transient Response of Solar collectors, Trans ASME, J. Sol. Energy Eng. 104(3), pp 165-172. Spire W (1992). Dynamic SDHW System Testing, Program Manual, Sektion Physik der LudwigMaximilians Universit~t Miinchen.
-
TRNSYS (1994). A Transient System Simulation Program, Version 14.1 Solar Energy Laboratory, University of Wisconsin, Madison, Wisconsin.
-
-
-
TRY (1982). Danish Test Reference Year, SBI Rapport 135, Statens Byggeforskningsinstimt.
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141
UNCERTAINTY IN ECONOMICAL ANALYSIS OF SOLAR W A T E R HEATING AND PHOTOVOLTAIC SYSTEMS
Ser.qio Colle, Samuel L. de Abreu and Ricardo ROther LABSOLAR- Department of Mechanical Engineering Federal University of Santa Catarina P.O. Box 476, 88040-900, Florian6polis, SC, Brazil E-mail: [email protected]
Abstract: The present paper focuses on the statistical analysis of the fraction of the energy demand supplied by solar energy based on the f-chart method for solar water heating systems. For photovoltaic systems, the saving cost is linear with the collector area. The uncertainty of the solar fraction is correlated with the monthly means of the global irradiation and the correlation coefficient between monthly means. Numerical examples for one location in Brazil and three locations in the United States are presented. These examples show that the uncertainty of the life cycle savings is significantly dependent on the uncertainty of the monthly means of the solar radiation data. The present analysis is intended to provide a basic procedure that could be useful to make a straightforward feasibility analysis of a solar system. This is particularly interesting to evaluate the investment risk associated with photovoltaic plants, for which the capital cost greatly overcomes the advantages in saving electric energy consumption from the utility grid. I. INTRODUCTION The production of PV modules has been increasing in the last years, while the production cost has decreased insofar, due to new manufacturing technologies and production scale factors. Thin film modules of amorphous silicon are offered in the market by a price of US$ 4,00/Wp or less (Curry , 1999). Government incentives in USA, Europe and Japan are expected to heat up the market, which means more investments in research and development for competitiveness. The production cost of US$ 2,55/Wp, thought to be reached in year 2004, is becoming realistic as reported in Curry (1999). On the other hand, the cost of energy derived from fossil sources and hydro are pressed to go up, due to the increasing penalty for environment degradation and pollution, the requirements for increasing investment in exploration and to the decrease of the availability of fossil fuels. Searching new alternatives to produce pollution free energy is nowadays included in government planning worldwide. Solar energy, in this context, has been considered a true competitive alternative for the near future. Prior to making a decision on any alternative energy project, one should look for an economical figure of merit. The techniques for economical analysis presently in use in solar energy projects are the Life Cycle Cost (LCC), Life Cycle Savings (LCS), Annualized Life Cost (ALC), Payback Time and Return of Investment (ROI), are described in Duffle and Beckman (1991), and many standard books in economics. Among these, a useful and straightforward technique for LCS to optimize solar heating and cooling systems is the P1-P2 method proposed in Brandemuehl and Beckman (1979). Sensibility analysis is useful in order to evaluate the effect of design parameter variation, as well as the effect of inflation, interest rate and fuel cost variation, and capital cost on the LCS. A complete analysis is also carried out in Brandemuehl and Beckman (1979). In the circumstance where the capital cost of any alternative energy plant becomes close to the threshold cost, the precise knowledge on the availability of the primary energy resources becomes of major importance. In the ease of PV
generation therefore, the availability of solar radiation data, its variability as well as the uncertainty of the monthly means of the global radiation, should be taken into account to evaluate the uncertainty of LCS. This paper presents an elementary analysis of the uncertainty of LCS, either as a function of the monthly means of global radiation or the monthly means of the total radiation incident on the tilted surface of the thermal collector or PV modules. The analysis will be carried out for water heating systems, and PV systems integrated to the utility grid, as reported in Riither (1998). Large uncertainty may arise from modeling correlations for the radiation incident on tilted surfaces. In spite of the fact that these uncertainties can be significant, the numerical analysis will take into account only the uncertainties arising from the monthly means of global radiation on the horizontal surface. The monthly means of global radiation are usually estimated from data of sunshine duration records or from pyranometer data collected in ground stations of meteorological services. The former are available in many countries for long-term periods, up to 30 years. The monthly means derived from sunshine duration records are less accurate than those measured by calibrated pyranometers. On the other hand, time series longer than 30 years of qualified data from pyranometers are seldom available, particularly for South American countries, as reported in Tsvetkov (1997). The assessment of solar radiation from satellites has become a useful way to derive monthly means of incoming global radiation on a horizontal surface. Bias errors less than 5 % and mean square errors around 7 % for monthly means are usually found by many authors (Zelenka et al., 1992; Stuhlmarm et al., 1990; Pereira et al., 1996; Pinker and Lazlo, 1992). Presently, data derived from satellites are seldom available for periods longer than ten years. The sampling of monthly means derived from satellites are therefore limited and statistically not representative. On the other hand, for many countries, satellite derived data are the only possibility to assess solar energy, as is the case of Brazil (Colle et al., 1999) and many South American countries. Therefore, before
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going to study the economical impact of solar energy in the energy market, one should take into account the effect of the uncertainty of the solar radiation data on the figure of merit of the economical
analysis.
available analytically in terms of the monthly mean. Therefore the correlation for monthly means given by Hay (1979) will be adopted here. This correlation can be written in non-dimensional form as a function of the monthly average of the daily clearness index K r as given in Appendix B. In this case Y/ can take the
2. UNCERTAINTY ANALYSIS
form
As mentioned before, two systems will be investigated in order to cover solar domestic hot water systems and PV applications as follows: Case A: Solar domestic hot water system (SDHWS) According to Duffle and Beckman (1991) for the P1-Pe method, LCS is given by the following equation
LCS--P1CF1F L - P 2 ( C A 4 +CE)
m
energy (US$/GJ) in the first year of the economical analysis, F is
(6)
where
(7)
v/(rr ) = (ra).
(1)
where A c is the collector area, CF1 is the cost of the auxiliary
n
Yi=FR(ZOt)n(FR'/FR)HoiNiAc ~t(KT)/L i
According to the theory outlined in Appendix A, the uncertainty of F L can be conveniently expressed as follows m
the annual solar fraction, L is the annual load (GJ), C A is the
8(F L)/(F L)=QFR(rCt)n(FR'/FR)AcHaN/(F L)
(8)
collector cost per square meter (US$/m:), Ce is the cost of the where system independent of the collector area, and Pl and P2 are economical factors, accounting for reducing the operational cost to the present value, financing, insurance cost, depreciation and other minor costs. The annual fraction F is expressed as 12
F=
12 Q = [ g(Yi)g(Yj) V (KTi) V (KTj)Pij i,j=l X
(N i / N ) ( N j / N)(diTzt'~ / H--a )(bT-1-'jI H'-a )]1/2
(9)
m
fiL~/L i=1
(2)
where g ( Y ) = Of /OY, u
T are given in Appendix B
and H a is the annual average of the monthly means H i . The
where L i is the monthly load (GJ). For liquid systems f/ is a
correlation coefficients for the monthly means H i and H j are defined according to Appendix A as follows
function of parameters X i and Yi given by 1 M f/ = 1 . 0 2 9 Y / -
0.065Xi
(Hik - H/')(H-'jk -H--~)/dridrj Pij = "-Mk=l
- 0.245Y/2
(3)
+ O.O018X/2 + 0.0215Y/3
(10)
where M
where O'i =[
X i = (FRU L )(FR'I F R )(Tref - Ta )At A e / L i
(4)
(nik _ ~ ; ) 2 / M ] l / 2 k=l
(11)
and
and M m
Yi = FR(ra)n(FR'/FR)(ra)/(ra)nHriNiAc /Li
The monthly mean of the solar radiation incident on a tilted
surface,
H n ,
H~.=
(S)
is related to the monthly mean of the global and the
diffuse radiation on horizontal surface. There are correlations derived from the hourly sums method (HSM), daily sums method (DSM), as reported in Behr et al. (1997), Reindl et al. (1990), and Hay and McKay (1985). These correlations however are not
Hik/M k=l
(12)
where M is the number of years of the sample for months ( i ) and
(j). The computation of p/j requires yearly series of qualified monthly means H i with stabilized statistics, which means M _> 30. The uncertainty analysis can be extended to M less than 30, if
ISES Solar World Congress 1999, Volume III
appropriate criteria are assumed to estimate the confidence interval (Coleman, 1989). The uncertainty of LCS can be derived from Equations (1) and (8) in the form o%CS / LCS = QCF1PIFR (VOt)n (FR' / FR ) A c H a N / LCS
(13)
143
It can be seen from Equations (13) and (16) that the ratios P1CF1FR(va)n (FR' / FR ) A c H a N / LCS and P1CelAcHaN / LCS
are meaningful economical parameters. These are proportional to the ratio of the maximum energy savings due to solar radiation in the first year of the economical analysis, and the life cycle savings. The relative uncertainty of LCS is seen to be proportional to the annual average of global irradiation H a and inversely
m
The uncertainty analysis relative to Hri can be derived in the
proportional to LCS.
same way as given for H i , for which case it takes the form 3. NUMERICAL EXAMPLES m
bZCS / LCS = QTCFIP1FR (Va)n (FR' / FR ) A c H a N / LCS
(14)
where Qr assumes the same expression of Equation (9) with g / ( K T) taken equal to the unity, and Pij in this case being the
In order to simplify the present analysis and to reduce the calculations, the uncertainty of monthly means is assumed to be ( t S i ) / H a (no bias) and equal to e g , constant for all months. Furthermore, the uncertainties of Hri as a function of H i are not
correlation coefficient of HTi and HTj.
taken into account. In the case the uncertainty of H i is assumed to vanish for 12 - p
Case B: PV system integrated to the utility grid The LCS in this case can be simplified by using the P1-P 2 method, once the average efficiency of the PV system r/i for each month ( i ) is known. In this case LCS is given by
months, p < 12, the calculations should be made for each case corresponding to the other p non-vanishing months. For p = 2, there are 12 ! / 10! 2 ! = 66 cases; for p = 3 there are 12 ! / 9! 3 ! = 220 cases, and so one. The total number of cases for all possible combinations o f p non-vanishing months is the binomial number 212"
12
LCS=P1CE1
rli(Tpi)HTiNiAc-P2(CaAc +CE)
(15)
i=1
With the assumption of the same uncertainty of H i for all non vanishing p months, Equations (13) and (16) become linear in e~.
where CE1 is the electrical energy cost in the first year of the
The slope of the resulting straight line of ULcs as a function of
economical analysis (US$/kWh) and Tpi is the average operating
e~, depends on the economical parameters, as the cost of the
temperature of the PV modules for month ( i ). The uncertainty of LCS in this case can be written as follows
auxiliary energy as well as the way the monthly means are distributed during the year, and on the correlation between these monthly means. In particular, microclimate changes due to seasonal human activities, i.e., forest burning and also due to the activities of volcanos can partially or totally impair the monthly means along the year. The impact of these activities on the uncertainty of LCS can also be estimated from Equations (13) and (16). The numerical examples are carried out here according to the following specifications:
OZCS / LCS = (P1CEIAcNHa / LCS) 12
rlirl./(Ni /N)(N./ /N)p O.
x[ i,j=l xg'(Kn)
u (Krj )(b71i / H a ) ( o C H j / H a ) ]
(16)
If it is assumed a bias error B i and a precision index S i for H i , the relative bias error is B i / H a , while for a 95% confidence interval for H i , the relative precision index is (tS i ) / H a , where t is the t-distribution of Student. Equations (13) and (16) hold to estimate the bias error BLCs , in which case 8 H i should be
Case A: SDHWS Collector area, Ac optimized for each location Annual load, L = 13.8 GJ FR(m). = 0.7 FRUL 5.0 W / m K pg = 0.2 - -
replaced by B i . These equations hold also to estimate the 95% confidence interval for LCS, in which case 8 H i should be replaced by tSi . The uncertainty of LCS is then given by ULCS 2 =BLcs 2 +(tSzcs) 2 . If no bias error B i is assumed, BLCs
vanishes and the relative uncertainty of LCS is then
ULcs = ( t S L c s ) / L C S ,
which is a function of the relative
uncertainty U i = (tS i) / H a .
Cost of the fuel in the first year, Cm = US$ 28,00/GJ Inflation of Cm, iv = 10% Discount rate, d = 8% Cost of collector area, CA = US$ 85,00/m 2 Cost independent of collector area, Ce = US$ 600,00 P1 = PW(Ne, i~ , d ) (non-commercial plant) P2 = 1 (the system is totally financed by the owner) Period of economical analysis, Ne = 20 years
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144
In the present analysis, the cost CE accounts for the cost of the reservoir. Customers may be interested in purchasing collectors in the circumstance they already have the reservoir (gas fueled or electrically heated). In this ease, CE would include only the installation cost, auxiliary pump and piping, and other minor costs. Case B: PV system This system has been in operation since 1997 at LABSOLAR, and has the following specifications (Riither, 1998): Power = 2 kWp Average monthly efficiency, r/= 5.3 % (measured) pg = 0.2 Cost of the electrical energy, CE1 = 10 r Inflation of CE1, iF = 10 % Discount rate, d = 8 % Capital cost = US$14.000,00 ( US$ 7,00/Wp) For the PV system chosen here, the threshold cost (for which LCS = 0) is US$ 4,6/Wp. Four locations are chosen here, one is the city of Campo Grande (20.45~ 54.62~ in Brazil and three locations in USA, namely, Miami (25.8~ 80.27~ Houston (29.98~ , 95.37~ and Los Angeles (33.93~ 118.4~ Miami and Campo Grande are cities located in subtropical areas. For the location of Campo Grande, the Brazilian Weather Service (INMET) provided the records of monthly means derived from measurements with pyranometers during the period between 1973 and 1990 (17 years). The monthly means derived from measured radiation for Miami, Houston and Los Angeles for a 30 years period is found in Marion and Willcox (1994). While the correlation coefficients PO" are estimated with confidence for the USA locations, these coefficients show a lesser degree of confidence for Campo Grande, since the statistics for the 17 years long time series was found to be not stabilized. Therefore it is necessary to verify the effect of the correlation coefficients p# on the uncertainty of LCS. Figures 1 and 2 show the results obtained for correlated (p# ~ 0) and uncorrelated ( p # = 1; i = j and
p/j = 0; i ~ j) monthly
means, for the SDHWS and PV system, respectively, for Miami and Campo Grande. These figures show that for e~ = 10% the uncertainty of the LCS differs in 2% for the SDHWS and around 8% for the PVsystem. This means that in the case of PV, when the capital cost is close to the threshold cost, the correlation coefficients should be significant in the evaluation of the uncertainty of LCS. The effect of the capital cost on the uncertainty of LCS is shown in Figure 3, for the city of LOs Angeles. It is seen from this figure that for a capital cost of US$ 4,00 /Wp, an uncertainty e~ of 10% corresponds to an uncertainty of LCS around 35%, while for a capital cost of US$ 3,00/Wp it is around 10 %. For a capital cost of US$ 2,00/Wp, the uncertainty of LCS becomes pretty small, around 5 % and for this case, e~ of 5 % corresponds to an uncertainty of LCS around 3%.
Since the capital cost of the SDHWS is relatively low, the uncertainty e~ has a small effect on the uncertainty of LCS, as shown in Figure 4. This is due to the relatively high value of LCS for the type of system chosen here. The effect of the months for which the uncertainty vanishes is shown in Figures 5, 6 and 7, for the PV system with capital cost equal to US$ 3,00/Wp. It can be seen from these figures that forp fixed non-vanishing uncertainties, all corresponding cases lie between two limiting straight lines, which correspond to the maximum and minimum for the set of all possible cases. The uncertainty of LCS for the PV system for the different locations chosen is shown in Figures 8 and 9. These figures show that for both the SDHWS and the PV system, the uncertainty of LCS depends on the location. This conclusion can be drawn for the USA locations chosen here. The results for Campo Grande is less precise, because of the lower confidence of the correlation coefficients of monthly means for this location. The effect of electrical energy cost on the uncertainty of LCS is shown in Figure 10 for the PV system with capital cost equal to US$ 3,00/Wp. It shows how the increase in the electrical energy cost leads to a decrease in the uncertainty of LCS. 0.20
0.15 I-
M.iauli ( t m c o n ' c l a t c d )
Miami ~ 0.10
0.05
Campo ~ n d e (uncorrdated) 0.00!
oo
.
.
o11
.
.
.
oi,
I
~a Fig. 1. Uncertainty of LCS for the SDHWS for correlated and uncorrelated monthly means. 1.0
0.8
Miami (correlated) 0.6
~'*
t
Miami(unco~]a~)
0.4
0.2
Cam~ (eorretated)
\c 00
011
po 0:3
0'.,
I 0.5
Fig. 2. Uncertainty of LCS for the PV system with capital cost of US$ 3,00/Wp for correlated and uncorrelated monthly means.
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145
Fig. 3. Effect of the capital cost on the uncertainty of LCS for the PV system for the location of Los Angeles.
Fig. 6. Uncertainty scattering of LCS of the PV system with capital cost of US$ 3,00/Wp, for p = 6 for the location of Los Angeles (924 cases).
Fig. 4. Effect of the capital cost on the uncertainty of LCS for the SDHWS for the location of Campo Grande.
Fig. 7. Uncertainty scattering of LCS of the PV system with capital cost of US$ 3,00/Wp, for p = 9 for the location of Los Angeles (220 cases).
Fig. 5. Uncertainty scattering of LCS of the PV system with capital cost of US$ 3,00/Wp, for p = 3 for the location of Los Angeles (220 cases).
Fig. 8. Effect of the location on the uncertainty of LCS for the SDHWS.
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The present analysis could be useful to determine the effect of the hourly variation of the electricity cost during the day, on the uncertainty of LCS for those cases of high effective load carrying capacity. The same approach used here can be extended, to analyze the uncertainty arising from the correlation relating the monthly mean of incident radiation on tilted surfaces to the monthly mean of global radiation on a horizontal surface. The uncertainty degree, levelly presented here is far underestimated, since the uncertainty associated with the correlation between global and diffuse radiation is also not taken into account in the present analysis.
1.0
0.8
Houston o.6L-
Miami
0.4
Caande 0.2
Los Angeles
0.0
011
012
0.3
E~
NOMENCLATURE
0'.4
0.5
Zc
C CE]
Fig. 9. Effect of the location on the uncertainty of LCS for the PV system with capital cost ofUS$ 3,00/Wp. 0.8
C~ d FR FR' FR 89
0.6
Ha
~
collector area (m2) C = 0 for non-commercial plants; C = 1 for commercial plants cost of electric energy in the first year of the period of economical analysis (US$/kWh) cost of the auxiliary energy in the first year of the period of economical analysis (US$/kWh) discount rate collector heat removal factor modified heat removal factor (=Fx) collector loss factor historical annual daily average of global radiation, derived from the monthly means H--/* (J/m2).
0.4 m
Hi
monthly mean of global solar radiation for month (i)
(J/m2)
m,
0.2
0 . 0 _1~"-
0.0
,
t
0.1
.
.
.
0'.2
0'.3
.
;.4
Hi
historical average of the monthly means H i
HT~
monthly mean of solar radiation incident on a tilted surface for month (0 (J/m2) inflation of the auxiliary energy daily clearness index = H/H0
0.5
En Fig. 10. Effect of the electrical energy cost on the uncertainty of LCS for the PV system with capital cost of US$ 3,00/Wp for the location of Los Angeles. 4. CONCLUSIONS The uncertainty analysis of the LCS for a solar domestic hot water system and a PV system is carried out. It is shown that the uncertainty of the monthly means of global radiation is important to estimate the uncertainty of the LCS of PV systems integrated to the utility grid, particularly in the case where the capital cost is close to the threshold cost. For a fixed value of the uncertainty of the LCS, there is a correlation between the uncertainty of the solar radiation data and the capital cost. The greater the capital cost, the smaller the accepted level of uncertainty of these data should be. The relative uncertainty of the LCS becomes sensitive with the uncertainty of the monthly means, but it is dependent on the value of LCS itself. However, the relative uncertainty of LCS is highly sensitive for cases of low LCS, i.e. for circumstances of low auxiliary energy costs or high capital cost. The effect of the uncertainty of a subset of months with known uncertainty in the monthly mean in the year is also shown for different subset cases.
KT Kr L Li LCS N
N, N, t'1 1'2=1 PWF Rt,
t
_ t Ta -Tref
m
monthly average clearness index = H / H0 annual load (GJ) monthly load of month (i) (GJ) life cycle savings number of days in the year number of days in the month (i) period of the economical analysis P1 = ( 1 - C t)PWF(N e, iF, d) for the case the owner pays the system cash, noncommercial plant, no depreciation value, no federal and state taxes and no insurance cost. present worth factor for a series of payments ratio between the monthly mean of beam radiation incident on the tilted surface and the monthly mean of the beam radiation incident on the horizontal surface t-student distribution effective income tax rate average monthly ambient temperature (~ reference temperature for f-chart (100 ~
ISES Solar World Congress 1999, Volume Iil
147
Greek symbols At total number of seconds in the month considered reflectance of the ground surrounding the collectors Pg
Behr, H. D. (1997), Solar radiation on tilted south oriented surfaces: validation of transfer-models, Solar Energy, Vol. 61, No. 6, pp. 399-413.
( ~a) (ra% (ra)d
Reindl, D. T., Beckman, W. A., and Duffle, I. A. (1990), Evaluation of hourly tilted surface radiation models, Solar Energy, Vol. 45, No. 1, pp. 9-17.
(~a)g
(ra)o co,
average transmittance-absorptance product (monthly) transmittance-absorptance product for beam radiation transmittance-absorptance product for diffuse radiation transmittance-absorptance product for radiation reflected from the ground normal transmittance-absorptanee product sunset angle for horizontal surfaces
REFERENCES Curry, R. (1999), Photovoltaic Insider's Report, Vol 18 No. 3, pp. 1-6. Duffle, J. A. and Beckman, W. A. (1991), Solar Engineering of Thermal Processes, 2nd Edition, Wiley Interscience, New York. Brandemuehl, M. J. and Beckman,W. A. (1979), Economic Evaluation and Optimization of Solar Heating Systems, Solar Energy, Vol. 23, No. 1, pp. 1-10. Riither R. (1998), Experiences and operational results of the first grid-connected, building-integrated, thin film photovoltaic installation in Brazil, Proceedings o f the 2nd World Conference on Photovoltaic Solar Energy Conversion, 6-10 July, Vienna, Austria, pp. 2655-2658. Tsvetkov, A. (1997), (personal communication at the BSRN Workshop of Budapest, May 1997), Reports of the World Radiation Data Centre- Leningrad (St. Petersburg) Zelenka A., Czeplak, G., D'Agostino V., Josefsson, W., Maxwell, E., Perez, R., Noia, M., Ratto, C., and Festa, R. (1992), Techniques for supplementing solar radiation network data, Report I E A - SHCP 90-1, Vol. 2, Int. Energy Agency. Stuhlmann, R., Rieland, M., and Rachke, E. (1990), An improvement of the IGMK Model to derive total and diffuse solar radiation at the surface from satellite data, J. Appl. Meteorology, Vol. 18, pp. 1172-1181. Pereira, E. B., Abreu, S. L., Stuhlmann, R., Rieland, M., and Colle, S. (1996), Survey of the incident solar radiation in Brazil by use of meteosat satellite data, Solar Energy, Vol. 57, No. 2, pp. 125-132. Pinker, R.T. and Lazlo, I., (1992), Modelling surface solar irradiance for satellite applications on a global scale, J. Appl. Meteorology, Vol. 32, pp. 194-211. Colle, S., Abreu, S. L., Couto, P., Mantelli, S., Pereira, E. B., Raschke, E., and Stuhlmann, R., (1999), Distribution of solar irradiation in Brazil derived from geostationary satellite data, presented at ISES 1999, Jerusalem, July 5-9.
Hay, J. E. and McKay, D. C. (1985), Estimating solar irradiance on tilted surfaces: A view and assessment of methodologies, Int. J. Solar Energy, Vol. 3, pp. 203-240. Hay, J. E. (1979), Calculation of monthly mean solar radiation for horizontal and tilted surfaces, Solar Energy, Vol. 23, pp. 301-307. Erbs, D. C., Klein, S. A., and Duffle, J. A. (1982), Estimation of the diffuse radiation fraction for hourly daily and monthlyaverages global radiation, Solar Energy, Vol. 28, pp. 293. Coleman, H. W. and Glenn Steele Jr., W. (1989), Experimentation and Uncertainty Analysis for Engineers, W i l e y - Interscience, New York. Marion, W. and Wilcox, S. (1994), Solar radiation data manual for flat-plate and concentrations collectors, N R E L - US Dept. of Energy 463-5607 DE93018229
ACKNOWLEDGMENTS Thanks are due to I N M E T - Brazilian Weather Service and to NREL - National Renewable Energy Laboratory for providing the radiation data. Thanks are also due to the students W. Nuemberg and A. Montenegro for helping with the computation of the radiation data statistics. The authors are indebted to CNPq for the support of this work and also for support for participation in the ISES 99 meeting. The authors are also indebted to the Alexander von Humboldt Foundation- Germany for funding the PV system integrated to the grid, from which the performance data were taken. APPENDIX A - BASIC UNCERTAINTY ANALYSIS Let f = f ( X i , X 2 ..... Xn) be a function of n variables. Associated to each variable there are a bias error B i and a variance Gi . The total uncertainty for a 95% confidence interval associated to an estimate of X i is Ui 2 = Bi 2 + (tS i)2, where S i is an unbiased estimator for a i , tSi is the precision index and t is the t-distribution of Student corresponding value, chosen for 95% confidence. Similarly, the uncertainty for f is defined as follows, U f 2 = B f 2 + (tS f )2 . The relationship between U f 2 and the uncertainties U 2 , i=1,2 ..... n according to Coleman (1989) is given
by
ISES Solar World Congress 1999, Volume III
148
= n
~f
,,j=,
~f
(A1)
f f j po ,sj
where U 2 = B 2 + (tS) 2 . For uncorrelated variables X i and X j , Le., p/j = 0 for i ~ j , the sum of Equations (A1) and (A2) lead to
and
n(~f ~2 (tSf)2 = n
~f
~f
~,s=~~x~ ~x y P~
y)
ef2=i=lL-~i
(A2)
(A10)
ei 2
The life cycle savings LCS is a function of the averages monthly means of the solar radiation incident on the tilted surface, Hri,
where
p~ =%/a~%
(~)
i=1,2 ..... 12. Hri by its turn is a function of the monthly mean of the global radiation incident on the horizontal surface, so that
givenby
O'/j is the covariance of X i and X j 1
~L CS_ = ~L CS_ 3H r~_ ~gHi ~H n ~gHi
N
Gij = lim - (Xik - It i ) ( X # - It i) N ~ * * N k=l
(A4) The general equation for the uncertainty of LCS is given by
where
bT_,CS = 1
(A11)
12 OLCS OLCS -- ---- Pij ~-Ii ~ l j i,j=l OHi OHj
(A12)
N
/ti = lim ~" Xa N--->~ k = 1
(AS) APPENDIX
B - CORRELATIONS m
The correlation of (Hay, 1979), between H T and H
is the expected value of X i . If the distribution of Eik = X i k - fli is normal, for N > 30 the
expressed in the dimensionless form as
estimator o f / t i is given by
m
-
1 = --
Xi
N
(rot) n HO
N X
k=l
ik
(~'a) n
m
(A6)
('t'a)g ~ T ( l _ c o S f l ) + (~"~)d ff(gT) + Pg (rct)n 2 ('ca)n
while cr/j is estimated by
m
1 - -
m
x { (K T - r 1
can be
T ))R b +
. _ -
m
(1 + cos fl)[1 - (K T - ~ ( K T) (B 1)
N
(Xik - ~ ' ) ( X j k - X i )
tr/j= N
(A7) B
k=l
m
where f0= H d / H o is expressed according to (Erbs et al., 1982) as follows
and cri is given by 1
cri = ~ -
N
- - -2
(Xik - X/) k=l
(A8)
e(KT) =
HO
[ 1.391KT-3.560KT 2 +4.189KT3-2.137K'r 4
In the case the same bias error B and precision index tS are assumed for all X i , it is easy to see from Equations (A1) and
=~
0.3 < K-'T - < 0.8, for cos < 81.4 ~ 1.311K'-r - 3.022~r 2 + 3.427Kr 3 - 1 . 8 2 1 K r 4
(A2) that
[ 0.3 < K-"T < 0.8,for cos > 81.4 ~ = ( n Of Of Pij) U2 Uf2 i,j=l ~Xi ~X j
(A9) m
m
n
where K r = H r / H o .
032)
ISES Solar World Congress 1999, Volume III
qz'(r'"r ) = [1- q}'(X'-r )] (ra)6 R'b ('/'a) n
('t'a)g ( 1 - c o s f l ) + ('t'a) d tp,(~r) +P g ('t'a) n 2 ('ta) n x{[Kr - e ( K T )]Rb + 2(1 + cos fl)[1- (K r - tP(Kr))]} +
('t'tr)d e ( K r ){[I-e'(K T)]/~'b
(ra).
-
+
033)
(1 + c o s f l ) [ t p ' ( K r ) - l ] }
where (o'(KT) = d_._r . dKr The monthly solar fraction f expressed as
(Duffle and Beckman, 1991) is
f = 1.029Y-O.O65X-O.245Y 2
+ 0.0018X 2 + 0.0215Y 3
034)
The partial derivative of f with respect to Y is g ( Y ) = Of/~gY = 1.029 - 0.49Y + 0.0645Y 2
034)
149
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ISES Solar World Congress 1999, Volume III
SOLAR POND AS AN ENERGY SOURCE FOR DESALINATION Uri Fisher Head R& D Dept., Ormat Industries Ltd., P.O. Box 68, Yavne, 81100, Israel, Telephone Number 972-8-9433777, Fax Number 972-8-9439901, E-mail :[email protected]
Abstract - A shortage of potable water in many populated areas around the world has already reached the point at which desalination of seawater is the only solution. A large number of arid zones, near the seashore, have the requirements for seawater desalination and are situated in a high enough solar radiation zone suitable for large scale utilization for power or heat production. Compatibility between the heat source and the desalination plant utilizing it, is a major factor in the economics of such cogeneration plant. The development of solar pond technology was accelerated in the eighties mainly because of the threat of increasing oil prices that until now seem to be false alarm. The aim was electric energy production that was evaluated against electricity generated from gas, coal and oil. Because of the strong emphasize on energy, the discontinuing of energy projects due to the inability to compete in power generation, shadowed other applications such as desalination that lacked the sense of urgency which approaches us today.
The design, construction and most important, the continuous maintenance of the solar pond have been tested and evaluated. Cost of construction, construction materials and cost of maintenance work and materials has been up-dated. Information in regards with updated flash desalination techniques is available in the market. However, the simplicity in control and operation and the small temperature difference per effect, made us chose the LT-MED as a compatible system for cogeneration of heat and water. This paper discusses the combination of these two proven technologies: Salt Gradient Solar Pond (SGSP) and MultiEffect Distillation (MED), for large-scale solar desalination. The solar pond produces heat at temperatures ranging from 60~ to 95~ Assuming availability of sea water for desalination at 25~ to 35 ~ this same water act as heat sink and allow for large enough number of desalination effects between the heat source- the bottom layer of the pond and the heat sink- the sea water. Salt gradient solar pond needs regular supply of high concentrated brine to compensate for the natural diffusion of salt from the high concentrated bottom layer to the low concentrated top layer via the middle gradient zone. The combination of the MED plant with the pond eliminates the need for an evaporation system as it supplies the necessary concentrated brine by the continuous flashing of vapor from the brine in the flash chamber that acts as a salt generator. At the same time the continuous supply of concentrate enhances the stability of the solar pond that perfectly matches the operating temperature range of the MED and together provides a very inexpensive and most competitive seawater desalination systen~ It should also be specifically stressed that the combined system is environmentally benign. No brines are spilled on the ground and salt that may be accumulated on the bottom of the pond can be regularly collected and disposed of in sacks after sun drying or sent to other users. Water cost for 10,000 m3/day plant is about $1/m3. The product is high quality water of about 25 PPM or less. If mixed
with 3,000 PPM brackish water to give a final product of 800 PPM, production will rise by 35% and water cost will drop to $0.75/m3. Keywords: Solar Energy, Solar Pond, Desalination, High Concentration, Thermal Energy
Background The first practical research on the solar ponds was initiated in 1958 by Dr. Bloch who was the director of development of the Dead Sea Works in Israel. There was a slow but important development work Weinberger (1965) also by Tabor and Weinberger (1980) who took the initiative to summarize the technical achievements of those years. Then in the Seventies and Eighties came an era of enhanced development of solar pond by Ormat. The aim was to create an alternative energy source for power production as reported by Doron in 1986, Tabor in 1987. There were some demonstrations of utilization of the heat for green houses or industrial use as in the case of the El-Paso pond that was and is still used for power and heat supply as reported by Hightower, 1987. The continuous low price of oil kept industry away from this subject and most of the academic institutes do not have the financial resources to run such a system and further develop it. From time to time we find that another pond has been constructed, see Hassab (1992), Alagao (1994). Unfortunately in most cases it is only a demo that never develops into a full operating systerrL The possible use of the pond energy for desalination has already been mentioned by Tabor in 1975 as also by Doron et al 1991, Gluckstern 1991 and Hoffman 1992. Since the water situation in Israel is very sensitive to the annual precipitation with hardly any reserves, it frequently creates public discussion. The agreements between Israel, Jordan and the Palestinian Authority enhance the feeling that the increase of water supplies is a crucial matter. The use of recycled water for irrigation may postpone the exact date when large-scale desalination will be essential but it is probably in the near future.
ISES Solar World Congress 1999, Volume III
The Eastern Mediterranean is not the only near crisis area. In many populated areas around the world shortage of potable water has already reached the point at which desalination of seawater is the only solution. Large number of arid zones, near the seashore, in North Africa, Greece, Israel, south Italy, the Persian Gulf etc. have already used their ground water potential and require seawater desalination. Those countries are situated in a high enough solar radiation zone suitable for large-scale utilization of solar energy. As mentioned, since the main aim resulting from the oil crisis in 1974 was electric energy production, the pond feasibility was always evaluated by comparing cost of electric energy from the pond against electricity generated from gas, coal and fossil fuels. As a result of the inability to compete in power generation and due to the strong emphasize on energy, the pond related projects were discontinued. This also shadowed other applications such as desalination that lacked the sense of urgency that approaches us today. Pond Technology Research on the Salt Gradient Solar Pond (SGSP) was initiated in Israel in 1958. A number of small demonstration ponds were built to test various operating regimes and parameters. In the late 50s a test pond reached temperature of 96~ This was a most encouraging achievement that contributed to the understanding of the hydrodynamics of the pond as reported by Weinberger 1964. Those findings served as basis for the Ormat pond in Ein Boquek that was the first to have been combined with power generation using an Organic Rankine cycle. The successful operation of this pond as reported by Doron and Tabor 1986, led to the construction of the largest pond ever built also by Ormat near the Dead Sea. Results of the operation of the 250,000 m2 SGSP and the 5MW power plant were reported at the first International Conference on solar ponds in Cuernovaca, Mexico 1987 by Tabor. The scheme of pond and power plant is given in Fig. 1. The major technical issues studied during the construction and operation of the large pond were: -Pond geometry and optimal size -Pond lining and leakage security -Brine leak detection -Heat Loss to the ground -Formation of gases under the pond -Initial pond filling method -Establishing the salinity gradient -Maintaining the salinity gradient -Limiting the upper mixed zone layer -Heat extraction method -Power generation -Wind protection -Pond clarity maintenance -Salt make-up
Intensive work was published by Zangrando1979, 1980 and others related to pond operation and control, see Swift1989. Other ponds built since then for example the El-Paso pond is still running and it serves as a study place for students from the University of Texas that share the operation with other consortium partners. See Swift, 1993.
151
Desalination technologies and the combination with SGSP. The solar pond can produce either heat or energy for desalination. In case of electric energy production the desalination processes suitable are the Vapor Compression (VC) or the reverse Osmosis (RO). Both need only electric energy for driving the process. Since we consider here largescale operation we will use only SWRO for the comparison. In case of heat processes, there is always need for pumping and the pond can supply both the heat and the power.
The solar pond produces heat at temperatures ranging from 60~ to 95~ Assuming availability of seawater for desalination at 25~ to 35 ~ this same water acts as heat sink and allows for large enough number of desalination effects
Figure 1. Solar Pond and Organic Rankine Cycle Power Plant. between the heat source- the bottom layer of the pond and the heat sink- the sea water. A lot of information in regards with updated desalination techniques is available in the market. However, the simplicity in control and operation that allows the MED plant to work between 40% to 120% of the load and the small temperature difference per effect, made us select the Multi Effect Distillation MED as a preferred system for water production. However, Multi Flash Distillation (MSF) and SeaWater Reverse Osmosis (SWRO) will be compared as well. Salt gradient solar pond needs regular supply of high concentrated brine to compensate for the natural diffusion of salt from the high concentrated bottom layer to the low concentrated top layer via the middle gradient zone. The combination with power production cycle has no influence on the salt concentration in the various layers of the pond however; the combination of the MED plant with the pond eliminates the need for an additional salt generation system. The continuous flashing of vapor from the brine in the flash chamber supplies the necessary concentrated brine to the pond.
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ISES Solar World Congress 1999, Volume Ill
ORC average thermodynamic efficiency: 6.15 % Pond electricity production: ................ 24.35 kWh/rn2/year Pond water production for the thermal processes is based on the average Economy Ratio (ER) Pond specific construction cost: 16 $/m2 Due to unknown financial source, the cost was calculated assuming change of interest rate between 2 % and 8%. In case of different numbers one can easily complete the table.
Table 1. Size and Cost Summary Process ER
Figure 2. Solar pond and MED desalination plant. |
. It should also be specifically stressed that the combined system is environmentally benign. No brines are spilled on the ground and salt that may be accumulated on the bottom of the pond can be regularly collected and disposed of in sacks after sun drying or sent to other users.
Comparison of water cost The three main desalination technologies are presented for comparison, i.e. MSF, LT-MED and SWRO. Heat is supplied by the pond for the MSF and LT-MED while electric power can be either supplied by the pond or by the grid. In case the pond supplies the electric energy, an additional pond area is considered. Since the pond is also storage of energy for non-operating plant hours, we can assume that the production will be recovered in later period and therefore the desalination plants operates 8760 hours per year. Numerous proposals for the construction of large-scale desalination plant based on solar pond were submitted in the last ten years to various government agencies in Israel. Costs were re-evaluated and the resultant figures were used for a renewed evaluation of water cost that follows here. It has also been decided to aim at a size that will combine the desalination plant with practical size and number of solar ponds and use it as a module for multiplication in case larger desalination plants will be considered.
General assumptions: Desalination Plant size: ... 10,000 m3/day Average global solar Radiation: . . . .
2,200 kWh/rn2/year
Solar pond thermal efficiency: .............. 18 % Power is produced by Organic Rankine Cycle (ORC).
MSF 5.5 |
Specific water . production Specific energy consumption Pond area for desalination Pond area for , power Total . pond area Specific cost of Desal. Plants Desal plant cost Pond cost +power ,production I Pond cost desal. , only Total cost=power , production Total cost desal only
L T - M E D S W R O units 10 |
|
3.382 , 4.5 . 1,079 674 , .
1,753 . 1,500
|
m3/m2/y
6.15 , 2.5 . 593 ! i374 , 968 . . 1,400
I 5.5 .
' kWh/m 3 i |
1,000xm 2
1,000m2
824 824
" 1,000xm 2
1,200
$/m3/day
15,000
14,000
12,000 $xl,O00
31,568
17,428
1 4 , 8 4 0 $xl,000
, 19,500 i
, 10,682
'
$xl,000 '
46,568
31,428
2 6 , 8 4 0 Sxl,000
34,500
24,682
14,840 $xl,O00
Table 2a- MSF water production cost
Plant costPower at $0.06/kWh O&M 4% of investment/yr Chemicals & consumables Total
Including P/P 46,568,000
Without P/P 34,500,000 0.275/m 3
0.515/m 3
0.3785/m 3
0.0505/m 3
0.050$/m 3
0.565/m 3
0.6985/m 3
ISES Solar World Congress 1999, Volume III
Table 2b :MSF Investment & annual costs 25 years plant life-time
Including P/P
Without P/P
Interest %
Return Rate %
2 4 6 8 2 4 6 8
5.1 6.4 7.8 9.4 5.1 6.4 7.8 9.4
Fixed water cost $/m3 0.650 0.816 0.995 1.199 0.482 0.605 0.737 0.888
Total water cost $/m3 1.210 1.376 1.555 1.759 1.18 1.303 1.435 1.586
Table 3a: LT-MED water production cost
Plant costPower at $0.06/kWh O&M 4% of investment/year Chemicals and consumables Total
Including P/P 31,428,000
Without P/P 24,682,000 0.155/m 3
0.3445/m 3
0.2705/m 3
0.050$/m 3
0.0505/m 3
0.3945/m 3
0.4705/m 3
Table 4b SWRO Investment annual costs 25 years plant life-time Interest Return Fixed water cost % Rate % $/m3 0.375 2 5.1 Includin gP/P 0.470 4 6.4 0.573 6 7.8 0.691 8 9.4 0.207 2 5.1 Without P/P 0.260 4 6.4 0.317 6 7.8 0.382 8 9.4
153
Total cost $/m3 0.822 0.945 1.017 1.135 0.864 0.917 0.974 1.039
water
The results of tables 2,3,4 can be observed in Fig. 3. It shows that for a 10,000 m3/day plant the LT-MED system competes very well with SWRO (Desalination only), that is considered today as a most competitive plant for sea water desalination. Both desalination plants end up with water cost around 1 $/m3. The salinity of product water of the SWRO system is about 500ppm while the salinity of the product of the MED plant is about 25 PPM or less. This high quality product is actually tasteless for drinking. If mixed with 3,000 PPM brackish water to give a final product of 800 PPM, production will rise by 35% and water cost will drop to $0.75/m 3. Summary: A salt gradient solar pond can supply heat for the production of desalinated water at competitive price. The LTMED is the most suitable desalination process to be combined with a solar pond.
Table 3b:Investment and annual costs. Assume plant lifetime 25 years.
Including P/P
Without P/P
Interest %
Return Rate %
2 4 6 8 2 4 6 8
5.1 6.4 7.8 9.4 5.1 6.4 7.8 9.4
Fixedwater cost $/m3 0.439 0.551 0.671 0.809 0.344 0.432 0.527 0.635
Table 4a :SWRO water production cost Including P/P 26,840,000 Plant costPower at $0.06/kWh 0.2945/m 3 of 4% O&M investment/year 0.100$/m 3 Membrane replacements 0.0505/m 3
Total cost $/m3 0.822 0.945 1.065 1.203 0.814 0.902 0.997 1.105
water
Without P/P 14,840,000 0.3305/m 3 0.162$/m 3 0.1005/m 3
0.0655/m a
Chemicals and consumables Total
0.444$/m 3
0.6575/m s
Fig. 3. Comparison of water cost for MSF, LT-MED and SWRO driven by solar pond.
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ISES Solar World Congress 1999, Volume III
References: Alagao F.B (1994) the design construction and initial operation of a closed cycle salt gradient solar pond. Solar Energy 53, (4) 343-351.
Doron B. (1986) Solar Ponds-Lessons learned kW power plant in Ein boqueque April Conference Anaheim. Doron B. (1991) Solar Pond as an actual desalination IDA conf. On desalination and Washington
from the 150 1986 ASME solution for water re-use,
Glukstern P. (1991) Potential use of solar energy for water desalination European seminar on new technologies for use of renewable energy sources in water desalination. Athens. 2628,September. Itassab M.A. (1989) Problems encountered in operating salt gradient solar ponds in the Arabian Gulf. Solar energy Vol. 43, No3, pp169-181 Bightower S. (1987) Installation and operation of the first 100 kW solar pond Power Plant in the US. International progress in solar ponds Cuemavaca, Mexico. Hoffman D. (1992) The application of solar energy for largescale seawater desalination. Desalination, 89, 115-184. Sargent, Stephen L. Solar Pond today, International solar ponds Vol.4, No 1, Feb. 1990
Swift A.H.P. (1989). Topics in gradient maintenance and salt recycling in an operational solar pond. ASME Mechanical Engineering Solar Energy division, pp391-400. Swift A.H.P. (1993) Final report Texas solar pond consortium project551, 1989-1993. UTEP Department of Mechanical Industrial Engineering, August. Tabor H. (1975), Solar Pond as heat source for low temperature Multi effect distillation plants. Desalinationl 7, 289-302 Tabor H. Weinberger H.Z. (1980) Non Convecting Solar Ponds. Solar Energy handbook, Chap.10. (Edited by Kreider) New York McGraw-Hill. Tabor H. (1987) The Beith Haarava 5MW solar pond power plant International conf. on solar ponds. Cuemavaca, Mexico. Weinberger, H. (1964) The Physics of solar pond. Solar Energy Vol.8, No 2. Zangrando F. (1979). Observation and Analysis of a full-scale experimental salt gradient solar pond. Ph.D. Thesis, University of New Mexico, Albuquerque. Zangrando F. (1980) A simple method to establish salt gradient solar ponds. Solar Energy 25,467-470.
ISES Solar World Congress 1999, Volume Ill
155
MULTISTAGE STILL
Judith Franco, Luis R. Saravia, Sonia Esteban Instituto de Investigaci6n en Energias No Convencionales, INENCO, Universidad Nacional de Salta - CONICET Calle Buenos Aires 177, Salta- 4400- Argentina, E-mail: [email protected]
Abstract - A new design for a passive atmospheric multiple effect solar distillation unit is proposed. Inclined glass surfaces with a 4 ~ slope and placed one over the other in an isolated box are used. The cold salty water is fed only in the upper stage and flows along each surface, falling from one stage to the next by gravity and reaching finally a heated tray at the bottom. Vapour condenses below each surface and produces the water evaporation in the upper side of the same surface. The stilrs bottom is heated using a simple 1.3 m 2 solar collector with a fresnel type concentrator. The collector is separated from the still and heat is transported from one unit to the other using a 4 kg aluminium slab, which is placed in the absorber and it is heated at a temperature about 350 C. The slab is then placed below the tray in an isolated box. Four slabs are used and they are changed periodically when the slab temperature drops below 180 C approximately. Several slabs can be used for heat storage if several collectors are built, allowing the use of the still during some hours at night, improving its daily productivity. Experiments have been performed with a prototype and the results are discussed and compared with the values obtained with another electrically heated prototype.
1. INTRODUCTION In Argentina, as in most countries in the world, water is a priority. In many regions most of the water is salty, and no potable water is available. Passive solar desalination units can provide a solution in isolated rural areas for small group of persons. The simplest systems are the greenhouse solar stills. Small units with a daily production in the order of 4 litres per square meter are used. These productions are quite low. The use of passive multistage stills could provide an alternative if a simple and low cost design is available. In the past we have proposed the use of a still with a vertical disposition for the stages (1). Each one is made using an inclined stainless steel sheet with a 30 ~ slope and covered with a cotton fabric to improve the water distribution on the surface. Water is fed in the upper stage, runs along the surface and falls by gravity from one stage to the next until it reaches a tray in the lower position. Water is heated in the tray and the produced vapour condenses in the stage over it, which is cooled by the falling water. This process is repeated in each stage. In a recent work performed with greenhouse solar stills (2), it has been found that very small slopes can be used, in the order of 4 ~ A smooth water film is produced on the condensation surface when the glass is carefully cleaned with ammonia. In this paper the use of low slope glass surfaces is proposed for the multiple stage stills, reducing considerably the height of each stage and allowing the elimination of the cotton fabric, which is always a source of maintenance problems. 2. N E W DESALINATION SYSTEM
2.1 The Still Module and The Experimental Arrangement The still body is a rectangular box (50 cm x 50 cm x 36 cm) entirely built with glass and insulated externally with 50 mm thickness polystyrene foam. Stainless steel is used for the lower tray where the temperatures are higher. The system has a tray in the lower position, being heated from below. Five stages using glass surfaces with a 4 ~ slope are used, as it is shown in a cross section in Fig. 1.
Fig. 1: Multistage still cross section. The cold saline water is fed from above and falls from one stage to the next by gravity until it reaches the tray, which is heated from below. The water excess, with a high salt content, is eliminated from the system in the tray. In the upper part of each glass surface the cotton fabric is substituted by small glass dikes 1 cm high and placed as shown in Fig. 1. Water is fed in the upper parts and each dike forms a small lake with and average 1/2 cm depth covering the entire surface, since the slope is quite low and a very small amount of water is necessary to cover the surface. On the upper stage the water runs freely to maintain the temperature as cold as possible; the water supply for this tray being independent from the others. The lower tray is heated from below with four aluminium slabs that are placed in an isolated camera; the dimensions of the slabs are (27.5 x 27.5 x 2) cm 3. Figure 2 shows a general view of the still without the glass box insulation. One of the storage slabs is seen in an intermediate position.
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Figure 2.- General view of the still without the upper insulation. 2.2 The Solar Heater Multistage stills work with good productivity if the tray temperature is higher than the one used in greenhouse stills, typically around 90 C. In the high slope prototype the wood heated the system. In the new one simple 1.3 m2 solar collector with a fi'esnel type concentrator is used. The collector is separated fi'om the still and heat is transported fi'om one unit to the other using a 4 kg aluminium slab, which is placed in the absorber and it is heated at a temperature about 350 C. The slab is then placed in an isolated box below the tray. Four slabs are used and they are changed periodically. Several slabs can be heated us storage elements if a larger
collector is built allowing the use of the still during some hours at night, improving its daily productivity. The reflector concentrator is of the fi'esnel type. It is made with small height cone trunks placed concentrically above a plane. Each cone trunk is manufactured with a highly reflective aluminium sheet. The reflector is placed on an equatorial mount so that a single axis is moved during the day. A second axis normal to the first one is adjusted biweekly as the sun declination changes along the year. Figure 3 shows a picture of the concentrator prototype and one of the slabs
Figure 3.- Shows a view of the concentrator and one o f the slabs
ISES Solar World Congress 1999, Volume III
This disposition has several advantages: a)
the cones are built very simply from a flat sheet in comparison with other forms as the parabolic one.
b)
The whole mirror is flat simplifying its transportation.
c)
between cone and cone there are grooves allowing a better control of the forces produced by strong winds.
The two axes are placed in the centre of the mirror, where the incoming solar radiation is blocked by the absorber. Vertical steel column fastened to the floor supports the mirror structure. The angle of the equatorial axis and the floor can be adjusted since it should be equal to the latitude in the place where the concentrator is used. Soldered steel pipes with a rectangular cross section are used to make the mirror fiat structure. Steel wires circles with a diameter equal to the final inner cone diameters are fixed to the structure. The cones cut from fiat aluminium sheets are fixed against the wire circles as shown in figure 4 and they adopt the conical form with the angle needed to concentrate the radiation on the absorber placed 76 cm above the fiat surface. The external diameter of the whole mirror is 1.50 m and the effective reflecting surface measures 1.3 m 2.
157
Temperatures inside the still were measured every five minutes with small thermistors thermometers connected to a computer The temperature measured in the tray is not the real water temperature since the thermistor was placed in contact with the base of the metallic tray and this temperature is a little higher. Slabs temperature were measures with K type thermocouples placed inside a small hole made in the slab. Fig. 5-a and 5-b shows water's tray and slabs temperatures vs. time for two different days.
500 450 400 350 9300 0 Q.
250
E 200 150
50 0
10:00
,
~
,
,
,
,
,
11:00
12:00
13:00
14:00
15:00
16:00
17:00
18:00
hour
500 450 400 35O L)
Aluminum m
i
r
r
o
~
J
300
4,-'
m 250 0 o.
E 9200
I
I---
150
Rectangular p i p e /
100 50 -
0 10:30
Figure 4.- Scheme o f the procedure fo fix the conical mirror to the structure. Sheets of aluminium 0.5 mm thick with a 0.86 reflectivity are used to build the reflectors. The aluminium heat storage slab rear surface is insulated usin~ a high temperature low-density ceramic blanket. A 0.3 x 0.3 m" glass fixed to the mirror structure protects the front of the slab and decreases the heat losses.
3.
EXPERIMENTAL RESULTS
The measured experiments were carried out heating each slabs with an electric heater up to 400 C to obtain uniform conditions, allowing the comparison of the results obtained at different times. The slabs are changed every one and half-hours approximately.
11:30
12:30
13:30
14:30
_
iii
15:30
....
16:30
17:30
Hour
Fig. 5-a and 5-b: Tray and slabs temperature during the experimental time interval for two different days. The upper curve with a serrated shape is the slab temperature, the vertical lines indicates when the slabs are changed. The middle curve is the water temperature in the lower tray, and the other is the temperature in the water of the upper tray that is maintained at ambient temperature It takes almost three hours to reach the operative temperature at 95 C. When this temperature is reached it is necessary to put an aluminium screen between the slabs and the tray to decrease the heat transfer and maintain the temperature constant. This behaviour can be seen in Fig. 5 between 13:00 and 15:00 hours and in 5 b between 12.30 and 15:00 hours, it takes a larger time before changing slabs. The distilled water production was obtained manually. Fig. 6 shows the distillate production for different average
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Table 1: Different values ofrp
temperatures that was measured in the lower tray. The temperature in the upper Way was kept constant around 18 C.
~
AT
Qs~bs
md
Qdestillate
(c)
(MJ)
(kg)
(MJ)
rp
Tavera ~e
1,8
236
3.57
2,25
5.40
1,51
97
1,6
236
3.57
1,8
4.32
1,21
97
245
3.70
1,85
4.44
1,20
98
185
2.80
1,35
3.24
1,16
97
209
3.16
2,7
6.48
2,05
96
1,4
~" 1,2
i ~0,8
In previous works (1) we have reported a value for rp around 2.7 for the same still heated electrically or with natural gas. This value did not considered the losses in the insulated box below the Way. The smaller rp measured here indicate that some losses in the isolated chamber are being produced.
Q 0,6 0,4 0,2 0 80
82
84
86
88 90 92 94 Average Temperature C
96
98
100
Rp vs Tmedia 2,50
Fig 6: Distilled water vs. average temperature 2,00
The best measured production is 1.8 It/hour at 96 C, the values for points on the right are smaller, probably due to observed vapour leaks from the Way. A performance ratio, rp, giving the relation between the heat needed to evaporate the produced water and the consumed energy was calculated from the experimental results. The results are shown in figure 7.
rp =
Qdistillate Q lab
1,50
i
AA 1,00
&
A
0,50
0,00 80
,
,
,
85
90
95
100
Tempemtura C
The amounts of heat were calculated as
Fig 7: The Performance ratio rp vs average temperature of the lower Way
Qdistillate - m d f~ md -
distilled water mass
1-2.4~
4. CONCLUSIONS
MJ
Kg
Qslabs = 4 m C p s A T m = slabmass = 4kg J
Cps = 9 4 5 ~
AT = Yfinal - Tinitia I Table 1, shows the different parameters used for the calculation of rv they were obtained during different days
The new still is quite compact and maintenance problems are kept to a minimum since it is completely built in glass, no cotton fabric is used and it is fed with salty water in a single point. The still starts to produce distilled water very quickly at the beginning of the day since its thermal inertia is quite low due to the small amount of water in the stages and the Way. The productivity of the still, about 1.8 kg/hour at the higher temperatures, is good. Values of rp lower than 2 indicate that heat losses in the insulated box should be better controlled. The experiments have shown that the aluminium slabs are too heavy for an easy manipulation at high temperature during long time intervals. A new design of the slabs is being considered to solve the problem.
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5. ACKNOWLEDGEMENTS This work was partially supported by the CONICET (Consejo de Investigaciones Cientificas y T6cnicas). J. Franco and L. Saravia are researchers from the CONICET. The authors appreciate the collaboration of R. Caso and C. Fernfindez from the Universidad National de Salta to build the prototype.
6. REFERENCES
(1) Franco, J., Saravia, L., A New Design For Passive Atmospheric Multistage Still, Renewable Energy, Vol. 4, N ~ 1, pp 119- 122, 1994. (2) Franco, J., Destilador De Baja Pendiente, Avances en Eenergia Renovables y Medio Ambiente Vol 1, No. 1, pp. 6568, 1997.
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DEVELOPMENT OF A SMART SOLAR TANK Simon Furbo and Elsa Andersen Department of Buildings and Energy, Technical University of Denmark, Building 118, DK-2800 Lyngby, Denmark
Abstract - Theoretical and experimental investigations of small SDHW systems based on so-called smart solar tanks are presented. A smart solar tank is a hot-water tank in which the domestic water can both be heated by solar collectors and by an auxiliary energy supply system. The auxiliary energy supply system heats up the hot-water tank from the top and the water volume heated by the auxiliary energy supply system is fitted to the hot-water consumption and consumption pattern. In periods with a large hot-water demand the volume is large, in periods with a small hot-water demand the volume is small. The investigations showed that the yearly thermal performance of small SDHW systems can be increased by up to about 30% if a smart solar tank is used instead of a traditional solar combi tank. The thermal increase is strongly influenced by the hot-water consumption and consumption pattern. Recommendations for future development of smart solar tanks are given.
1. INTRODUCTION Almost all small solar heating systems for domestic hot water supply, SDHW systems, for single-family houses in Denmark are single-tank systems based on a combi hot-water tank. The domestic water in the eombi hot-water tank can be heated both by the solar collectors and by means of an auxiliary energy supply system. The water at the top of the combi hot-water tank is heated to a required temperature by means of the auxiliary energy supply system. In this way the top of the tank is always kept at a high temperature level. The volume of the water at the top of the tank heated by the auxiliary energy supply system is determined by the design of the tank. In the marketed Danish combi tanks this volume is sufficiently large for families with relatively large hot-water consumption. For marketed solar tanks with total volumes between 155 1 and 390 1 the top volume is situated in the interval from 601 to 1601. Measurements by Otto et al. (1997) have shown that most Danish families today have a relatively small hot-water consumption of about 60-150 l/day. The average hot-water consumption for a family is about 100 l/day. Measurements have also shown that the hot-water consumption and the consumption pattern vary strongly from family to family, and that the hot-water consumption and consumption pattern are not the same for a specific family during all periods of life. Further, increased water price and water saving equipment will most likely result in decreased hot-water consumption in the future. Furthermore, the hot-water consumption is normally not known before solar heating systems are installed. Obviously it is very difficult to choose the volume of the combi hot-water tank and the top volume of the tank in the right way. It is also obvious that the marketed solar tanks are oversized for typical hot-water consumption. Theoretical investigations have shown, Furbo and Shah (1996) and Shariah and L6f (1997), that the thermal performance of typical solar heating systems based on combi tanks can be strongly influenced by the hot-water consumption pattern and that the thermal performances of combi tank systems are much smaller than the thermal performances of preheating solar
heating systems with tanks which can only be heated by solar collectors. The ideal solar heating system from a thermal and energysaving point of view is therefore based on a preheating tank which can only be heated by the solar collectors and an auxiliary energy supply system built into the hot-water pipe from the tank close to the tapping locations. The auxiliary energy supply system heats up the domestic water instantaneously to the required hot-water temperature during tappings. In this way the thermal performance of the solar heating system is maximized since the operation temperature of the solar collectors is reduced to a minimum and the heat loss from the hot-water pipe and from the auxiliary energy supply system is minimized. However, a large power supply from the auxiliary energy supply system is required in order to maintain a reasonable hot-water comfort. For instance, a power supply of 20 kW is needed to heat cold water from 213~ (cold water temperature in March in Denmark) to 5013~ for a tapping flow rate of 6 l/rain. With a tapping flow rate of 12 1/min. 40 kW is needed. The maximum power supply from typical oil-fired boilers or natural gas burners for one-family houses in Denmark is about 20 kW. Therefore the hot-water comfort will not be sufficiently high for typical boilers/burners. Consequently the pure preheating system is not attractive in most houses. 2. SMART SOLAR TANK 0 PRINCIPLE The advantages of the pure preheating system O the large thermal performance of the solar collectors and the small heat loss from the auxiliary energy system n are to a certain extent also obtained in systems making use of a so-called smart solar tank. Investigations have thus indicated, Furbo and Shah (1997) that the thermal performance of small SDHW systems can be increased if a smart solar tank is used as the heat storage instead of a marketed combi tank. Fig. 1 shows schematic illustrations of the auxiliary energy supply system of a typical marketed combi tank and of a smart solar tank.
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marketed solar tank
smart solar tank
3. TESTED SYSTEMS Three small low flow SDHW systems have been tested under the same realistic conditions side-by-side in a laboratory test facility. Two of the systems are based on differently designed smart solar tanks and one system is a traditional system. Fig. 2 shows schematic illustrations of the three systems. All three systems are based on inexpensive vertical mantle tanks used in many domestic hot water systems in Denmark. The solar collector fluid from the solar collector enters the top of the mantle and returns from the bottom of the mantle to the solar collector. Electric heating elements are used as the auxiliary energy supply systems.
J Fig. 1. Schematic illustration of the auxiliary energy supply system of a typical marketed solar tank and of a smart solar tank. In the marketed solar tank the constant top volume of the tank is always heated to a required temperature by the auxiliary energy supply system. In the smart solar tank the auxiliary energy supply system can be built into a side-arm from the middle to the top of the tank. In periods with energy supply from the auxiliary energy supply system heat is transferred from the auxiliary energy supply system to the domestic water in the side-arm. By means of thermosyphoning in the side-arm/tank loop the hot water is transferred to the top of the tank. In that way the tank is heated from the top. The energy supply from the auxiliary energy supply system can be controlled in such a way that the energy content in the top of the tank during all hours can have a predetermined (variable) minimum quantity. In periods with a large hot-water demand the energy content can be large and in periods with a small hot-water demand the energy content can be small. That is: The water volume heated by the auxiliary energy supply system is fitted to the hot-water consumption and the consumption pattern. In most periods the hot top volume is much smaller in the smart solar tank than in the marketed solar tank. The heat loss of the smart solar tank is therefore smaller than the heat loss of the marketed solar tank. Further, the solar volume of the smart solar tank is greater than the solar volume of the marketed solar tank and the thermal performance of the solar heating system is increased if a smart solar tank is used instead of a marketed solar tank.
For the smart solar tanks the electric heating elements are built into a side-arm, which in one system, by means of a plastic pipe from the middle to the bottom of the hot-water tank, connects the middle of the hot-water tank to the top of the hot-water tank. In the other system the side-arm connects the middle of the mantle to the top of the mantle. In periods with the electric heating element in operation heat is transferred from the electric heating element to the domestic water/solar collector fluid in the side-arm. By means of thermosyphoning in the side-arm/hotwater tank loop or in the side-arm/mantle loop the heat is transferred to the domestic water located at the top of the hotwater tank. For increasing duration of the operating time of the electric heating element the volume of the water at the top of the tank heated is increasing. The data of the solar collector used in each of the three systems are given in Table 1 and the data of the tested systems are given in Tables 2 and 3.
Electrlc heatlnc element
/
a r m
electric heating
J
] I: ,.~
heating : ~ element
Cold water
Hot water
Cold water ~
Trad/tional system
Fig. 2. Schematic illustrations of the three tested systems.
J
-
/
A
.~
~
-"
Hot ware
Thermosyphovd~,_m in the eide-erm/tenk loop
Col
Hot water
Thermosypho~tn_m in the eide erm/menUe loop
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Table 1. Data of the solar collector used in the tested systems. Area
3.00 m 2
Efficiency for small incidence angles
[] Fq0.756 []4.37 FqTm []Ta [30.010 (Tm []Ta) 2 E E
Incidence angle modifier
1-tg3s ([3/2)
Heat capacity
5000 J/K m 2
Tilt
45E]
Orientation
south
Table 2. Data of the solar collector loop and control system used in the tested systems.
Traditional system
Smart solar tank system
Smart solar tank system
Thermosyphoning in the sidearm/tank loop
Thermosyphoning in the sidearm/mantle loop
Solar collector loop Pipe material
Copper
Copper
Copper
Diameter
10/8 mm
10/8 mm
10/8 mm
Length of pipe from solar collector to storage, outdoor
10.0m
10.0m
10.0m
Length of pipe from storage to solar collector, outdoor
13.3 m
13.3 m
13.4 m
Length of pipe from solar collector to storage, indoor
5.1 m
5.1 m
5.1 m
Length of pipe from storage to solar collector, indoor
4.6 m
4.6 m
4.5 m
40% (weight) propylene glycol/water mixture
40% (weight) propylene glycol/water mixture
30% (weight) propylene glycol/water mixture
Volume flow rate in solar collector
10"Sm3/s
10Sm3/s
10Sm3/s
Power of circulation pump
35 W
50 W
65 W
6 K/2 K
6 K/2 K
6 K/2 K
Solar collector fluid
Control system Differential thermostat control with one sensor in the solar collector and one at the bottom of the mantle start/stop difference
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Table 3. Data of the solar tanks in the tested systems.
Traditional system
Smart solar tank system
Smart solar tank system
Thermosyphoning in the sidearm/tank loop
Thermosyphoning in the sidearm/mantle loop
Solar tank Tank material
Steel St 37-2
Steel St 37-2
Steel St 37-2
0.175 m 3
0.175 m 3
0.175 m 3
1.484 m/0.394 m
1.484 m/0.394 m
1.484 m/0.394 m
0.003 m
0.003 m
0.003 m
0.009 m 3
0.029 m 3
0.058 m 3
0.700 m/0.425 m
0.700 m/0.473 m
1.285/0.473 m
0.002 m
0.002 m
0.002 m
Hot-water tank Domestic water volume Height/diameter Material thickness
Mantle Volume Height/diameter Material thickness Location
The mantle surrounds the hot-water tank. The upper 0.081 m 3 and the bottom 0.009 m 3 of the hot-water tank are not surrounded by the mantle
The mantle surrounds the hotwater tank. The upper 0.081 m 3 and the bottom 0.009 m 3 of the hot-water tank are not surrounded by the mantle
The mantle surrounds the hotwater tank. The upper 0.009 m 3 and the bottom 0.009 m 3 of the hot-water tank are not surrounded by the mantle
Auxiliary energy supply system
Upper 0.072 m 3 of the hotwater tank is heated to 50[X~ by the electric heating element
Side-arm from middle to top of hot-water tank. Electric heating element built into side-arm
Side-arm from middle to top of mantle. Electric heating element built into side-arm
Volume in side-ann: about 0.4 1/min.
Volume in side-arm: about 0.3 l/min.
The electric heating element is in operation if the energy content of the domestic water with temperatures higher than 50[X~ is too small to cover the hot-water demand completely with a minimum tapping temperature of 45[X~ and if the difference between the time and the predetermined tapping hours is smaller than 2 89hours
The electric heating element is in operation if the energy content of the domestic water with temperatures higher than 5 0 ~ is too small to cover the hot-water demand completely with a minimum tapping temperature of 45[X~ and if the difference between the time and the predetermined tapping hours is smaller than 2 89hours
1060 W
1120 W
1140 W
Mineral wool
Mineral wool
Mineral wool
Top
0.25 m
0.25 m
0.25 m
Side
0.05 m
0.05 m
0.05 m
Bottom
0.05 m
0.00 m
0.00 m
Power supply of electric heating element
Insulation Material Thickness
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4. TEST RESULTS
The three systems were tested side-by-side under the same realistic conditions: The solar irradiance on the collectors and the daily hot-water consumption of 183 1 is the same for all three systems. An energy quantity of 2.44 kWh, corresponding to 61 1 of hot water heated from 10IX2 to 45EE, is tapped from each system three times each day: 7 am, 12 am and 7 pm. The tests were carried out during April-June 1999. So far 8 test periods of 4-5 days' duration have been carried out. The measurements showed that the thermal performance of the traditional solar heating system is always very close to the thermal performance of the smart solar tank system with thermosyphoning in the side-ann/mantle loop. In some periods the traditional system performs best, in some periods the smart solar tank system performs best. The measured net utilized solar energies for the 8 test periods are given in Fig. 3 for the traditional solar heating system and for the smart solar tank system with thermosyphoning in the side-arm/tank loop. The net utilized solar energy is defined as the tapped energy from the solar tank minus the energy supply to the electric heating element.
The smart solar tank system with thermosyphoning in the sideann/tank loop has always a higher thermal performance than the traditional solar heating system. For the whole test period of 35 days' duration the net utilized solar energy was 9% higher for the smart solar tank system than for the traditional solar heating system.
5. CALCULATIONS
A simulation model for the smart solar tank system with thermosyphoning in the side-arm/tank loop has been built up and validated by means of measurements. The yearly thermal performance of small SDHW systems based on differently designed and controlled smart solar tanks were calculated with the model. Also calculations for traditional solar heating systems based on marketed solar tanks were carried out. All the systems taken into calculation are identical except for the solar tank. The data of the system taken into calculation is given in Table 4. The data of the two marketed solar tanks are given in Table 5 and the data of the smart solar tank are given in Table 6. Different daily quantities of hot-water consumption are assumed. The water is heated from 10~ to 50~ A third of the daily hot-water consumption is tapped three times each day. The weather data of the Danish Test Reference Year is used in the calculations. Fig. 4 shows the calculated net utilized solar energy and performance ratio of the systems. The net utilized solar energy is the tapped energy minus the energy supply from the auxiliary energy supply system. The performance ratio is the ratio between the net utilized solar energy for the system with the heat storage in question and the net utilized solar energy for the system with the Danlager 1000 heat storage. The thermal performance of the system is increasing for increasing hot-water consumption. For small hot-water consumption the thermal performance is not strongly influenced by the storage volume of the marketed solar tanks and by the consumption pattern. For increasing hot-water consumption the influence of both the storage volume and the consumption pattern on the thermal performance is increasing. If, for instance, hot water is only tapped during evenings the thermal performance of the systems with the marketed solar tanks is relatively small and the thermal performance is relatively strongly increased by increasing the storage volume.
Fig. 3. Measured net utilized solar energy for the traditional solar heating system and for the smart solar tank system with thermosyphoning in the side-arm/tank loop for 8 periods of 4-5 days' duration.
The thermal performance of the solar heating system with the smart solar tank is higher than the thermal performance of the system with the marketed solar tank with the same total tank volume. The smart tank system can have a thermal performance up to about 30% greater than the thermal performance of the traditional system with the same total tank volume.
ISES Solar World Congress 1999, Volume III
The smart solar tank is especially attractive if hot water is only tapped in the evenings. The thermal advantage is somewhat smaller if hot water is only tapped in the mornings and the thermal advantage is smallest if water is tapped in the morning, at noon and in the evening. Fig. 5 shows calculated net utilized solar energies of the system with different smart solar tank designs. The tank is designed as indicated in Table 6 with only one parameter changed at a time. The daily hot-water consumption is 1601 and hot water is tapped at 7 am, 12 am and 7 pm.
165
From the figure it is obvious that the side-arm outlet should be placed as high in the tank as possible with regard to the hotwater comfort and that the control system should stop the supply from the auxiliary energy supply system when the energy content at the top of the tank is as low as possible, of course also with regard to the required hot-water comfort. A variable flow rate in the side-arm resulting in a constant inlet temperature to the tank a little higher than 50[Z; or a high constant flow rate results in the highest thermal performance.
Table 4. Data of the SDHW-system taken into calculation. Solar collector Area
3m 2
Efficiency for small incidence angles
[3 = 0.75 - 5.40. (Tin- Ta)/E
Heat capacity
7000 J/Kin2
Tilt
45 ~
Orientation
South
Solar collector loop Pipe material
Copper
Diameter
12/10 mm
Length of pipe from solar collector to storage, outdoor
1.5 m
Length of pipe from storage to solar collector, outdoor
1.5 m
Length of pipe from solar collector to storage, indoor
3.5 m
Length of pipe from storage to solar collector, indoor
3.5 m
Heat loss coefficient of pipe
0.25 W/mK
Solar collector fluid
40% (weight) propylene glycol/water mixture
Volume flow rate in solar collector loop
7.5-10.6 m3/s
Power of circulation pump
3O W
Control system Differential thermostat control with one sensor in the solar collector and one at the bottom of the mantle Start/Stop difference
10K/2 K
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Table 5. Data of two marketed solar tanks taken into calculation. Name Type Tank material
Danlager 1000
Daulager 2000
Mantle tank Steel St 37-2
Mantle tank Steel St 37-2
0.189 m 3 1.082/0.500 m 0.003 m Upper 0.080 m 3 of the hot-water tank is heated to 50.5~ by the auxiliary energy supply system
0.265 m 3 1.450/0.500 m 0.003 m Upper 0.089 m 3 of the hot-water tank is heated to 50.5~ by the auxiliary energy supply system
0.007 m a 0.395/0.525 m 0.003 m The mantle surrounds the bottom part of the hot-water tank. The upper 0.094 m 3 and the bottom 0.019 m 3 of the hot-water tank are not surrounded by the mantle
0.012 m 3 0.715/0.525 m 0.003 m The mantle surrounds the bottom part of the hot-water tank. The upper 0.109 m 3 and the bottom 0.019 m 3 of the hot-water tank are not surrounded by the mantle
PUR-foam 1.7 W/K
PUR-foam 2.3 W/K
Hot-water tank Volume Height/diameter Material thickness Auxiliary energy supply system
Mantle Volume Height/diameter Material thickness Location
Insulation Insulation material Heat loss coefficient
Table 6. Data of the smart solar tank taken into calculation Mantle tank Steel St 37-2
Type Tank material
Hot-water tank 0.189 m 3 1.082/0.500 m 0.003 m
Volume Height/diameter Material thickness
Mantle 0.007 m 3 0.395/0.525 m 0.003 m The mantle surrounds the vertical sides of the hot-water tank. The upper 0.094 m 3 and the bottom 0.019 m 3 of the hot-water tank are not surrounded by the mantle.
Volume Height/diameter Material thickness Location
Side-arm and auxiliary energy supply system Side-arm location Auxiliary energy supply system Power of electric heating element Volume flow rate in side-arm in periods with energy supply Control system
801 located above the side-arm's outlet pipe. Side-arm inlet connected to top of hot-water tank Electric heating element built into the side-arm 1200W 6.67[:106m3/s The electric heating element is in operation if the energy content of the domestic water with temperatures higher than 50~ is too small to cover the hot-water demand completely with a minimum tapping temperature of 50[]2, and if the difference between the time and the predetermined tapping hours is smaller than 2.5 h.
Insulation Insulation material Insulation thickness
Top Side Bottom
PUR foam 0.10m 0.05 m 0.05 m
ISES Solar World Congress 1999, Volume III
167
Hot water tapped at 7am, noon and 7pm
Hot water tapped at 7am, noon and 7pm 1.3
-.- o a . ~ e ,
1300 @
1200
C9
1100
,-'C'
lOOOl
_o .m
_~ ~1ooo
1.2
1.15
o
>,
o
-.4-Danlager 1000 -4,--Danlager2000 Smart tank
1.25 -*- Smart tank
~
1.1 1.05
_m Q
9
O.
600
z
0.95 4oo 0
0.9
,
~
,
50
100
150
0
200
50
150
200
Hot water tapped at 5am, 6am and 7am
Hot w a t e r tapped at 5am, 6am and 7am 1.3
14oo
-o- Danlager 1000 Danlager 2000 -*-" Smart tank
1300 1200
0
100
Hot water consumption [I/day]
Hot water consumption [i/day]
-~- Danlager 1000 -,,- Danlager2000 -*- Smarttank
1.25
_o .m
r ,~,11oo (g ~ 1 0 0 0
1.2
1.15
Q
~
1.1
m
[E 1.115 700 9 Z
o
600
0.95
5OO 4oo 0
,
,
,
50
100
150
0.9 200
0
,
i
,
50
100
150
Hot water consumption [I/day]
200
Hot water consumption [I/day]
Hot water tapped at 7pm, 8pm and 9pm
Hot w a t e r tapped at 7pm, 8pm and 9pm 1.3 Danlager 1000 -'- Danlager 20001 -dr- Smart tank
1300 1200
_~..~~
Z
O
r Q
--=-Danlager I000 --- Danlager2000 -,-- Smarttank
1.25
O
11oo
,.,~, (g =1000 O >~
~
1.2
1.15
o
~
1.1
m 9
800
~ 1.05
9 Z
600
a.
a
0.95
500 40o 0
,
l
,
50
100
150
Hot water consumption [I/day]
0.9 200
0
50
100
150
200
Hot water consumption [I/day]
Fig. 4. Yearly net utilized solar energy and performance ratio as a function of the hot-water consumption for different consumption patterns and heat storage.
ISES Solar World Congress 1999, Volume III
168
1350
>,
1300
cO Q
1250
e==
1200
m : 4,,
1100
Z
4O5O
0
1000 0.4
1300 1250
~'~
1200
=
1100
0 Z
0.45
0.5
0.55 0.6 0.65 0.7 M a n t l e h e i g h t [m]
0.75
1350 / ..._.-~-----, ,_.____..__ ,._ c9 o
9 0
K,7
0.8
1050 1000
0.65
0.2 0.4 0.6 0.8 1 Side-arm outlet pipe's distance from t o p o f tank [m]
Constant hl~t temperatureto the hot water tank fn:xn the side-arm
o
1250
i: 4)
1250
,Iz
ID
m "~
1100
Z
=
1100
o Z
4O50
1050
1000 50 2 4 6 8 10 Power of auxiliary energy s u p p l y system [kW]
65
70
75
80
85
90
95
1350
>~
1300
ro Q
1250
.ram m
1200
m ~
1100
u
Z
50
Inlet t e m p e r a t u r e t o t h e h o t w a t e r t a n k f r o m t h e s i d e - a r m [ ' C ]
1350
=3
55
1200
J
7
1100
Q
1050
Z
1000 1 2 3 4 5 Lower limit of energy content of consumption water with t e m p e r a t u r e s h i g h e r t h a n 50=C [ k W h ]
6
1050 1000 0.3
0.5
0.7
0.9
1.1
1.3
1350
~)
1250
0 C Q
~'~'U
1200
~'~
1150
m 1050 z
lOO0 95o 0
50
1.5
1.7
V o l u m e flow rate in side-arm [I/min]
100
150 Tank Volume
200
250
300
[I]
Fig. 5. Yearly net utilized solar energy of the solar heating system with differently designed smart solar tanks.
1.9
2.1
ISES Solar World Congress 1999, Volume III
6. DISCUSSION AND CONCLUSION
Investigations of smart solar tanks based on a mantle tank with a side-arm in which an auxiliary energy supply system is built in have been carried out. Both tanks with the side-arm connected to the hot-water tank and to the mantle have been investigated. The investigations showed that the tank with the side-arm connected to the hot-water tank is the best tank from a thermal point of view. Further, the investigations showed that the thermal performance of solar heating systems can be improved by up to about 30% by making use of such a smart solar tank. The thermal advantage of smart solar tanks is largest if hot water is not tapped during the light hours of the day. It is estimated that the costs of a typical small SDHW system will be increased by about 3% by making use of a smart solar tank. Consequently the performance/cost ratio can be improved by up to about 25% by making use of a smart solar tank. So far, detailed investigations have only been carried out for smart solar tanks with the auxiliary energy supply system built into a side-arm. However, smart tanks can be designed differently. Preliminary investigations indicate that the auxiliary energy supply system can be built into the hot-water tank in such a way that thermal stratification is built up as good as or even more efficiently than in tanks with the side-arm connected to the hot-water tank. These investigations will be finished during the summer of 1999. From the autumn of 1999 the two most promising designs of smart solar tanks will be tested in two small SDHW systems in practice. The thermal performance of the systems will be measured during the first year of operation. In this way possible operation or durability problems connected to the designs of smart solar tanks will be elucidated. It will also be elucidated if the hot-water comfort can be accepted by the consumers and if the thermal performances of the systems in practice are as good as expected. Based on the promising results it is recommended to start work to develop smart solar tanks based on other auxiliary energy supply systems than electric heating elements. In order to further improve smart solar tank systems it is also recommended to start work to develop a smart control system both for the energy supply from the auxiliary energy supply system and for the pump in the solar collector loop. Most likely, the system performance can be somewhat increased if the flow rate in the solar collector loop is controlled in such a way that water in the top of the hot-water tank is heated by the solar collector to a directly usable temperature. The flow rate in the solar collector loop will therefore vary from one period to another.
169
REFERENCES
Furbo S. and Shah L.J. (1996) Optimum solar collector fluid flow rates. EuroSun '96. 10. Internationales Sonnenforum Proceedings, Freiburg, Germany. Book 1, 189-193.
Furbo S. and Shah L.J. (1997) Smart Solar Tanks [3 Heat Storage of the Future? Proceedings of ISES 1997 Solar World Congress, Taejon, Korea.
Otto W, Nielsen J.E. and Dalsgaard Jacobsen T. (1997) Ydelsesstatistik for mindre brugsvandsanl~eg [3 erfaringer fra det femte ~ s rrfilinger 1996. Danish Solar Energy Testing Laboratory.
Shariah A.M. and L6f G.O.G. (1997) Effects of auxiliary heater on annual performance of thermosyphon solar water heater simulated under variable operation conditions. Solar Energy 60, 119-126.
170
ISES Solar World Congress 1999, Volume III
THERMAL
MODELLING
AND PERFORMANCE
PROCESSES
UNDER
PREDICTION
OF DRYING
OPEN-SUN-DRYING
H.P. Gar.q and Rakesh Kumar Centre for Energy Studies, Indian Institute of Technology, Hauz Khas, New Delhi - 110 016, India Tel.: +91-11-6861977, Fax: +91-11-6862037, E-mail: [email protected] Abstract - An analytical model has been developed for the drying characteristics of any product under open-sundrying(OSD). The model is based on the theory of 'Generalised drying curve'(GDC) in terms of receding front. The developed model can be used to ascertain the drying characteristics of any product under OSD. The model has included all the climatic and product parameters explicitly. The main concern of the present study is to estimate the effect of product thickness and climatic conditions on the drying rate and drying time. The numerical calculations have been made for the climate of Delhi. For the calculations, the chosen crop is grapes(initial moisture: 82%). The results have been plotted for both s u m m e ~ a y ) and winter(December) conditions of Delhi. It is noticed that under open-sun-drying conditions, the drying rate depends significantly on the product thickness and climatic conditions. The thickness of the product is taken(or kept) as small as possible for fast and quality drying under OSD. Also, the developed model for drying characteristic under OSD has been validated with published experimental observations on typical crop under Delhi climatic conditions.
1. INTRODUCTION
final moisture:
12%). The results are plotted for both
summer(May) and winter(December) conditions of Delhi. It is The most widely used method for crop drying is open-sundrying(OSD). This is simplest method for product dehydration. In OSD, the product is spread in a thin layer on the horizontal ground and exposed directly to the solar radiation, wind and other atmospheric conditions. In this type of drying, heat is transferred from the surrounding air and the sun to the exposed stnface of the product. A part of this heat is travelled to the product interior to rise its temperature and remove the moisture from the product interior to its surface. The remaining heat is utilized to evaporate the moisture from the product surface to the SUl'rounding air. This process of heat and mass transfer has occurred simultaneously in OSD. The rate of drying depends on the number of external parameters(solar radiation, ambient temperature, wind velocity and relative humidity) and the internal parameters(initial moisture content, type of the crop, mass of the product per unit exposed
seen that under OSD conditions, the drying rate depends on the product thickness and climatic conditions. The thickness of the product is taken(or kept) as small as possible( from 2-4 cm) for fast and quality drying. The present model on OSD has been validated with the experimental results of typical crop(mango) under Delhi climatic conditions. 2. MATHEMATICAL ANALYSIS
area, etc.). Some theoretical and experimental study on OSD are reported in the literature (Garg, 1987). However, the available study is not sufficient for full understanding of the drying processes in OSD. The principle of drying in OSD is different from solar drying. There are enough literatures available on solar drying (Muhlbauer, 1981, 1986; Garg and Kumar, 1998; Cfarg et al, 1998) whereas, very few work has been done on OSD. Sodha et al (1985) has developed an analytical model for OSD. The developed model has considered a variation in the product temperature along its thickness and has applied thin layer drying equation for moisture evaporation. The analysis has not taken into account the heat capacity effect of the product, product thickness and the quantity of the product per unit exposed area to the sm'rounding air. In this paper, the analysis and the numerical results for a typical crop are presented under OSD. The developed model has used the receding front phenomena in which product can be considered into two parts, viz wet part and dry part. Earlier, the study on 'Generalised drying curve'(GDC) was proposed by Ratti and Crapiste (1992) and extended by Chou et al (1997). The present study is the extension of the work of Chou et al. Also, the model has included all the climatic and product parameters explicitly. The developed model can be used to identify the drying characteristic of any product under OSD. The numerical calculations have been made for the Delhi climate. For the calculations the chosen crop is grapes(initial moisture: 82% and
Product 9
-
L .....
~
r,-,
lffftt
Lo..
; :~
(a) Fig. l(a) Schematic of open-sun-drying procedure The schematic of open-sun-drying(OSD) is shown in Fig. 1.. Following assumptions are made in the present analysis: (1) Thermal properties of dry product is constant. (2) The product is considered as a material of uniform thickness. (3) There is no heat conduction in the product slab. (4) No volume shrinkage of dried product. (5) No temperature and moisture gradient along the product thickness. (6) The temperature in the ground at 4 m depth is constant and taken as 24 ~ for Delhi conditions. The heat and mass balance in the OSD is expressed in terms of the following equations.
ISES Solar World Congress 1999, Volume III
171
2.1 Equation for moisture evaporation The present study is based on the concept of receding front phenomena. Ratti and Crapiste has developed this concept on the basis of 'Generalized Drying Curve'(GDC). In this model, the GDC is independent on the drying conditions, type of the product to be dried and is a function of the moisture content in the product. This model is fuaher simplified by Chou et al and used in the present study. In this model, the product has been considered into two parts viz., dry part and wet part. These two parts are separated by an arbitrary curve known as receding front. During the drying process, the receding front moves from dry part to wet part. The position of the receding front at any moment of the drying is shown in Fig. l(b).
1=11+12+13+14
(6)
in which I is solar radiation absorbed by the product, I1 heat losses to the product due to convection and radiation, 12 heat stored as a sensible heat in the product, 13 heat flux losses to the ambient due to evaporation and !4 heat flux conducted into the ground. The above Eq. in terms of different temperatures on simplification is written as,
Tp ( ~TL) ( "-~-) ~M . mpcp'-'~'-=Qp-hg(Tp-Ta)-hfg t)TG/y'aPo -KG'--~- =
h
2.3 Equation for ground temperature The temperature variation inside the ground To is characteristic of the heat conduction Eq. as,
Ms -x ----_____
zl
(7)
_~_x
Wet part
/
O2TG igTG K G ~)y2 = P G CG Ot
_=Wsat
(8)
Mo This conduction Eq. is applicable with the following boundary conditions,
(b) Fig. l(b) Schematic of receding front model
Tp/y=o= TG/y=O ,
The moisture profile in the product at any time is represented by the following relations,
To get the instantaneous values of various parameters (moisture content, product temperature, ground temperature), the developed heat, mass and temperature balance equations are solved by using finite difference technique. Equations for moisture evaporation, product temperature and ground temperature are re-written as, For moisture content,
M=Mo
O
( M-Ms ) (l-z) = Mx-Ms 1-x
(1)
x
TG/y=4m= 24~
(9)
(2)
MJ+I = M j From above two Eqs., we get,
kg At Ap (WJsat- WJa) mp (l + Bim (O) )
(10)
For product temperature
M=Ms+(Mx-Ms)(1-z) 1-x
x
(3)
Following the analysis developed by Chou for identical case of drying, the drying rate is expressed by
aM _ Ap kg [ ]Wsat + - ~ -mWa -~)l. bt mp
(11)
(WJat WJa) KG( Tj-TG(1)j )] _
_
Ay
(4)
in which W~t is the saturation humidity of air at the product temperature. Bim is the biot number and it depends on the diffusion coefficient. ~ is Generalised drying parameter(GDP) and its approximated value is calculated by using r = 1- M
TY+I Tip+ At [QJg_hg(T j-TJa) - ( hfgkg ) P = mpCp l+Bim(#p)-
(5)
For ground temperature
TG (i )j+l = TG(i )j + At K G Paca
(12)
( TG(i + I ) J - 2 T G O ) J + T G O - 1 ) j ) A x2
Mo 2.2 Equation for product temperature The product is assumed to be a compact slab of dry product and moisture. The quantity of the product per unit exposed area depends on its thickness and density of the crop. The heat balance over a unit area of the product slab is written as,
where i=2, N-l, N is the number of last ground layer at 4 m depth from the surface. For the eonvergence of Eq. 11, the following condition would be satisfied
ISES Solar World Congress 1999, Volume III
172
A t <_Pa ca( Ax)2
V o p
(13)
2 kG
3 m/s 5.67xl 0s W/m2-K4 101325 Pa
and Ax >_~ff 2 kG At ) Pac6
(14)
the value of Ax is calculated using above equation. 3. RESULTS AND DISCUSSION The numerical calculations on the basis of developed model for the drying characteristics of high moisture grapes have been carried out under open-sun-drying(OSD). The calculations have been made for two extreme conditions of Delhi climate viz. May(summer) and December(winter). The input climatic parmneters (global solar radiation, ambient temperature and relative humidity) for each chosen month are given in Table 1 and are taken from Mani and Rangarajan (1981). The values of the other thermo-physical parameters are given in the Table 2. The computer simulation has been carried out in FORTRAN77. The values of heat transfer coefficients are calculated by using Duffle and Beckman (1991). Table 1.
The effect of product thickness on the drying time and drying characteristics are given in Fig.2(a) for the month of May. The chosen typical product thicknesses are 2, 4, 6 and 8 cm. As seen from Fig.2(a), the higher product thickness increases drying time and reduces drying rate. The more product thickness means more product mass corresponding to same exposed area~ As thickness increases from 2 to 8 cm, the amount of product per unit exposed area has changed from 16.8 kg to 67.2 kg. It is also seen from this Fig. that for very high product thickness(8 cm), the product is not reached to its final moisture even after 5 sunshine days.
0
.9--
.
Solar radiation (W/m2) May Dec.
Ambient Temperature (0{2) May Dec.
6 7 8
30 177 383
0 3 85
26.3 27.1 29.3
9.7 9.5 9.8
46 46 43
85 86 86
9 10 11 12 13 14 15 16 17 18 19
581 746 865 925 920 844 719 547 352 158 27
259 425 543 605 605 538 418 255 86 3 0
31.8 34.1 35.7 36.9 37.7 38.3 38.5 387 38. 37.5 35.9 Table 2.
12.3 14.9 17.3 19.0 20.1 20.7 20.9 20.7 19.7 17.2 15.9
37 34 30 26 24 22 20 20 19 19 22
86 76 65 56 49 43 40 38 38 41 50
Parameters/constant Lo Mo pg Po kg K~ Cgp Co Ca Cv
Relative humidity(%) May Dec.
Numerical values 2, 4, 6, 8 era(for May) 1, 2, 3, 4 cm(for Dec) 0.82(db) 840 kg/m3 2050.56 kg/m3 0.418 W/m-~ 0.519 W/m-~ 2800 J / k g - ~ 1840 J/kg~ 1005.7 J/kg-~ 4190 J/kg-~
Mantra: MD/
O.7. ~
nmlm4m : 1R~.
0.6. ~ ~
.5.
~
0.4-
~
(
m
r
"
a .4 e
0.20-10
Time
"
0.8-
lb-
~
~
,/o ~
s'o houra (trs)
do
TO
Fig. 2(a) Variation of moisture content with time for various thicknesses under summer(May) conditiom The variation of the average product temperature with time for the month of May are plotted in Fig.2(b). The different craves represent the product temperature for different thicknesses. Only marginal temperature difference has been observed as the thickness of the product changes from 2 to 8 cm. It can be understood as the maximum fraction of total heat supplied to the product is used in the moisture removal and evaporation. The processes of heat and mass transfer continues upto the final moisture content and beyond that rapid increase in the product temperature has been observed. The profile of the moisture evaporation and product temperature for the month of December for the same product under different climatic conditions are plotted in Fig. 3(a) and 3(b) respectively. In this case, the product thickness are taken as 1, 2, 3 and 4 era as input heat is less. The pattern of the moisture evaporation and product temperature are almost identical to the Fig 2(a) and 2(b) and only difference in their magnitudes. Therefore, for the uniform, fast and quality drying under OSD, the thickness of the product is very important and is taken as small as possible. Also, climatic conditions for OSD are also equally important for quality of drying.
ISES Solar World Congress 1999, Volume III
Fig. 2(b) Variation of product temperature with time for various thicknesses under summer(May) conditions
173
Fig. 4 A comparison of theoretical and experimental results under OSD 4. CONCLUSIONS
Fig. 3(a) Variation of moisture content with time for various thicknesses under winter0)ecember) conditions
The following conclusions have been drawn from the present study: The developed analysis for the drying characteristics are very simple and can be used for any crop and location. The thickness of the product is kept as small as possible(2- 4 cm) for fast and quality drying under OSD. As thickness of the product slab has changed from 2 to 8 cm, the product has reached to its final moisture level(about 12%) in more than 60 hrs(it is 10 hrs for 2 cm thickness). The rate of drying for higher thickness is even more slow in winter conditions. The experimental validation of the present model has proved its utility to ascertain drying characteristics of any crop under OSD. NOMENCLATURE
A~ % cG Ca Cv
h~ h~ k~ kgl
I% Lo Fig. 3(b) Variation of product temperature with time for various thicknesses under winter(December) conditions The present model is also validated with the experimental observations of Sodha et al (1985). In Sodha et a1(1985), a series of experimental study under OSD was reported. The chosen product for the experimental study was mango(initial wt.---9 kg, product thickness about 1 em.). This amount of the product was exposed to the solar radiation and wind. The numerical calculations on the basis of present model for mango is also carried out for the experimental conditions. The variation of moisture content under theoretical model and experimental observations (Sodha et al, 1985) are plotted in Fig. 4. The close agreement in the theoretical and experimental results has proved the utility of the present model.
mp M
Mo Ms P
Q~ r~ T~ TG
v W~
Wa Pg Po
exposed area of the product slab (m2) product specific heat (J-kgl-~ "1) specific heat of ground (J-kg-1-~ specific heat of air (J-kgq-~ "1) specific heat of water vapour (J-kgl-~ "1) heat transfer coefficient from the product to air (W-m'2-~
1)
heat of evaporation of water (J-kgq) mass transfer coefficient (kg-m2-s"1) thelTnal conductivity of product(W-m"l-~ thermal conductivity of ground (W-m-l-~ -1) product thickness(m) product mass (kg) instantaneous moisture content(db) initial moisture content(db) moisture content on the drying from(db) atmosphedc pressure(Pa) solar radiation on the product (W-m-2) product temperature (~ ambient temperature (~ ground temperature (~ wind ~ m - s "1) saturation specific humidity of the air at the average product temperature~kg-kgl) ~ i f i c humidity of the ambient air (kg-kg"1) generalised drying parameter(dimensionless) density of product (kg-m"3) demity of ground (kg-m-3)
174
ISES Solar World Congress 1999, Volume Ill
REFERENCES
Chou, S.K., Hawlader, M.N.A., Chua, ICJ. and Teo, C.C.(1997), A methodology for ttmnel dryer chamber design, Int. J. Energy Res. 21,395-410. Duffle, J.A. and Beckman, W.A.(1991), Solar Engineering of Thermal Processes, 2nd ed., Wiley, New Yorlc Garg, H.P.(1987), Advances in Solar Energy Technology, (VoL3): Heating Agricultural and Photovoltaic Applications of Solar Energy, D. Reidel Publishing Company, The Netherlands. Garg, H.P. and Kumar, Rakesh(1998), Studies on semi-cylindrical solar tunnel dryer: Year round collector performance, Int. J. Energy Res. 22, 1381-1395. Garg, H.P., Kumar, Rakesh and Datta, G.(1998), Simulation model of the thermal performance of a natural convection type solar ttmnel dryer, Int. J. Energy Res. 22, 1165-1177. Mani, A. and Rangarajan, S.(1981), Solar Radiation over India, Allied Publishers, New Delhi. Muhlbauer, W.(1981), Solar drying of agricultural products, Proc. of Rural Development Technology, 25-29 May, Korea Advanced Institute of Science and Technology, Seoul, pp. 415-433. Muhlbauer, W.(1986), Present status of solar crop drying, Energy Agriculture 5, 121-137. Ratti, C. and Crapiste, G.H.(1992), A generalised drying curve for shrinking food material', in Drying'92, Part A, A. S. Mujumdar ~d.) Elsevier, 864-873. Sodha, M.S., Dang, Aman, Bansal, P.IC and Sharama, S.B.(1985), An analytical and experimental study of open sun drying and a cabinet type chic'r, Energy Convers. Mgmt. 25(3), 263-271.
ISES Solar World Congress 1999, Volume III
175
MEDIUM SCALE SOLAR CROP DRYERS FOR AGRICULTURAL PRODUCTS Oliver St C Headley and William Hinds Centre for Resource Management and Environmental Studies, University of the West Indies, Cave Hill Campus, P.O. Box 64, Bridgetown, Barbados. Telephone 246 417 4316, Fax 246 424 4204, E-mail oheadley~hotmail.com
m
the heated air is then passed through the crop. The smaller solar dryers use natural convection or chimneys for air circulation, but for solar collectors of more than 10 m 2, forced convection is usually necessary. Dryers which use an existing building are economically very attractive and Hollick (1998) has described a dryer in India for drying 30 tonnes per day of sesame seeds, chilly and coriander using a collector of 1120 m 2 attached to an existing fossil fueled dryer to replace about 40% of the conventional fuel with solar energy. He calculates a payback time of 2.1 years. In this paper, we report on two medium scale dryers which were developed for drying onions (Allium eepa), hay and similar crops which require a relatively low drying t e x t u r e - less than 45~ We also give a comparison with solar dryers of similar size which are used for other products such as fruit and timber. Dryers tend to be product and site specific and the design is therefore very dependent on what the client wants to dry. In the section below, we give an example of the basic procedure which one follows in the design of a medium scale solar crop dryer.
m
2. D R Y E R D E S I G N
1. INTRODUCTION Solar energy is abundant in the Caribbean islands and Table 1 shows its variation over the year for selected Caribbean territories from Barbados northward. During a typical dry season day, even a small island receives terawatt hours of energy. Table 1: Solar Radiation is k W h / m 2 at four Caribbean Sites
Adams Airport, B'dos
VC Bird Airport, Antigua
9
9
13.0~ 59.5~ Jan
m
9
5.6 m
Mar n
Apr m
m
4.93
5.56
5.46
5.86 9
5.67 9
5.01 m
m
5.89 9
4.94
m
May
6.1
5.85
Jun
5.9
5.75
Jul
6.0
5.90
Aug
6.1
5.87
5.07
6.07
Sep
5.7
5.27
4.79
5.61
9
m
9
9
Oct
5.3 9
9
Nov 9
4.95
5.1 |
Dec
|
|
5.05 9
4.48 9
4.22 W
9
4.50 9
3.92 9
m
5.99
4.56
4.49
4.8 9
9
|
~m
|
5.88
4.56 |
9
5.82
4.83 9
9
5.06 9
9
6.2
4.32 9
m
m
9
18.3~ 66.0~
3.90
4.94
6.0 n
9
9
m
San Juan, Puerto Rico
17.5~ 88.2~
4.55
9
!Feb
9
17.1~ 61.8~
5.1
i
i
Belize City, Belize
9
4.13 9
9
Solar drying is one of the oldest techniques employed in agriculture and solar crop dryers have been built by the Solar Energy Project at the University of the West Indies since 1973, [Headley and Springer (1973), Headley et al (1986), Tang Kai et al (1993), Balladin et al (1997), Balladin and Headley (1999)], and during that time we have dried a wide variety of agricultural crops and timber species such as sorrel (Hibiscus sabdariffa), bananas (Musa sapientum), yams (Dioscorea alata), red cedar (Cedrela odorata) and mahogany (Swietenia maerophylla). Dryers have varied in size from the 2.2 m 2 wire basket dryer to the 149 m 2 roof collector. Most solar dryers use solar air heaters and
For a drying temperature of 40 - 45~ an unglazed collector may be employed since the radiative heat losses to the sky are relatively small at these temperatures. The roof of an existing barn is converted into a solar collector after it is painted flat black and fitted with a ceiling which becomes the duct through which air, the heat transfer fluid, is pumped. The advantage of using the roof of an existing farm building as a solar collector is that the dryer costs less than if it employs a dedicated collector. The first dryer which we describe is of this type and uses a corrugated galvanized steel roof of 149 m 2 and a fan which supplies air at 5.6m3/s with a pressure drop of 747 Pa. Air is sucked into the ceiling space from one end of the roof and is ducted to the fan which then blows it into the plenum chamber at floor level. This dryer is owned and operated by Mr Patrick Bethell of Friendship Plantation in Barbados. Bags of onions or bales of hay are placed on the top of the plenum (which is built on the floor of the barn) and whose top is perforated to allow the air to flow through it and then through the crop. The crop is stacked so that heated air does not bypass it and escape to the ambient without removing moisture. This dryer does not employ an auxiliary heating system since Bethell did not think it was necessary. Given the airflow rate, psychrometric calculations using the procedure described by Exell (1980) show that drying will take place even when the temperature rise is low. Preliminary calculations assumed that hay was to be dried at 800 bales per week. Each bale has a mass of 18.15 kg (40 lbs) After field drying the moisture content of the hay is 40~ (wet basis) and this is to be reduced to 15% during drying. The relevant equation to determine the mass Mw of water lost is
ISES Solar World Congress 1999, Volume Ill
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Mw Where and
= Mc [(WI
(1)
WF)/(IO0- Wr)]
-
by the equation P =
WI is the initial moisture content WF is the final moisture content Mc is the initial mass of the crop.
(3)
Where VA is the airflow rate in m3/s, 1~ is the pressure drop in Pa and v is the fan efficiency, taken as 0.7 for this calculation. The fan power requirement is therefore 3.1 x 747/0.7 = 3300W or 4.4 hp. For extra capacity, we oversized the fan and the one fitted delivers 5.66 m3/s at a pressure drop of 747 Pa. This dryer's maximum capacity is ten tormes of hay. The aperture area of the intake to the roof space which is the entry point for air into the solar collector is 12.19 m x 0.102 m = 1.24 m 2. The linear air flow rate is therefore 5.66/1.23 = 4.57 m/s. Determination of the Reynolds number allows us to decide if the flow is turbulent or laminar. The Reynolds number is given by the expression
Both W1 and WF are taken on a wet basis. Mw is therefore 800 x 18.15 [(40 - 15)/(100 - 15)] = 4270 kg. If one assumes that each kg of water requires 2.5 MJ for its removal, then the total heat needed is 4270 x 2.5 = 10,676 MJ. Exell's procedure for calculating the amount of water which can be removed by the airstream is then employed using a psychrometric chart. Assuming an input air temperature of 25~ (dry bulb) and a relative humidity of 70%, the psyehrometric chart shows that its humidity ratio is 0.0141 kg water/kg dry air. When the solar collector heats it to, say, 40~ (dry bulb), the humidity ratio remains constant. If on passing through the crop, the air absorbs moisture until its relative humidity is 90%, the psychrometric chart shows the humidity ratio to be 0.020 kg water/kg dry air. The change in humidity ratio is therefore 0.020 - 0.0141 = 0.0059 and the corresponding dry bulb temperature is 28.2~ From the gas laws PV = MART
V#OP/*
Re = PVLDH/~t (4) Dn is the hydraulic diameter of the duct, = 0.2 m ~!is the density of the air = 1. lkg/m 3 VL is the linear velocity of the air stream = 4.57 m/s ~t is the air's dynamic viscosity = 1.79 x 10"Skg/m.s
Where
(2)
Where
P is the atmospheric pressure = 101.3 kPa V is the volume of air in m 3 M^ is the mass of the air in kg T is the absolute temperature in kelvin, and R is the gas constant = 0.291 kPa m3/kg K For a humidity ratio increase of 0.0059 kg water/kg dry air, each kg of water will require 1/0.0059 = 169.5 kg dry air. For this calculation, the absolute temperature is 28.2 + 273 = 301.2 K and the volume of air needed to remove 1 kg of water is 169.5 x 0.291 x 301.2/101.3 = 146.6 m3; hence 4270 kg will require 4270 x 146.6 = 626,047 m 3. For a drying time of seven days and 8 hours operating time per day, one has 8 x 7 x 3600 = 201,600 seconds. The air flow rate is therefore 626,047/201,600 = 3.1 m3/s. From previous experience with conventional dryers, it is known that the hay bales in the stack produce a pressure drop of about 3" water gauge = 76.2 mm water gauge = 747 Pa. The fan power requirement P is given
i r duc ts dor collector
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r-
-
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, I
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,
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"-
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.
.
.
.
.
.
.
.
-
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9
ISES Solar World Congress 1999, Volume III
Re = 1.1 x 0.2 x 4.57/1.79 x 10-5 = 56,063 which means the flow is turbulent which will give good heat transfer. The Nusselt number may be calculated from the Reynolds number by using the equation Nu = 0.0158 Re ~ = 0.0158 x 56,0630.8
(5) = 99.45
The heat transfer coefficient hc then be found using the equation hc = Nu k/DH (6) Where
I)n is the hydraulic diameter, and k is the thermal conductivity of air = 0.029 W/mK Here DH is twice the plate spacing = 2 x 0. lm = 0.2 m. hc = 99.45 x 0.029/0.2 = 14.4 W/m2K. The procedure for finding the top loss coefficient and the efficiency factor for the collector follows Duffle and Beckman's method. The radiative heat transfer coefficient from the plate to the sky was calculated to be 6.39 W/m2K given a plate ten~eramre of 45~ a plate emissivity ofO.95 and an ambient temperature of 28~ assuming that the sky temperature is equal to the ambient temperature. The convective heat transfer coefficient was calculated to be 24.7 W/m2K, using McAdams equation for a wind velocity of 5 m/s. These combined to give a top loss coefficient Ux of 5.07 W/m2K. The bottom loss coefficient UB for the collector was found to be 2.15 W/m2K given that the duct base was 19mm thick and that its thermal conductivity was 0.041 W/mK. The overall loss coefficient UL = lax + UB = 7.23 W/m2K. The collector efficiency factor F' was calculated to be 0.72, the dimensionless capacitance rate was found to be 9.28, the collector flow factor F" was 0.95 and the collector heat removal factor F R was 0.68. The second dryer uses a glazed solar collector array of-- 40m2 (20 panels of 1.95 m 2 each) which is connected to a fan by horizontal and vertical ducts. Two horizontal ducts then direct the flow into the drying chambers. By means of the ducting system, the fan sucks air through the solar collectors and forces it into the containers where the onions are placed on racks. One of the containers is insulated and lined with concrete, while the other is simply insulated. The cross section of the air heating space in each solar collector measures 25.4 mm by 914 mm which gives an equivalent diameter of 126 mm for a circular duct. This was the first step in the calculation for determining the pressure drop in the system and the fan power requirement. Standard procedures such as that described by Malik and Buelow (1976) were used for this determination. The fan is 2.2 kW (3 hp) and is rated to deliver 1.16 m3/s at a pressure drop of 747 Pa (3" water gauge). A linear air flow rate of lrn/s in each collector gives a throughput of 0.92 m3/s in the complete array of 40 collectors. The procedures described by Duffle and Beckman (1991) were used to determine the collector efficiency factor. The collector was glazed since we needed to raise its efficiency by reducing radiation and convection losses. The two standard 6 metre (20 It) shipping containers, one of which was modified to enhance heat storage capacity with a lining of concrete blocks for increased thermal mass formed the drying tunnel. One of the advantages of a drying tunnel is that the product to be dried can be easily loaded unto trolleys and loaded into the dryer. This greatly simplifies materials handling.
177
This dryer is fitted with electrical resistance heaters to supply heat during rainy or cloudy periods and to allow the dryer to be used at night. Even though electricity is 26r in Antigua - which is the second highest electricity tariff in the anglophone Caribbean - the Central Marketing Corporation (CMC) of Antigua preferred to use this heat source since they consider it to be more reliable and less dangerous than liquefied petroleum gas even though the latter is cheaper. The maximum capacity of this dryer is four tonnes of onions and a diagram is shown in the Figure. 3. PERFORMANCE As can be seen from Table 1, the monthly available solar energy at the V. C. Bird International Airport varies from a high of 5.87 kWh/d in July to a low of 4.22 kWh/d in December. At an efficiency of 40%, the heat output from the collector array is 5.87 x 3.6 x 1.95 x 40 x 0.4 = 659 MJ/d in July. If 2.5 MJ are required to evaporate a kilogram of water, then the solar heat input has a drying capacity of 279 kg of water per day. However the fan delivers air 1.16 m3/s which amounts to 33,400 m 3 for an eight hour day. The inherent drying capacity of the input air is determined from its temperature and relative humidity, this is determined using the same procedures that described above for the 149 m 2 dryer.
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Table 2: Comparison of four medium scale solar dryers
Territory
Antigua
Barbados
Trinidad
Belize
Operator
CMC
Patrick Bethell
David Richardson
Soren Sorensen
Solar collector type
Singly glazed
Bare plate, no glazing
Singly glazed (corrugated plastic)
Singly glazed
Collector area
40m 2
149 m E (1400 sq. R)
30m 2
100 m 2
Air circulation system
1 fan, 3 hp (2238 Watts), 1.16 ma/s
1 fan, 5.66 m3/s (12,000 cfm)
4 fans, 350W each
1 fan, 4.7 m3/s (~ 10,000 cfm)
Dryer configuration
Tunnel, two shipping containers
Perforated floor, upward air flow
Two stack, crossflow
Drying tunnel with 12 trolleys
Main products dried
Onions, herbs on occasion
Hay, onions on occasion
Timber
Mango, papaya, banana, pineapple
Load capacity
7 tonnes on trays
350 bales fresh hay (10,000 kg, ~22,0001b)
18 m 3 in two stacks with stickers
4500 kg (10,000 Ib) fresh fruit
Drying time
~ 1 week
~ 21 days (3 days at 7 h/d)
3 weeks, two weeks with continuous backup
24 h (mango), 30 h (papaya), 72 h (banana)
Backup heater
Electric resistance
None
Waste wood burner
Waste wood burner with flue gas heat exchanger
Cost, US $
40,000 (1997)
4,200 (1985)
~20,000 (1992)
15,000 (1986)
In one run, a tonne of herbs was dried from an initial moisture content of 85% (wet basis) to a final moisture content of 10% (wet basis) in one day. More detailed testing was unfortunately interrupted by the hurricane season and the dryer lost five of its collector panels when Hurricane Georges struck Antigua in September 1998. We have been negotiating with the CMC to determine who will pay for the hurricane damage, this has yet to be resolved. Table 2 compares the Antigua dryer with three other medimn scale dryers in the Caribbean. These dryers are all designed to dry loads of several tonnes of agricultural products. Many agricultural products are currently dried using fossil fuels. In Trinidad, a large copra dryer on a coconut estate dries 9100 kg of copra in 34 hours using 660 litres of diesel fuel for heating and a l0 hp fan for air circulation. During this time, it consumes 660 litres ofdiesel fuel at a cost of $145 US and $9.30 US of electricity (254 kWh at 3.67r US/kWh). Since the fuel cost is much greater than the electricity cost, it clearly pays to reduce the fuel cost by having a solar collector mounted on the roof of the drying barn and using it to supply hot air to the dryer. The payback period for this modification will be examined under the section on economic considerations.
4. ECONOMIC CONSIDERATIONS
In Barbados, the cheapest of the fossil fuels is natural gas. For a small consumer, the price varies with the amount consumed according to the figures given in Table 3. Table 3: Natural Gas Price for a Small Consumer
Natural Gas Consumed (m3)
Price (US$)
The first 150
0.65 i
The next 2250
o.60
The next 5300
0.58
The next 8700
0.57
ISES Solar World Congress 1999, Volume Ill
Assuming a heating value of40MJ/m 3 and a burner efficiency of 75%, one can calculate the annual cost of the solar heat which is produced by the solar collector. If the collector is assumed to have an average efficiency of 40% over the year, noting from Table 1 that the average annual insolation in Barbados is 5.33 kWh/m2d; assuming 300 days of operation per year, then from a surface area of 149 m 2 it should deliver 5.33 x 0.4 x 149 x 300 x 3.6 MJ - 343,081 MJ. In terms of natural gas heating delivered from the burner, this would be 343,081/(40 x 0.75) = l l,400m 3. From Table 3, this is somewhat more than the sum of the first three rows, which comes to $1.83 US, but less than the sum of the four rows which is still only $2.40 US. With a dryer cost of $15,000 US, the repayment period based on replacing fuel with the solar energy collected is15,000/2.4 = 6250 years. Clearly solar energy for this dryer cannot compete with natural gas! If we consider using solar heat to replace some of the diesel fuel used by the copra dryer in Trinidad, a similar sized solar collector giving a similar anual output of heat the (the solar regime in Trinidadis not much different to that of Barbados), then with a heating value of 40MJ/kg and a burner efficiency of 0.75, one has 343,000/(40 x 0.75) = 11,433 kg or about 12,700 litres. This is worth $2790 US, hence the payback period for a $15,000 investment is 5.4 years, which is much more resonable. Dr Irnran McDoom and his graduate student at the St Augustine campus of the University of the West Indies are now looking a producing a solar dryer for a ccr~mut estate.
5. CONCLUSIONS Solar crop dryers are a cost effective solution to some of the problems of food preservation in sunny climates. The solar dryer with the least initial capital cost is one which uses an existing farm building or adds a solar air heater to an existing conventional crop dryer. While multi-crop dryers may seem to be an ideal solution, the fact is that most operators prefer to have a dryer which is dedicated to one or two crops or to a specific kind of crop, fruit for exan'q31e, since the compromises inherent in a multi-purpose dryer often result in reduced efficiency for its primary product. ACKNOWLEDGMENTS The research, development and outreach work reported on in this paper was supported partly with funding from the Inter American Development Bank under their financial assistance programme for the development of Science and Technology at the University of the West Indies and partly by the European Development Fund under Project Number 5100.51.94.176. REFERENCES
Balladin D. A., Headley O., Chang-Yen I. and McGaw D. (1997). Extraction and evaluation of the main pungent principles of solar dried West Indian ginger (Zingeber officinale Roscoe) rhizome. Renewable Energy, 12, 125 - 130. Balladin D. A and Headley O. (1999). Evaluation of solar dried thyme (Thymus vulgaris Linnt) herbs. Renewable Energy, 17, 523 - 531.
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Duffle J. A. and Beckman W. A. (1991). Solar Engineering of Thermal Processes, Second Edition, pp 296- 301. WileyInterscience, New York. Exell R. H. B. (1980). A simple solar rice dryer, basic design theory. Sunworld, 4, 186 - 191. Headley O., Harvey W. O'N. and Osuji P. O. (1986). Simple solar crop dryers for rural areas, In Proceedings oflSES Solar World Congress, 22-29 June, Montreal, Canada, Bilgen E. and Hollands, K. G. T.(Eds) 2, 1082 - 1086, Pergamon Press, New York. Headley O. and Springer B. G. F. (1973). A natural convection solar crop dryer, In Proceedings of the ISES/UNESCO Solar
Energy Conference, "The Sun in the Service of Mankind," Paris, Paper No. V 26. Hollick
J. C. (1998). Commercial scale solar drying.
Renewable Energy, Vol. 16, pp 714 - 719. Malick M. A. S. and Buelow F. H. (1976). Hydrodynamic and heat transfer characteristics of a heated air duct. In Heliotechnique and Development, Kettani M. A. and Soussou J.(eds) Vol 2, pp 3 - 30. Development Analysis Associates, Cambridge, Massachusetts. A. Tang Kai, I. A. McDoom and O. Headley (1993). A solar timber kiln for artisans, In Proceedings, ISES Solar Worm Congress, 23 -27 August, Budapest, Hungary, Vol. 8, pp 43 47, Farkas I. (Ed.) Hungarian Solar Energy Society, Budapest. Keywords: Crop drying, solar air heaters, unglazed collector, thermal mass, forced convection.
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THE MARSTAL CENTRAL SOLAR HEATING PLANT: DESIGN AND EVALUATION
Alfred Heller Dept. of Buildings and Energy, Technical University of Denmark, Building 118, DK-2800 Lyngby, Denmark, Phone: +45 45 25 18 91, Fax: +45 45 93 17 55, e-mail: [email protected] Jochen Dahm Dept. of Building Service Engineering, Chalmers University of Technology, S-41296 Gothenburg, Sweden, Phone: +46 31 772 1150, Fax: +46 31 772 1152, e-mail: [email protected] Abstract H A model for simulation of Central Solar Heating Plants (CSHP), including different control strategies is built in TRNSYS. The model is partly calibrated with data from the solar heating plant in Marstal, Denmark. Good agreement between measurement and calculation is obtained. Furthermore, a model applying TRY data for Copenhagen is prepared. Using this model three control strategies to operate the collector circuit are evaluated. The first two strategies are applying variable flow in the collector circuit to obtain a constant temperature from the collector field. In Marstal this method was chosen to obtain a high solar coverage during the summer season and to save electrical energy for pumping. The third is a more conventional control strategy operating the collector pumps at a constant flow rate. The results from model calibration and simulations are presented and discussed.
1
INTRODUCTION
2
THE MARSTAL PLANT
Background and motivation
Description
The central solar heating plant (CSI-IP) in Marstal was put into operation in the autumn of 1996. The first experiences were presented at the last ISES World Conference in Korea by Heller and Furbo (1997). Specific for the Marstal solar heating plant is the control strategy to operate the collector circuit. The motivation for the applied control strategy is the ambition to supply 100% of the heat demand during the summer season and to minimise the consumption of electrical power for pumping. These two goals led to a rather complex control strategy and variable flow in the collector circuit. To cover the summer heat demand by solar only, the solar heat gain must at least equal the demand and at a minimum temperature, the district heating forward temperature. Therefore, the solar circuit volume flow is controlled to achieve this forward temperature. Above the temperature criteria, the control strategy has to compensate for the cycle time of the solar circuit fluid, which can be up to one hour and hereby to ensure smooth flow control. Applying an efficiency expression generated by a solar collector test, the collector return temperature can be controlled sufficiently. In this paper the development of a simulation model for the Marstal plant and the calibration using measurements is described. Applying more general TRY weather data for Copenhagen the control of the collector circuit is evaluated with regard to solar coverage during summertime and electricity consumption for pumping.
Marstal, a little village on the Island of Aeroe in Denmark, comprises approximately 1250 households getting twice as much in summer time due to tourism. The local utility Marstal District Heating A.m.b.A. decided in 1991, encouraged by national laws demanding the reduction of CO2-emissions, to utilise solar heating. The district heating (DH) system, mainly servicing single-family and row houses with yearly 26 GWh (94 TJ) heat, was prepared in the late 70's for low-temperature operation. The heat supply temperature was set to 72~ The return temperature is about 44~ in summer time and about 35~ in winter- time. The central solar heating plant consists of 8064 m2 solar collectors an~anged in two blocks of 32 rows each. Each row comprises 10 Ar-Con HT collector modules tilted 40 degrees facing south. All connection pipes, except the connection between the collectors, are buffed underground applying standard district heating enterprising. Three Gnmdfos NF 100= 65-200/191 pumps run up to two at a time with a varying mass flow rate from 20 up to 160 m3/h. An insulated steel tank with 2,100 m 3 water satisfies the heat demand of three to five summer days. The tank is connected to the plant by two sets of inlet- and outlet diffuser anm~gements. Most equipment is doubled to ensure reliability. A diesel driven cogeneration plant (150 kWel) is installed for backup. Figure 1 shows the principle of the central solar heating plant.
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Figure 1. Principle drawing of the Marstal central solar heating plant. The control and monitoring system is based on PLC's connected to a PC network. The PC's are equipped with InTouch terminal software. 240 parameters have been monitored since 1997, captured in daily reports on which the current analysis is based. The control of the collector circuit in more detail is as follows: Summer operation: The solar collector efficiency is calculated online using tested collector parameters and measured solar irradiation on the collectors. As soon as the result of the calculation exceeds zero, the primary solar collector circuit is started at a minimum flow rate of 20 m3/h. The secondary collector circuit starts operation, as soon as the primary collector return temperature exceeds the tank bottom temperature. After starting the secondary collector circuit the flow is controlled in order to obtain the minimum required district heating supply temperature plus a margin leading to a set temperature of 80~ In case the maximum flow of 160 m3/h in the collector circuit is attained, the temperature in the network increases. Winter operation: Operation start of the primary and secondary collector circuit as for summer operation. As opposed to summer operation the flow is controlled in order to obtain a minimum solar collector return temperature of 50~ Further, a gravel pit storage was built in 1998 as a research and demonstration project. In 1999 the collector field was enlarged with 1000 m 2 of solar collectors. Due to control problems the three pumps on the primary side are replaced by two pumps running one at the time. Due to the fact that this paper is based on data from 1997, the latest changes are not taken into consideration.
Measurement of solar collectorfield performance The measured performance is shown in for a sunny summer's day.
Figure 2. Measured data for the Marstal central solar heating plant, 23-07-1997. Top: Ambient temperature in degrees Celsius, and total solar irradiation on the 40 degrees tilted collector surface in W/m 2. Middle: Volume flow rate and solar circuit power, measured on the secondary side of the heat exchanger. Bottom: Curve (1) shows the temperature to the collector circuit, (2) the temperature back from the collector circuit at the heat exchanger and (3) the average outlet temperature at the collector rows measured right after leaving the last collector module. The operation modes for the primary collector circuit pump and the secondary pump transporting the heat to the district heating net or the storage tank are shown as well.
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The most significant findings from are: [3
[3
[3 [3
The pump in the primary circuit is started approximately 1 89 hours before the secondary pump, preheating the collector. A control based on solar irradiation (start at 100 W/m2) would start preheating approximately half an hour later. During the morning and evening hours the collector volume flow is controlled to obtain the desired collector return temperature of 80~ In the afternoon, the flow rate reaches maximum and the temperature back from the collector field exceeds this temperature. Simultaneously the collector forward temperature rises. For a nice summer day as shown in the maximum solar collector power is 4.5 MW. For a clear day the Marstal summer control strategy operates well with a stable start-up in the morning and stable shotdown in the evening.
In similar plots for days with fluctuating solar irradiation (not shown) many start-stop situations on the primary and the secondary side of the solar circuit earl be observed. Anyhow, it can be concluded that the Marstal control strategy for summer and winter operation works well.
Measured distn'ct heating load The measured forward and return temperatures of the district heating network are shown in Figure 3 dependent on ambient temperature.
temperature is higher and about the same as usually occurring in DH networks. The total measured heat load of the Marstal district heating network is 26 GWh in 1997. During the summer season the DH load amounts to 2 GWh. In Figure 4 the measured district heating power for 1997 is plotted versus the ambient temperature.
Figure 4: Measured district heating power versus ambient temperature, 1997. Figure 4 shows a clear coherence between ambient temperature and heat demand of the district heating network. The heat demand decreases linearly from a maximum of 7 MW at 0~ ambient temperature to about 1 MW at 17~ ambient temperature. Above 17~ ambient temperature only network heat losses and domestic hot water answer for the network heat demand. From this it can be concluded that the space heating demand is dominant during the heating season, while the domestic hot water demand is dominant during the summer period. As well, it can be seen that the demand during summer time is relatively constant due to the size of the network. During 1997 no values were recorded at design ambient temperature (-12~ which indicates that 1997 was warmer than expected by design rules. Extrapolating the total heat demand of the network results in a design heat demand of about 10 MW or 8 kW/residential unit.
Monthly measuredperformance Figure 3. Measured forward and return district heating network temperatures versus ambient temperature, 1997. From Figure 3 the main findings are that the DH forward temperature is constantly about 72~ In summer, at times with a large solar heat gain, the network forward temperature varies due to the control of the solar circuit. The return temperature varies dependent on the ambient temperature. For ambient temperatures below 15~ the space heating demand increases and the return temperature comes down to 32~ At higher ambient temperatures the return temperature increases to a maximum of 46~ From this it can be concluded, that the Marstal district heating network operates at low temperatures, advantageous for the operation of a solar plant. Especially in winter and at times with space heating demand the network return temperature is low compared to other similar DH networks. Anyhow, in summertime and at higher ambient temperatures the return
In Figure 5 the measured performance for the Marstal central solar heating plant and district heating network is shown.
Figure 5: Measured heat balance for the Marstal central solar heating plant and district heating network for 1997.
ISES Solar World Congress 1999, Volume III
According to Figure 5, the yearly solar fraction for 1997 is 15.4%, while the highest monthly solar fraction is 95% (August). Due to the rather small storage tank with a heat capacity of 3-5 summer days it is not possible to cover 100% and auxiliary backup is required. For this case a small auxiliary oil boiler is used. The yearly solar collector gain is 450 kWh/m2,a. 3
183
Using a dynamic fitting software (W. Spirkl) the average deviation between measured and calculated power is 55 kW for the heat exchanger and 345 kW for the collector field regarding the four chosen days. For these days the collector operates at an average power of about 3000 kW. Figure 6. shows the measured and calculated collector field power as well as the difference for June 23, 1997 (one of the four days used for calibration).
SIMULATION MODEL
To evaluate the Marstal solar heating plant a simulation model was built in TRNSYS. The simulation model includes a specific load model for the plant, a detailed model of the collector circuit to evaluate the solar collector control and a model of the auxiliary backup of the plant. In this chapter the simulation model is described and calibrated using the measurements of the district heating load and the central solar heating plant.
10o
I ~P-m~md
I
~ , 3000
~
L..
i=
i
Not used for parameter identificationl
Weather data Solar radiation is measured as global irradiation in the plane of the collector. Most (solar collector) calculation models require knowledge of the direct and diffuse part of the solar irradiation. To be able to use the measured global irradiation for simulation purposes it is therefore required to calculate the direct and diffuse part on the horizontal. Here, the isentropic sky model by Liu and Jordan (1960) was used to divide the global radiation into its direct and diffuse parts. Calculating the diffuse and direct part of the global radiation introduces uncertainty. Especially the location of the plant (maritime climate) might cause different conditions as valid for the used Liu and Jordan isentropic sky model. However, for simulation and evaluation purposes this calculation model is considered to be sufficient. Solar Collector Circuit As one of the major parts regarding thermal performance of the plant, the solar collector circuit is investigated and a solar circuit model is developed. Major model components are the MFC-collector model by Isakson (1993), the Multyport Storage by Driick and Pausehinger (1995), TRNSYS TYPE 140 and the standard TRNSYS TYPE 5 heat exchanger (TRNSYS, 1996). To account for heat losses of the collector pipelines (480 m forward and 250 m return) and for the time delay in the primary collector circuit a ground pipe model (ALTENER, 1997) is introduced. The circulation pump model is the standard TRNSYS TYPE 3 model (TRNSYS, 1996). The heat loss parameter values of the pipe component are taken from the manufactm'er's data sheet. The average heat loss for typical operation conditions is about 30 W/m (forward and return pipe together) for a temperature difference of 60~ between pipelines and surrounding soil. To calibrate the solar circuit simulation model, four days, i.e. two days representing winter and two days representing summer operation mode, were picked out from the measurements. The solar circuit heat exchanger was calibrated for a constant UAvalue. Identified parameter values of the solar collector model are the zero loss efficiency, the first order loss coefficient and the effective heat capacity. All identified collector parameter values represent the performance of the primary collector circuit including all components as described above.
0 +"-"""~ 0 2
-40
.OO
-80
:
4
6
8
10
2 4 -100
12
Figure 6. Measured vs. calculated power and difference, 23. June 1997 (one of the four days used for calibration). From Figure 6, it can be observed that the largest difference between measured and calculated power occurs at start-up and shut-down of the collector field. One reason is among others the uncertainty introduced by dividing global irradiation into its direct and diffuse parts. Note: The start-up phase in the morning was skipped for parameter identification, Due to the control strategy for summer operation the collector operates during most of the time at the same temperature (~60~ collector average temperature). Therefore, the validity of the identified model parameters is limited to this operation range. It would have been advantageous to identify the model parameters using a wider range of operation conditions. To calculate the collector field pump electricity consumption a typical pump characteristic was applied (based on product data from GRUNDFOS) for each of the two pumps in both primary and secondary collector circuits. Figure 7 shows the applied pump characteristic. 1.0
P~ / Poem=,= 1.3"V / V2n~x- O.4*V/ Vrr=x + 0 . ~ / ~ 0.8
throttle
variable RPM
41". . . . . . . . . . . . . . . . . . . . X
E 0.6 n
-~ 0.4 13.
0.2
J
lowest possible RPM
0.0 0.0
0.2
0.4
I
i
0.6
0.8
1.0
V/ Vm~
Figure 7: Typical pump characteristic (GRUNDFOS data).
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The maximum power, Pma~, for one primary solar circuit pump is 15 kW and 2.2 kW for one secondary pump. It is assumed that the second pump starts when the first pump approaches maximum flow.
Load model The district heating load curve has a large influence on the overall performance of the solar heating plant. Especially for solar plants with diurnal heat storage the return temperature and the domestic hot water (DHW) load during summer time are of importance. To calibrate the simulation model of the solar heating plant the applied load model has to be calibrated. As described, the DH load in Marstal mainly consists of residential buildings. Most of the buildings are single-family houses. The applied reference Space Heating (SH) load was calculated for one typical residential building using TRNSYS TYPE 56 and multiplied by the total number of 1250 connected buildings. The reference building is standard-insulated and built as is typical for this region. The SH load for one reference building is 15.3 MWh/a using the measured weather data for Marstal in 1997. The DHW load for one house (two residential units) is about 4.5 MWh/a. To account for visitors on the island during summer the DHW demand increases from Maj to July by 70% and decreases until the end of August. To account for distribution heat losses two ground-buried pipe models (4000 m) are applied using the manufacturer's heat loss data for an average pipe diameter of DN 100. Figure 8 shows that the calculated return temperature is about 3~ lower than measured during summer time (hours 0 to 3000). One major reason for this is a better performance of the modeled DHW substations and the assun~tion of a single reference building type representing the whole variety of buildings in the village, defining the load model.
remembered that only one reference building type is applied as load model.
Figure 9: Measured and calculated duration curves for the DH load. From Figure 9 shows that the calculated DH load curve is in accordance with measured DH load curve. In total (including DHW, SH and pipe heat losses), the measured DH load amounts to 26.1 GWh compared to 26.2 GWh. The difference is less than 1%.
Solar heating plant model Combining the DH load model with the solar circuit simulation model, adding a heat store, a boiler and a co-generation plant model together in one simulation model the yearly overall performance of the solar district heating plant can be predicted. Applying the measured weather data and the DH load model as described, the measured and calculated solar gain and districtheating load can be compared for the period of one year.
Figure 10: Measured and calculated DH load and solar gain on a monthly basis using measured weather data of Marstal 1997.
Figure 8 : Measured and calculated duration curves for network temperatures. Figure 8 shows it can be seen, that the calculated return temperature is about 3~ lower than measured during summer- time (hours 0 to 3000). One major reason for this is a modeled better performance of the modeled DHW substations as measured. As well it should be
The maximum relative difference between measured and calculated solar energy on a monthly basis is 60% and o~urs in December. The maximum relative difference for the load is about 20% and occurs during the summer at times of low demand. Anyhow, the absolute difference is little on these occasions. Thus, the relative difference on a yearly basis is small (0.2% load, 1.3% solar). The specific solar gain is calculated to be 452 W/m2,a which differs about 1% from the measured value of 446 W/m2a.
ISES Solar World Congress 1999, Volume III
Regarding the investigations as described, the simulation model of the solar heating plant is taken as being calibrated.
Reference model
185
Figure 13 shows that during wintertime the electrical energy is relatively large in relation to gained solar heat. In this period the collector circuit of the Marstal plant operates at constant flow and no savings of electricity for pumping can be noted.
To investigate the performance of the solar heating plant in a more general way a reference model was created. Since the spring and summer of 1997 were warmer than an average year for this region and to make the study comparable to other investigations, the weather data was exchanged against the test reference year for Copenhagen.
Figure 13 : Solar thermal energy (left, thick bars) and pump electrical energy (right, thin bars)
Figure 11 Calculated energy balance for the Marstal solar heating plant using weather data for Marstal 1997 and Copenhagen TRY (reference model). In Figure 11 shows the calculated energy balance for the solar heating plant using weather data for Marstal 1997 and Copenhagen TRY. It can be observed that the district heating load, using TRY weather data, during spring is higher than the district heating load for Marstal 1997. This is due to a higher average ambient temperature in the spring of 1997 compared to Copenhagen TRY. The solar heat gain during the winter period is small compared to the heat gain during the summer months. Further, using TRY data additional auxiliary heat during the summer months is required. Using TRY weather data results in 9.5% (50 kWNm2/a) lower solar heat gain and 1% lower solar fraction compared to Marstal 1997. The calculated solar circuit pump electricity for TRY-conditions is shown in Figure 12 and Figure 13.
Figure 12 Electricity for the solar circuit pumps: (left, bars). Fraction of electrical energy on thermal solar gain and (right, line) operation mode, 1= variable flow, summer operation.
Figure 13 shows that the monthly electricity usage for pumping which is about constant for all months. Based on this observation it should be considered, taking the low solar heat gain during winter time into account, to apply the summer operation mode all the year round or even switching off the plant from November until the end of January. Switching off the solar circuit during this time causes no major decrease of solar heat gained on a yearly basis (about 3% less or 100 MWh) and saves about 16 MWhel in electrical energy. Note: The calculations could not be compared with measurements due to lack of confidence in the measured result. 4
OPERATION MODE EVALUATION
This chapter focuses on an initial evaluation of the solar circuit operation. As described, in Marstal two specific operation modes are used, one for winter and one for summer operation. Here, both modes are studied regarding solar fraction during the summer period and electrical energy used for solar collector circulation pumps. Further, a conventional solar collector control is investigated for comparison. For this purpose additional simulations using the simulation model and TRY weather data for Copenhagen were carried out: For the first simulation the summer operation mode as used in Marstal was applied all the year round. By using the summer mode, which means running the collector at a high temperature even during the winter period (80~ collector return), a lower solar gain is expected during this time. For the second simulation the winter operation mode as used in Marstal was applied all the year round. Using this mode a higher solar gain is expected, since in this mode the collector always operates at the lower winter set operation temperature (50~ collector return) even in summer time. For the third simulation a conventional operation using constant flow and a simple difference temperature thermostat control is applied. Here, the primary collector circuit is started at a constant, maximum flow as soon as the collector temperature exceeds the DH return temperature. The secondary pump starts at a constant, maximum flow as soon as the collector return
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temperature at the heat exchanger exceeds the DH return temperature. In this operation mode the collector always operates at the lowest temperature possible. Due to constant flow in the collector circuit this operation is expected to require more pump electricity than the Marstal operation modes. Figure 14 shows the solar fraction, defined as the ratio between solar gain and total district heating load, for the originally applied operation mode, for summer and winter operation mode all the year round, as well as for the conventional controller operation.
600
s./Isx
500 .~
400
'-
300
i... r -6 r
200
//
i1'.
/-
100
-Ul ~
5o =
oTl---
original
,
9 x\
o
1,'
_
"'.~.
-~
I
I
,
,
,
,
,
|
,
,
,
Figure 15 9 Solar gain for the different control strategies on a monthly basis.
....
:30 2~
Summer oPera,on
--Winter o~on
[TRY]
Winter operaUon
,
.~."" "
constant flow, simple contxol 0
100 II ~ ~ l
~ ,
!
!
I
Fm~q F i g u r e 14 9 Solar fraction for the different control strategies on a monthly basis. Applying the summer operation all the year round does not differ significantly from the original Marstal operation. Considering the winter operation applied all year an increase of solar fraction (about 5%) can be achieved during the summer period. Regarding the conventional controller mode it is interesting to note that the solar fraction is slightly higher during spring and autumn, but significantly lower during the summer period. One reason for this is that the conventional controller operates always at a constant, maximum volume flow rate of the collector field. As a result, especially during low load periods in summer, the temperature from the collector field is lower than the minimum required network forward temperature, but the volume flow is higher than demanded by the district heating network. In that case an auxiliary boiler heats up the part of the volume flow which goes directly to the network at a minimum network forward temperature. The remaining volume flow from the collector field is directed to the heat store. It should be noted that the applied flow, 160 m3/h or 0.33 litre/min.m2, i.e. maximum flow used in Marstal, is a rather high flow rate (Dahm, 1997). Figure 15 shows the solar gain for the different control strategies.
From Figure 15 we find that a modification of the Marstal strategy, running summer conditions during the whole year, leads to slightly lower solar gain in periods with poor solar irradiation. Opposite leads a winter control strategy to higher solar gain during periods with high solar irradiation. The conventional operation strategy leads to the highest production in spring but decisive lower solar gain in summer and no difference in autumn and winter. This result may be explained by the fact that high flow rate leads to a less efficient way of storing heat in the storage tank. No conclusions can be made on this observation yet.
control strategy Marstal original Summer operation all year Winter operation all year Constant flow, simple control
solar gain [kWh/mZ,a] 402 388 434 413
T a b l e I 9 Specific solar gain for the control strategies. Table 1 shows that the simple control strategy has a higher solar gain compared to the original. However, applying the winter strategy all the year round increases the yearly solar gain by about 7%. The solar circuit pump electricity for the three operation strategies is shown Figure 16. ,..., 18 p .....
/ 9
"-~ 14
/
tf""
g lO
..."J
,,._
\
\
\
\
~ 8 9
4
_
•
---uem~ ~
. . . . sumnw operam
--Winter
. . . . cemtant ~ w , simple
cCeme~
F
0
Fmq
Figure 16 9Pump electricity for the different control strategies. It can be seen that the summer mode results in about the same pump electricity as the original mode. Applying winter mode increases the pump electricity compared to the summer mode by
ISES Solar World Congress 1999, Volume III
50% during summer time. As expected the pump electricity is highest for the constant flow controller mode. Compared to all Marstal operation modes the constant flow controller mode needs up to three times more electrical energy. It can be concluded that the control strategy has a major influence on solar gain, as well as on pump electricity, and that it should be possible to find an optimum by performing a more systematic study also taking alternate system layouts into account. 5
SUMMARY
The main goal of this study is an initial evaluation of the solar collector operation strategy of the Marstal solar heating plant. For this purpose a simulation model was developed and calibrated. The focus of the model calibration is on the solar collector circuit and the district heating load. Due to the measured operation of the solar collector field, which is a constant collector return temperature, the collector field operates at most times in the same temperature range. The calibration is limited to this range. The load model was calibrated using the measured district heating network power and temperatures. The difference between measured and calculated load is less than 1% and the network temperatures are within +3~ of the measured values. For the solar plant simulation model, taking the measured weather data for Marstal in 1997, the deviation between measured and calculated energies is small and amounts to 0.2% for the load and 1.3% for solar energy on a yearly basis. The specific solar gain is calculated to be 452 W/m2a which differs about 1% from the measured value of 446 W/m2a. Running this model with TRY weather data for Copenhagen three different solar collector control strategies, specific for Marstal, were investigated and compared with a simple constant flow control. From this initial study it can be concluded that the highest solar gain can be achieved by applying a modified Marstal control strategy, winter mode. Here the collector field return temperature is lowered compared to the Marstal strategy all the year round. Using this mode the highest solar fraction is obtained during the summer season. Using a simple constant flow control the solar gain is slightly higher than by using the original control, but the pump electricity is twofold on a yearly basis. However, the goal to achieve 100% solar coverage during the summer season cannot be achieved by any of the studied control strategies, as the heat storage is too small. AKNOWLEDGEMENTS
The Danish Energy Agency supported this work. Thanks also to the Marstal District Heating, Flemming Ulbjerg, RAMBOLL, the designer of the Marstal plant and Jan-Olof Dalenb/ick for critical reading. REFERENCES
Altener Project (XVIF4.1030/AL/67/95/AUS): "Dynamic Simulation Model for Analyz~g Combined Solar-Biomass District Heating Plants", Vienna, April 1997.
187
Dahm, J.: "Evaluation of a Solar Heating System for a Small Residential Building Area", Document D39:1997, Dept. of Building Services Engineering, Chalmers University of Technology, Gothenburg, Sweden, 1997. DF software, InSitu Scientific Software, Wolfgang Spirkl, Germering, Germany. Driick H. and Pauschinger T., "Multiport Storage- Model for TRNSYS", Institut ftir Thermodynamik und W/irmetechnik, University of Stuttgart, Germany, 1995. Heller A. and Furbo S., "First Experience from the World Largest Fully Commercial Solar Heating Plant", ISES Solar World Congress, 24-30. August 1997, Taejon, Korea. Isakson P., "MFC 1.013 Matched Flow Collector Model for simulation and Testing, User's manual", Department of Building Services Engineering, Royal Institute of Technology, Stockholm, Sweden, 1993. Kristiansen F., "Provning af solfangers effektivitet og driftssikkerhed foretaget for Prgvestationen for solenergi, solfangerfabrikanten: AR-CON", Report 91-26, Thermal Insulation Laboratory, Denmark's Technical Highschool, 1993. Liu B.Y.H. and Jordan R.C., "The Interrelationship and Characteristic Distribution of Direct, Diffuse and Total Solar Radiation", Solar Energy, Vol. IV, July 1960, pp.l-19. TRNSYS, Reference Manual, Solar Energy Laboratory, University of Wisconsin-Madison, July 1996.
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A COMBINED EJECTOR COOLING AND HOT WATER SUPPLY SYSTEM USING SOLAR AND WASTE HEAT ENERGY
B.J.Huang Department of Mechanical Engineering, National Taiwan University, Taipei 106, TAIWAN Tel: +886-2-2363-4790 Fax: +886-2-2364-0549 e-mail: [email protected]
V.A.Petrenko Department of Mechanical Engineering, National Taiwan University, Taipei 106, TAIWAN and Odessa State Academy of Refrigeration, Odessa 270026, UKRAINE Tel: +886-2-2738-6997 e-mail: [email protected]
Abstract
In the present study, we proposed a combined solar ejector cooling and hot water supply system that simultaneously utilizes solar energy, condensing heat of ejector cooling system, and waste heat from boiler and generates cooling capacity and hot water for process application. It is shown that, for a factory with steam boiler capacity 10 tons/hr, the combined solar ejector cooling and hot water system can deliver a cooling capacity of 40 kW and process hot water at 92~ at a flow rate 20,000 kg/h. The ratio of the process heat supply to the cooling capacityfis shown to be larger than 35.
I.INTRODUCTION An ejector cooling system (ECS) can be powered by lowgrade energy at a temperature higher than 70~ (Huang et al, 1985; Shchetinina et al, 1987; Sun and Eames, 1995). Heat energy in this temperature range can be easily obtained from flat plate solar collector or industrial waste heat. In the previous study (Huang et al, 1998), we have developed a solar ejector cooling system (SECS) using R141b as the working fluid. The experimental COP for a ECS alone can reach as high as 0.5 at generating temperature fig) 90~ condensing temperature (T=) 28~ and evaporating temperature (Te) 8~ (Huang et al, 1999). For a SECS, we obtained a high overall COP, around 0.22. Food, beverage, textile and fiber industries use a lot of hot water in the temperature range of 50-100~ for the process. Hot steam and cold water consumption in these industries are also enormous. In addition to solar energy, waste heat exhausted from the steam boiler in the flue gas (around 200oc) can be used to power the ECS too. For a ECS, a low condensing temperature is required in order to obtain a high entrainment ratio of the ejector and the COP. Hence, the makeup water in industry provides a good medium for condenser cooling of a ECS. In the present study, we then propose a combined ejector cooling and hot water supply system (CECHS) using solar and waste-heat energy.
ECS, solar energy, and waste heat; (2) hot water is supplied for process application at a higher temperature 90~ (3) the solar 24~
Tc Pc CONDENSER
32~
EVAPORATPR EJECTOR
,9
8~
L COLD WATER FOR AIR-CONDITIONING
GENERATOR Tg
Pg
CITY WATER *
2.DESIGN CONCEPT OF A COMBINED SOLAR E J E C T O R S Y S T E M S F O R C O O L I N G AND H O T WATER SUPPLY
30 oC
90~
96~ 1
= AUXILIARY HEATER
m
6 ~0~ SOLAR
Shown in Figure 1 is the schematic diagram of the CECHS. The makeup water from city line or ground (24~ is used to cool the condenser of the ECS and heated to a higher temperature around 30~ The warm water is then heated by the solar collector to a temperature around 60~ In order to drive the ECS, waste heat is used to heat the hot wter up to 96~ so that Tg of ECS can reach 90~ for better performance. The hot water temperature at the exit of the generator of ECS remains at 92~ which can be used as a process heat source. The following typical features of the CECHS can be identified: (1) three energy sources are used, that are the condensing heat of
CITY WATER STORAGE
92~
STORAGE
30~
Fig 1 The schematic of the combined ejector cooling and hot water supply system using solar and waste-heat energy. collector efficiency can be maintained at high efficiency (0.70 in average) at low inlet water temperature (30~ (4) the COP of
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the ECS can be kept at its optimum value (around 0.48) at low condensing temperature (32~ (Huang et al, 1998, 1999); (5) cooling capacity is obtained simultaneously for air conditioning; (6)at the time when no cooling is required, process heat still can be supplied from solar and waste-heat energy. 3.DESIGN ANALYSIS OF A COMBINED SOLAR EJECTOR SYSTEMS FOR COOLING AND HOT WATER SUPPLY The COP of a ejector cooling system (ECS) is defined as the ratio of the cooling capacity Qe to the energy input to the regenerator Qg. The power input to the mechanical pump is negligible as compared with the heat input to the generator and, therefore, is not included in the theoretical calculations. Hence,
COPEcs = Qe
Qg
(1)
For a CECHS, the total quantity of heat absorbed by makeup water after passing through condenser, solar collector, auxiliary heater and generator is determined from the energy balance relation:
Qhw = Qc + Qsc + Qah - Qg = Qe + Qsc + Qah where the solar energy gain Qsc = rlscAsclT 9
(2) (3)
The total temperature rise of the makeup water is AThw = ATwc 4- AYwsc 4- AYah - ATwg
(4)
where ATwc,ATwsc and ATah are the temperature rise at the condenser, solar collector and auxiliary heater, respectively, ATwg is the temperature drop in the regenerator. From energy balance for the ejector cooling system (ECS), we obtain
Qc = Qe + Qg = Qg (1 + COPEcs )
(5)
ATwg =
(6)
Therefore,
ATwc 1+ COPEcs
substituting equation (6) into (4) we obtain
AThw = COPEcs ATw~+ ATws~ + ATah 1+ COPEcs
(7)
Equation (5) results in
~w =
Qe l + COPecs cwATwc COPecs
(8)
Therefore,
Qsc = n~wCwATsc = QeATwsc(1 + COPzcs ) ATw~COPEcs
(9)
189
Asc - QeATwsc(1 + COPEcs ), rlsclr ATwcCOPEcs
(10)
where Ir is the solar radiation intensity incident upon the collector slope. It is seen that the required collector area increases with decreasing COPecs and solar radation intensity It. The ratio of the process heat supply to the cooling capacityfis defined and derived as
f = Qhw =14 ATwsc(1 + COPecs ) + ~Qah Qe aTwc COPecs Qe
(11 )
The performance of a CECHS can be estimated from the above equations. Table 1 presents some of the designs of a combined ejector cooling and hot water supply systems.
Table 1 Design of a combined ejector cooling supply system. Hot water consumption, n~w (kg/h) 1,000 Solar collector area, A~c(m2) 64 Cooling capacity of ECS, Qe (kW) 2 Generation heat of ECS, Q~ (kW) 4.2 Condenser heat recovered from ECS, Qc 6.2
and hot water 2,500 20,000 158 1,280 5 40 10.4 84 15.4 124
(kW) Solar eners), heating rate, Qsc (kW) 31 Waste heat used, Qct (kW) 37.2 Hot water temp. for process, The,(~ 92 Process heat supply, Qhw(kW) 70.2 Fraction of process heat to cooling load, 35.1
77 93.4 92 175.5 35.1
620 744 92 1,404 35.1
f The results in Table 1 are obtained from the following assumptions. The cold water from city line or ground (24"C) is used to cool the condenser of the ECS and heated to a higher temperature around 30"C. The warm water is then heated by the solar collector to a temperature around 60"C. In order to drive the ECS, waste heat is used to heat the makeup water up to 96"C so that Tg of ECS can reach 90"C for better performance. The hot water at the exit of the generator of ECS is the lowered to 92"C which can be used as process heat. The unique features of the CECHS include: (1) three energy sources are used, that are the condensing heat of ECS, solar energy, and waste heart; (2) hot water is supplied for process use at a higher temperature 920C; (3) the solar collector efficiency can be maintained at highest efficiency (0.70 in average for solar radiation at 700 W m"z) at low inlet water temperature (300C); (4) the COP of the ECS can be kept at its optimum value (0.48) at low condensing temperature (32"C); (5) cooling capacity is obtained simultaneously for air conditioning. It is shown that, for a factory with steam boiler capacity about 10 tons/hr, a CECHS can deliver a cooling capacity of 40 kW and process hot water (92"C) at a flow rate 20,000 kg/h, assuming boiler efficiency 80%. The ratio of the process heat supply to the cooling capacity f is shown to be larger than 35. This means that CECHS is suitable for the factory that uses much more process heat energy than cooling energy. 4.CONCLUSION
Combining equation (9) and (3), we obtain the required solar collector area Asc:
In the present study, we proposed a combined solar ejector cooling and hot water supply system that simultaneously utilizes
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solar energy, condensing heat of ejector cooling system, and waste heat from boiler and generates cooling capacity and hot water for process application. It is shown that, for a factory with steam boiler capacity about 10 tons/hr, CECHS can deliver a cooling capacity of 40 kW and process hot water at 92"C at a flow rate 20,000 kg~, assuming boiler efficiency 80%. The ratio of the process heat supply to the cooling capacity f is shown to be larger than 35. This means that CECHS is suitable for the condition that the factory uses much more process heat energy than cooling energy. NOMENCLATURE Asc solar collector area, m2 COPecs coefficient of performance of ejector cooling system, dimensionless solar radiation incident upon the collector slope, Wm "2 /r mass-flow rate, kg s "1 n~ Q thermal energy, kW temperature, *C T temperature rise of makeup water, *C AT solar collector efficiency, dimensionless I"1sc Subscripts all auxiliary heater condenser C e evaporator g generator hot water hw inlet in outlet out solar collector sc water w REFERENCES
Sun Da-Wen, Eames I.W.(1995). Recent developments in the design theories and applications of ejectors - a review. J. Inst Energy 68, 65-79. Huang B.J., Jiang C.B. and Fu F.L. Ejector performance characteristics and design analysis of jet refrigeration system. ASME JEngngfor Gas Turbines and Power 107, 792-802. Huang B.J., Chang J.M., Wang C.P. and Petrenko V.A. (1999). A 1-D analysis of ejector performance. Int.J.Refrig. 22(5), 354364. Huang B.J., Chang J.M., Petrenko V.A. and Zhuk ICB. (1998). A solar ejector cooling system using refrigerant R141b. Solar Energy 64(4-6), 223-226. Shchetinina N.A., Zhadan S.Z. and Petrenko V.A. (1987). Comparison of the efficiency of various ways of heating the generator of a solar-ejector Freon refrigerating machine. Geliotekhnika, 23(4), 71-74.
Acknowledgement - The present study was supported by the National Science Council, Taiwan, ROC, through Grant No.NSC88-2212-E-002-050.
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A SOLAR STILL W I T H M I N I M U M INCLINATION AND C O U P L E D TO AN OUTSIDE C O N D E N S E R A. EI-Bahi and D. I n a n , Physics Engineering Department Hacettepe University, Beytepe, 06532 Ankara- Turkey Fax: + (90-312) 426 85 44 E-mail: [email protected] Abstract: An evaporator of a basin-type solar still was connected to an outside passive condenser, and the basin of the still was covered with a (3 mm) glass sheet inclined to the minimum. Direct solar radiation heated the saline water and evaporated it. A fraction of the resulted vapor condensed on the inner glass cover surface, while the rest was purged and diffused to the outer condenser to be condensed there. Depending on the energy balance equations a theoretical analysis was derived. The proposed solar still worked perfectly and the total daily yield was more than 6 l/m 2 .d, (sun radiates the still for 8 hours only, because of the obstacles). Solar radiation, ambient temperature and the temperatures at different points in the evaporator and in the condenser were measured accurately. The efficiency during June, July and August was improved up to 75%. When the solar still was operated without condenser the yield decreased to 70% of that with the condenser. Keywords: Solar energy, Solar still, Saline water, Solar distillation
1. INTRODUCTION Fresh water is the primary requirement to assist the life in our Universe. However while water covers about three quarters of the earth's surface, only 3% is fresh water in lakes, rivers and underground, and not all of this limited quantity is suitable for drinking, because of the salt concentration and environmental pollution. Water treatments are usually needed, and water desalination is necessary to provide fresh water from brackish or sea water. Survey of solar desalination systems was done by Sotres Kalogirou, 1997; and J. Ayoub and R. Alward, 1996; discussed the seriousness of emerging water scarcity and the desalination plants * ISES Member. already build in some locations in the world. In 1980; Malik et al., illustrated a several types of solar stills. A historical review for solar desalination was given by A. A. Delyannis and E. Delyannis, 1984. The shortage of fresh water, the pollution of drinking water sources and the increase of human population made the demand for fresh water to be increased rapidly. The desalination of saline water from the sea is the most attractive solution to overcome this serious water shortage. However desalination is energy intensive, and because of the scarce of wood quantity and oil crises, solar desalination is the promise cost effective solution as expected by many authors, where solar energy is renewable, safe, free and clean energy. Fortunately, in the areas where fresh water is more shortage, solar energy is more abundant during all of the seasons, the south of the Mediterranean terrain is an example. Among other types of solar distillation, solar stills are adequate for providing fresh water from the sea or brackish water, for a single house or a small committee, if the required quantity in the range of 200 m3/day. Although basin-type solar stills production is low, they have the advantages of simple
design, construction, and less technology, hence, easy to be maintained if required. In this paper a basin-type solar still with an extra condenser, A. E1-Bahi and D. Inan 1997 and 1998, was modified for more improvements, and it starts on the period, June to November 1998. 2. EXPER/MENT A basin-type solar still with minimum inclination and connected to an outside condenser (Fig. 1) was designed and constructed at Physics Engineering Department, Hacettepe University, Ankara ( 39 ~ 57' N )-Turkey. The solar still of 1 m 2 basin area, made of 0.35 mm galvanized iron sheet, covered with 3 mm transparent glass, inclined 4~ the south, the base (painted black) and the vertical sides of the evaporator were insulated from outside with 50 mm polyurethane foam and enclosed in a wooden box, to minimize the heat loss. Steenless steel reflector was added to increase the incident solar radiation on the glass cover and to make a shadow for the condenser. Thermoeouples were put in different positions to measure temperatures glass (Tg), ambient (TO, water (Tw), vapor (Tv) and that of condenser (Too). Solar radiation (Hs) was measured by a pyranometer ( CM 11/14 ) fixed at 5 cm above the glass cover. Water capacity in the still was 20 kg ( of 4% salinity ) of simulated sea water. Data Equisition System PCL-812 PG. reads data every 5 minutes and a P.C. unit stores all the measured data. The yield of fresh water was measured by a scaled flask over one hour period. Sun radiation absorbed directly by water and by the black base surface. The solar thermal energy heated and evaporated the saline water, a fraction of the resulted vapor condensed on the inner glass cover surface, while the rest was purged ( due to pressure difference ) to the outside condenser, where it condensed and collected at its base.
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Fig. 1. Schematic o f the Solar Still. 3. THEORETICAL ANALYSIS. An analytical expression for the hourly yield of the system has been derived and the following assumption have been taken into consideration: The physical properties of the water and water vapor are not change in the operating temperature range, the system is processed in a quasi-steady state condition, there is no water or water vapor leakage and the thermal heat capacity of the glass and the insulation is negligible. There is no temperature gradient along the thickness of the water mass or that of the glass cover, and the heat transfer coefficients are constant during the assumed small time interval. The energy balance equations Dunkle 1L V. [8], for the distillation unit components can be written as follows: for the glass cover, hl(Tw-Tg)=h2(Tg-Ta) (1) for the condenser, h'l(Tw-Too)=h'2(Too-Ta ) (2) for the water mass, xg (x~ ns +h3 (Tb-Tw)=MwdTw +hi (Tw-Tg)+h'l(Tw-T~o)
dt
(3) for basin liner, l:g~I-Is=h3(Tb-Tw)+h4(Tb-Ta) From equations (1-4) we can have:
(4)
+C~ ---zg~ ~+(~g~w h3I~)/(h3+h4)+CiT,
dr.
" - - " - +C ~=j'i(I-I,,T~)
(5)
dt
where, C,=(
hlh 2
hl'h2
h3h4
+ ~ + ~ ) / M w hI + h2 h; + h 2 h 3 + h 4
and fi= [a:gtxwI~ + 0;g *,wh3 Hs) / (h3+h4) +C1 Ta]/Mw
Equation (5) could be solved to give: Clt ) + TwO Tw=(1 e
e
Clt
(6) Two is the initial water temperature at time t = 0. The hourly distillate water can be given
by:
heg(Tw - Tg)+ heco(Tw - Tco) r~
L
*3600
(7) where, L is the latent heat of vaporization. By integrating equation (7), the total daily produced fresh water could be calculated. 4. RESULTS AND DISCUSSION The proposed solar still was first operated on June, 1998. However the distilled yield produced by solar distillation is directly proportional to the water temperature (vapor source) and inversely to the condenser temperature, in the processing still, the evaporator air space (average width 6cm) was minimized and the water depth was 2 cm in order to decrease the thermal capacity of the still. Solar energy heats the brine directly and through the black base, (a small fraction of this energy was reflected and absorbed at the glass cover), the heated saline water evaporates, and the resulted vapor partially transfers to the glass cover (inner) surface due to convection currents formed in the aperture between the water surface and the glass cover, the condensed fresh water collected by a tray at the lower side of the glass, while the rest of the vapor purges (due to the pressure deference between the evaporator and the condenser) to the condenser to be condensed and collected at its base. Solar radiation absorbed by the salty water and by the black basin-plate heated the brine up to more than 70~ at the midday, and the resulted vapor condensed at the inner glass surface and at the condenser. Fig.2 shows the solar intensity and the hourly yield produced on 12th
ISES Solar World Congress 1999, Volume III
distilled (6.52 1/m2.d) of a summer day are shown in Fig. 3. The temperature of the basin water, glass cover, condenser and the ambient versus time are shown in Fig. 4, where we can easily see the small difference between the brine and glass temperatures, i.e. the quantity of heat transfers to the glass by evaporation (which depends on the water and glass temperature difference) will be decreased and consequently more vapor transfers to the condenser, where its temperature is much lower than that of the glass and brine surface. The average temperatures of the brine, glass cover, condenser and the ambient temperature and their maximum values of a summer day(s) are shown in Table 1.
June 1998. It can be seen from this typical data that the total distillate increasing more sharply at the first four hours reaching a maximum at the hour 14:00, and the condensation continuous for a few hours after sunset, due to the brine thermal capacity. The glass cover temperature at the first period of operation was low enough for more vapor to be condensed on its inner surface (higher than that condensed at the condenser) while midday after the condensation continued approximately the same, up to sunset, where after that the condensation on the glass (it was again more cold) would be partially higher. The hourly accumulated yield collected by the glass cover and that by the condenser and the total fresh water
1400
'
I
'
I
'
I
'
193
I
'
I
'
I
'
1000
I
1200 800
60
1000 9 N
9..
600
E8oo E
;O .,,.o
O
- o9 6 0 o
400
400
3 200
200
9
I
8
'
I
10
'
I
12
'
I
'
14
I
16
'
I
18
'
I
20
T i m e ( hours )
Fig. 2. The solar intensity and hourly distilled water versus time.
0
194
ISES Solar World Congress 1999, Volume III
7000
6OOO
m
500O
m
"10 N
E E v
4000
"o (D
/
3000
i9 f - ~
'
'
i
.
'
'
i
.
i
"0
(D
2000
E 1000
i
.
10
I
,
I
12
.
14
i
.
16
i
.
18
20
~
Time ( hours )
Fig. 3. The daily accumulated fresh water by the condenser ( Doo),by the glass cover ( Dg ) and the total accumulated ( Doo+ Dg ).
80 70
0
60
Tw
50
Tg
v
v
40 (D
Q.
E (D
I--
30
,oF
20
0 00:00
,
I 04:00
,
I 08:00
,
I 12:00
,
I 16:00
,
I 20:00
,
i 24:00
Time ( hours ) Fig. 4. The temperatures of the brine Tw, glass cover Tg, condenser Too and the ambient Ta versus time.
195
ISES Solar World Congress 1999, Volume III
Table 1. The average and maximum temperatures of the basin water, glass cover, condenser and the ambient of a summer day(s). date of operation
water temperature (~
glass temperature (~
condenser temperature(~ )
ambient temperature (~
)
)
17. June 1998
63
60
52
28
21. July 1998
66
61
55
30
24. August 1998
67
64
57
33
Maximum values
80
65
60
36
other meteorology parameters. efficiency was calculated by,
The dependence of the efficiency on solar intensity, ambient temperature, insulation, wind velocity and water depth was studied by many authors. The efficiency is strongly depends on solar intensity and to a limit value on the ambient temperature, insulation depth and wind speed. The efficiency is increasing with increasing ambient temperatures is shown in fig. 5, it ranges from 65% to more than 75%. In this figure we can see some of the same efficiency values for the same ambient temperature improving the dependence of efficiency on
The
given
system
n(%)= Q,
H0
where; Qr ( MJ/mVd ) is the energy required to evaporate the specific daily yield and H0 (MJ/m'/d) is the measured total solar energy incident on the glass cover surface of the evaporator.
85 80+
+
75 + +
+
+
+
~~~"~'~'~+ +
+ +
o~ 70
~
O r
+
65-
W 60~
+
+
I
'
++
++
i
55-
'
I
26
24
'
28
I
30
'
I
32
'
I
34
Ambient tem perature ( ~ ) Fig. 5. The efficiency of the still versus ambient temperature.
5. CONCLUSION A modified basin-type solar still with limited thermal inertia is suitable for a single house to provide it with fresh water distilled from sea or brackish water. The proposed solar still is simple for construction and can be
built from local materials, and shows no complicated operation or maintenance. Purging a fraction of vapor to the integrated condenser minimize the pressure (inside the evaporator), and the formation of vapor droplets on the inner glass surface and consequently the reflection and absorbing of the solar radiation (in the vapor) are
ISES Solar World Congress 1999, Volume III
196
decreased, i.e. more solar energy is transmitted to the water, the glass temperature is also lowered and more evaporation is improved. The inclination of the glass cover was minimized and the unit responses to the solar radiation in a shorter time, the water temperature and the evaporation rate were enhanced. The efficiency of the solar still was improved up to more than 70%, and the distilled fresh water was up to 7 1/mE .d, for Ankara meteorological conditions and less than 8 hours sunshine, because of the surrounding obstacles. If this unit was operated in the south of Mediterranean terrain, a higher yield would be expected, where the meteorological conditions (solar intensity and ambient temperature) are more suitable for solar distillation. NOMENCLATURE g/c glass or condenser. lh heat transfer coefficient by convection. heat transfer coefficient by evaporation. hr heat transfer coefficient by radiation. hi heat transfer coefficient by evaporation, convection and radiation from water surface to glass (W/m s K). h'l heat transfer coefficient by evaporation and convection, from water surface the condenser (W/ms K). hz heat transfer coefficient by convection and radiation, from the glass cover to the ambient (W/m" K). h'2 heat transfer coefficient by convection and radiation from the condenser surface to the ambient (W/m s K). h3 heat transfer coefficient from the base plate to the water (W/ms K). h4 heat transfer coefficient from the base plate to the ambient (W/ms K). Mw thermal water capacity (J/K). water absorptivity. x g , Xw transmissivity of the glass and the water respectively. eg/c, ew emissivity of glass or condenser and water respectively. Stefan-Boltzrnan constant (w/ms 1(4). APPENDIX The heat transfer coefficients by evaporation, convection and radiation are defined as follows (J. A. Duffle and W. A. Beckman 1980):
( Pw - Pg/c )( Tw + 273) h~ = 0.88 [(Tw- T~/o~-
268.9X103 -- P~
]1/3
(T~ + 273) 4 - (Tg/~ + 273) 4 h~ = e,fr<~ [
]
where,
1 Eelf = [
ew
d-~--
1] "1
~g/c Pw m Pg/ c
he = 16.276 x 10"3xhewx ( and the radiation heat transfer coefficient to the ambient is,
(Tg/~ + 273) 4 - ( Ta + 273) 4 hr = Eg/eO [
]
REFERENCES 1- Sotres Kalogirou, Survey of solar desalination systems and system selection, Energy Vol. 22 No. 1 (1997), 6981. 2- J. Ayoub and R. Alward, Water requirements and remote arid areas: the need for small-scale desalination, Desalination 107 (1996), 113-147. 3- Malik M. A. S. et al. (1982), Solar Desalination, Pergamon Press, Oxford. 4- A. A. Delyannis and E. Delyannis, Desalination, 50 (1984), 71. 5- Hassan E. S. Fath, High performance of a simple design, two effect solar distillation unit, Desalination 107 (1996), 223-233. 6- A. E1-Bahi and D.Inan, Analysis of a solar still built at Hacettepe University, ISES 1997, Solar World Congress, August 24-29, 1997 Taejon, Korea. 7- A.E1-Bahi and D.Inan, Analysis of a parallel double glass solar still with separate condenser, Renewable Energy, 17 (1999) 509-521. 8- Dunlde R. V., Solar water distillation : the roof type still and multiple diffusion still, Int. Dev. Heat Transfer, 5, p. 895 (1969). 9- Duffle J. A., and Beckman W.A. Solar engineering of thermal processes. 2nd ed. New York: Wiley Interscience, 1991.
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197
MODELLING OF A THERMOSYPHONALLY DRIVEN DISCHARGE UNIT OF A STORAGE TANK Ulrike Jordan, Klaus Vajen, Brigitte Knopf, Astrid Spieler FB Physik, Universit~t Marburg, D-35032 Marburg, Phone: (49)/6421/282-4148, Fax: -2826535 [email protected]
Frank Hilmer Wagner & Co. Solartechnik, Ringstr. 14, D-35091 CSlbe, Germany A b s t r a c t - In this paper a natural convection discharge unit with an integrated regulation device for the storage water
flow rate is presented and discussed in terms of the dynamic behaviour of the discharge unit. Temperature oscillations are arising at the domestic water outlet. The influence of the domestic water flow rate and of the temperature distribution in the storage on temperature oscillations is shown. Solutions found for the reduction of the oscillations refer to the time delay occuring at the valve due to the heat capacities, the pressure drop coefficient of the storage water flow at the valve, and the auxiliary temperature. It is shown that the oscillations can be suppressed entirely. However there is a conflict of aims between the energy efficiency of the discharge unit, due to an enhanced stratification, and the reduction of oscillations by means of an increased auxiliary back-up temperature and a changed pressure drop characteristic.
1.
INTRODUCTION
In the recent years solar installations serving the hot water demand for both, domestic hot water (DHW) and space heating, called solar combisystems, were applied increasingly in Central Europe. Meanwhile, a wide variety of different system designs have been developed. Also, an IEA-Task (International Energy Agency) was established to work on this topic. In solar combisystems, additionally to the domestic water cycle of the traditional small solar plant (for one-family houses), another hydraulic cycle needs to be integrated. Unlike bigger solar plants, in which mostly two separated storage tanks are applied, a big variety of new combistores have been developed for small solar systems. Most of these systems are buffer stores. The domestic water is heated by the space heating water in the storage tank, possible with a tank-in-tank storage, or with external, or internal heat exchangers. The aim when integrating the domestic water cycle into the buffer store should be a good temperature stratification, in order to enhance the overall performance of the solar system (e. g. Sharp et al., 1979, and Phillips et al., 1982). Combistores with one or more internal heat exchangers placed at different positions in the store have been investigated by Dahm et al., 1998. Another new type of discharge units developed for solar combistores uses internal heat exchangers with a natural convection cycle as the primary flow within the store. Density differences arising inside the tank, produced by the heat exchanger as the heat sink, are used for the driving forces of the fluid flow. Several new systems, applying tipped coil- and spiral heat exchangers, were introduced by Driick et al., 1998, and Leibfried, 1998. In natural convection combistores the heat exchanger is placed in the upper part of the store (fig. 1). The heat exchanger is surrounded by a containment, with inlets for the storage water on top. The bottom of the containment is connected to a vertical pipe. During a discharge, hot water inside the containment is cooled. Due to its higher density, compared with the surrounding water outside the containment, the storage water flows down along the fins of the heat
Fig. 1. Natural convection combistore. Two water 'columns' are
Fig. 2. (a) Regulation device, composed of a valve placed in a vertical tube below the heat exchanger and a thin cylinder containing an expansible material. (b) angle of the valve ~ . (c) A mixer for the domestic water, customarily placed at the domestic water outlet.
exchanger and through the pipe to the bottom of the tank. In this way, two water 'columns' are formed, producing the driving forces for the circulation of the storage water. One problem arising in natural convection combistores is that the flow rate of the storage water strongly depends on the temperature distribution in the storage tank. In this paper a natural convection discharge unit with an integrated regulation device for the storage water flow rate is presented and discussed in terms of the dynamic behaviour of the discharge unit. The regulation device is composed of a valve placed below the heat
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ISES Solar World Congress 1999, Volume III
Fig. 3. Experimental setup. A 500 1 storage tank made of acrylglass is used. (a), (b), (c) : Positions of thermocouples.
Fig. 4. Measurements of the temperature stratification in the storage during a discharge cycle. Without valve (a) and with valve (b). Initial storage temperature: about 52~ DHW-flow rate: about 7.5 l/min (a) and about 10 l/rain
(b).
exchanger (fig. 2). The valve is connected with a thin cylinder containing an expansible material that is placed at the domestic water outlet. For stationary conditions, the angle of the valve is a (linear) function of the domestic water outlet temperature. Therefore, the pressure drop at the valve increases, if the DHWtemperature rises. This leads to a decrease of the storage water flow rate and storage water outlet temperature of the heat exchanger. In this way, high storage temperatures in the bottom part of the storage can be avoided. 2.
EXPERIMENTAL FACILITIES
The experimental set up of the storage tank is shown in figure 3. A 500 1 storage tank made of acrylglass is used. Thermocouples are placed at a vertical ledge in the tank, at the inlet and outlet positions of the heat exchanger, and at the bottom of the vertical tube. The flow rate of the domestic water is measured by a magnetic-inductive flow meter and the angle of the valve, with a potentiometer, placed at the axis of the valve. A comparison of the stratification of the storage water temperatures during discharge measurements done with and
Fig. 5. Regulation loop of the domestic water outlet temperature.
temperatures have been established. The domestic water was drawn with a constant flow rate of about 8 and 10 l/rain, respectively. 3.
MODELLING
The simulations were carried out with the object orientated simulation tool SMILE (Dehmel et al., 1997), which was developed at the Berlin Technical University. The differential equations were solved with a Runge-Kutta Method. An adaptive time step control was applied. The storage was modeled with 14 nodes. These storage temperature layers were used to describe the vertical temperature distribution in the storage, which is decisive for the driving forces. Due to the short time periods regarded for the simulations (several minutes), the heat transfer between the temperature layers as well as thermal losses were neglected. The regulation loop for the domestic water outlet temperature is shown in figure 5. The temperature distribution
Fig. 6. (a) Angle of the valve in dependence of the temperature of the expansible material for steady state conditions. (b) Response of the angle of the valve (prototype) to step changes of the DHW-temperature. Start-DHW-temperature: 62~ final DHW-temperatures between 45~ and 57~
ISES Solar World Congress 1999, Volume III
Fig. 7. Comparison of measured and calculated functions of the outlet temperatures of the heat exchanger. T~,r was measured at the top and at the bottom of the vertical pipe in the storage. The positions of the thermocouples are shown in figure 3. in the store, the DHW-flow rate, and the inlet temperature of the domestic water cause disturbances in the DHW-temperature. The loop can be subdivided into two parts. First, a loop is formed by the valve, the pressure drop component, and the heat exchanger. The angle of the valve reacts to a change of the DHW-temperature with a certain time delay and then influences the pressure drop of the water in the vertical pipe. This effects the storage water flow rate, which again influences the performance of the heat exchanger. The regulated value of the first loop is the DHW-temperature. A second loop is formed by the heat exchanger, the driving forces, and the pressure drop. The driving forces of the primary flow of the heat exchanger depend on the temperature distribution in the storage as well as on the temperatures in the enclosure and the vertical pipe. The latter strongly depend on the heat transfer to the domestic water and the friction acting on the flow. The regulated values of this loop are the velocity and the temperature of the storage water in the vertical pipe. A common approach to model a density driven circulation loop is the one dimensional steady state pressure equation for incompressible flow, as the balance of the pressure and frictional forces (e. g. Dahl et al., 1998). g (Ppipe-Pstore)dh=
pipe
P pipe 2 Pstore 2 2 V pipe -- ~store 2 V store
with lPpipe >> Vstore, and p , v
(1)
~ = ~(q~,~stor,d,...),
199
Fig. 8. Measurement of the DHW-temperature, the angle of the valve, and the DHW-flow rate. The time resolution is 5 s.
On the left hand side, the driving forces for the storage water flow in the discharge unit are expressed by the density differences of the water columns shown in figure 1. The fi'ictional forces on the right hand side are described by the kinetic pressure drop of the storage water. The pressure drop coefficient is a function of the geometry of the discharge unit, the angle of the valve, as well as the flow rate of the storage water. Due to the fact that the velocity of the storage water in the vertical pipe is much higher than the upwards velocity of the storage water in the tank, the second term on the fight hand side in Eq. (1) can be neglected. For stationary conditions there is a linear correlation between the angle of the valve and the temperature of the expansible material (fig. 6a). The time delay T occuring for the reaction of the valve to a change in the domestic water outlet temperature, caused by the heat capacity and conductivity of the expansible material, can be described by Eq. (2). "Cq~+ ~ = k ( Tdom,out - T b )
(2)
The value for T was determined experimentally by the response of the angle of the valve to step changes of Tao,,~o.t(fig. 6b). The heat exchanger is modeled with the energy balance (Eq. (3a) and (3b)). Whereas the differentiation of the time dependent variable was carded out by the simulation tool SMILE, a discretization in the space coordinates was necessary. A one-dimensional approach with 20 nodes for each fluid, using backward differences was applied. In this way, the heat exchanger was treated as a counter flow heat exchanger.
(3)
and ~" : density, velocity, and pressure drop
coefficient, respectively, of the storage water outlet temperature below the heat exchanger in the vertical pipe (index: pipe) and of the surrounding storage water in the tank (index: store).
C ~om .Ord~.
bt
i
~TsLr
C stor .
Ot
. UA. ( r~o m
. = v A . ( rAom.
i ) - J~dom Cp "( rdom - r~om) i-1 rstor
r;Lr . )+
,,orC .
(r;'to
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ISES Solar World Congress 1999, Volume Ill
Fig. 9. Calculated DHW-temperature during a discharge. Initial storage temperature: 60~ domestic water flow rate of 5, 10, and 201/min.
In the model, the vertical pipe was divided into four nodes. Since the valve causes the storage water flow rate to approximate the domestic water flow rate, the heat capacity rate (UA) was described as a function of the domestic water flow rate, solely. Validation of the model is shown in figure 7. The DHWtemperature and the storage water temperature measured during a discharge at the top and at the bottom of the vertical tube (at the positions a, b, and c in figure 3) are shown. The time delay of the temperature drop at the top and at the bottom of the pipe as well as the values of the temperatures show good correspondence. 4--
PROTOTYPE CONTROL DEVICE OF THE STORAGE WATER FLOW RATE
The time delays, arising due to the heat capacities of the system's components, cause oscillations of the DHWtemperature at the beginning of a discharge. A measurement made with a prototype of the valve is shown in figure 8. The flow rate of the domestic water was about 15 l/rain. For the measurements, no mixing devices at the DHW-outlet were employed, which are usually applied in solar water heating systems (figure 2c). At the be~nning of the draw-off, the domestic water outlet temperature has nearly the same value as the storage water temperature in the containment, as long as the water content of the heat exchanger is taken. Meanwhile, the storage water inside the containment is cooled, since there is only a small flow rate of the storage water due to a high pressure drop caused by the valve. This leads to a minimum of the outlet temperature, 5 K below the value in equilibrium. In the following, the influence of parameters concerning the valve's operation will be discussed.
4.1 Domestic waterflow rate The influence of the DHW-flow rate on the minimum DHWtemperature T=~, is shown in figure 9: The higher the DHWflow rate for equal initial storage temperature distributions, the higher the amplitude of the oscillation. The DHW-temperature under stationary conditions T ~ increases, if
Fig. 10. Calculated DHW-temperature with and without mixer. Without the mixer, the flow rate in the heat exchanger is constant, whereas with use of the mixer, the total DHW-flow rate discharged is constant. In this case a variing part of the water is bypassing the heat exchanger.
the DHW-flow rate decreases. This also means, that the higher the flow rate, the less energy is inserted into the bottom of the storage tank during a discharge.
4.2 Storage water temperature dis~'bution Although there is a high amplitude of the oscillation (T~,Train) for high storage temperatures and high flow rates, the absolute value of T~. is smaller for low than for high storage temperatures (for fully mixed tank): T~.(95~
> T~.(60~
for I ~ m < 22 1 / min
However, if two or more different temperature layers are taken into account, no uniform tendency can be formulated concerning the behaviour of the minimum value of the DHWtemperature in dependence of the mean storage temperature. At a homogeneous storage temperature of 95~ and a flow rate of l0 1/min the curve is critically damped, for smaller flow rates the curves are overdamped. Due to the influence of the storage temperatures on the oscillations, the set temperature for the auxiliary heater always has to be considered, when the behaviour of the valve is investigated.
Fig. 11. Calculated DHW-temperature oscillations for different time constants: x = 20 s, 10 s, and 0 s (theoretical case) and with a differential component (PD). Initial temperature of the storage: 60~ DHW-flow rate: 15 l/min.
ISES Solar World Congress 1999, Volume III
4.3 Mixer Measurements done with a standard mixing device show that temperatures above the DHW-set temperature can be smoothed, temperatures below the DHW-set temperatures, however, can not be avoided. Simulations done with two different flow rates show, that Tmi, only increases slightly, whereas the time, at which the minimum occurs, is shifted (fig. 10). 4.4 Time constant The influence of the time delay of the regulation loop caused by the heat capacity of the expansible material is shown in figure 11. The simulations show, that the values for Tmi, can be decreased distinctly by a reduction of the time constant and an improvement of the heat transfer beween the DHW and the expansible material. Other than that, the oscillations could be avoided with an ideal regulation device, containing a differential (e. g. electrical)component. 4.5 Pressure drop characteristic Furthermore, the pressure drop characteristic, resulting from the geometry of the vertical pipe and the containment, influences the oscillations. Simulations show, that with a change of the pressure drop characteristic, the oscillations can be avoided (fig. 12). However when the oscillations are reduced, the values of the storage water flow rate, the storage water outlet temperature, and T~q= rise. Therefore, there is a conflict of aims between the efficiency of the valve, influenced by the storage water outlet temperature and the comfort for the user due to a well damped course of the DHW-temperature.
5.
CONCLUSIONS
A regulation device for the density driven storage water flow of the discharge unit of a combistore was investigated. Measurements showed, that the stratification can be improved decisively with the regulation device during a discharge cycle. The investigations presented were focused on DHWtemperature oscillations caused by the regulation device. The oscillations can be described well by a model, composed of the one-dimensional pressure drop equation, a differential equation to decribe the dynamic behaviour of the valve, and a 20 node heat exchanger model. The validation of the model showed good accuracy between measured and calculated values. The influence of the domestic water flow rate and the temperature distribution on the oscillations were shown. The temperatures at the top of the storage came out to play an important role for the oscillations. With a mixer, which is usually applied in a solar system, the DHW-minimum temperature could not be increased perceptibly. Values which directly influence the oscillations are the time constant, describing the time delay of the valve due to the heat capacity and conductivity of the expansible material, and the pressure drop characteristic of the storage water flow in the vertical pipe. The time constant of the valve's reaction could be reduced by more than 50 % compared to the prototype. By changing the pressure drop charcteristic, a sufficiently damped course of the DHW-temperature can now be reached and the oscillations can be avoided. Furthermore, the simulations showed that there is a conflict of aims between the energy efficiency reached by the system
0 o
=3
57
----------
55
............................
201
53
t~ L_ ID
51
E
49
............................
(D "*7'
47
............................
I
45
prototype
121 43
~
0
30
60
90
120
150
180
210
time / s
Fig. 12. Calculated DHW-temperature for different pressure drop characteristics of the valve. Initial storage temperature: 60~ DHW-flow rate: 10 l/min.
and an extensive reduction of oscillations. Latter can be reached by a higher auxiliary temperature or by changes of the pressure drop characteristic. NOMENCLATURE % C d g h k
specific heat capacity, J/gK heat capacity, J/K diameter, m gravity constant, rn/s2 storage height, m transfer coefficient, 1/K
~: T UA v
mass flow rate, kg/s temperature, K heat transfer coefficient area, W/K velocity, m/s
I&
volume flow rate, mVs
Greek P (o T
Indices b dom, out dom equ min pipe stor store
density, kg/m3 angle of the valve, o time constant of the expansible material, s pressure drop coefficient
begin of the regulator temperature interval of the valve domestic water outlet domestic water in the heat exchanger nodes equillibrium at minimum of oscillation inside the vertical pipe storage water in the heat exchanger nodes storage water layers
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ISES Solar World Congress 1999, Volume III
REFERENCES
Dahl, S. D., Davidson, J. H. (1998). Mixed Convection Heat Transfer and Pressure Drop Correlations for Tube-In-Shell Thermosyphon Heat Exchangers with Uniform Heat Flux. Journal of Solar Energy Engineering. Vol. 120, pp. 260269. Dehmel, K.-H., Klein-Robbehaar, C. (1997) Universit~rer Forschungsschwerpunkt 4: Entwicklung rechnergestiitzter Simulationshilfsmittel zur Beschreibung des Betriebsverhaltens komplexer energiewandelnder Systeme, Berlin. Driick, H., Hahne, E. (1998) Test and Comparison of Hot Water Stores for Solar Combisystems. In Proceedings of EuroSun ISES Europe Solar Congress, 14-17 Sept. 98, Portoroz, Slovenia, Vol. 2, III.3.3. Dahm, J., Bales, Ch., Lorenz, IC (1998). Evaluation of Storage Configurations with Internal Heat Exchangers, Solar Energy, Vol. 62, No. 6, pp. 407-417. Leibfried, U. (1998) Kombispeicher mit ThermosiphonWErmetauschem ftir Warmwasser. In Proceedings, 8. Symposium thermische Solarenergie, 13-15 May 98, Staffelstein, Germany, pp. 39-44. Phillips, W. F., Dave, R. N. (1982). Effects of Stratification on the Performance of Liquid-Based Solar Heating Systems. Solar Energy. Vol. 29, No. 2, pp. 111-120. Sharp, M. K., Loehrke, R. I. (1979). Stratified Thermal Storage in Residential Solar Energy Applications. Solar Energy. Vol. 3, No. 2, pp. 106-113.
ISES Solar World Congress 1999, Volume III
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PERFORMANCE OF TRANSPARENTLY INSULATED SOLAR PASSIVE HOT WATER SYSTEMS N a r f n d r a D. Kaushika and Kalvala S. Reddy Centre fi)r Energy Studies, Indian Institute of Technology Delhi, Hauz Khas, New Delhi - 110 016, India, Ph. +91 -i 1-686 ! 977 Ex. 5006. Fax" 9 l- i 1-6862037, [email protected]
Ab~lract - This paper presents the design and performance characteristics of TIM insulated solar ICS water heaters with water/ground/ .,,and and concrete as storage materials; it uses computer simulation approach based on accurate determinations of the criteria of convection suppression, solar transmittance and thermal loss reduction characteristic.,, of TIM device. The simulation model is validated with prototype field experimental ob.~rvations and is used to evolve the optimum system design and trade-off characteristic.,,. The TIM cover system characteristics significantly influence the overall system perii~rmance. The effectiveness of several configurations of TIM cover system as a comparative study, has therefore been investigated. The results seem to favour absorber perpendicular (honeycomb) configuration over others. The system performance tends to level at a honeycomb cover depth of 7.5 cm. Compounding of honeycomb w~th an air layer 112 mm thicknes.,,) tends to improve the performance, the honeycomb cover depth of 5 cm is near optimum in this c~mliguration. The absorber parallel configuration is simple in practical realisation: it may be recommended for application in passive .~olar water preheaters. TIM insulated ground integrated collector storage water heating system has also been investigated.The system consists of a network of pipes embedded in a concrete slab whose top surface is blackened and covered with TIM device and bottom is insulated by the ground. Solar gain (solar collection efficiency of 30-50 ch corresponding to temperature of 40-60 ~ and diurnal heat storage characteristics of the system are found to be of the right order of magnitude for solar air/water heating application.,, 1. INTRODUCTION For many years now air-filled honeycomb devices have been considered for use as transparent insulation in solar ponds ( Lin. 1982; Ortabasi et al.. 1983; Kaushika et al.. 1983:Sharma and Kaushika. 1987:Schaefer and Lowre). 1992). hot water systems (Kaushika and Banerjee. 1983; Goetzberger. 1984; Rommel et al.. 1987). buildings (Goetzberger. 1984: Kaushika et al.. 1992) and other integrated collector-storage systems (Gordon. 1987:Kaushika et al. 1990). These applications have opened up a .,,ignificant area of developing transparent insulation matenals in parallel with the conventional(opaque) insulation materials.The configurations based on the storage water tank seem yen.' suitable for domestic and industrial applications. Kaushika and Banerjee (1983) suggested and analysed the configuration, which consists ,)t storage tank 'Cul'rt)id' in shape, transparently insulated at the surface and covered with opaque insulation at all other sides. Subsequently, Goetzberger and Rommel(1987) examined the prospects of such a system for application in central Europe. A cubic storage water tank using transparent insulation at its surface as well as side walls (Kaushika and Sharma. 1994) and a simulated well stratilied tank made of tubular subunits (Schmidt et al. 1988: Schmidt and Goetzberger. 1990) have also been considered and solar energy gain up to 40% has been reported. The ICS solar water heater having water tank of triangular cross-section and with transparent insulation such as Methyl Methacrylate(MMA) on the top and sides has been studied by Prakash et al. (1994). The above thermal analyses/evaluations of the performance of transparently insulated ICS solar water heaters are in general based on discrete measurements of solar transmittance and heat losses across the TIM cover system. In simulation models very little attention has been paid to the tbrmulations of ~lar beam and diffuse radiation transmittance and thermal loss reduction characteristics of the TIM cover system. This paper is Intended t() present design and perlormance data ~dTIM insulated solar ICS ~ater heater.,, with water, ground. ~and and concrele a.,, ,,forage material.,,: it u.~s c()mputer simulation approach ba.,,ed on accurate deternunatlons of the criteria ol convection suppresslt)n, solar
transmittance and thermal loss reduction characteristic.,, of TIM device. 2. T H E S I M U L A T I O N M O D E L Consider a TIM insulated integrated-collector-storage solar water heater which involves the solar heating of storage water tank. cuboid in shape, having TIM cover on the top surface and opaque insulation on all other surfaces. The solar radiation, after transmission through the TIM cover is absorbed by the top surface ()f tile tank. Part of the absorbed energy is used to heat the water and the remaining energy being lost to the surroundings by conduction, convection and radiation. The energy balance Ior water can be written as:
dTw(t) M
-
S '(t)- Qt.(t)
(1)
dt or
dT(t)
(2)
+ E T ,(t)= F(t)
dt The total heat loss(QL(t) ) from the storage water is given as:
Qc(t)
=
U L[T(t)-
T(t)]
(3)
Where U L is overall heat loss coefficient and is expressed as: U,_ = UT+UR+U.~ (..I)
u,.
E:--, M and
F(t)=
s :(t). U,.T(:t)
M-MC
M .M,C,
Eq..(2) is linear differential equation with integration factor eEr. Applying initial condition T.(t) = T . , at t = O. At small interval of
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time(t). F(t) may be regarded as constant (F). The solution is obtained as : l
T(t):
~-F[I-c
-El']
- T
e
.wr
(5)
This expression may be used to calculate mean water temperature as a function of time. The radiant energy(S'(t)) reaching the absorber plane at time t is given by:
S '(t) - Ib(t) 1~ (z a)b 9 I,(t) R, (, a). 9 [Ib(t) - l,(t)] R (x a ) .
(6)
The fiwmulations of transmittance-absorbtance products corresponding to beam radiation (I: a)h. sky diffuse radiation ( 1:a )a, and ground diffuse radiation ( z a ) ~ for the cover system embodying the TIM device were discussed in Kausika and Arulanantham ( 1996)and Kaushika and Reddy (1999) and have been adopted in the present work. The UL values may be computed by using the concept of thermal network.The steady-state energy transfer between the absorber surface at 1", and the bottom cover of TIM at T, invoives the heat transfer (a) through TIM (compound honeycomb), (b) between top of the TIM and tempered glass cover and (c) between tempered glass cover and ambient air.The collection efficiency of the system ts the ratio of energy collected by water mass to solar radiation received on absorber plane during time t and, is given by :
the bolt on cross-bar according to the requirement of placing the absorber plane perpendicular to solar rays at noon. The top surface of the tank is painted with black paint to absorb the .q)lar radiation and covered with TIM to reduce the heat losses. The temperature of water inside the tank is measured by Copper-Constantan(typeT, Copper(+)/Constantanr thermocouples. The CopperConstantan thermocouple wires are arranged in the form of a multichannel probe, wherein the junctions are placed at a separation of 10cm to measure the vertical temperature distribution in the water tank. The thermocouple probe is connected to a SC-7501 multi logger (IWATSU ELECTRIC Ltd., Japan). The water temperatures have been recorded at an interval of one hour. The water was not drained out from the tank for two days. The time-history of temperature development in the tank at various heights was recorded.The experiments were carried out tn December ( Dec. 11-13, 1996, a winter month) at New Delhi. The temperature gradient builds up during the day and tends to diffuse during the night. The system exhibits significant retention of heat dunng off-sunshine hours. The experimental variation of mean temperature of water in the tank ts compared with simulation results which indicate an excellent agreement between the experimental observations and simulation model results (Fig. ! ).
t
fQ.(t)dt
q_
(7)
o
t
A fs(t)dt 0
3. EXPERIMENTAl, VALIDATION O F SIMULATION MODEl,
3. ! Fabrication of TIM Cover System The TIM cover system for the proposed water heater is fabricated from extruded cellular strips supplied by ArEI Energy Lid, Israel. The product is in the form of cellular strips of (70 x i.6 x 5 cm) and (70 x !.6 x 10 cm) sizes. The square cells are of widths 3 mm and 4 ram. The strips were glued to form a cellular array. The gluing process was carried out manually using pure liquid Chloroform (CHCI~, which is a solvent for lexan. Agarwal Chemical Industries, New Delhi) as adhesive. The cellular matrix was finally encapsulated in a tray made of transparent polycarbonate sheet of 0.5ram thickness. In the covet" system, an air layer of 12ram is maintained by placing the structured sheet ribs in the bottom of the tray.The cover system is easy to handle and has sufficient built-in strength to maintain rigidity 3.2 Experimental Set-Up A prototype field experiment system has a tank of 25 Iitrcs (66 x 45 x 8.5 cm) with a rectangular cross-section; it is made from an 18 gauge galvaniscd iron sheet, it is covered with mineral wool insulation at its sides and bottom and encased in a w(mdcn box. It ts kept in an inclined position and facing due south. The inclination of the system can be adjusted manually by changing
Fig. 1 4. SYSTEM O P T I M I S A T I O N AND TRADE-OFFS Experimentally validated simulation model may be used for the derivation of optimum design parameters and trade off characteristics. For a given solar absorber area, the temperature in the tank will vary with the capacity of the tank. These variations of the storage water temperature as a function of capacity of the tank, selective and black absorber characteristics, cell width and thickness of TIM configuration are portrayed in Figs.2 (a, b). The results may be used for trade off between the system efficiency and required hot water temperature. The effectiveness of a honeycomb cover system on the development of mean water temperature in the tank has also been investigated. Two configurations of the cover system are considered: (i) honeycomb cover system which consists of the encapsulated cellular array (ii) compound honeycomb cover
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5. PERFORMANCE COMPARISONS
Fig. 2 (a)
The ~lar transmittance and heat loss reduction characteristics of TIM cover system considerably influence the overall thermal performance of ICS solar water heating system. We therefore investigate the effectiveness of various kinds of TIM cover systems. The cover system consists of TIM device compounded with mr layers of near critical Rayleigh regime at its top and/or bottom. Following configurations of TIM device placed between the top tempered glass ( S m m thickness) cover and the absorber have been considered. I. Absorber-Parallel Configuration of T I M Device (CI) Air Layer (near criticalRaylcigh regime) - single cover (C2) Glass Sheet (2 m m thickness) (C3) Polycarbonate Sheet (0.5mm thickness) (C4) Double Wall Structured Polycarbonate Material (6ram thickness, G E Plastics) (C5) Double Wall Structured Polycarbonate Material (10ram thickness. G E Plastics) 2. Absorber- Perpendicular Configuration of T I M Device (C6) Cellular Array(Honeycomb) of 5 cm thickness ~C7) Encapsulated Cellular Array of 5 c m thickness compounded with 12 m m air layer (C8) Encapsulated Cellular Array of 10 cm thickne.,,.,, compounded with 12 mm air layer All the configurations have been experimentally tested as well as analysed by the simulation model. The mean temperature of water in the storage tank as measured experimentally as well as predicted from the simulation model is considered. A comparison of performance characteristics of absorber parallel configurations of TIM in terms of solar gain efficiency defined by equation (7) is presented in Table-!. Results indicate an excellent agreement between the experimental ob~rvations and simulation model results. The performance of solar ICS water heater with cover system embodying a double walled structured sheet of 10ram (GE plastic product) as TIM excels over others. The comparison of performance characteristics of ~lar ICS water heater with absorber- perpendicular structures of TIM cover systems is summarised in Table-2. The cover system C7. made of comp()und honeycomb of 5 cm thickness and having 12 mm air laycl ~ts top and bottom corresponds to relatively higher solar collection-storage efficiency of the water heater. Table-i Performancecomparison of transparently insulated (absorberparallel structure) solar ICS water heater. r
Fig. 2 (b)
syslem which consists of an encapsulated air layer (12mm thickness placed in the bottom region) and the cellular array; in this geomeffy the air layer remains in near critical Rayleigh regime and provides additional insulation without affcclfing the solar transmittance of the cover system. The system performance tends to level at a honeycomb cover depth of 7.5cm. Compounding of honeycomb with an air layer tends to improvc the ped'ormance: the honeycomb cover depth of 5cm is near optimum in this configuration.
Dine of Exp.
Insolation kWh/day-
TIM
P e r f ~ c c Characteristics
m"
28-I 1-96 01-12-96 30-I 1-96 02-12-% 13-05-98
5.28 5.96 5.91 6.32 6.15
CI C2 C3 C4 C5
T,,
T,,,~
T,,
qc %
14.2 14.0 15.3 15.5 29.0
47.7 52.8 51.2 50.5 56.0
23.1 29.5 31 7 31.4 47.0
14.33 22. II 23.59 21.40 30.97
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Table-2 Performance comparison of Tmnspareutl), insulated (absorber- perpendicular structure) solar ICS water heater.
/ Dme of Exp.
Insolation kWh/dayI112
TIM
Performance Charmm27 May | 1999sties J Storage water temp. 'C
I 1-12-9~
5.49
C6
09-12-~
5.72
C7
05-,2-9(
5.25
C8
T.,
r.~
T.~
9.2 9.0 li.0
47.6 ~,5., 39.3
31.8 33.1 Y).7
(qc) qt
on a winter day at New Delhi. The adjustment of the inclination is often considered as a liability by the users, in the cuboid shape, the tanks of large capacity may also be used: they may be segmented or made from tubular subunits to simulate well-stratified tanks of good structural stability. In fixed tilt configurations several design variations are possible. For example one could use a triangular shaped tank wherein the honeycomb and absorber plane are inclined say perpendicular to sun rays at winter noon.
35.35 35.82 31.10
in Fig.3 the diurnal variations of storage water temperature as obtained from the validated simulation model for all the configurations arc portrayed. All the variations correspond to same radiation and atmospheric air data shown therein. These results further support the inferences of preceding sections. The relattve merits of various TIM cover systems can also be judged l?om their ~)lar transmittance (': a) and heat loss reduction ~t.',.) characteristics. For this purpose, the effective transmittanceabsorptance product(t a h, of TIM cover systems is experimentally measured at normal solar angle of incidence. The Ut is theoretically calculated from thermal network analysis and experimentally evaluated from mght-time cooling of water temperature in storage tank as follows: During the oil-sunshine ~eriod. the solar intensity term S'(t) is lave Ter'it;' :q;'o'i:f'n: ' _ : : i w: '
,T)]
(8)
From eq.(4, i ). night time Ut. may be expressed as: [
Ut
=
_M T
T(t)- T ( t ) lr~ [ T,,(t-1)- T(t)
(9)
Where M = M,, C,, + M, C, The theoretical and experimental values of overall heat loss coefficient and transmittance-absorptance product for TIM cover systems are summarised in Table-3.The multiple wall structured sheet provide good heat loss reduction but simultaneously cut the solar radiation whereas the TIM devices made of absorber perpendicular structure provide good heat loss reduction as well as high .solar transmittance resulting in relatively higher solar gain of the hot water systems.The results on system performance charactenstics as well as UL and (a :)dr values seem to favour absorber perpendicular configurations over absmlmr parallel configurations. The integrated-collector-storage units have often been used as solar water preheaters. Our expenments were perfmmed during the December (winter month at New Delhi). The indinmion of the tank was adjusted such that the absorber plane was peqmndicular to sun rays at noon. It was found that the hot water lemperatures of 50-60'C could bc attained with solar energy alone. It has also t~en found that the unit using encapsulated compound lameycomb ;L',TIM holds promise of use as 100% solar fraction water heater
Fig. 3
C i- Air layer : C2- Glass sheet: C3- Polycarbonate sheet C4-Structured shee (6mm); C5- Structured sheet (10 ram) C6-Ceilular array (5cm): CT-Encapsulated TIM (Scm) C8- Encapsulated TIM ( I 0cm) Ib(t)- Beam radiation Id(t)- Diffuse radiation Ta(t)- Ambient temperature
6.TIM INSULATED GROUND ICS WATER HEATER ~
6. i System Configuration And Approach The concrete-ground integrated-collector-storage system consists of a concrete slab. a pipe network and is placed in the ground. The PVC pipe-network is of 20ram outer diameter and l.Smm wall thickness "and embedded inside the concrete slab. The top surface of the slab is painted with black (c = 0.9) paint to absorb solar radiation and covered with Transparent Insulation Materials (TIM) device to reduce the top heat loses. The TIM cover system
ISES Solar World Congress 1999, Volume III
consists of a compound honeycomb. The ground b assumed as semi-infinite media and has a low thermal conductivity to provide adequate insulation from the bottom. The side surfaces of the collector are perfectly insulated. Solar radiation passes through TIM cover and is absorbed at the absorber plane. The absod)cd energy is transmitted to the fluid(water) flowing Ihrough the pipe network. ,An exploded view of ground TIM -ICS solar water heating system is illustrated in Fig.4. The temperature distribution in the ground ICS system may be estimated by solving the Fourier heat conduction equation with appropnate initial and boundary conditions characteristics of system geometry.The ground ICS is assumed to have constant therrr~)physical properties. In this numerical method, the temperature ts calculated at certain discrete points of space and time. The space and time derivatives are converted into finite diflcrcnces and a set or linear simultaneous algebraic equations arc obtained which may bc solved by matrix algebra (Reddy c1 al.,
207
20cm; Ihe variation of heat gain efficiency as a function of slab thickness al.~) supports the infercnoc. It is advL~blc to consider a slab thickness of 10-25cm In this regard, wc have. therefore.
1998). 6.2 R.'.~uhs and Discu.vsion The above mathematical model may be used to evaluate the thermal performance of TIM insulated ground ICS solar w a l ~ Table-3 Thermal and optical characterktics of v i r i o ~ TIM cover systems for ICS solar water heater.
Tl~'oretical (W/re" "C)
Exp.
TIM
(Who" "C)
Covcr
UT Ci
C2 C3 C4 C5 C6 C7
C8
6.23 3.73 3.43 2.63 2. I I 1.93 1.73 !.33
UI,
U~
Uc
0..S9 0.59 0.59 059
0.38 0.38 0.38 0.38
7.20 4.68 4.40 3.60
059
0.38
3.08
0.59 0.59 0.59
0.38 0.38 0.38 ,,
2.90 2.70 2.30
,,,
(T=),~ at 8=O
UL
7.9:t.0.25 5.6:t.0.10
0.7:59 0.658
4.5r
0.696 0.565 0..S65
4.4:t.0. I I 3.8:t,0.12 3.4:t0.13 2.8r 2.6~-0. I$
0.625 0.588 0.430
l:ig. 4
heating system. The values of thermophysical parammers used in the analysis are as follows: (a) Cover system:Compound Honeycomb (b) Concrete:K, = 1.75 W/m K, C, = 880 J/kg"C and p, = 2242 kgJm' ( ~ Ground: K~ = 0.56 W/m K. C~ = 1840 J/kg"C and Pc = 20:50
considered concrete slab of 25cm thickness for further computer runs. The final water temperature corresponding to concrete slab thickness of 25cm and the pipe network at different depths are sho~) in Fig.@ It is observed that the rise in water temperature is higher at lower depths but the thermal storage effect is better at
k~m'
larger depths: middle of the slab is considered a good compromise.
(d) Common Brick:K,, = 0.72 W/m K. C, = 1884 J / k ~ and p, = 1922 kg/m ~ (d) Iron impregnated sand:K, = 2.41 WIm K, C. = 2860 J/kg"C and p,, = 2466 kg/m -~ (e) Mass flow rate of fluid:0.001 kg/s The solar radiation data and ambient temperature data used m simulation model are corresponding to the month of December 1996. The diurnal variations of final water tempcf'~urc for the surface slab of 4(km thickness and Ih pipe network placed at various depths for dillcrcnt collection-storage materials arc tlluslralcd in Fig.5 (a.b.c). The u.~ of highly condm.live storage material n)provcs the thermal performance. For exan~le, iron impregnated sand (ratio I: 19) as collection.storage malerial of Md,lr ICS system cxc'cls over ground, c o l l l n l o n brk.'k and c[mc'rcle. ll~ u.~ of such material incrca.~s the total system r The rt.~tn waler tcmwraturc ts negligible for the slab thickness ove[
208
ISES Solar World Congress 1999, Volume III
Fig, 5(a) Concrete
Fig, 5(b) Iron im1~n~gn~i~d
Fig 5 (c} Total ground
Fig~ 6
ISES Solar World Congress 1999, Volume III
NOMENCLATURE A,
C, C. l~(t) I,j( t )
M M, M.: Qi-ill Qt.~t I Rr, R,
Rr Sit~ S'<tp "|"
"l"
1"., Till T.,II i ['i
I!, ['I (:ix it, ( ": ix ),t~
(~
Area of the solar absorber surface of the tank (m:) Specific heat of tank material (i/kg "C) Swcific heal of water (J/kg "C) Solar beam radiation at time t (W/m:) Solar diffu.~ radiation at time t (W/m') Number of grid lines along width of concrete slab Mass of the tank (kg) Mass of the water in the tank (kg) Retrieved heat flux of heater at time t (W/m: I "l',ual heat loss from the system at time t (W/m:) Tilt factor (or beam r',Kliatnon Tilt lacl()r for diffu.~ radiation Tih factor fi)r reflected radiation Solar irradiatmn at time t (W/m:) Solar irr',gtiation reaching the absorber at time t (W/m:) l"nn~ duration between two successive observations (sect Average ambient air temperature ("C) Average water temperature in the tank CC) lnntnal water temperature in the tank ('~) Ambient air temperature at time t r W:iter temperature at time t CCI Bottom heat hiss coefficient W/m-"C ()~erall heal loss coefficient (W/m: K) Side heat loss coefficient (W/m-" K) Top heat loss coefficient (W/m-" K) Tran.~mittance-absorptance product for beam radiation "rransminancc-absorptance product for ground diffuse radiation "I'ransmlnancc-absorptance pr(~luct for sky dill'use radiation
REFERENCES
209
Kaushika, N.D. and Reddy, K.S. (I 998). Thermal design and field experiment of transparesnt honeycomb insulated integratedcollector-storage solar water heater. Applied Thermal Engineering. Vol.19, pp. 145-161. Kaushika, N.D. and Arulanantham, M.(1996). Transminanceabsorptancc product of Solar glazing with transparent insulation matenals. Solar Energy Mat. and Solur Cell.~. 44, 383. Kaushika. N.D. and Avanti. P. 11996). Temperature distribution in the ground ICS system with TIM. In Proceedings of National Solar Energy Convention, 96, 80. Kaushika, N.D. and Sharma, P.P. (1994). Transparent honeycomb insulated solar thermal systems for energy conservation. Heat Recovery Systems & CHP, 14. I. 37. Kaushika, N.D., Sharma, P.K. and Padmmapriya, R.(1992). Solar thermal analysis of honeycomb r(~f cover system for energy conservation in an air-conditioned building. Energy & Building. 18. 45-49.
Kaushika. N.D., Ray. R.A.. and Padmaprnya, R.~i990). A honeycomb ~ l a r collector and storage system. Energy Corn's. Magml. 30, 127. Kaushika, N.D. and Banerjce. M.B. ( 1983 ). Honeycomb solar pond; evaluation of applications, hi Proceedings of ISES Solar World Congress, Perth, Australia, 246. Kaushika, N.D., Banerjee, M.B. and Yojana Kant.(1983). Honeycomb solar pond collector storage system. Ener~tv. 8.883. Lin E.I.H. (1982). A saltlcss solar pond. in Prr~'eedings of
ISES(Amerwan section of ISES).225. Arulanantham. M. and Kaushika. N.D. (1996). Coupled radiative and condu, thermal transfers across transparent honeycomb insulation lli.,;Cllals. AI,plied Thermal l='ngg. 16, 3,209.
Onabasi, U.. Dyksterhuis, F.H. and Kaushika, N.D. (1983). Honeycomb stabilised sahlcss solar pond. Solar Energy. 31,229.
Avann. P., Arulananiham. M. and Kaushika, N.D. (1996). Solar lhcnnai analysis of ground integrated collector/storage system with transparent insulation. Applied Thermal Engg. 16, ! !. 863.
Prakash, J., Kaushika, S.C., Kumar, R. and Garg, H.P. (1994). Performance prediction for a triangular built-in-storage-type solar water heater with transparent insulation. Energy. 19, 8. 869.
Duffle. J A. :rod Beckman, W.A. (1991). Solar Engineering of Tiu,ruud Proc'es.,~c's. 2n ~ edn, pp. 54-59. Wiley inierss
Reddy, K.S., Avanti, P. and Kaushika, N.D. [1998). Finite-time thermal analysis of ground integrated-collector-storage solar water heater with transparent insulation cover. Inlernalional Journal of Energy Research. in Press.
New York Gucizh
Goct/,berger. A. 11984). Seasonal storage of thermal energy with radnatnvclv he;ucd storage walls. Int. J .Solar Energy. 2. 52 l. Gordun. J.M. (! 997) l~wv heal loss double glazed windows. Energy 12. 1333-1336
Rommel, M.V., Winwar, V. and Goctzbcrger. A. (1987). Thermal energy collection and storage with radiatively heated walls. In Proceedings of Advances in Solar Energy Technology (eds. W.H.Bloss and F. Pfisterer) ISES. Hamburg. 2. Scheatcr. R. and L~)wrcy. P. (1992). The optimum design of honeycomb .~(llar ponds and a comparn.,~)n with salt gradient solar ponds. Solar [~wrg.v. 48, 69.
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Schmidi[. Ch. and Goetzbcrgcr, A. (1990). Single-tube intcgratcd collector storage systems with transparent insulation and involute reflector. Solar Energy. 45.2, 93. Schmidt. Ch., Goe(~hcrgcr. A. and Schmidt, J. (1988). Test rcsult.~ and evaluation of in(cgratcd toilet(or storage system.,, with transparcnl insulation. S~)lar Eners 41,5, 487.
Sharma. M.S. and Kaushika. N.D.c 1987). Design and performance charactenstics of honeycomb sohu"pond. E~zergyCmzvers.Mgmt. 32. 345.
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THERMODYNAMIC STUDY OF A REGENERATIVE WATER DISTILLER
Gustavo Koury Costa and Naum Fraidenraich Depto de Energia Nuclear, Universidade Federal de PE, Av. Prof. Luiz Freire, 1000,Recife, PE, 50740-540, Brazil,+55 (801 2718252), Fax +55 (081 271-8250), E-mail: 2 [email protected] / [email protected] A b s t r a c t - The need of water for human consumption in arid regions where there is a great shortage of non saline sources has motivated the study of new desalination technologies. The present work introduces a water distiller whose main characteristic in relation to other units is the high production allied to a good efficiency. This equipment that promotes the continuous reuse of the water condensation heat is then called Regenerative Distiller. A theoretical study of the thermodynamic process involved is made, resulting in a numeric model for the project of new units accordingly to the required demands. The theoretical results are also compared with experimental data obtained through the test of a built prototype put into operation by the FAE/DEN group of the Federal University of Pernambuco.
1.1NTRODUCTION Equipment for water desalination has been widely studied all over the years. Such facilities can be u s e d , for example, for providing water for human and animal consumption in places where water is present with a considerable degree of salinity. The north-eastern part of Brazil is a region whose underground water sheets are quite saline. Amongst the available processes, distillation occupies a traditional position due to its relative simplicity, its versatility in terms of production capacity as well as the easiness for obtaining energy in the form of heat. The study of distillers finds, therefore, fundamental importance even in face of other modem processes like Reverse Osmosis. Due to the relative simplicity of construction and to the low operation temperatures, atmospheric distillers (i.e., distillers which operate at atmospheric pressure) have been widely studied [Baumgartner et al. (1991), Fraidenraich et al. (1995), Heschl, O. and Sizmann, R. (1987) and Rom,Sn, R., C. et al. (1993)]. The low operation temperature (maximum of 90~ makes it possible for renewable energy sources (like solar energy, firewood, bagasse and bio-gas) to be used. In chronological order of appearance, one can mention the solar still, multi-stage distiller and the regenerative distiller. The multi-stage and the regenerative distiller can work, either with solar energy or other thermal energy sources, unlike the solar still which operates exclusively with direct incidence of solar light. Also, in the multistage and regenerative distillers the condensation heat is recovered in order to improve the equipment performance, whereas in the solar still is lost to atmosphere. This work develops a mathematical model of a type of regenerative distiller with vertical condensers and evaporators placed parallel to each other. Closed form analytical expressions for the main distiller parameters are obtained and numerical results for the equipment performance are then compared with experimental data (Figueiredo, 1997). 2. OPERATING PRINCIPLE OF THE REGENERATIVE DISTILLER Figure 1, helps understanding the working principle of the distiller.
Fig. 1. Schematic representation of the regenerative distiller. Water temperatures are also indicated. Saline water enters the condensers at room temperature (Tci) and is warrrted as it goes up to the heater (Teo). Temperature is then risen up to its nominal (working) range (Tei). As water flows down through the cotton sheets (evaporator), it evaporates, condensing on the condenser's surface, opposite. Distilled water is then collected on the side gutters and the rest of the water from the sheets is collected on a tray from where it can be pumped up to a storage tank. Water temperature at the evaporator output is Teo. The distiller is placed inside a wooden container. Thermal insulation surrounding the walls and a reflective film adhered to its internal surface keeps the distiller heat losses at reasonable levels. 3. ANALYSIS OF THE DISTILLER PERFORMANCE 3.1 Energy ej~ciency
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The energy efficiency of a distiller c a n be measured by the Gained Output Ratio (GOR), defined by the following equation (Baumgarmer et al., 1991):
mdc+ mdw .~,
(4)
GOR = me .Cpw.(Teo - Tci ) + QL
GOR =
(1)
mdt.2,
If mass and heat losses are negligible, the ideal GOR, from Eq. (4), can be written as
me.Cpw.(Tei - T c o )
where
GOR) i= .
thd~.2
(5)
me.C, .(r o
me mass flow rate of saline water (saline water input). thdt mass flow rate of distilled water (distilled water output).
It is worth noting that, if the heat losses QL are negligible,
t latent heat of vaporization (600kcal/kg). Cpw specific heat of water (liquid).
It's important to note that malt in Eq.(1) takes into account all the distilled water produced by the distiller. This will include the condensed vapour on the inside walls which will not contribute effectively to heat regeneration. In fact, the use of the GOR for the distiller efficiency measurement may raise some controversies regarding its thermodynamic meaning. A thorough analysis may be found at Figueiredo Melo (1997) and Costa (1998). Thus, Eq. (1) may be written as follows,
"
mdc+ ma~
~.I (2)
GOR = me X:pw.(Tei - Tco )
where malt has been divided into m dc , which stands for the water condensing on the condenser and mdw, which represents the water collected on the distiller's insulating walls. We can notice that only mdc effectively contributes for energy regeneration. Another modification can be made in Eq. (2) so that temperature of the incoming water flux, Tci, may appear instead of Tco which is far more difficult to measure. An overall energy balance gives:
me.Cpw.(Tei - Teo ) = me.Cpw.(Tco - Tci ) + QL
(3)
Eq. (3) shows that the temperature difference, At, between evaporator and condenser is uniform along the evaporator length. In the next section, we describe the heat and mass transfer processes in the distillation chamber assuming that no such losses exist. Once the solutions for this model are obtained, we incorporate real effects in order to reproduce experimental results as closely as possible. 4. MODELLING THE DISTILLATION CHAMBER The operational parameters of the distiller, defined experimentally, are the temperature of the incoming water flux Tci, the mayJanum operating temperature, Tei and the saline water flow. To determine the gained output ratio (Eq.(5)), the distilled water production (at the condenser and walls) and the evaporator output temperature, Teo, must be obtained either experimentally or theoretically. To model the distilling chamber, one dimensional heat and mass balance equations has been set up for an infinitesimal element of volume, taken between evaporator and condenser (Costa, 1998). The energy balance is shown in Fig. 2, where, for simplicity, only one condenser facing one evaporating sheet is represented. We also made the assumption that heat and mass exchange between each evaporator and condenser is not influenced by the presence of the other sheets (Costa, 1998), so that the process may be treated as shown in the figure.
Y1
I
mi "(y).c~ti(y)
lkb.dy.Ahd~
,yx.exzxyx:~o-x.,~
Condenser
m ~"(y +dy). ce, h (y+ dy)
where QL represents the energy losses through the insulating walls.
~__ Evaporationsheet
Equations (3) and (2) allow us to write a practical expression for the GOR, Fig. 2. An energy balance in an infinitesimal area of the distillation chamber.
213
ISES Solar World Congress 1999, Volume III
Considering a sheet of width b and a global heat transfer coefficient ht 1, we can write for the energy balance:
r n ( y ) ~ p w J ( y ) - rn(y + dy)~CpwJ(y + dy)
Tei 9
"
(10)
he(614.8-t )
reo = h t.dy:b.At
(6)
We, then make use of equations (9) and (10) to obtain the distilled water production on the condensers ( m dc ):
for a distiller with n evaporation sheets 9A similar procedure can be taken for the mass balance. This will result in Eq.(7) below:
9
~ n ( y ) - rn(y + dy) _ heJg.At.dy
mdc _ A
(7)
m
n
dt
mo =me.
h v
[
Teo - Tci 614.8- Tei q Tei 614.8 - Teo dt (614.8- Tei )h he(614.8-t) 2 Teo
where h v is the enthalpy difference between the liquid and the
(11)
saturated vapour.
Equations (9) and (11) allow us to determine the ideal GOR (Eq.(5))
Energy and mass balance equations (6) and (7) yield a system of two differential equations in two unknown functions, that is, the temperature t(y), along the evaporator sheet and the mass flow of saline water, rh(y). Initial conditions for water flow
( 614.8 - Tei GOR) i = Cpw.(Teo - Tci) ~1 - 614.8- Teo
(12)
and temperature are rh(0)= rheand t(O) = T e i . Integrating the equations all over the cotton sheets, we obtain an implicit relation for the output evaporator temperature, Teo and for the
It's important to remember that equation (12) has been obtained under the consideration that there are no heat losses. This allowed us to make the temperature difference (Teo - Td )= (Tel - T~o ) constant. Table 1, below, gives
output water flow as a function of the evaporator length, L. The difference between the incoming and the outgoing mass flow rate, gives the distilled water at the condenser walls 9
L=me.h...........~ h v h t _ T e i nJ~.At he
!
dt h ht reo he -----P-~- - Cpw he
some values of me , A
t
,~2
mdc , thdc A th e
hypothetical value of ( T e o - Tci)= 26~
(8)
Tel equal to 90~
and G O R ) i
for an
and a temperature
The values are obtained as a function of
reo.
The relation between ht and he remains practically constant over the temperature range in which the distiller operates. Average
Table 1 Values obtained from the use of equations (9), (11) and (12) as a function of Teo.
m
values of h v (h v ) and Cpw have been used as well. With
I
these considerations, equation (8) turns into: GOR)
me = A
Te~ - Tc~ Tei dt (614.8 - Tei )h "o he(614.8-t) 2
(9) 36 40
Teo
where A represents the total evaporating area. Developing the energy and mass balance equations (Eq.(6) and (7)) in another way and making the same considerations we made to obtain equation (9), it is possible to determine the outgoing mass flux:
" _
0,093 0,087
5,59 3,61
0,28
0,081
2,56
0,36
0,074
1,90
6,30
0,43
0,068
1,46 !
1,44 2,39
0,13 0,21
44
3,50
48
4,79
52 56
8,11
0,49
0,061
1,13
60
10,31
0,56
0,054
0,88
64
13,10
0,62
0,69
5. HEAT LOSSES 1
The global heat transfer coefficient takes into account the contribution of convection (he) and evaporation (he). Radiation will not be considered for its contribution is not significant at this temperature range.
The assumption of a constant temperature difference between evaporator and condenser is essential for solving equations (6) and (7). Nevertheless, losses do occur in the distiller and one
ISES Solar World Congress 1999, Volume III
214
must account for them. The method used has been to correct equation (1) which, in turn, will yield equation (4), as shown before. The term thdc in Eq.(4) can be calculated by Eq.(11) and the heat transferred through the walls, QL, can be estimated as follows:
QL
=
Tei-Tenv
Ruw
+
Te~ +leo ~ - T 2
Rsw
env
+
Teo_ Teta,
that the ideal GOR can also be written as (Fraidenraich et al., 1993)
he/Cpw GOR)i = ~ the l A
(15)
in such a way that the single operational parameter available is (the / A ). If this ratio decreases the GOR increases, but at the (13)
Rlw
where Te,,v is the temperature outside the equipment, Ruw,
same time the amount of distilled water goes to zero with ( the / A ), as will be shown in the next section. 6. COMPARISON WITH EXPERIMENTAL RESULTS
Rsw and Rlw account for the thermal resistance of the upper wall, side walls and the lower wall, respectively. Once the heat transferred, QL, is determined, one can then estimate the amount of condensing vapour on the walls by: mdw =
QL he .
2
ht
(14)
It is important to note that an attempt has been made to correct equation (1) and still make use of the very simplifying supposition of a constant temperature difference between the evaporator and the condenser. One could also try to solve Eq.(1) directly but the question of how this temperature difference would behave might still not be answered. So, one starts from an ideal process and make the proper adjustments in the end. This is an approach which has been widely used in many branches of Engineering and has yielded quite reasonable results most of the times.
A prototype of such distiller was built and put into operation by the FAE/DEN Group of the Federal University of Pernambuco. Data were collected and a wide assort of tests of this equipment can be found at Figueiredo Melo (1997). The prototype was tested with one, five and eleven cotton sheets, each sheet with an evaporating area of 0.6mz. The whole evaporation set was thermally coupled to two side condensers, made of copper. The temperature of the water Ta was kept around 900C with the aid of an electrical heater and a temperature controller. Graphs of the GOR as well as the distilled water production ( thdc / A ) versus (th e / A ) could then be plotted and the results can be seen in figure 4.
Fig. 3. A comparison between GOR and GOR)i plotted as functions of the saline water mass flow rate. Figure 3, above, compares the GOR, calculated by Eq. (12) with the one given by Eq.(4). One observes that the effect of the losses term in equation (4) is to bring the curve to a point of maximum, instead of the growing asymptotic behaviour, in the ideal case. The fact that the GOR increases when the ratio ( the / A ) decreases, deserves some comments. It can be shown
Fig. 4. A comparison between experimentally measured GOR and theoretical values, for two ranges of the parameter ( the / A ) : a) 0 to 4 and b) 0 to 21.
Theoretical values are plotted in full lines alongside with the experimental data. It's plain to observe, however, that the
ISES Solar World Congress 1999, Volume III
numeric model fairly represents the real situation, despite the simplifications which were made. Another observation is appropriate here. Fig 4.b shows different dot styles for the experimental data. This is so, because those data were obtained from different experiments. Round dots were obtained from tests on the 11 cotton sheet distiller. Square dots came from the 5 sheet distiller and the remainders from the one sheet distiller. Figure 4.a shows data from the 11 sheet distiller. The continuity of experimental data from different distillers reveals that, to a considerable extent, the equipment won't be affected by the number of evaporator screens, being strongly influenced by the ratio ( t h e / A ) instead. This was an assumption made to formulate our model which, in a certain way, was confirmed by the distiller's tests. The distilled water production, plotted against ( t h e / A ) is shown in figure 5. Again, concordance between the experimental data and theory is observed. It's important to note, however, that the production of water tends to stabilise as the feeding water flow grows bigger. This calls our attention to the fact that a maximum GOR will not correspond to the greatest production of distilled water. Therefore, one has to balance the energetic efficiency against the non-saline water output.
215
The efficiency of this distiller is thoroughly dependent on its feeding water flow per unit area ( me/A ), which gives us a clue as how to vary the number of evaporator screens and control the flow rate me to adjust the parameters according to the results wanted. Anyhow, the good concordance between theory and experiment gives us confidence in the model to pre study future units. The prototype tested, with a size of 0.8 x 0.8 x 1.1 m 3 and an evaporating area of 13.3 2 , is able to produce 1.8 litres/h of distilled water operating at a GOR of 2. REFERENCES Barbosa, E. M. de S., Fraidenraich, N. and Tiba, C., (1995) Desalinizagho de ~gua: desenvolvimento e testes de um prot6tipo de destilador multiefeito, Proceedings of the XIII
Brazilian Congress and II lbero American Mechanical Engineering Congress, Belo Horizonte, MG, (in portuguese). Baumgartner, T., Jung, D., Ktssinger, F. e Sizmann R., (1991) Multi-effect Ambient Pressure Desalination with Free Circulation of Air, In proceedings of ISES 1991 Solar World Congress, Denver, Co, USA, Vol. 2, Part II. Figueiredo Melo, A. G. de, (1997) Estudo experimental de um destilador de ~gua do tipo regenerativo, M.Sc. Thesis, Nuclear Energy Department, UFPE, Recife, PE, Brazil (in portuguese). Fraidenraich, N., Costa, G. K. e Barbosa, E. M. de S., (1993) Analysis and design criteria for regenerative water distillers,
Proceedings of the XIXth Meeting of the National Association of Solar Energy, ANES, La Paz, Baja California Sud, Mtxico. Franco, J. e Saravia, L., Ensayo de un destilador atmosftrico de tipo multiefecto de tamafio familiar con calentamiento a lefia
Proceedings of the 16th Meeting of the Argentine Association of Solar Energy, ASADES, and 7th Latino-American Congress of Solar Energy, ALES, La Plata, Argentina (In spanish). Heschl, O. e Sizmann, R., (1987) Solar sea water desalination with a high efficiency multi-effect solar still, In Proceedings of ISES Solar World Congress, Hamburg. Fig. 5. A comparison between theoretical and measured distilled water production per evaporating area (mdt/A ), as a function of the saline water input per evaporating area ( me/A ).
Finally, it should be remarked that the prototype tested, with a size of 0.8 x 0.8 x 1.1 m3 and an evaporating area of 13.3 m2, is able to produce 1.8 liters/h of distilled water operating at a GOR of 2. 7. COMMENTS AND CONCLUSIONS The numeric model presented showed to be trustworthy and suitable for predicting the behaviour of other units.
Gustavo Koury Costa, (1998) Estudo termodinfimico de um destilador de ~gua do tipo regenerativo, M. Sc. Thesis, Nuclear Energy Department, UFPE, Recife, PE, Brazil (In portuguese). Roman, R., Corvalfin, R., Ponce, D. e Doria, J., (1993) Sistemas Multiefecto de Una Etapa para Purificacion de Aguas Salobres, Proceedings of the 16th Meeting of ASADES and 7th Latino-American Congress of Solar Energy, ALES, La Plata, Argentina (In spanish).
216
ISES Solar World Congress 1999, Volume III THE PERFORMANCE AND ANALYSIS OF A MULTIPLE-EFFECT SOLAR STILL UTILIZING SOLAR AND/OR WASTE THERMAL ENERGY
A.I. KUDISH*, E.G. EVSEEV*, L. HORVATH** and G. MINK'* *Solar Energy Laboratory, Department of Chemical Engineering, Ben-Gurion University of the Negev, Beer Sheva 84105, ISRAEL; Tel: +972 7 6461488, Fax: +972 7 6472916, E-mail: [email protected] **Research Laboratory of Materials and Environmental Chemistry, Chemical Research Center, Hungarian Academy of Sciences, 1525 Budapest Pf 17, HUNGARY; Tel: +36 1 325 5992, Fax: +36 1 335 7892, E-mail: [email protected] Abstract - The performance of an air-blown, multiple-effect solar still designed to recycle the thermal energy of condensation has been studied in three operational modes, viz., driving forces, utilizing: (i) solar energy; (ii) waste thermal energy and (iii.) both solar and waste thermal energy. The still glazing was a double-walled polycarbonate sheet, a transparent insulation material (TIM) and the still area was 1 m2. A solar simulator providing a constant irradiation intensity of 650"~_10 Wm2 was used, which facilitated the inter-comparison of the system performance under different conditions. The waste thermal energy was simulated using a feedstock reservoir maintained between 86 and 90~ As a result of the low thermal mass of the still, steady-state temperatures and yields were achieved within one hour after start-up. The performance of the solar still, operating under constant energy input (i.e., a constant irradiation intensity) was determined mainly by the flow rate of the air stream, which functions both as a mass and energy carrier. The optimum range of the air flow rates, under all modes of operation, was determined experimentally. Mass and heat balances utilizing experimental results and referring to optimum operating conditions were also performed. 1. INTRODUCTION The utilization of solar energy for the distillation of brackish or saline water has been practiced for a very long time. Various types of solar stills and solar-assisted desalination units have been designed and investigated. A number of manuscripts have been published on this subject, which include a classic one by Talbert, et al. (1970), Malik, et al. (1982) and Kudish (1990). In arid zones, solar distillation can be an ideal source to produce fresh water from saline water, for both human consumption and agriculture. The main disadvantage of the solar stills presently available is that their productivity per unit area is relatively low. The fixed capital investment cost of a solar desalination plant is roughly proportional to the still area; consequently, increasing the productivity per unit area by recycling the thermal energy of condensation of the distillate can be of paramount importance. The performance of an air-blown, multiple-effect solar still consisting of an upper evaporation chamber and a lower condensation chamber has been analyzed and reported in detail by Kudish, et al (1997) and Mink, et al. (1997, 1998). These analyses suggested that it would be possible to utilize low grade waste thermal energy, when available at the site, as the driving force in the distillation process. This would allow the still to operate 24 hours a day by utilizing solar energy and/or waste thermal energy during the daytime and waste thermal energy during the night, i.e., nocturnal distillation. In the present paper we shall report on the experimental results obtained when operating the still under three different operation modes, i.e., driving forces: (i.) only solar energy, (ii) waste thermal energy and (iii.) both solar and waste thermal energy. 2. EXPERIMENTAL Experimental setup The solar still prototype under investigation is shown schematically in Figs. 1.a and 1.b. It is essentially of the tilted-wick genre in the form of a thin rectangular box divided into two chambers (upper evaporator and lower condenser) by a central metal sheet. The central metal
sheet does not extend across the full length of the still but leaves a slot of 10 mm between its top end and the still's upper extremity. The metal sheet also functions as (i) the support for the wick (a black porous material), which covers it on the upper chamber side; (ii) the surface to which a serpentine tube is in contact with in the lower chamber side. This serpentine tube serves as a conduit to transport the feedstock to the upper chamber and also functions as a heat exchanger for preheating the feedstock prior to entering the upper chamber. The spacing between both the upper chamber still glazing and lower chamber backside and the central metal plate is 12 mm.
Mode of operation when using solar energy The mode of operation of the solar still is as follows: (1) ambient air is pumped into the upper chamber at the bottom of the tilted still and sweeps the water vapor evaporated from the tilted wick into the lower chamber via the slot at the top of the tilted still. The maximum temperature of the air stream is measured at this point, above the slot, prior to reversing direction and entering to the lower chamber. In the lower chamber it serves as the hot fluid in what is essentially an air-liquid heat exchanger. (2.) the major portion of the water vapor in the air stream entering the lower chamber condenses either on the backside of the metal sheet supporting the wick or on the serpentine tube. The latter enters at the bottom of the lower chamber and transports the feedstock to the upper chamber. (3.) the feedstock, which enters the serpentine tube at a flow rate in excess of the rate of evaporation from the wick, is preheated during its passage through the serpentine tube and exits at the top edge of the central plate. It then passes over a weir and flows by gravity down the wick. (4.) the distillate and humid air stream exit the lower chamber and enter a gasliquid separator. This distillate, defined as primary, is collected, whereas the humid air stream enters an external heat exchanger in order to recover the water vapor remaining in the air stream, defined as secondary distillate, prior to venting to the ambient.
ISES Solar World Congress 1999, Volume III
217
Fig. 1.a. Schematic diagram of the experimental unit. The temperature probes are indicated by numbers 0 - 15.
Fig. 1.b Top view of the lower chamber with serpentine tube and baffles (not drawn to scale). In the upper chamber the air stream flows countercurrently to the feedstock flowing down the wick by gravity, whereas in the lower chamber it flow countercurrently to the direction of the feedstock flowing upward through the serpentine tube (of., Fig. 1.b). Due to the nature of this solar still, viz., that the upper chamber glazing does not function as a condensation surface and, in fact, any such condensation is detrimental to still performance, a double glazing is used to reduce thermal energy losses via the glazing to the ambient. We have used a solar grade, double-walled, 10 mm thick polycarbonate sheet was used as the solar still glazing.
It is a non-wetting polymeric transparent insulating material (TIM). It has been reported previously (Kudish, et al. (1997) and Mink, et al. (1997, 1998)), that a relatively large fraction of the thermal energy of condensation of the process was successfully recycled to both preheat the feedstock and heat the backside of the evaporation plate, which separates the two chambers. Consequently, a two- to three-fold increase in distillate yield was achieved relative to that reported for conventional type solar stills.
ISES Solar World Congress 1999, Volume III
218
Mode of operation when using solar energT and/or waste thermal energy The waste thermal energy was simulated by means of a conventional heater that maintained the feedstock reservoir at a temperature between to 86 and 90~ In this mode of operation, since the feedstock was preheated externally, it entered the still at the upper part of the evaporation chamber (of., Fig. 1.a); i.e., the feedstock did not enter via the serpentine tube. Consequently, the thermal energy recycle from the lower chamber to the upper chamber pronly via a single process, viz., condensation on the backside of the central metal plate.
The absolute humidity of air in the vicinity of the slot was calculated from the material and energy balances on the still. A data acquisition system served to monitor and store the temperature data from the sixteen thermistors and to calculate differential and cumulative yields at variable time intervals. It consisted of a PC with an A/DD/A converter card, electronic measuring and magnetic valve control unit, temperature sensors and a digital balance with a RS232C serial interface. The data acquisition, control and analysis s o ~ e was developed specifically for this study. The parameters under investigation were the feedstock and the air flow rates.
Experimental conditions and procedure The tilt angle of the solar still module was set at 20 ~ throughout this study. Similarly, the solar simulator provided a constant radiation intensity of 650-&-_10 Wm z and its tilt angle was also 20 ~. The differential and cumulative yields from Separator I and II were measured automatically by a type PT 6 Satorius electric balance with an accuracy of • g. The temperatures were measured at 16 strategically positioned locations, of., Fig. 1.a, with an accuracy of • ~ using calibrated temperature sensors of the silicon base type KTY 11-2A.
3. RESULTS AND DISCUSSION Solar radiation driving force As a result of the low thermal mass of the solar still, steady-state temperatures and yields were achieved within one hour of start-up in all experimental runs. All results reported in this study refer to steady-state operating conditions and the mass and energy flows were normalized to a 1 m z still area.
Table 1. The temperature profile of the solar still at steady-state as a function of air flow rate. The position of the thermistors are shown in Fig. 1.a (numbered 0 to 15). Feedstock flow rate = 2.96 kgm2hl; irradiation = 650-k10 Wm 2 and ambient temperature = 25.8 -26.6~
UPPER CHAMBER
Air Flow Rate
T1
T2
(kgm'2h-1) (oc) (~
LOWER CHAMBER
T~o~
T~j
Tx~
~X,o~
T~
(~
(oc) (~
%
T8
(~
(~
(~
Cc) Cc)
Tll Cc)
T12 T13 Cc) Cc)
T14 Cc)
T15 (~
96.9
97.4
96.9
94.9
52.5
48.9
34.0
30.4
24.1
62.4
"I'3
T4
"1'5
"1'6
'1'9
T10
0.32
68.6
95.1
96.3
96.2
97.0
0.58
67.6
91.9
93.6
93.6
94.4
94.2
94.5
94.0
91.7
68.2
68.2
60.7
57.1
21.9
61.9
1.02
64.9
87.4
89.8
89.8
90.7
90.7
90.7
90.0
87.2
69.9
69.9
65.6
64.1
28.6
60.4
1.24
62.6
84.6
87.4
87.6
88.4
88.5
88.4
87.5
84.4
67.5
67.5
64.1
63.0
32.1
58.4
1.76
59.1
80.1
83.7
83.7
84.7
84.8
84.3
83.5
79.9
64.4
64.4
62.0
61.3
34.1
55.7
2.21
57.5
77.1
81.3
81.2
82.1
82.1
81.6
80.7
77.0
63.1
63.1
60.9
60.4
35.0
54.5
75.9
75.1
71.1
59.6
59.6
58.1
57.8
36.0
51.5
70.6
69.9
66.0
56.0
56.0
54.9
54.6
35.4
48.7
3.27
54.0
71.1
76.4
75.9
76.6
76.7
4.51
51.2
65.6
72.2
71.3
71.6
71.3
The temperature profile for the still operating under a solar energy only driving force as a function of the the air flow rates, expressed as kg bone dry air per m 2 still area per hour, is given in Table 1. The solar still productivity, as a function of air flow rate in the range from 0.32 to 4.51 kgrn2h 1 is reported both in Table 2 and in Fig. 2 in terms of the primary (I), secondary (II), and total (~) distillation rate. The primary distillation rate refers to that condensed within the lower chamber during the thermal energy recycle process. The secondary distillation rate is that obtained by passing the humid air stream exiting the solar still
through an external heat exchanger prior to venting to the ambient. It is apparent from Table 2 that with regard to still productivity there exists an optimum range for the air flow rate, approximately between 1 and 3 kgm'2h"1, for the system under investigation. It is also observed that the ratio of secondary to primary product increases with increasing flow rate, i.e., the primary decreases and the secondary increases with increasing air flow rate (ef., Fig. 2). The reason for this optimum air flow rate range has been discussed in our previous papers (Kudish et al (1997) and Mink et al. (1997, 1998)).
219
ISES Solar World Congress 1999, Volume III
Table 2. Still productivity (I- primary; II- secondary; E- total distillate) as a function of air flow rate. Feedstock flow rate = 2.96 kgrn-2h-1; irradiation = 650-kl 0 Wm -2.
II
Z
Air Flow Rate (kgrn-2h-1)
(kgm-2h-1)
(kgm'2h-1)
(k~-~h -')
0.32
0.68
0.01
0.69
0.58
0.93
0.04
0.97
1.02
0.97
0.13
1.10
1.24
0.96
0.15
1.11
1.76
0.87
0.20
1.07
2.21
0.84
0.27
1.11
3.27
0.75
0.29
1.03
4.51
0.65
0.32
0.97
I
1.20 ,.;. t--
"
100 -..--.----~
/inmm_=m_/
E 0.80 ( A
-~ 0.60 "~
2
0.40
~____
0.20
~
~
~-.-----~
= II
m..lm,~~-,, ~
0.00 ' 0.00 0.50 1.00 1.50 2.00 2.50 3.00 3.50 4.00 4.50 5.00 Air Flow Rate (kgm2h ~) Fig. 2. Productivity in terms of primary, secondary and total distillation rate as a function of air flow rate. Feedstock flow rate = 2.96 kgm2hl; irradiation = 650-kl 0 Wrn2.
Waste thermal energy driving force-exposed glazing In this mode of operation the feedstock flow rate was maintained at 9.3 kgm'2h-1. The feedstock was preheated between 86 and 90 ~ by a heat exchanger, in order to simulate an external source of waste thermal energy. As mentioned previously, the preheated feedstock entered the upper chamber at the top of the central metal plate, of., Fig. 1.a and did not pass through the serpentine tube. The experimental results are summarized in Fig. 3 and in Tables 3 and 4. The waste thermal energy utilized, qwas~, which is defined as the heat released by the preheated feedstock in the upper chamber, q ~ : d , is reported in Table 4. This term is defined as the difference between the thermal energy contained by the feedstock entering the upper chamber q ~ and that of brine exiting the upper chamber qbme, qwas~ = q ~ s , ~ = qty, - q ~ e = c~mfTf~ - c ~ ~ T b m e , (1) where ~ was determined form a material balance, viz., as a difference of the feedstock flow rate and the rate of evaporation in the upper chamber. It is observed that increasing the air flow rate results an increase in both productivity and utilized waste energy, qw~t~. Also, it is apparent from the results that at a reasonable air flow rates (i.e., around 3 kgm2h 1, where the parasitic energy requirement of the air pump is still relatively low) a total productivity rate of about 0.7 kgh 1 was obtained.
ISES Solar World Congress 1999, Volume III
220
Table 3. Temperature profiles at steady-state as a function of air flow rate when utilizing only waste thermal energy for preheating the feedstock. Ambient temperature = 23.6 -26.0~ feedstock flow rate = 9.3 kgrn2h l and feedstock inlet temperature = 86 - 90~
Air Flow Rate
UPPER CHAMBER i
T2 (~
% (~
49.7
65.7
44.6
2.02
Tl,out Tsopj Tyon Tx,out Tbrinv
LOWER CHAMBER
(~
T5 (~
T6 (~
T7 (~
T8 (~
% (~
T10 (~
69.0
73.7
84.7
80.0
74.9
70.2
65.0
64.5
67.4
71.8
83.3
77.9
73.2
68.8
42.4
62.9
65.7
69.9
82.3
76.2
71.6
3.12
'~ 37.0
58.3
61.2
65.7
80.7
72.6
4.11
35.0
54.6
57.8
62.3
78.3
4.90
34.4
53.5
56.5
60.9
7.07
31.1
49.1
51.9
56.7
T1
(kgm2h "l) (~ 1.20
T4
(~
T12 (~
T13 (~
T14 (oc)
T15 (~c)
55.9
57.5
51.9
49.8
28.3
54.5
63.8
54.5
56.4
52.4
50.9
28.9
52.6
22.7
62.6
53.6
56.0
52.7
51.2
27.8
51.5
67.5
19.1
58.6
50.1
52.2
50.3
50.0
29.5
46.7
70.2
64.3
16.7
56.1
48.7
50.7
49.1
48.5
30.6
45.3
78.1
69.0
62.9
15.4
54.8
48.1
50.0
48.6
47.9
30.7
43.3
77.7
66.3
59.0
11.4
50.9
45.7
46.6
45.5
44.9
31.0
41.4
Tll
||
1.56 ||
Table 4. Productivity rates and calculated waste thermal energy input, qw~,o, at steady-state as a function of air flow rate when utilizing only waste thermal energy. Ambient temperature = 23.6 - 26.0~ feedstock flow rate = 9.3 kgm'2h"1 and feedstock inlet temperature = 86 - 90~
Air Fl~ (kgm'2h'l)
I
lrradiati~ (Win-2)
qw~ (Wm"2)
I
(kgm2h'l) ....
[
n
I
z
(kgm'2h'l) . . . . (kgm'2h'l)
1.20
.....6
356
0.34
0.05
0.39
1.56
0
380
0.39
0.08
0.48
2.02
0
395
0.41
0.12
0.53
3.12
0
452
0.52
0.17
0.69
4.11
0
468
0.53
0.19
0.72
4.90
0
490
0.55
0.22
0.77
7.07
0
512
0.58
0.25
0.82
Waste thermal energy driving force & insulated glazing During the nocturnal distillation mode of operation, it seemed to be logical to further reduce the top losses via the double-glazing by covering it with an insulating material. A 5 mm thick polyurethane foam with a reflective aluminum foil on one side was placed upon the still glazing. The results of these experiments are
reported in Fig. 4 and in Tables 5 and 6. It is observed that the use of this additional insulation on the still glazing resulted in both higher operating temperatures and productivity rates at comparable air flow rates. However, the increase in the total productivity was only in the range of 0.1 kgm2h 1.
ISES Solar World Congress 1999, Volume Ill 1.00
221
r
0.90
, ~ t a
0.80
,v.|
r-
/,.,
0.70
E
,~,0 ~
T-'
"
0.60
o) v
0.50
. m
>
0.40
. B
O "O
0.30
Q.
0.20
2
~'J II
,..~!~"--"-"
0.10 0.00 0.00
1.00
2.00
3.00
4.00
5.00
Air fl0w rate (kg
m2
6.00
7.00
8.00
h~)
Fig. 3. Productivity as a function of air flow rate when utilizing only waste thermal energy. Ambient temperature = 23.6 26.0~ feedstock flow rate = 9.3 kgm2h 1 and feedstock inlet temperature = 86 - 90~
Table 5. Temperature profiles at steady-state as a function of air flow rate when utilizing only waste thermal 23.7 - 25.9~ feedstock flow rate = 9.3 kgm-:h 1 and feedstock inlet temperature = 86 - 90~ UPPER CHAMBER
Air Flow
LOWER CHAMBER
Tt,o=
T~,I
Tx~
Tx.o~
T~i~e
Tll
T12
T13
T14
Tls
||
Rate
T~
% (~
(~
(oc)
(oc) (oc) (oc) (oc) (~
81.4
78.3
29.6
70.7
61.0
63.4
58.0
55.8
25.5
59.3
84.3
79.4
76.3
73.2
68.7
58.6
60.8
57.1
55.7
30.0
56.2
73.5
83.1
77.7
74.5
26.4
66.8
57.5
59.8
57.0
55.6
30.2
54.4
64.8
68.5
81.1
74.3
70.0
21.9
61.7
53.0
55.2
53.2
52.4
30.6
48.4
57.9
61.4
65.3
79.4
72.0
66.9
19.5
59.3
51.7
53.9
52.3
51.5
32.9
47.8
36.7
56.8
60.1
64.1
78.6
71.2
65.8
18.3
58.0
51.3
53.0
51.7
51.0
32.8
45.7
32.0
50.6
53.9
58.2
77.3
67.0
60.3
12.8
52.7
47.4
48.5
47.4
46.7
33.2
42.7
TI
(kgm2h "1)
T2
T3
T4
Ts
(~
(~
(oc) (~
(oc) (~
56.3
71.9
74.2
77.6
85.0
48.6
69.8
72.4
75.5
45.7
67.6
70.3
38.4
61.8
37.2
Ts
TlO
T9
("c)
w,
1.05 ||
1.55 2.00 M
3.15 ||
4.11 4.58 ii
6.99
Table 6. Productivity rates and calculated waste energy input, qwasto, at steady-state as a function of air flow rate when utilizing only waste thermal energy for preheating the feedstock and placing an insulating cover on the still glazing. Ambient temperature = 23.7 - 25.9~ feedstock flow rate = 9.3 kgm2h -1 and feedstock inlet temperature = 86 - 90~ Air Flow Rate (kgm'2h'l)
~ ~
Irradiation
1.05
l[
II
qwaste (win -2)
(kgm'2hd)
i
1
(kgm-2h-1)
(k~-:h -1)
0
308
0.36
[
0.07
0.43
1.55
0
348
0.46
0.13
0.55
2.00
0
375
0.51
0.16
0.67
3.15
0
442
0.64
0.17
0.81
4.11
0
451
0.61
0.24
0.85
4.58
463
0.60
0.26
0.87
6.99
504
0.64
0.27
0.91
(Wm'2)
Z
222
ISES Solar World Congress 1999, Volume III
Hybrid mode of operation, simultaneous use of both solar and waste thermal energy as the driving forces In these experiments a constant feedstock flow rate of 5.7 kgm-2hl was used. The feedstock was preheated in the range of 86 to 90~ by the external heat exchanger to simulate the waste thermal energy prior to entering at the top of the central metal plate. The results are summarized in Tables 7 and 8 and in Fig. 5. It is observed that in this mode of operation, for an air flow rate in the range of 2 - 3 kgm2h "1, total productivity rates as high as 1.57 kgm2h "l were achieved. These productivity rates are more than 40% higher than those obtained in the solar only operation mode (of., Tables 2 and 8). This enhanced productivity is due mainly to the
fact that in the hybrid operation mode the temperature and the vapor content of the air stream that exits the lower chamber is much higher and consequently, the amount of distillate condensed in the external condenser (II) is significantly enhanced. It is also observed that in this hybrid mode of operation that the temperature and therefore, the water vapor content of the saturated air stream that exits the condenser and vented to the ambient was relatively high. The external condenser used in this study was incapable of recovering a major fraction of the water vapor content of the exiting air stream prior to venting to the ambient, viz., it was not efficient enough for the system when operating in the hybrid mode.
1.00 0.90
~o-.-r
0.80 "r
0.70
E
0.60
///
(3)
0.50 >
0.40
--------~ I
.,.,..
"~
0.30
ja.~a---
"0
t~ II
0.20
0 Q.
0.10
||
0.00
t
0.00
1.00
2.00
3.00
4.00
5.00
6.00
7.00
8.00
Air flow rate (kg m "2 h~) Fig. 4. Productivity as a function of air flow rate when utilizing only waste thermal energy and placing an insulating cover on the still glazing. Ambient temperature = 23.7 - 25.9~ feedstock flow rate = 9.3 kgm'2h"1 and feedstock inlet temperature = 86 - 90~ Table 7. Temperature profiles at steady-state as a function of air flow rate when operating in the hybrid mode. Ambient temperature = 28.4 - 34.4~ feedstock flow rate = 5.7 kgm2hl; feedstock inlet temperature = 86 - 90~ and irradiation = 650-&10 Wm 2. Air Flow Rate
UPPER CHAMBER T1
(kgm'2h-1) (%)
LOWER CHAMBER
Tto=
T~.I
Tx.in
T~o~
T~
T2
T3
T4
T5
T6
T7
T8
T9
T10
Tll
T12
T13
T14
T15
(~
r
(~
r
(~
r
(~
(~
r
r
r
(~
(~
r
1.33
72.2
95.(i
97.6
97.5
99.i)
97.4
97.1
48.5
94.1
84.7
87.6
86.1
84.9
51.5
73.4
2.03
62.5
88.5
91.7
91.6
94.3
92.1
91.2
42.4
87.1
78.2
80.5
79.3
78.3
49.3
67.4
3.52
56.9
80.8
85.5
85.4
89.5
86.7
84.8
36.3
80.1
73.3
74.2
72.9
71.6
48.9
60.6
4.02
56.6
79.4
83.8
83.8
90.9
86.3
83.4
34.9
78.5
72.5
73.3
72.3
71.1
51.2
60.7
223
ISES Solar World Congress 1999, Volume III
Table 8. Productivity rates and calculated waste energy input, qwme, at steady-state as a function of air flow rate when operating in the hybrid mode. Ambient temperature = 28.4 - 34.4~ feedstock flow rate = 5.7 kgm-2h-1; feedstock inlet temperature = 86 - 90~ and irradiation = 650!-_10 Wm -2.
II
Irradiation (Wm -2)
qwaste
I
(Wm:)
(kgm-2h -1)
II (kgm-2h-1)
E (kgm-Eh-1)
1.33
650
251
0.95
0.53
1.48
2.03
650
290
0.98
0.60
1.57
3.52
650
330
0.93
0.63
1.57
4.02
650
331
0.87
0.64
1.51
Air Flow Rate (kgm-2h-1)
1.8 1.6 -
iT_, t--
E
1.41.2-
1.0-
,L
0.8 "O
2 n
0.6-
,-II |l
l
0.40.2
0.0
-
-
1.0
1.5
2.0
2.5
3.0
3.5
4.0
4.5
Air Flow Rate (kgm2h ~)
Fig. 5. Productivity as a function of air flow rate when operating in the hybrid mode. Ambient temperature = 28.4 34.4~ feedstock flow rate = 5.7 kgm2hl; feedstock inlet temperature = 86 - 90~ and irradiation = 650-2-_10Wm -2.
Inter-comparison of the modes of operation The results of this experimental study are summarized in Fig. 6 with respect to the total distillate rate as a function of the air flow rate for all the modes of operation investigated. It is not recommended to operate this prototype solar still at relatively high air flow, since this will result in an increase in the parasitic electric energy required to drive the air pump. The optimum air preheating the feedstock in the lower chamber (i.e., flow through the serpentine tube) even in the hybrid mode of operation. 4. CONCLUSIONS The performance of an air-blown multiple-effect solar still consisting of an upper evaporation chamber and a lower condensation chamber has been analyzed in three modes of operation, i.e., driving forces: (i) solar energy; (ii) waste thermal energy and (iii) hybrid, both solar and waste thermal energy.
flow rate for both the hybrid and the nocturnal distillation modes is observed to be in the range of 2 - 3 kgm'eh1. We believe that a productivity in excess of 20 kgmeday "1 may be achieved when operating the still in arid zones under these conditions, i.e., hybrid mode with nocturnal distillation. It may be possible to further enhance the productivity, based upon the results of this study, by In all modes of operation the performance of the still was determined as a function of the flow rate of the entering air stream and the optimum range of the air flow rates were determined experimentally. The optimum air flow rate for both the hybrid (during the daytime) and the nocturnal distillation modes was in the range of 2-3 k g m 2h'l. Based upon the experimental results of this study, a productivity in excess of 20 kgm2day "1 may be achieved when operating the still in arid zones under these modes, i.e., hybrid mode with nocturnal distillation. It may be possible to further enhance the productivity, based upon the results of this study, by preheating the
ISES Solar World Congress 1999, Volume Ill
224
feedstock in the lower chamber (i.e., flow through the serpentine tube) even in the hybrid mode of operation. 1.8
-
1.6
-
Such a mode of operation will be studied in the future.
Sol~tr +wa ~te 1.4
4r
)
-
"7
"=
W a s t ( ; + in,.;ulate,]
~
J
0.8
91,=,l
Sohr
/
1.2
t'q
o
0.60.4
-
0.2
-
c~
.__ ~ - - J k - -
~
~
------
~ ' - - - -
"-'-
...-'~v
waste
~-j
_
0
0.5
1
1.5
2
2.5
3
3.5
4
4.5
5
5.5
6
6.5
Air Flow Rate (kgm2h "1) Fig. 6 Inter-comparison of the total productivity as a function of the air flow rate for all modes of operation investigated. The experimental conditions for each mode of operation are defined in the text. MENCLATURE c heat capacity ( J k g l K "1) m mass flow r a t e ( k g m 2 s "1) q thermal energy normalized to unit still area (Wrn2) qw~ waste energy input = q l ~ normalized to unit still area (Wm2) q~l=soa thermal energy released by the preheated feedstock in the upper c h a m b e r (Win"2) T temperature (~
Subscripts a brine f I II in out sep u w X
ambient brine drain-off feedstock primary distillate secondary distillate stream entering chamber stream exiting chamber vapor/liquid separator upper chamber water external heat exchanger/condenser
Acknowledgment- This research was supported under
Agriculture. Parker. B.F. (Ed). Vol. 4. pp. 255-
Grant No. TA-MOU-95-C15-050. US-Israel Cooperative Development Research Program. Office of Agriculture & Food Security. Center for Economic Growth. Bureau for Global Programs. Field Support and Research. USAID. One of the authors L. Horv t h acknowledges also the support of the Hungarian OTKA Fund, project No F-025 342.
294. Elsevier Science Publishers B.V. Amsterdam. Kudish, A.I., Evseev, E.G., Aboabboud, M.M., Horv~ith, L. and Mink, G. (1997) Heat transfer processes in an air-blown multiple-effect solar still with thermal energy recycle. Proceedings of the ISES Solar World Congress, Vol. 6,. Taejon, Korea, pp. 158-167. Malik, M.A.S., Tiwari, G.N., Kumar, A. and Sodha, M.S. (1982) Solar Distillation. Pergamon Press. Oxford, pp. 175.
REFERENCES Kudish. A.I. (1990)
Energy
in
Water Desalination.
Agriculture-Energy
in
In Solar World
ISES Solar World Congress 1999, Volume III Mink, G., Aboabboud. M.M., Horv~ith., L., Evseev, E.G. and Kudish, A.I. (1997) Design and performance of an air-blown solar still with thermal energy recycle. Proceedings of the ISES Solar World Congress, Vol. 6, Taejon, Korea, pp. 135-144. Mink, G., Horv~ith, L., Evseev, E.G. and Kudish, A.I. (1998) Design parameters, performance testing and analysis of a double-glazed, air-blown solar still with thermal energy recycle. Solar Energy 64, 265-277. Talbert, S.G., Eibling, J.A. and L f., G.O.G. (1970) Manual on Solar Distillation of Saline Water. R&D Progress Report No. 546. US Department of Interior.
225
226
ISES Solar World Congress 1999, Volume III
PERFORMANCE AND ANALYSIS OF A MULTIPLE-EFFECT SOLAR STILL UTILIZING AN INTERNAL MULTI-TUBULAR HEAT EXCHANGER FOR THERMAL ENERGY RECYCLE G. MINK*, L. HORV/~TH*, E.G. EVSEEV ~ and A.I. KUDISH** *Research Laboratory of Materials and Environmental Chemistry, Chemical Research Center, Hungarian Academy of Sciences, 1525 Budapest Pf 17, HUNGARY; Tel: +36 1 325 5992, Fax: +36 1 335 7892, E-mail: [email protected] **Solar Energy Laboratory, Institutes for Applied Research, Ben-Gurion University of the Negev, Beer Sheva 84105, ISRAEL.; Tel: +972 7 6461488, Fax: +972 7 6472916, E-mail: [email protected] Abstract - To achieve a relatively high productivity at reduced investment costs, an innovative, air-blown, multi-tubular
solar still module was fabricated l~om readily available, corrosion resistant materials. Performance studies were made as a function of the air flow rate (in the range of 0.36 - 5.29 kgm2hq) utilizing a solar simulator providing a constant irradiation of 630-!-_10 Wrn2 and a constant feedstock flow rate of 2.5 kgm2h1. A simulation model has also been developed to describe the heat and mass transfer processes occurring in this prototype solar still and it was validated by the experimental data. It has been found, both experimentally and by the simulation model, that the total amount of thermal energy recycled (to preheat the feedstock and to directly heat the evaporating surface) is a maximum at an air flow rate of-- 1.28 kgm-2hl; corresponding to a maximum still productivity of 0.97 kgm2hq. 1. INTRODUCTION The consumption of water unfit for drinking is a major health hazard in rural areas. To supply these rural people with water free of salinity and/or pathogens is an urgent task to be solved. Solar distillation could be an ideal source of fresh water production, however, the crucial problem is that the productivity per unit area of the traditional solar stills is low. In addition, the fixed capital investment cost of a solar desalination plant is roughly proportional to the still area. Consequently, there are two possible approaches to overcome these limitations; to either increase solar still productivity per unit area and/or decrease the fixed capital investment per unit area. In our previous papers 13 the performance of an airblown, multiple-effect solar still that offers a significant increase of the still productivity per unit area at marginal incremental costs has been presented and analyzed, viz., the first approach. In this still, consisting of an upper evaporation chamber and lower condensation chamber, a large fraction of the heat of condensation of the distillate is successfully recycled both to the evaporation plate and to preheat the feedstock. The aim of the present work was not to enhance the productivity of the above still but to simplify its construction and thereby make it a more economically viable alternative, viz., the second approach, reducing the fixed capital investment cost per unit area. This was achieved by utilizing relatively inexpensive and corrosion resistant materials in the construction of the solar still. The new solar still design, though significantly different in appearance from the original, utilizes the same heat transfer processes to obtain the final distillate product. In spite of its simple construction, the heat and mass transfer processes occurring within the still are numerous and mutually interrelated. Another aim of
this work was to develop a mathematical model, utilizing non-linear differential equations, which is capable of simulating still performance under both transitional and steady-state conditions. 2. SOLAR STILL DESIGN The solar still module under investigation is shown schematically in Figs. 1.a and lb. It consists of a bottom and edge insulated, thin, rectangular plastic tray (L = 1.84 m; b = 0.54 m; H = 12 mm; area = 1 m2), 40 glass tubes (Dolt = 7 mm; D~, = 5 mm; L = 1,8 m) covered with a black wick and a plastic serpentine tube (Dout = 4 mm; Dr, = 3 mm; L =20 m). The still glazing is a solar grade polycarbonate double-walled sheet (10 mm thick), a transparent insulation material (TIM). A low pressure variable speed air pump was used for the air, the mass and thermal energy carrier, and a peristaltic pump for the feedstock. The still shown schematically in Fig. 1, operates in the following manner: (i.) ambient air at temperature To enters at the lower extremity of the still; (ii.) evaporation occurs from the wick, which is also wets the external surface of the glass tubes; (iii.) the air stream achieves both its highest temperature and vapor content at the upper extremity of the still; (iv.) the air stream is directed into the longitudinal glass tubes at the top of the still and then flows downwards, i.e., it reverses direction, and most of its vapor content condenses on the inner surface of the glass tubes and the thermal energy of condensation is conducted via the tube wall to the wet wick to enhance the rate of evaporation from the wick; (v.) the enthalpy of air stream, at temperature Tr, entering the lower compartment, which unifies the air streams exiting the 40 glass tubes, is utilized to preheat the feedstock, prior to its entering the evaporation chamber. The feedstock flows through a 5 m long black tube packed within this compartment to facilitate the heat exchange process.
ISES Solar World Congress 1999, Volume III
(vi.) in the evaporation chamber the feedstock is further heated as it flows through another 15 m of the black serpentine tube, positioned above the glass tubes, before exiting the serpentine tube onto the black wick at the uppermost part of the evaporation chamber. (vii) the air stream and distillate exiting the solar still enter a gas-liquid separator (Sep. 1) to collect the primary
distillate (I), in order to determine the amount of distillate which condenses within the still and is a measure of the efficiency of the thermal energy recycle process; (viii) the air stream exiting Sep. 1, still containing vapors, then enters an external heat exchanger and a second separator (Sep. 2), where the secondary distillate (II) is collected.
Fig. 1.a Schematic diagram of the multi-tubular solar still and the experimental setup. temperature probes are indicated by numbers 0 to 10.
Fig. 1.b
227
The location of the
Simplified scheme of the top view of the multi-tubular solar still. The actual number of glass tubes is 40.
ISES Solar World Congress 1999, Volume Ill
228
3. EXPERIMENTAL The solar still was tilted at an angle of 20 ~. To facilitate the parametric sensitivity studies (viz., the dependence on air flow rate) a solar simulator, also tilted at 20 ~ was utilized. It provided a constant solar radiation intensity of 630-2_10 Wm 2. A 50 mm thick polyurethane foam was used to insulate the edge and bottom of the solar still (not shown in Figs. 1.a and 1.b). The differential and cumulative yields from Separator 1 and 2 were measured automatically with a type PT 6 Satorius electric balance with an accuracy of +1 g. The temperatures were measured at 11 locations with an accuracy of +1 ~ using calibrated temperature sensors of the silicon base type KTY l l-2A. The absolute humidity of air entering the glass tubes was calculated from material and energy balances on the still. The feedstock was pumped into the still at ambient temperature, which varied between 23 and 27~ The feedstock flow rate was kept constant at 2.5 kgm2h "1 throughout the experiments and the still performance was investigated in an air flow range between 0.36 and 5.29 kgm-2h1.
A data acquisition system served both to monitor and store the temperature data from the eleven thermistors and to calculate differential and cumulative yields at variable time intervals. It consisted of a PC with an A/D-D/A converter card, electronic measuring and magnetic valve control unit, temperature sensors and a digital balance with a RS232C serial interface. The data acquisition, control and analysis software were developed for the study. 4. RESULTS AND DISCUSSION Transitional stage of the distillation Athe mass and energy flows were normalized to a 1 m 2 still area. Due to the low thermal mass of the still, the steady-state yields and temperatures were achieved within one hour of start-up in all experiments. A set of experimental result, typical of those obtained, are shown Figs. 2 and 3. The transitional and steady-state temperatures corresponding to air temperatures T1, T2 and T5 in the evaporation chamber (cf., Fig. 1.a) are shown in Fig. 2.
90 _AAAAAAAAAAAAAAAAAAAAAAAA 80 --
70-A 60 -0.
All I
50
;
iii IIIIIIIIIIIIIIIIIIIIIIIIIIII
&I I
iI
IT1 i T2
40 ~-i~I~l!
AT5
30 20 0
I
I
I
i
i
I
20
40
60
80
100
120
140
Time (min) Fig.2.
Variation of the air temperatures T1, T2 and T5 with the time of irradiation. Feedstock flow rate = 2.5 kgm2h -1, irradiation = 630s Wm "2.
The approach to steady-state distillation rates, primary (I), secondary (II) and total (Z), as a function of the time lapse since start-up is presented in Fig. 3. The primary distillation rate refers to that condensed within the glass tubes and in the feedstock preheating compartment of the still. The secondary distillation rate is that obtained by passing the saturated air stream exiting the solar still through an external heat exchanger prior to venting to the ambient.
It is observed in Fig. 3 that there is a time delay of 27 minutes before the first breakout of distillate from the still. This is caused by the fact that when starting with a dry still, distillate first appears in the separator only after all the inside surfaces have been wetted and the small but not negligible dead volumes of the system are filled with the distillate. Once steady state is achieved, approximately after one hour, a random variation in the measured yields is observed. This is due to the high surface tension water which flows into
ISES Solar World Congress 1999, Volume III
separator 1 both continuously and periodically, in the form of rivulets. Therefore, steady state yields have been defined as an average of the data measured during
Fig. 3
229
at least one hour of operation under steady state conditions.
Still productivity as function of time since start-up. Feedstock flow rate = 2.5 kgm-2h1, irradiation = 630-2_10 Wrn-\ r~
Still performance under steady state conditions The temperature profile of the still at different air flow rates expressed in kg bone dry air per m 2 still area per hour is reported in Table 1, where T~mb is the ambient temperature; TVREnEATthe temperature of the
preheating compartment; TSEPa the temperature of the first separator; and Tx,m and Tx,o~t the air temperatures entering and exiting the external heat exchanger, respectively.
Table 1. Temperature profile of the multi-tubular solar still at steady-state as a function of air flow rate for a constant feedstock flow rate of 2.5 kgm2h 1 and irradiation = 630-2_10 Wm 2. The position of the thermistors are shown in Fig. 1.a. Air Flow
Tamb
CHAMBER
Rate (kgrnEh1)
To (~
T1 (~
T2 (~
T3 (~
T4 (~
T5 (~
(~
(~
(~
(~
0.34
23.5
80.3
97.1
99.1
97.8
98.3
80.1
57.7
51.9
25.8
73.6
65.9
63.6
27.3
TpREHEhT TSEP.I T6
T7
Tx,t~
Tx,out
T8
T9
1.28
24.7
76.2
91.1
93.6
90.2
92.8
2.53
25.0
65.5
75.7
82.3
84.0
84.6
65.6
57.8
56.3
26.8
5.29
27.1
57.2
62.3
70.9
73.4
75.2
60.1
53.3
52.8
28.8
The still productivity, as a function of air flow rates in the range from 0.34 to 5.29 kgm-2h-1 is reported both in Table 2 and in Fig. 4 . It is observed that for the primary distillate rate, which is a direct measure of the thermal energy recycle efficiency, there exists an optimum range of air flow rates in the vicinity of 1.3 kgm2h 1. It is also observed that the ratio of secondary to primary product increases with increasing flow rate, i.e., the primary decreases and the secondary increases with increasing air flow rate (cf., Fig. 4). The reason
for an optimum air flow rate has been discussed in our previous papers 2"4 and it will be also shown later in section 5 that the simulation model also predicts a maximum for the primary distillation at the same air flow rate. The heat and mass balances on the still were calculated from the experimental data assuming that the air stream, after its temperature in the glass tubes dropped to the dew point of the air that enters into the glass tubes, was always saturated while passing
ISES Solar World Congress 1999, Volume III
230
through the glass tube to the ambient. Consequently, at T9. The dew point temperature of the air and the saturated air exits the glass tubes at temperature T6, mass flow of vapor as calculated from the mass and enters separator 1 at TT, the external heat exchanger at energy balances on the still are reported in Table 3. Ts, the separator 2 at T9 and vented to the ambient also Table 2. Still productivity as a function of air flow rate. Feedstock flow rate = 2.5 kgrn2h-1; irradiation = 630s Wm 2.
, (kgm2h Flow-1)te II(kgm-2h I -1)
II
Z
(kgm2h "1)
(kgm'2h'l) '
0.56
0.05
0.61
1.28
0.82
0.16
0.97
2.53
0.68
0.19
0.87
5.29
0.66
0.29
0.95
0.34
~
Fehler! Keine gfiltige Verkniipfung.
Fig. 4.
Still productivity as a function of air flow. Feedstock flow rate = 2.5 kgm2h "1, irradiation = 630-~_10 Wm 2.
Table 3. The mass flow of vapor at steady-state as a function of air flow rate. Feedstock flow rate = 2.5 kgm2hl; irradiation = 630s Wm "2. Air Flow Rate (k~l:l2h -1)
TlxmX
TDw
(Ts)
(~
(~
m v,m .B
mv,m~xD
mv~
mv~ r
(k~rm'2h"1) (kgm-2h-1) 0cgm-2h-1) (kgm-2h-1) (kgm-2h-1)
0.614 0.617 0.003 91.9 98.3 1.000 0.013 0.987 84.3 92.8 0.901 0.927 0.026 74.2 84.6 2.53 1.084 1.022 0.062 65.0 74.2 5.29 A _ the calculated dew point of air stream turning into the glass tubes at T.~; B _ the mass flow of vapor carried by the entering air stream at Ta; c _ evaporaion rate in the evaporaion chamber; D _ the mass flow of vapor at Tm~; F _ the mass flow of vapor entering the exchanger at Tx:,; F _ the mass flow of vapor exiting the exchanger at Tx,out and vented to the ambient. 0.34 1.28
The following comments are based upon the above analysis presented in Table 3" 1. The calculated dew point temperatures are below the observed fluid temperature Tm~x,, which suggests that the mass transport between the wet wick and the air stream is not intense enough. 2. The operational optimum is observed in the vicinity of an air flow rate of 1.28 kgm-2h"1 since (i) the total productivity is high (0.97 kgm2hl); (ii) the load on the heat exchanger is moderate (0.18 kgm2h "1 vapor at a relatively high fluid temperature (63.6 ~ (iii) the amount of vapor vented to the ambient is marginal (0.03 kgm2h q) and (iv) because the air flow rate is low, the parasitic electrical energy required to drive the air pump is marginal (< 1 W per m 2 still area), which in rural areas might be provided by PV panels.
.
0.057 0.180 0.247 0.414
0.007 0.030 0.057 0.134
Increasing air flow rate results in a higher evaporation rate but lower condenser efficiency and thereby, greater thermal energy and water vapor losses which are vented to the ambient. Additionally, at higher flow rates the parasitic electrical energy requirement increases.
5. SIMULATION MODEL OF THE MULTITUBULAR SOLAR STILL A simulation model of the multi-tubular, air-blown solar still was developed. The temperatures of the air stream and feedstock at each node in the thermal analysis of the solar still system were modeled by the appropriate energy and mass balance relationships. This resulted in a model consisting of a set of nonlinear energy transfer equations. The model was solved numerically, using an explicit predictor-corrector difference scheme assuming "steady-state conditions"
ISES Solar World Congress 1999, Volume III
by using sufficiently small time intervals during which the feedstock flow rate and ambient temperature are assumed to be constant. The following assumptions have been made in this analysis: 1. The water film is replenished locally on a continuous basis, is stagnant and very thin; thus its thickness and temperature are assumed to be that of the wetted wick. 2. The system is considered to be uni-dimensional in the flow direction of the individual streams. 3. The Lewis number for the air-water mixtures in the operational temperature range is assumed to be equal to 1.0, yielding l~m-----Cs 9 4. In theevaporator, as well as in the internal heat exchanger in the lower compartment, a single-row serpentine geometry is assumed. Therefore, the moist
231
air stream is assumed to pass in a cross-flow pattern over the feedstock tube. 5. In order to simplify the model by eliminating the j-th surface temperatures Tj (j=g, t or w) and to preserve at the same time a reasonable accuracy for the longwave radiative exchange processes (Duffle and Beckman, 1980), the net flux containing the above temperature is expressed by Qrj = Ej(~.( T 4- Tj 4) = hrj "( T - Tj), where hrj = 413jO-[( Tj,in + 2T + Tj,out )/4] 3 and T is the corresponding air stream temperature, either above or below the surface node.
Mathematical model The total thermal energy transferred through any j-th surface temperature Tj (j=g, t or w) is given by (2) O0) + U0),i~'( Tin- Tj) = Aout/Ain-U0),out-( T]- Tout), where Q0) is the sensible thermal energy transfer due to either the condensation/evaporation process or the incident solar radiation heat flux.. The total temperature difference above and below any surface is given ATtot- Tin- Tout = ATe, + ATout; AT~n= Tin- T3, ATout = Tj- Tout.(3) The differences from Eq. (3) may be calculated formally in term of ATtot, by substituting for the corresponding surface temperature as ATin = (O(j),,in+Aout/Ai~-U0),out )-1 [Aout/Ain.O0),out ATtot- Q0)], (4) ATout = (U0),,i~+Aout/Ai~'U0),out)-1 [U0),i~ ATtot + Q0)](5) The glazing, tube wall and wetted absorber temperatures may be substituted from the corresponding mass and energy balances (Veza et al., 1993; Kudish et a1.,1997). Consequently, the energy transfer equations in the evaporator are obtained as follows: (pc/5)~OTa~/& + ma/bw~" O((Ca+ W2 c, )Ta,2)/Ox2) - Qev,2 - Qoo~t~ - Qcon,g + Ag/Aw,2-O2 (Uamb + 02 )'l[Uamb(Tamb - Ta.2) + Ocon.g + Gg] + UE(U1 + Din,1/Dout,l'UE)'l'[Din,1/Dout.l'U](Ta.1- Ta,2) + Qev,2+ Din,1/Dout3-Qcon.1 + Gw] + At/Aw,2" UE(U2 + Oin,t/Oout,t'Uf)'l'[oin,t/Oout,t'Uf(Tf,2 -Ta.2) -Qcon.t,2 - Gt], (6) (pc Din t2/4Dout,t)f.2OTf.E/& + me Cy//1;Dout, t'OTf,E/OXt= Qeon,t~ + Gt + lJr (Din,t~o:~tUf ~-U2)"1 [U2(T2-Tf) - Qoon,t~- Gt], (7) ma ()W2/t) X2 "- hm,2b2" (Ws,2 ((Ta,2+Tamb)/2) - W 2 ) . ( 8 ) Similarly, the energy transfer equations within the glass tube are given by the following, where X 1 -- -X 2 to account for the reverse in flow direction: (pc Di~ 12/4Dour 1)a,lOTa,1/& +ma,1/~Doult,l" O((ca+W1 Cs )Ta,1)/~Xl) = - Qcon,1 + [.Jl(Din,1/Dout,lOl + U2) [ W2(Ta,2 - Ta,1) + Oeva + Q0o~,1+Gw],(9) ma,1 OW1/~Xl- h~l- ~Do,~l" (Ws, l((Ta, l+Tamb)/2)- W1). (l 0) The energy transfer equations for the air stream, feedstock and humidity in the internal heat exchanger are analogous to Eqs. (6) - (8) but with a view to being concise they are not detailed in this manuscript. Equations (6)-(10) were solved by imposing the following initial and boundary conditions: Ta,l(X,0) = Ya,2(x,0) - Tamb, Yf,2(x,0)= Ywag0, W2(x=0) - Wamb, (11) Ta~lx=O= Zamb, Tf,2Jx=O= Twat,i~, Ta,llx=L= Ta,2lx--L, Wllx=L- W2]x=L-(12) The definitions for the heat and mass flux and corresponding transfer coefficients applying to Eqs. (6)- (10) are: Qev = ~, (Ta)M~v, Q~o.= ~(Ta)Mr (13) U2 = hr + 0.9-a-4-[(Ta,1 + 3"Ta.2)/4]3, (14)
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232
Uamb= hamb+ 0.9"6"4"[(3"T arab+ Ta,2)/4]3, 1/U1 = 1/he,1 +Din, l-log(Dout,1/Din,1)/2Kw + 1/heon, l, (16) 1/Uy = 1/hf + Din,t'log(Dout,t/Din,t)/2~:w, 1/hamb= 1/hg + 1/ha + 2-~ig/rg, (18) hv,d = 5.7 + 3.8Vamb; (19)
(15) (17)
where hg (cf., Eq. (18)) is determined from the Nusselt number correlation for natural convection between two parallel planes as proposed by Buchberg et al. (1976): Nug = 1 + 1.446(1 - 1708/Ra*), for 1708 < Ra*< 5900, (20) and Ra* = 2gp 2(Ta~ - Tamb) ~a3cOs0"Pr/(3T~2 + Tamb)Ba2. (21) The heat transfer coefficient for forced convection in the laminar flow regime is determined from the following correlation (Heaton et al., 1964), Nu~j = 5.4 + 0.0019[RejPrDrq/L]l71/(1 + 0.00563[RejPrDnj/L]L17), j=1,2, (22) which is valid for Rej< 2.3"103. The values for l~j determined by Eq. (22) were corrected for the effect of simultaneous heat and mass transfer by applying the Ackerman correction (Treybal, 1980). The forced convective heat transfer coefficient hf for laminar flow is estimated from the Nusselt number (Kreith, 1976), Nuf = [3.65 + (0.0668RefPrDm, t/Lt)/(1 + 0.04(RefPrDm,t~) 2/3. (23) The overall heat transfer coefficient through the bottom of the still to the ambient is given by Oloss
= 1/(Sim/lq~s + ~)b/Kb). (24)
The condensation heat transfer coefficient at low mass fluxes inside smooth horizontal tubes is given by Chato (1962) with the correlated coefficients proposed by Singh et al.., (1996): 3 lko~,~ = 0.0925-[p~pw- Ps)g~* ~:w/D~B~Ta,1Ta~)]~/4, ~*= ~, +0.68 Cw(Ta,1- Ta~). (25a)
(25)
The heat transfer coefficient for condensation from the saturated air stream on the inner glazing surface is given by (Kreith, 1976) as h~
= 0.725[ps(Ps- Pa)g sin0~3/2l-aBs(T~- T arab)]v4 .
(26)
The heat transfer coefficient for condensation from the moist air stream onto the surface of the feedstock tubes is given by (Kreith, 1976) as
heon,t = 0.943[pw(Pw- ps)g~:ff/2L1B~(Ta,1-Tamb)] 1/4.
(27)
The mass transfer coefficient hm is determined utilizing assumption 3. The rates of evaporation and condensation are calculated using the following relationships: M ~ = hm,co,,[W~2 - Ws (0.5(T~2 + Tamb))], (28) Moo~,tj = hm,~[W~j - W, (0.5(T~j+ Tfj))], j = 2,3, (29)
M~= h~ov[Ws (0.5(T~2+T,,0)- W~2],
(30)
where the value of Ws at the given temperature Ta is calculated from Ws = 0.622- V/(V- Ps). (31) The saturation pressure of the water vapor P~(Ta,) is evaluated from an empirical formula which was derived by a least square analysis of the data taken from the steam tables (see Elsayed, 1983). Numerical solution We have used the "predictor-corrector" difference scheme 0Vlarchuk, 1975) along each x-direction for the numerical solution of Eqs. (6) - (10), in the following format: (Ti n+l/2- Tin)/0.5At + f (W") (Tin+'/2- Ti-1 n+l/2)/AX = F(Tin), (32) 0.5-(Tin+l _ Tin + Ti 1n+l - Ti.I n )/At + f (W n+l/2)(Tin+l/2 . Ti-1 n+l/2)/Ax = F(Ti n+l/2), (33) where i=l,...,I and I=L/N is the number of mesh points.
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function of air flow rate are shown in Fig. 5 and the dependence of the primary distillate on the air flow rate is shown in Fig. 6. In both cases, there is good agreement between the measured and calculated values. It is important to note, that the simulated primary distillate rate exhibits a maximum value, corresponding a maximum in the thermal energy recycle efficiency, at the same location as that measured.
At first, we utilize Eq. (32), a linear but strongly stable scheme, to obtain intermediate values of Tn+1/2 and apply Eq. (3) to refine the calculation by correcting for the whole step. Convergence was achieved for the initial air and fluid temperatures with various gridsteps. The results shown in Figs. (5) and (6) were obtained with grid-steps Ax = 0.1m and At = 0.3s. The measured and calculated air temperature T5 and T6 as a
120 100
G"
~
T5meas
80c
e
600) [-.,
T5calc T6meas
40
e
T6calc
20 I
I
I
I
I
1
2
3
4
5
Air flow rate (kgm2h "1)
Fig. 5
Measured and simulated values for T5 and T6 as a function of air flow rate.
6. CONCLUSIONS An innovative, air-blown, multi-tubular solar still fabricated from readily available, corrosion resistant materials has been studied experimentally and a simulation model describing it was developed and validated by the experimental data. The performance testing was done using a constant feedstock flow rate of 2.5 kgrn-2h1 (sufficient to maintain the wick completely wetted). A maximum in the steady-state productivity ( 0.97 kgrn2h-1) was
observed as a function of the air flow rate at 1.28 k g m 213-1, both experimentally and predicted by the simulation model. These optimum operating conditions correspond to the highest thermal energy recycle efficiency of the system, a relatively low thermal load on the external heat exchanger and relatively low parasitic electrical energy requirements to drive the air pump.
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1 I
0.9 0.8
~,= '~ ..~ ~9
0.7 0.6 0.5 0.4 o 0.3 0.2 0.1 0
--4-- Imeas "
0
I
I
I
I
I
1
2
3
4
5
Icalc
Air flow rate (kgm-2h1) Fig. 6
Measured and simulated values of the primary distillate rate as a function of air flow rate.
NOMENCLATURE
arnb
A b c D(Dn) G g H h hm L M m N P Q T U V W~
surface area (m2) width (m) heat capacity (Jkg'lK "1) diameter (hydraulic diameter) (m) solar radiation (Wm"2) gravitational constant (ms "E) height (m) heat transfer coefficient ~m-2K "1) mass transfer coefficient (kgs-lm-2K1) length (m) mass flow rate (kg m2s 1) air/water flow (kg sl) number of mesh points in the chamber pressure (Nm-2) thermal energy (Wm2) temperature (K) overall heat transfer coefficient (Wm2K1) linear velocity (ms 1) saturated humidity for air at T (kg v-(kg dry
b c
W
air humidity (kg v.(kg dry air)"1)
air) "1)
Greek thickness (m) surface tilt angle thermal conductivity (WmqK1) latent heat of vaporization (Jkg-1) dynamic viscosity (kgm-ls-1) density (kgm-3) Stefan-Boltzmann constant (Wm2K "4)
Subscripts a
air
con ev f g in ins out r s t w wd 1 2 3
ambient bottom convection condensation evaporation feedstock glazing inlet/inside insulation outlet/outside radiation saturated tube wick/water wind internal heat exchanger evaporator external heat exchanger
Acknowledgment- This research was supported under Grant No. TA-MOU-95-C 15-050, US-Israel Cooperative Development Research Program, Office of Agriculture & Food Security, Center for Economic Growth, Bureau for Global Programs, Field Support and Research, USAID. One of the authors Mr. L. Horvfith acknowledges also the support of the Hungarian OTKA Fund, Project No. F-025 342. REFERENCES Aboabboud, M.M. and Mink, G. (1993) Solar still of improved efficiency. Proceedings of lSES solar World Congress, Vol. 4, Budapest, Hungary, pp.319324. Kudish, A.I., Evseev, E.G., Aboabboud, M.M., Horvfith, L. and Mink, G.(1997) Heat transfer processes in an
ISES Solar World Congress 1999, Volume III
air-blown, multiple-effect solar still with thermal energy recycle. Proceedings of the ISES Solar World Congress, Vol. 6, Taejon, Korea, pp. 158-167. Mink, G., Horv~ith, L., Evseev, E.G. and Kudish, A.I. (1998) Design parameters, performance testing and analysis of a double-glazed, air-blown solar still with thermal energy recycle. Solar Energy 64, 265277. Mink, G., Aboabboud, M.M., Horwith, L., Evseev, E.G. and Kudish, A.I.(1997) Design and performance of an air-blown solar still with thermal energy recycle. Proceedings of the ISES Solar World Congress, Vol. 6, Taejon, Korea, pp. 135-144. Buchberg, H., Catton, I. and Edwards, D.K. (1976) Natural convection in enclosed spaces: A review of application to solar energy collection. Trans. of ASME, Jr. Heat Transfer 98, 182-188. Duffle, J.A. and Beckman, W.A. (1980) Solar Energy of Thermal Processes, Wiley Interscience, New York, 762 pp. Veza, J.M.,and Ruiz.,V. (1993) Solar distillation in forced convection. Simulation and experience. Renewable Energy 3, 691-699.. Heaton, H.S., Reynolds, W.C. and Kays, W.M. (1964) Heat transfer in annular passages. Simultaneous development of velocity and temperature fields in laminar flow. Int'l. J. Heat & Mass Transfer 7, 763-781. Kreith, F. (1976) Principles of Heat Transfer, 3rd edn, Harper & Row, Publishers, New York, 656 pp. Marchuk, G.I.(1975) Methods of Numerical Mathematics, Springer-Verlag, New York, pp. 316. Treybal, R.W. (1980) Mass Transfer Operations, 2nd edn, McGraw-Hill, New York, 717 pp.. Chato, J.C. (1962) Laminar condensation inside horizontal and inclined tubes. ASHRAE Journal 4, 52-60. Elsayed, M.M.(1983) Comparison of transient performance predictions of a solar-operated diffusiontype still with a roof-type still, Journal of Solar Energy Engineering 105, 23-28. Singh, A., Ohadi, M.M., and Dessiatoun, S.V., (1996) Empirical modeling of stratified-wavy flow condensation heat transfer in smooth horizontal tubes, ASHRAE Transactions 102, 596-603.
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Modelling Thermosyphon Solar Water Heaters with Mantle Collector-Loop Heat Exchangers Graham L. Morrison, Gary Rosengarten and Masud Behnia Department of Mechanical and Manufacturing Engineering, University of New South Wales Sydney, Australia 2052 [email protected] Abstract- This paper describes the characteristics of horizontal mantle heat exchangers for application in thermosyphon solar water heaters. A new correlation for heat transfer in horizontal mantle heat exchangers with bottom entry and exit ports was used to predict the overall heat transfer and stratification conditions in horizontal mantle tanks. The model of the mantle heat exchanger tank was combined with the thermosyphon solar collector loop model in TRNSYS to develop a model of thermosyphon solar water heaters with collector loop heat exchangers. Predictions of stratification conditions in a horizontal mantle tank are compared with transient charging tests in a laboratory test rig. Predictions of daily energy gain in solar preheaters and systems with in-tank auxiliary boosters are compared with extensive outdoor measurements and the model is found to give reliable results for daily and long term performance analysis.
1. INTRODUCTION The thermosyphon solar water heater is the principal product concept in most major solar water heater markets. Thermosyphon systems with open loop connection between the tank and the solar collector have been widely adopted for both low pressure and pressurised water supply systems in climates that do not experience freezing conditions. Without freeze protection these systems are limited to tropical climates or locations that never experience frost. Freeze protection in open loop thermosyphon systems can be provided by water dump valves, electric heating in the collector header and by tapered riser tubes that control the growth of ice so that a rigid and expanding ice plug is not developed. Although all these techniques have been used in commercial products, they are not suitable for widely traded products that may be installed in any climate zone. The only inherently freeze tolerant systems are drain-down systems or closedloop collector circuits with a heat exchanger between the collector and the tank. The drain down concept is difficult to implement in a thermosyphon system if a pressurised water tank is used, however, drain-down thermosyphon systems are widely used in China for systems that provide hot water only in the evening. There have been numerous studies of the performance of closed loop thermosyphon systems based on tube heat exchangers inside vertical and horizontal storage tanks (Mertol et al. 1981, Webster et al. 1987). Tube heat exchangers can provide adequate heat transfer between the collector loop and the tank, however internal tube heat exchangers for thermosyphon operation are difficult to construct. The heat exchanger configuration that has been widely adopted for horizontal tank thermosyphon systems is the mantle or annular concept shown in Fig 1. A mantle heat exchanger is easy to construct, provides large heat transfer area and with appropriate design can promote thermal stratification in the storage tank. The primary advantage of mantle heat exchangers for thermosyphon systems is that they provide little resistance to thermosyphon circulation and maintain the simplicity of the thermosyphon concept. Mantle heat exchangers are also used for vertical tank, pumped circulation systems (Furbo, 1993, Baur et al, 1993, Shah and
Furbo, 1997). If a mantle heat exchanger had been used in the study by Webster et al (1987) instead of the eight immersed copper tubes, the heat transfer area would have been two and a half times larger and the heat exchanger penalty would have been significantly reduced. Furbo (1993) compared low flow solar water heating systems with a range of heat exchanger configurations and found that mantle heat exchangers outperformed immersed coil and external shell and tube heat exchangers in vertical tank systems. Although many manufacturers of thermosyphon solar water heaters use horizontal mantle heat exchangers there is very little information on the performance characteristics of this type of heat exchanger. Manufacturers of horizontal tank systems usually take a conservative approach to sizing and use the largest possible mantle (full c~umference and full length of the storage tank). A wide range of mantle widths and positions of the hot inlet pipe have been used. Systems designed to operate as solar pre-heaters typically have the hot-inlet pipe mounted near the top of the annulus. Systems with in-tank electric boosting have the hot-inlet pipes mounted below the level of the electric heater or have both the inlet and outlet mounted in the bottom of the annulus, Fig 1.
Fig 1. Full circumference mantle heat exchanger on a horizontal tank, showing alternative collector return points.
ISES Solar World Congress 1999, Volume III
Baur et al (1993) also studied vertical mantle heat exchangers for pumped circulation systems using an empirical heat transfer correlation for laminar flow in the mantle gap developed by Mercer et al (1967) and a finite difference solution for the energy exchange between the mantle fluid and the tank contents. Based on experimental assessment, Baur concluded that there was little difference in the performance of pumped circulation solar water heaters incorporating vertical tanks with mantle heat exchangers or external heat exchangers. Inlet connections to the upper level in the mantle can result in substantial heat loss due to reverse circulation at night unless the input pipe has insulation equivalent to the tank insulation. High level connections may also result in poor circulation under low radiation conditions due to opposing buoyancy in hotter upper levels of the annulus. If a single tank mantle system with an upper level electric booster is used, the input level to the mantle must also be below the level of the boost element to avoid heat dumping at night and low circulation during the day. Due to the combination of these effects in many systems a low level input point is typically used for horizontal tank mantle systems. For a low level connection to the annulus (Fig. 1) the flow in the annulus is a combination of forced circulation from the collector loop and internal natural convection within the annulus. This paper presents a model of horizontal mantle heat exchangers and the integration of this model into the TRNSYS dynamic simulation package.
237
tank conditions. The flow structure in a bottom entry mantle depends on the temperature of the inlet flow relative to the thermal stratification in the inner tank. Numerical simulation of the flow structure in the mantle is shown in Figs 2 and 3 for mixed and stratified inner tank conditions. In Figs 2 and 3 half of the curved mantle has been unwrapped so that the flow field can be represented on a 2D projection. For all operating conditions there is a complex flow structure near the impinging inlet which results in 10 to 15 % of the total heat transfer taking place near the inlet port. For a uniform inner tank and mantle inlet temperature higher than the top of the inner tank (Fig 2) the mantle flow covers the full circumference of the heat exchange surface. For stratified inner tank conditions and a mantle inlet temperature less than the top of the tank, the mantle flow only rises to its thermal equilibrium level, Fig 3. For stratified conditions and a mantle inlet temperature colder than the top of the tank a recirculation zone develops above the mantle inlet as shown in the bottom right comer of Fig 3, where the inlet is displaced slightly from the end wall.
2. MODELLING THERMOSYPHON SYSTEM PERFORMANCE Simulation models for mantle heat exchangers on vertical tanks have been developed by and Baur et al (1993), Furbo (1993) and Shah and Furbo (1997). The model proposed by Baur has been implemented in the TRNSYS simulation package, and a heat exchanger model based on turbulent heat transfer correlations has been implemented in the WATSUN simulation package. The modelling procedure is the same for top entry mantles on either vertical or horizontal tanks. However, for horizontal tanks with low level or bottom entry into the mantle there is a significantly different flow structure due to mixed forced and free convection processes inside the mantle. Horizontal mantle tanks usually have a much narrower annular gap than vertical systems, typically 5 mm for horizontal systems and 20 mm or more for vertical systems. The flow entry port normal to the heat exchange surface in a horizontal mantle heat exchanger also results in higher local heat flux levels near the entry, due to impingement effects. The flow structure in a horizontal mantle with bottom entry and exit points has been investigated experimentally and numerically by Nasr et al (1997, 1998) and Morrison et al (1998). Due to the large area of a full circumference mantle and low flow rates inherent in thermosyphon solar water heaters the flow in horizontal mantle heat exchangers is usually in the developing laminar flow regime. For flow rates corresponding to the high end of thermosyphon circulation, mixing induced by the impinging inlet flow has been observed by Rosengarten et al (1997). The flow structure in a horizontal mantle has been studied by Morrison et al (1998) for both mixed and stratified inner
Fig 2. Simulated flow structure in a horizontal mantle heat exchanger with bottom level entry and exit for mixed inner tank temperature 30~ and a mantle inlet temperature of 50~ (half of the horizontal mantle circumference unwrapped).
Fig 3. Simulated flow structure in a horizontal mantle heat exchanger with bottom level entry and exit for stratified inner tank conditions 20~ - 40~ and a mantle inlet temperature of 30~ (half of the horizontal mantle circumference unwrapped).
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3. HEAT TRANSFER CORRELATIONS Heat transfer in vertical mantle heat exchangers in solar water heaters has been modelled by Baur et al (1993) using a correlation (eqn 1) for developing laminar flow between parallel flat plates developed by Mercer et al (1967).
0.!6 Rew Pr N u = 4 ~ 6 § .............................................................................................
(])
Nu =Nusselt numbeT based on .logmean temperature difference hw =
............... k
given in eqn (1) is used in a thermosyphon model the additional iteration loop to find the log mean temperature for the mantle slows the convergence process. Although the log mean temperature difference form of the Nusselt number has the attraction of a constant value for fully developed flow, it is difficult to apply in a thermosyphon loop simulation model. Mercer at al (1967) also suggested a more direct solution for heat transfer using a Nusselt number based on a heat transfer coefficient defined in terms of the temperature difference between the inlet fluid and the wall. This modified form of Nusselt number is referred to as a non-dimensional heat flux to differentiate it from the customary log-mean temperature difference form of the Nusselt number. The modified Nusselt number N~ based on inlet temperature difference is given by
h = average convective heat transfer coefficient
q w Twall k
Nu w D
and QfhHLATIAr
(2)
Tin I-I
where q = average heat flux over the heat transfer surface.
R % = Reynolds number based on mantle gap width ~w
Hw# Pr = Prandtlnumber = :PeP k
A
=
(
(T. - r.a )
ln (To-T~mu)
The form of correlation function that is applicable to this definition of nondimensional heat flux can be derived as follows. The variation of fluid temperature in a heat exchanger with a constant wall temperature is given by
T.o-T~a ATo ~ 1 HL~I , r_..-
= at.
=
(3) )
(r. -
T,,~u = temvemure of heat ~ansf= surface T., = fluid inlet temlmmture To ffi 9
outlet temlmratme from the mantle
= flow rate through the ~ e
A = heat transfer ~ w ffi ~
area
o f mantle gap
H
= width of mantle- Imrpendicularto flow d ~ o n k = thermal conductivity of fluid ] . / = viscosity of fluid
Cp= specific heat of fluid The results of the simulation model were compared with experimental data reported by Furbo and Berg (1992). For vertical entry into the top of the mantle Baur found that a correction factor of 1.8 had to be applied to Mercer's heat transfer correlation for flow between flat plates. The heat transfer coefficient in eqn (1) is based on the log mean temperature difference and hence requires knowledge of both the inlet and outlet temperatures for the flow through the mantle. Although this type of correlation has been successfully implemented for pumped circulation systems by Baur et al (1993) it presents numerical difficulties if used in models of thermosyphon solar water heaters. Numerical solution of thermosyphon collector-loop circulation through a mantle heat exchanger requires an iteration for the collectorloop flow linked to an iteration for the tank temperature stratification. If a mantle heat transfer correlation of the type
whcrc L = length of mantle if flow dim~on.
The ovcralt heat transfr is given by
Q-mp(ro - r.) =
(4)
Aro)
The modified n o n d ~ i o n a l heat franker coef~i,nt t h ~ becomes ~. Q I w
u~= HLAr.-~
Although the wall temperature in a mantle heat exchanger will not be constant, the form of the correlation given in eqn (5) could be expected to apply to other boundary conditions if the mean wall temperature is used. Conventional heat exchanger effectiveness functions cannot be applied to mantle heat exchangers because the free convection flow rate on the tank contents side of the heat exchanger wall is not usually known. Heat transfer in horizontal mantles with the entry and exit points at the bottom of the mantle 9(Fig 1) has been studied by Rosengarten et al (1998,1999b). Rosengarten measured heat transfer in a vertical-slot mantle heat exchanger with bottom entry and exit ports, for a range of flow rates, mantle widths and stratification conditions on the heat exchanger surface (set by stratification in the inner tank). Rosengarten also used
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The variation of the modified Nusselt number N u w with Reynolds number Rew in a typical horizontal mantle is compared in Fig 5 with the Nusselt number for developing flow between parallel fiat plates with one plate heated. This comparison indicates that the heat transfer in a horizontal mantle with mixed convection conditions is significantly higher than for heat transfer from flow between parallel flat plates. It is interesting to note that the increase of heat transfer in a horizontal mantle is similar to the increase of 80% observed by Baur et al (1993) for vertical mantles with top entry. Although the geometry of these two cases is very different the flow in the vertical mantle is also strongly influenced by buoyancy effects, Shah et al (1999), hence a similar heat transfer correlation may apply for both orientations of the mantle.
manUe z
2.5
Fig 4. Comparison of measured heat transfer in a horizontal mantle heat exchanger and the modified Nusselt number correlation, eqn (6). a detailed CFD model of the mantle and the inner tank to determine the distribution of heat flux in a horizontal mantle. The experimental and numerical data for overall heat transfer was correlated using eqn(5) with a value of h selected to give the best fit to a wide range of measured and simulated data. For typical conditions in thermosyphon solar water heater mantle heat exchangers ( 1 [] Rew D 1 O0 , w / L [2 0.02 and w / H [2 0.07 ) Rosengarten found that a value of h =392 W/m2K gave the best fit to the measured data, as shown in Fig 4.
Rosengarten et al (1999b) also developed a stratification correlation factor (St) to account for non-uniform wall temperatures due to stratification in the inner tank. The form of the correlation for bottom entry and exit mantles is
(6) whr162162
St = 0.93- 0.050+ 0.1202 o = (T= -
-
T~ao,, , = temperature aI the bottom of the mantle wall
Tm,~. = mean wall t e m ~ Tjn= fluid inlet t e m ~
to the mantle
z0
. . . . . . . .
1.5 z c
"1.0
0.5
0.0
T
0
2O
40
60
80
Re,adds N~ Re.
Fig 5. Comparison of heat transfer in a horizontal mantle heat exchanger and in a parallel channel heat exchanger with heat transfer through one isothermal boundary.
4. MODELLING OF STRATIFICATION IN A SOLAR PREHEATER The local heat transfer rate around the circumference of a horizontal mantle heat exchanger depends on the flow structure in the mantle and the variation of heat-transfer area with height. Due to the circular cross section of a horizontal tank a large proportion of the heat transfer area is in the top and bottom sections of the tank. If a horizontal tank divided horizontally into 20 equal mass segments then 20% of the heat transfer area will be in each of the top and bottom segments, for a 40 element tank approximately 15% of the area is in each of these elements. The development of stratification in a horizontal tank with a mantle heat exchanger was measured in a scaled 57 litre tank instrumented with thermocouple grids in the inner tank (Fig 6).
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ISES Solar World Congress 1999, Volume Ill
and the effect of jet impingement opposite the bottom inlet results in a relatively uniform temperature zone in the bottom of the tank. 4.1 Simulation model
Fig 6. Laboratory test rig used to measure mantle heat exchanger characteristics.
Fig 7 Measured and simulated heating of a horizontal tank with a mantle heat exchanger. Mantle inlet temperature 41.7~
The collector loop hot water flow into the mantle was supplied from a temperature-regulated source. A three-hour heat-up test was monitored for steady mantle flow conditions. The tank was started with a slight temperature stratification (25.3~ to 27.5~ and the mantle flow was set at 0.582 litres/min and a temperature of 41.7~ The tank temperature stratification was monitored over a 3 hour period as shown in Fig 7. As a result of the relatively large wall area in the top of the tank a hot stratified layer develops in the top of the tank as soon as the mantle heating starts. The mantle effectively acts as a stratification promoter and directs a significant part of the heat input to the top of the tank. High heat transfer into the bottom of the tank, due to the large bottom contact area
A thermal model of a ho"rtzontal mantle tank was developed using the correlation proposed by Rosengarten et al (1999b) for heat transfer in the mantle passageway. Convection heat transfer inside the tank was modelled as free convection along a vertical plate. Flow along the inner surface of the tank wall was assumed to be laminar and to stratify without mixing the contents of the inner tank. The heat exchanger model was developed in the TRNSYS package (Kline et al (1996)) as part of the TYPE38 stratified tank routine. The flow in the mantle gap was divided into horizontal segments in line with the segments used to model the inner tank. The TYPE38 tank model in TRNSYS uses a plug flow concept to follow the fluid into and out of the tank, however when a collector loop heat exchanger is included the only flow in the tank is the load flow. The TYPE38 model for the tank was retained for consistency with the well-established open-loop thermosyphon model in TRNSYS. Energy transfer between the mantle and the inner tank was modelled by dividing the tank into equal-mass elements of 2% of the tank contents. The predictions of this model for the steady heat-up tests are shown with the test data in Fig 7. The model required solution time steps of less than 0.2 hours before the predictions were independent of time step. This is less than the 0.5 hour time step commonly used for TRNSYS modelling of solar water heaters. The agreement between the model predictions and the measured data was generally good, although the temperature gradients above the bottom mixed layer differed from the observed data. This error is due to the simplified stratification assumption that is used to model the natural convection process over the inner surface of the mantle wall. The model correctly predicted the initial rapid temperature rise in the top of the tank due to the direction of the hot mantle inlet stream to the top of the mantle during the initial hour of heating. Once the temperature in the top of the tank approached the mantle inlet temperature (after one hour of heating), the flow in the mantle gap could not reach the top of the heat exchanger surface and more heat was directed to the bottom of the tank.
5. THERMOSYHON CIRCULATION IN A SOLAR COLLECTOR LOOP WITH A MANTLE HEAT EXCHANGER A common form of solar water heater is the horizontal-tank close-coupled thermosyphon system shown in Fig 8. The mantle heat exchanger is formed as part of the tank and then covered with insulation. The flow connection to the mantle in Fig 8 is a bottom entry configuration (see Fig 1).
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Fig 8. Close-coupled thermosyphon solar water heater with a mantle heat exchanger in the collector loop. The advantage of bottom connections to the mantle is that heat loss due to reverse flow at night are suppressed compared to systems with top level input to the mantle. A model of thermosyphon solar water heaters with collector loop heat exchangers was developed by Morrison (1994) from the detailed open-loop model in TRNSYS. Bickford and Hittle (1995) compared this model of a solar water heater with measured data and showed that the model over predicted collector energy gain by up to 10%. The model also over estimated the degree of stratification in the storage tank, when it was operated as a preheater. This model has now been extended to include the heat transfer correlation developed by Rosengarten et al (1999b) and improved analysis of conduction down the shallow depth of the tank contents and through the tank walls. In the original TRNSYS horizontal tank model the equivalent conductivity of the tank walls and tank contents was quantified by a single value for the effective conductivity. This is correct for a vertical tank however, for a horizontal tank the effect of wall conduction increases significantly for the top and bottom sections of the tank. Heat conduction in the walls of a tall vertical tank has only a minor effect on thermal stratification however, for a horizontal tank the wall heat conduction has a substantial impact on thermal stratification in the top and section of the tank. The variation of effective thermal conductivity with depth for a 3 mm wall thickness steel tank is shown in Fig 9. As a result of the high wall effect in a horizontal tank stratified conditions cannot be maintained for extended periods in the top 10 to 20% of a horizontal tank. A new model of a thermosyphon solar water heater with a collector-loop heat exchanger was developed in TRNSYS, using the TRNSYS solar collector model (TYPE1), the TRNSYS stratified tank model (TYPE38) and a new heat exchanger routine integrated within the TRNSYS thermosyphon loop model (TYPE45). The heat exchanger model and the TYPE45 thermosyphon collector loop model both allow for the temperature dependent properties of propylene glycol that would usually be used in the collector loop for freeze protection.
Fig 9. Effective thermal conductivity as a function of depth in a horizontal tank. 6. COMPARISON WITH OUTDOOR MEASURED PERFORMANCE The performance of two mantle-tank thermosyphon solar water heaters installed on an outdoor test rig, were monitored over an extended period and compared with the new TRNSYS model. The water heaters were installed on a roof with a 34 ~ slope (latitude angle) and were operated either as a solar preheater or as an integrated system with in-tank electric auxiliary boosting. During the preheater tests the systems were filled with cold water at the start of each day and allowed to operate throughout the day without any draw off. In the early evening the tanks were discharged and the net useful energy collected over the day determined by integrating the heat discharge rate:
, cp
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242
1"able 1, Charaete~ics of close-coupled ~ h o n solar water heaters used in the assesunent program.
50
,j/,,,St
"o40 volume heat loss UA
300 litres 2.8 W/K 450 rnm
dimneter
Mantle gap
5mm 1.8 m
,
Cone~or efficiency System No l, selective absorber.
>30
F sss SS
.~
re -
sssJ Sr
IsSSSJJ~SS
s ~ sss JSS
<_,
=
j~s
"5 10
E
.~
System No 2, black absorber.
= 0 ~ 6 - 7.15(T- T~)/I -
,jss SJSS
0
i
i
20
30
40
50
Measured energy delivery MJ/day
I= incident radiation
T = mean fluid tempenm~ Ta - ambient t e m ~ r e
6.1
i
10
Fig 11. Comparison of measured and simulated daily energy gain in a solar preheater, 300 L tank, 4m 2 collector, system 2.
Pre-heater simulation
Comparison of measured and simulated performance of a thermosyphon heat-exchanger system operated as a preheater is shown in Figs 6 and 7 for systems 1 and 2 respectively. The data for system No 1 includes summer and winter conditions, the data for system 2 is from a single 15 day period in summer. The results in Figs 10 and 11 indicate that the simulation model is able to predict the daily energy gain of a mantle-tank solar preheater with a one standard deviation error of 1.7 MJ. The largest errors occur for high radiation conditions when the system reaches temperatures above 65~ at the end of the day and normally minor effects such as pipe losses start to become significant. 6O
~, 40 >=
~
>, 30
0 0
10
20
30
40
50
Measured energy delivery MJIday
Fig 10. Comparison of measured and simulated daily energy gain in a solar preheater, 300 L tank, 4m 2 collector, system 1.
60
6.2 Boosted system simulation
The two close-coupled, mantle-tank thermosyphon systems were operated as integrated systems with the auxiliary booster activated so as to provide a continuous hot water service. The systems were installed on an outdoor test rig and operated with simulated domestic load conditions over a period of 18 months. The systems were identical except for the collector quality described in Table 1. The auxiliary booster was located in the middle of the tank and controlled by a thermostat set at 65~ The auxiliary input was energised at all times under the control of the thermostat in the middle of the tank. A constant daily load distribution was used throughout the year however, the daily load was varied each month to simulate domestic load conditions in a temperate climate as defined in Australian Standard AS4234 (1994). The simulation model used the measured load volume and cold water temperature at each load interval as inputs. The thermostat temperature was determined from the temperature of the morning load periods. The simulation results for daily auxiliary energy use of system 1 are compared with the measured daily energy use in Fig 12 for a six month test period spaning mid-summer to mid-winter in Sydney, Australia. The scatter in the daily results is partially due to slight differences in activation times of the auxiliary heater just before or after the midnight division between days. The one standard deviation error for simulation of auxiliary energy use is 2 M J/day. The monthly average solar contribution relative to a conventional electric water heater supplying the same loads is shown in Figs 13 and 14. The annual energy savings F R was defined as
ISES Solar World Congress 1999, Volume III
F R [3 AuxNs [] Auxs A ux NS
243
(6)
where AuxNs [3auxiliary energy used by a conventional non-solar water heater. Aux s [] auxiliary energy used by the booster in the solar water heater. The simulation of the monthly energy savings over the 18 month test period for system 1 showed a 2 percentage point error however, the energy savings over the full 18 month test period was only 0.5 percentage points of relative energy savings.
Fig 13. Comparison of energy savings of system 1 relative to a conventional electric water heater.
7. CONCLUSIONS
Fig 12. Comparison of measured and simulated auxiliary energy use for system 2 with a boost element in the top half of the tank.
The performance of mantle heat exchangers for closecoupled thermosyphon solar water heaters has been characterised for system configurations with bottom entry and exit ports in to the mantle. A model of a thermosyphon solar collector loop incorporating a collector-loop mantle heat exchanger has been developed in the TRNSYS solar modelling package. The model predictions of stratification development in a solar preheater tank were tested in a controlled laboratory rig and shown to give reliable results for overall heat transfer and the development of stratification in the storage tank. The model was also assessed against outdoor test results for thermosyphon solar water heaters with collector-loop heat exchangers. Predictions of daily energy gain for a solar preheater was found to have an average uncertainty of 1.7MJ/day. Predictions of auxiliary energy use for in-tank boosted systems showed an average error of 2 MJ/day in auxiliary use however, the long term energy savings was found to give very reliable results.
8. REFERENCES
AS4234 (1994). Australian Standard. Solar water heaters domestic and heat pump - calculation of energy consumption. Baur J. M., Klein S. A. and Beckman W. A. (1993). Simulation of water tanks with mantle heat exchangers. Proceedings ASES Annual Conference, Solar93, 286-291.
Fig 14. Comparison of energy savings of system 2 relative to a conventional electric water heater (no test data in January).
Furbo S. and Berg P. (1992) Calculation of the thermal performance of small hot water solar heating systems using low flow operation. Thermal Insulation Laboratory, Technical University of Denmark.
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ISES Solar World Congress 1999, Volume III
Furbo S. (1993). Optimum designed heat storage for small low flow systems. ISES Solar World Congress, Budapest Hungary. V5, 117-122.
Shah L.J., Morrison G.L. and Behnia M. (1999). Characteristics of vertical mantle heat exchangers for solar water heaters. Solar99 ISES Israel 1999.
Klein S.A et. al. (1996). TRNSYS 14.1, User Manual. University of Wisconsin Solar Energy Laboratory.
WATSUN 13.1 1 9 9 2 . Users manual and program documentation. WATSUN Simulation Laboratory, University of Waterloo, Ontario Canada.
Mercer W. E., Pearce W. M. and Hitchcock J. E (1967) Laminar forced convection in the entrance region between parallel flat plates. ASME J of Heat Transfer V89, 251-257. Mertol A., Place W. and Webster T. (1981). Detailed loop model analysis of liquid solar thermosyphons with heat exchangers. Solar Energy V27,367-386. Morrison G. L., Nasr A., Behnia M. and Rosengarten G. (1998). Analysis of horizontal mantle heat exchangers in solar water heating systems. Solar Energy V64, 19-31. Morrison G. L., Nasr A., Belmia M. and Rosengarten G. (1997). Performance of horizontal mantle heat exchangers in solar water heating systems. ISES Bi-annual Conference Taejon Korea, V2,149-158. Morrison G.L. (1994) TRNSYS extensions for Australian solar water heating systems (TRNAUS). Report 1994/FMT/1 Kensington, University of New South Wales, 1994. Nasr A., Morrison G. L. and Behnia M. (1997). A parametric study of an annular heat exchanger with application to solar thermosyphon systems. ICHMT, International Symposium on Advances in Computational Heat Transfer, Cesme Turkey, 299-307. Nasr A., Morrison G. L. and Behnia M. (1998). A parametric study of horizontal concentric heat exchangers for solar storage tanks. J of Computer Modeling and Simulation in Engineering. V3, 269-274. Rosengarten G., Morrison G. L. and Behnia M. (1997) Understanding mantle heat exchangers used in solar water heaters. Australian and New Zealand Solar Energy Society, Solar97 Conference. Rosengarten G., Morrison G. L. and Behnia M. (1998) Mixed convection in a narrow rectangular cavity with application to horizontal mantle heat exchangers. 11th International Symposium on Transport Phenomena, The Pacific Center of Thermal-Fluids Engineering Taiwan 126-131, Rosengarten G., Behnia M. and Morrison G. L. (1999a) Some aspects concerning modelling the flow and heat transfer in horizontal mantle heat exchangers in solar water heaters. International Journal of Energy (in press). Rosengarten G., Morrison G. L. and Belmia M. (1999b) Mixed convection in a narrow rectangular cavity with bottom entry and outlet. Submitted to Int J Heat & Mass Transfer. Shah L. J. and Furbo S. (1998). Correlation of experimental and theoretical data for mantle tanks used in low flow SDHW systems. Solar Energy V64,245-256.
Webster T., Coutier J., Place J. and Tavana M. (1987). Experimental evaluation of solar thermosyphons with heat exchangers. Solar Energy V38,219-231.
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BRIDGING THE GAP: RESEARCH AND VALIDATION OF THE DST PERFORMACE TEST METHOD FOR CEN AND ISO STANDARDS - Project
Results
-
Daniel J. Naron and Marinus Rolloos Renewable Energy and the Built Environment, TNO Building and Construction Research, P.O.Box 49, NL-2600 AA, Delft, The Netherlands, Tel: +31 15 2695249, Fax: +31 15 2695299, e-mail: [email protected]
M. J. Carvalho INETI, Estmda do Paso do Lumiar, 1699 Lisboa Codex, Portugal Tel: +351 1 716 2712, Fax: +351 1 716 3797, e-mail: [email protected] Abstract - The Dynamic System Testing (DST) method is one of the methods chosen in the preliminary CEN/TC 312 European quality standards [1] to measure the energetic perfonmnce of Solar Domestic Hot Water (SDHW) systems. These European standards make a reference to the DST procedure as defined in the Draft International Standard ISO/DIS 9459-5 [2]. The European SMT project 'Bridging the Gap - Research and lntereomparison on the DST test method' was started to ftn~er research, improve and experimentally validate the DST performance test method, in order to further support use of the method in CEN and ISO standards. The project objectives have been structured into three Work Packages (WP):
Work Package 1: Definidon of Scope Clear demarcation, definition and widening of the scope of the test method, allowing for as many systems and conditions as possible. Fine-timing of the present description of the procedure.
Work Package 2: Con~arison with CSTG method Comparison of the DST method to the CSTG method [3], leading to correspondence factors suitable for comparing DST results with CSTG results.
Work Package 3: ~ m e n t a l
validationprogramme
Experimental validation of the DST method; intercomparison tests in a number of recognised laboratories throughout Europe. This paper covers the final results achieved in Work Package 1, and the preliminary results of Work Package 2 and some preliminary results of Work Package 3.
1.
INTRODUCTION
Performance test methods for solar domestic hot water (SDHW) systems provide designers, manufacturers, installers and users with information how to represent, how to measure and how to compare the thermal performance of these systems. Suitable SDHW test methods must be able to predict a (reliable) long term (annual) performance atter a (some) short term measurements in order to be broadly applicable. The Dynamic System Test (DST) method is one of those methods. After being developed in the IEA task 3 (Solar Heating and Cooling Programme), being worked on in the Dynamic System Testing Group and in IEA task 14 (same programme), the DST method now has the status of ISO Draft International Standard (DIS) [2] and is referred to from preliminary CEN European standards [ 1]. Making the validation of the DST method complete, a project has been approved by the CEC Standardisation, Measurement and Testing Programme (SM&T), in order to 'bridge the gap' towards CEN standardisation.
It is in this SM&T project ('Bridging the Gap') that ten recognised laboratories throughout Europe have performed: 9 simulations in order to define the scope of the DST method (Work Package 1). 9 a comparison with the CSTG method; the other test method described in CEN (Work Package 2). 9 experimental validation programme (Work Package 3) In this paper the final results of WP1 and the preliminary results of WP2 and WP3 of this SM&T-DST project are reported. 2.
DST TEST METHOD
In dynamic system testing, a mathematical SDHW model is used in order to collect as much information as possible from the available measuring data [4]. The measuring data are being obtained by a series of short outdoor tests on a SDHW system. A dynamic computer model in which this SDHW mathematical model is implemented is used for 'parameter identification', which characterises the SDHW system (being tested) in terms of model parameters (see table 1). In order to predict the annual thermal performance (energy saved by the SDHW system), these identified parameters and this (general) SDHW computer model are used. In figure 1 a schematic overview of the DST procedure is outlined.
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J symbol
unit
'physical meaning'
Ac*
m2
Uc*
Wm-2K-1
Collector loop heat loss parameter
Uv
jm-3K-1
Wind velocity dependency of Uc* (if applicable)
Us
WK -1
Overall heat loss coefficient of the store
Cs
MJK-1
Heat capacity of the store
Effective collector area
Auxiliary fraction of the store (if applicable)
i
DL
Draw-off mixing parameter
Sc
Collector loop stratification parameter
RL
KkW-1
Thermal resistance of load side heat exchanger (if applicable)
"Table 1: List of model parameters, characterising a SDHW system.
Outdoor Tests on the SDHW system
Data Processing using .qeneral SDHW model
Parameters describing characteristics of tested SDHW system
Performance Calculation using the ueneral SDHW
Yearly Performance Prediction for specific climate and load conditions
Outdoor tests: S~sol,a;S~sol,b;S.~store;S~aux
Fig. 1: Schematic outline of the DSTprocedure on SDHW systems. One of the advantages of the method is the 'black-boxapproach', meaning that there is no need for internal
measurements or special knowledge of the system. More important, very different kinds of SDHW systems can be tested with this same method. Furthermore, the DST test results are independent on the location being tested and the performance of the SDHW system can be predicted for any climate and load profile. 3.
WORK PACKAGE 1: DEFINITION OF THE SCOPE
The objectives of WP1 are: Fine-tuning of the present description of the DST procedure to an expected reproducibility of 5 to 10 %, with predictions for different climates and hot water loads. Clear demarcation, definition and widening of the scope of the test method, to allow for as many systems and conditions as possible.
3.1 Overview of work done The most efficient way to systematically investigate the boundaries of ~ application of the DST method is to use 'Simulated Test Data' for various sites and systems. The idea is to replace the data obtained from a real DST (outdoor) test on a SDHW system by a set of simulated test data. These simulated test data are generated with a detailed simulation model of the SDHW system. The detailed model is run on real climatic data under the same conditions as they would occur in a real test, this will generate data files that are equivalent to real measurement data files. These simulated test data files are then used as input for the DST data processing (see figure 2).
Real climatic data
Detailed simulation model of the SDHW system
Simulated test data (to be used DST predictions) Fig. 2: Producing simulated test data The principal advantage is ~ a h'e~' performance of the simulated SDHW system can be calculated, by running the detailed simulation model on one whole reference year. This enables c o ~ of the DST p e r f ~ prediction against this Year calculated performance, to find absolute errors in the DST predictions. This much quicker and cheaper way of producing 'measurement data' means that much more situations can be assessed and the boundaries of the scope of the DST method can be defined much clearer.
ISES Solar World Congress 1999, Volume III
3.2 Calculations The three participants in WP 1, TNO (Netherlands), DTI (Denmark) and ITW (Germany) and the fourth voluntary partner INFA-Solar (Germany), have investigated the boundaries and limits of the DST test method [5]. With the 'simulated test data approach', the precision of the DST method was verified in an absolute way. The following issues have been investigated: 9 SDHW system type; Forced circulation systems, thermosyphon systems or ICS system High flow or low-flow systems Preheat or auxiliary heated systems 9 Collector type Flat-plate black -, flat-plate spectral selective - or evacuated tube collectors with two-phase heat transfer Incident angle dependency of the collector efficiency 9 Store influences Vertical heat conductivity in 'horizontal' store Worst-case analysis: strongly non-linear heat losses o f the collector and low store temperatures during Ssol,b sequence 9 Climate Testing climate (i.e. Stockholm - northern climate, Brussels (Uccle) - central/marine climate, Davos - mountain climate, Athens - mediterranean climate) Climate for performance prediction Testing season (winter, spring, summer, autumn) Hot water load Wind dependency The influence of the auxiliary heater set temperature Error analysis: influence of systematical errors on sensors used for DST measurements Extrapolations of the DST result into performance predictions of identical but differently sized systems 3.3 Results It was found that for almost all considered cases, the DST procedure as described in [2] will lead to precise test results. These are well within +10% of the solar fraction for different testing and prediction climates, hot water demands and for the system types common on the European market. Those cases where this precision was not reached, have been isolated and analysed which has led to several conclusions. From some of these conclusions suggestions for improvements in the DST procedure can be formulated; other conclusions have led to a clearer definition and demarcation of the scope of the test method. The following results are recommended to take into consideration in order to demarcate the range in which the method can be accurately used.
247
Low-flow systems Simulated tests on (extreme) low-flow systems showed that DST can handle these systems very well, especially when the solar heat exchanger or manifold is extended over the total length of the store.
Incident Angle Modifier In general the absorption and transmittance of a collector cover depends on the angle of incidence of the radiation. This effect is not taken into account by the DST method. A common way to express the Incident Angle Modifier (IAM) is the Ambrosetti equation: Typically the incident angle modifier coefficient of a single
glazed flat-plate collector is in the range from r = 0.25 to 0.4. Testing in spring and autumn of systems with higher r-values (r>0.4) will lead to higher annual performance predictions. 9 Simulations show that for r < 0.4, the errors were satisfactory. 9 DST give systematic errors for SDHW systems with strong IAM (r > 0.4). However, when the irradiance is corrected for the IAM during data processing and performance prediction, these systematic errors disappear completely. Therefore the scope of the DST procedure can be broadened to include systems with 'strong' IAM, provided that their IAM is determined and used for correction of the irradiance during both the parameter identification and performance prediction.
Fig. 3: Correction for incident angle dependency in DST
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Overheating during testing Sensitivity to extrapolations *
9 9
Extrapolations of the (design load) heat demand, ranging from 50% up to 200%, sometimes lead to increasing errors. These errors stay always precise to within • at the 95% confidence limits. Extrapolation to other climates than the test climate leads to no effect on the prediction error. Extrapolations of DST predictions to performance predictions of identical, but differently sized (collector, store) systems seem to be precise.
Conductivity effects Several simulations have been performed on stores with different conductivity behaviour from the atvdliary part downward to the solar heated part. DST does not explicitly take this effect into account, but adapts the value of the parameter f ~ to simulate a larger part of the store being heated by the atvdliary heating.
3.4 Recommendations for improvement of DST The following changes are recommended in order to improve the accuracy of the DST procedure.
Negative system outputs In the DST DIS [2], it is said that during performance prediction any negative system output (cooling instead of heating of the water) is to be ignored. However for systems where the store is located outdoors, this is not physical: cold winter nights will allow the store temperatme to go down (= negative system output). According to TNO calculations for ICS systems, ignoring these negative system outputs leads to larger prediction errors, than when the negative system outputs are included. Therefore it is recommended that these negative outputs are no_.~tignored and that the standard text is changed to reflect this.
Wind dependency of heat losses The heat losses of a collector depend on the air velocity around the collector. Not taking into account this wind dependency (although the DST model describes a special, optional parameter uv quantifying this dependency), this sometimes lead to underpredictions in cases for strongly wind dependent collectors. With respect to the treatment of the air velocity (wind) surrounding the collector, it is recommended to use the option Wf,,~ (of the DST software) for SDHW systems with glazed collectors. Investigations on the wind speed during testing showed that for collectors with spectral selective absorbers a difference between mean wind velocity during testing and yearly mean wind velocity for the location of the performance prediction of +_2 m/s can be accepted. For collectors with black absorbers either a maximum difference of +1 m/s either the option W~ should be required. Note: This recommendation implies that all SDHW systems should be tested with a forced air velocity (fans) around the collector.
When the overheating protection mechanisms of the SDHW system are activated during the test, this destroys the precision of the test. Therefore this must be avoided at all times. Small improvements in the testto achieve this are suggested.
Conductivity effects; H/D Ratio of Store It is recommended to limit DST testing of SDHW-systems with auxiliary heating to a certain minimum ratio Store Height / Store Diameter (H/D) I. Note 1: It is the (solar) system configuration which requires this; this is not a limitation of the DST method. Note 2: This recommendation implies that thermosyphon and ICS systems with a horizontal store should be tested as 'solar-onlysystems'.
Auxiliary set-point temperature during testing For systems with integrated auxiliary, the present procedure requires the auxiliary set-point temperature to be set at 45 ~ for Ssol,b days and at 80 ~ for Ssto~ and Saux sequences. It is probably better to require the same set-point temperature in all sequences (close to the temperature used in real operation). Calculations suggest that switching on the auxiliary is only required in the Saux sequence, and is not necessary for the Ssotb sequence.
Changing the 12 MJ/day requirement for 'valid' test days For systems with high heat losses and test periods with irradiation just over 12 M J / m 2 for valid Sso~b days, a (certain) threshold temperature is sometimes not reached. This may lead to insufficient variability in the test data and thus to precise results. These high uncertainties can be prevented by testing according to the following extra requirement: "If the temperature of the water withdrawn in a S, ol.b sequence is always below a threshold temperature (to be specified later), the sequence shall be extended until two consecutive days with higher irradiation (for instance 15 MJ/m e) are included'. This will guarantee sufficiently that the system reaches higher temperatures and the heat loss parameters are fitted with sufficient precision. Note: This will occur only occasionally, so normal testing duration is not affected. Very cold test periods Tests performed during very cold circumstances (average outdoor temperature below or around O~ during testing) may result in higher prediction errors, especially for systems with high heat losses (systems with non-selective absorbers, uncovered collectors). R is recommended that for such systems, a lower limit is required on the outdoor temperature during testing.
1TO be specified later in this project
ISES Solar World Congress 1999, Volume III
Sensitivity with respect to sensor errors It has been extensively investigated how sensor errors would affect the DST results. 9 The DST method is insensitive to random errors in the sensor readings. 9 DST is sensitive to systematic errors in sensor readings 2. It appears that the systematic errors allowed in the DST/DIS 9459-5 [2] still can lead to deviations up to +5-6 % (absolute) at 95% confidence limits. Therefore it is recommended that the allowed sensor deviations in the DST/DIS should be reviewed to see if they could be formulated more strictly. Reducing DST parameters The DST model seems to be overdetermined. Testing (some) SDHW systems lead to parameter values which are either negligible, e.g. Sc=0, DL=0, or highly cross-correlated, e.g. Uc* with Us. It might be necessary to reduce the amount of parameters.
3.5 Conclusions Work Package 1 Work Package 1 has been very succesfull, because it has resulted into improvements and a clear demarcation of the test metho& The accmacy has been found mostly within :L5% and always within +10% (if the improvements have been implemented). 4.
WORK PACKAGE 2: COMPARISON W I T H CSTG M E T H O D
The objectives of WP2 are: Comparison of the DST method with the CSTG method, which is also used in the CEN, leading to correspondence factors which enable comparison of DST results with CSTG results.
4.1
Overview of work done
The four participants 1NETI (Portugal), CSTB (France), NCSR (Greece) and UWCC (UK) have been carried out both DST tests as well as tests using the CSTG method on SDHW systems [6]. 9 INETI tested two thermosyphon systems, each with nonselective fiat plate collector. During the tests fans have been used in order to take care of wind influences. 9 CSTB tested one ICS system with double glazing and tubular tank (surrounded by minors). 9 NCSR tested two forced circulation systems, one with a fiat plat collector, one with an evacuated tube collector. 9 UWCC tested one ICS system; only a CSTG method, both indoor and outdoor. TNO has carried out simulations according to the 'simulated test data approach'. The data analysis has been carried out by TNO and INETI.
2 This recommendation applies not only to the DST method but also to similar methods that use short-time measurements to predict yearly energy gain: ISO 9459 part 2 and ISO 9459 part 3.
249
4.2 CSTG test method The SDHW systems, described above, have been tested according to the CSTG method, described in [3]. Also the system characterisation has been done using the CSTG method in order to determine the three parameters and their associated uncertainties of the model:
Q = a 0 + a 1 +a2(tamb(av ) +tmains) Where: Q [MJ] H [MJ] t [~
(2)
Daily energy gain from SDHW system Daily solar irradiation on collector Temperature (ambient / mains)
In ISO 9459-2 (1995) a calculation method for estimating the system long term performance has been formulated. The calculation procedure includes two load patterns: 9 Load pattern 1: Load determined by the volume of daily hot water consumption 9 Load pattern 2: Load determined by a minimum useful temperature limit for the hot water consumption; when the outlet store temperature is lower then this minimum value no water is extracted from the store. Using these load patterns it is not possible to compare the CSTG outcome with the DST outcome. In order to be able to compare the CSTG result with the DST result, a load pattern 3 must be defined: 9 Load pattern 3: Load determined by maximum energy needed (see also [7]). Load pattern 3 is in agreement with the Reference conditions of prEN 12976-2 (1997) [1]. One will use load pattern 3 comparing DST with CSTG
4.3 DST test method Because the CSTG test method cannot deal with auxiliary energy, the DST test must be carried out with (possible) auxiliary heater switched off. The systems were tested according to DST test method, i.e., ISO/DIS 9459-5. The test sequences performed for comparison with CSTG test method were: 9
Ssol, a
9
S~o1,bauxiliary off
9
Sstor c
The parameter identification as well as the LTPP has been carried out using the Insitu Software Package - version 2.7a.
4.4 Comparison of test results Based on the values of LTPP done according to DST and CSTG test methods, a comparison was possible. The percentage difference between the results of the two test methods is calculated according to:
Difference = Qosr - Qcs~ x 1oo (%)
Q~
(3)
In the case of thermosyphon and ICS systems, the DST values are almost always higher than the CSTG values. The differences are not higher than 15%. Values between 5% and 13% are observed for one of the (2) thermosyphon systems and ICS system. For the other thermosyphon system differences are lower than 5%.
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Figure 4 gives a graphical representation of the predicted annual thermal performances (using CSTG and DST) of those systems mentioned.
Parameter adoption of CSTG method For the determination of the model's parameters al, a2 and a3, the SM&T group recommends to use Weighted Least Squares Regression in order to take into account the uncertainty of the measurements of Q, t~y) and tm,~.
The determined uncertainty of parameter ~ gives guidance for the adoption of a simplified model: Q = alH + a2(t~,d~y tmains) i.e.
If a3 < t (V,95%) ~a3 ~
Fig. 4: Yearly thermal performance of thermosyphon and ICS systems predicted according to the CSTG and DST method. The results obtained for forced circulation systems will not be reported here; more details can be found in w Simulations.
4.5 Simulations Simulations have been carried out on an ICS and a forced circulation system, using the 'simulated test data approach'. These simulations have led to the following findings: 9 The discrepancies (differences between CSTG and DST) are small (within 5%). 9 Also CSTG does not take incident angle dependency into account. In order to have a proper CSTG-LTPP one has to correct for strong incident angle dependency (r>0.4 in equation (1)). 9 DST predicts a slightly higher thermal performance of a SDHW system than CSTG.
4.6 Recommendations to CSTG (1SO 9459-2) The following adaptations concerning CSTG method are recommended: Procedure of LTP prediction of CSTG method The group recommends a change in the LTP calculation procedure on point 9.4 of ISO 9459-2 standard. The energy delivered by the system must not exceed the energy demand given by:
Energy demand = VioadPw Cr~ (tload- tmai~ instead o.o.o~thepresent defined draw ~ in ISO 9459-2 Validation of the test method done earlier already considered this type of load. Simplification of the CSTG method Tests show that there is no need to perform test for determination of the 'mixing draw-off profile' The normalised draw-off profile is recommended to be used in all corresponding equations in LTP calculation.
the simplified model can be adopted. t (v,95%) is the 'student distribution' for v degrees of freedom and 95% confidence level. The uncertainty of parameter a9 (determined from either the complete model or simplified model) gives guidance to the need of more test days for a good determination of this parameter. i.e. 9 If a2 < t (v,95%) 6~2 ~ more test days are needed Resulting in a2 > t (v,95%) 6a ;this will correspond to a larger value range for (t~day)- tmm~). 5.
WORK PACKAGE 3: EXPERIMENTAL VALIDATION
The WP3 objective is: Experimental validation of the DST method by means of intercomparison tests in a number of ten recognised laboratories throughout Europe.
5.1
Overview of work done
The participants of WP3 have carried out DST tests in order to fia~er validate the DST test method. A second goal op the tests in WP3 is a field check on the results coming from WP1. The tests have been carried out on the below mentioned systems. 9 1NETI (Portugal) have been carried out DST tests on a thermosyphon system (with non selective flat plate collectors) 9 NCSR (Greece) tested a thermosyphon system and a forced circulation system; both flat plate collectors; both stores have a electrical auxiliary heater installed 9 SPF (Switzerland) carried out three DST tests. Two tests on one same ICS in order to detect (possible) seasonal influences. One other DST test have been carried out on another ICS (tested by CSTB (France) in WP2). These two ICS systems were preheat systems 9 DTI (Denmark) has been carried out a DST test also on an ICS; the same ICS SPF has tested earlier. 9 FGH (Germany) has been tested one thermosyphon (preheat) system. 9 SP (Sweden) has tested a forced circulation system (which has been tested by NCSR before) and a thermosyphon system. The thermosyphon system has been tested by FGH before. 9 Infa Solar (Germany) has tested a forced circulation system. This is a low-flow system with a external load side
ISES Solar World Congress 1999, Volume III
heatexchanger. An auxiliary heater could be installed directly to the storage tank. ITW (Germany) has carried out four DST tests. Two tests on one same thermosyphon system (again) in order to detect possible seasonal influences. One test on a thermosyphon system tested before by 1NETI and one test on a forced circulation system earlier being tested by TNO (The Netherlands). TNO has also carried out four DST tests. One test on the thermosyphon system earlier being tested by INETI (for WP2). One test on the forced circulation low-flow system (Infa Solar) One test on the forced circulation system with evacuated tubes earlier tested in WP2 by NCSR (see figure 5). one test on a forced circulation drain-back system; this system has been shipped to ITW afterwards Another two systems, a thermosyphon system and a forced circulation system have been validated back in 1996 by TNO and ITW. These result will also be used for this Work Package.
9
9
9 9 9
5.3
251
Because of correlation between the Uc*-parameter and the Us parameter, the collector loss parameter (Uc*) can (but does not need to) be omitted. Thermosyphon systems being tested as solar only systems, show a good mutual agreement tested at the two different laboratories. One can detect a significant Sc parameter only when thermosyphon systems are concerned. LTPP predicted for cold climates might be questionable because of possible freezing problems. Thermosyphon systems (with electrical auxiliary) as well as store in which the electrical auxiliary take care of (extreme) mixing of the store will decrease the solar gain enormously. DST can detect this effect very well. Recommendations
Mounting influence The mounting of a solar system is not stated clearly (enough) in CEN (or ISO). This can affect the test result and therefore the long term performance prediction.
Load Side Heat Exchanger It is recommended to exclude Solar Domestic Hot Water Systems with an external load side heat exchanger from the scope of ISO 9459-5 until sufficient experience is available for these system types.
Auxiliary power for performance prediction It is recommended to amend the specifications for the auxilia~-heater-thermal-performance to be used for Long Term Performance Prediction in the reference conditions of prEN 12976 and 12977 (table B 1): A d d ~ change at r e m a r k s : . . , i f not specified otherwise by the manufacturer. Change: '100 + 30 Wattper litre. . . ' into "150 • 50 Wattper litter... " Fig. 5 : W P 3 test on an f o r c e d circulation system with an evacuated tubes collector
5.2
Preliminary WP3 Results
Although not all participants have finalised their reports testing at the time of writing this paper, some preliminary results can be formulated. 9 The DST test procedure give good results testing preheat systems (ICS, thermosyphon as well as forced circulation systems). 9 DST is good in predicting a SDHW system with internal auxiliary heatexchanger. 9 As stated in WP1, DST has difficulties to predict the performance of an ICS with a high heat loss coefficient of the store. 9 The method to correct for incident angle dependency (e.g. for ICS systems) works out well. Assumed that the incident angle dependency is well know, this is however at this stage a potential for errors; it is recommended to integrate this into the DST software.
Energetic Performance Representation The (DST-) data processing software uses Watt as the unit for energetic performance representation per given time interval (e.g. one year). The respective prEN specifies that MJ/year should be used (to report) the energetic performance. It is suggested that in the future, either one of these units will be used.
Component Testing Experimental investigations within Work Package 3 of this SM&T project showed a very promising agreement between the component test method (CTSS), prEN 12977 and the DST method. 6.
CONCLUSIONS
This project 'Bridging the Gap' has been divided into three work packages. Three general conclusions are:
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WPI: Definition of the scope The accuracy and the reproducibility of the DST method is considered as good. A statistical analysis of the results has led to an accuracy of • an in critical cases up to • of the solar fiacfon for different testing and prediction climates, hot water demands and for the system types common on the Etwo~an market. Those cases where this accmacy was not reached, have been isolated and analysed which has led to several conclusions, leading to suggestions for improvements in the DST procedure. WP2: Comparison with the CSTG method Based on the values of Long Term Performance prediction done according to the DST and CSTG test methods, the comparison between the two methods showed, for thermosyphon and ICS systems, differences on an average 78%; those differences lead up to 15% for more critical cases. Simulations lead to promising results, showing that CSTG and DST appear to be comparable. WP3: Experimental validation The comparison of results within the validation programme confirms the accuracy figures obtained by theoretical investigations in Work Package 1. This accuracy is on the average 5% and leads up to 10% for problematic cases. However, a series of tests have also led to significant and unacceptable discrepancies between test results. The analysis of these cases is not finished so far. Therefore the SM&T-group can only give an preliminary assessment of the DST method.
REFERENCES
[1] [2]
[3]
[4]
prEN 12976-2 (1997), Thermal solar systems and components. Factory made systems - Test Methods. ISO 9459/DIS-5 (1996), 'Solar heating - Domestic water heating systems - Part 5: System performance characterization by means of whole-system tests and computer simulation'. ISO 9459-2 (1995), 'Solar heating- Domestic water heating systems - Part 2: Outdoor test methods for system performance charaeterisation and yearly performance prediction of solar only systems'. Naron, D. J. Van der Ree B. G. C. M. Rolloos (1998) Bridging the Gap: Reasearch and Validation of the DST Performance Test Method for CEN and ISO StandardsProgress and Preliminary Results -. In Proceedings of EuroSun 98, September 14-17, Portoro~ Slovenia, Goetzberger A. and Krainer A. (Eds), pp. 111.3.10-1 111.3.10-7, The Franklin Company Consultants,
Birmingham. [5] [6]
[7]
Naron, D. J., Van der Ree, B., (1999) 'Bridging the Gap' Final Report of Work Package 1: Definition of Scope, TNO, Delft, The Netherlands. Carvalho M. J.,Busearlet C., Marshall R., Mathioulakis E. Van der Ree B. (1998) Factory Made Systems Thermal Performance: Comparison of Test Methods. In Proceedings of EuroSun 98, September 1417, Portoro~, Slovenia, Goetzbe rger A. and Krainer A. (Eds), pp. 1/1.3.2-1 - 111.3.2-6, The Franklin Company Consultants, Birmingham. Carvalho M. J., (1998) Interim Report: W P 2 Comparison with CSTG test method, INETI, Lisbon, Portugal.
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253
R E S E A R C H ON A N E W TYPE OF HEAT PIPE VACUUM TUBE S O L A R WATER HEATER Zhu Ning and He Zinian Beijing Solar Energy Research Institute, No.3 Huayuan Road, 100083 Beijing, China Tel: 86(10) 62001022, Fax: 86(10) 62012880, E-mail: [email protected]
ABSTRACT Anew type of heat pipe vacuum tube solar water heater that can be placed at a very small tilt angle is introduced. A series of experiments was made to test the new system, which indicates that the new system has the same performance as conventional one. A further theoretical analysis of this phenomenon as also made in this paper.
1. INTRODUCTION
2. SOLAR WATER HEATER DESCR/PTION
It is well known that heat pipe vacuum tubes have been used in solar water heater for many years owning to their advantages of anti-freezing, anti-corrosion and low heat loss. However, the tubes are always placed with a tilt angle of more the 15 ~ in north-south direction. This is because it is regarded that the gravity-assisted heat pipe can only work with a minimum tilt angle of 15~ and the condenser must be higher than the evaporator, which obviously limits the application of this type of solar water heater. A new type of heat pipe vacuum tube solar water heater is investigated which can easily installed both in south wall or in the outside surface of balcony and can work well with a very small tilt angle (o._<_30).
The new solar water heater consists of heat pipe vacuum tube, storage tank, rack and circulation pipes, Figure 1 shows a layout of the system. The solar collector connects the storage via circulation pipes to form a natural circulation system. The solar collector fixed on the rack is hung along the outer wall of the balcony in the east-west direction to fit the structure of building, while the storage tank is installed inside the balcony even inside the house to reduce heat loss in cold weather.
Fig. 1. Layout of the new solar water heater system
2.1 Heat pipe vacuum tube Heat pipe vacuum tube is mainly composed of heat pipe, absorption plate, glass tube and metal sealing cover, as shown in Figure 2. Heat pipe is also composed of evaporator and condenser. The evaporator is placed inside the glass tube to draw heat from absorption plate; the condenser releases thermal energy to heat water in circulation pipe.
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Fig.2. Layout of heat pipe vacuum tube It is usually demanded that the heat pipe vacuum tube should be placed at a 15~ tilt angle to guarantee the water formed in condenser flow back to evaporator, otherwise heat pipe can not work. How can the heat pipe in this new solar water heater work in such a small tilt angle a? It is because the author changes the heat pipe's structure. The condenser is bent upwards to form another slope angle 13 (I3--350) in the system, Figure 3 shows the layout. The liquid pillar formed in the
condenser and the slight inclination of the evaporator jointly made the heat pipe work well. Under this condition, the condensate flows along the bottom surface of the pipe (puddle pumping) instead of being distributed uniformly over the wall surface as in the case of vertical and high-tilt operation. And it is also because the heat flux transferred by the heat pipe is little comparing with conventional one.
Fig.3. Layout of the bent heat pipe 2.2 Heat conduction unit
In order to connect condenser of the heat pipe vacuum tube and circulation pipe, the authors design a special heat conduction unit as shown in Figure 4.
Fig.4. Layout of heat conduction unit
ISES Solar World Congress 1999, Volume III
3. EXPER/MENT RESULT AND DISCUSSION 3.1 Single tube test Several single vacuum tubes were tested either at a tilt angle of a=3 ~ and a slope angle of 13=35~ in east-west direction or at a tilt angle of a=40 ~ in north-south direction. Each tube heated a small box filled with 2kg water. In 1.5 hours, each box's water was heated from 20~ to 75~ around. The result indicates they have the same thermal performance. 3.2 System test According to Chinese national standard (GB/T 12915-51), the daily efficiency and heat loss coefficient of the whole system were tested. Solar irradiance is measured by an EPPLEY
255
Model PSP Pyranometer placed at the same tilt angle with the absorption plate in vacuum tube. The pyranometer has linearity of_+0.5% and temperature dependence of +1%. Five thermocouples were installed inside the storage tank in proportion. The accuracy of the thermocouple is _+0.1~ A DATATAKER Model DT 600 data acquisition device conducted data acquisition for solar irradiance, water temperature within storage tank and ambient temperature. The new solar water heater is composed of 5 pieces of vacuum tubes with a absorber area of 1.13m2, water capacity of 59.1kg. Table 1 gives the solar isolation, water's temperature variation and ambient temperature on Aug. 26, 1998.
Table 1 Solar isolation, water temperature and ambient temperature on Aug. 26, 1998 Time 8:30 9:30 10:30 11:30 12:30 13:30 14:30 15:30 16:30
Solar isolation(MJ/m 2) 0.00 1.92 4.76 8.37 11.85 15.13 18.20 20.57 21.99
Water Temperature(~ 17.0 18.5 23.6 30.8 38.8 46.6 53.6 58.8 61.3
The daily efficiency can be calculated using these data.
M.Cp .(Te -T~) Od =
A.H
rid --daily efficiency M--water capacity Cr--specific heat Te----fmal water temperature Ts----initial water temperature A--absorber area H--solar isolation The calculated daily efficiency of the new solar water heater is 45.9% and the average heat loss coefficient is 0.766W/m2~ which is the same as the conventional one. Further tests, such as the one to find the best combination of the tilt angle a and the slope angle 13 and the one to determine the best liquid charge amount are lett for future work.
4. CONCLUSION Both experiments and theoretical analysis prove that heat pipe vacuum tube can work well in the new way mentioned above especially in the case of low even heat flux on the evaporator. The result can make solar water heaters integrate with building much better than before.
REFERENCE Yasuhiro Kamotani, Performance of gravity-assisted heat pipe at small tilt angle, 1977
Ambient Temperature(~ 24.1 28.1 31.6 33.5 32.7 32.9 33.0 32.1 31.7
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Solar Process Heat : Distillation, Drying, Agricultural and Industrial Uses Brian Norton Centre for Sustainable Technologies University of Ulster Newtownabbey BT37 0QB N. Ireland E-mail [email protected]
Abstract - The major applications of solar thermal conversion are discussed. Solar process heat systems are now mature technologies. Solar detoxification and advanced solar-powered manufacturing processes are the subject of research. The goal of much present research is to reduce costs by technical refinement and the use of new materials. The history of solar industrial and agricultural process applications is presented together with the principles underlying each generic application and practical examples described.
1
Introduction
A solar energy thermal conversion system comprises conceptually a solar collector, intermediate heat storage and a means of conveying the collected heat between these and to the application. The solar collector is usually selected in terms of outlet temperature, as illustrated in Figure l, matched to the process heat requirement. Where and/or at times when the collector outlet temperature is less than that at the process inlet, additional heating is provided. Even where feasible technically it may be not be economically optimal to seek solar fractions of unity. Many solar energy applications involve sensible energy storage. Only occasionally have latent or chemical energy storage systems been considered. For processes that continue during the night or during periods of insufficient insolation, providing auxiliary heating has been frequently found to be more viable economically than providing sufficient energy storage to enable solar fractions close to unity to be achieved. An exception is the concomitant solar energy collection and storage provided by a non-convecting solar pond. Auxiliary heating is also required where the magnitude and duration of the direct component are insufficient to render feasible a concentrating solar energy collector providing directly the desired application temperature. The availability of direct insolation also delineates the geographic locations for applications predicated on the use of concentrated solar energy. In general vacuum tube collectors are used in medium temperature thermal conversion systems. Non-evacuated inverted absorber concentrators, as shown in Figure 2, (Yadav et al., 1996) have been developed as lower cost and potentially more durable options. Non-imaging compound parabolic trough medium-temperature solar collectors can harness the diffuse components. The inverted absorber compound parabolic concentrator (IACPC) reflects insolation to a downward-facing absorber from which convective heat losses are thus suppressed (Kienzlen et al, 1988; Norton et al, 1991). The IACPC, in
addition to use as water and air heating collectors, has applications in integrated collector-storage water heaters (Tripanagnostopoulos and Yianoulis, 1992), solar distillation and solar cooling (Norton et al, 1997). Solar Hot Water Provision
Solar hot water provision is a mature technology whose design is well understood (Duffie and Beckman, 1991; Norton, 1992). Further adoption, where it is economically viable, is limited largely by non-technical factors such as availability of capital investment, market and trading structures and customer awareness. Solar Thermal Electricity Recent developments in direct steam generating solar collectors, intended for electricity production, will remove the present need when generating steam from a solar thermal system for a collector circuit with heat transfer oil and oil-tosteam heat exchanger (Ahnanza and Lentz, 1998). The adoption of direct solar steam generation will render solar steam production more efficient and possibly more cost effective. Again system control will be an ~ r t a n t issue (Lippke, 1996) however the practical limiting factor to the diffusion of this technology is likely to be the commercial availability of absorber tubes coated with high temperature selective surfaces (Lanxner and Elgat, 1990; Zhang et al., 1998). Solar thermal electricity (in common with other renewable electricity generation) offers lower life cycle greenhouse gas emissions when compared with fossil fuel electricity generation as shown in Table 1 (Norton 1999).
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I
COLLECTOR TYPE
CONCENTRATION INDICATIVE RATIO. C1 TEMPERATURES
i
NAME
FOR DIRECT
SCHEMATIC DIAGRAM
Non-convecting
solar pond
i
>n~ < Z 0 h-<
=
:
or) n~ i,I m n,"
~
, T (K)
INSOLATION
C~ 1
300 < T < 360
o cD
, n~ <
[]
Flat-plate absorber
< _J I,
C.~I
' 300
C~ 1
320
_
L~
Evacuated : envelope n
n
Compound parabolic reflector
Z
i
rY H ~
I
i
13z~0
I , C <S
m
I
0
m
I
_
9
5~C.~15
3/+0
1 5 < C <1+0
31+0
I
o
Parabolic reflector
Or )
I-
X
0
_J
~ m , DI
LLJ __I C9 Z Or )
Fresne[ reflector
O Z ,,,(
,.,Z. li,
,,
/ _
C) < t'Y t--rY < __J O Or)
Parabolic
I0.~C < 1+0
10 <
3z, O ~
T < 5~+0
C 950
31+0
9T < 5~,0
dmsh 100 .~ C < 1000
i
Or)
0
n
{-}
reflector
'
i
-
Cylindrical reflector
x <
<560
<
Spheric al bow[ reflector
~
i
m cF
l
o u~ m <
3z,.O < T
9 1200
l
100 < C < 300
: 3z, O . , T < 1000
I-Z '
O
13.
Heliostat field
@ FIG. 1
100 < C
9 1500
/~00
91"
93 0 0 0
:
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ISES Solar World Congress 1999, Volume III
horizontal - "~ ~ ~ rot\ ~ ". / ~ f'~ % "-/ U;rray\ \/ Transparent / Aperture - - ~ ~ f
/
~
Absorber
I " 9 ( ~reflector I ~ ~ Reflector- I_ "~-" "~l~ ~ [~ ~ / J-Primary
~
| . Primary ~ t~ poro~,c / ~ I reflector ~ Primary~~~..~/ cavity
,/
Tertiary "
/ --
d 'q - - Tertiary J plane W_~__ reflector
' / //" "" ~
/"
cir
~
"-- / ~ /
~ " Secondary cavity Secondary circular arc reflector
FIG. 2
Table 1 Life-cycle Greenhouse Gas Emissions from Renewables and Fossil Fuel Generated Electricity
Energy Crops (g/kWh)
CO2 SO2 NOx
Solar Thermal
Wind
Future Practice
(g/kWh)
(g/kWh)
(g/kWh)
(g/kWh)
30-40 0.08-0.16 1.1-2.6
30-33 0.06-0.08 0.40-0.55
7-9 0.02-0.03 0.06-0.07
98-167 0.20-0.34 O. 18-0.30
20-30 0.50 0.23
6.5-9.1 0.02-0.09 0.02-0.36
Oil Best Practice (g/kWh)
Gas Combined Cycle Gas Turbine (g/kWh)
Coal
Best Practice
4
PV
Current Practice
(g/kWh)
CO2 SO2 NOx
Hydro
Flue Gas Desulphurisation & Low NOx
954.60 11.82 4.34
986.50 1.49 2.93
817.90 14.16 3.99
Water Treatment
Water extracted from streams, rivers and lakes in tropical locations frequently contains pathogens and thus requires disinfection before human consumption. A solar ultra-violet disinfection process, has been developed for this purpose (Wegelin et aL, 1994). However as some bacteria have enzymes that repair their DNA after ultra-violet radiation damage, ultra-violet treated water can only be consumed safely within one day. Solar thermal pasteurisation, in which the water
429.800 0.494
Diesel Embedded (g/kWh)
772.00 1.55 12.30
Nuclear (g/kWh)
6 0.02-0.03 0.06-0.07
is heated to 348 K for ten minutes destroys, rather than damages, all pathogens, does not require the water to be filtered and has fewer operational constraints than chlorination. Though solar pasteurisation is effective and incurs little requirement for maintenance, it has been shown to have higher life-cycle costs than competing options (Burch and Thomas, 1998) further research to reduce initial capital cost is thus essential for the adoption of this potentially very important application of solar energy.
ISES Solar World Congress 1999, Volume Ill Many arid regions have underground brackish water resources or are close to sea water and have high annual levels of insolation. The production of potable water using solar energy has thus been well-researched and, in remote or isolated regions adopted practically. Fundamentally there are three potable water extraction processes that can employ solar energy: (i) distillation in which water is evaporated using solar heat, it condenses separated from its mineral content. This can be at atmospheric pressure in various diverse forms of passive basin stills (Malik et al., 1982) or in a multiple effect process. In a vapour compression system, water vapour is compressed adiabatically producing a superheated vapour. This is first cooled to saturation temperature and then compressed, using mechanical energy, at constant pressure, (ii) reverse osmosis, in which a pressure gradient across a membrane causes water molecules to pass from one side to the other, larger mineral molecules cannot cross and (iii) electrodialysis: a selective membrane containing positive and negative ions separates water from minerals using solar-generated electricity. The optimization of solar multi-effect desalination plant design is very sensitive to the collector cost. El Nasser (1994) has shown that a reduction in collector cost by 50% could result in a decrease in the minimum cost of potable water from US$4.77/m 3 to US$3.03/m3 and minimum water cost is always achievable with an evaporator having the highest performance ratio that corresponds to the largest number of effects possible. The comparative cost of full and partial solar desalination systems was considered in a detailed case study of 20,000 m3/day and 200,000 m3/day systems (Gluerkstem, 1995) using non-convecting salt gradient solar ponds and dual-purpose solar electric power stations. Two desalination technologies were also considered; multi-effect distillation and hybrid multieffect/reverse osmosis. The use of a solar pond to power a desalination system was found cost effective for favourable site conditions. Fully solar desalination systems where found to be more competitive than partial solar systems at very low solar field cost and/or very high electricity prices. Such conditions prevail in remote arid regions due to the low cost of land and distance from grid-connected electricity. However, in such locations it may be challenging to provide effective support for operation and maintenance. The experience gained during construction and operation of the solar distillation plants built on a Greek island has been described by Delyannis and Belessiotis (1996).
Solar Drying Of Crops Natural circulation solar dryers, (Ekechukwu et al) an illustrative selection of which is provided in Figure 3, constructed from wood, metal and glass sheets and used to dry the full range of tropical crops have been evaluated extensively and quite widely used (Brenndorfer et al., 1985). Cassava is a major part of the staple diet in many tropical countries. It contains cyanogenic compounds that must be destroyed in processing for it to be safe for human consumption. The destruction of cyanogenic compounds in cassava during solar drying has been studied by Monroy-Rivera et al. (1996). When three cassava varieties were dried, the results showed that the loss of total and bound cyanides observed varied between 80% and 100% of the initial cyanogenic compound content. The concentration of cyanide residue was invariably less than 40 parts per million on a dry
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basis. Water diffusion into the interior of the chips was found to influence greatly the elimination of cyanogenic compounds.
6 Solar Refrigeration Solar refrigeration is employed to cool vaccine stores. The need for such systems is greatest in peripheral health centres in rural communities in the developing world. Lacking grid-connected electricity the vaccine cold chain can be extended to these areas with autonomous solar energy vaccine stores. Most of this provision is met by compression cycle refrigerators powered by photovoltaics. Solar thermal refrigeration systems have been subject of extensive research (Critoph, 1998) but have, to date, only rarely (see, for example Worsoe-Schmidt, 1984) led to commercially produced systems. 7
Livestock Building Heating
Solar energy may be used for the space heating of agricultural buildings. The guiding principles are similar to the solar space heating of non-agricultm'al buildings: first conserve energy, then adopt passive means of solar energy collection, distribution and storage and only then consider active solar technologies. The use of active solar has been helped where the construction of farm building roofs can be modified readily to hold air-heating solar collectors. Low-cost roof-based airheating solar collectors have tended to one of two alternative specifications: one with transparent plastic film cover over a black plastic or metal absorber, the other fabricated from metal with glass covers. Plastic collectors cannot operate above 330 K and due to ultraviolet degradation have a limited effective operational life, often as short as one year. Farm-built metal solar air-heaters range from unglazed low-temperature units for animal husbandry through to double-glazed medium temperature systems for crop drying (Norton, 1992). The principal attraction of farm-built air-heating collectors is the low initial investment required. The disadvantages are potentially poor fabrication quality control leading to poor performance and the frequent lack of optimised sizing of system components.
8
Conclusion
Many solar thermal applications are viable economically now in favourable climate and use contexts. More would be so if for those applications nearing economic viability the economic externalities associated with potential for CO2 abatement were given tangible value by appropriate fiscal intervention. Solar materials processing technologies require further research and development. Solar thermal process heat delivery though a mature technology still also requires intensive programmes of component improvement and demonstration.
References Almanza R. and Lentz A. (1998) Electricity production at low powers by direct steam generation with parabolic troughs. Solar Energy, 64, 115-120. Brermdorfer B., Kennedy L., Bateman C.O., Mrena G.C. and Wereko-Brobby C. (1985) Solar Dryers: their role in post harvest processing. Commonwealth Secretariat Publications. London. Burch J. D. and Thomas K.E. (1998) Water disinfection for developing countries and potential for solar thermal pasteurisation. Solar Energy, 64, 87-97.
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Critopy, R. E. (1998) Delyannis E. E. and Belessiotis V. (1996) Solar application in desalination: the Greek Islands experiment. Desalination, 100, 27-34 Duffle J.A. and Beckman W.A. (1991) Solar Engineering of Thermal Processes. 2nd Edn. Wiley Interscience, New York. Eames P. C. and Norton B. (1993) Validated, unified model for optics and heat transfer in line-axis concentrating solar energy collectors, Solar Energy, 50, No 4, 339-355. Ekechukwu et al E1 Nashar A. M. (1994) System design optimization of a solar desalination plant using a multi-effect stack type distillation unit. Desalination, 97, 587-618 Gluerkstem P. (1995) Potential uses of solar energy for seawater desalination. Desalination, 101, 11-20
Norton, B. (1998) Renewable Electricity- What is the true cost? IEE Power Engineering Journal, 13, 6-12. Tripanagnostopoulos Y. and Yianoulis P. (1992) Integrated collector-storage systems with supperssed thermal losses, Solar Energy 48, No l, 31-43./ Yadav Y. P., Yadav A. K., Anwar N., Eames P. C. and Norton B. (1996) Fabrication and testing of a line-axis compound parabolic concentrating solar energy collector, Renewable Energy, 9, 572-575. Wegelin M., Canonica S., Mechsner K., Tleisehmann T., Pasaro F. and Metzler A. (1994) Solar water disinfection: scope of the process and analysis of radiation experiments. Journal of Water Supply Research and Technology - Aqua, 40, 154-169.
Worsoe-Schmidt P. (1984) Solar refiigeration at village level based on a solid-absorption cycle. Proceedings of the Solar
Energy Society of India National Solar Energy Convention, Kienzlen V., Gordon J. M. and Kreider J. F. (1988) The reverse flat plate collector: a stationary, non-evacuated, lowteclmology, m e d i u m - ~ a a ' e solar coUector,ASME Journal of Solar Energy Engineering, 110, No 1, 23-30. Laplaze D., Bernier P., Flamant G., LeBrun M., Brunelle A. and Della-Negra S. (1996) Solar Energy: Application to the production of Fullerenes. Journal of Physics B, 29. Lanxner M. and Elgat Z. (1990) Solar selective absorber coating for high service temperatures produced by plasma sputtering. SPIE 1272 Proceedings of Optical Materials
Technology for Energy Efficiency and Solar Energy Conversion/X, 12-13 March, The Hague, Netherlands, 240249. Lippke F. (1996) Direct steam generation in parabolic trough solar power plants: numerical investigation of the transients and the control of a once -through system. ASME Journal of Solar
Energy Engineering, 118, 9-14. Malik M.A.S., Tiwari G.N., Kumar A. and Sodha M.S. (1982)
Solar Distillation.. Pergamon Press, Oxford, U.K. Mills D. R. and Guitronich J. E. (1978) Asymmetrical nonimaging cylindrical solar concentrators, Solar Energy, 20, No 1, 45-55. Monroy-Rivera J.A., Angulo O., Sanchez T. and Lebert A. (1996) Elimination of cyanogenic compounds of cassava &tfing solar drying, Drying Technology, 14, 2371-2385. Norton B., Eames P. C. and Yadav Y. P (1991) Symmetric and asymmetric linear compound parabolic concentrating solar energy collectors. The state-of-the-art in optical and thermophysieal analysis, International Journal of Ambient Energy, 12, No 4, 171-190. Norton B. (1992) Solar Energy Thermal Technology. SpringerVerlag, Heidelberg, Germany. Norton B., Eames P. C., Yadav Y. P. and Gfifliths P. W. (1997) Solar concentrators for rural applications, InternationalJournal
February, Bhopal. Zhang Q-C., Zhao K., Zhang B-C., Wang L-F., Shen Z-L., Zhou Z-J., Lu D-Q., Xie D-L. and Li B-F. (1998) New Cermat solar coatings for solar thermal electricity applications. Solar Energy, 64, 109-114.
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BRACKISH WATER DISTILLATION WITH PLANE MICROPOROUS MEMBRANES D R I V E N BY T E M P E R A T U R E D I F F E R E N C E Luis A. Odicino, Jos~ Marchese, Daniel A. Perell6
Universidad Nacional de San Luis, Laboratorio de Energia Solar, Ejtrcito de los Andes 950, 5700 San Luis, Argentina Telephone and Fax (+54) (652) (25109) E-MAIL [email protected] Graciela Lesino
Universidad Nacional de Salta, Buenos Aires 117, 4400 Salta, Argentina Abstract - The methods used to produce distilled water are of great interest for its meaning in all human activity. Water distillation by microporous membranes is a relatively new process widely investigated. The phenomenon lies mainly in that the surface of a microporous membrane is tangentially fed with an aqueous solution previously heated (feeding side). The other side of the membrane (permeate side) is at a temperature lower than that of the feeding solution. The brackish water of our region is not apt for human consumption due to two main components: fluorine and arsenic in the form of sodium fluoride and sodium arsenate, respectively. In particular, sodium fluorine occurs in mean concentrations in the order of 7 - 8 ppm, being in the order of 1.4 to 2.4 ppm the concentration recommended by the World Health Organization. In this study, fluorine is the saline component selected to be removed. Distillation tests are carried out on three synthetic solutions, each one having different concentrations of sodium fluoride: Two temperature differences of trans-membrane are chosen: ATI=I 0~ with Tu=40~ and AT2=30~ with TM=60~
1. INTRODUCTION One of the most active fields for purification of brackish water as well as hard fresh water is represented by the microporous membranes. These, constructed in distinct materials, in theory are considered a means of separation of two phases. The usual ways of performance are by means of a pressure gradient together with a concentration gradient. One of the ways of producing a pressure gradient is by means of a temperature gradient. This fact induces that separation by means of those membranes becomes attractive in the field of taking advantage of solar energy, as those gradients don't need having elevated values in order that the membrane may become active. In a previous paper we explored the feasibleness of the method by developing a distillation cell, in the present paper we have performed experiments intended in discerning the behavior of a microporous membrane constructed in the laboratories of the UNSL confronting one of the most common contaminants of the spring water in our region, as is fluorine. For the purpose we detected the maximum values of natural waters of the region, and, establishing controlled samples with concentrations including this maximum and lower values, concentrations of: 5 ppm, 7 ppm, and 9 ppm were selected, and a series of experiments with various values of water volume and temperature gradients were performed, all of them consistent with temperature values possible to be obtained with the medium solar radiation of the region. The polymeric membrane has a plane geometry and was synthesized in our own university laboratory by the use of polysulphones and the reversal phase process. This type of membrane is asymmetric with the following structural characteristics in the dense layer or skin: 0.1 - 0.3 Bm thick, mean pore diameter 20 - 30 A, mean porosity 3.5%, and 10000 daltons molecular weight cut-off (Marchese, Jos6). As we have explained in former papers (Odicino et al.), on one side of the membrane we have the concentrate circulating while on the other side we carry away the distillate that is to be
condensed by means of a controlled air flux counterdirected to the feeding. The experimental construct is shown in the figure:
Fig. 1 Scheme of the experimental construct
2. THEORETICAL MARGIM 2.1 MASS TRANSFER There are three mechanisms of mass transfer in processes of permeating membranes, which are: Knudsen type diffusion, Poiseuille type flux, and a Fick type diffusion model. Each of these models obeys the following equations:
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Nx = 1 , 0 6 4 ~
(M) ~
~
(P~ - P,,) ............ (I)
r2m M (pZ _ p,~).
Ne = 0,063 rEr/RT
................
(2)
1 DeM
By means of measurement of permeate values, the value for N, and the fixings of temperatures we could find the value of K for the cell and in particular the membrane. 2.2 HEAT TRANSFER Heat transfer is performed applying two main mechanisms: transfer of latent heat that accompanies the vapor flux, and by conduction through the membrane. The flux of latent heat simply is
Y~. Er RT (P~ - P'') ....................... (3) where: NK, NF, Ne represent the mass flux in the Knudsen diffusion model, in the Fick diffusion, and the Poisseuille type flux model, respectively. M is the molecular weight, R is the universal gas constant, T is the absolute temperature, ~i the membrane thickness, 11 is the viscosity, D is the molecular diffusion coefficient, Y~. is the medium logarithm of the molar inert gas fraction, P are the vapor pressures at both sides of the membrane, e is the porosity of the membrane, x is the tortuousity factor that is assigned the value 2, and r is the mean pore diameter. The calculus that permits knowing the mean free trajectory of the water molecules under mean working conditions yields the result that this is around 50 to 60 times the medium membrane diameter, so that we can discard the Poiseuille type flux model and also the molecular diffusion model. Consequently the Knudsen type diffusion model is applicable, as the flux will be dominated by the collisions of the water molecules with the pore wall. The former equations suggest that the permation flux through the membrane obeys the following simplified equation:
Nw
with N for mass flux of permeate, and Hv for vaporization heat. In our case it is that of pure water.
Q - NH
(8)
Heat flux from conduction obeys the expression
k.
-
+
(1 - e)k,
..............
(9)
In this expression l~n is the effective thermic conductivity of the membrane that is calculated in a similar way as a composite wall. Here 1%and ks are the eonductivities of the gas inside the pores, in our case air, and the conductivity of the solid the membrane is composed of, respectively, e is the porosity of the membrane. Total heat transferred through the membrane is
r , . - 7;.) ........... (lO)
d P , T,
= C~.-~.(
~ooooooooooooo
, - T2) ............... ( 4 )
By means of the Clausius-Clayperon equation
dP dT
PM2 IrM= R T 2
............... (5)
and an approximation of the behavior of the pressure
e=e
t
9
Beside that there are the conduction coefficients corresponding to the border layer that evolves at both sides of the membrane, hfood and h~. An electric analogon can be outlined as follows:
TM_45 )
. .................. (6) Q
Tout
T2
hfoed
t
we can reach a general expression of kind: T1
N = K ( ' I ; - T2) ............................. ( 7 ) where constant K depends on the structural parameters of the membrane and will also be subject to temperture and flux conditions.
%,
QV
In this analogon it is important to emphasize that conduction heat implicitly assumes knowledge of surface temperatures of the membrane, Tm and Tout, that in general are unknown.
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Applying the principles of heat transfer and energy conservation it is possible make eq. (11) a function of the effectively measured temperatures that are the temperatures measured in the interior of the liquids. In consequence, we can put
~ _ . . . . e
ILl
~
e
--
- - = - - delT.12.C - - e - - DeI.T.30C
8
o
,?.
Q r = H(T~ - 1"2) ....................
(11)
o
4
where H is the effective coefficient of heat transfer of the membrane. Pay attention to the fact that this coefficint is dependent on TM, since N is dependent on this parameter, and its dependence will be larger or smaller owing to the type of flux we have to consider (Poiseuille, Knudsen, or Fick, or maybe a combination of them). H gets the shape
3. EXPERIMENTAL DESIGN The experimental design is equal to the one proposed by Odicino et al. At the present opportunity in distilled water controlled quantities of fluor were dissolved in order to obtain three concentrations, one corresponding to the highest value found in natural waters of the region, one comparable to the highest value permitted by the WHO, and one value between both. The concentrations so selected are 9 ppm, 7 ppm, and 5 ppm. For each of these concentrations water circulation was performed under controlled feeding quantities as well as with a variety of thermic gradient values, compatible with the results of former experiments, selecting the temperatures that offered the best yield of permeate. Those temperatures are AT1 = 10~ with a T ~ D = 40~ and ATE = 30~ with a TMED= 60~
EXP
1 2 3 4
FEED FLOW Cm3/s 5 8.3 12.5 16.7
6
x uJ
9
4
Tout oC
Tinput oC
35.1 35.1 35.1 35.1
34.6 34.8 34.9 34.9
18.0 19.0 18.9 18.6
9
I
8
9
I
9
10
I
9
12
I
14
9
I
9
16
FEED CM 3 / SEG)
Figure 1 According to an experience with 9 ppm of fluorine, AT = 12 ~ and AT =30 ~ In the included table the values obtained by experiment for a particular experimentation are shown, in this case the group of measurements corresponding to those performed with concentrations of 9 ppm of fluorine and a AT1 = 12~ approximately. The relative water quantities were of 0.3 lts/m, 0.5 Its/m, 0.75 Its/m, and 1 lt/m. Having in mind the dimensions of the cell, those quantities correspond to Reynolds numbers lower than 2500, permitting us to assure that the water flow is laminar in all cases. Two temperature values of entrance and two temperature values of exit are noted, as well as for the feeding as for the dragging air, and for calculation purposes the mean values at each point are noted. The values are noted every 30 minutes, and also in each time interval the permeate liquid abundance, while the feed flow is controlled permanently by means of a water flow meter positioned in the circuit. Stabilization of the values of temperature and permeate is achieved in the course of two or three measurement intervals. By this means the effective operation conditions are reproduced to which the cell will be exposed when working with a feeding heated by solar radiation.
AIR
Tin oC
I
6
Toutput ~C 27.0 26.0 26.0 26.0
PERM.
Qv
Q
cm3/s
Watt
Watt
Watt
8.2 10.21 10.10 8.86
2 0.25 0.40 1.65
10.5 10.46 10.5 10.5
x l 0 -3
3.6 4.5 4.4 3.9
Table 1 Experimental values corresponding to a determined value of AT =12~ for four different flows. Moreover heat leak values of the system are indicated.
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4. CONCLUSIONS
REFERENCES
The results show: in the first case, we need 52 mJ/h for a m2 of membrane. At San Luis (33.2 Lat. S, 66 Long W and 800 W/m 2 mean radiation on horizontal surface) we need about 23 rn2 of horizontal fiat collector (11=0.8). In the second ease, we need 208 mJ/h and about 140 m 2 of horizontal flat collector (11=0.5). The productivity is about 220 cm3/h of distillate water per m2 of membrane and m 2 of flat collector in the first case and about 60 cm3/h in the second case. In the graphic shown below the construction proposal of a solar distillation system with membrane and use of a low cost solar collector is designed. The system seems feasible, but the selectivty of the membrane and its relation to temperature will have to be considered. Though the permeate has low salt values, of 9 ppm of fluorine in the feeding, the permeate has fluorine values between 3 and3.5.
Odicino L. A., Ochoa N. A., and Lesino G. Water Distillation With Microporus Membranes Driven With Solar Radiation.Proceeding Eurosun 1998. Lawson K. W., and Lloyd D.R. Membrane Distillation II Direct Contact MD. Journal Of Membrane Science 120(1996) 123-133 Lawson K. W., and Lloyd D. R Membrane Distillation Review Journal Of Membrane Science 124(1997) 1-25 Hogan P.A., Sudjito, Fane G. A. and Morrison G. L. Desalination by Solar Heated Membrane Distillation. Desalination 81 (1991) 81-90 Fasulo A. J., Perello D. A., and Follari J. A Solar Water Reservoir for Domestic Use..- WREC 1998 Marchese Jest, Membranas Procesos con Membranas, Editorial Universitaria San Luis 1995
Sketch of a distillor based on membrane, using a low cost water heater
Hemicell and membrane
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EFFECTIVE SOLAR ENERGY UTILISATION - MORE DEPENDENT ON SYSTEM DESIGN THAN SOLAR COLLECTOR EFFICIENCY
J. Rekstad, L. Henden, A.G. Imenes, F. Ingebretsen, M. Meir Department of Physics, University of Oslo, P.O. Box 1048 Blindern, N-0316 Oslo, Norway Tel.: +47 22 85 64 75, FAX: +47 22 85 64 22, [email protected]
B. Bjerke, M. Peter* SolarNor AS, Erling Skjalgssons gate 19B, N-0267 Oslo, Norway
Abstract - A new concept for a combined space heating and domestic hot water solar system based on polymer solar collectors and low temperature utility system is presented. Experimental results from laboratory as well as from systems in the field demonstrates that the performance is comparable to the performance of conventional systems with high efficiency collectors connected to high temperature utility systems. These results agree with the results of simulations carried out for different system configurations and operating temperatures.
1. I N T R O D U C T I O N
increased load from space heating compared to the domestic hot water (DHW) demand.
Before introducing solar energy a critical evaluation of conventional energy systems in domestic buildings seems appropriate. Most of the existing solar energy systems consist of solar collectors, heat stores and controllers adapted to the conventional heating systems. These systems are results of a long lasting development governed by unlimited availability of low cost and high quality energy carders. The fundamental needs to be served are represented by the desired indoor temperature of 20-23~ and a hot water temperature which ranges from 30-50~ The system discussed in this paper attempts to cover these needs by a sufficient, but minimum heat carrier temperature. Heat exchangers are avoided, the heating surface is maximised by using the floor area as heating surface, and to apply heat buffers of considerable sizes. The paper discuss the different elements in such a system, based on the measurements from series of experiments from laboratory as well as from systems in the field. The results are compared with simulations by means of a program which computes the energy flow hour by hour. The simulation code includes a synthetic weather generator which provides the opportunity to easily study the impact of various temperature and solar radiation condition. 2. D E S I G N O F T H E C O M B I N E D SYSTEM
2. This increased load demands larger solar collector area than conventional DHW-systems. Hence the collector cost becomes very significant in the total cost-picture. The use of low cost and consequently less effective collectors (especially at high temperature) becomes necessary. 3. The system has built in a hierarchy for the use of the solar heat. The first demand to cover is the energy required to heat the domestic hot water from cold water temperature up to room temperature. The second demand is to cover (all or parts) of the energy for space heating. Excess solar heat is used for heating the domestic hot water from room temperature up to desired taptemperature. These design criteria focus strongly on minimising the system temperature. The construction and properties of the various system elements are discussed below.
SOLAR
The system shown in figure 1 is the result of a design process where the following aspects have been important 1. The solar system will, in particular in cold climate as in Norway, have a relatively low solar fraction due to the
* present address: Steinbeis Transfer Zentrum, Berlin
Fig. 1. Design of a combined solar space and DHW heating system for low temperature heating
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3. T H E S O L A R
COLLECTOR
The development of a simple building-integrated polymer collector has been completed in collaboration with General Electric Plastics [ 1]. The collector, shown in figure 2 consists of a front sheet of polycarbonate (twin-wall), and an absorber of modified polyphenylenoxid (PPO) plastics with internal channels filled with ceramic granulates. The heat carrier is pure water, which trickles downwards through a number of parallel channels inside the absorber. These channels are filled with ceramic granulates. The effect of these granulates is to enable good heat transfer from the front side of the absorber to the water, and to fill the vacant space between the granulates completely with water by removal of the air without a subsequent pressure build-up.
Fig. 3. Collector efficiency of a solar collector in PPO-plastics. Due to the temperature limitations of the PPO-material, the collector has no radiative selectivity in order to avoid overheating at standstill. Hence the efficiency drops considerably for high temperatures. This feature makes the focus on a low operational temperature even more important than if the collector had high efficiency at high operational temperature. The aim of this presentation is to show that the net energy gain expressed in terms of the total system efficiency, is high in spite of a moderate collectorperformance at high temperature.
4. L O W T E M P E R A T U R E S P A C E H E A T I N G
Fig. 2. Solar collector in polymer plastics The result is a modular building system of low weight (10 kg/m2) and flexible with regard to adaptation to various building forms and areas. Freezing and boiling is prevented by control of the circulation pump, since the water drains automatically out of the collector and into the store when the circulation stops (drain-back system). The solar collector efficiency is shown in figure 3.
The expression <> refers to the temperature of the heat carrier in the system, and not on the comfort temperature required by the user. In principle a low temperature heating system can be realised in different ways, the key point is large heating surfaces where the heat transfer is dominated by heat radiation rather than convection. Support has been obtained by the Norwegian Association for Allergy and Asthma which promotes hydronic floor heating as the heating principle in new, well insulated buildings from health point of view. Floor heating is applicable in various floor constructions. Figure 4 shows a cross section of a floor heating system in a typical Norwegian wooden floor construction.
Fig. 4. Floor heating system in a typical Norwegian wooden floor construction
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Tubes of cross-linked polyethylene plastics (PEX) equipped with heat distributors in aluminium are fixed under the top layer of the floor, and provide a sufficient thermal coupling between the heat carrier and the upper surface of the floor. Figure 5 shows typical experimental results for the relation between the heating power per area unit and heat carrier temperature in a modem wooden floor. Even better thermal coupling is normally obtained in a concrete floor. The present insulation standard in Norway demands a heating power up to approx. 30 W/m: floor area. This demand can easily be fulfilled by a floor heating system, which may transfer, under acceptable temperature conditions, up to 50 W/m:.
Fig. 6. The heat buffer store: The pump unit for the solar loop is placed to the left. The floor circulation pump is place to the right. A DHW-preheating tank in the centre.
Fig. 5. Relation between the heating power and heat carrier temperature in a modern wooden floor. The heated floor area is 175 m e.
A disadvantage with low temperature heating systems is the limited power dynamics, associated with the temperature restrictions on the floor surface and the considerable heat capacity of the floor. This aspect is discussed later.
5. T H E H E A T S T O R E The heat store is justified by the unstable solar radiation, but also by a more general need for better matching between power supply and power demand. Hence the heat store is considered as the interface between any source system and the system of utilisation. Not only solar, but also heat pumps, biomass and even electricitydriven heating systems gain flexibility from the introduction of heat store capacity at the end-user level. The heat stores used in the investigations reported here, are square-shaped water tanks in aluminium or stainless steel. The store is shown in figure 6.
It is typically equipped with pump modules for floor heating, for circulation through the solar collector, auxiliary heat source and a pre heating tank for domestic hot water. As seen from fig. 1, we omit conventional heat exchangers between the store and the solar collector loop, as well as the floor heating system. The store is connected to the atmosphere through a small aperture and therefore pressureless. The top volume of the store (typically 5 % of the store volume) is reserved for the thermal expansion and reservoir for water occupying the collector under operation of the solar system (drain-back system). This simple approach has proven to be very feasible; several components usually needed in a floor heating and solar system can be dropped, and the start up procedure of a new built-up system (fill up with water, removal of air) becomes much simpler. The low temperature approach puts less emphasise on the temperature stratification aspect than in most other solar store designs. The stratification obtained during the charging is negligible, while stratification develops during discharge. The dimensions of the heat store are determined from a criterion of 15 hours storing capacity under the coldest winter period, leading to 1000 litres store pr. 80 m2 floor area. This ratio is valid for the south part of Norway, milder climates require smaller store volumes. 6. T H E T E M P E R A T U R E C O N T R O L L E R The performance and usability of the system described above depends on the quality of control of the various processes and functions. A special controller has been developed for this concept due to a series of new aspects presented by the use of polymer collectors, the drain back feature and the floor heating characteristics. The poor thermal conductivity of the plastic material requires a thermal sensor which monitors the solar radiation in addition to the collector temperature. This ,,modified" collector temperature is
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compared with the store temperature. A special pumping procedure is applied for deposition and transportation of the air which after long stand still is occupying the volume between the granulates in the collector. The floor heating system (or more general low temperature heating system) has a poor power dynamic. In other words, the system has a considerable thermal inertia. Temperature control based on room thermostats and feed-back loops may in this case often cause temperature oscillations with negative effects on comfort and energy economy. The control strategy applied is to determine in advance the heat demand from the available information on ambient temperature and solar radiation, and to deliver the required energy for the next time period in terms of an ,,energy pulse". The duration of this pulse is calculated from the available information about the floor heating system. The control principle is shown in figure 7, while figure 8 gives a typical example of the quality of temperature stabilisation which is achieved with this technique.
Fig. 9. Measurements at the Sol-Lab, 12. 05. 98, active collector area 4.05 m 2, buffer store volume 2851 The characterisation of the collector-store system based on the present concept has been explored under controlled laboratory conditions as a part of this study. Collectors in polymer plastics with 4.03 and 5.3 m 2 active area, tilted in 30 ~ (towards south), have been used in combination with heat stores containing 280 litres and 1000 litres respectively. Results obtained with these combinations are shown in the figures 9 and 10, and are summarised in table 1.
Fig. 7. Temperature regulation by sequential pump operation.
Fig. 8. Quality of temperature stabilisation achieved with the control principle illustrated in fig. 7. 7. EXPERIMENTAL RESULTS
7.1. Laboratory experiments A small test laboratory with approx. 20 m 2 floor area, insulated according to Norwegian building standards, has been built at the Department of physics, University of Oslo. The laboratory is equipped for full scale solar heating and radiative cooling experiments [2 ].
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Table 1. Results from solar laboratory measurements. (Sol-Lab) Date
Aeon Ve Tmb Tstar (m2) (litres) (~ (~
12.05.98 23.07.98 01.03.99 05.05.99 02.03.99 02.05.99 04.05.99 06.04.99
4.05 4.05 5.30 5.30 5.30 5.30 5.30 5.30
280 280 280 1000 280 1000 1000 1000
19 20 8 17 8 13 17 17
9.5 20.0 44.0 18.5 39.0 31.0 51.0 43.0
Tstop (~ 52.5 58.0 54.4 41.5 51.5 46.5 59.5 55.5
Oso~ ~ (kWh/m2)
F (%)
3.45 3.05 2.30 1.41 0.77 3.40 1.90 2.70
7.2
48
6.8 3.8 3.9 7.2 7.0 7.2
34 37 20 47 27 38
269
The energy monitoring is based on calorimetric measurements of the well insulated heat store, where the temperatures are measured in intervals of 10 minutes at different levels. These temperatures are stored together with simultaneous values of the solar radiation, ambient temperature and the floor heating temperature, in a SolDat portable datalogger. The datalogger is programmed and started by means of a PC, which also contains programs for read out and analyses of the data in the logger. Figures 11 and 12 show these quantities measured during two representative time sequences.
The system is effective in the temperature range up to approx. 60~ which covers the desired working temperature for domestic hot water and room heating supply. The results are used as a standard for evaluation of the results obtained for systems mounted in the field as described below.
Fig. 11. Measurements in the field, week 20 (1998), house A The data refers to a house (A) with 262 m 2 heated floor area and a DHW demand of 4.500 kWh. The solar collector of 31.7 m 2 aperture area coupled to a 3000 litres water storage. The collector tilting angle is 36 degrees. The auxiliary heat source in this system is an oven for biomass (wood, pellets), also connected to the heat store by means of a heat exchanger.
Fig. 10. Measurements at the Sol-Lab, 4.-5.05.99, active collector area 5.3 m 2, buffer store volume 10001, flow rate 4801/h 7.2. Experimental results obtained in residential solar installations Around 100 systems with a design as described in chapter (2) here have been installed in Norway during the last couple of years. A few homes have been objects for energy measurements and evaluations, partly financed by the Norwegian Directorate for Water Resources and Energy (NVE). Some of the results are presented here, while a complete reports of these studies is published in [3].
Fig. 12. Measurements in the field, 27.4.-22.5.98, house A The second house 03) has a solar collector of 28.2 m 2 area, tilted in 27 degrees (south-faced), and connected to a 2000 litres heat store. The heat load is represented by a 180 m 2 floor heating system, a swimming pool of 25.000 litres (not in use during the measurements) and DHW. Auxiliary heat is provided by a 6 kW electric heating element. A part of the floor heating system in this
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house is installed in a workshop with a room temperature in the range of 10 to 15 ~ The results from two houses are summarised in table (2). Table 2. Results of energy measurements. House A
House B
Measuring periods Number of days Days with solar gain Total solar gain Average solar gain pr. day Maximum output in one day
06.03.-22.04.98 44 34 1585 kWh 36.0 kWh 2.91 kWh/m2
27.02.-24.05.98 87 67 3036 kWh 34.9 kWh 3.11 kWh/m2
temperature. The latter is a consequence of the system of utilisation (floor heating, radiators) and the efficiency of eventual heat exchangers. The tap-temperature of the domestic hot water is 55~ The store volume used in the comparison shown in table 3 is 3000 litres. The <> collector shows less dependency of the system temperature then the polymer collector, the difference in delivered solar energy by the two systems is 15 % at 25 ~ system temperature (representative for floor heating). For a conventional radiator system ( T ~ n = 60 o) the difference is more than 30 %.
Table 3. The measurements reveal the response of the solar system on various climatic conditions. The net annual solar energy gain is determined by comparison of the measured results with simulations for a whole year based on this response and on average radiation and temperature conditions. The net solar gain is 7.900 kWh/year and 9000 kWh/year for house A and house B respectively. The reason for a higher output from the system in house B is the additional load during summer time due to the swimming pool.
8. SOLAR GAIN VERSUS COLLECTOR EFFICIENCY The results of the measurements reported above show that the solar gain obtained for the combined space and DHW heating systems based on the described low temperature design, are comparable with results obtained for high temperature systems with more effective, but considerably more expensive collectors than the present polymer collector. Different system configurations have been studied by means of a simulation program which calculates hour by our the energy flow and temperature conditions from information on solar radiation, ambient temperature and various heat loads. A typical Norwegian one-family house with an annual DHW demand corresponding to 4.600 kWh and room heating demand of 12.000 kWh/year is used for the comparison. The climate is characteristic for Oslo, where the annual solar radiation is approx. 1000 kWh/m: year and the average ambient temperature in January is -10~ The season for space heating is from September to May. Two collectors are compared; the present polymer collector and an <>collector with maximum efficiency of 81% and an overall heat loss coefficient of 3.0 W/(m:K). A collector area of 20 m:, oriented toward south with a tilting angle of 45 ~, is used for both systems. The performance is studied as a function of heat store volume and characteristic system
Comparisonof performance of systems with high and moderate collector efficiency
Tsymm(~
25 40 50 60
Solar energy gain (kWh) Presentsystem Ideal system
5.700 5.100 4.700 4.300
6.600 6.100 5.900 5.600
The <
REFERENCES
[1] J.Rekstad et al. (1998). Building integrated solar systems, Proceedings 5th European Conference in Solar Architecture and Urban Planning, Bonn, Germany [2] A.G. Imenes et al. (1998). The Sun-Lab at the University of Oslo, Annual Report of the Nuclear Physics and Energy group at the University of Oslo, 1998. [3] J. Rekstad et al. (1999). Kombinerte soloppvarmings-anlegg for varmtvann og romoppvarming - m~dinger av energiutbytte, Report 5/99, SolarNor AS, Erl. Skjalgssons gt. 19B, 0267 Oslo, Norway.
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DYNAMICAL MODEL FOR A SOLAR STILL VALIDATED FROM OPTICAL AND THERMAL PARAMETERS MEASURED EXPERIMENTALLY Eugenio Torijano, Alejandro V~zquez, Hernando Romero-Paredes, Alejandro Torres, Juan J. Ambriz, Eugenio Torijano Jr. Area de Ingenieria en Recursos Energ6ticos, Universidad Aut6noma Metropolitana-Iztapalapa, A.P. 53-540, Iztapalapa, C.P. 09340, M6xico, D.F., Tel. (52) 5724 4644, Fax: (52) 5724 4900, E-mail: [email protected] Abstract - A differential dynamical model of a solar still of three stages is proposed to simulated laboratory prototypes, in order to optimize their design. The model is generated from differential mass and energy balances in each one of the three stages; it is composed of seven coupled ordinary differential equations that correspond to the metallic basin and the water in it and two glass reservoirs acting as basin and condensing surfaces. Temperatures in the interfaces between water bodies and metallic basin and glass surfaces and water bodies, ambient temperature, solar global radiation and wind velocity were monitored. Heat transfer coefficients from metallic basin to water (or glass basin to water) and heat transfer coefficients from water surface to glass condensing surface were obtained from experimental data, and then included into the model. The set of ordinary differential equations was solved numerically by means of subroutines of DGEAR with approximation of the jacobian for backward finite differences. The predicted and experimental temperature profiles were compared and show a good accordance during the simulation period. Toward the end of the solar day they show an important thermal inertia, which is very close in both cases.
1. INTRODUCTION There are a number of papers related with design, operation and modeling of solar stills prototypes (Shoda et al, 1980), they show high variety of geometrical configurations and performances. Some of the most successful models are for the stationary state and provide acceptable solutions for this case. Thus for a more complete acknowledgment in the solar distillation field and in order to reduce time and costs involving in construction and testing of prototypes, a model describing the phenomenon, in a complete form, the performance of any kind of still could be a very useful tool. 2. OBJECTIVE The aim of this work is to develop a better approximation to a complete dynamical simulation model for solar stills of different number of effects, for any time, which includes those of the stationary state.
effects, the system was tested during four hours periods centered at solar noon. These values were introduced into the model as parameters in the energy balance equations. Temperatures were measured by means thermocouples fixed to both surfaces, lower and upper, of every glass plate, inside every water body and another one on inner surface of the metallic plate. Convective heat transfer coefficients was obtained from a heat transfer simulator developed by Holman (J.P. Holman, 1997) using temperature values previously measured; also ambient temperature was measured. Electric signals coming from the thermocouples and global solar radiation were recorded using data acquisition board in a personal computer. Wind velocity was measured with a analogic cups anemometer. A set of ordinary differential equations, ODE's, was generated from mass and energy balances considering every one of the surfaces involved. The set of equations describing the system is: Metal basin:
3. THEORETICAL ANALYSIS AND METHODOLOGY The first step in developing the model is to make the following assumptions: 1. Heat capacities of glass and water in the still are not negligible. 2. There is an important fraction of incident solar radiation absorbed by glass and water together. 3. Thermal losses in the lateral faces are negligible, (north and south faces are insulated) and there is not thermal losses from the bottom. 4. The heating of the water mass is uniform in each basin. Solar radiation attenuation due to three glass plates and two water bodies (upper and middle levels) was measured putting a Kipp pyranometer in the bottom of the still and another one outside it. First, for only the upper glass cover, then the upper and middle glass covers and so on. Comparing both measures it was found the total attenuation percentage due to combined glass and water bodies. In order to minimize the zenithal angle
dr,7 3 3 A1 mpCpp -di- =Otp'C~'gn s + u~(r~ -r,)+u~(ro-r,)
(1)
Lower water film:
m, Cp,, - - ~ = tt,, ~ H ,
-Up~ (T,, - Tp)+ U~g, (Tgl - T~ ) (2)
Lower glass cover:
I I mgcp,, aTgI -] =otg'r
(r~-r,)+ (r,-r~)(3)
Middle water film:
mw Cp ~ ~
=otwrw~H , -Ug,,~,(T., -Tz,)+U~. ~(Tg, -Tw, ) (4)
Middle glass cover:
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dr,~ q -AIm, Cp,~--~ =~,~:.,'r,H. -U.,.a.(T,2- T,,.~)+ U,,~(T,~- T,,)(5)
1[
water bodies are shown in figures 2 to 5, the corresponding measured values are shown in figures 6 to 9.
Upper water film:
dT,~] =awrgH, -Ug~,~(T,~ -Ts~)+U~3 (Tg, -T,~)
-~ m,~Cp,~ - ~
(6)
Upper glass cover:
m. Cp., --~ :%H.-U.~,(T., -T.~)+U..,~176
(7)
The numerical solution of this ODE's system is accomplished using the numerical subroutines of DGEAR with approximation of the jacobian for backward finite differences (Hindmarsh, 1975). The results, experimental and simulated, are presented graphically in the differents figures. 4. RESULTS AND DISCUSSION Table 1 shows some average values for incident solar global radiation (out and inside the still), measured in winter time, for three glass covers and two water bodies. Time of the day 12:00 12:07 12:10 12:15 12:20 12:24 12:30 12.35 12:40 12:44 12.50 12:55 13:02 13:06 13:10 13:14 13:28 13:31
Solar radiation outside (W/m2) 695.6 748.4 758.3 718.6 700.0 734.7 711.2 740.9 698.8 718.6 748.4 742.2 739.7 732.3 745.6 738.2 729.8 714.9
Solar radiation inside (W/m2) 314.4 343.4 346.3 315.0 312.2 316.4 322.1 315.0 305.1 299.6 323.6 320.7 332.1 336.4 324.1 325.5 322.6 319.1
Fig. 1. Measured solar global radiation.
Attenuation (%) 54.9 54.6 54.3 56.1 55.4 56.9 54.7 57.4 56.3 58.3 56.7 56.7 55.1 54. I 56.5 55.9 55.8 55.3
Table 1. Averages attenuation of solar radiation input (percentages). The following design parameters for the materials were taken. Xg = 0.78; Xw= 0.99; % = 0.22; t ~ = 0.01; % = 0.75; A1= 0.25 m2; A2 = 0.27 mE; A3 - 0.58 m2; Cpp - 501.8 J/Kg K; Cpw = 4174 J/Kg K; Cpg = 819.7 J/Kg K. Convective heat transfer coefficients obtained were: Upw = 4.0 W/m2K; Uwlgl = 1.13 W/m2K; Uw2g2= 1.69 W/m2K; 2 Uw2~= 1.72 W/m2K; Uglw2- 3.79 W/mK; Ug2w32- 3.79 W/mEK 28th June was chosen as typical day for evaluating the model. Measurements of the solar global radiation for this day are shown in fig. 1. The simulation results for the glass covers and
Fig. 2. Simulated temperature for upper glass cover
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Fig. 3. Simulated temperature for middle glass cover.
Fig. 4. Simulated temperature for lower glass cover.
Fig. 7. Measured temperatures on both sides of middle glass.
Fig. 5. Simulated temperature profiles for water bodies.
Fig. 8. Measured temperatures on both sides of lower glass.
Fig. 9. Fig. 6. Measured temperatures on both sides of upper glass.
Temperatureprofiles for the water bodies inside the still.
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The temperatures measured in the water layers are not so different like those that model predict, probably due to the fact that exist a better heat transfer between glass and water; then, water temperatures resemble glass profile temperatures. 5. C O N C L U S I O N S
Fig. 10. Vapour temperatures inside the still.
The model proposed to describe a dynamical state, for a three stages. Experimental validation shows an adequate agreement once the maximum and final operation conditions are reached, however, the model description for the initial phase of operation is not well described. Future work will be focused to obtain better values for materials properties and heat transfer coefficients. The mathematical model proposed not describe the phenomena presented during the operation like fog and excessive growing of water layer over condensing surfaces. 6. REFERENCES
(1975). Applications of EPISODE: An experimental package for the system of ordinary differential equations. Lawrence Livermore
Hindmarsh A. C., Byme G.D.,
Laboratory Report UCID - 30112. Holman J. P., (19977)dteat Transfer. Me Graw Hill. Porta M.A., Chargoy N. and Fern~mdez J.L., (1997).
Extreme operation conditions in shallow solar stills.
Fig. 11. Temperature gradient between both sides in each glass cover. The experimental results shown in figures 10 were used to evaluate the convective heat transfer coefficient. The figure 11 shows that non exist zero gradient in glass plates as is assumed in other works in the literature. Temperatures predicted by the model for the glass covers are relatively high due to the value considered for absorption in them. Upper glass cover temperature has a slightly less value than the others due to convective losses to air. Respect to water temperatures, the model predicts a higher one in the lower level caused by the high absorption in the metallic plate; middle and upper water temperature are similar. However, temperature profiles for water are very different of those for glass covers: they show higher thermal inertia toward the end of solar day than glass covers. In the laboratory sets, the heating of glass covers occurs more slowly that predicts the model and is less sensitive to solar radiation variations, but once the glass covers reach the higher temperature, they stay almost stable. Upper glass cover temperature is also the lower one, but it presents the higher thermal gradient between its lower and upper surfaces.
Solar Energy 61, 4, 279- 286. Shoda M.S., Nayak J.K., Tiwari G.N. and Kumar A., (1980a). Double basin solar still. Energy Conversion, 20, 23. Shoda M.S., Kumar A., Shingh U. and Tiwari G.N., (1980b). Transient analysis of a solar still. Energy Conversion, 20, 191. Vfizquez A., Torijano E., Romero-Paredes H., Torres A. and Ambriz J.J., (1998). Unsteady Treatment for a Three-Effect Solar Still. Renewable Energy, 15/16, 2251-2254.
7. N O M E N C L A T U R E i = Index to describe the effect i = 1,...,3 (bottom=l) t = time (s) Ta = ambient temperature (K) T# = glass cover temperature at effect i (K) Tw~= water temperature at effect i (K) Tp = absorber plate temperature (K) Hs = solar intensity (W/m:) Cp# = thermal capacity of glass i (J/kg K) Cpw~= thermal capacity of water i (J/kg K) Cpp = thermal capacity of plate (J/kg K) m# = mass of glass at effect i (kg) mw~ = mass of water at effect i (kg) mp = mass of absorber plate (kg) A# = area of glass at effect i (m 2)
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Awi = area of water at effect i (m 2) Ap = area of absorber plate (m 2) Upw = heat transfer coefficient from plate to water (W/m 2
Ir Upa = heat transfer coefficient plate to ambient (W/m 2 K) Uwig~= heat transfer coefficient from water to glass cover i (W/m2 K) U~wi+l = heat transfer coefficient from glass cover i to water i+l (W/m 2 K), i = 1,2 Ucga = heat transfer coefficient from glass cover to ambient ( W / m 2 K) Urg~ = radiative heat transfer coefficient from glass cover to ambient (W/m 2 K)
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CHARACTERISTICS OF VERTICAL MANTLE HEAT EXCHANGERS FOR SOLAR WATER HEATERS L.J. Shah Department of Buildings and Energy, Technical University of Denmark DK-2800 Lyngby, Denmark Phone +45 45 1888, Fax +45 45 934430, Email: [email protected]
G.L. Morrison and M. Behnia School of Mechanical and Manufacturing Engineering, The University of New South Wales, Sydney 2052, Australia Phone + 61 2 93854127, Fax + 61 2 96631222, Email: [email protected], [email protected] Abstract- The flow structure in vertical mantle heat exchangers was investigated using a full-scale tank designed to facilitate flow visualisation. The flow structure and velocities in the mantle were measured using a particle Image Velocimetry (PIV) system. A CFD simulation model of vertical mantle heat exchangers was also developed for detailed evaluation of the heat flux distribution over the mantle surface. Both the experimental and simulation results indicate that distribution of the flow around the mantle gap is governed by buoyancy driven recirculation in the mantle. The operation of the mantle was evaluated for both high and low temperature input flows.
1. INTRODUCTION Energy transfer from a closed-loop pumped-circulation solar collector circuit to the primary water tank in a solar water heater can be achieved using a range of heat exchanger systems. To obtain optimum system performance it is necessary to promote thermal stratification in the storage tank and to minirnise pumping power for the collector loop. The most effective way of promoting stratification is to avoid designs based on water jets entering the tank. A range of stratification promoters or diffuser manifolds mounted inside the storage tank have been proposed and adopted in Germany and Switzerland. However, the diffusers have not been generally adopted due to their cost and difficulties of fitting them into the tank. Distributing the collector loop heat input over the tank wall surface area can also be used to minimise stratification disturbances. Distributed heat transfer minimises disturbing convection flows in the tank and results in most heat transfer occurring near the thermal equilibrium level of the collector loop inlet flow. Wrap-around coils and mantle tank systems are the simplest and cheapest means of producing high heat exchanger effectiveness while promoting stratification. Due to limited contact area on the collector-fluid side of a wrap-around coil, turbulent flow is needed to achieve the required heat transfer. This can only be achieved with high flow circulation rates in the collector loop. The mantle tank concept provides the largest heat transfer area and effective distribution of the collector loop flow so that most heat transfer occurs near the thermal equilibrium level between the solar collector fluid in the mantle and the contents of the tank. This paper addresses the performance of mantle heat exchanger tanks using visualisation of the flow structure in vertical tanks and numerical simulation of the flow in the mantle and in the core tank. 2. BACKGROUND Modelling solar water heating systems incorporating a mantle tank requires data on the heat transfer coefficients for the flow in the mantle passageway. Heat transfer in annuli, which are geometrically similar to mantles, has been examined in several
studies. However, there have been few reported investigations of heat transfer in large annuli. Nagendra et al. (1970) investigated free convection heat transfer in vertical concentric cylindrical annuli both analytically and experimentally. They used the approximate double boundary layer model for vertical parallel plates to obtain heat transfer correlations for vertical annuli. To check their correlations they tested two annuli with an inner diameter of 0.008 m and an outer diameter and length of [0.018, 0.279 m] and [0.0254, 0.287 m], respectively. Water was used as the fluid medimn, and the inner cylinder was maintained at a high temperature and the outer cylinder was at a lower temperature. Nusselt versus Rayleigh number correlations were developed for heat transfer across the annulus. Kubair and Simha (1982) analysed free convection heat transfer in vertical annuli by using the double boundary layer approach. To check their correlations they tested annuli with inner diameters in the range of 40 to 60 mm, outer diameter of 100 mm and a length of 0.2 m. Water and mercury were used as the fluid mediun~ The inner cylinder was maintained at a high temperature and the outer cylinder was at a lower temperature. They also developed Nusselt number correlations as a function of Rayleigh number. E1-Genk and Rao (1989) presented heat transfer data and correlations for hydrodynamically developed but buoyancyassisted (thermally) or opposed flows of water in vertical annuli with constant temperature boundaries. They gave data for forced and mixed laminar flows, forced transitional and turbulent flows and natural laminar and turbulent flows. Their experimental investigations were for two annuli with inner/outer cylinder diameters of 12.7/38.1 mm and 25.4/44.6 ram. The lengths of the annuli were 0.360 m and 0.9 m respectively. Nusselt versus Rayleigh and Reynolds number correlations were presented for the different cases. They also studied the effects of buoyancy induced instability on heat transfer in annuli using dye-injection flow visualisation and temperature-tracing techniques as reported by E1-Genk and Rao (1990). They correlated their data in the form of Nusselt versus Rayleigh number. Khan and Kumar (1989) conducted a numerical investigation to evaluate the effects of diameter ratio, s = rout,,/rmr and aspect ratio, A = (height o f annulus)/(rou~er-ri~e,), in natural
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convection of gases in vertical annuli. The inner cylinder was a constant heat flux boundary, and the outer cylinder was a constant temperature boundary. The cases considered were in the range of 1 [3 A n 10 and 1 [3/?[3 15. They also developed an average Nusselt (versus Rayleigh) number correlation. Rogers and Yao (1990) studied the hydrodynamic stability of mixed convection in an annulus with a constant heat flux boundary condition on the inner annulus and an insulated outer boundary. They found that the parallel flow assumption could not be used. They presented a Nusselt-Rayleigh number correlation for data in the region between linear stability and Ra=600. The annulus gap was used as the length scale:
Nu rq 1.02 ERa 0.28
for
1O0 [3Ra [3 600
particle tracks from each section were assembled to get the full flow field over each wall. The flow was measured in the centre plane of the mantle gap. Fig. 3 shows a photograph of the glass tank and mantle heat exchanger. Front:
400 Top:
(1) 400
Dhimdi and Bolle (1997) numerically determined heat transfer coefficients for natural convection in vertical annuli for large Grashof numbers using water as the heat transfer medium. Heat transfer was correlated with radius ratio and aspect ratio for constant temperature boundaries. They found that the Nusselt number was independent of the boundary curvature. Shah and Furbo (1998) evaluated heat transfer in the context of solar heating systems. In their study, the fluid in the annular gap was warmer than both walls, which results in heat transfer from the fluid to the walls instead of from one wall to another. Their analysis resulted in a local Nusselt-Rayleigh correlation, where x is the distance from the top of the annulus:
Nu x rqO.46[~axDc/Dhy~ ~ ~s
for
490
101
45
400
490
45
Fig. 1: Simplified sketch of the glass mantle tank. Dimensions are in [mm].
Ra~ rqlO" (2)
I I
Apart from the work by Shah and Furbo (1998) the boundary conditions used in previous studies of heat transfer in annuli are not applicable to heat transfer in mantle tank applications. For example, a uniform temperature boundary is impossible to obtain in a mantle tank just as a uniform heat flux boundary is hard to achieve. Hence, most available correlations for heat transfer in vertical annuli are not appropriate for calculating heat transfer in mantle tanks in which heat transfer is from the fluid to both walls of the mantle (useful heat transfer to the inner wall and heat loss from the outer wall). In this paper, a mantle heat exchanger is investigated experimentally using a Particle Image Velocimetry (PIV) system and numerically using Computational Fluid Dynamics. The ultimate aim is to develop heat transfer correlations for vertical annuli, with boundary conditions appropriate to vertical mantle tanks. 3. EXPERIMENTAL APPROACH A vertical mantle tank was constructed from glass for flow visualisation and velocity measurements with an optical PIV system. A square section tank was necessary to allow the velocity measurements in the narrow mantle gap. The tank used in this investigation is like a standard vertical mantle tank used in Denmark for small low-flow solar domestic hot water systems. The mantle and the inner tank volume, as well as the ratio between the solar and the auxiliary volume are similar to commercial solar water heater tanks used in Denmark. During the experiments, the tank was insulated with 24 mm PUR-foam. Fig. 1 shows a schematic of the mantle tank and Table 1 gives detailed heat transfer properties of the components. In order to get sufficient resolution with the PIV system the mantle was divided into smaller sections (Fig. 2), and images of
I
E
i
I
I '-I I
I
M7
M6
I I
M8
M5
'
I
%
M4
i
I fmm M12
Mll Mll
I
Fig. 2: The mantle is divided into 12 sections.
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[3 [3 [3
Volume flow rate in the mantle. Temperatures in the inner tank, T1 to T5 (Fig. 4). Ambient temperature.
Temperatures were measured using copper/constantan thermocouples placed directly in the fluid in order to get fast reaction time. A data logger was used to control the data sampling and temperatures and mantle flow rate data were stored on a PC.
Fig. 4: Measuring points in heat storage test facility. Fig. 3: Glass model mantle tank. 4. PARTICLE IMAGE VELOCIMETRY SYSTEM 124.0 Inner tank volume [1] 21.5 Mantle volume [1] 54.9 Inner tank volume over mantle [1] 54.9 Inner tank volume inside mantle [1] 14.2 , Inner tank volume under mantle [1] i, 0.901 ' Total heiglat of inner tank [m] 0.4 Total hei~ht of mantle [m] 0.012 Tank wall thickness [m] 0.012 Mantle wall thickness [m] 0.033 Mantle gap width [m] Heat transfer area between mantle and tank [m 2] . 0.64 ,
,
9
|
,
,
|
,
|
|
,,
|
|
|
|
9
,,
|
|
,
Glass materialproperties Specific heat [J/kg-K] Density [kg/m3] Thermal conductivity [W/mK],,
Insulation ~ UR-foam) properties:
,2250 , ,840 i .1 ! |
Specific heat [J/kg.K] .1045 . Specific mass capacity [kg/m3] : 70 Thermal conductivity [W/mK] 0.033 Table 1" Data for the mantle tank and insulation rt~aterial "
The mantle was supplied with a constant inlet temperature and a constant flow rate during the heat transfer tests. To obtain heat transfer data for the mantle the following variables were measured: El Mantle inlet temperature. [3 Mantle outlet temperature.
Particle Image Velocimetry (PIV) is an experimental technique, which combines flow visualisation and digital image processing to measure the fluid velocity over a wide section of the flow field. The basic principle of PIV is to determine fluid flow velocity indirectly by analysing the motion of seed particles in the flow. Small neutral density particles are used, so the velocity of each seed particle can be considered to be the same as the fluid velocity. This technique has the advantages of non-intrusive and instantaneous measurement of complete flow fields. The PIV system is based on continuous recording of the position of particles in a slice of the flow illuminated with a laser light sheet using a digital video camera. The video images were processed on a PC using the 'ScionImage '1 sottware package. The PIV system in these tests consisted of a 30 mW HeliumNeon Laser with a wavelength of 635 nm and a Panasonic digital video camera. Plastic particles with a density of 990 kg/m3 and a diameter of approximately 100-200 [3m were used to visualise the flow patterns. Fig. 5 illustrates the PIV system set-up. The PIV system was developed at the School of Mechanical and Manufacturing Engineering, University of New South Wales, based on the principles described in Dahl et al. (1995). The continuous video sequence is divided into a number of separate pictures (frames) at a specified time interval. The required time interval between the frames captured from the video depends of the fluid velocities to be measured. If the velocities are high, a short time interval between the frames is needed, and if the velocities are small, a low frame rate is
1
www.scioncorp.com
ISES Solar World Congress 1999, Volume III
needed. For the mantle tank experiments in this investigation 1 to 4 frames per second were used. Fig. 6 shows a typical frame from a video recording of particles in the mantle gap. It appears that the particles in the fluid are not significant however, a clear interpretation of the particles can be obtained with appropriate adjustment of the image (Fig. 7). Particle tracking is achieved by plotting a number of flames "on top of each other" as shown in Fig. 8 to give a good interpretation of the flow field. In Fig. 9 the resulting velocity vector field is shown alter evaluation of particle traces using the Scion soltware, which determines the co-ordinates for each particle in the flow field. The basis of particle tracing is to locate particle co-ordinates in 3 consecutive frames separated by equal time steps. Vectors representing the particle movements between pairs of the images can then be calculated as noted in Dahl et al. (1995):
Pi [3 ~CtD2 [2 Xi
t[]l
,
yt~
[3 Yi
Omax
279
is a maximum angular direction change
Poyj, P~j, P~., P~.
are horizontal and vertical distances
respectively between particles [m] These limits are used to avoid errors from false readings caused by particles entering and leaving the illuminated slice of the flow and by that giving a clear illustration of the 2D flow field. The size of the limits depends on the flow structure and velocity.
tal (3)
and p j F] ~cj tD3 [3 x tD2 , y ytD3 rq ytn2
To avoid errors from "false" particles that move in and out of the illuminated section of the flow, the vectors must fulfil certain conditions:
Ip~ rqp j ] [3 zTp.~
Fig. 6: A captured frame from the movie.
and (4)
arctan P c~. V]arctan p ~ V1s , ~ P c~j P ~i where P i , P j are velocity vectors [m/s]
~Oma x is a maximum velocity change [m/s]
Fig. 7: Using the threshold facility to clarify the particle images.
Fig. 8: Particle trajectories illustrate the flow pattern.
Fig. 5: Particle Image Velocimetry equipment.
Fig. 9: The resulting velocity field alter evaluation of particle traces.
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4. FLOW MODELLING A simulation model of the flow in a mantle heat exchanger and inside the inner tank of the glass test tank was developed using the CFD code CFX (CFX International, 1997) to solve the flow and energy equations. A three-block structure was used to model the mantle tank. One block for the top of the inner tank, one block for the mantle and middle part of the inner tank and one block for the bottom of the inner tank. The mesh was concentrated in high gradient regions near the walls. Table 3 shows the number of grid elements used in the model and Fig. 10 shows the vertical grid distribution. Numerical solutions were obtained for laminar flow with the Boussinesq assumption for modelling of buoyancy. Under relaxation parameters were used on velocities and viscosity to achieve convergence. The velocity-pressure coupling was treated using the SIMPLE algorithm and the QUICK diseretisation was used for the convection terms. A time step of 1 second was used. Heat losses from the tank were modelled for an ambient temperature of 20~ combined with an outside heat transfer coefficient of 3 W/(m2K). The mantle inlet temperature, and volume flow rate were specified to match the test conditions. Axis
50~ Case 2 is for an initially mixed inner tank (20~ with the mantle supplied with constant flow and inlet temperature (0.0082 kg/s, 50~ for four hours followed by a step change of inlet temperature to 31 ~ for a further 0.5 hours. 5.1 Casel During the heating, flow patterns in the mantle and inner tank are visualised using the PIV system. Fig. l l shows the conditions during the test. Mean values of the mantle inlet temperature, mantle flow rate and ambient temperature were used as boundary conditions for the CFD modelling. In terms of the flow patterns in the mantle and inner tank, this case is very interesting, as we have a high mantle inlet temperature and an initially cold tank. This could resemble a typical situation in a SDHW system: atter a large draw-off from the heat storage, the sun shines and heats up the collector fluid resulting in a high mantle inlet temperature that heats the cold tank.
CFX Model Top
Middle
Bottom
I 25 46 15 J 40 70 40 K 23 38 23 Total 159,160 Table 3" Number of cells (not including the boundary nodes)
in the CFD model.
Fig.
11:
Experimental
boundary
conditions
and
tank
temperatures at transient Case 1.
Fig. 10: Vertical grid distribution in the CFX-model 5. RESULTS Two heating situations were studied; Case 1 is for an initially mixed inner tank (20~ with the mantle supplied with a constant flow rate of 0.0082 kg/s and an inlet temperature of
5.1.1 Flow dism'bution After 30 min the flow structures were observed for the inlet side, front side and outlet side of the tank. Fig. 12 shows an averaged plot of 60 video images of the particles in the flow separated by 0.2 second and Fig. 13 shows the corresponding vector plot. The plot is in the centre plane of the mantle gap. The flow structure in Figs 12 and 13 shows that the inlet stream initially flows around the top of the mantle, as would be expected. However, the flow then reverses and flows back towards the inlet at a level below the main inlet stream. During this reverse motion the fluid is cooled and drops down the mantle heat transfer surface and develops a return stream along the bottom of the mantle towards the outlet port. The velocity in the middle zone is low as it moves down from the buoyancy driven reverse flow under the main inlet layer at the top of the mantle. In the bottom of the mantle, the velocities are higher as the flow moves towards the outlet port. The maximum velocity in the visualised plane is approximately 8 mm/s. Good agreement was found between the observed and simulated flow distribution in the mantle. The CFX flow prediction illustrated by velocity vectors in Fig. 14, shows the same flow pattern as observed in the P IV measurements. There is the same reverse flow, dead centre zone and suction from the outlet port. The simulation model indicated a maximum velocity of 7.5 mm/s in the mid-plane of the mantle.
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Fig. 12: Casel: The PIV flow field over one half of the mantle circumference. The picture is a combination of 12 PIV scans of the mantle walls in the mantle centre plane. The flow visualisation results from the addition of 60 video frames separated by 0.2 seconds.
Fig. 13: Casel: Measured velocity vectors over one half of the mantle circumference in the mantle centre plane. The picture is combination of 12 PIV scans of sections of the mantle walls.
Fig. 14: Casel: Flow in the mantle centre plane illustrated by velocity vectors predicted by CFX.
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distribution over the inlet wall, front wall and outlet walls are very similar except in the reverse flow layer near the top of the mantle. The heat transfer is highest in the top 10% of the mantle and then decreases as the flow moves down the mantle. Fig. 17 shows the temperature differences between the mantle fluid and the tank wall at the same locations as the heat flux profiles. The temperature difference distribution over inlet wall, front wall and outlet walls are very similar to the heat transfer distribution, which shows that the temperature differences are dominating the heat transfer.
5.1.2 Heat transfer Fig. 15 shows the simulated temperature contours on the mantle side of the heat-exchange wall. Apart from near the inlet and at the very top, the temperature profiles are almost horizontal. Minor distortion of the temperature contours is due to the comers of the square mantle used in this study. Apart from near the top and bottom, the contours are close to each other. This is consistent with the flow structure in Fig. 14, which shows that the middle of the mantle is a uniform low velocity zone. Fig. 16 shows the simulated heat flux profiles down the middle of the three walls of the square test tank. The heat flux
Fig. 15: Casel: Simulation of tank wall temperature contours.
Fig. 16: Case 1: Simulated heat flux profiles down the centre of
Fig. 17: Case l: Simulated temperature difference profiles down
each wall 0.5 hrs after start of heating.
the centre of each wall 0.5 hrs after start of mantle heating. The temperature differences are calculated as the mean mantle fluid temperature in the plane minus the tank temperature.
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5.2 Case2 For this test the tank is heated by a constant flow of 0.5 1/min with inlet temperature of 50~ entering the mantle. After one hour, the inlet temperature is reduced to 31~ which is lower than the outlet temperature at the end of the first hour of heating. The inner tank temperatures (T1, T2, T3, T4 and T5) at this time are shown in Fig. 18 together with the test temperatures and flow rate before and after the step change in the mantle inlet temperature. The flow structure in the mantle was visualised using the PIV system. Mean values of the mantle inlet temperature, flow rate and ambient temperatures were used as boundary conditions for the CFD modelling of this case. In terms of the flow patterns in the mantle, this case is similar to the mantle inlet temperature in solar water heaters on 'typical afternoons', when the irradiance is low. This test was designed to assess if cooler flow into the mantle following a high-temperature heating period will de-stratify the tank. The mantle heat exchanger effectively promotes stratification with inlet temperatures higher than the temperature in the top of the mantle. However in SDHW system operation high inlet temperatures are not guaranteed. Therefore, it is of interest to investigate how the mantle fluid behaves when cooler fluid flows into the mantle following a high-temperature heating period.
283
5.2.1 Flow distribution Fig. 19 shows an averaged plot of 60 video images of tracer particles in the centre plane of the mantle separated by 0.25 second and Fig. 20 shows the corresponding velocity vectors determined by particle tracing. The CFX flow predictions in the middle of the mantle gap are shown in Fig. 21. Due to the lower inlet temperature, the inlet stream drops down immediately as it enters the mantle and passes along the bottom section of the mantle. This stream induces a large re-circulation zone in the top two-thirds of the mantle volume by entraining fluid from the top of the mantle as it drops down the centre of the inlet wall. The PIV results show very chaotic flow under the inlet port. There also appears to be mixing between the re-circulating flow and the chaotic flow beneath the inlet port. This pattern is not seen in the simulation results in Fig. 21 as these calculations are based on a laminar flow model. 5.2.2 Heat transfer Fig. 22 shows temperature contours on the outside of the mantle heat exchanger wall. Below the inlet, the colder incoming fluid drops down the mantle and reduces the wall temperature. However, on the front and outlet side walls, the profiles are horizontal again, which indicates that the flow structure has spread around the mantle. The temperature profiles are distorted at the square corners of the square mantle tank due to mixing of the flow as it passes around the comer.
Fig. 18: Case 2: Experimental boundary conditions and tank temperatures.
Fig. 23 shows the simulated heat flux profiles down the middle of the three walls of the square test tank. The heat flux distribution over inlet wall, front wall and outlet walls are quite different compared to Case 1. The heat transfer is still highest in the top 10% of the mantle, but in the middle of the mantle the heat transfer only varies from 600 to 800 W/m2. For Case 1 the heat transfer varies from 100 to 1,100 W/m2 in the middle zone. For Case 2, Fig. 24 shows the temperature differences between the mantle fluid and the tank wall at the same locations as the heat flux profiles. The temperature difference distribution over inlet wall, front wall and outlet walls are very similar to the Case 2 heat transfer distribution. These results show that the middle zone is mixed to some extent. The mixing is due to the large re-circulation zone induced by the inlet flow dropping down to the bottom of the mantle.
Fig. 19: Case2: Measurement of the flow field in the mantle centre plane. The picture is assembled from 12 sectional pictures of different sections of the mantle walls. The particle streak paths were generated by the addition of 60 video frames separated by 0.25 seconds.
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Fig. 20: Case2: Velocity vectors in the mantle centre plane obtained from PIV measurements.
Fig. 21: Case2: Simulated flow pattern in the mantle centre plane.
Fig. 22: Case2: Simulation of temperature contours on the mantle wall.
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Fig. 23: Case2: Simulated heat flux profiles down the centre of
Fig. 24: Case 2: Simulated temperature difference profiles down
each wall three minutes after the step change in the mantle inlet
the centre of each wall three minutes atter the step change in
temperature.
mantle inlet temperature. The temperature differences are calculated as the mean mantle fluid temperature in the plane minus the tank temperature.
6. DISCUSSION AND CONCLUSIONS The experimental investigations of the flow structure in vertical mantle heat exchangers has shown that for high temperature input into a mantle there is a recirculating buoyancy driven flow in the top 20% of the mantle. Flow continually enters the recirculation zone from the hot inlet point and is then distributed around the circumference of the mantle from the bottom of the recirculation zone. Good agreement was obtained between the measured flow patterns in a mantle tank designed for flow visualisation and a CFD model of the tank. Although the size of this tank is representative of typical commercial mantle tanks it has a higher thermal resistance in the mantle wall than standard steel tank systems. The simulation model was used to investigate the heat flux distribution over the mantle. Very high heat flux levels were indicated in the top 20% of the mantle with an approximately linear variation of heat flux, outside of the recirculation zone at the top of the mantle. The heat flux is uniformly distributed around the mantle outside of the recirculation zone at the top. Experimental and numerical evaluation of the flow structure and heat transfer in a mantle tank when there is a drop of inlet temperature has indicated that colder inlet flow drops down the mantle gap without disturbing existing stratification in the inner tank. The colder inlet flow results in a much deeper recirculation zone in the mantle however, the colder inlet drops down the mantle to its thermal equilibrium level without disturbing existing stratification in the inner tank.
REFERENCES CFX International, (1997): CFX, Release 4.2. 8.19 Harwell, Oxfordshire OX11 0RA, UK Dahl J., Hermansson S.E. and Veber P. (1995): Use of Video-based Particle Image Velocimetry Technique for Studies of Velocity Fields in a Water Heat Storage Vessel. Experiments in Fluids. Vol. 18, pp. 383-388. Dhimdi S. and Bolle L. (1997): Natural Convection: The Effect of Geometrical Parameters. Proceedings of the 4th National Congress on Theoretical and Applied Mechanics, pp. 43-46. Leuvren, Belgium 1997.
E1-Genk M.S. and Rao D.V. (1989): Heat Transfer Experiments and Correlations for Low Reynolds-Number Flows of Water in Vertical Annuli. Heat Transfer Engineering, Vol. 10, No. 2, pp. 44-57. E1-Genk M.S. and Rao D.V. (1990): Buoyancy induced instability of laminar flows in vertical annuli. International Journal of Heat and Mass Transfer, Vol. 33, pp. 2145-2159. Khan J.A. and Kumar R. (1989): Natural Convection in Vertical Annuli: A Numerical Study for Constant Heat Flux on the Inner Wall. ASME Journal of Heat Transfer, Vol. 111, pp. 909-915. Kubair V.G. and Simha C.1LV. (1982): Free Convection Heat Transfer to Mercury in Vertical Annuli. International Journal of Heat and Mass Transfer, Vol. 25, pp. 399-407. Nagendra H.R., Tirunarayanan M.A. and Ramachandran A. (1970): Free Convection Heat Transfer in Vertical Annuli. Chemical Engineering Science, Vol. 25, pp. 605-610. Rogers B.B. and Yao L.S. (1990): The Effect of Mixed Convection Instability of Heat Transfer in a Vertical Annulus. International Journal of Heat and Mass Transfer, Vol. 33, no. 1, pp. 79-90. Shah L.J. and Furbo S. (1998): Correlation of Experimental and Theoretical Heat Transfer in Mantle Tanks used in Low Flow SDHW Systems. Solar Energy, Vol. 64, No. 4-6, pp. 245-25
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A FOR
SYSTEM
DECENTRALIZED F r a n k S~ite,
FOR
SOLAR
APPLICATIONS
PROCESS
HEAT
IN DEVELOPING
COUNTRIES
Bernd Harrier, Klemens Schwarzer
Solar-Institut Juelich, Fachhochschule Aachen (University of Applied Science), Ginsterweg 1, 52428 Juelich, Germany Tel + 49 2461 99 32 37, Fax + 49 2461 99 32 35, [email protected]
Keywords: solar thermal system, process heat, decentralized
systems (e.g. storage-collector). The advantages of the system
system, developing countries
are:
1. Introduction Today firewood is one of the most commonly used energy supplies in developing countries. In many African countries firewood covers more than 90% of the energy demand. The growing population and improvements in the living standard lead to a higher consumption. The result is an abusive cutting of the trees and a deforestation of whole regions ("desertifieation"), for example the Sahel and Sudan zone in Africa. To preserve the environment, the firewood has to be replaced in spite of the growing energy demand. This could be done by using gas or oil but few countries have own resources. On the other hand most developing countries, especially African countries have a high solar insolation. Solar energy could cover a big part of the energy demand. In the regions concerned by desertification firewood is used only in food conservation and preparation. Other technics, like wood-gasifiers for industrial energy supply are not available. So, when talking about the reduction of the firewoodconsumption, you have to think of food-technologies. In this paper a system for solar process heat for community kitchens, bakeries, post harvest treatment and daily production is presented.
2. Process Heat
high security because of the pressureless operation a storage cuts load peaks working in short periods of low radiation or at night temperature may be more than 200~ may be operated as a thermosyphon system a backup heating may be included Three solar collectors, a storage unit, a backup-heating and different loads have been examined in experiments and theoretical investigations (see figure). The components are connected by a circuit with a heat transfer medium (oil). For the design of the system numerical models have been implemented in two different computer programs (TRNSYS and MATLABSimulink).
Systems
Technical systems installed in developing counlries have to fulfill other requirements than those operating in industrialized countries. The reasons are as follows: 9 infrastructure for transport of goods is less developed 9 energy supply is not assured (Isolated grid systems) 9 water supply can be a problem 9 maintenance and supply with spare parts is a problem 9 the climate is less favorable for food preservation Due to the transport problems the systems have to be decentralized. The power demand is lower than for central systems used in industrialized countries. The Systems have to offer a high security, low demand of maintenance and should only depend on locally available energy sources.
3. Solar Process Heat A solar process heat system which fits the demands has been developed at the Solar-Institute Juelich. This has been done mainly in the thesis of one of the authors [Hafner 1999]. A system with separated components and a heat transfer circuit has been chosen. Here it is easily possible to adapt each component to the demand, which is very difficult for combined
Figure 1: The components of the solar process heat system
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4. Solar C o l l e c t o r s Fix-Focus Parabolic Collector A new type of collector, a Fix-Focus Parabolic Concentrator with an absorber has been developed and tested. As reflector, the Fix-Focus Parabolic Reflector of Scheffier [Scheffier 1997] has been used (see figure 2). The specifications of the reflector are given in table 1. Surface Aperture optical efficiency diameter of focus reflector material tracking
7 m2 5.5 to 6.5 m 2 0.6 0.3 m polished aluminium automatic (clockwork)
and closed cavities [Stine 1989]. In some experiments the opening was covered by a glass-window. The performance without window was better, the window has only advantages at temperatures higher than 500~ whereas the collector will be operated at maximum 300~ At these high temperatures the effect of lower heat losses with window is more important than the optical losses of the glass. The second cavity-absorber has a cylindrical geometry and uses all inner wall as aborber. For the design of the cavity-absorber the geometry of the reflector has been investigated with a ray tracing program. As for the first absorber the aim was to reduce the radiation to a value which is not overheating the oil. The result is a cylinder with a 0.35 m diameter and a depth of 0.35rm The experimental results gave a much better performance of this second absorber (see table 2).
Table 1: Specifications of the Fix-Focus Parabolic Reflector
/ / direct
radiation
yes no no
0.3 0.39 0.51
efficiency absorber 0.5 0.64 0.91
T
/
T
/ \
/
//\
/
1 2 3
efficiency absorber + reflector 0.33 0.42 0.55
jacketed vessel
~ \
\
t
window of efficiency glass system
Table 2: Comparison of the two cavity-absorbers
9
fixed-focusparabolicmirror
measurement
T
\
-
The model for the thermal performance has as input the solar radiation and the optical efficiency. For the convective losses a model for open cavities adapted from Leibfried [Leibfried 1994] has been used. The radiative losses are calculated by a radiation exchange between the aperture surface of the opening and the sky. The energy balance is written in a finite-volume model, based on the work of Patankar [Patankar 1980]. It has been shown by Hafner [Hafner 1999] that there is a good accordance for this model with measurements.
Flat-plate Collector
absorber kT
Figure 2: The Fix-Focus Parabolic Concentrator In the absorber the heat flux has to be limited for not overheating the heat transfer medium oil. A cavity absorber is used for this purpose. The oil passes through the jacket of the double-wall of the absorber. Calculations have shown, that a maximum of 1.5 W/cm2 can be removed by free convection in the oil. The flux of 5 W/cm2 in the focus is reduced by the geometry of the cavity. Two cavity absorbers have been tested experimentally. The first one has a flat absorber of 0.5 x 0.5 m ~-which is situated about 0.4 m behind the focus. All other inner walls of the cavity are made of aluminum reflectors. The opening of the cavity has the diameter of the focal area (0.3 m). The results of the measurements showed a poor performance of the absorber. The reason is that the convection in the open cavity still plays an important role for the losses. This fact was also shown by numerical investigations with models for convection in open
The flat plate collector used in the system is a high temperature flat plate collector with double glazing and external reflectors (see figure). This collector has been developed for solar cookers [Hafiaer, Schwarzer 1993]. It has proved its reliability and good performance in several field tests and applications [Schwarzer 1996]. In the experiments the collector has reached 240~ at 1000 W/m2 solar radiation. Table 3 gives the collector performance data achieved in the tests. maximum Temperature (at 1000 W/m2) 240~ heat loss coefficient (k)3.0 W/(m2K) optical efficiency 0.7
Table 3: data of the flat plate collector The numerical model is also based on the finite-volume approach. As the system might be operated at very low flowrates a 2xN model has been chosen. It has been shown [Wittwer 1996] that for this case the 2xN model agrees better with measured data than a simple one-dimensional model. For the calculation of the reflectors a new model has been implemented [Hafner 1999]. The approach is similar to that of [McDaniels 1975], but the model is not limited to east-west orientated reflectors but it may calculate one reflectors at each of the four sides of the collector and the collector or the reflectors might even be tracked during the day.
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5. Parabolic trough collector This type of collector has not been tested experimentally since the performance is well known [Trieb 1995] and numerical models are available. It is used for comparison with the two new collectors in the numerical investigation on the system. A parametric model based on the efficiency of the collector is used which has been adapted to parabolic troughs from the fiat-plate collector model of Isakson [Isakson 1991 ].
6. Storage A pebble-bed oil storage is used in the system for storing the energy from the collector and releasing it dependent on the demand of the load. The pebbles in storage have two functions: The volume of oil used for the system is reduced, expensive heat-transfer oil is replaced by cheep pebbles. The volume of oil in the storage is reduced so the thermal expansion is also lower. As the system used an external expansion vessel, the amount of oil exchanged with the uninsulated vessel is reduced. For the investigation a 360 liter storage has been installed in an experimental setup (see figure 3). The cylindrical storage is equipped with Pt-100 temperature sensors in three radial positions (wall, half-radius, center) and seven horizontal layers equally distributed from bottom to top. It is possible to measure both the vertical and the radial temperature distribution in the storage.
Figure 4: temperature profile at normal charging Figure 3: experimentalsetup for testing the storage Figure 4 shows a cut of the storage and the temperature profile during charging. The temperature of the oil entering the top is higher than the top layer of the storage. The flow entering the storage at the center immediately distributes over the crosssection of the storage. In the ease of an inversed thermocline, if the entering fluid is colder than the top-section of the storage, the cold fluid falls down in the center and distributes only when it reaches a layer with the same temperature (see figure 5).
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shows the temperature in the storage during charging in the seven layers of the temperature sensors and the results of the models with variable and constant density. Specially the bottom layer in the constant model has a big difference to the measured data (up to 20 K, which is 16 % compared to the temperature difference between inlet and outlet).
Figure 6: Comparison of the models with constant and temperature dependent density with measured data At high flow rates the model calculates the heat transfer solid fluid. As shown in figure 7, the single-phase model has big differences to the measurement at high flow rates which occur specially during discharging.
-
Figure 5: temperature profile with an inversed thermocline The pebble-bed oil storage may easily be operated as a stratified charging and discharging storage which has some advantages for solar systems (e.g. low collector inlet temperature). Two models of the storage have been implemented. The first one is a two-dimensional finite-element model which was used for detailed investigations of the inversed thermocline and the position of in- and outlet. It has been shown that even for large diameters (more than 2 m) with the inlet pipe ending at the top center the flow distribution in the storage at normal charging is an almost perfect plug-flow. For the inversed charging there is a narrow column of cold fluid passing through the hot layers until it reaches the layer of its temperature. The second model is, as a result of the first model, a onedimensional plug-flow model. In this model, as in the first one, it was necessary to take the temperature dependency of the density into account. The model with constant density showed big differences in comparison to the measurements. Figure 6
Figure 7: One-phase models compared to a measured discharging of the storage
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The system is located at on 35 ~ latitude (Mediterranean region).
The solar fraction, defined as
f=
Qaux Oconsumption
for the different configurations and conditions is shown in figure 9.
Figure8: Two-phase models compared to a measured discharging of the storage
7. System Optimization With new mathematical models it is possible to design the process heat system according to the weather data and the load profile. The energy consumption of the backup heating has been chosen as the target function to be minimized in an optimization. For the optimization an evolutionary strategy has been used. Conventional algorithms like gradient or simplex algorithms are not functional for the target function with a lot of local minima. The local minima are due to the discretisation of models in space and time. Three different load temperatures which provide a constant load of 2 kW combined with 3 different weather conditions and the three collectors have been used. Table 4 and 5 give the characteristics of the load and the weather data.
Figure 9: solar fraction for the different systems and conditons Figure 10 gives the optimum storage size for each of the collectors at different load- and weather-conditions.
geography: :35 ~ latitude weather :amb. temperature 20~ sum of daily radiation 4, 5 and 6 kWh/(m2*d) time :march (average sun position)
Table 4: weather data 1. inlet temperature desired temperature massflow 2. inlet temperature desired temperature massflow 3. inlet temperature desired temperature massflow
Table 5: load profiles
20~ 100~ 45 kg/h 80~ 120 ~ 90 kg/h 120~ 140~ 180 kg/h
Figure 10: optimum storage size Especially for the flat plate collector there has been a strong influence of the massflow rate on the best system. For lower
ISES Solar World Congress 1999, Volume III
load-temperatures (profile 1) it is a high-flow system with 50 kg/(m2h), for high load-temperatures (profile 2, 3) it is a low-flow system with 15 kg/(m2h).
8. Conclusion It has been shown that even with limited technical possibilities a collector, the fix-focus parabolic collector, can be built with a performance similar to high-quality parabolic trough collectors.
9. Outlook A test of the system under real operation conditions has to be done. A long-term performance test will show the reliability of the system.
Literature Hafner, B. Modellierung und Optimierung eines solar betriebenen ProzeBw~irmesystems, Shaker Verlag, Aachen, 1999 Isakson, Per Matched Flow Solar Collector Model for TRNSYS, Users and Programmers Manual, November 1991 Leibfried, Ulrich Optische und w~metechnische Untersuchung eines parabolischen Solarkonzentrators mit ortsfestem Hohlraumempf~inger, VDI-Fortschrittsberichte Reihe 6 Nr 317, VDIVerlag Diisseldorf, 1994 McDaniels, D.K.; Lowndes, D.H.; Mathew, H.; Reynolds, J.; Gray, R. Enhanced Solar Energy Collection using Reflector-Solar Thermal Collector Combinations, Solar Energy, Vol.17 No.5, p. 277-283, Pergamon Press, New York, 1975 Patankar, S.V. Numerical Heat Transfer and Fluid Flow, Series in computational methods, Hemisphere, New York, 1980 Schwarzer; Hafner; Krings; von Meer Simulation and dimensioning program for solar cookers with temporary storage and comparison with measurement, ISES Solar World Congress, Vol. 5, p. 473-478, Hungarian Energy Society, Budapest, 23.-27.8.93, 1993 Schwarzer, Klemens; Krings, Thomas Demonstrations- und Feldtest yon Solarkochern mit tempor'~em Speicher in Indien und Mali, Shaker Verlag, Aachen, 1996 Schemer, W. Installation and testing of a solar steam generator for cooking in India, GTZ Final Report, 1997 Stine, W.N.; McDonald, C.G. Cavity Receiver Convective Heat Loss, Proceedings, ISES Congress, ISES, Kobe, Japan, 1989 Trieb, Franz Solar Electricity Generation, DLR, July 1995 Wittwer, C.; Rommel, M. Implementierung eines 2x4 Knotenmodells eines Kollektors und die Validierung anhand von MeBdaten bei Normal- und Lo-Flowbetrieb, Sechstes Symposium therm. Solarenergie, Tagungsband, Seite 202, OTTI, 8.-10.5.96, Staffelstein, 1996
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PERFORMANCE OF A CASCADE OF THE FLAT PLATE COLLECTORS TEOLAN TOMSON*
Estonian Energy Research Institute Paldiski Road 1, Tallinn, 10137, ESTONIA Tel.: +372 2 537 097, Fax: +372 6 460 206 E-mail: teolan@,anet.ee Abstract- For the solar domestic hot water (DHW) system performing with serpentine flat plate collectors (FPC) in the single-pass regime, a cascade of FPCs with different properties connected series in the hydraulic duct can be recommended. This cascade provides a gain up to ~10% compared to the conventional DHW system with the medium properties of its FPC. 1. INTRODUCTION It is reasonable to design solar domestic hot water (DHW) systems with the maximum specific yield. One of the possible options here is to use the single-pass regime of heat carrier ("fluid") in a cascade of two (or more) flat plate collectors (FPC). Fig. 1 shows the layout of a cascade. The conventional solar flat plate collector (FPC0) is replaced by two (or more) collectors (FPC1 and FPC2) with feasible selected characteristics connected series in the hydraulic duct. To study the features of a cascade we shall presume their active (specific) surfaces to be equal Ao =Ac2 =Ac0/2 = 1 m2 while they together
equal to the surface area of a conventional DHW system. We shall designate FPC1 as a "cold" and FPC2 a "hot" collector and simulate the behaviour of the system to f'md out the specific yield for a statistical (Estonian, 59~ summer day for the simplifications described in [1]. A peculiarity of the cascade is the same (specific) flow rate mjs valid for both FPC. Theplugflow model of the storage tank behaviour is considered in the present paper. Our task is to provide the increased specific yield of the cascade, which is defined as the ratio ~>1 of produced energies in the cascade and in the conventional system during a certain time interval (hour, day or season).
Solar collectors
m, 60 ~
m, 60 ~
Ac2 = A c l = Aco /2
Solar collector
T (h) + 50 K
Storage tank M
/r( I
T(h) + 50 K
It(h) ~
I
Storage tank M T(h) = 10 ~
T(h) = 10*C
FPC1 ,~
Ac1
m, 1 0 o c
m, 10~ Fig. 1. Layout of the cascade.
2. THEORETICAL BASIS The properties of the cascade will be compared with the properties of the conventional DHW system by the simulation of the useful energy produced for a day and for a current hour during the daytime. As our task is to
" ISES MEMBER
compare the cascade with the conventional DHW system, we shall consider the conventional DHW system to be a "cascade" of two equal FPCs with the properties of FPC0 and calculate its useful energy by the same method.
ISES Solar World Congress 1999, Volume III
The useful energy converted to heat from the solar radiation on the specific surface area during a certain time interval (hour) is accordingly the basic law Q~ = s - u~ . ( r~ - Ta) = b " F~ " ( , a ) - F~ " U~ " ( T , -
ra).
(la)
Here FR'(za)=r/0 and instead o f UL'Tp=FR'UL" 7',. in Eq. (la), we shall consider FR'UL" Ti=kl " Tm as a simplification. Correspondingly
Qum lT']~o- kl "(Tm- Ta)
293
absorber plate temperature Tp describes precisely the instantaneous value of the FPC efficiency. Calculations based on Tm cause some error, but it is small for the FPCs with a long hydraulic duet where (Tp-T,~)<<(Tm-T,.) and which are suitable for the single-pass performance. Due to the comparison of two values derived by the same method, the result is not much influenced by the said simplification. According to the theoretical basis, both particular temperature rise steps can be estimated
(lb)
ATI=[IT'Ool-kll'(Ti - Ta)]/[mf'Cp+ where kl=FR 9Ut, is the slope o f the linearised efficiency curve and Tm is the (fluid) mean temperature. The formula (1 b) is used in the simulation model. On the other hand, useful energy produced during the same time interval can be calculated as
AT2--[IT. no2-k12.(T o - Ta)]/[mf'Cp- k12/2 ].
The flow rate my and temperature rise step A T are ever inversely proportional variables. Investigations show that except the very low values of the specific flow rate my<0.01 k g s l r n 2, the temperature profile along the hydraulic duct of the FPC can be considered linear and therefore the mean temperature of the fluid in the FPC Tm may be calculated via the average value of the temperature rise step A T for each FPC. The
To
= r, + aT1 + aT2
Solahart S-250 TeknoTerm Sonnenkraft Arcon ST NESTE AEE Sole AS BATEC Aidt Miljo Modulsolar Transelektro Transelektro formal formal formal formal formal
Country
Model
Australia Denmark Sweden Austria Denmark Finland Austria Greece Denmark Denmark Italy Hungary Hungary -
M S-250 SK500 ST GVC 2.5 K 16 WASKO BA 22 sel. LF 2.5 LS 1.5 NK- 101 NK-201 a b c d e
(4)
(5)
and its output energy (for the specific surface area) is
(6)
Q , : [mf" Cp . (AT1 + ATE)] / 2.
The system of equations ( 3 ) - (5) is recursive and can be calculated by the iterative method only. The said simulation model is valid in the range my> > k12 / (2-Cp). At my--->k12/(2. Cp) the expected output temperature cannot be provided and have to be reduced to provide the temperature balance.
Table 1. Properties of the marketed flat plate solar collectors. Producer
(3)
The output temperature o f the cascade can be found by
(2)
O~ = m , . G . a r .
k11/2]
Formula of the original efficiency curve 0.82 - 2.8"AT~IT- 0.01 "AT2/IT 0.68 -- 1.9"AT~IT-O.023"AT2/IT O.76 - 2.98"AT~IT- 0.0165"AT2/IT 0.77 - 3.3"AT/IT- O.O08"AT2/IT 0.82 -4.6"AT~IT 0.73 - 3.59"AT~IT- 0.014"AT2/IT 0.87 - 4.AT/IT- O.O3.AT2/IT 0.772 - 3.994"AT~IT-0.0216"AT2/IT 0.77- 5.3"AT~IT 0.7 - 5.2"AT~IT- O.024"AT2/IT
7/o 0.82 0.68 0.76 0.77 0.82 0.73 0.87 0.772 0.77 0.7 0.8 0.85 0.75 0.7 0.7375 0.775 0.8125 0.85
The slope kl, m2Kl 3.2 2.8 3.6 3.8 4.6 4.2 5.1 4.9 5.3 6.2 10 8.5 6.5 3 4.25 5.5 7.25 9
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3. SELECHON OF FLAT PLATES Flat plate solar collectors (FPC) with different properties (Table 1) are available in the market. Their main features can be characterised by the initial value of the efficiency 7/0 and the slope of the linearised efficiency curve k l. The said properties are correlated (Fig. 2). In Table 1 and Fig. 2 the efficiency 7/o and the slope of the linearised efficiency curve k l are given for the formal FPCs 'h"..."e" used at the simulation, too. In Fig. 2 two lines are shown: the statistical trend of the selected quantity of real FPCs (solid line) and the trend (dashed line) of the formal FPCs with the increased slope to study the influence of different FPCs on the cascade.
10
kl, kWm2Kq
optical cover) r/01 is feasible, but this FPC has a rapidly descending characteristic with the high value of k l l and its performance at the inlet temperatures over 30- 40 ~ is ineffective (Fig. 3). For heating water at the inlet temperatures over 20-40~ a FPC with the low value of k12 (having an advanced optical cover, but with the lower initial value of efficiency and ineffective performance at the lower temperatures) is feasible. Fig.3 shows the instantaneous efficiency of FPCs "a. . .,. 'c" and "e" for the temperature rise step AT=25K (valid for the single-pass regime in a FPC with the serpentine piping). The best result can be obtained if the DHW system could use the envelope of said curves. It is possible in a cascade of two (or more) different FPCs. In the case of the single pass system, the
e .-1.0 . p O ~" 4"ss
9
9
d
0.8
C .,.'0"
-,. m .,. . , . .,. .,.
a.,4" .-9- "9~
b
s "''"
9 9
9
9
9
0.6
9
------
a
--- c
0.4 0 0.68
I 0.73
I 0.78
I 0.83
i 0.88
a
@ A T = 25 K
e
170
~T~,K
0.2 5
10
15
20
25
30
35
40
45
Fig. 2. Correlation of the FPC data. Fig. 3. Instantaneous efficiency as a function of the inlet temperature. In the case of the single-pass regime, the inlet temperature of the FPC cascade remains (nearly) constant at the low value ( 7 - 1 5 ~ of tap water. In our calculations we consider the input temperature to be constant T,.= 10~ Also, the outlet temperature is (nearly) constant at the required high temperature level (50-60~ and we consider in our examples the output temperature to be constant To=60~ The high value of the global temperature rise step A T = AT1 + AT2 = 50 K requires the use of FPC 1 and FPC2 with the serpentine piping. To stabilise the outlet temperature for the variable solar radiation conditions, a circulation pump with the regulated speed is required, but it is not a big problem. Nearly constant outlet temperature can be provided in a thermosyphone DHW system with a high internal hydraulic resistance (this value is provided by the serpentine piping of FPC) also. 4. THE ENERGY YIEID OF A SINGLE FPC IN THE TEMPERATURE RANGE
For heating water at the inlet temperatures T~=(7-15)~ a FPC with the high initial efficiency value (having a light
FPC "e" is feasible as FPC1 (Fig. l) for the inlet
temperatures T/1"q'25~ and FPC "a " is feasible as FPC2 (Fig. l) at the
7,2>25~
(which
is
ever
satisfied
T~2-- T i + AT1 > 25 ~
A cascade of both FPCs connected in the hydraulic duct in series makes it possible to use each FPC ("a" and "e") in their preferred range and provides the maximum of specific yield of the system, compared with the same collector surface area of the FPC0 with the medium characteristic '~:" (a compromise between the light and advanced optical cover). The said positive effect can be provided by the calculation of the daily or/and the current energy yield. 5. INFLUENCE OF THE FPC PROPERITIES ON THE DAILY YIELD Two different regimes for the cascade were simulated and compared on the basis of daily energy yield simulation:
ISES Solar World Congress 1999, Volume III
1. The case with constant flow rate mf=const (and nonconstant output temperature To= var). 2. The case with constant output temperature To=const (and nonconstant flow rate my= var). When the average values of output temperatures for both cases are kept equal, the daily useful energy is equal, too. The influence of the FPC characteristics on the calculated daily yield at Irm~=0.6kW/rd 2, Ta= 15 ~ is given in Table 2.
Table 2. Comparison of the FPC data to the energy yield. Conditions FPCI: FPC2:
efficiency factor ~ ("gain"), which is the ratio of the produced energy by a l m2 of the collector surface area during a current time interval (an hour, for instance) in the cascade (AQ,1+AQ,2) to the same in the FPC with the medium characteristic: = (aQ.~ + aQu~) / aQuo >1. (7) The simulation example (Fig. 5) is calculated for FPCs with the formal (but realistic) data: FPCI: 'b", FPC0: '}" and FPC2: "a".
1.4
modest advanced conventional cascade cascade "c . . . . d . . . . e" "c . . . . b . . . . a"
Q., kWhrn "2 if To = const
2.13
2.28
2.42
Q,, kWhm2 if mr= const
2.15
2.26
2.42
average gain (e
1
1.07
1.13
T~(z-
1.21 1.1
.I
10
f
~
~\
60.5
4 "1-
-*--m~2
~- 59.5
2 -[
----rooo
I
8.30
10.30
12.30
14.30
h
Fig. 4. Current output temperature and flow rate in the cascade of similar "00" and diferent "12" FPCs.
6. INFLUENCE OF THE AMBIENT CLIMATE ON THE ENERGY YIEI~ The said positive effect can also be provided by the
: @0.6/15 \\
I\
~
- - - : @0.45/15 - - - : @0.75/15///
\\\ \.~
1/
/
"
." I
8.30
12
\
1.3 ~ - k
1.o
Due to the variable solar irradiation It(h) and ambient temperature Ta(h) during the daytime 8 < h < 16 in the first ease To(h) will vary and in the second case my(h) will vary (Fig. 4). The energy gain of the cascade depends on the inversely proportional data of FPCs, used in the cascade (Table 1). The efficiencies and the particular temperature rise steps are not equal for FPC 1 and FPC2 in the cascade and in case similar FPCs are used in the cascade, the "hot" one works ever ineffectively.
295
I
10.30
I
I
12.30
I
I
14.30
h
Fig. 5. Oltrent efficiency factor of the cascade in a day.
The said efficiency factor ~r increases at the low value of solar radiation being important for the Nordic countries and for the utilisation of solar energy in the morning and afternoon hours as demonstrated in Fig. 5 for the daily performance at the irradiance values I r , , ~ {0.75, 0.6, 0A5} kWrli 2 and ambient temperature Ta= 15~ In Fig. 5 the outlet temperature of the cascade is constant 60~ The dependence of the daily average value of the efficiency factor ~ is presented in Fig. 6 as function of daily average temperature Ta and the maximum value of the solar irradiance Irmax. The character of the functions may to mean that at the lower latitudes a cascade is out of importance, but it's an option to help the utilisation of solar energy in Nordic conditions. An optimisation of the cascade (and its flow conditions) has to be made to find the possible limit of the cascade gain for the selectively matched fiat plate collectors in a DHW system.
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296
30
To, ~
25
I r ~ , kWm "2
"~.4P
20
00, 001, 002
0.9 0.g
15
0.7
10 ~
"~o
9Ir,,~ t
1.1
1.05
0.5
t
1.15
0.6
gain ~'e 1.2
Fig. 6. Average efficiency factor of the cascade depending on climate conditions. CONCLUSIONS In the DHW system performing in the single-pass regime, a cascade of two flat plate collectors with different data is feasible. The "cold" fiat plate collector has to be with light and the "hot" one with the advanced optical cover. The gain of the cascade increases with the difference of parameters of the said collectors increasing. The cascade can be recommended just for the modest solar conditions, available in Nordic countries. The positive effect does not depend on the flow conditions of the heat carder, but the regulated flow rate allows to supply hot water with the constant temperature and the latter could be preferred.
initial value of the
efficiency
~e
ratio of the produced energy: cascade versus conventional DHW system AQ~o,AQ,1, AQ,2 the amount of energy generated during a time interval, kWhm2 temperature rise step of a FPC, temperature aT, aT1, aT2 difference, K Ao Acu, Acl, Ac2 collector area, m2 specific heat, klkgqK1 c. FR collector heat removal factor instantaneous time, hour h (hourly) irradiation on the tilted surface, It, It(h) kWhm-2 irradiation amplitude on the tilted surface, Ir.~ kWhm-2 slope of the linearised efficiency curve, kl, kll, k12 kWm-2K-1 heated water mass, kg m g, g~, g2 storage tank, stored water mass, kg specific flow rate,kgslm "2, l ~ ' l m "2 mf(h), my useful heat energy, kV~qlm"2 Qu useful absorbed energy, kWhm2 S fluid temperature, ~ r(h) ambient air temperature, ~ Ta(h) input temperature of the fluid, ~ T,, T~I,T~ (fluid) mean temperature, ~ T. output temperature of the fluid, ~ To, To(h) absorber plate (mean) temperature, ~ r, uL coefficient of the collector overall thermal losses, kWrn2K1
AcknowledgementsmWe thank the Estonian Scientific Fund for supporting these investigations with the grant G-3133 in 1997 and 1998.
NOMENCLATURE za
transmittance-absorbtance product
REFERENCES Tomson T. (1995) Comparative analysis of conventional and splitted solar domestic hot water systems. Proc. Estonian Acad. Sci. Engin. 1, 2, 183- 198.
ISES Solar World Congress 1999, Volume III
A SOLAR A B S O R P T I O N AIR-CONDITIONING PLANT USING HEAT-PIPE EVACUATED TUBULAR COLLECTORS HE Zinian and ZHU Ning Beijing Solar Energy Research Institute, Beijing 100083, China
Abstract.A solar-powered absorption air-conditioning plant with 100 kW cooling capacity has been successfully designed and constructed in Shandong Province, China. The plant consists of heat-pipe evacuated tubular collector array, LiBr-H20 absorption chiller, cooling tower, water storage tanks, circulating pumps, fan-coil units, auxiliary oil-burned boiler and control system. The solar collector
array using 2160 heat-pipe evacuated tubular has a total aperture area of 540 m 2 .This paper introduced design characteristics and measuring performance of the plant which has a multifunction of space cooling in summer, space heating in winter and domestic water heating in other seasons. Thermal efficiencies of the collector array are respectively 40% for space cooling, 35% for space heating and 50% for domestic water heating. It was found that the cooling efficiency for the entire system is around 20%.
1. INTRODUCTION Electric power required for providing airconditioning takes a very large portion of total electric power consumption in the world. For this reason, various solar-powered air-con- ditioning systems have been investigated (Bong et al.,1992.Yeung et a1.,1992. George Lof,1993. Back et a1.,1997) Among them, absorption air-conditioning systems were commonly utilized. The solar cooling has an obvious advantage that the most cooling demand is matched with the strongest sunshine in summer. Besides, solar absorption cooling can be combined with solar space heating and solar water heating so that this comprehensive system will increase the economic benefit of solar air-conditioning. The LiBr-H20 absorption chillers have been widely commercialized. The chiller requires moderately high inlet temperature. In view of thermal performance of the system, a higher inlet temperature will result in a higher COP value. Fortunately, heat-pipe evacuated tubular collectors, developed by Beijing Solar Energy Research Institute (BSERI) and recently produced by Beijing SUNDA Solar Energy Technology Co., Ltd. in China, can meet this requirement ( He,1997 ). They have been used for a solar absorption air-conditioning plant incorporated with a LiBr-H20 absorption chiller at an operating temperature about 88~ This paper described the design characteristics of the completed solar air-conditioning plant, and the primary measuring results under conditions of space cooling,
space heating and domestic water heating.
2. SITE OF THE PLANT The solar absorption air-conditioning plant has been successfully designed and constructed in Yintan Tourist & Holiday-Spending Zone of Rushan, Shandong Province, China. Rushan is located at the southeast end of Shandong Peninsula, which is 360north latitude and 1210east longitude. It is 100 km from Weihai in the east and 150 km from Qingdao in the west, bordering on the Huanghai Sea in the south. In this area, the annual average daily solar radiation is around 17.3 MJ/m 2, the annual average air temperature is around 12.3~ maximum air temperature in summer is 32.1~ the minimum air temperature in winter is-7.8~ Under the local climate, both space cooling and space heating are required for comfort in summer and winter. The solar air-conditioning plant was installed in the Solar Energy Hall of the Chinese Renewable Energy Popular Science Park in this Zone. The Hall is a two-storey building with a construction area over 1000m2 and was architecturally designed to meet requirements of solar collector placement, as shown in Figure 1.
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3. SYSTEM DESIGN
3.1 Layout
Fig. 1. Full view of the solar air-conditioning plant in Rushan, Shandong Province, China The solar air-conditioning plant consists of evacuated tubular solar collector array, LiBr-H20 absorption chiller, cooling tower, water storage tanks, circulating
pumps, fan-coil units, auxiliary oil-burned boiler and control system. Figure 2 shows a layout of the plant.
Fig.2. Layout of the solar absorption air-conditioning plant
ISES Solar World Congress 1999, Volume Ill 3.2 Solar collector array In order to supply the absorption chiller with higher inlet temperature, heat-pipe evacuated tubular collectors have been used. Each evacuated tube has an outer diameter 100mm and a length 2000mm.
The evacuated tube is mainly composed of heat-pipe absorber plate, glass envelope tube, metal sealing cover, getter and others, as shown in Figure 3. The heat-pipe consists of an evaporator section and a condenser section.
To obtain more solar irradiation over a day, evacuated tubes with a semicylindric absorber plate were selected. Theoretical calculation and measuring data show that
the semicylindric absorber plate increases energy gain 10-14% more than the flat absorber plate (He et al., 1997).
1
2
3
4
5
-I i i
J
1. condenser section 3. glass envelop tube 5. evaporator section
2. metal sealing cover 4. absorber plate 6. getter Fig.3. Configuration of heat-pipe evacuated
As heat-pipe technology is applied to evacuated tubes as well as "dry connection" between evacuated tubes and manifolds is utilized to modules, the heat-pipe evacuated tubular collector has many advantages, such as freeze resistance, fast start-up, high pressure bearing, thermal shock endurance, etc. The solar collector array using 2160 heat-pipe evacuated tubes has a total aperture area of 540 m 2 and a total absorber area of 364 m 2. The collectors were arranged in 9 rows, in which 7 rows were installed on an inclined south-facing roof and 2 rows were installed on a fiat roof with a tilt angle 350. To reduce flow resistance within the system, front 4 rows and back 5 rows were respectively connected in parallel. Then these two parts were connected in series. The solar system is driven by the circulating pump P1.
3.3 Cooling chiller The air-conditioning plant applies a Model LCC-03 lithium-bromide absorption chiller that is made by Dalian SANYO Refrigeration Co., Ltd in China. The chiller has a maximum cooling capacity of 176 kW ( 50 USRT ). The solar collector away supplies hot water at 88. to an inlet of the chiller and the water leaves the chiller at 83.. The chiller produces chilled water at 8. and the water returns to the chiller at 13.. The cooling water temperature through the chiller is 31. and 37. successively. The hot water, chilled water and cooling water through the chiller is respectively driven by circulating pumps P2, P3 and P5. 3.4 Storage tanks There are totally 4 storage tanks in the plant. Volumes of tank 1, tank 2, tank 3 and tank 4 are successively 8 m 3, 4 m 3, 6 m3and 10m 3.
The tank 1 and tank 2 are called as hot water storage tanks and used to store the hot water produced by the solar collector array. The smaller tank 2 aims to reach a specified temperature for the chiller in the early morning. The tank 3 is called as a chilled-water storage tank and is used to store the chilled water to reduce heat losses of the storage tank because temperature difference between the ambient temperature and the chilled water temperature is much smaller than that between the hot water temperature and the ambient temperature. The chilled water between the tank 3 and fan-coil units is driven by the circulating pump P4. These three tanks are also applied to store the hot water for space heating in winter. The largest tank 4 is called as a domestic hot water storage tank. It supplies hot water by means of a heat exchanger inside the storage tank. It also can be used to rescue the plant from overheating when necessary in summertime.
3.5 Auxiliary boiler To ensure an all-weather operation for the plant, the auxiliary oil-burned boiler was installed. The boiler has a rated thermal power of 350 kW with a rated outlet water temperature of 95~ A designed thermal efficiency of the boiler is approximately 88%. 3.6 Controlsystem The control system consists of temperature sensors, electro-actuating valves, network control modules and operator working station. There are totally 9 temperature sensors, 2 three-way valves and 14 two-way valves in the system. Positions indicated by the temperature sensors are as follows: T1 outlet from collector array
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300 T2 T3 T4 T5 T6 T7 T8 T9
outlet from tank 1 inlet to collector array water in tank 2 water in tank 1 inlet to chiller outlet from tank 3 inlet to tank 3 water in tank 4
The control system has three main functions: First, it serves to open / close valves and switch on / off circulating pumps to operate the plant for space cooling or space heating or domestic water heating as well as to arrange storage tanks for minimizing heat losses. Second, it serves to light up / off the auxiliary boiler to compensate the thermal power provided by the solar collector army when water temperature is below the specified level. Third, it also serves to prevent the plant from overheating when necessary in summertime and from freezing when required in wintertime.
4. MEASURING PARAMETERS AND INSTRUMENTS In order to determine performances of the plant, following parameters must be directly measured: solar irradiance, inlet / outlet temperatures and water flowrates for the solar collector army, cooling chiller and fan-coil units, water temperatures within storage tanks and ambient temperature. Solar irradianee is measured by an EPPLEY Model PSP pyranometer on the plane of the solar collector army. The pyranometer has linearity of ~0.5% and temperature dependence of fi1%. Inlet / outlet water temperatures for the solar collector army, the cooling chiller and fan -coil units are measured by Pt 100 resistance thermometers. The accuracy of the thermometer is ~0.1~ Water flowrates for the solar collector array, cooling
chiller and fan-coil units are measured by vortex flowmeters. The flow-meter has an accuracy of ~1%. A c o m p u t e r program has been developed for data acquisition and processing so that not only instantaneous flowrate but also accumulated flowrate can be simultaneously read out. Furthermore, instantaneous heat flux and accumulated heat flux can be immediately calculated, incorporating the instantaneous flowrate with the corresponding temperature difference between inlet and outlet of each facility. Data acquisition for solar irradianee, water temperatures within storage tanks and ambient temperature are conducted by a DATATAKER Model DT 600 data acquisition device.
5. RESULTS AND DISCUSSION The solar absorption air-conditioning plant has been operated since November 1998. Its performances for different purposes were measured: space cooling in June 1999, space heating from January to March 1999, and domestic water heating in May 1999.
5.1 Spacecoolingperformance In June 1999, the measuring data indicated that the outlet temperature was generally over 88.on a sunny day, which is necessary for operation of the chiller. Figure 4 shows the variation of solar irradiance with time on June 25. The solar irradiance ranged from 380W/m 2 to 1000
W/m 2 during the day.
The
maximum value at noon was 1047 W/m2. The solar isolation from 8:30 to 16:30 was 20.3 MJ/m2. The variation of solar power gain with time on the same day is shown in Figure 5. The total solar energy gain was 3264.9 MJ.
Fig. 4. Variation of solar irradiance with time.
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301
200
~.~
150
v
._c tO C~
8
0
7:12
8:24
9:36
10:48
12:00
13:12
15:36
14:24
16.'48
Time (hours)
Fig.5. Variation of solar power gain with time As a definition, thermal efficiency of the solar collector array is the total solar energy gain divided by the product of the solar isolation and the total absorber area. Using all above measuring data, the collector array thermal efficiency was 44.5%. Figure 7 shows the cooling power in the whole day. It varied from 20kW to 90kW. This implies that the
The variation of the collector exit temperature with time is given in Figure 6. As expected, the collector exit temperature was kept at around 85~ during most of the day on June25. coefficient of performance (COP) of the chiller ranged from 0.4 to 0.7.
160.00
1
140.00
120.00
loo.oo t_ O~
80.00
0')
.E 0 0
6O.OO
o
40.00
20.00
0.00
I
7:12
8:24
I 9:36
I
I
I
I
I
I
10:48
12:00
13:12
14:24
15:36
16:48
18:00
~rne (hours)
Fig. 7. Variation of cooling power with time The cooling efficiency will be the cooling energy divided by the solar isolation. On June 25, the cooling efficiency of the entire system was 22%. The fact that the cooling efficiency of the entire system is higher than that published in previous literatures, is owe to the higher thermal efficiency of the collectors and the higher COP of the chiller.
Figure 8 is the temperature variation in the tank 2 on June 25, It can be seen that hot water temperature within the take 2 rose very fast and reached 88~ at 9:40without supply of the auxiliary boiler, which met the requirements of the chiller.
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302
o
,-.9
80
f
J
'
I
70
60
I
I
8:09
7:55
I
8:24
I
8:38
I
8:52
9:07
I 9:21
I 9:36
I 9:50
10:04
Time (hours)
Fig. 8. Temperature increase in tank 2
5.2 Spaceheatingperformance The measurements of solar irradiance, solar power gain, collector exit temperature and temperature in the tank2 for space heating, with the same methods as that for space cooling, were conducted from January to March 1999. Similarly, thermal efficiency of the
collector array can be calculated. Table 1 gives the solar isolation, solar energy gain and collector thermal efficiency on Jan.15, Jan.16, Feb.25, Feb.26, Feb.27 and Mar. 1. It indicates that the thermal efficiency of the collector array is approximately 35% for space heating.
Table 1 Solar isolation, solar energy gain and collector thermal efficiency Solar energy gain
Solar isolation (MJ/m2)
Date
1859.9 1840.7 953.3 1251.7 2966.6 2175.3
16.41 15.93 7.85 9.75 24.45 17.13
Jan.15 Jan.16 Feb.25 Feb.26 Feb.27 Mar.01
Collector efficiency (%)
(MJ) 31.2 31.8 33.5 35.3 33.2 35.0
Lowest ambient Temperature (*C)
-2.9 -2.1 4.8 7.2 0.5 6.9
within the tank 2 on Feb. 27. Hot water temperature in the tank 2 also increased quickly Fig. 9. Temperature increase in tank 2 (Feb. 27) without supply of the auxiliary boiler.
58.0.
54.0.
5.3 Domesticwater heatingperformance
52.0-
The similar measurements and calculations have been done for domestic water heating purpose. On May 25, water temperature inside the tank 4 changed from 18.2~ at 8:00 to 44.2~ at 10:30, witch was suitable for domestic use. The water temperature increment in the tank 4 along with solar isolation is shown in Table 2.
0 "-'50.0 ~.48.0, I-46.0 44.042.0-
40.o 8:24
i
8 38
i
8 52
'
9:.07
'
9".21
~
9:36
'
9.50
~me (Hou
Figure 9 shows the temperature increase
10:04
At that moment, the solar isolation was 5.83 MJ/m2, the solar energy gain was 1088.6 MJ, and thus the collector thermal efficiency was 51.3%. According to these data, it is evident that the plant can supply approximately 32 m 3 of domestic hot water above 440C per day.
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Table 2 Water temperature in tank 4 and solar isolation (May 25 )
Time 8:00 8:30 9:00 9:30 10:00 10:30
Upper 19.8 25.6 30.3 35.1 40.1 45.7
Water temperature(~ Lower Middle 16.4 18.5 21.8 24.3 26.6 29.3 31.7 34.2 36.8 39.2 42.2 44.7
6. CONCLUSION The solar absorption air-conditioning plant with 100 kW cooling capacity using heat-pipe evacuated tubular collectors has been designed, constructed and operated in Shandong Province, China. The performance of the plant has been primarily measured and analyzed. The operating results show that the solar absorption air-conditioning plant can be used for space cooling in summer, space heating in winter and domestic water heating in other seasons. This multifunctional plant increases the economic benefit of the solar airconditioning system. The measured thermal efficiencies of the solar collector array are approximately 40% for space cooling, 35% for space heating and over 50% for domestic water heating. It indicates that heat-pipe evacuated tubular solar collectors are suitable for applications at relatively high operating temperature and quite low ambient temperature. As a smaller storage tank is specially adopted, hot water temperature can be raised from 70~ to 88~ in the early morning in summer to meet the requirements of the chiller, and from 40~ to 55~ in the early morning in winter for space heating. The chilled-water storage tank is obviously useful for reducing heat losses of the tank because temperature
Average 18.2 23.9 28.7 33.7 38.7 44.2
Isolation (MJ/m 21 0.00 0.94 1.94 3.02 4.31 5.83
difference between the ambient temperature and chilled water temperature is much smaller than that between the hot water temperature and the ambient temperature. The measured COP of the chiller is approximately 0.70 under solar-powered conditions. The cooling efficiency of the entire system is around 20% under local circumstance REFERENCE Back N.C., Shin U.C. and Jeung S.H. (1997) Study on the Solar Absorption Cooling and Heating System, Proceeding of lSES Solar World Congress 4, Taejon, Korea. Bong T.Y., Ng K.C. and Tay A.O. (1987) Performance Study of a Solar-powered Air-conditioning system, Solar Energy 39, 173-182. George Lof (1993) Active Solar System, The MIF Press, Cambridge, London, England. He Z.N. (1997) Development and Application of Heat Pipe Evacuated Tubular Solar Collectors in China, Proceeding of lSES Solar World congress 2, Taejon, Korea. He Z.N., Ge H.C., Jiang EL. and Li W. (1997) A Comparison of Optical Performance between Evacuated Collector Tubes with Flat and Semicylindric Absorbers, Solar Energy 60, 109-117. Yeung M.R.,Yuen P.K.,Dunn A.and Cornish L.S.(1992) Performance of a Solar-powered AirConditioning system in Hong Kong, Solar Energy 48, 309-319.
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ADVANCED FUZZY CONTROL OF THE TEMPERATURE IN THE TEST CHAMBER
Borut Zupan6i6 and Igor ~;krjanc Faculty of Electrical Engineering, University ofLjubljana, T1-2a~ka25, 1000 Ljubljana, Slovenia, Tel. No.: +386 61 1768 306, Fax No.: +386 61 126 46 31, Email: [email protected]
Aie.~ Krainer and Bo.~tjan Furlan Faculty of Civil Engineering, University ofLjubljana, Jamova 2, 1000 Ljubljana, Slovenia, Tel. No.: +386 61 1768 604, Fax No.: +386 61 125 06 88, Email: [email protected]
Abstract-The paper deals with a systematic approach to the control system design with a final goal to efficiently control some living conditions in a test chamber. The first step of each systematic approach is mathematical modelling. It was based on theoretical approach, which means that the model was developed on the basis of physical laws, energy equilibrium laws etc. The mathematical model was implemented in MATLAB Simulink environment as a simulator entitled KAMRA. The structure is modular, the robust numerical algorithms give accurate results and fast simulation runs. As simulator is implemented in MATLAB environment, it is very easy to implement arbitrary pre simulation and post simulation processing. The validation was made by real measurements as test chamber is equipped with all needed sensors. The developed simulator gives an ideal environment for the design and validation of different control structures. Four fiuzy logic controllers were proposed for efficient control of indoor temperature: for heating, for cooling, for coordination of both subsystems and for roller blind positioning. However a good dynamic response to the reference changes and appropriate disturbances elimination was only one requirement. Energy consumption and some other comfort living conditions (e.g. daylight illumination, ...) were also taken into consideration. Simulation gave applicable information for control system implementation with industrial hardware.
1. INTRODUCTION From prehistoric times bioclimatic conditions in buildings were of extreme importance for pleasant and healthy feeling. As such they represent a process with inexhaustible possibilities for the studying of new control design approaches. The development of new information technology enables new artificial heating and cooling solutions and realisations and harmonized combination with natural resources. So the gab between natural and artificial environment is becoming smaller and smaller. Recent outcomes of the control engineering area are more and more applicable on different areas due an incredible development of software and hardware technologies. Beside traditional PID algorithms which are implemented in industrial hardware for many years, it is also possible to use more advanced techniques, e.g. fuzzy logic, neural nets, expert systems, genetic algorithms, adaptive and multivariable control etc.
There are several important steps in the so called control system life cycle. In the design phase mathematical modelling is an unavoidable phase at least for moderate and severe process complexity. It is also very important that mathematical model is developed with particular aims. Of course models must be verified and validated by extensive number of process measurements. Only well tested models can be transformed into user friendly simulators and efficiently used in control design procedures. So the first stage of control system design is performed by the aid of process simulator. Different control strategies from traditional PID controllers, ON/OFF controllers, lead-lag compensators to more sophisticated approaches which originate
from artificial intelligence area e.g. fuzzy logic or neural net controllers can be tested in simulation environment. The solution which gives the convenient results with regard to design specification (e.g. fast response, zero steady state error, efficient disturbance rejection, ...) is finally implemented with industrial hardware. In this stage many additional functions must be included beside by simulation developed control algorithm. This are functions for operator's interaction, for safe and reliable operations, for monitoring and some more complex supervision functions, for fault detection and diagnosis etc. Unfortunately appropriate simulation results do not guarantee that the control system will satisfactory operate with industrial hardware and real process. Most frequently the reason is a bad mathematical model. However the reason can be also in significant differences between control algori~m implementations in simulation and implementation environments (Zupan6i6, 1998b). 2. MODELLING AND SIMULATION 2.1 Principles of theoretical modelling The designed building simulator (entitled A M R ) is based on theoretical mathematical modelling approach. From many reasons mathematical models are the most suitable and the most widely used category of models. They are concise, unambiguous and unique interpretable, while their manipulation and the evaluation of alternatives are relatively inexpensive. Mathematical model can be defined as a mapping of relationship between physical variables of a system to be modelled into corresponding mathematical structures. The essence to theoretical mathematical modelling lies in the
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decomposition of the studied system into particular subsystems, which must be as simple as possible. The corresponding relations between chosen subsystems must then be determined on the basis of different balance equations and physical laws for the area under investigation. In the case of technical systems modelling, the known mass, energy and momentum balances are most frequently used, which gives the overall model expressed by difference or differential equations (Matko et al, 1992). The theoretical modelling of heat dynamics of a room was based on the analyse of thermal conduction, thermal convection and solar radiation and on appropriate energy balance equations (Kladnik, 1987, Krainer and Kladnik, 1995, Zupan6i6 et al, 1998a, Furlan et al, 1998). The properties of the envelope are treated as time-varying parameters as they are variable by their own nature. The variable nature is especially worth for the openings in the building envelope. Frequently openings are equipped with different less or more sophisticated shading systems (e.g. roller blinds), what enables at least changing the shading ratio of opening or even their geometry. In some cases high-tech glazing (electrochromic, fotochromic .... ) are used, where optical characteristics of glass could be changed. The automatic adaptation of such envelope properties appears as a new great opportunity of indirect controlling of the indoor living space parameters according to the current outdoor conditions. 2.2. Important features of the simulator KAMRA Simulator KAMRA can be used for different purposes. In this case it was used for control system design purpose, so only input/output relations will be presented in more detail. The inputs of the simulation model are the outside conditions as well as dynamical parameters of envelope: Variable outdoor (weather) conditions: 9 the outdoor air temperature, 9 the temperature of the terrain, 9 global solar radiation, 9 level of cloudiness and 9 ratio of diffuse/direct radiation. Changeable properties of the building's envelope are: 9 the opaque elements: thermal capacity and resistance of these elements can be changeable, 9 the transparent elements (windows): geometry of openings, optical characteristics of glass and resistance of fill between glass panes are variable, 9 interior properties: absorption, emission coefficients of walls and thermal capacity of furnishing are variable, 9 other characteristic: changeable orientation, Additional heating and cooling: the power of heater and ventilator. The outputs of the simulation model are: 9 the indoor air temperature and interior heat flow 9 the walls, windows and surface temperatures There are following important features of the simulator: 9 It is possible to simulate rectangular building with arbitrary walls, floor and ceiling composition. 9 The opaque elements of the building envelope are floor, ceiling, walls and they are composed of 5 layers, which enables adequate thermal description of different envelope structures.
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In each wall one window of rectangular shape could be placed. All windows in the model are supposed to be double-glazed and filled with different gases. 9 The inner space of the building can contain furniture and equipment. The ratio of furnishing/surface of the envelope is flexible, the material properties of furniture are optional. 9 The solar radiation is composed of direct radiation and diffuse solar radiation. The ratio direct/diffuse radiation (DDR) in the model is flexible. 9 The level (CLD) of cloudiness is also an attribute of the outside conditions, as it affects the final amount of direct and diffuse radiation. 9 The orientation of the building is optional parameter and it is defined by the declination angle between real (geographic) south and the direction of building' s axes. The following suppositions are considered in the mathematical model: 9 The whole mass of the inner air is supposed to have uniform temperature. In reality the temperature of the air in the inner space is position dependent function, but the temperature used in calculation is an average temperature of the whole air mass. 9 Temperature changing in directions along the wall or window surfaces are neglected, thus the conduction problems through the envelope elements can be treated as one-dimensional crosswise through them. 9 Whole mass of furnishing/equipment is heated only by the surrounding air. ARer the development of the mathematical model and the simulator concept the appropriate programming tool had to be selected. The most important requirements for the appropriate selection were as follows: 9 Modular and transparent syntax. Model can easily be understand and modified as well. 9 Modern graphic user interface should enable that modelling and simulation unskilled users can efficiently experiment with the model. Users must concentrate to thermal problems instead of problems with modelling, simulation, programming etc. 9 High numerical accuracy and robustness. 9 Fast simulation. 9 Portable models. The selected environment should be a widely spread one, used not only on academic institutions but also in industry. So developed models can be easily transferred between different computers, groups or institutions. 9 With regard to the developed simulation concept the capability for the inclusion of continuous and discrete submodels into the simulation model must be presented. 9 In the chosen environment different toolboxes must give powerful possibilities not only for simulation but also for analysis, design, graphical results presentation etc. 9 If possible, control structures obtained by off line simulation and design can be automatically coded for appropriate target hardware giving efficient real time implementations (Zupan/~it, 1998b). 2.3. Implementation in MATLAB-Simulink environment
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Having in mind the itemised requirements the mathematical programming environment MATLAB-Simulink with appropriate Toolboxes (for control, fuzzy logic .... ) was used. Fig. 1 shows the highest hierarchical level of the modularly constructed simulation model, which is already prepared to serve as a test cell for control system development and validation.
of Ljubljana. Fig. 2 shows the scheme of the test chamber with basic sensors and actuators used in the verification and validation of the simulator KAMR .
Fig. 2. Scheme of the test chamber with basic sensors and actuators used in the verification and validation of the simulator KAMRA.
Fig. 1. Simulation scheme of the simulator in MATLABSimulink environment. In the block Initialization all the parameters about the materials, geometry of window, orientation, geographic location and starting simulation time are given. So the simulation of the behaviour in the case of different materials, orientations, geographic location, position and number of windows and period of the year can be performed. Outdoor temperature and Global solar radiation are defined with appropriate data files obtained from real measurements. They can easily be defined by some other signals from Simulink library. Temperature of terrain, Direct radiation and Cloudiness are prepared as constants or step signals. However in the presented model Direct radiation and Cloudiness are not independent (e.g. if the parameter of direct radiation changes from 0 to 1, the parameter of cloudiness changes from 1 to 0). Up to now described signals are treated as disturbance inputs. The last three inputs signed with Heater, Ventilator and Blind are control inputs as they will be fed by controller signals in order to assure the appropriate indoor temperature. Indoor temperature is the model output or from the point of control system the controlled variable. Of course the simulator can be easily modified so that also other variables of the model can be influenced or monitored. 2.4. Simulator validation The verification and validation of the simulator is one of the most important tasks in each modelling cycle. It is mainly based on the comparison of the measured and simulated results. The verification and validation of the simulator KAMRA was performed with a real system (chamber)- a testing cell built on the roof platform of the Faculty of Civil Engineering, University
The test chamber is a box with all dimensions lm. The south wall is completely glazed, double-glazing is composed of two layers of standard clear glass and air fill, the thickness of wooden frame is 5cn~ The roller blind is as external PVC blind and the alternating window geometry was realised by moving the blind to desired position. Walls, floor and ceiling are composed of dry wall panel lcm, mineral wool 8cm dry wall panel 2cm (from outside). Internal walls are painted in light grey colour. The box is shifted off the ground and the roof is ventilated in order to avoid overheating caused by direct radiation on the roof. Measured values for outdoor conditions were global and reflected solar radiation and outdoor air temperature. Pyranometer CM-6B (Kipp&Zonnen delft BV) was used for measuring solar direct and reflected radiation. Termocouples type T was used for measuring temperature. The temperature of indoor air defines thermal response of the object and it was also measured with termocouples type T. Window size was expressed as ratio of shaded area and whole glazing area. For the purpose of collecting of different samples of the outdoor environment conditions, some series of measurements were executed in different seasons of the year. Position of blind was changed randomly in different time intervals independent of the outdoor conditions. Several measurements were used for appropriate final parameters tuning of the theoretical model of the test chamber. Another set of measurements was used for simulator validation. Simulations were obtained with the measured outdoor temperature and global solar radiation as input variables taken from the experiments as well as with the signal for blind moving r e , m e (see Fig. 3). The comparison of the simulated indoor temperature and the measured one is presented in Fig. 4. The error between calculated and measured values is acceptable in the range of 5-20%. Mainly it is caused by unexpected ventilation heat-losses through some cracks in the dry wall panels and by the influence of wind.
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temperature, radiant flows . . . . ). On the other hand feedback control can also be implemented on numerous ways. From simplest ON/OFF control algorithms to widely used PID algorithms, compensators, cascade controlers, fuzzy controllers to most sophisticated adaptive, predictive and multivariable controllers. The latter are probably not justifiable as the process is not enough complicated and the solutions would be too expensive. Among other possibilities preliminary studies (Zupan~i~, 1998a) indicated that conventional approaches (e.g. PID, ON/OFF . . . . ) give acceptable results in conjunction with particular working regimes. But due to very different regimes, nonlinear behaviour, time varying characteristics, several conventional controllers with appropriate switching mechanisms should be developed. Our preliminary investigations confirmed that fuzzy control approach could represent more appropriate solution. 3.2. Fundamentals of fuzzy logic Fig. 3. Blind position, global solar radiation and outdoor temperature.
Fig. 4. Comparison of the measured and simulated indoor temperature.
The development of fuzzy logic in the late 70s has provoked many applications and enlargements of a simple fuzzy controller (Babu[ka, Verbruggen, 1996, ~krjanc et al, 1996, ~krjanc et al, 1997a, ~krjanc et al, 1997b, Kav~ek et al, 1997). Further investigations pointed out certain advantages over traditional (e.g. PID) approaches. Beside satisfactory robustness in case of more complicated nonlinear and time varying systems probably the most important advantage refers to design principle, which is very similar to human reasoning. This principle is based on simple conditional rules. The design and tuning of fuzzy controllers consists of membership function design and definition of fuzzy rules. However no general tuning methods exist so the design approach is based on real or simulation trial and error experiments. So the control system can be designed also without precise mathematical model (with direct experiments on real process). Using Fuzzy control toolbox ANFIS in MATLAB environment the trial and error experimentation approach is very easy. The designer must select appropriate membership functions, fuzzy rules and some other parameters. With this the controller is automatically prepared for Simulink simulation environment. So it can be easily combined with the simulator KAMRA.
3. CONTROL SYSTEM DESIGN AND VALIDATION
3.3. Proposed control scheme
As mentioned the final aim was to develop a control system for pleasant and comfort living conditions but also for economic energy consumption. The comfort was mainly determined with indoor temperature. However the appropriate daylight was also important as it can significantly influence comfort behaviour and energy consumption as well. It is well known that optimization of each technical system starts with different and contradictory demands so the most important step in the design procedure is to choose appropriate criterion functions.
Fig. 5 depicts the Simulink simulation scheme of heat dynamic control system. The heater and ventilator are controlled by two fuzzy systems, each of them consists of two fuzzy controllers. The inputs of both fuzzy systems are control error and error derivative. With such approach a fuzzy logic controller can be treated as a dynamic controller. Obviously heating and cooling (ventilation) are not simultaneously in operation. This operation is harmonized by the third fuzzy feedforward controller entitled fuzzy logic coordinator. The output of this system is the third input of the fuzzy controllers for heating and ventilation. The blind positioning is controlled by the fourth fuzzy logic feedforward controller with the following inputs: solar radiation, direct solar radiation and day index (indication of summer and winter period respectively).
3.1. Possible control strategies Different control strategies can be taken into account: from feed forward open loop control schemes to closed loop regulation (feedback control). Feed forward is usually a cost effective solution but in our case unsuitable as the system is exposed to significant disturbances (time varying external
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For the clearness and simplicity only two inputs, which are important for dynamic characteristics - control error and its derivative are presented. Appropriate membership functions and fuzzy rules were obtained with many experiments. Fig. 7 shows the membership functions for Fuzzy PD part for control error, for its derivative and for controller output (heating power).
Fig. 7. Membership functions for PD controller.
Fig. 5. MATLAB Simulink simulation scheme of the heat dynamics control system. 3.4. Fuzzy logic controller for heating Fuzzy logic controller for heating consists of a parallel structure of proportional derivative (I'D) and proportional integral (PI) part. It is a kind of fuzzy PID controller. PD part enables fast response and appropriate damping and PI part eliminates the steady state error. The appropriate simulation scheme is presented in Fig. 6.
Fig. 6. Simulation scheme of the fuzzy logic controller for heating.
The mnemonics in conjunction with membership functions have the following meaning: N negative error ZE zero error P positive error DEC decreasing error INC increasing error ZE zero heating LH low heating HH high heating After the inputs and output are appropriately determined with fuzzy sets, the appropriate rules between inputs and output should be defined. For Fuzzy PD part 6 rules were used: 1. If (E is N) and (ED is DEC) then (HPD is ZE) 2. If (E is N) and (ED is INC) then (HPD is ZE) 3. If (E is ZE) and (ED is DEC) then (I-IPD is ZE) 4. If (E is ZE) and (ED is INC) then (HPD is ZE) 5. If (E is P) and (ED is DEC) then (I-IPD is LH) 6. If (E is P) and (ED is INC) then (HPD is H ~ E means error, El) means error derivative and HPD means heating (contribution of PD part). Beside membership functions and fuzzy rules, the following methods in fuzzy approach were selected: Decision method for fuzzy logic operators AND: MIN Decision method for fuzzy logic operators OR: MAX Implication: MIN Agregation: MAX Deification: CENTROID (centre of gravity) The same methods were used in all other fuzzy controllers.
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Fig. 8 shows the membership functions for Fuzzy PI part for error, error derivative and for output respectively. NI-G
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3.5. Fuzzy logic controller for cooling Cooling is in simulator KAMRA modelled as a negative heat flux from the ventilator. The structure of fuzzy logic controller for cooling is very similar to the heating one. It also consists of a parallel structure of PD and PI part. The membership functions are mirror images of membership functions for heating.
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The described fuzzy controllers were validated with several simulation studies. Heating was tested with the following parameters: ~ o| I I I I I I I I I winter (January 31) 0.05 Season: 0.03 0.02 0.01 0 0.01 0.02 0.03 0.04 0.05 0.04 Error derivative Outdoor temperature: 0~ i 1~ ~ J ~ i~ I I Temperature of the terrain: 0~ Global solar radiation: 400 W/m2 ~ 0.5 Parameter for direct radiation: 1 I Parameter for cloudiness: 0 0 I I I I I i I 3 2.5 2 1.5 1 0.5 0 0.5 Ventilator power: 0W Output Roller blind: shut Fig. 8. Membership functions for PI controller. Fig. 9 shows the transient response of the indoor temperature when the reference temperature is changed from 18 ~ to 20 ~ The mnemonics in conjunction with membership functions have the following meaning: 2O NEG negative error SN small negative error ZE zero error SP small positive error POZ positive error IN increasing of indoor temperature Z no change in error 18i DE decreasing indoor temperature 2.5 2.55 2.6 2.65 2.7 D decreasing of output Time [s] xlO~ SD slow decreasing of output NA no action Fig. 9. Indoor temperature as the response to the reference SI slow increasing of output change. I increasing of output The relations between inputs and output are determined with It can be noticed, that the control system has appropriate the following fuzzy rules: performance. The response to the step change is fast, the 1. If (E is Z) and (ED is DE) then (I-IPI is NA) overshoot is small (about 0.04 ~ The steady state error 2. If(E is Z) and (ED is Z) then 0-IPI is NA) vanishes in reasonable time. This confinm the integral action of 3. If (E is Z) and (ED is IN) then 0-1PI is NA) the controller. 4. If (E is SP) and (ED is DE) then 0-IPI is SI) Cooling was tested with the following parameters: 5. If (E is SP) and (ED is Z) then (HPI is SI) Season: summer (July 15) 6. If (E is SP) and (ED is IN) then 0-IPI is NA) Outdoor temperature: 30 ~ 7. If (E is POZ) and (ED is DE) then 0-IPI is I) Temperature of the terrain: 20 ~ 8. If (E is POZ) and (ED is Z) then (HPI is I) Global solar radiation: 500 W/m 2 9. If (E is POZ) and (ED is IN) then (I-IPI is SI) Parameter for direct radiation: 1 10. If (E is SN) and (ED is DE) then 0-IPI is NA) Parameter for cloudiness: 0 11. If (E is SN) and (ED is Z) then 0-IPI is SD) Heating power: 0W 12. If (E is SN) and (ED is IN) then 0-IPI is SD) Roller blind: shut 13. If(E is NEG) and (ED is DE) then 0-IPI is SD) Fig. 10 depicts the indoor temperature and the cooling power 14. If (E is NEG) and (ED is Z) then 0-IPI is D) of the ventilator when reference temperature changes from 18 15. If (E is NEG) and (ED is IN) then (HPI is D) ~ to 20 ~ and back to 18 ~ E means error, ED means error derivative and I-IPI means heating (contribution of PI part). ~o.sF
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Fig. 10. Indoor temperature and cooling power as responses to reference temperature change. Cooling power is signed by negative values. At the first reference change the ventilator completely stops for a short time. In this period the indoor temperature increases due to the high outdoor temperature. It can be noticed that lower reference temperature needs higher cooling power. The control system again assures zero steady state error. For the sake of simplicity heating and cooling systems were tested independently. It is obvious however that the appropriate switching coordination is needed in order to prevent simultaneous heating and cooling actions and to assures bumpless transfers from cooling to heating phase and minimal energy costs. 3.7. Harmonisation of heating and cooling with fuzzy logic coordination It is very important to appropriately define the season for heating and the season for cooling. Unfortunately theirs starting and ending dates are not known in advance or fixed. The heating period is sometimes extended in May and the cooling period in October. The best additional information for economic heating and cooling is the difference between the reference and outdoor temperature. If this difference is positive, the cooling should be suppressed and if it is negative, the heating should be suppressed. So the fuzzy logic coordination seemed to be a right solution for the mentioned problem. It is implemented as a feedforward control as the action of the controller does not depend on the indoor temperature. The first input is the day index. January 1 means day index 0 and December 31 day index 365. The one year range was split to three periods: Jan. 1 -May 15, May 15Oct. 1, Oct. 1 - Dec. 31. All three ranges are covered with three trapezoidal membership functions with very small overlapping periods. The second input is the difference between the reference and outdoor temperature. It is fuzzyfied with three symmetrical membership functions in the range -30 ~ to 40~ The output variable is described with three symmetrical membership functions (active heating, inactive heating and cooling, active cooling). The fuzzy inputs and outputs are linked with nine fuzzy rules.
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functions for different roller blind shading areas. The input and output signals are linked with sixteen fuzzy rules. Ten rules are used in conjunction with cooling season. Here it is possible to achieve considerable energy savings as roller blind suppresses the warming due to moderate and intensive global or direct solar radiation. Beside energy consumption criterion the daylight illumination is also taken into account. During heating season however the priority was given to living conditions in comparison with energy consumption. In the case of clouded sky the roller blind remains opened. In sunny weather however the fuzzy logic shades the window at the most for 50%.
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3.10. Simulation tests for roller blind Different simulations were performed for cooling and heating periods. In the example presented in Figs. 12-14 the influence of global radiation disturbance to indoor temperature in the heating period was studied. The following experiment parameters were used: Season: summer (July 15) Outdoor temperature: 25 ~ Temperature of the terrain: 20 ~ Global solar radiation: 200 W/m 2 Parameter for direct radiation: 1 Parameter for cloudiness: 0 During observation the reference temperature is changed from 18 ~ to 20 ~ and back to 18. ~ In the moment t=-50000s the global solar radiation is changed from 200 W/m 2 to 500 W/m 2. Fig.12 depicts the reference and indoor temperature. Fig. 13 shows the global solar radiation disturbance and the appropriate reaction of roller blind. At the beginning the roller blind is already lowered to app. 0.6 due to direct solar radiation. At the moment when the disturbance occurs the reaction of the roller blind is instantaneous as it is controlled with feed forward approach. Both diagrams i n Fig. 14 illustrates however the cooling power, which is needed for appropriate control. It can be noticed, that the cooling is completely switched off for a short period after the reference changes to 20 ~ At t=50000s the fuzzy controller in the cooling system intensifies the cooling power. All these actions cause that the radiation disturbance does not effect the indoor temperature (see Fig. 12).
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The advanced fuzzy control system, which was designed with simulation, was finally implemented with industrial programmable logical controller MITSUBISHI. The control schemes are defined in a special progranmaing environment IDR BLOK on PC computer. This is a graphical block oriented editor, so the programming is very similar as in Simulink. IDR block library contains many elementary arithmetic blocks (e.g. summer, gain .... ), as well as more control oriented blocks (e.g. analog and digital inputs and outputs, P ID controller, fuzzy controller . . . . ). This low level control functions were improved with supervision level which was realized by Factory link. As
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this application is linked to the Microsoft Access database a very complex and flexible experimenting environment was obtained. The described implementation will be systematically tested in near future. 5. CONCLUSIONS
Krainer A., Kladnik R. and Perdan R. (1997). Light and Thermal Energy Coordination in Building, PLEA 1997 Book of Proceedings, Kushiro, Japan. Matko D., Karba R. and Zupancic B. (1992). Simulation and Modelling of Continuous Systems: A Case Study Approach, Prentice Hall Int., New York.
The life cycle of complex control systems consists of many complicated phases. In this paper some of them, which demand high level of knowledge about modelling, simulation and control system design approaches as well as about modem software and hardware technology, were discussed and implemented. The development of mathematical modelling is probably the most crucial part. As the validation of the model was successfully accomplished, it was later used for the design of control system based on fuzzy logic. The proposed control scheme which consists of two feedback and two feedforward controllers is only the first attempt with very promising results. A lot of work has to be done with more detailed evaluation of the proposed control system in the simulation and implementation environment. The experiments must accurately evaluate all influences to energy consumption and living conditions as well. There are many new possibilities for new modelling and control system design approaches. Experimental modelling or the combination of theoretical and experimental modelling (e.g. with neural nets) can result in more accurate model for control design purpose. There are of course many other possibilities for control system design. It is expected that the best results will be achieved with the combination of traditional methods (e.g. PID ) and methods that originate in artificial intelligence (e.g. fuzzy logic, neural nets, genetic algorithms, expert systems .... ).
~krjanc I., Kavgek-Biasizzo K. and Matko D. (1996a). Fuzzy Predictive Control Based on Fuzzy Model, Proceedings of 4~
European Congress on Intelligent Techniques and Soft Computing, EUFIT-96, Vol.3, pp. 1864-1869, Aachen. ~krjanc I., Kav~ek-Biasizzo K. and Matko D. (1996b). Fuzzy Predictive Control based on Relational Matrix Models, Computers chem. Engng, Vol.20, pp. $931-$936, Elsevier Science Ltd. ~krjanc I., Kav~ek-Biasizzo K. and Matko D (1997a). RealTime Fuzzy Adaptive Control, Engineering, Applications of Artificial Intelligence, Vol.10, No.l, pp.53-61, Elsevier Science Ltd. ~krjanc I. and Matko D. (1997b). Fuzzy Adaptive Control versus Model-reference Adaptive Control of Mutable Processes, Methods and Applications of Intelligent Control, Edited by Spyros G. Tzafestas, Kluwer Academic Publisher, Dordrecht.
Zupan6i6 B. and Klop6i6 M. (1995). Environment for the Simulation and Design of Control Systems, Proceedings of the Session "Software Tools and Products", Eurosim Congress '95, Ed. F. Breitenecker, I. Husinsky, Vienna, pp. 147-150.
REFERENCES
Autar R. and Bakker L. G. (1998). Smart integrated control of lighting and solar shading for oJ~ces, Solar control, EU-Joule III Project, Contract JOR3CT960113, TNO Building and construction research, Delft.
Zupan6i6 B., Krainer A. and ~krjanc I. (1998a). Modelling, Simulation and Temperature Control Design of a Test "Chamber", Proceedings of 1998 Summer Computer Simulation Conference, Ed. M. S. Obaidad, F. Davoli, D. DeMartinis, Reno, USA, pp. 173-178.
Babu~ka P,. and Verbruggen H. B. (1996). An Overview of Fuzzy Modelling for Control, Control Eng.Practice, Vol.4, No.ll, pp. 1593-1606.
Zupan6i6 B. (1998b). Extension software for real-time control system design and implementation with MATLAB- Simulink, Simulation Practice And Theory (6)8, pp. 703-719.
Duffle J. A.and Beckmann W. A. (1991). Solar Engineering of Thermal Processes- 2ndedition, John Willey&Sons, Inc.
ACKNOWLEDGMENTS
SOLAR CONTROL-integrated approach to solar techniques, Furlan B., Krainer A. and Perdan R. (1998). Measurements of thermal Response of Test Object with variable Geometry of Openings, Comparison to Computer Simulation, EuroSun98 Book of Proceedings Vol. 1, Portoro~, Slovenia.
CE DGXII JOULE-THERMIE, JOR3CT960113 (1996-1999)
Intelligent Control System for efficient use of Energy and indoor climate Conditions in the Building, Ministry of Science ofrep. Slovenia, L2-7695
Kladnik R. (1987). Theory of KAMRA, publication, Faculty of Civil Engineering, University ofLjubljana, Slovenia. Krainer A. (1994). Toward Smart Buildings, TEMPUS Joint European Project JEP 1802, Building Science and Environment-Conscious Design, Module 1: Design Principles, London.
The Interaction between dynamical Openings and Building Envelope, Ministry of Science ofrep. Slovenia, J2-9080
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XX. Solar, Thermal and Photovoltaic Concentrating Collectors
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DESIGN AND CONSTRUCTION OF A LINE - FOCUS PARABOLIC TROUGH SOLAR CONCENTRATOR FOR ELECTRICITY GENERATION
Bakos G.C., Adamopoulos D. and Tsagas N.F. Democritus University of Thrace, Department of Electrical Engineering and Electronics, Laboratory of Energy Economy, 1 Vas. Sofias Str.- Tel. 30 054179725,Fax. 30 054179734, 67100 Xanthi - GREECE, E-mail: [email protected] [email protected]
Soursos M. SENERS Energy Systems, 16 Kleovoulou Str., 11744 Athens - GREECE
Abstract - The design of a parabolic trough concentrator (PTC) used for electricity generation is presented in this paper. PTCs are the preferred type of collectors used for steam generation due to their ability to work at high temperatures with a good efficiency. Parabolic trough collectors are frequently used for solar - thermal applications because temperatures of about 300 oc can be obtained without any serious degradation in the collector efficiency. An important issue about solar concentrators of this kind is the thermal behavior analysis. Moreover, the conversion efficiency is determined. The different parts of the parabolic trough concentrator such as the metal frame, the parabolic mirrors, the solar energy absorption system are described. The sampling and control unit is tested on a similar small - scale PTC model which is developed and installed inside the Energy Economy Laboratory. A simulation program is developed which combines the level of solar irradiance, meteorological data, orientation and dimensions of the parabolic mirrors aiming to a better evaluation of the system efficiency. The advancement in this project is the use of parallel processing (real time processing) using OCCAM language for the system control and analysis of various parameters. Suggestions for system performance improvement are given concerning the choice of heat transfer fluid. The possibilities about combined operation of such a solar system using an MHD generator rather than conventional generators are examined.
I. INTRODUCTION The design of a parabolic trough concentrator (PTC) used for electricity generation is presented in this paper. PTCs are the preferred type of collectors used for steam generation due to their ability to work at high temperatures with a good efficiency. Parabolic trough collectors are frequently used for solar - thermal applications because temperatures of about 300~ can be obtained without any serious degradation in the collector efficiency. A typical application of this type is the Southern California power plants known as Solar Electric Generating Systems (SEGS). About 2500000 m 2 collectors have supplied more than 4000 GWh of electricity into the Californian grid [Kearny and Price, 1992, Jesch, 1998]. Current commercial PTC thermal electricity generation uses collectors which comprise an evacuated - annulus receiver consisting of an inner stainless steel tube mounted in a concentric evacuated cylindrical glass envelope which serves to minimize convective and conductive losses [ Norton, 1992, Kalogirou, Lloyd and Ward, 1997 ]. An important issue about solar concentrators of this kind is the thermal behavior analysis. Moreover, the conversion efficiency should be determined. The trough concentrator was designed following appropriate strength and pressure criteria. The different parts of the
parabolic trough concentrator such as the metal fimne, the parabolic mirrors, the solar energy absorption system are described. Furthermore, the sampling and control unit is tested on a similar small - scale PTC model developed and installed inside the Energy Economy Laboratory. The peripheral electronics, including the interface to PC and the sampling software, were developed and tested. Theoretical results, produced by an appropriate simulation program, are compared with experimental results taken from the experimental facility which is installed outdoors and is available at a specially formed area nearby Democritus University of Thrace. The system performance, under the specific environmental conditions and ground configuration noticed in the area of Xanthi in Northern Greece, was tested. A new simulation program is developed which combines the level of solar irradiance, meteorological data, orientation and dimensions of the parabolic mirrors aiming to a better evaluation of the system efficiency. Another important issue is related to the thermal loss of these facilities. The reduced performance of the experimental facility in comparison to the theoretical system is due to the existence of irregularities in the experimental solar receiving system and the quality of the heat transfer fluid which, in the case of experimental application, is tap water. Suggestions for system performance improvement are given. For instance, a good choice of heat transfer fluid should be
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made. The transfer fluid could be steam, to be used directly in a steam turbine, or a proper oil. Although it seems to be an expensive solution, the use of a thermochemical storage medium is under discussion. Another important issue related to these systems is their cost especially when they are intended to be used in a commercial scale. The overall cost of the PTC system described in this paper is estimated at approximately $10000 and is a unique application for Greece. It is designed to be used only for research purposes. The research work in our Laboratory is also focused on the possibilities of a combined operation of such a solar system with a non - conventional electricity generation system such as MHD (magnetohydrodynamic) generator [Mankins, 1996]. Some commercial companies and local authorities showed great interest to sponsor this effort since it is the first solar application of this type in Greece. 2. DESCRIPTION AND FUNDAMENTALS OF THE PTC SOLAR SYSTEM The experimental facility is installed at a specially formed area nearby Democritus University of Thrace. The sampling and control unit is tested on a similar small - scale PTC model developed and installed at Energy Economy Laboratory. Concentrating collectors of that kind present operational difficulties compared to fiat - plate collectors [Duffle and Beckman, 1991]. Except at the very low end of the concentration ratio scale, they have to be oriented to 'track' the sun so that beam radiation will be directed onto the absorbing surface. There is also need for maintenance in order to retain the optical system operation in a satisfactory level for long temporal periods in the presence of dust, nasty meteorological conditions, oxidizing or other corrosive atmospheric components. The total facility of the PTC consists of the following parts: a) metal support frame of the PTC b) parabolic mirrors c) solar radiation absorbing system (pipes) d) solar orbit "tracking" system e) sampling system- control panel 2.1 Metal support frame of the PTC A schematic diagram Of the metal frame is given in Fig. 1. It consists of two fixed base parts and three movable parts. The first movable part is a turret which is resided to the low base by a couple ofjoints. This gives the possibility for the turret to move (spherical coordinates - v movement). There is a small carriage arranged onto the turret which slips from the top to the bottom point of the turret. One of the two edges of the parabolic mirrors rotation axis is mounted onto this carriage. The other edge is supported from an horizontal barrier which is arranged through a couple of joints to the large base of the syster~ The combination of the turret - carriage system is used as a "tracking" system of the seasonal solar orbit. When the small carriage moves onto the turret, the last one starts to incline from the vertical position leading to the inclination of the parabolic mirror rotation axis from the horizontal position. This happens in order to achieve a perpendicular direction, for the rotation axis, to the incident solar in~adiance. This is very important in order to achieve maximum efficiency of the system.
Fig 1. A schematic diagram of the metal frame. The various movable and immovableparts are depicted. It is noticed that the seasonal level of solar orbit is different regarding to the horizontal level of the site. The turret height is the appropriate one, so when the small carriage reaches the top point, the achieved inclination of the turret at that moment, brings the rotation axis of the parabolic mirrors to the appropriate slope from the horizontal level (ground). This leads to the conservation of the verticality for the incident solar radiation. This boundary case appears for the lowest observed solar orbit. There is a parabolic metal lattice supported to the rotation axis of the parabolic mirrors. The metal lattice's size is corresponding to the total area of the parabolic collector surface. It is used for the rotation of the parabolic mirrors (daily tracking of the sun). The metal lattice is constructed covering the appropriate mechanical criteria of strength. When the rotation axis of the mirrors is horizontal, then a rotation within 180 degrees is achievable. Tracking of the daily solar orbit is achieved by rotation of the horizontal metal lattice around the PTC axis independently from the vertical movement of the system used for the seasonal solar orbit tracking. A natural view of the facility is given in Fig. 2. 2.2 Parabolic mirrors The total collector surface consists of 4 parabolic mirrors of 3m2 (1640 m m x 1700 ram) each. The incident solar radiation is focused to a light line (focus point) which is parallel to the rotation axis, at a distance of 1700 ram. Mirror thickness is 4 mm, which is small compared to the total magnitude of its surface. These mirrors where imported from the German company Flachglas Flagsol GMBH. They have high reflectance (98 %) and their efficiency was tested at companies' laboratory.
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and the mirrors can be modified for a few centimeters in such a way that the tubes are positioned to the focus line. This kind of regulation takes place only once when the parabolic mirrors are installed and is not repeated during the measuring procedure. The order of magnitude for the absorbing pipes cross section was chosen according to the broadness of the area which the collected solar radiation occupies on the focus line (magnitude of sun's image).
Fig. 2. A natural view of the metal frame in the facility of the PTC systc~n. The points through which the parabolic mirrors are mounted onto the metal frame allow micro - scale adjustments to achieve an exact focusing of the incident solar radiation. 2.3 Solar radiation absorption system Solar radiation absorption system consists of metal pipes (tubes) which are placed parallel to the rotation axis. Their placement coincides with the focus point. The pipe cross section is 4 cm and the wall thickness is small so that quick transfer of energy from solar radiation to the working fluid is achieved. The outer pipe surface is covered with a special black paint (selective surface) which increases the absorbance of the incident solar irradiance and reduces, simultaneously the reflectance. The most important issue about the absorption system of that pipes, is that they are covered externally with glass tubes of equal length, concentrically placed. These tubes have a 9 cm cross - section. The area between the glass tubes and the absorbing pipes is hermetically closed. The internal pressure is less than the atmospheric pressure. The usefulness of this section is significant since it reduces the convection heat losses. The external glass tube reduces the heat losses due to radiation, because the produced heat from the metal surface of the absorption pipe is "trapped". At the edge of the absorbing pipes, a flexible tube is used for the conveyance of the heat transfer fluid. A proper oil is planned to be used for the large scale facility, but for the experimental operation of the small scale facility, tap water was used. For the circulation of the fluid, a regulating flux pump is used ( provided with a fluxmeter). The absorption pipes are supported to the focusing line by a number of trapezoidal metal light - weighted barriers, whose axis of symmetry is perpendicular to the rotation axis. At the support points of the tubes, there are joints which allow two axis regulation. This is very helpful in order to improve the focus quality. In other words, the distance between the tubes
2.4 The solar orbit tracking system The support system of the parabolic mirrors is constructed in order to track the sun. The orientation of the parabolic trough collector need to be the appropriate one, so that the incident solar beam will fall always perpendicularly. The orientation is achieved by two independent movements of the whole system, as it was referred to the previous paragraph. The first movement is for "tracking" the solar daily orbit. It is achieved by the use of a DC motor of 1,5 KW. The power of the motor is transferred to the rotation of the PTC axis through a gear system with a significant reduction ratio (1: 6000). This is necessary to accurately orient the optical system, since the mirrors should continuously change their position according to the slow movement of the sun in real time. The second movement is also achieved by a motor of the same nominal power. This movement corresponds to the seasonal solar orbit "tracking". The philosophy of the second movement of the PTC system was given in the previous paragraph. The motor movement is achieved trough the Pulse Width Modulation ( PWM ) technique. 2.5 Sampling system - Control panel The sampling system monitors three parameters: temperature, solar radiation density and heat - transfer fluid flux. The light intensity measurement is also used for the positioning of the optical system [ Bakos, 1991 ]. There were developed appropriate programs and interface systems using sensors and amplifiers for the experimental results. The position checking is being done using an absolute rotary encoder. The control panel provides the possibility to move the system manually or automatically through the computer. Normally, the system operation is taking place through the computer. There is continuous evaluation of the pre - referred parameters together with continuous storing of data through the interface system to the computer. 3. SYSTEM M O D E L L I N G 3.1 Thermal analysis In this part of the report, the thermal analysis of the PTC system is described. Also a description of the developed simulation program used for the calculation of the efficiency and the fluid outlet temperature as a function of solar radiation intensity and characteristic parameters of the PTC is given. The energy absorbed by the heat transfer fluid per unit area is given as follows: q = h(Tp - Tf) where h = heat transfer coefficient
( 1)
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dT
Tp =pipe temperature Tf --fluidtemperature The useful energy absorbed by the fluid can be calculated from the difference between the incoming solar radiation and the total heat loss:
q=
Ief - ( F t + Fp )-(Tp - T a )
Fp = passive area thermal loss factor T a = ambient temperature We notice that the actual absorbed solar radiation Ief is evaluated from the incident solar radiation Io using the appropriate equations: Ief = I o 9cosCO where V&is the hour angle evaluated for the current relative position of the sun to the earth :
h - [ I e f - ( T f - T a)-(F t + Fp)]
h+Vt +Vp
4
d . u . Cp ~ - ~ q dy def
(4)
d = fluid density u = fluid speed Cp = specific heat
where D is the diameter of the absorber pipe. The conversion efficiency n is given from the following equation: Q
n = ~
3.20n'ginal system simulation According to the thermal analysis described in the previous section, an appropriate simulation program was developed. It is a conventional sequential (serial) computer program The input data to the simulation program is shown in Table 1. The results were taken for a line - focus PTC (small - scale facility) of an active pipe length of 1.7 m. There were many graphs produced and conclusions derived concerning the PTC operation. In this paper some of them are presented. These results were taken for an indicative summer day, with flux rate of 50 g/s. Table I
1
h
collector's width W absorber's pipe diameter D1
[Ief - ( T e f - T a ) ( F t + Fp)]
def ducp h+F t+Fp
(5) If we place: -
-
-
4 ~
def
---
+Fp
Parameter
d ef = collector's active diameter From equations ( 3 ) and ( 4 ):
C
h+Ft
(7)
S c = collector's active surface (m 2)
where
~
Q =V(DL. h .[Ief - ( T f - T a)-(F t + Fp)]
(3)
The temperature increase d Tf of the heat transfer fluid in a
B
(6)
I o = incoming solar radiation ( W / m 2)
length dy of the absorber pipe when energy q is absorbed, is given as follows:
A
e_BL
where
where t is the solar hour. From equations ( 1 ) and ( 2 ) we take :
dy
C
I o "S c
co =15~ (t-12)
4
C
To =-~+1 Ti - ~ l .
absorber pipe length L. The values of B and C can be calculated from the pipe and heat - transfer fluid characteristics. The useful energy Q absorbed from the pipe is given from the following equation:
F t = active area thermal loss factor
dT
and after numerical integration:
Equation ( 6 ) gives the outlet temperature To of the heat transfer fluid as a function of the inlet temperature T i and the
I ef = absorbed solar radiation
dTf
=C-B-Tf
(2)
where
q=
dy
~
1 ducp
A-h (Ft 9 +Fp) h+Ft +Fp A-h
h+Ft +Fp
[Ief 9 + T a .(F t + Fp)]
then equation ( 5 ) can be written as follows:
Value 1.64m 0.04m
shield's diameter D2 transmission coefficient 12 absorption coefficient 1. reflectance r active area thermal loss factor F t
optional (0.3 m)
passive area thermal loss factor Fp
0.2 W/m 2 K ~
ambient temperature pipe length L flux rate u solar radiation I o
Ta
0.85 0.92 0.9 2 W / m 2K~
290 K 1.7m 50 g/s selection
In the simulation program there is an option to put a proper parabolic piece of metal upwards the absorber pipe which is often called shield. The shield is aimed to reduce the heat
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losses from the absorber. The absorber loses radiation only in directions unprotected by the shield [Twidell and Weir, 1990]. The PTC facility is not provided with such a shield, but there is a possibility to evaluate the concentration ratio of the collector with or without shield and find out the difference from derived output of the simulation program. Since the shield use increases the concentration ratio, the PTC energy conversion efficiency is also increased. The mathematical formulas for the evaluation of the concentration ratio for both of these cases are: W-D
With a protecting shield: C r =
2
3.3 Current system simulation The advancement in this project is the use of parallel processing (real time processing) using OCCAM language for the system control and analysis of various parameters. The simulation software is being written also in OCCAM language, running on a T800 transputer of the 1NMOS company. 1NMOS Transputer is a high performance single chip computer whose design facilitates the construction of parallel processing systems. The Transputer executes OCCAM programs more or less directly [Pountain D., 1987].
(8)
BEGIN
vC]DlZ~ 3600
i
where
INPUT Ti Io I
C r 9concentration ratio D 1" absorber's pipe diameter D 2" shield's diameter W" collector's width z angle representing the uncovered absorber part
:
W Without shield" C r = ~ (9) V(lD1 The output of the simulation program is the concentration
I
INPUT PTC PARAMETERS DATA .J FOR EACH HOUR OF THE DAY FIND ACTUAL SOLAR RADIATION FOR COLLECTOR ORIENTATION
ratio C r , thermal conversion efficiency n, output temperature T o and produced thermal power P. In Fig. 3 there is a schematic diagram for a PTC provided with protecting shield. In the simulation program there is also derived an evaluation of the maximum attainable output temperature Tmax for ideal quality of the PTC operation. For instance, when the PTC is provided with a protecting shield, Tmax could be achieved when the shield allows radiation to move towards the mirror side. Of course, this temperature cannot be practically achieved because parabolic trough collectors are not perfectly parabolic and part of the useful heat is removed as long as the heat - transfer fluid pass through the absorber pipe. The flowchart of the original simulation program is shown in Fig .4.
Y
o
FIND CONCENTRATION RATIO (SHIELD UNPROTECTED) I
L~ FINDTo ~-J FIND Tmax FIND ENERGY ABSORBED FIND CONVERSION EFFICIENCY
~sTE
THE~ ULTS /
GO TO THE NEXT HOUR
Fig. 3. Scheme of a shield protected PTC where W is the PTC width and z is the angle representing the uncovered absorber part.
Fig. 4. The original simulation program flowchart. The angle mentioned into that flowchart, concerns the z angle for the uncovered part of the absorber when a protecting shield is applied to the PTC.
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The advantage of this approach is that the PC can control and analyze the system input parameters (such as collector positioning, weather parameters etc.) simultaneously leading to a better system efficiency evaluation. It can be used as an input data, information from measurement elements, concerning ambient temperature, sky clearness index, wind speed etc. Variations of these parameters would entrance continuously into the OCCAM language program attaining real time processing evaluation. The main idea is represented in Fig. 5.
line expresses the same factor for a PTC system without shield protection (concentration ratio value is 13).
Fig. 6: Output temperatureTo versus incident solar radiation Io. The high solar radiation gives an increased efficiency of the PTC system. It is also expected that a concentration ratio increase leads to the conversion efficiency increase. The difference is obvious for low values of Io. When we have high values of Io the efficiency difference for these two values of the PTC concentration ratio is getting smaller but is still significant (a 10% difference is not considered negligible). The conclusion is that using a well designed protection shield in a proper position, results an increased efficiency for the PTC system.
Fig. 5: The current simulation program flowchart. Input data is given continuously. 3.4 Simulation results In this paper, four graphs concerning the small scale PTC system operation are presented. In Figure 6, the output temperature To variation, as a function of incident solar radiation Io is shown. Continuous line represents the simulation result while the dotted line corresponds to the output temperature experimentally attained, for the small scale PTC system. As it was normally expected, the output temperature increases with the incident solar radiation. We also observe a linearity between these two magnitudes for high values of Io. For Io = 800 W/m2, the simulated output temperature is To = 440 ~ while the actual value is To = 416 ~ Given that the working fluid was tap water this result i s considered encouraging. In normal conditions, steam production is the output of the system for high values of Io (at the middle of the day or early alternoon). In Fig. 7 the thermal conversion efficiency variation n, as a function of the incident solar radiation is given. The dotted line shows the variation of the factor n, for a shield - protected PTC (concentration ratio value is 32) while the continuous
Fig. 7: Thermal conversion efficiency n versus incident solar radiation Io. In the following graph (Fig. 8) the absorbed thermal power P variation is shown as a function of the incident solar radiation Io. The dotted line represents the variation of the power P, for a shield- protected PTC (concentration ratio value is 32) while the continuous line expresses the same magnitude for a PTC system without shield protection (concentration ratio value is 13). The absorbed thermal power is higher when concentration ratio is increased, as it was expected, because the conversion efficiency factor is higher. Considered that the simulation program concerns a small - scale PTC facility using tap water, the derived level of the absorbed thermal power is encouraging.
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In Fig. 9 there is a comparison between simulation and experimental (actual) results for the PTC system operation. The thermal conversion efficiency n as a function of the incident solar radiation Io is shown.
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is still under experimental operation and improvements should be applied. 4.2 Magnetohydrodynamic and solar energy combinan'on Based on the existing work which has been done so far, there is a strong intention to form a PTC system used as a heat source for a magnetohydrodynamic (MHD) generator rather than a conventional generator. A schematic diagram of an MHD generator is given in Fig. 10. One of the most promising plasma applications is the magnetohydrodynamic (MHD) generators. At this kind of generators, there is an overheated gas flow with high velocity through a proper magnetic field B , leading to the production of an electromotive force E. The magnetic field is developed into a duct which is called MHD channel (Fig. 10). The
Fig. 8: Thermal power absorbed (P) versus incident solar radiation Io. The continuous line represents the numerical results committed by the simulation program. The dotted line represents the experimental results of the small - scale PTC system for the corresponding solar radiance values Io. It is expected that the actual performance of the PTC is reduced compared to the simulation procedure.
intensity B of the magnetic field should be significantly high. The magnetic field source is a superconducting magnet or field coils which is preferred for experimental applications (Fig. 10). During the MHD power production procedure, conversion of thermal energy into electricity is taking place. The applied magnetic field direction is perpendicular to the direction of the plasma flow. As a result, we have reduction of the total hot - plasma energy during its movement into the MHD channel. This energy reduction leads to the electricity production. The derived amount of electric energy can be driven to a proper load. In figure 10, the incoming plasma to the MHD channel has velocity u directed along the X axis, the magnetic field direction is along the Z axis while the direction of the produced electric field is along the Y axis (see figure 10). The MHD duct (channel) is rectangular shaped. The electrodes used for the electric field development produced by the plasma movement are the MHD channel walls. The operation principle of the MHD generator is shown in Fig. 10. The incoming hot - plasma is partially or fully ionized so the main idea is to drive the corresponding charges to the electrodes. That is achieved because of the properly directed applied magnetic field. If the electrodes are not attached to an output load there is no closed electric circuit neither electric current. For the case of the closed circuit, the general form of Ohm's law gives:
where J Fig. 9: Thermal conversion efficiency n versus incident solar radiation Io. Simulatedand experimentalcomparison. 4. ELECTRICITY PRODUCTION 4.1 General aspect Having a maximum output temperature To= 440 ~ is quite difficult to produce superheated steam directly and drive it to a conventional steam turbine for electricity production. The alternative solution is to use the produced thermal power into a preheat cycle aiming to the reduction of the fuel quantity needed for steam superheating. It is our intention to extend the PTC facility by purchasing additional mirror and absorber components. The whole system
is the current density, 11 the specific plasma
conductivity, E and B the electric and magnetic field intensity which influence the moving charge carriers having velocity u . It is easily derived that the developed electrode voltage is: AV=u.B-z Plasma velocity ( u ) is reduced into the MHD channel. Part of the plasma kinetic energy is transformed to electric energy. Half of the total power derived from the gas plasma flow is consumed for the internal plasma heating. The rest of it, is provided to the output load. Basic demands for the system operation are high plasma flow rate and high conductivity. It is also important to have a high plasma density in order to assure a sufficient level of input power. When the plasma ionization degree is high, the demand for high conductivity is fulfilled. Therefore, it is intentional to introduce a plasma seeding with low ionization potential, like Cesium.
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type of generator, at the inlet of MHD channel the nonequilibrium plasma clots are initiated by high power electron beam. After that, a supersonic gas flow drags this layers through the cross magnetic field of high intensity and produce electric energy. The numerical simulation has shown that enthalpy extraction ratio near 42% and isentropic efficiency 85% could be achieved [Slavin et al, 1996]. For the creation of a solar MHD generator on the basis of solar concentrator there is a fundamental scientific problem which must be solved: the development of heat source which consists of solar concentrator and heat accumulator for space conditions at working body temperature (probably noble gas) of about 2000 ~ 5. CONCLUSIONS
Fig. 10: Schematic diagram of a magnetohydrodynamic generator (MHD channel) and its operation principle. M H generators drive significant amount of power through a small resistor. That means high current intensity. Therefore, use of conversion devices is necessary to control that power. MHD power production is an open research field. According to many researches which have been taken place worldwide, it is estimated that MHD generators could give a reliable solution to the energy problem which will appear during the 21 st century. Currently, there is cooperation on this issue with foreign institutes. MHD electricity production using solar power seems very promising especially for space applications where the incident solar radiation is significant compared to the incoming irradiance which reaches the earth. The project of space power plant using solar radiation as a source of thermal energy is very attractive. However, there is a problem of cooling at the working cycle of a heat engine in space conditions. Removal of heat is possible only through thermal radiation. Emission of thermal energy from the surfaces of radiators to outer space, is effective only at the temperature about 600 ~ for power higher than 1 MW, that determines the temperature of heater in about 2000 ~ at efficiency level near 30%. Unique opportunities occur by using an M I - I generator combined to a concentrator of solar energy aboard of a space power plant, because MHD generator can effectively operate working body temperature of about 2000 ~ Until now, the effective closed cycle MHD generator has not been created. Experiments conducted in Tokyo Institute of Technology on the MHD facility of a Hall type, the best result have been obtained: the enthalpy extraction ratio has reached 38%. This result satisfies the requirements for perspective industrial MHD power plants. However in Hall generator it is difficult to obtain the satisfactory results on the second major parameter which is isentropic efficiency [Okamura, 1994]. This should not be less than 70%. It is easy to fulfill this condition in Faraday- type MHD generator. The new idea for realization of Faraday type closed MHD generator using stratified recombinated plasma flows of the noble gas has been suggested [Slavin et al, 1996]. In a given
In this paper, there was a description of a parabolic trough concentrator (PTC) design and construction. There was also a comparison between simulation and experimental results for a small- scale model of the PTC system. During the experimental procedure, the PTC was moving in two directions. In one direction the axis of PTC was vertical to the incident solar radiation following the seasonal solar orbit variation and in the other direction, it follows the solar daily orbit variation. These movements are almost impossible to be realized simultaneously in a large scale PTC solar system. In this case, the absence of one of these movements costs an additional heat loss part of incident solar energy. Similar behavior is observed between numerical and experimental results. The difference noticed is very logical and depends on technical reasons. The absorbing pipes used in the experimental measurements were provided only with glass covers without vacuum and their quality was poor. As it was mentioned earlier, the heat - transfer fluid used in the experiment was tap water with low anti-corrosion protection. For the numerical calculations high quality vacuum pipes (Farnell - Philips) and better quality heat- transfer fluid than water was considered. The possible uses of a similar larger system could be hot water production for industrial use and electric energy production for combined use with natural gas systems. There is a strong intention to form a PTC system used as a heat source for a magnetohydrodynamic (MHD) generator rather than a conventional generator. This research field seems very interesting especially for space applications. The described PTC system can reach a conversion efficiency level of 35 - 45 %. This could probably increased even more by applying proper improvement techniques. In the experimental procedure there was also a serious disadvantage ; the reduced dimensions of the PTC system introduces difficulties towards the precise behavior estimation of large scale facilities. REFERENCES
Bakos G.C. (1991). Three dimensional (3D) acoustic and vision systems. Ph.D. Thesis, University of Liverpool, UK. Duffie J.A. and Beckman W.A. (1991). Solar Engineering of thermal processes. 2~dEd., J. Wiley & sons, New York.
ISES Solar World Congress 1999, Volume III
Jesch L. Solar thermal power . Renewable Energy World, 1998, 1, p. 52-53. Kalogirou S., Lloyd S. and Ward J. (1997). Modeling, optimization and performance evaluation of a parabolic trough solar collector steam generation system. Solar energy, Vol.60, No 1, p. 49 - 59. Keamey D.W. and Price H.W. (1992) Solar thermal plantsLUZ concept. In Proceedings of 2*d Renewable Energy Congress, Reading UK, Vol.2 ,p. 582-589. Mankins J.C. (1996). A flesh look at space solar power. In Proceedings of 31st Intersociety Energy Conversion Engineering Conference (IECEC - 96), Washington D.C., v.3, p. 451 - 457. Norton B. Solar energy thermal technology (1992). Springer Verlag, Heidelberg, Germany. Okamura T. et al (1994). Review and new results of high enthalpy extraction experiments at Tokyo Institute of Technology. In Proceedings of SEAM - 32, Pittsburgh, USA. Pountain D. (1987). A tutorial introduction to OCCAM programming. 1NMOS Company, UK. Slavin V.S., Danilov V.V. and Sokolov V.S. (1996). Closed cycle MHD generator with non - uniform gas plasma flow driving recombined plasma clots. In Proceedings of 31 ~t Intersociety Energy Conversion Engineering Conference (IECEC - 96), v.2, p. 836-841, Washington D.C., USA. Slavin V.S., Lobasova M.S., Finnikov K.A., Danilov V.V. and Sokolov V.S. (1996). Numerical simulation of MHD process in the planned experimental facility with non uniform gas plasma flow driving recombined plasma clots. In Proceedings of 12ta International Conference on MHD electric power generation, Yokohama, Japan. Twidell J.W. and Weir A.D. (1990). Renewable energy resources. 2=dEd., University Press, Cambridge, p. 130- 133.
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THE DUCT SELECTIVE VOLUMETRIC RECEIVER: POTENTIAL FOR DIFFERENT SELECTIVITY STRATEGIES AND STABILITY ISSUES Xavier G. Casals Departamento de Fluidos y Calor, Universidad Pontificia Comillas-ICAI, C/Alberto Aguilera 23, Madrid, 28015, Spain, Tel./915422800 ext 2366, e-mail: [email protected] Jose Ignacio Ajona Departamento de Energia Solar, Viessmann, c/Volta 4, Poligono Industrial San Marcos, Getafe, 28906, Spain, Tel./91-6820911, email: [email protected] Abstract - Recently much theoretical and experimental work has been conducted on volumetric receivers, but not much attention has been paid to the possibilities of different selectivity mechanisms to be used in order to minimize radiation thermal losses which are the main ones at high operating temperature. In this paper we present a duct volumetric receiver model and its results, which allow the evaluation of different selectivity strategies such as: conventional e/a, geometry, frontal absorption and diffuse/specular reflection. In recent work on volumetric receivers based on simplified models, it has been concluded that the duct volumetric receiver is inherently unstable when working with high solar flux. We didn't find any unstable receiver behaviour even at very high solar fluxes, and conclude that a substantial potential for efficiency improvement exists if selectivity mechanisms are properly combined.
1. INTRODUCTION In order to achieve high thermodynamic conversion efficiencies, modem combined power cycles or advanced gas turbine cycles demand high quality thermal energy input, with maximum working fluid temperatures in the range of 1700 K. When trying to solarize or hybridize such a power cycle, high temperature solar thermal collector-receiver technologies are required if no limitations are to be imposed on the maximum conversion efficiency or solar fraction. Therefore, we need a solar thermal concentrating technology able to work at high temperatures with a high receiver thermal efficiency. Amongst the different technologies available for electricity generation with solar thermal power plants, the central receiver technology is one that seems to have a high technical potential for power generation in the 100-200 MWe range, because of introducing the least limitations on the temperature level at which solar energy can be introduced in modem thermodynamic power cycles. From the central receiver technologies available, volumetric receivers seem to be the most appropriate to reach these objectives. Different types of volumetric solar receivers have been tested since 1981. However, meama'ed receiver efficiencies and air outlet temperatures have not been as high as one would e ~ t from the volumetric concept. Early results from Odeillo (Menigault, Flamant and Rivoire, 1991) were ofhth = 66 % at Tso = 500 ~ for packed bed receivers in 1981 and lath= 73 % at Tgo = 600 ~ for honeycomb structure receivers in 1983, being unable to work at higher air outlet temperatures because of overheating at the front surface. This problem was overcome with fluidized bed receivers, but with lower effieieneies: lath= 47 % at Tso = 800 *C. From 1987 on, tests on different volumetric receivers have been done in the PSA (Almeria-Spain). With the Sulzer metal wire receiver (1987) one of the highest effieieneies was measured giving lath= 75 % at Tgo = 700 ~ (Becker, Cordes and Bfihemer, 1992). The ceramic foil receiver (B6hemer and Chaza, 1991) operated at Tgo = 782 ~ with lath= 59 % , and the ceramic foam
receiver at Tgo = 730 0(3 with lath = 54 % (Chavez and Chaza, 1991). Although these first kind volumetric receivers were not optimized, and operating solar irradiations were under 0,8 MW/m 2 (for tests in PSA), these efficiencies look rather low. Operation with modem gas turbines would demand higher air outlet temperatures, and therefore lower receiver efficiencies, in order to reach a significant solar fraction, but overall eonvertion efficiency would be very low. Higher gas outlet temperatures have been measured in the last years for both open and dosed volumetric receivers. In (Pitz-Paal, Morhenne, and Fiebig, 1991) results are given for several open volumetric receivers tested in the solar fumance of DLR (Cologne-~y) with non-homogeneous solar irradiation averaging and peacking up to 1,3 and 1,9 MW/m2. The highest air outlet experimental results where of lath = 82 % at Tso = 800 0(2 with the corrugated foil receiver, 1~ = 38 % at Tgo = 1100 ~ with the SiC honeycomb receiver, l~h = 7 1 % at Tso = 820 *(2 with the wire mesh receiver and hth = 58 % at Tso = 810 ~ for the ceramic foam receiver. In (Karni et al, 1996) the highest air outlet temperatures are reported for the DIAPR closed volumetric receiver with secondary concentration, attaining lath= 74 % at Tso = 1200 *C operating at 4 MW/m2 solar irradiation. To overcome the limitations of low operating tempertaures and efficiencies, one thinks of introducing selective strategies. In (Becker, Cordes and B6hemer, 1992) it was suggested that a high potential existed for receiver efficiency improvement by using selective coatings with a high a/e ratio. However, the first try to benefit from selectivity effects was proposed in (Flamant, Menigault and Olalde, 1987), with a different strategy: a two-slab closed selective volumetric receiver, based on disposing a semitransparent volumetric material (best results were obtained with a silica honeycomb) above a packed bed of absorbing particles. After theoretically and experimentally optimizing the receiver, they measured lath= 70 % at Tgo = 814 ~ operating with a 0,67 M W / m 2 almost uniform solar flux in a solar fumance (Variot, Menigault and Flamant, 1992). The same selective effect was proposed in (Pitz-Paal, Morhenne and Fiebig, 1991) for an
ISES Solar World Congress 1999, Volume III
open receiver, using the geometry of a duct volumetric receiver. The first experimental results from this receiver (Pitz-Paal and Fiebig, 1992) were ofhth = 60 % at Tgo = 750 ~ with 0,6 MW/m2 average solar irradiance. Recently much theoretical and numerical work has been conducted on volumetric receivers, but no attention has been paid to the possibilities of different selectivity mechanisms to be used in order to minimize radiation thermal losses, which are the main ones at this high operating temperature. The numerical model we have developed accounts for the main radiative, convective and conductive heat transfer fenomena, and allows for the analysis of different selectivity approaches, such as conventional a/e selectivity, geometry, frontal absorption and diffuse/specular reflection, allowing the axial variation of thermophysical and thermo-optical receiver properties to explore further selectivity mechanisms that could arise from it. We conclude that there exists a high potential for receiver efficiency improvement, mainly by introducing solar especular behaviour in the duct walls. Conventional a/e selectivity applied to the inner tube walls doesn't have a high effect on receiver efficiency, but it does have a high effect in the frontal surface exposed by the receiver. Minimising this frontal surface by sharpening the tube walls at the inlet is one of the most important issues to be able to benefit from the full potential of the other selectivity strategies. In recent work on volumetric receivers (Kribus, Ries and SpirE, 1995), (Pitz-Paal et al, 1997) it has been concluded that the duct volumetric receiver is inherently unstable when working with high solar flux (above aprox. 1 MW/m2). The models on which these conclusions have been drawn are ot~en based on simplified radiation exchange formulations and do not include all flow inertial effects, non developed flow entrance friction and convection phenomena, temperature dependence of therrnophysical properties other than viscosity and radiation selectivity mechanisms. Many of these assmnptions are not appropriate for the duct receiver, and can si~ificantly modify the temperature distribution along the duct, and thus its performance. In the model we developed we have tried to overcome most of these limitations, and when exploring stability issues, we didn't find any unstable behaviour even at very high solar fluxes ( 10 MW/m2).
2. THE DUCT RECEIVER MODEL The conjugate heat transfer problem in the receiver may be formulated by an integro-differential second order equation if convection effects are determined through Newton's law of cooling with a heat transfer convection coefficient based on experimental correlations. This equation, together with the proper integral energy and momentum balances in the air flow, permit computation of wall and air temperaturs along the duct. Spectral effects are considered by allowing different surface behaviour for the solar and infrared radiation exchanges. The infrared radiative terms are formulated with the enclosure theory (Siegel and Howell, 1992), on such a way that allows for convenient linealization in the iterative resolution procedure. Solar irradiation terms are formulated using conventional configuration factors (Siegel and Howell, 1992) when duct walls have a solar diffuse behaviour, and specular configuration factors (Lin and Sparrow, 1965), 0Labl, 1977) when duct walls have solar-specular
325
behavior. Dependence of all thermophysical and thermodynamical air properties (r, ca, , m , k) with temperature is retained in the formulation, because, due to the high air temperature increments when flowing through the receiver its properties can experiment variations up to 400 % for conductivity, 200 % for dynamic viscosity, 80% for density and 60 % for specific heat. Air flow will be laminar due to the small tube diameters considered (2-8 mm), and depending on the Reynolds and Rayleigh numbers, forced or mixed convection heat tansfer may be found (Shah and London, 1978). Rayleigh number will be higher at tube inlet, and secondary flows could develop due to bouyancy forces. Also, due to the small L/D ratios considered, significant portions of tube lenght will have a thermally and hydrodynamically developing flow with important effects on heat and momentum transfer. All of this is taken into account by using the appropriate convection and friction correlation to evaluate the Nusselt number and friction factor (Holman, 1997), (Shah and London, 1978). The integro-differential equation is linealized and iteratively numerically solved with a second order finite difference scheme. Appropriate relaxation factors are included to reach convergence of the linealized iterative scheme due to the non linealities introduced by radiative infrared terms, and convergence studies performed to reach proper wall and gas ten-q3emturedistributions. The results presented here are obtained with a one-chnanel model for an open volumetric receiver, and are therefore strictly valid for uniform irradiation at receiver inlet, which is approximated by the use of secondary concentrators. When solar irradiation at receiver aperture is not uniform, the fundamentals of the receiver behaviour are still in the one chanel model, but the results for parallel channels have to be combined in order to get a precise description of receiver performance as in (Spirkl, Ries and Kribus, 1997) and (Pitz-Paal et al, 1997). Nevertheless, as our focus is on finding potential improvements by introducing selectivity strategies, and all this information is included in the one channel model there is no point on combining parallel chanel results, which moreover depend a lot on solar flux distribution in receiver aperture, and therefore on concentrator geometry and configuration. To discuss the receiver performance and the effect of the different selecivity stategies, we find it convenient to introduce the equivalent emissivity defined by Eq. (1), which is the ratio of the radiative receiver losses to the losses that would exist to mantain the inner cavity at the end of the tube with its temperature without tube wall. Therefore this parameter permits to evaluate the selective capabilities of the duct receiver.
=
--eq
q rad,loss
(i)
a . jr~ . D2 //4 . (y4+2_ T1 )
3. DIFFUSE SOLAR ABSORPTION We'll begin by considering the duct solar receiver with a diffuse reflection of incident solar irradiation. This is the case of the volumetric receivers tested up till now and therefore will allow us to evaluate the ability of the model to predict receiver performances as well as the possibilities for performance improvement when introducing selectivity strategies in these
326
ISES Solar World Congress 1999, Volume III
receivers. When solar radiation undergoes diffuse reflection in the duct walls, in order to minimize reflection losses it is necessary to have a high solar absorption, and therefore most of the radiation is absorbed in the inlet tube region. This leads to a negative axial temperature gradient in almost all the tube with a decreasing absolute value as one moves towards the end of the tube. This is so as well with and without frontal absorption. As a consequence, temperatures in the inlet region and thus radiative losses are very high, and the selectivity potential of the duct geometry completely diluted. The selectivity strategies for these receivers involve appropriate choosing of emissivity and absorptivity in the tube surface as well as on the tube front, and thermal conductivity. High conductivities will improve receiver performance because of allowing a conductive transport of thermal energy in the front towards the tube end, reducing therefore wall temperatures in tube entrance (radiative losses), and allowing a higher air temperature outlet. In fact, wall conductivity is one of the factors that most affects receiver equivalent emissivity, but its effect saturates at values in the order of 50 W/m-K. High solar absorptivities will improve receiver performance because of reducing reflection losses, in spite of generating higher inlet tube temperatures and eoq. The reduction in reflection losses is so important with diffuse solar behaviour that the effect of high solar absorptivity is more important than the effect of low infrared emisivities. Low infrared emissivities will improve receiver performance because of reducing e, a. The tube inlet region is the one that has the higher axial temperature gradients, solar absorption, and view factor of the aperture, therefore the influence of a , e and k is mainly restricted to this region. When the tube leading edge is not sharp, frontal absorption becomes very important, increasing the wall inlet temperature and axial ~ t u r e gradient, and therefore radiation and reflection losses, with a very important reduction in receiver efficiency. In this situation, thenno-optieal properties of the tube edge become very important and dilute the effect of thereto-optical properties in the tube wall. High solar frontal absorption keeps on being very ~ r t a n t because of its reduction in reflexive losses in spite of its increase in radiative losses. Low frontal infrared emisivity is also very important because of reducing era. Tube wall a, e, and conductivity follow the same trends as before, but they are less important than ale and ele. In Table-1 we can find the effect of the different parameters on a diffuse tube without frontal absorption, while in Table-2 we present these results in the ease of frontal absorption. The base ease referred at these tables, as well as in Fig.l, is given by IJ i l l 5 ; D=2mm ; e--0,8 mm; R~=150 ; a---ale=0,8, e=ele=0,2 ; k=-50 W/inK. Solar irradiation is 4 MW/m 2 in the case without frontal absorption and 2,4 i W / m 2 with frontal absorption giving air outlet temperatures in the order of 1800 K in both cases.
T~ (K) 9
Base Case
i~
eel
0,79
1,04
!Dhth
De~l
|
1808
9
I
Modification (%) D T s o m
m
a = 0,95
1,73
mm
.9
m
a = 0,2
~
- 16,49
I
a = 0,2 - 0,8
0,00 -3,70 m
-3,70 m
-9,77 m
1,38 9
k =2-50 W/mK
0,00
-5,77
48,53
-5,77
48,53
- 15,11
166,62
2,16
-16,26
-.0,81
7,74
m
k = 100 W/mK mm
0,00
m
k = 2 W/mK mm
-10,44
!
e = 0,2 - 0,8 mm
-25,31 m
e = 0,8 m 9
0,90
I
,d
mm
2,73 mmM
m
-0,52
Table-l: Solar diffuse without frontal absorption.
In Fig.1 we present some wall and gas temperatme distributions for a diffuse receiver with frontal absorption and different design parameters in order to illustrate the above-mentioned comments. The receiver performance in the different cases from Fig. 1 may be found in Table-2. In Fig.2 & 3 we present the effects of conventional e/a selectivity on hth and e, a for a diffuse receiver with frontal absorption working at different air outlet temperatures (solar fluxes). As we may see, for increasing gas outlet temperature, the effect of e increases, and the one of a decreases. Receiver efficiencies on Fig2 are in the order of those obtained in volumetric receiver tests. Ts, (K) Base Case m
1841 m
Modific. (%) 99
hth 0,69 m
Tle (K)
1,32
2118
Deeq
DTle
9
) DTp m
%q
Dhth 9
9
I
i
ale - 0,2 mm
-17,94
9
ale = 0,95
3,85
ele = 0,8 9
mm
-4,19
m
9
0,73
-9,30
4,18
-6,50
72,00
9
, -8,26
-19,4
9
6,05
-6,01
m 9
a = 0,2
-27,3 9
9
-12,8
m
mn
-9,63
-5,05
9
9 J
a = 0,95 ii
0,91 i
82
e = 0,8 mm
i 1,43
-3,32 9
m
i
-5,16 9
1,08
0,51
ml
-3,90
36,19 9
1
m|
I
k = 2 W/mK mm
- 10,64 9
k = 100W/mK m 9
- 16,3 m
1,56 9
2,44 9
20,45
155,7 l
m
In
-3,18
- 15,3 9
Table..2: Solar diffuse with frontal absorption
9
iron
ISES Solar World Congress 1999, Volume III
qdA,, ~i~ I d l ~
; I~
327
design desicions. In Fig.4 we present such a plot to evaluate the appropriate tube diameter in a solar diffuse with frontal absorption duct volumetric receiver. Such a figure should be complemented with Fig.5 which provides air outlet temperature.
n-tO0;~K
26OO
IJD~8 D -'2imi
O - 0,8 man
Z;k,,k 4,,,O0WbIK;T
qu/A,,2Mmtn
~,,MItK
o
[---o--- T. I I b " 0,2
tm
S"
----~-- a ,, O~ - - - B - - k m3 l l ~ K
400
in"
400
o- % ' U
l i b " OJI
180 UD,, t8
300 (phg 0
~ " 0'8 imi
M
~) aM
o
u
o,4
u
u
t.2
1
--o--D" 8 n
1010
--x--D,, 8 i m
]d.
Fig.l: Diffuse with front absorption. Temperature distributions. 0 k,,kb,,B)W/al.K;Ri
D,,180;T
al,,li~lK;
Ib,~8;
1100
800
1800 Tpl~
- o - - p,0,0; --~e-0~;
a=0.O 000,3
-x-
,,..0,3
~ 0,o ;
21100
Fig.2: Diffuse with front absorption. Effect o f conventional ;electivity at different solar fluxes on receiver efficiency.
Till " 8 4 3 K ; k1r
80 ; iilo ,, 0.8 ; elo ,, 0,2 ; RED,, 100 ;
1,0 k " 80 W/m.K
1.8 1.7
UD,, 18
1.8
D . . 2 mm
e,, 0.O mm
qr I 1.8 1,4
I --o--e- 0,O ;
1.3 1.2
",,0il
a - 0,2
-•
I~ ;
~o,,
0,2 : - - 0il
I
Y
1.1
1
On Fig.4 we may see that there is a very pronounced bend on the curves, wich means that once this point is reached, for a 5ttle increment in receiver efficiency, we'll get a huge increment in fan power consmnption to have the air flowing through the receiver. Clearly there will be a point from which it won't be worthwhile to pursue higher receiver efficiencies because extra fan power consumption will exceed extra energy generation, and will therefore have a negative impact on the overall plant efficiency. For two receivers working at the same efficiency, the best choice will be the one of lower pressure loss, and therefore, the envelop of curves in Fig.4 will define the most apropriate diameter to work at each receiver efficiency. This means that for the range of receiver efficiencies between 60% and 80 % the best choice seems to be D = 4 - 6 rnm. However, when solar receiver is coupled to a thermodynamic cycle, outlet air temperature has also influence on overall efficiency, and therefore the information on Fig.5 is also ing)ortant. Here we see how for a given air outlet temperature the smaller the receiver diameter, the higher the receiver efficiency, but what is not seen in this figure is that pressure loss for a given outlet temperature becomes much higher as receiver diameter is decreased.
I---~o" U ; - " U qo#~,,2MIm
1 ?00
U
md mass flow.
o ,,0.8 mm
O,4
U
Ng.4: Solar diffuse with frontal absorption. Effect of diame~
UD,, t8
"~'~,,~
O,4 ha,
u
0,6
U
eb,,0,2
1200
1700
2200
Tp00
U
S;k"k
b"O01id(;T
a"E48K
-
an" II b - O,il
U-
Fig.3: Diffuse with front absorption. Effect of conventional ~electivity on equivalent emissivity.
0,7-
i"
U-
%=0,2
I J D - 18
When designing a duct volumetric receiver decisions have to be made on the receiver geometry for a process optimization. As well as in the case of conventional compact heat exchangers (Kays and London, 1964), two of the main factors to be considered are heat transfer capabilities and pressure loss, which will mean fan power consumption. A good receiver design means a balance between these two factors. Since the most relevant effect of good heat tansfer to the working fluid is receiver efficiency, we propose a plot of pressure loss vs. thermal efficiency as a good tool to take
0.4-
e-'0~n
---G--D -, 2 r i
U-
--Z~-D,, 4 mE
U-
- - X - - D " e nan
0.1 -
--O---D,, II i i
0O00
1000
1000
30OO
mOO
1",,00
Fig.5: Solar diffuse with front absorption. Effect of receive] diameter.
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ISES Solar World Congress 1999, Volume Ill
4. IDEAL SPECULAR SOLAR REFLECTION We have seen that the main reason that spoils receiver efficiency are the high radiation losses that exist when avoiding excessive reflection losses. All of this is strongly associated to the diffuse behavior of duct wall to solar radiation. Indeed, a diffuse reflection is good for infiared radiation because part of the radiation emission goes towards the inside of the duct, allowing for a reabsorption of this energy. But the diffuse behaviour with solar radiation is responsible for the high dominating reflection losses that appear when increasing solar reflectivity with the aim of reducing font wall temperatures. This problem would dissapear with a specular solar reflection. Indeed, specular reflection of solar radiation will strongly increase the duct capability to transport solar radiation towards the inside of the duct. However, if a mnqace is specular for solar radiation, it also will be specular for the infrared longer wavelength radiation, and the benefit of diffuse infrared reflection would be lost, with an increase in radiation losses. We propose to overcome this limitation by using a composed surface consisting of a solar specular coating deposited on a substrate, and a semitansparent coating over it, which allows solar radiation to go throuh, but which is opaque and diffuse to infiared radiation. Such surfaces are familiar to solar concentrating technologies, since they are being developed to irro.lement on heliostats or other solar reflectors. In such applications the function of the semitransparent coating is to protect the specular coating, without bothering at all how it behaves with infiared radiation, because heliostats don't deal with such radiation. With this concept, the top performance would be expected from the perfect specular reflector (r~ = 1), in which all the solar energy incident on the aperture would be transported through the duct without any loss for being absorbed in the cavity at the end of the duct, that behaves almost like a black body if the ratio of its inner surface to the cross duct section is high enough. Collected solar energy would be transported through the duct by infrared radiation and conduction mechanisms, and transfered convectively to the air flowing inside. In such a duct volumetric receiver axial temperature gradients are always positive, whatever the duct wall infi-ared emissivity and conductivity, with the lower wall temperatures located at the tube inlet. Thermal efficiencies are incredibly high, and equivalent emissivities incredibly low. With ideal specular reflection all the selective potential of the duct geometry comes out. The geometry selectivity is so good that the best value for the rest of thermo-physical and thenno-optical properties will be the ones that less spoil geometry selectivity. The lower the thermal conductivity, the better the efficiency, because now a positive axial temperature gradient exists, and therefore, conduction heat transfer will tend to increase inlet wall temperatures with the associated increase in equivalent emissivity. In fact is enough to keep low conduetivities in the inlet duct region. Thermal conductivity is the parameter with the highest effect on this receiver performance, by being the one that most effectively can destroy the selective effect of the non-isothermal cavity. The effect of infrared emissivity on this receiver is less important, and cualitatiely depends on the value of thermal conductivity. With low conductivities, receiver performance improves when increasing infrared emissivity, which is due to the higher infrarred absorption in the inner part of the duct, decreasing the amount of infrared radiation that reaches the front
duct side, and therefore emission losses. However, with high conductivity, temperature in the front duct wall is higher (conduction is the main heat transfer mechanism in this case), and therefore high inflated emisivity increases radiative losses. However, the effect of infrared emissivity is less in'~rtant as conductivity increases. Therefore, we conclude that with the perfect solar specular duct receiver there exists a high potential for efficiency improvement, and if conductivity is low conventional selectivity mechanisms with low e are negative for receiver performance, while if conductivity is high the potential improvement with low e is very small. To illustrate all of this, in Table-3 we present the effect of different design parameters on receiver performance. The base ease is without frontal absorption and given by L/D=l 5 ; D = 2 mm ; e = 0,8 mm ; R ~ = 150 ; k = 2 W/mK ; e = 0,8, and the solar flux is 3,14 MW/m 2 , in order to achieve 1800 K air outlet temperature. We further insist on the high receiver efficiency (lath= 99,4 % at Tgo = 1800 K), showing the high potential for the specular solar reflection-infiared diffuse reflection concept. Like before, when two values appear on a parameter the first one is associated to the inner duct half, while the second one to the outer duet half.
Tg,00
e,q
Base Case
1800
0,994
0,0044
Modific. (%)
V~
D~
De,~l
k = 50 W/mK
-2,89
-4,33
2621
k= 100W/inK
-4,61
-6,98
6078
k=-50-2W/mK
-0,26
-0,10
196
e = 0,2
-0,70
-0,99
142
e = 0,2 - 0,8
-0,09
-0,12
14
e = 0,8 - 0,2
-0,52
-0,62
96
||
, l|
: ||
l|
,i
Table-3: Ideal solar specular reflector.
5. SPECULAR SOLAR REFLECTION When we consider non ideal specular solar reflection, most of the incident solar energy will be absorbed in the inlet duet side because of multiple reflections ocuning to the rays that come at high angles. Therefore, the most important effect of specular reflection is limited to this part of the duet, having less effect on receiver performance if for manufacturing or materials reasons we have to limit the specular behaviour at the inlet region. If the cavity at the end of the duct does have a finite solar reflectivity, as r, from the tube walls is increased, reflection losses will increase, but with L/D high enough these losses can be reduced to very small values. Increasing r, improves solar transport, reducing wall temperature in the inlet region and generally increasing air outlet temperature. However, this depends on ~ a t i o n value: For high irradiations higher r, gives higher receiver efficiencies because radiative losses dominate over reflective losses, but at lower solar irradiations this is otherwise. With r~ above r, = 0,8 there are
ISES Solar World Congress 1999, Volume Ill
positive axial wall temperature gradients in all the duct lenght, and therefore receiver performance is favoured with low wall conductivities. However, for smaller values of solar reflectivity there are negative axial wall temperature gradients, increasing receiver efficiency with high conductivity for the same reason as in the solar diffuse duct. For rs = 0,8 axial wall temperature gradients are almost zero, and therefore axial conductivity has little effect. Infrared emissivity has less effect on receiver performance, which increases slightly if low infiared emisivities are used in the duct inlet region. As the L/D ratio increases, reflection losses and solar transmission decrease, with an increasing equivalent emissivity. All this produces an optimum value of L/D for each solar irradiation and mass flux. For rs = 0,8, ReD=150, k=5 W/ml~ we get (L/D)opt = 15 for a wide range of solar irradiations, but receiver efficiency is almost the same for higher values of L/D. Small effects come from an axial variation of these properties. When solar absorption occurs in the front duct edge, a negative axial wall temperature gradient exists near the tube inlet, which becomes positive luther on if r~ is high enough. The effects of frontal absorption are the same as for the solar diffuse receiver. In Table-4 we quantify the effects of different parameters on the solar specular without frontal absorption (q~/A = 3,6 MW/m:), while in Table-5 we do the same for the solar specular with frontal absorption (q~/A = 2,2 MW/m2),being the reference case given by L/D = 15 ; D = 2 mm ; e = 0,8 mm ; ReD = 150 ; e = ele = 0,2 ; r~ = 0,85 ; k = 50 W/mK ; ale = 0,8. while in Figures 6 and 7 we see the effect of solar reflectivity at different air outlet temperatures (solar ~ a t i o n s ) , for the solar specular without frontal absorption, on receiver efficiency and equivalent emissivity,
329
in the front duct part, and diffuse solar behaviour in the inner duct region, generally reflection losses and equivalent emissivity increase, and therefore receiver efficiency decreases, although if we keep high solar absorptivity in the inner region, the efficiency reduction is not very important. Nevertheless, with high r~, receiver efficiency improves when introducing low solar reflectivities (diffuse or specular) in the duct's inner region because of reducing reflection losses, which are significative in this case. This opens the way to reduce the temperature operation limitations imposed by the solar-specular/infiared-diffuse material on gas outlet temperatures. Still these limitations are rather important. With r,,e=0,8 in the outer duct region and rs,d = 0,2 in the inner duct region, if the maximum operating temperature of the specular material is 900 ~ solar irradiation has to be kept below 1,5 MW/m2 , and air outlet temperatures under 1140 IC The higher the specular solar reflectivity, the less important the material temperature limitation. |_
Tie (K)
.. -
nu
Base Case
nn
1852
0,76
n
Modific. (%) " e = 0,8 ii
.. k = 2 W/mK k=-100W/mK ,. i rs = 0,5
i
DTt~ n
-2,36 1 ! -2,29
~
a
n
TgoOK)
hth
eeq
0,07 -3,96
Base Case
1800
0,901
0,46
Modif. (%)
DTgo
Dhth
Deeq
II
-1,94
-3,02
36,76
k = 2 W/mK
1,36
2,14
-49,02
k=-IO0 W/mK
-0,50
-0,78
14,75
r, = 0,5
-3,19
-4,96
88,84
10,25
1,99
-0,49
45,69
5,78
-35,56
-5,48
-17,11
-21,94
3,21
4,78
73,96
-6,18
- 11,22
1,15
-0,30
-0,34
l
-27,47 9
ale = 0,95
3,89 l
9
6,12 l
ele = 0,9
-5,07 9
ele = 0,1
0,94 9
kle=100W/mK
0,024 l
I"
9
0,037 l
kle= 2 W/mK
ll
9
1,46
l
ii
l
-7,84
l
.. ii
ii
-18,06
-0,12
1
,|
e=0,8
3,54
l
-3,09
l
l
ale = 0,2
ii
||
2,26
31,33
l
9
l
DTle
n
l
rs = 0,95
i
Dem
' -3,56
-2,55
1902
m
-3,66
0,05 l
0,81 a
Dhth
m
l |_
nn
-1,25 9
l
-1,95 |
13,86
16,27 9
n
nI
Table-5: Solar specular with front absorption.
r~ = 0,95 2,50 3,93 -54,05 n' Table-4: Solar specular without front absorption.
6. STABILITY ISSUES As it may be seen in these tables without solar frontal absorption, the real solar specular duct volumetric receiver has very high efficiencies in spite of a considerable incease in equivalent emissivity. However, frontal absorption reduces a lot the receiver efficiency, and therefore will be the main limiting factor to benefit from this concept. Tube inlet edges should be sharpened with the minimum manufacturable and structurally admissible slope in order to tend towards the no front absorption limit. When exploring the posibilities of combining specular reflection
In the last years analytical (Kribus, Ries and Spirkl, 1995) and numerical (Pitz-Paal and Hoffschmidt et al, 1997) studies have claimed that duct volumetric receivers have inherent limitations on the maximum solar flux they can manage due to unstable behaviour, although this has not been experimentally confirmed. The hot points that have been observed experimetally are likely to be due to the synergic effects that anise when exposing parallel receiver channels to non-uniform solar irradiation, and problably have nothing to do with unstabilities.
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330
First of all we have to point out that the unstabilities which have been mentioned are only relative, and don't mean that one couldn't work with such receivers without any problem Indeed, the unstable behaviour depends on the characteristic pressure lossmass flow curve of both the fan and the receiver, and the condition for such unstability to affect the system is that in the working point the fan curve slope is higher than the receiver curve slope. One speaks about unstable branch in the receiver curve when the slope is negative, and about unstable branch on a fan curve when the slope is positive. Most fans do have an unstable branch and they are used without any unstable behaviour, so one shouldn't panic if a receiver curve would come out to have an unstable branch in its characteristic curve. A bit of engineering care (mostly in the selection and operation of the fan) should be enough to guarantee a stable behaviour of such receiver in all working conditions. T~,, 643 K ; I ~
o" le0 ;
e-. 0~ ;
r,o,, 0,2 k - E0 W/InK UD,, 16
US
#-
D - - 2 mm e,, U mm
us --G-- r m , , U
U
- - ~ - r , . - US -o--r,.- U - x - t u - 0,2
0,rS 0,7 800
1800
1000
2000
21500
Tp(K)
Figure~: Solar specular without front absorption. SolaJ -effectivity and solar flux effects. T#,,E43K;Re
a,,150;
e,,0,2;
r,o,,0,2
1.1. k , , 80 W/inK
-
/
I.A) ,, 18 D , , 2 mm
11,7
-
o m 0mS IIIIll
r u , , 0,8
..-b--r u , , 0,118 -•
11,1
,~"
0,2
D
tO00
11110
2000
2500
1"1,00 Rgure-7: Solar specular without front absorption. Effect ol ~olar reflectivity on equivalent emissivity. ,
However, the hypothesis on which are based the models that were used to predict such unstabilities have important simplifications of the heat and flow processes that occur in a duct volumetric receiver. As the unstability has its origin on the dynamical viscosity dependence on temperature, and the hypothesis in this models may affect the temperature distribution inside the duct, one can not conclude that such unstable branches exist without a more detailed analysis. On the other hand, in ( P i t z -
Paal and Hoffschmidt et al, 1997) they only find the unstable branch in the duct receiver after applying the model to several volumetric receivers, being this the one for which the model is less appropriate, and justified the fact of experimentally not finding the unstable branch by the radial conduction effects. With our model, we didn't find any unstable behaviour even at much higher solar fluxes, and without introducing radial conduction effects. More specifically, the model simplifications in (Kribus, Ries and Spirld, 1995) that could affect the stability are: - No inertial effects in momentum equation (asymptotic limit for Reynods ~ 0). - Temperature effects only on density and viscosity. - No radiative exchange calculation inside the receiver, and therefore no reabsorption mechanisms for infrared radiation inside the receiver matix. - Uniform solar absorption in the receive matrix, while most of the absorption is in the font part, even for a solar specular duct. - Infinite convection coefficient. - No conduction in receiver matrix. The numerical receiver model used in (Pitz-Paal and Hoffschmidt et al., 1997) is more sophisticated. It accounts for solar absorption, radiative, convective and conductive heat transfer in a one channel model, and then goes further by combining these results to evaluate the effect of non uniform solar irradiation and radial conductivity. However, it is still based on some simplifications that are not appropriate for the analysis of the duct volumetric receiver. - lnfi'ared radiative exchange modeled with a simplified version of participating media with isotropic scatering and specular reflection at the duct end, which is good, and in fact one of the few things one can do, to model porous media, but not a duct receiver. - No spectral dependence ofthenno-optical properties. - It doesn't include the entrance length effects on heat transfer, friction and fluid acceleration. These effects are equivalent to the inertial effects that in this same reference are identified as stabilizing factors. - It doesn't include temperature dependency of other fluid properties than density and viscosity. - It doesn't include inertial effects in the momentum equation except for the shape pressure loss. - It also doesn't consider the possibility of secondary flows and mixed convection transfer to occur, although with small duct diameters Rayleigh numbers won't be high enough for it to happen at the solar ~ a t i o n they used. We applied our models to evaluate pressure losses under different circumstances, and never found an unstable branch in the receiver characteristic curve, in spite of calculating with much higher solar fluxes (up to 10 MW/m2) than in these references. However, most receiver characteristic curves show an inflection point that reminds of the mechanism that seemed to generate the unstable branch in the other models. There are several factors that increase the tendency of this inflection point towards unstability, but without ever reaching it. We observed that the tendency towards unstability grows as air inlet temperature is lowered, L/D is increased, or duct diameter is reduce& The first of this factors affects through the higher variation of air viscosity through the duct, and the two last ones affect because the less influence of the entry lengh effects. In Figure-8 we present some results of receiver
ISES Solar World Congress 1999, Volume III
characteristic curves for a diffuse duct receiver with fontal absorption working at different solar fluxes. The pressure loss is presented on a non dimensionalized way to better appreciate how tendency towards unstability grows as higher solar irradiations are used. But still at 2,5 MW/m2 the inflection point is far from creating an unstablre branch. In Figure-9 we present the results for a specular receiver with front absorption and different L/D ratios working at 2 MW/m 2. k - k ssuS0'filnl(;
om osou0,2; itse-0,8 ;
/
100
00
r u -0B
-x--qo/A 9U IdW/m t --n--qoJ~- 1 MVlln " -u--q,~-U~ 9
(SL)
00 i
L/D~6
Id
Do|l e"Um
6
21)
331
conventional e/a selectivity doesn't have a very important effect. For both specular or diffuse solar reflection, frontal solar absorption reduces a lot receiver efficiency, diluting other selectivity mechanisms. Sharpening of the leading edge with such a small angle as structurally posible, and conventional e/a selectivity stategies in the frontal edge have to be used to get higher receiver efficiencies. A plot of receiver presure loss vs. receiver efficiency is introduced as a useful tool to take design decisions as appropriate values o fL/D and D. When using a receiver model incorporating the effects that were left out in former studies, we didn't find any unstable branch on the receiver characteristic curve up till 10 MW/m 2 solar irradiation, which is much higher of what is needed to work with modem gas turbines. Therefore we conclude that there is no inherent limitation on the maximum solar irradiation a volumetric receiver can manage without presenting possibilities of unstable beheaviour. NOMENCLATURE
0
2
4
8
8
ae ~ t) Figure..8: Solar diffuse with frontal :urves at different solar fluxes. q,/h-2~ lOO
t ; T 6,,648K;
absorption. Characteristic
ro-0,O;k,,Ido,,O01WmK
-
-o--Lq) - S0 --x--LA~ - 80 ~LJD
~)
-" 18
--I~--IJD - 8 M.
D-'Rmm
I
o " O,ll mm
/
B
e-
el,-U
11)
0
2
4
8
a : Duct wall solar absorptivity. Leading edge solar absorptivity. Dl~t: Total air pressure loss through receiver D : Duct diameter. e: Duct wall thikness. e : Infrarred wall emissivity. elo : Leading edge infrared emissivity. e~q :Equivalent emissivity. hth : Receiver thenml efficiency. k: Thermal wall conductivity. kle : Thermal wall conductivity on the between leading edge and first wall node. L : Duct length. qs/A : Solar irradiation on receiver aperture. rs: Solar reflectivity. r~e: Specular solar reflectivity. rs,d" Difi~e solar reflectivity. Tgi : Inlet air temperature. "['go : Outlet air temperature. ale:
8
Figure-9: Solar specular with front absorption. Effect of L/E "atio on characteristic receiver curve.
7. CONCLUSIONS In this paper we have explored the potential of different selectivity strategies to improve solar receiver efficiencies. For solar diffuse duct receivers, wall conductivity plays an in'gmrtant role and conventional e/a selecivity works although can't bring a very in~rtant efficiency improvement. Important effects are restricted to the entrance region of the duet receiver. A very high potential for efficiency improvement is associated to the new concept of solar-specular/infiared-diffuse wall behaviour. Again the main effect of thennophysical properties is located in the duct entrance region. Small wall conductivities also have a significant effect on receiver efficiency improvement, but
REFERENCES Becker M., Cordes S., B6hemer M. (1992) The Development of Volumetric Solar Receivers. In Proceedings of the 6th Int. Syrup.
on Solar Thermal Concentrating Technologies, MAdrid, Spain, pp.945-952 B6hemer M., Chaza C. (1991). The Ceramic Foil Volumetric Receiver. Solar Energy Materials 24, pp.182-191 Chavez J.M., Chaza C. (1991). Testing of a Ceramic Absorber for a Volumetric Air Receiver. Solar Energy Materials 24, pp. 172-
181 Holman J.P. (1997). Heat Transfer. 8th ed., McGraw-Hill Flamant G., Meningault T., Olade G. (1987). Nouveau Dispositif d'Abnsorption S61ective de l~Energie Solaire Concentr6e par des Lits de Particules. C.R.Acad.Sci.Paris t 304, SbT"e 11, 3, pp.689-
332
ISES Solar World Congress 1999, Volume III
694. KalTli J., Rubin IL, Kribus A., Doron P., Sagie D. (1996). Test Results with the Direct-Irrdaiated Annular Pressurized Receiver. In Proceedings of the 8th Int. Symp. on Solar Thermal Concentrating Technologies, 6-11 October, Cologne, Germany, M.Becker, M.BiJhmer (Eds), pp.607-621, C.F. Mailer Kays W.M., London A.L. (1964). Compact Heat Exchangers. 2nd ed., McGHraw-Hill Kribus A., Ries H., Spire W. (1995). Inherent Limitations of Volumetric Solar Receivers. Solar Engineering-VoL1, ASME 1995, pp.649-655 Lin S.H., Sparrow E.M. (1965). Radiant Interchange Among Curved Specularly Reflecting Surfaces-Application to Cylindrical and Conical Cavities. Journal of Heat Transfer, May 1995, pp.299-307 Meningault T., Flamant G., Rivoire B. (1991). Advanced HighTemperature Two Slab Selective Volumetric Receiver. Solar Energy Materials 24, pp 192-203 Pitz-Paal R., Morhenne J., Fiebig M. (1991). A New Concept of a Selective Solar Receiver for High Temperature Applications. Solar Energy Materials 24, pp.293-306 Pitz-Paal IL, Fiebig M. (1992). First Experimental Results from the Test of a Selective Volumetric Air Receiver. In Proceedings of the 6th Int. Syrup. on SOlar Thermal Concentrating Technologies, Madrid, Spain, pp.277-289 Pitz-Paal IL, Hoffschmidt B., B6hmer M., Becker M. (1997). Experimental and Numerical Evaluation of the Performance and Flow Stability of Different Types of Open Volumetric Absorbers Under Non-Homogeneous Irradiation. Solar Energy, vol.60, Nos. 3/4, pp.135-150. Rabl A.. (1977). Radiation Transfer Through Specular Passages-A Simple Aproximafion. lnt. J.Heat Mass Transfer, vol.20, pp.323330 Shah ILK., London A.L. (1978). Laminar Flow Forced Convection. Acadmic Press Siegel IL, Howell J.IL (1992). Thermal Radiation Heat Transfer. 3rd ed., Taylor & Francis Spirkl W., Ries H., Kribus A. (1997). Performance of Surface and Volumetric Solar Thermal Absorbers. Journal of Solar Energy Engineering, May 1997, VoL119,pp.152-154 Variot B., Meningault T., Flamant G. (1992). Modelling and Optimization of a Two-Slabs Selective Volumetric Solar Receiver. In Proceedings of the 6th Int. Syrup. on Solar Thermal Concentrating Technologies, Madrid, Spain, pp.325-345
ISES Solar World Congress 1999, Volume III
333
A PARABOLIC DISH CONCENTRATOR FROM A TELECOMUNICATION ANTENNA: OPTICAL AND THERMAL STUDY OF THE RECEIVER
Claudio A. Estrada Centro de Investigaci6n en Energia, UNAM, A.P. 34, Temixco, Morelos, 62580, M6xico. Tel. (052 73) 25 00 48, 25 00 44, e-mail: [email protected], [email protected]. Rubtn Dorantes Departamento de Energia, UAM-A, Av. San Pablo No. 180, Col. Reynosa, Tamaulipas, 02200 Mtxico, D.F.
Eduardo Rinc6n Facultad de Ingenieria, UAEM, Cerro de Coatepec s/n, Toluca, M6xico.
Abstract - In Mtxico, Telcom, a big telecommunication company has been removing, due to the installation of new systems, hundreds of parabolic antennas made by aluminum and other light materials without further future use. Now, it is being investigated the adaptation of those parabolic antennas as parabolic dishes (see fig. 1) to be used as solar concentrators for thermal conversion of solar energy. This paper, which is a continuation of Estrada et al (1), presents theoretical advances of the research focused on the determination of the optimal shape of the receiver and the thermal losses associated. It is intended to have low cost, parabolic dishes with high surface's reflectance and high receiver's efficiencies. The antenna's dimensions are 3.32 m in diameter, with focal length of 0.83 m, which gives an aperture angle of 90*. For the theoretical analysis, a software package for facilitated optical analysis of 3-D distributed solar energy concentrators called CIRCE2 (2) was used. The flux distributions for a cylindrical and a conical shape receivers were analyzed and the error function of the concentrator was determined giving a value of a = 2 mrad. The peak local concentration for the cylindrical receiver was found to be 2040 suns and the geometrical concentration was 860 suns. With these concentrations it is possibly to think in temperatm'es as high as 1000 K or even greater. So far, the results indicates that a simplified and cheap systems can be built with those antennas to allow the conversion of solar to thermal energy for high working temperatures.
1. INTRODUCTION Some researchers in Mexico are interested in finding a new use for several parabolic antennas that have been removed by a telecommunication company in Mtxico named Telcom, due to the installation of new systems. Aluminum and other light materials make the antennas. The dimensions of each one are 3.32 m in diameter (an aperture area of 8.66 m:) with a focal length of 0.83 m, which gives an aperture angle of 90 ~ and with a total weight of 210 kg. The Japanese NEC Company made those antennas in 1964 and it is estimated that there are around 300 of those antennas in Mexico. The Autonomous Metropolitan University got free ten antennas and all of them are in very good conditions. Now, it is being investigated the adaptation of those parabolic antennas as parabolic dishes to be used as solar concentrators for thermal conversion of solar energy. The goal is to transform the antennas into low cost, parabolic dishes for solar thermal conversion with high surface reflectance and efficient receivers. In a previous work, Jimtnez et al (1997) showed that the antenna aluminum surface can be polished until get an average reflectance of 0.75 between 400 a 700 nm (visible region) and 0.92 for the infrared region until 3 000 rim. The cost of preparation and polish for each antenna's surface was estimated to be 200 USD, including materials and labor. To have a good surface reflectance is just the first step in the adaptation of the parabolic antenna as a solar concentrator. Now, it is necessary to quantify the deviation of the surface from a perfect paraboloid that will determine the form and dimensions of
the optimal receptor needed to capture the concentrated solar energy. This paper presents advances in the experimental and theoretical studies to determine the focal zone's dimensions and the optimal shape of the receiver and the thermal losses associated. 2. DISH EXPERIMENTAL CHARATERIZATION The ideal focal zone of the antenna as a solar concentrator is an ellipsoid, which has major and minor axes of 1.54 cm and 0.77 cm respectively. These values come from the solar rim angle (16' = 4.653 mrad) and the antenna's dimensions (rim angle of 90*). Due to manufacturer and polish processes the actual antenna's focal zone is greater and different to that ellipsoid. To determine experimentally the deviation from the ideal focal zone a laser scanning technique was used, see figure 1. Estrada et al (1998), showed that using a 100 cm long rod of 1 cm in diameter located at the principal axis of he paraboloid and the laser scanning technique, it was possible to determine the height of the focal zone around 6.3 cm. Additional to the rod, two more bodies were located as receivers at the focal zone of the concentrator and the laser scanning process was applied for each one. One was a cylinder with a diameter of 5 cm and a height of l0 cm. Another was a cone with a diameter of 10 cm and a height of 6 cm. Fig. 2 shows the three receivers and the laser beams hitting on them.
334
ISES Solar World Congress 1999, Volume III
Fig. 2. Concentrator receivers with the reflected beams hitting on them. a) long rod, b) cylinder and c) cone.
Fig. 1: Parabolic dish in vertical position
For the two other bodies, it was found that 97% of the reflected beams hit the receivers. The geometrical concentration for the cylinder was found to be 860 suns, while for the cone it was 700
Concentrator imperfections such as slope errors, surface roughness, random facet misalignment, etc., have impact on the actual normal at the concentrator surface. These imperfections give rise to error distributions, which assign a probability to the chance that the actual surface normal will take a given direction. Concentrator imperfections can usually be model with 2-D elliptic normal error distributions. If it is assumed that the errors are truly random without a predisposition in any direction, it is appropriate to model them with 1-D circular-normal distributions. These distributions, special case of the elliptic-normal distributions, are used for most error types. The equation representing this circularsymmetric Gausssian distribution error is 1 E1D(r) = 2 m 2
exp[
r2 - 2-'~" I
(1)
SUNS.
It is important to note that the optimal rim angle for a paraboloid concentrator is 45* and its receiver is normally design as a cavity with a fiat aperture perpendicular to the paraboloid principal axis. However, the antenna dish has a rim angle of 90 ~ implying that an external receiver should be uses to capture the concentrated solar energy. This explains the shapes of the receivers used. 3. DISH T H E O R E T I C A L CHARATERIZATION The theoretical analysis of the parabolic concentrator was done using a software package to facilitate optical analysis of 3-D distributed solar energy concentrators developed by Romero [2] and named CIRCE2 (Convolution of Incident Radiation with Concentrator Errors). This software allows a user to efficiently build and analyze a variety of point-focus solar concentrator. CIRCE2 was set up for each of the three receivers along with the paraboloid dish parameters. For each case, the program was run and the flux distributions along the exterior receiver's surfaces were obtained. Because the axisymmetrical configuration of the receivers the flux distributions are just a fimction of the Z coordinate (see fig. 2).
where r is the local radial coordinate where the solar ray is hitting the surface and 6 is the standard deviation or dispersion parameter associated to the concentrator. Eq. 1 also called the error function, characterizes the optical errors of the solar concentrator and the parameter ~ defines how big is the deformation of the reflected sun cone due to the surface's errors. For an ideal concentrator ~ is zero and thus Em is from Eq. 1 equals to zero. To run CIRCE2 it is necessary to define the value of a. Also, Estrada et al (1998), showed that using CIRCE2, it was possible to determine a ~ value of 2 mrad for the antenna which produces a width distribution of 6.1 cm being in agreement with the 6.3 cm found experimentally. Therefore, it is established that the dispersion parameter a associated with the antenna dish is 2 mrad. Fixing G to the value of 2 mrad, CIRCE2 was set up for the cylinder and the cone receivers. Fig. 3 shows the flux distributions as a function of the axial Z-coordinate. The maximum local concentration obtained for the cylinder and the cone were 2040 and 907 suns respectively. In both cases, the distributions have a maximum at the intermediate zone of the receiver's surface loaded to the positive direction of Z. It is clear from this figure that the cylindrical receiver is the one that gave the higher concentration and the one that should be recommended as a system's receiver.
ISES Solar World Congress 1999, Volume III
Q = mCp
AT At
335
(3)
where Q is the energy incident into the cylinder, m and Cp. are the mass and the heat capacity of the cylinder, AT is the difference between the average temperature of the solid and its initial temperature, and At is the increment in time.
Fig. 3. Cylinder and cone flux distributions as a function the axial Z-coordinate for o = 2 mrad.
4. T H E R M A L CHARACTERIZATION For the thermal characterization, an iron steel cylinder 5 cm in diameter and 7 cm long was located at the concentrator "s focal zone. Six thermocouples type K were located in the cylinder. Three of them .3 cm from the cylinder's surface, and the other three 2.5 cm from the surface, that is, at the cylinder's center. Figure 4 shows the thermocouple positions and named them. Fig.5 Temperature distributions for thermocouples T2, T3 and T4, for several times.
For the experiment shown in Fig 4. m = 1.66 kg, Cp= 473 J/kg, Tp = 98 C (Tp is the weighted average temperatures of the solid), To = 35 C and At = 10 sec., then Q = 4.946 kW. The energy incident at the concentrator was 822 W/m 2, and the concentrator area is 8.6 m 2, thus the input energy was 7.069 kW. This gives a thermal efficiency of the concentrator of 0.7 % for this conditions. This could be understood as the upper limit for the thermal efficiency of the concentrator, for a receiver working a much higher temperature, the thermal efficiency is expected to decrees. Fig. 4. Cylindrical receiver and thermocouple positions.
The concentrator was aligned with the solar rays and the concentrated solar radiation was allowed to strike the cylindrical receiver. Fig 4 shows the temperature distributions for thermocouples T2, T3 and T4, for several times. It is clear comparing Fig. 3 and Fig. 4., that the flux distribution is in correspondence with the rise in temperature at the first seconds of the transient processes. If we assume that at the first seconds of the solar exposition, all the losses by convection and radiation are neglected, then all the energy incident to the receiver will increase the internal energy of the receiver's mass, that is, the energy balance gives Q=mCp
or
dT dt
(2)
5. CONCLUSIONS The laser scanning technique as presented in this paper was adequate and it allowed finding the height of the focal zone to be around 6.3 cm. The error function of the concentrator was determined giving a value for the dispersion parameter (6) of 2 mrad. The flux distributions for a cylindrical and a conical shape receivers were analyzed and it was found that the cylindrical receiver is the one that gave the higher concentration and thus the one that should be recommended as a receiver for the dish concentrator. The pick local concentration for the cylindrical receiver was found to be 2040 sun and the geometrical concentration was 860 suns. And experiment was conducted to show that a upper efficiency of 70% can be reached by this concentrator. It is possibly to think in temperatures as high as 500 C or even greater, even though the efficiency can be low. So far, the results indicates that a simple and cheap systems can be built with those antennas to allow the conversion of solar radiation to thermal energy for medium working temperatures. The next step is define a simple and cheap structure to support the
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antenna with a control system, as well as the heat exchanger for the receiver to be used in a useful thermal processes. A cost effective study should be made.
6. REFERENCES Jim6nez, M.R., F. Hemfindez and R. Dorantes, Construcci6n de un Concentrador Solar a Partir de una Antena Parab61ica., Proceedings of the 21 st National Week of Solar Energy, Mexican Solar Energy Society, 1997. Romero, V.J., Cice2/Dekgen2: A Software Package for Facilitated Optical Analysis of 3-D Distributed Solar Energy Concentrators. Sandia National Laboratories, SAND91-2238, 1994. Estrada, C.A., R. Dorantes, F. Hem~indez, M. Martin and E. Rinc6n, 1998. A Parabolic Dish Concentrator From a Telecommunication Antenna. Proceedings of the 1998 Annual Conference, American Solar Energy Society. Vol 1, pags.267-270. Editado por R. Campbell-Howe, T Cort6z and B. WilkinsCrowder ASES, USA.
ISES Solar World Congress 1999, Volume Ill
337
EFFICIENCY IMPROVEMENT OF PARABOLIC TROUGH COLLECTORS BY MEANS OF ADDITIONAL END REFLECTORS
TH. FEND, J. LEON l, P. BINNER, R. KEMME 1, K.-J. RIFFELMANN AND R. PITZ-PAAL Solare Energietechnik, Deutsches Zentmm ftir Luft- und Raumfahrt e.V. (DLR), Linder H6he, D-51147 K61n, Germany, Phone: +49 2203 601 2101, Fax: +49 2203 66900, E-Mail: [email protected] 1plataforma Solar de Almeria, Carretera de Senes s/n, Tabemas (Almeria, Spain) An additional collector element designed for the attachment to existing north south aligned parabolic trough collectors was tested on the Spanish Test-Center Plataforma Solar de Almeria (PSA). It may be installed at the north end of a collector module and reduces end losses occurring during non-normal incident solar radiation. Thermal efficiency tests were performed by nmning a test loop including a 48 m parabolic trough collector of the LS-3 type. It could be shown by calculations and experiments, that with the use of the additional reflector collector efficiency increases significantly. From the data acquired during the experiments an annual efficiency could be calculated, which is 42.6 % compared to 39.9 % without the additional reflector. Abstract.
1. I N T R O D U C T I O N incident angle (Dudley et.al., 1994). One consequence, which has an influence on these factors, is the occurrence of end losses. They are caused by radiation penetrating through the aperture area at the north end of the collector (for the northern hemisphere), which cannot be concentrated on the absorber tube. Additionally, north-south oriented troughs with a one axis tracking system loose radiation penetrating through the north end of the collector. As has been described in more detail in previous publications (Binner, 1997; Fend et. al., 1998b), an additional end reflector at the north end of the collector may be used to eliminate both kinds of these losses (Figure 1). Calculations show that flux on the absorber tube increases by nearly a factor of two in a zone near the additional mirror.
Parabolic trough collectors concentrate direct solar radiation on a line focus, where absorber tubes convert radiation into heat, which is transported by a liquid medium, most frequently oil or water. The generated thermal energy may be used for process heat or electricity generation (Becker and Gupta, 1995). For economic reasons parabolic trough systems follow the sun by means of one axis tracking systems, which leads to non-zero angles of incidence of the solar radiation in most times of the day. When the troughs are aligned in a North-South direction, the annual average of this angle of incidence is minimized. However, the angle impacts the optical efficiency of the collector. Geometrical and optical losses are usually characterized by so called IAM-factors (Incident _Angle Modifier) as functions of the
/
,, , , / / / /
,
9
,,/,,/,,/
,
~X
/
/
receiver J r J"" i
/
/
concentrator
I end reflector
Figure 1: Reflection at the front side of the collector with end reflector
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338
Measurements carried out on a parabolic trough test-loop with an aperture of 2 m2 and an 0.6 m2 end-reflector confirmed these calculations (Fend et. al., 1998). The biggest plant using parabolic trough technology is in operation in the SEGS-plants at Kramer Junction, California, since 1987 (Cohen et.al., 1998). Since this time the plant has produced an average annual electricity output of approximately 400 GWh. 2. DESCRIPTION COMPONENTS
OF
THE
TEST
LOOP
Israeli company LUZ-Intemational Limited and is the third generation of this collector technology. It consists of parabolic collector modules of 48 m length and an aperture width of 5.76 m. 112 thick glass mirror facets produced by Flachglas AG, Germany concentrate direct solar radiation on a heat collecting element (HCE) consisting of an evacuated glass envelope with an inner steel tube, which is coated with a selective black surface (see Table 1 for further technical data). The HCE is designed for Temperatures of up to 400~ Usually oil is used as a heat transfer medium. One module of this collector type was erected at the Plataforma Solar de Almeria (PSA) in 1997 for testing purposes (Figure 2).
2.1 CollectorThe so-called LS-3 collector used at the SEGS plants mentioned above was developped by the US-
Figure 2: Photograph of the 50 m LS-3 test-loop on the PSA
Table 1:LS-3 Collector Technical Data Aperture Width Total Aperture Focuslength Total Collector Length Number of Mirror Facets
[m] [m [In] [m]
Dimensions of Mirror Facets
[m]
Number of Receiver Elements Length of Receiver Element Outer Absorber Tube Diameter Absorber Tube Wall Thickness Glass Envelope Diameter Glass Envelope Wall Thickness Thermo Fluid Maximum Outlet Temperature Maximum optical efficiency Maximum Collector Efficiency @400 ~
[in] [mini [mini [mini [mm] [~ [%] [%]
5,76 272,5 1,71 48 112 1,70xl,63 1,70xl,50 12 4 70 2,25 115 3 400 78 68
ISES Solar World Congress 1999, Volume III
2.2 Test Loop The structure of the test-loop is drawn schematically in Figure 3. An electrical heater is integrated to enable a quick warm up to working temperatures of up to 400~ As a heat transfer medium Syltherm 800 is used, which is
339
stable at temperatures of up to 400~ The pump enables maximum volume flows of 10 m3/h. Volume flow, collector position, heater and cooler can be controlled via Computer.
Figure 3: Schematic drawing of the PSA's LS-3 test-loop
2.3 Measurement and Data Acquisition Temperature of the thermofluid is measured at the collector inlet and outlet as well as in the middle of the collector by thermocouples. A Volume flow sensor is integrated behind the pump. System pressure is measured before the pump. Signals from these sensors as well as Direct Solar Radiation (DNI) and some more weather data are monitored by a Data Acquisition System (DAS), which is connected to a PC. Data from the DAS is acquired every second, 5 minute average values are generated and stored.
2.4 End-Reflector For the LS-3 test-collector described above a 6 m2 end reflector was designed and installed in April 1998 (Figure
4). Special attention was paid on costs, lightweight and simplicity of installation. The collector consists of 8 facets of high specular aluminium coil material with average solar reflectance values of 88-90 %. Durability and optical performance properties have been investigated in previous investigations (Fend et.al., 1998a). The facets were glued on aluminum frames, which were connected to steel supports by screws allowing a precise adjustment with respect to the geometry of the concentrator and the absorber. Each facet has been adjusted separately with a laser beam simulating the sun's radiation from different incident angles. The steel supports were easily connected to the LS-3 by screws without dismounting any part of the framework of the existing collector (Figure 3).
Figure 4: Front and backside view of the end reflector for the LS-3 test loop at the PSA
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4. RESULTS AND DISCUSSION Efficiency tests have been performed in February, April and July 1998. Due to accuracy problems of the tracking unit and some misalignment of the LS-3 mirror facets, the optimum possible solar to thermal efficiency of more than 60% was not reached. Efficiency rl of the collector module was calculated from volume-flow 17, density p , specific heat c t, and temperature data of the heat transfer fluid as well as from Direct Normal Incident Solar Radiation data (I) using
q
~
#. p.c
A denotes the aperture area of the collector. Five days of reference testing were followed by 3 test days with endreflector. At each day incident angles from about 40 ~ in the morning to 0 ~ at solar noon were observed. As an example, measured data of one day reference testing are shown in Figure 5. After a phase of additional electrical pre-heating in the morning, an inlet temperature of 250 ~ C was adjusted. Maximum efficiency of this day (44%) was reached at 2.00 p.rm, 30 minutes aider solar noon. Due to geometrical losses efficiency in the morning hours is markedly lower.
"(L., -T,n) I.A
Figure 5: Efficiency, Volume Flow and Temperature data measured at the LS-3 test-loop at the PSA on February 19, 1998
ISES Solar World Congress 1999, Volume III
-
1.2-
~
~
~
~
.
.
.
.
.
! end reflector
,
..u
~. e~
.
--
~i
t
.
...............
*' 0.4 Q >
341
,m
t~ 0
m,,
~
'
no lend reflector
Im
0 0
10
20
30
40
Incident angle
Figure 6: Relative Efficiency as a function of incident angle of the LS-3 Collector at the DISS-Reference loop of the PSA comparing the performance of an additional end-reflector with the original configuration To characterize the optical properties of a parabolic trough collector, solar to thermal efficiency rl is often replaced by relative efficiency rl/rl0 with 110 denoting collector efficiency for 0 = 0 ~ incident angle of the solar radiation. Relative efficiency values rl/rl0 are shown in Figure 6 as a function of the incident angle. Each data point was taken from the above mentioned 8 testing periods. In order to generate the data points in Figure 6, 20 min average values were taken from the data of each day, which was monitored in 5 min intervals. Finally the data points were fitted with second degree polynomial functions. The positive effect of the end-reflector is significant. An efficiency increase of up to 20% can be observed. 5. CONCLUSIONS A prototype of an end reflector could be tested successfully on the LS-3 test loop at the PSA. Thermal tests have shown, that collector efficiency can be increased significantly. Furthermore it could be shown, that high specular aluminium coil material can be successfully applied in parabolic trough technology. To calculate a possible annual efficiency of a collector of the LS-3 type, the data shown in Figure 6 were applied on a north south collector located 35 ~ northern latitude. An annual efficiency of 42.6 % was estimated compared to 39.9 % without an additional end-reflector. However, tests were performed on PSA's LS-3 test loop with an overall length of 50 m. Commercial plants usually consist of longer troughs. Assuming an annual DNI of 1800 KWh/m 2 the end-reflector would produce an annual yield of 13400 KWh. Costs of the end-reflector have been estimated to approximately 1500 Euro.
References Becket, M., Gupta, B. (1995) Solar Energy Concentrating Systems, p25, p40, C.F. Miiller, Heidelberg Binner, P. (1997) Wirkungsgradsteigerung von Parabolrinnenkonzentratoren durch Zusatzreflektoren, DLR, IB 375-97/01 Dudley, V.E., Kolb, G.J., Sloan, M., Keamey, D. (1994) SEGS LS-2 Solar Collector, Sand94-1884 Duffle, J.A. Beckmann, W.A. (1991) Solar Engineering of Thermal Processes, John Wiley & Sons, Inc., New York, p 22 Fend, Th., Jorgensen, G., Bthmer, M., Kr/imer, T., Rietbrock, P.M. (1998a) First Surface Aluminium Mirrors: an Assessment for Solar Outdoor Applications, Proc. Eurosun '98, Sept. 14-17, Portoroz, Slowenia Fend, Th., Le6n, J., Bthmer, M., Binner, P., Deidewig, F., Kemme, R. (1998b) Development and Test of an End Reflector for Parabolic Trough Collectors, 9th Int. Symp. - Solar PACES - Solar Thermal Concentrating Technologies, 22-26 June 1998, Odeillo - Font-Romeu Geyer, M., Hollfinder, A., Aringhoff, R., Nava, P. (1998) H/ilfte des weltweit produzierten Solarstroms. Sonnenenergie 3, 33-37
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342
EXPERIMENTAL PERFORMANCE OF A PV V-TROUGH SYSTEM N. Fraidenraich and E. M. de S. Barbosa
Research Group on Alternative Sources of Energy- FAE, Universidade Federal de Pernambuco- UFPE. Av. Prof. Luiz Freire 1000 - Cidade Universit~a, 50 740 540 Recife-PE/Brasil Phone / Fax: +55 081 271 (8252) / (8250) e-mail: [email protected]
Abstract - We describe the design procedure and the experimental facility built to asses the technical viability of a Vtrough tracking PV system. Experimental results for radiation collected at the cavities' aperture and absorber regions are presented. The optical properties of the cavity are described theoretically and compared with experimental results as well as the output electric power when the PV system is connected to a bank of batteries. Both, optical and electrical estimations have a good agreement with experiment. Thus, the simple theoretical tools used to analyze the system's performance can be reliably used for design and evaluation purposes. Experimental results, represented as a function of diffuse to hemispherical radiation ratio, yield a set of empirical correlations which have been used to simulate the system behavior along the year. Results of simulation show that the annual benefit in electric energy due to tracking is equal to 26 % and the increase of electric energy output for the V-trough system is 72%, both with respect to an horizontal fixed collector.
1. INTRODUCTION
to determine the cavity parameters (Fraidenraich, 1998).
Photovoltaic systems combined with V-trough cavities and N-S solar tracking c a n have their energy output significantly increased when compared to fixed fiat plate systems. If manufacturing costs of V-trough concentrators and tracking mechanisms are kept low, the cost-benefit relation can be rather favorable. Previous results, obtained with a small prototype of photovoltaic V-trough tracking generator, working in the local sunny climate, shows that the energy output almost doubles as compared to fixed fiat plate PV systems (we verified an average increase of 52%, due to concentration, and 30 % increase due to solar tracking) (Fraidenraieh and Krenzinger, 1997). Several works have been published analyzing the benefits of PV systems coupled to Vtrough cavities. The performance of the largest photovoltaic system with tracking and mirror boosters (Carrisa Plains, CA) is described by Berman and Mitchell, (1989). The potential of this technology for different climates has been analyzed by Nann, (1990) and the economic potential for European climate, for systems with passive tracking, has been studied by Klotz (1995). The objective of this work is to evaluate the performance of a PV system with tracking and concentration. Experimental results, represented as a function of diffuse to hemispherical radiation ratio, yield a set of empirical correlations used to simulate the system behavior along the year. We describe, at first, the design procedure of the V-trough cavity. Immediately, we present results for the collection of solar radiation at the cavity absorber and aperture as compared with hemispherical solar radiation. We analyze then the optical efficiency of the cavity, the module's temperature when working with and without concentration, the electric output of the PV generator and finally we present a simulation of the monthly and annual performance of the system.
Design criteria, satisfied by the concentrator cavities are: (a) Light distribution on module's surface is uniform Co) Heat at the absorber region is dissipated by passive means (c) Small deviations of tracker-sun alignment are allowed, still satisfying uniformity of illumination on absorber region. The concentration ratio (C) and vertex angle ( 9 which minimize the cost of energy are chosen as the best option in terms of cavity design. For the local climate (Recife, latitude 8,05 South, Kh=0.55) the geometric parameters of the cavity, optimized with that procedure, are C = 2.2 and 9 = 30 ~ The optical behavior of the cavity is properly described by the angular acceptance function, F(0i), which gives the fraction of light rays, incident on the aperture at angle 0 i , able to reach the absorber. The angle 0i denotes the projection of the incidence angle on the eavity's cross section. The function F(0 i ) , for a cavity with uniform illumination at the absorber, is illustrated in Fig. 1. At incident angles 10il <--tt the illumination of the absorber region, where the modules are located, is uniform (Fig. 1). The ~t angle is a design parameter, allowing for a deviation from normal incidence without altering illumination uniformity. It can be calculated as (Fraidenraich, 1998) 2
sin(u165
- sin(y)
tantt = C -21 C-1
(1) sin(u165
+ cos(y)
which, for C-2.2 and ~ / = 30 0, is equal to 3 0 or 12 minutes 2. THE PV V-trough SYSTEM
2.1 Design procedure and cavity parameters In a previous paper we describe an optimization procedure
of tracking tolerance. As can be observed in Fig. 1 the angular acceptance function at the origin is smaller than 1.0. That means that, even at normal incidence, a fraction of incident light is
ISES Solar World Congress 1999, Volume Ill
rejected. For
incidence
angles
[0il ___~t
the
angular
tan to = C +...~1tan C-1
acceptance function can be calculated as:
Ioil < B
F(Oi ) = 1 + 2cos(2~) C
2.2
10il_<..
the width of the "plateau" ( B ) the largest the fraction of light rejection (1 - F(0 i) ). The optical efficiency of the cavity is
[0il-<~ .
I=
o o
1.0 0.9
= Q
0.8 0.7 0.6
,i
q,,,
0
= "Q.
o
0.4
: I,.,. m m
0.3 0.2
: o) r ,<
0.0
r.)
~t
~0.5
q) U.
0.1 I
0
20
40
(3)
For C=2.2 and ~ =30 0, to is equal to 57 ~
(2)
yielding a value of 0.909 for C=2.2 and ~ = 3 0 ~ Light rejection is the price to pay for the benefit of uniform The largest illumination within the angular interval
limited then by the value of F(0i) at the interval
343
60
i~do~ mM~ ~ (dogma) Fig. 1. Angular acceptance function of the V-trough cavity with C=2.2 and W =30 ~
Experimental facility
The whole system, with four cavities and two PV modules each, is able to track the sun path around one or two axis. The reflective surface of the cavities are built with fiat commercial mirrors, 3 mm thick and 3 m long, supported by a plastic sheet for environmental protection combined with an aluminum structure for mechanical rigidity. The photovoltaic generator, made up of eight 50 Wp series connected modules (C-Si), supplies electric energy to an accumulator during the day and to the load at night. The V-trough generator is shown in the Fig. 2, bellow in left and it's principal physical characteristics in Tab. 1. Evaluation of the system's performance considers: 9 Net gain of collected solar radiation due to tracking, as compared to fixed systems; 9 Comparison of collected solar radiation by the Vtrough tracking PV generator with tracking and fixed systems; 9 Effective concentration ratio of the V-trough cavity; 9 Temperature increase of photovoltaic modules as a function of collected solar radiation level for tracking V-trough PV generators. Comparison with fixed systems; 9 Output of electric energy.
Tab. 1. Physical characteristics of the V-trough generator Parameters N ~ of cavities
04
Concentration ratio (C)
2.2
Vertex angle ( ~ )
30
Aperture area
0.77 m 2
Absorber area
0.43 m 2
N ~ o f modules (2/cavity) Type
08
Mono-crystalline (C-Si)
Power per module (W)
50 Wp
Electric connection
Series
Total occupation area
Finally the largest angle at which there is light acceptance ( to ) (Fig. 1) can be calculated as:
Characteristic
12.18 m 2
Total Weight
270kg
Length
3.00 m
Width
4.06 m
344
ISES Solar World Congress 1999, Volume III
3. EXPERIMENTAL RESULTS
3.1 Tracking gain and radiation flux at cavity absorber The figure below illustrates the time variation of collected solar radiation at the module's surface of the generator, at the concentrator aperture and the hemispherical and diffuse solar radiation on the horizontal plane. The experimental data represent the average of 10 minutes measurements. At this particular day, the net gain due to tracking was equal to 38 % and the collected solar radiation at the module's surface is 2.2 times the solar radiation collected by a fixed horizontal plane. Due to the local small latitudes, the horizontal plane will be considered as a fixed reference plane. It is worth noticing the low level of diffuse radiation ( H ~ for this particular day, equal to 1.41 kWh/m2 for a global radiation of 6.57 kWh/m2 (Hh), which gives an ~ h fraction equal to 0.215.
Since both, tracking and concentration devices mainly benefit fi'om beam radiation, it is expected that their performance will be strongly dependent on the ratio of diffuse to hemispherical radiation. Thus, as a background reference, we represent in Fig. 4 daily values of the (I-I~/I-Ih) ratio, registered during our experiment's period, against the clearness index (Kt). The gain due to tracking (HA/Hh), for a series of daily measurements, has been plotted in Fig. 5, where the symbol HA denotes the solar radiation collected on the cavity aperture. The tracking gain varies from values around 1.2, for cloudy days (H h =3.0kWh/m 2 ), up to 1.35 in very sunny days ( Hh =6.5kWh/m 2 ). Even there is a clear correlation between the tracking gain and solar radiation level, a more general relation, not dependent on the particular place where the tests have been made, can be obtained by plotting the tracking gain against the relation of diffuse to global radiation (HadHh), represented in Fig. 6. The variability of data, for a representative range of values of ~ h (from 0.16 up to 0.80), is considerably reduced when compared with the plot in Fig. 5.
Fig. 3. Optical flux on the concentrator absorber plane (1), collected solar radiation at concentrator aperture (2), global (3) and diffuse radiation on horizontal plane (4), during a good local day (Recife, 8.05 South latitude, Brasil).
Fig. 5. Daily collection gain due to tracking against horizontal solar radiation level.
Fig. 4. Ratio of diffuse to hemispherical radiation against daily cleanness index Fig. 6. Gain due to tracking Vs relation between diffuse and hemispherical solar radiation
ISES Solar World Congress 1999, Volume III
The relation of radiation flux at absorber and aperture is also plotted against H~/Hh (Fig.7). A maximum effective concentration of 1.60 is obtained for a day with I-I~/Hh equal to 0.16, decreasing steadily towards one as the ratio Hdh/Hh increases.
345
temperatures registered in three different positions in the modules.
Fig. 9. Temperature increase above ambient of photovoltaic modules as a function of radiation flux on absorber area.
Fig. 7. Effective concentration ratio against fraction of diffuse to global solar radiation.
Finally the ratio of the radiation flux on the surface of the photovoltaic module and hemispherical solar radiation, is shown in the next fig. 8.
The regression coefficient relating temperature increase and radiation flux is approximately equal to 0.019, figure similar to coefficients quoted in the literature, but for lower radiation levels. Note that radiation levels registered in Fig. 9, obtained as ten minutes averages, reach 1700 W/m z. For comparison we represent the temperature increase of a fixed system together with that of the tracking V-trough concentrator, Fig. 10, both against horizontal insolation. The coefficient of the module's temperature for the V-trough tracking system is now affected by the ratio of the radiation flux on the cavity's absorber region over the horizontal insolation (< I a / I h >) ( --- 1.65).
Fig. 8. Relation of flux radiation on absorber over hemispherical solar radiation compared with the fraction of diffuse to hemispheric solar radiation.
The radiation flux on the surface of the photovoltaic module reaches values 2.0 times the horizontal solar radiation level in very good days (Hd/Hh =0.26), decreasing to 1.35 for I-Id/Hh = 0.80. For a typical local day with I-ld/Hh equal to 0.40 the relation Ha/l-Ih is approximately equal to 1.85.
3.2 Temperature of the modules Temperature increase above ambient, measured each ten minutes, are plotted in Fig. 10 against radiation flux on the absorber area (Ia). Results shown are the average of
Fig. 10. Temperature increase above ambient of photovoltaic modules as a function of horizontal radiation flux, in tracking V-trough and fixed system
3.3 Effective concentration ratio The performance of the PV generator depends strongly on the optical quality of the V-trough cavity, associated to mirrors reflectivity, manufacturing precision and structural stability during operation of the cavity's geometry. Tracking precision is also important in terms of system performance. The relation between radiation flux at the absorber and collected solar radiation at the aperture (Ha/HA), called
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346
"effective concentration ratio", is a measure of the optical performance of the cavity and has been plotted, in Fig.11, against (HdA/HA). In the same figure, experimental results are compared with a theoretical estimation of the relationship between those parameters, obtained with a procedure described in the next section. 1.81.6
I , ~
99 9
9
Trendlineef
1.4 12
angle modifier is taken equal to 0.92. Thus the beam optical efficiency is equal to: Tlbca v =0.79 Taking into account reflection losses in the mirror walls, the interchange factor between aperture and absorber is equal to 0.43 (Fraidenraich, 1995). Writing (HbA) as (HA- Hd~, the combined optical efficiency for beam and diffuse radiation is equal to: rlcav = (0.79.(HA-HdA) + 0.43.HdA)/HA
~1.0
= 0.79 -- (0.79-0.43) HdA/HA
(6)
yielding for the effective concentration ratio (Ha/HA) the following expression
0.6
o.4
Ha/HA = C.0.79 - C.0.36 HdA/HA
(7)
0.2
When C=2.2 is substituted in Eq. (7), it becomes
0.0
i
0.OO
i
0.20
|
i
0.,40
i
I
0.60
i
i
0.80
Fig. 11. Effective concentration ratio vs diffuse to total collected solar radiation at cavity's aperture. Experimental tendency line and theory.
4. OPTICAL EFFICIENCY OF V-TROUGH CAVITY Based on the optical properties of V-trough cavities a theoretical estimation of the relationship between Ha/H A and HdA/H^ Can be made. If we call PI(0) the aperture fraction occupied by beam radiation reaching the aperture at normal incidence and making one reflection before reaching the absorber, it can be shown that for our cavity (C=2.2, ~4/=30~ it is equal to 0.455 (Fraidenraich, 1998). Similarly, the aperture fraction which corresponds to light rays reaching the aperture at normal incidence and the absorber without reflections, called Po(0), is also equal to 0.455. Nine percent of light rays are rejected at normal incidence. The fraction of diffuse radiation transferred from aperture to absorber is called exchange factor (EAa). Given the beam radiation at the aperture, the light flux reaching the absorber can be calculated as: Aa.Hba = AA.HbA.[Po(0) + PI(0). P .K( 0 a )]
(4)
or Hba = C. HbA. rlbcav
(5)
where (Aa) and (Hba) are absorber area and daily beam radiation at absorber region, (AA) and (I'IbA) are aperture area and daily beam radiation at aperture of the optical cavity and
(K(0a)) is the incidence angle modifier for light rays reaching the absorber at angle Oa . The reflectivity of a 3 mm thick back silvered mirror is equal to 0.80 and the incidence
Ha/HA = 1.74 - 0.79.HdA/HA
(8)
Equation (8) is represented in Fig. 11. It can be observed that it lies well within the set of experimental points with a slope slightly higher than the experimental tendency line. This is a first verification about the optical quality of the cavities under working conditions. A second one, including all the PV modules of the system can be obtained analyzing the electric output.
5. ELECTRICAL OUTPUT OF THE PV SYSTEM The electric energy generated by the photovoltaie system charges a set of batteries, the number of which was selected according to what can be considered the best match between generator and load. Obviously, this match depends on a number of factors, the solar resource between them, and how the most frequent solar radiation levels are related with the efficiency curve of the generator. In Fig. 12 we show the electric power and its correspondence with the radiation flux on the module's surface along a day. The synchronism between radiation flux on the absorber and electric power indicates that, within tracking tolerance, the system follows properly the sun position. The voltage output of the generator varies between 100 to 115 V, increasing as the accumulator charge increases. When the highest voltage is reached, a charge controller cuts the current off. Since the generator is formed by an arrangement of eight modules in series, the voltage per PV module varies between 12.5 and 14.4 V and the electric current at this voltage takes values close to the short circuit current of the module. Non linearities in the behavior of the PV generator arise when the batteries charge increases and the module's voltage approaches 14 V. Theoretical calculations show that the generator efficiency around this voltage and Ia in the interval between 1000 and 1600 W/m~-varies between 9.0 and 10.4 %. Experimental values are around 88-90 % of theoretical estimations (8.0 to 10.0 %). The 4 to 1 1 % difference can be arxx)unted for by mismatch losses, originated
ISES Solar World Congress 1999, Volume III
by dispersion of electrical characteristics of PV modules, series resistance of generator-load connection cable, optical non-uniformities and finally the model prediction accuracy. A detailed analysis shows that mismatch losses and differences between theory and experiment on the optical efficiency account for almost 7 % of the difference in generator efficiency. Since the performance of the PV generator depends closely on the overall optical quality of the concentrators, the deviation observed permits to conclude that the optics of the system lies within tolerance requirements (3 to 4 %).
4. Daily
average
values
347
of
Icoll
are calculated
as
(Hcoll/2 ~coll ), where 2 ~coll is the collection period; 5. The photovoltaic arrangement operates at the maximum power point; Results are shown in Figs. 13 and 14.
Fig. 13. Monthly and annual average of solar radiation collected by fixed, tracking and tracking with V-trough systems. Fig. 12. Radiation flux at the module's surface (1) and electric power generated by the system (2).
6. PERFORMANCE SIMULATION Using the experimental results ~already described we can simulate the system performance. Basic aspects of the procedure are: 1. 2.
3.
For fixed systems, the collected solar radiation is equal to the solar resource (average Hh); For tracking collectors we calculate HA from experimental results (trend line in Fig. 6) and similarly for collectors with V-trough concentrators and tracking (Fig. 8); From experimental results we determine the temperature coefficient ((X T ), relating temperature increase above ambient and collected solar radiation Tm - Tam b = o~T .Icoll
(12) SUMMARY AND CONCLUSIONS
Module efficiency is then obtained as rlPV = [1 - 15((ZT.Icoll + (Tamb - Tref))]
(13)
As shown, experimental results yield a coefficient a T =0.019. For the coefficient (15) a value equal to (0.005 ( ~
can be adopted (Siegel et al., 1981).
Measurements of ambient temperature in our laboratory suggest an average value of Tamb - Tre f ---5 ~ ;
Fig. 14. Monthly and annual average of eletric energy produce by fixed, tracking, tracking and tracking with V-trough systems.
9 The procedure developed in this work allows to evaluate the collection of solar radiation at the cavities' aperture and absorber regions, and the optical and electrical properties of the V-trough tracking PV system quite accurately. 9 Theoretical estimation of the optical efficiency of the cavity for beam radiation is equal to 79%. For global radiation it goes from 74 %, for a ratio of diffuse to hemispherical radiation of 0.15, to 57 % when that ratio
348
ISES Solar World Congress 1999, Volume III
increases up to 0.6. The theoretical prediction compares fairly well with experiments. 9 Differences between measured and estimated electrical output are, on the average, equal to 8 % and can be explained by electrical mismatch and optical losses. Since the performance of the PV generator depends closely on the overall optical quality of the concentrators, the difference observed permits to conclude that the optics of the system lies within tolerance requirements (3 to 4 %). 9 Simulation of the system's performance, based on the empirical correlations obtained from experiment, yield for the city of Recite an annual benefit equal to 26 % and 72 %, for tracking and tracking and concentration respectively.
REFERENCES
Berman, E. and Mitchell, K. W. (1989).Photovoltaic power plants: present and future. In Proceedings of the 4th
International Photovoltaic Science and Engineering Conference, Sidney, NSW Australia, The Institution of Radio and Electronics Engineers of Australia, Edgecliff, NSW, Australia 2027. Fraidenraich, N. (1998) Design procedure of V-trough cavities for photovoltaic systems. In Progress in photovoltaics: Research and Applications. Vol. 6, No. 1, pp. 43-54. Fraidenraich, N. e Krenzinger, A. (1997) Experimental performance of a PV prototype with N-S tracking coupled to V-trough cavities. In Proceeding of 14th European Photovoltaic Solar Energy Conference and Exhibition, Vol. l, pp. 348-35 l, Barcelona, Espanha. Fraidenraich, N. (1995) Analytic solutions for the optical and radiative properties of non-accepted light radiation of Vtrough concentrators. Applied Optics, Vol. 34, pp. 4800-4811. Klotz, F. H. (1995) PV systems with V-trough concentration and passive tracking. Concept and economic potential in Europe. In Proceedings of the 13th European Photovoltaic Solar Energy Conference and Exhibition, Nice, France. Klotz, F.H., Noviello, G. and Samo, A. (1995) PV V-trough systems with passive tracking: Technical potential for Mediterranean climate. In Proceedings of the 13th European Photovoltaic Solar Energy Conference and Exhibition, Nice, France. Nann, S. (1990) Potential for tracking photovoltaic systems. Solar Energy, Vol. 45 pp. 385=393. Siegel, M. D., Klein, S. A. and Beckman, W. A. (1981) A simplified method for estimating the monthly=average pefformarlce of photovoltaic systems. Solar Energy, Vol. 26, pp. 413-418.
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PERFORMANCE ANALYSIS OF A COMBINED PHOTOVOLTAIC/THERMAL (Pvrr) COLLECTOR WITH INTEGRATED CPC TROUGHS H.P. Garfl and R. S. Adhikari Centre for Energy Studies, Indian Institute of Technology, Hauz Khas, New Delhi - 110 016, India Tel.: +91-11-6861977, Fax: +91-11-6862037, E-mail: [email protected] Abstract - In the present investigation a theoretical analysis has been presented to study the performance of a hybrid PV/T collector coupled with a CPC. In the design several CPC troughs are combined in a single collector panel. The absorber of the hybrid PV/T collector under investigation consists of an array of solar cells for generation of electricity, while collector fluid circulating past the absorber provides useful thermal energy as in a conventional fiat plate collector. In the analysis, it is assumed that solar cell efficiency can be represented by a linear decreasing function of its temperature. Energy balance equations have been developed for the various components of the system. Based on the developed analysis both thermal and electrical performance of the system as a function of system design parameters are presented and discussed. Results have been presented to compare the performance of hybrid PV/T collector coupled with and without CPC. 1. INTRODUCTION Solar energy is a valuable form of energy which has the potential to meet a significant proportion of world' energy needs. The major applications of solar energy include solar thermal and solar photovoltaic systems. Solar collectors are designed to generate thermal energy, however, photovoltaic cell produces electricity directly from solar energy. During photovoltaic energy conversion, thenml energy is also generated which results increase in cell temperature. This is well known fact that the efficiency of PV cells drops as the temperature of cell increases. As a result, this heat generation within a photovoltaic module has traditionally been viewed as hindrance to efficiency and subsequently efforts have been made to transfer the thermal energy of PV module to the surrounding air. It has been shown that this heat can be harnessed by combining both the thermal and photovoltaic system in a hybrid Photovoltaic/Thermal (PV/T) energy system. The normal flat plate collector is converted in to a hybrid system by pasting solar cells directly over the absorber plate. In the literature, a number of theoretical as well as experimental studies (Floreschetez, 1979; Hendrie, 1980; Hendrie and Raghuraman, 1980; Cox etal., 1985 ; Chandra et al., 1983 ; Agarwal, 1989) have been made on the hybrid PV/T systems. Garg and his collaborators (Bhargava et al., 1991; Garg et al., 1991; Agarwal and Garg, 1994; Garg et al., 1994 and Garg and Agarwal, 1995) have carried out detailed analytical and experimental studies on hybrid PV/T air and liquid heating systems. Sopian et al. (1995) has presented a steady state model for the performance prediction of single and double pass PV/T air collector. A research project (1995) to explore and identify the PV hybrid concept has been carried out by IT Power and New Castle Photovoltaic Application Centre, commissioned by the Joint Research Centre at ISPRA. Recently, Bergene and Lovvik (1995) has proposed a detailed physical model of hybrid PV/T thermal system using water as a heat transfer fluid for the prediction of system performance. Studies on hybrid PV/T systems are being made at Indian Institute of Technology 0IT) in cooperation with All India Council of Technical Education (AICFE), New Delhi, India. Earlier studies made by present authors (Garg and Adhikari, 1997, 1998) include the development of steady state and transient simulation models to predict the performance of hybrid PV/T air heating collectors. These studies have shown that hybrid PV/T systems have great potential in terms of their system efficiencies.
It has been also observed that in a hybrid PV/T air heating system, a large area of solar cells is required, therefore, one must look for some avenues for enhancing the electrical output of the system keeping the cell area at a low level. It has been envisaged that a promising cost reduction technique could be its effective coupling with a solar concentrating device. As a result, the output of solar cell would be increased as more solar insolation would fall on the cell. Moreover, for a given electrical output, the system cost would be less, as optical concentrating devices are cheaper than the solar cells. In the past, several novel designs of optical concentrating device for solar energy collection have been developed in order to reach high collecting temperatures with high efficiency. The Compound Parabolic Concentrator (CPC) is the most promising high concentration device which produces elevated operating temperature and represent the optimum optical properties (Winston, 1974; Rabl, 1976). Recently, present authors (Garg and Adhikafi, 1998) have made optical design calculations of a CPC suitable for its coupling with a hybrid PV/T system Results have been presented for a full and tnmcated 3X CPC. In the present investigation a theoretical analysis has been presented for the modelling of thermal and electrical processes of a hybrid PV/T collector coupled with CPC. The system under investigation consist of several CPC troughs combined in a single PV/T collector panel. Based on detailed heat transfer analysis the energy balance equations have been developed for each component of the system. Thermal and electrical performance of the system as a function of system design parameters are presented and discussed. 2. THEORETICAL ANALYSIS The system under investigation along with associated thermal energy network have been shown in figure 1. The system consist of several troughs combined in a single collector panel, a metallic black absorber plate and a well insulated rear plate. In this configuration inlet air is passed through the passage between absorber and rear plate. The absorber plate is either a black paint or selective surface and solar cells are pasted over it. The adhesive used for pasting solar cells is characterized by high thermal conductive and good eleelric insulating material.
350
ISES Solar World Congress 1999, Volume III
0p_f = hp_f (Tp- Tf)
(10)
0b_f = hb_f (Tb- Tf)
(11)
(~b_f = hb_f (Tb- Tf)
(12)
Based on the energy transfer mechanism, the steady stateenergy balance equations for cover, solar cell,absorber plate,back plate and working fluidcan be written respectivelyas :
Fig.1. Hybrid PV/T collector coupled with CPC along with associated thermal energy network. In order to sin~lify the theoretical analysis, following assumptions have been made: 1. Steady state energy transfer has been achieved. 2. The temperatures of different system components v i z . transparent cover, solar cell, absorber plate, rear plate vary along the direction of working fluid flow only. Heat capacity of transparent cover is very small and, therefore, neglected. There is no leakage of air from the system. Side losses from the system are negligible.
2.1 SolarEnergy Interception Part of the incident solar radiation absorbed by transparent cover, solar cells and absorber plate are given by (Hseich, 1981) Qi,g = CR. Itot ttg (1 + Z'gpp PR 2n)
(1)
(13)
0I,s - 0s_p "0s_g "(~e : 0
(14)
0I,p + 0s-p" (~p-g - 0p-f "0p-b = 0
(15)
(~p_b- Ob.f- (~b_a=0
(16)
rh Cf dTf B
dx
-
-~6p-f""~()b--f=
0
(17)
where
(~e = r/s(~l,s
(18)
3. ANALYTICAL SOLWrlON Solving eqs. (13-16) for temperature of the transparent cover, solar cell, absorber plate and rear plate and substituting theses values in eq.(17), one obtains following linear differential equation: ~~,x.......A,dTf + P Tf (x) = q dx
(19)
Now solving this equation for fluid temperature with the help of initial boundary condition,
where Itat = IB + ID
(1)
(~i,p = CR. Iuvg p~ gg~p (1 +(pppgPR2n)/CR)(1-PF)
(3)
{~I,s= CR. Iu Vgp [ g as [ 1 + (Ps Pg PR 2n)/CR] PF
(4)
Tf (x) =Ti
at x = 0
one obtains,
Tf (x) = q + (Ti - q ) e"px P P
(20)
The fluid temperature can be averaged over collector length and can be calculated as
where Iu = (IB + ID)/CR
(~I,g + (~s_g+ 0p.g" (~g,a=0
(5)
2.2 Energy Balances Various energy fluxes across the system are given as follow:.
1L Tf =-~- 0 Tf (x)dx
m
(21)
(~g-a = hg-a (Tg- Ta)
(6)
one obtains following expression for average fluid temperature
(~s.g = hs.g (Ts- Tg)
(7)
(~p.g = hp-g (Tp- Tg)
(8)
-1 Tf = q + "-7"(Ti - q) (1- e"pL ) p pL p
0s_p = hs.p (Ts- Tp)
(9)
applying boundary condition that
(22)
ISES Solar World Congress 1999, Volume Ill
Tf(x)=To
Table 1. Thermo-physical parameters
at x = L
outlet temperature can be calculated as follows: To = q + ( T i - q ) e -pL P P
(23)
Parameter
Value
Parameter
Value
% %
0.05 0.90 0.90 0.85 0.90(Paint) 0.15(Selective) 0.15 0.05 0.85 0.90
Ti Tr fir g Cf
25.0 C 25.0 C 10% 0.0050 m 1008.0 J/Kg C
m PF IB ID
100 Kg/m2 h 0.50 800 W/m2 100 W/m2
oq
3.1 Heat Transfer Coefficients The forced convective heat transfer coefficient hp.f and hb.f were calculated using the relation given by Tan and Charters (1969) which includes the effects of the entrance and exit length in the air flow passage. The convective and radiative heat transfer coefficients from absorber and solar cell to transparent cover 1 ~ and l~g have been calculated using the relations developed by Rabl (1978). The convective and radiative heat transfer coefficients from wansparent cover to ambient lag_~; the conductive and radiative heat transfer coefficients hp.b, hb-a were calculated using the relations given by Duffle and Beckman (1991). 3.2 Performance Parameters Various performance parameters of the system can be calculated as follows:
eg ep 0g xg
T h e r m a l E f f i o l e n c y (q~) 70 j ~-I00 Kg/m ~ h PF-0.80
60 ~ 9
"o
" ~
I 0.01
0
I 0.02
"
I O.0a
I 0.04
(Ti - Te)/Ito t ( ~
(24)
"
"
.
I 0.06 me/W)
I 0.0"6
-
I 0.07 '
.
.i
i ;;::'"
.
.4~!08
'.f
Fig. 2. Thermal efficiency curves for a hybrid PV/T collector coupled with a CPC (selective coating).
Solar cell efficiency (Florschuetz, 1975)
r/s = r/r [ 1- fir (Ts- Tr)]
8elective Absorber.
. . 9
40
2
Ih Cf (To - Ti) CR. Itot
....
30
Thermal efficiency
r/t =
351
(25)
T h e r m a l E ff i c i e nc y (%)
70, dr-100 Kg/m = h PF-O.50
where
Pc- 2.0 m
60
Non-eeleotive Absorber
i
50
(Ts -Tr)
40
Where T*s is the solar cell temperature at which its efficiency drops to zero.
3O
20 0
Electrical efficiency
r/e
Qe CR. Itot
(26)
4. RESULTS AND DISCUSSION For the quantitative appreciation of simulation model, calculations have been made for PV/T air heating systems coupled with and without CPC. An algorithm has been developed to predict the working fluid (air) temperature as a function of absorber length. System performance has been discussed corresponding to different design and operational parameters. Various thermo-physical parameter and weather data used in the calculation have been summarised in Table 1. The optical design parameters of a tnmcated 3X CPC have been taken from Garg and Adhikari (1998).
0.01
0.02
o.oa 0.04 0.05 (TI-Ta)/Ito t ( ~ mZ/W)
0.06
0.07
0~08 ", ...
:
Fig. 3. Thermal efficiency curves for a hybrid PV/T collector coupled with a CPC (non-selective coating). Fig.2 and 3 show the thermal efficiency curves for PV/T collector with and without CPC configurations for a fixed duct depth, collector length, mass flow rate and packing fraction (Area covered by solar cells) corresponding to the absorber with and without selective coating. It has been observed that for both nonselective and selective absorbers, initially the performance of the system without CPC is better, however, after a certain point the system with CPC performs better. It implies that the integration of a CPC with PV/T system is appropriate for the application in the higher temperature range. For both the configurations, with and without CPC, the system
ISES Solar World Congress 1999, Volume III
352
performs better in the case of selective absorber. This is due to reduction in radiative losses from absorber to glass cover. The results of the parametric studies have been presented in the subsequent figures. Fig.4 shows the variation of thermal and electrical output as a function of air-mass flow rate corresponding to the system configurations with and without CPC. Thermal
Energy,
Electrical
W/m a
Energy,
W/m =
lOO.
2000 .
1800
NOMENCLATURE
PF-0.60
_ AO-2.0 m 2
1600
0R-2.88
80
1400 60
1200 1000
. 4 o :.
800 CRII.0:
600
...
400
,o
200
- . , . , -.i'-
9.
0 i _ 0
It has also been observed that the system coupled with CPC always performs better in terms of both the thermal and electric output. It is to be noted that as material cost increases with the integration of more number of cells and the CPC in the system, the final decision in this respect must be based on the costeffectiveness of the system by minimising life cycle costing (LCC) of the systen~
i
I
100
200 Alr mass Thermal
i
l
i
300 400 flow rate..Kgl m=h
Energy
----" Electrical
9'
800 . Energy -
6 0 0..... .~(
:'~
Fig. 4. Variation of system performance output as a function of air mass-flow rate. It has been seen that increasing mass flow rate increase the thermal and electrical output. It is also quite evident that the performance output of the system with CPC is quite higher than without CPC. The effect of packing fraction of solar cells on the performance of the system has been shown in the fig. 4. This indicates that increasing area covered by solar cells increase electrical output of the system quite rapidly. However, the thermal output remains more or less same. The system coupled with CPC shows better performance in terms of both the thermal and electrical output. Thermal Energy, W/m ..... m- m o Ir41/nt2 h
~'
Electrical
Energy,
A b B C CR g h I L m n p,q PF Q T U
area, m 2 duct depth, m duct width, m specific heat, J/Kg ~ concentration ratio absorber-reflector gap width heat transfer coefficient, W/m2 ~ solar irradiance, W/m2 collector length, m mass flow rate Kg/h m2 average number of reflection for radiation passing through CPC inside the acceptance angle coefficients of linear differential equations for fluid temperature Packing fraction, area covered by solar cells, % energy flux, W / m 2 temperature, ~ heat loss coefficient, W / m 2 ~
Greek letters a absorptivity TI efficiency e emissivity p reflectivity
W/m =
Fig. 5. Effect of packing fraction (area covered by solar cells) on performance output.
Subscripts a ambient B beam radiation b rear plate D diffuse radiation e electrical f working fluid (air) g transparent cover I solar irradiance i inlet o outlet p absorber plate R reflector r reference s solar cell tot total
5. CONCLUSIONS
ACKNOWLEDGEMENTS
The developed steady state model predicts the thermal and electric performance of a hybrid PV/T air heating collector coupled with a CPC. Results show that the coupling of CPC with a PV/T air heating collector depends on the temperature range for which system is designed. Parametric studies show that the thermal and electric output of a PV/T system increases with increase in collector length, air mass flow rate and packing fraction (cell density), and decreases with increase in duct depth.
Authors duly acknowledge the financial support provided by All India Council of Technical education, New Delhi to carry out the present investigation.
2ooo[
1.oof
~
-100
2.0 m2
1600
1200 _
80-
~
~
60
looo 8 0 0 ~"
40..
6oo t
...~/
i .
"200
o 0.0
' 0.2
, 0.4
Area covered -'--- Thermal
, 0.6 by solar
Energy
-"-'-
o:L.
9 0.8
1.0~.~"'
cells, Fraction Electrical
Energy
~;ii 9
. ~i:; "~.L
ISES Solar World Congress 1999, Volume III
353
REFERENCES
Garg, H.P. and Agarwal, ILK. (1995). Some aspects of a PV/T collector: forced cerculation flat plate solar water heater with solar cells. Energy Convers. Mgmt. 36 (2), 87-99.
Agarwal, ILK.(1989). Studies on Solar Photovoltaic-Thermal (PV/T) Systems, Ph.D. Thesis, Indian Institute of Technology, New Delhi, India.
Hendrie, S.D. (1980). Evaluation of combined photovoltaic / thermal collectors. In Proceedings of lSES International Congress and Silver Jubilee, 28 May - 1 June 1980, Atlanta, GA.
Agarwal, ILK. and Garg, H.P. (1994). Study of a photovoltaicthemaal system-thermosyphonic solar water heater combined with solar cells. Energy Convers. Mgmt. 35, 605-620
Hendrie, S.D. and Raghuraman, P. (1980). A comparison of theory and experiment photovoltaic/thermal performance. In
Proceedings of 14th IEEE Photovoltaic Specialists Conference, January 1980, San Diego, CA.
Beregene, T. and Lovvik, O.M. (1995). Model calculation on a flat-plate solar heat collector with integrated solar cells. Solar Energy 55(6), 453-462.
Energy, 27, 19.
Bhargava, A.K., Garg, H.P. and Agarwal, ILK. (1991). Study of a hybrid solar system-solar air heater combined with solar cells. Energy Convers. Mgmt. 31,471-479.
Sopian, K., Yigit, K.S., Liu, H.T., Kakac, S. and Veziroglu, T.N. (1995). Performance analysis of photovoltaic thenml air heaters. Energy Convers. Mgmt. 37, 1657-1670.
Chandra, R., Goel, V.K. and Ray Choudhuri, B.C. (1983). Thermal performance of a two-pass PV/T air collector', In Proceedings of Solar Energy Society of India (SESI), Dec. 15-18 December, 63-69, Baroda, India.
Tan, H.M. and Charters, W.W.S. (1969). Effect of themml entrance region on turbulent forced convective heat transfer for an asymmetrically heated rectangular duct with uniform heat flux. Solar Energy 12, 513-516.
Cox, C.H. and Raghuraman, P. (1985). Design considerations for flat-plate photovoltaic/thennal collectors. SolarEnergy 35, 227.
IT Power Ltd. (1995). Hybrid Photovoltaic/Thermal Concepts. Final Report Produced for EC Joint Research Centre, ISPRA, June 1995.
Duffle, J.A. and Beckman, W.A. (1991). Solar Engineering of Thermal Processes, 2~'t edn, Wiley InterScience, New York.
I-Isieh, C.K. (1981). Thermal analysis of CPC collectors. Solar
Rabl, A. (1976). Comparison of solar concentrators.
Solar
Energy 18, 93. Florschuetz, L.W. (1975). On heat rejection from terrestrial solar cell arrays with sunlight concentration. In Proceedings of IEEE Photovoltaic Specialists Conference Records, May 1975, 318326. Florschuetz, L.H. (1979). Extension of the Hottel-Whiller Bliss model to the analysis of combined photovoltaic/thermal flat plate
collectors. In Proceedings of lCSE Solar energy Society Meeting, 15-20 August, Winnipeg, Canada. Garg, H.P. and Adhikari, R.S. (1997). Conventional hybrid photovoltaic/thennal (PV/T) air heating collectors : steady state simulation. Renewable Energy 11,363. Garg, H.P and Adhikari, ILS. (1998). Transient simulation of conventional hybrid photovoltaic/thermal (PV/T) air heating collectors. Int. J. Energy Research 22, 547-562. Garg, H.P. and Adhikari, ILS. (1998). Optical design calculations for CPC's. Energy 23(10), 907-909. Garg, H.P., Agarwal, ILK. and Bhargava, A.K. (1991). The effect of plane booster reflectors on the performance of a solar air heater with solar cells suitable for a solar dryer. Energy Convers. Mgmt. 32, 543-554. Garg, H.P., Agarwal, ILK. and Joshi, J.C. (1994). Experimental study on a hybrid solar photovoltaic-themml solar water heater and its performance predictions. Energy Convers. Mgmt. 35, 621633.
Winston, IL (1974). Principles of solar concentrators of a novel design. Solar Energy 16, 89.
ISES Solar World Congress 1999, Volume III
354
AN ASTIGMATIC CORRECTED TARGET-ALIGNED SOLAR CONCENTRATOR FOR SOLID STATE LASER PUMPING Mordechai Lando, Jacob Kagan, Boris Linyekin, Ludmila Sverdalov and Grigory Pecheny Laser Division, Rotem Industries. P.O.B. 9046, Beer-Sheva 84190, Israel Tel. (972)7657-1460, Fax (972)7655-5984 e-mail: [email protected]
Yaakov Achiam NRCN, P.O.B. 9001, Beer-Sheva 84190, Israel, Tel. (972)7656-7798, Fax (972)7656-8770
Abstract-Solar pumped lasers are candidates for wireless power transmission in space, free space optical communication and terrestrial photochemistry. Solar pumping of lasers requires a highly concentrated solar energy. This may be obtained by a combination of a primary concentrator with f/D > 2 and an additional non-imaging concentrator. As the primary concentrator we have designed and constructed a novel tower concentrator. A 3.4m diameter primary mirror was mounted on a commercial two-axis positioner. Unlike the common zenith mounting, the positioner fixed axis is directed southwards, 32~ above the horizon. With this novel mounting, the concentrator is the first implementation of the Astigmatic Corrected Target Aligned (ACTA) design, which flattens the irradiation density variations during the day. 61 primary mirror segments are each mounted on a separate two-axis mount, and aligned to compensate for astigmatism. The segments are spherically curved with 17m radius of curvature, while their vertexes are placed on an 8.5 m radius spherical cap. A four-segment plane mirror reflects the light towards a horizontal focal plane. We have measured the focal spot power distribution during a full day and found good agreement with optical design calculations. Also, absorbed solar power into a rectangular 8.9x9.1 cm2 aperture were within :k30% of the average during four hour period near solar noon. Peak solar concentration in the focal plane exceeded 400 suns.
1. INTRODUCTION Solar pumped lasers may find applications at a variety of space technologies as well as on earth (Brauch et al 1991, Duchet 1992, Hall 1992). Being virtually the only energy source in space, solar energy may be used to pump solid state lasers either directly or indirectly. In indirect pumping, solar cells are used to generate electric energy, and diode lasers would convert the electric power into a pumping light. Direct solid state laser pumping with solar light is inherently more efficient, much simpler, and more reliable. Existing solar energy facilities are too large for efficiency demonstration of solar pumped lasers, and their operation is too expensive for durability studies. Therefore, for the purpose of solar pumped laser development we have designed and constructed a special solar concentrator. To pump solid state lasers, high solar concentrations are needed (Thompson et al, 1993). The desired concentration may be obtained by combining primary concentrator with f/D>2 with a non-imaging optics concentrator (Gleckman 1988). For the primary concentrator, we chose a tower configuration, with a folding mirror, which directs the solar light towards a fixed horizontal plane. Such a tower configuration is more convenient for conducting experiments than a dish or a cassagrainian configuration, in which the focal plane moves with the solar orbit. In a tower configuration, the deflection of the solar light from its incoming direction causes astignmtic aberration and focal spot increase. With a primary mirror segmentation, the focal spot may decrease to a minimum at certain hours and for a certain season by segment alignment (canting, hereafter). Usually, this is done for the solar noon, and the focal spot area
increases in the afternoon and in the morning. Several realignments are required over the year to compensate for sun seasonal inclination variation. The focal spot increase is reduced by a novel design, Astigmatic Corrected Target Aligned (ACTA) heliostat (Ries and Schubnell 1990, Zaibel et al 1995). We adopted the ACTA concept for our concentrator, and demonstrated its first implementation. The next section compares conventional mounting with an ACTA configuration, and presents the predicted solar concentration for both configurations. The concentrator construction and the performance are depicted in the third section. 2. THEORY We start with a qualitative explanation for the difference between conventional and ACTA mounting. Then, we give the design parameters used in the optical analysis, and present its results. In a conventional tower configuration, the fixed azimuth axis is vertical, and the elevation axis is horizontal, as shown in Fig. l a. Fig. lb shows an ACTA tower configuration, in which the fixed axis is tilted southwards to the target folding mirror and the second "elevation" axis is normal to the fixed axis. To discuss both configurations, we point out that in tower configuration incidence plane is defined by the centers of the primary mirror, the sun and the target. Incidence angle is the angular deviation of the target direction from the sun direction. For both configurations, the incidence plane is normal to the ground at canting time of solar noon. To simplify the discussion we designate the central highest, lowest, leflmost and rightmost segment mirrors at canting time as Up, Down, Left
ISES Solar World Congress 1999, Volume III
and Right segments. The canting process is symmetrical relative to a symmetry line which connects the centers of the Up and Down segments. Consequently, the Left and Right segments, for example, are aligned towards the central segment with the same angles, but from opposite sides.
355
canting continues to be effective, and astigmatic correction deteriorates only due to incidence angle slow variation. Before presenting the optical analysis results, we describe the primary mirror in more detail. The 3.4 m diameter primary mirror is mounted on a two-axis positioner. The positioner fixed axis is directed southwards, 32~ above the horizon. The primary mirror is composed of 61 hexagonal segments, 360 mm between sides, each mounted on a separate two-axis base. The segments are spherically curved at 17 m radius of curvature, while their vertexes are placed on an 8.5 m radius spherical cap, in similarity to former solar furnaces (Diver et al 1983, Thompson et al 1992). The optical design calculation were done for a semi-annual canting regime, with a winter canting and a summer one. The winter canting is for 22~ incidence angle, compatible with solar noon on November 7 and on February 7. The summer canting is for 42~ incidence angle, optimal for the solar noon on May 7 and on August 7. The focal point distribution for several hours during the day, and for various days during the year was calculated by OMEK OPTICS using in house ray tracing code. Fig. 2 compares calculated focal distribution with experimental ones, which will be discussed in the third section. The summer canting calculations are for 2 hours before solar noon (a), solar noon (b) and 3 hours after solar noon (c), all during June 23.
Fig. 1 Tower configurations: Conventional versus Astigmatic Corrected Target Aligned mounting
When the sun moves from its southward direction, the canted primary mirror revolves around the two axes to follow solar orbit. However, the exact mode of revolution is different for each configuration. In the conventional configuration, the Up, Down, Left and Right segments stay always in their respective highest, lowest, leftmost and rightmost positions, while the incidence plane stops to be normal to the ground. Consequently, the incidence plane is not anymore situated symmetrical to the primary mirror, the astigmatic correction deteriorates, and the focal spot size increases. In ACTA configuration, the incidence plane contains the target oriented fixed axis. To follow solar orbit, the primary mirror revolves around the fixed axis, keeping the symmetry line on the incidence plane, while it rotates around the "elevation" axis to compensate for the slowly varying incidence angle. Thus, the
Fig. 2 Calculated versus measured light distribution at focal plane at June 23. a. 2 hours before noon. b. solar noon. c. 3 hours after solar noon. Calculation square frame sides are 200, 200, 300mrn, respectively. Fig. 3 shows the hourly variation of the focal spot power density for May end, demonstrating that power density is within :k25% of the average power density during four hours around solar noon. For comparison, we also present calculated focal spot power density for conventional tower configuration of an otherwise identical design, clearly manifesting the advantage of ACTA over the conventional mounting.
356
ISES Solar World Congress 1999, Volume III
Fig. 3 Peak solar density at focal spot during the day for ACTA (solid line) and conventional (dashed line) mountings
3. CONSTRUCTION AND PERFORMANCE The primary mirror was mounted on a commercial two-axis ORBIT Advanced Technologies AI~035-1SL antenna positioner. Serving as a basis for the positioner, a steal construction beam was fixed to a concrete base, which was forged to a concrete roof of a three stage building. For southward orientation, we looked at Polaris with a theodolite, and used time dependent astronomic calculation to correct for Polaris north deviation at Rotem Industrial Park.
Fig, 4 The 61 segments primary mirror with individual bases on hexagonal steal frame and 9 parallel beams The primary segments are individually mounted on a two-axis base. The base mounts are connected to a hexagonal steal frame through 9 parallel beams as shown in Fig. 4. A large steal
construction contains the experiment booth and serves as a rigid base for the folding mirror and for a 150X112 cm2 optical table located inside the booth. The construction is rigidly fixed to four concrete cubes, which are forged to the concrete roof as well. The folding mirror frame can rotate around an east-west and around south-north axis. Four glass-on-aluminum plane mirrors serve as reflecting surfaces. Mechanical attachment enables mutual alignment of the four segments to constitute a single plane. The solar light enters the experiment booth through a special hole in its roof, and the focal spot plane is the entrance plane for the collecting device, which is a Compound Parabolic Concentrator (CPC) for solid state laser pumping. Before installing the folding mirror, we have mounted a Doraboard 1200 target on the direct focal plane, and used removable plastic covers on the segments to expose the target to a separate segment at each time. This was used initially for canting, and later to evaluate the hourly focal spot variation. To that end, we have exposed five segment mirrors; the central mirror and the aforementioned extreme segments. The resulted focal spot envelope on the Doraboard target was video recorded, and a SPIRICON beam analyzer was used to evaluate the intensity distributions. Typical results compare well with the calculation results of the former section, as shown in Fig. 2. To follow the solar orbit, we used the ORBIT option of the positioner. The solar orbit in the zenith-north-east coordinate system is calculated using an astronomical software, US Naval Observatory MICA1.5 (1998). An OMEK OPTICS code (ZAVIT1-2.0) bisects the incidence angle to account for reflection, and transforms the bisector coordinates to a southwards tilted coordinate system. The calculating PC computer sends a weekly table of coordinates to the positioner controller for off-line tracking. A 2 kVA Gamatronic Uninterrupted Power Supply (UPS) drives the positioner to ensure fast removal of the high intensity focal spot under electrical grid failure. To monitor the ACTA concentrator operation, we have installed three auxiliary systems. The first one is a small meteorological station equipped with a Young 06201 wind tracker, Exetech temperature and humidity digital meter, and EPLAB NIP pyrheliometer on ST-1 equatorial mount. The second system is a vision system with which the worker inside the booth keeps control of the positioner motion. One of its two video monitors shows the scene of Fig. lb, while the other one views the focal spot plane. The third auxiliary system is a cooling system, which removes excess heat from various part of the CPC and laser head, and monitors temperatures and cooling liquid flow rates. National Instruments DAQ AT-MIO-64E card and LabVIEW 5.01 software are used for data acquisition, data presentation and calorimetric calculations. Detailed description of the cooling system is due to appear in a forthcoming report on the Rotem solar pumped laser. With the cooling system, we could measure the absorbed solar power into the 8.9X9.1 cm2 CPC aperture. In Fig. 5 we show the measured absorbed solar power on May 19 as a function of the solar hour. The power is shown to be within :~.30% of the average from 10:00 to 14:00, in good agreement with the calculation of Fig. 3. It should be noted that solar insolation
ISES Solar World Congress 1999, Volume III
variation was not taken into account in the calculation, while the insolation manifested a +4% variation between 10:00 and 14:00 (solar time), which affected the absorbed power to the same extent.
357
Gleckman P., Achievement of ultrahigh solar concentration with potential for efficient laser pumping, Applied Optics,
27,4385-4391, 1988. Hall R. B., Lasers in industrial chemical synthesis,
Laser
Focus, pp. 57-62, September 1992. US Naval Observatory Multiyear Interactive Computer Almanac 1990-2005, Willman-Bell, Richmond 1998. Thompson G. A., Krupkin V., Oron M., Yogev A., High power solar pumped solid state lasers, in CLEO, (OSA), 11,
Q 500 a..~ 400 300
590-592, 1993. Thompson G. A., Krupkin V., Yogev A., Oron M., Solar Pumped Nd:Cr:GSGG parallel array laser , Optical
200
0 ~ .o <
100 0
Engineering 31, 2644-2646, 1992.
9
10
11
12
13
14
15
Solar time (hours)
Fig 5 Measured absorbed solar power into the CPC aperture during May 19 4. CONCLUSIONS A tower configuration is advantageous over a dish configuration for the purpose of research and development of a high concentration device due to the easier access to the fixed target. Its main drawback is that focal spot size varies during the day and frequent canting is needed to overcome seasonal inclination variation. We have successfully implemented an ACTA tower configuration, and demonstrated that it flattens the daily irradiation intensity variations. This achievement renders the ACTA configuration an attractive option for future high concentration solar energy collector. 5. ACKNOWLEDGMENT We would like to thank Prof. A. Yogev, Dr. V. Krupkin, Mr. D. Sagie and Prof. J.M. Gordon for fruitful discussions, and Dr. Z. Burshtein for critical reading of the manuscript. 5. REFERENCES
Brauch U., Muckenschnabel, J. Opower H., Wittner W., Solar pumped solid state lasers for space to space power transmission,
Space Power 10, 285-294,1991. Diver R. B., Carlson D.E.E., Macdonald F.J., Fletcher E.A., A new high-temperature solar research furnace, jr.
Solar Energy Engin. 105, 288-293, 1983. Duchet M., Cabaret L., Laurens A., de Miscault J. C., Space power supply networks using laser beams", Space Power
11, 241-250,1992.
Ries H. and Schubnell M., The optics of a two-stage solar furnace, Sol. Energy Mater. 21, 213-217,1990. Zaibel R., Dagan E., Kami J., Ries H., An astigmatic corrected target-aligned heliostat for high concentration, Solar Energy Materials and Solar
Cells 3 7, 191-202, 1995.
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NONIMAGING FRESNEL LENS CONCENTRATORS FOR PHOTOVOLTAIC APPLICATIONS Ralf Leutz *?, Akio Suzuki'*, Atsushi Akisawa* and Takao Kashiwagi* *Tokyo Uniw.'rsity of Agriculture and Technology, Department of Mechanical Systems Engineering 2-24-16 Naka-cho, Koganei-shi, Tokyo 184-8588, Japan t email ralf~star.cad.mech.tuat.ac.jp, phone/fax +81-42-388-7076 ** UN ESC'O, Bouvin 3.26 SC/EST, 1, rue Miollis, 75732 Paris (~edex 15 A b s t r a c t - This study aims at clarifying the role of color aberrations in a novel nonimaging Fresnel lens of moderate concentration, intended for photovoltaic applications where homogeneous illumination is imperative. Refraction does indeed lead to color aberrations, but these are eliminated by rays of different color mixing on the absorber due to the nonimaging nature of the lens. Additionally, the manufacturing of the lens prototype, its working principle, and preliminary tests under the sun and under the moon are explained. An introduction deals with the metaphorical separation of photovoltaic concefitration by means of lenses on the one hand, and solar thermal concentration with mirrors on the other.
1
PHOTOVOLTAIC OR CONCENTRATION?
SOLAR
THERMAL
When we had completed the design simulation of a nonimaging Fresnel lens solar concentrator, we thought of it a.~ being a direct competitor to the Compound Parabolic (~onccntrator (CPC). We still think this to be true, but the title of this paper states that this paper is inw:stigating the role of our Fresnel h:ns in photovoltaic applications. Two questions arise: First, why dues 'refractive lens' sound like 'pv', and 'reflective mirror' like 'solar thermal'" And second, assuming the distinction to be merely historical, why are we considering and testing our nonimaging lens for use in photovoltaic conversion of sunlight? Boes and Luque (1992) try to illuminate why lenses have been used almost exclusively in photovoltaics, and mirrors in solar thermal systems. They point out that Fresnel lenses are offering more flexibility in optical design, thus allowing for uniform flux on the absorber, which is one of the conditions for efficiency in photovoltaic cells, l"urthermore, flat. Fresnel lenses are said to be less prone to manufacturing errors, since the errors at front and back faces of the prism are indeed partially self-correcting, while an angular error in the mirror's slope leads to twice this error in the reflected beam. This is true for flat Fresnel lenses, where the front faces of the prisms blend into a horizontal surface, a , d also for shaped lenses, in particular nonimaging lenses. On the other hand, imaging Fresnel lenses are still w:ry prone to movements of the focal poi,t due to nonparaxial incidence, especially when compared to ,o,imaging mirrors, which have been availabh: longer than nonimaging lenses. Ideal nonimaging concentrators are offering uniform radiation on flat absorbers, the main characteristic
of 'ideal' being the condition that the first aperture of the concentrator be filled out completely by uniform radiation, or radiation from a Lambertian source. Only then the second aperture (the absorber) will receive uniform flux. The sun itself may qualify as Lambertian approximation, although its brightness is not uniform, and wavelength dependent brightness changes significantly from its center to its outer areas. Since nonimaging concentrators are designed according to one or two pairs of acceptance half angles, the concentrator accepts light other than the almost paraxial rays of the s,n (acceptance half angle 0 = 0.27~ af,d concentrated flux is not uniform. Secondary concentrators can be used to make the flux on the a b ~ r b e r more uniform, but the price to be paid usually is rejection of at lea.st som(: rays. Imaging Fresnel lenses may be designed a.sphericaily, and with corrections in each prism for uniform flux, but both focal forshortening a[,d longitudinal focal movements require high precision tracking. Prisms split white light into its coh)r components. Refraction indices are wavelength depeltdent, and true uniform flux will remain an illusion, although we will see that our nonimaging Fresnel lens mixes colors at the absorber. A complete discussion of these topics can be found in Grilikhes (1997), Boes and Luque (1992), and Welford and Winston (1989). Although the authors' approach comes from different directions according to the field the are most familiar with, no clear technical link between 'lens' and 'pv' or "mirror' and 'thermal' could be established. Historical aspects are apt to throw more light onto these metaphorical connections. The ability to concentrate has been known for both lenses and mirrors for millenia, l'resnel h:nses made of glass have been used soon after their practical discovery by Jean Augustin l:resnel in 1748 as
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collimators in lighthouses. The reason for their success was that they were considerably lighter than singlets and absorbed less radiation than their oil covered mirror predecessors. Even today, lighthouse lenses are manufactured from glass to withstand the high temperatures present. Parabolic mirrors, on the other hand, had been used for large scale solar thermal applications since the beginning of the 20th century: in 1913 a 35 KW ...... h collector field consisting of 1,233 m 2 parabolic troughs was installed in Eqypt for irrigation, before World War I destroyed further efforts in solar thermal power generation ((;ra.,~se r al., 1991). The first idea for the collection of solar energy for industry and recreation comes from Leonardo da Vinci. who in 1515 proposed a parabolic mirror four miles across. In 1866, Mouchot designed and ran a solar powered steam engine (see Kaneff, 1996). Solar air and water heating for housing application was tested and documented for four houses in the United States by the end of World War II {1,6f 1992); and thermodynamic properties of solar energy collection were well researched. The photovoltaic effect was discovered already in 1839 by Edmond Becquerel (see Green, 1992). More important, howew:r, was the coincidence that the world saw both the invention of practical Fresnel lenses due to the availability of acylic plastic and the development of efficient silicon solar cells for applications in space in the early 1950's. Polymethylmetacrylate ( P M M A ) i s a lightweight, clear, and stable polymer with optical characteristics close to those of glass, and superior utilizabilily for the manufacturing of Fresnel lenses. The properties of P MMA were reported during World War II by Johnson (see Oshida, 1961). Since the late 1940's experimental Fresnel lenses were build mainly for optical applications for electrotechnology, such as sensors (Miller ct aL, 1951). From fine beginning, Fresnel lenses and photovoltaics were the domain of companies and large research institutions. The link between both fields may have been electrotechnology where experiences in pv and optical sensors are overlapping. Confusingly, the nonimaging concentrator CPC' had been invented in 1965 for the reflection of (.'erenkov radiation onto a sensor, and it took more than a decade for it. to become the metaphor for solar thermal energy collection. The CPC found other applications in astronomy (in combination with a lens, llildebrand, 1983J, and laser t(:(:lnnology. The advantages of nonimaging concentrators were realized by the solar t.hermal community. Some researcin work was carried out in the development of nonimaging l"resnel lenses (Collares- Pereira, 1979; Kritchman et al., 1979; l,orenzo and I, uque, 1981), with the two h~rrner works aiming at solar thermal applications, but earlier failures with imaging lenses (Harmon, 1977) led lenses into thermal oblivition. Modern solar thermal markets are established, and concepts have been developed that do not include Fresnel lenses. Research institutions and companies
359
are organized in strict separation of solar thermodynamic and solar electrotechnical departments, reinforcing the stat us quo. Similarly, mirrors never found their way into the photovoltaic community and market. The imaging Fresnei lens of ()'Neill (1978) is the only commercially introduced concentrator technology for photovoltaics, that we know of. It has never been published in the predominantly thermal Solar Energy Journal. In fact, it seems that no other lens specifically designed for photovoltaics has been made public in any journal, or patent, an indication that only ordinary imaging Fresnel lenses are used for solax cells. This paradigma has only recently been broken with the EIJCLIDES project developed by Sala et al., including Luque, which uses parabolic trough concentrators and bifacial cells. Having said all this, and having found only historical reasons for a distinction between photovoltaics and solar thermal developments, why do we follow the same trodden path, and design and test a novel nonimaging Fresnel lens for photovolt ales? The answer is: The novel nonimaging Fresnel lens has been designed with a thermal application in mind. Thermal requirements differ from those in photovoltaics. Medium temperatures can be achiew:d by reducing conductive and conw:ctive heat losses, tracking is more problematic due to transport of the working fluid, and 'hot spots' pose less problems. Absorber design is of some difficulty as its shape is often not flat, heat pipes or fluid operations have to be instalh:d. Testing the performance of the collector is relatively simple in photovoltaics, although concentrator cells must be applied, but output is easily measured, whereas solar thermal application testing requires larger collector arrays. The decision to produce a prototype with acceptance half angle pairs of a cross-sectional 0 = +2 ~ and a perpendicular ~, = 12 ~ fell based on the potential of easier recognition of optical errors when absorber and angles are chosen smaller, and tlne geometrical concentration ratio is selected higher.
2
NONIMAGING FRESNEL AND MANUFACTURING
LENS
DESIGN
The design of tile nonimaging Fresnel lens shown in Fig. 1 has been described in detail in Leutz et al., 1999. Based on the priciples of edge rays, and minimum deviation prisms, under the condition of a smooth outer surface, the optimum shaped Fresnel lens has been found ill a numerical simulation. The design of the line-focusing lens is based on the definition of two pairs of acceptance half angles: 0 in tile cross-sectionM plane (the plane of the fixtures in Fig. 1), and ~, in the plane perpendicular to it. A rlumber of potential manufacturers in Japan, Germany and the USA were contacted concerning the prototyping of
360
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gular segment similar to those formed by the spokes of a wheel, but without its circular slnape. The lens to be the first prototype was ctnosen to be of acceptance half angles 0 = 2 ~ and r = 12 ~ T h e lens was truncated at half height based on previous findings t h a t found the performance of a truncated lens only slightly inferior to the one of a full lens (Leutz t t al., 1999).
Figure '2: Preparations for manufacturing the Fresnel lens prototype. The prisms of an optimum shaped lens (a) are moved and rotated to form a flat sheet (b). Prism tips and grooves are arranged equaldistant to a centerline for ease of moulding. Resized prisms are brought back into shape
(c).
Figure 1" The first prototype of the nonimaging Fresnel le,s under the sun of Tokyo, May 1999. A c c e p t a n c e half angles 0 = 2 ~ t/' = 120.
a Fresnel lens with given specifications. In the first round of consultations, glass as lens material (i.e. the integration of the lens and a vacuum tube) was ruled out a.s impractical. The second round ruh:d out extrusion as manufacturing method for a prototype, and tbund the. cost for a mould for the arched lens prohibitively high. The mould for a shaped lens must. feature a collapsable core due to tl,e undercut formed by some of the prism tips. Finally, it was decided to have the lens manufactured as lint sheet, to be bent into shape later on. A Tokyo based manufacturer was found, and the following specifications were obtained. The lens was to be of a size of 400 • 400 ram, moukled into PMMA of thickness 1.0 mm, with the distance in heigtnt between prism tips and groow.~s being svnaller than 0.5 mm. All prisms of the lens should be designed in such a way that tlney had a common cent erline, i.e. a horizontal line crossing all prisms from which the distances to each one prism's tip and groow: we:re equal. Contrary to common imaging designs, the prisms in ttnis lens are l,ot equaldistant when assembled horizontally. In the shaped version of the lens, each prism cow'rs a an-
T h e lens is prepared for manufacturing by simulating its width according to the maximum dimensions given. T h e absorber width is found accordingly. T h e number of prisms and their coordinates in the shaped lens are calculated under the restrictions of given maximum groove depth. It is l,elpful to be able to have an additional degree of freedom in the simulation, which is the possibility to backstep, i.e. starting a new prism from a given point on the front face of the previous one, thus avoiding the thickness of the lens to bc zero at. the grooves. T h e effects of the 1.0 mm acrylic sheet for refraction (refraction at a plane parallel plate) are dismissed as insignificant on the grounds of being very small. In a first step (see Fig. 2a), the prisms are moved into a horizontal position, and rotated until their front faces form a smooth flat line. Ttne second step deals with the changes necessary due to the centerline requirement. A prism is chosen to serve as reference for setting the position of the centerline. Prisms from the reference prism towards the center of the lens are increased in size in order to have their back faces (which are almost parallel to their front faces, making the prism very 'thin') cross the centerline at a point where the elevation distances between groove and tip are equal. Outward prisms are decreased in size until the same condition is fulfilled. T h e centerline condition facilitates ease of pressing the lens shape into a P M M A sheet. The coordinates of tlnese prisms (Fig. 2b) are given to the manufacturer.
ISES Solar World Congress 1999, Volume III
In a third step (Fig. 2c), the resized prisms of the flat lens are rearranged into the arched shape for two purposes. Small differences are obserw:d in comparisot, to the original shape clue to the resized prisms. The new lens shape must be known to first, produce a frame into which the bent lens is to be fixed and, second, to evaluate the newly shaped lens by ray tracing.
3
PRELIMINARY LENS
TESTS
OF
THE
02/12-
The prototype of the nonimaging Fresnel lens with acceptance half angle pairs 0 = 2 ~ and ~, = 1'2~ has not been available long enough to permit the conduction of detailed tests. However, the lens was mounted on a test rig, and its previously simulated optical design properties could roughly be confirmed. T h e r e never had been great concern about the verification of the characteristics of the original optimum shaped lens, but we were satisfied to see that the lens manufactured as fiat sheet fulfilled all expectations when bent into shape. It can be stated that the amount of material to bc displaced between the prims when the lens is bent into shape does not influence the optical properties of the lens in any visible way.
361
defined against a dark sky, whereas the sun appears larger, with a solid angle of 5.7 ~ per definition filled out by beam radiation. Light entering the concentrator shows on the absorber not only bright against the dark background, but can be described a.s offering a clear threshold between direct rays and darkness. When concentrating sunlight during the (lay, the region around the sun is of almost the same brightness as the sun itself making tracing of rays in the concentrator difficult, bright light and bright absorber are sometimes difficult to distinguish. The light reflected from the moon is 'cold' light, its intensity is too low to heat up the absorber. On the other hand, sunlight geometrically concentrated by a factor of 20, may damage a d u m m y absorber, or measuring device. Still, the sunlight reflected from the moon at night does haw: a similar spectral distribution to sunlight coming directly from the sun during the day, since the moon lacks an atmosophere t h a t could interfere with its characterization of a grey body.
Table 1: The suit and the moon as light sources for the testing of solar concentrators. Average radii, distances to earth, solid angles on 1 May 1999, 21:32 hours, in Tokyo, and relative brightness.
Sun
Radius, m l)ist.ance to earth, m Solid angle (1 May 1999), 0 Relatiw: brigthncss*
0.695- 109 150 9109 0.5291 770
Moon
1.73-10 {; 384 910 r 0.4913 1
*Apparent brightness, or m a g n i t u d e in visible light when celestial body in opposition, i.e. opposite side of the earth from the sun, usually closest, to earth, and best visible. In a first experiment, the lens prototype was exposed to the light of the almost full moon ( 9 9 ~ full on 1 May 1999, whet, the experiment was conducted). Utilizing the sunlight, reflecte(l from the moon for mea.suring the optical properties of a solar concentrator offers some advantage ow:r using the rays of the sun directly. The moon appears to haw: almost the same size as the sun when seen from the earth (Tab. 1). T h e moon's diamel.c.'r is 400 times smaller than the sun's, but the moon is on average 400 times closer to earth than the sun. T h e image of the moon is clearly
Figure 3: Testing the Fresnel lens under the 99% full moon. 1 May 1999. 21:20-21:35, Tokyo, Japan, at 35.50N. The lens is seen from its lower right side. T h e see through absorber, seen from its back, and almost filled out with light, appears in the lower left corner of the picture. A photograph of the moon over the lens is shown in Fig. 3. The see through absorber appears in the lower left corner. The photograph captures a time series of ten exposures over a period of fifteen minutes. T h e moon appears over the longitude of the same location on earth not every 24 hours like the sun (not taking into account analemma), but about fourty minutes later. Titus, the movement of the moon ow:r a point, on earth is slightly slower than that of the s , n . Not 150 of solid angle are covered every hour, but only approximately 14.5 ~ During the 15 minut(..s of photographic exposure in Fig. 3 the moon ha.s covered a solid angle of some 3.5 degrees. Th(: absorber has almost fully been covered with light concentrated by the lens, which has art acceptance half angle of +2.0 ~ a.s expected. Furthermore, the lens has been inclined towards the south, or in the perpendicular direction
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to its cross-section. As will be shown later, aberrations of the refractive behaviour of the lens are smallest in this position at the perpendicular design angle ~, = - 6 1 2 ~ A similar test was conducted under the sun at clear skies few days after the moon experiment (Fig. 1). With the increased brightness, the exact focal area was hard to distinguish from its immediate surroundings, but the optical properties of the lens were confirmed.
~E
1600
"
8O0
~d
400
~
o
o~
100
....
O r m
E
Contrary to imaging lenses under similar conditions we could not observe any color aberrations, which are common near the focal point of imaging devices. Of course, any prism refracts rays depending on wavelength, but the nonimaging lens mixes those rays in the absorber plane, and no color lines accompanying the focal area are observed.
0
I
I
,
.
REFRACTION AND WAVELENGTH PENDENT DISPERSION
| .
n = n(,~)
(1)
Dispersion occurs due to the color-dependent refraction. Shorter wavelengths (ultraviolet, blue) are refracted furtt, er off the surface normal as longer wavelengths (red, infrared). In most cases, three wavelengths are used to describe the dispersion of an optical material. These are the spectral lines for helium at 587.6 nm (yellow light), and hydrogen at 486.1 nm (blue) as well as 656.3 nm (red light), respectively (Shannon 1997). When only one refractiw: index is given, usually the D-line at 589.2 nm is used (.Jenkins, White 1981). "l:he refractive index can be plotted as a function of the wavelength for a material. This function can only be expressed empirically. The most common approach and the industrial standard since Schott abandonned its Schott dispersion formula (Shannon 1997) is called Sellmeier formula. This formula, found in 1871, is not entirely empirical but has a physical basis in describing the dispersion of uncoupled molecules (they are assumed to response with resonance to the passing light waves, and in turn alter the velocity of the light).
n =
1 +
A2 _ ,,/,.j
(2)
3=1
where the wavelength A is in #m. The refractive index of polymethyimetacrylate is sightly smaih:r than the one
.,
.
! .
w \
r
i
.
1.54 1.52 ~ 1.5 ~ 1.48 ~" 1.46
~io
~is
io
Wavelength, 104m
DE-
The speed of light in a medium varies with color. Since the refractive index n of a material is defined as ratio of the speed of light in vacuum to the speed of light in a material, the refractive index is a function of the color of light, i.e. its wavelength.
~ .
PMMA.u , , - . ~ a , ~ ~ /
ols 4
"
1200
Figure 4: Top: Solar terrestrial spectral irradiance (IEC, USA, AM=I.5; Amakawa and Kuwano (1994)). M i d d l e : Transmittance of general purpose acrylic and acrylic with enhanced transmittance for ultraviolet rays (straight and broken lines, Fresnel Technologies, 1995), dotted line authors's measurements of 1.0 mm sample. B o t t o m : Refractive indices of BK7 glass (Schott, 1992) calculated with the Sellmeier formula, and polymethylmetacrylate PMMA ( O s h i d a , 196 l) calculated with the H a r t m a n n formula. All data plotted as function of wavelength.
presented for BK7 glass (see Fig. 4}. The wavelength dependent refractive index may also be calculated with the empirical Hartmann formula. Its accuracy is limited in comparison to the Sellmeier formula, but this should not be of great concern for solar energy applications, since in the visible spectrum accuracy is sufficient. The Hartmann formula is explained for PMMA by Oshida (1961) with the constants for acrylic. A in A.
n = no +
C 93.42 = 1.4681 + X - A0 A - 1,235
(3)
Data for PMMA obtained with Hartmann's formula has been plotted in Fig. 4. Transmittance of PMMA almost reaches that of BK7 glass over the whole solar spectrum. Data for a sample of d = 3.2 mm analysed by Fresnel Technologies (1995) is reproduced i~, Fig. 4. A sample of d = 1.0 mm has been measured for comparison. Further measurements of a ten times thicker sample of general purpose polymethylmetacrylate show that reflection at the surface, and not absorption within the material is the leading cause for transmission losses. Reflection accounts for less than 10% of transmission losses if the angle of incidence is kept below 55 ~ (Jans, 1979). The 02/12 lens is a maximum of d = 1.28 mm at its largest prism.
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363
light on any given point on the prisms' surfaces contributing to the light hitting the absorber. Not all light,, even after the losses on and inside the lens body are accounted for, reaches the absorber. Some rays miss, because of a combination of design principle and refractive laws. The design angles for the lens are the maximum combinations of the acceptance half angles, +0/~p, and -0/r The perpendicular angle ~p is symmetrical along the 2D-lens, but some rays entering at *Pin < *P, and maximum 0, are missing the absorber because refraction does happen in the perpendicular plane as well as it does happen in the cross-sectional plane of the lens.
Figure 5: Dependency of refractiw: index of PMMA relative humidity and temperature.
on
The refractive indices of pla.st.ics are temperature and humidity dependent. The effects of changes in relative humidity are smaller than the effects of chav,ges in ambient temperature on the refractive index. Kouchiwa (1985) examined those changes for an injection molded polymethylmetacrylate singlet lens after having exposed it for a period of 30 days to changes from a defined standard condition of 211~ and 65% relative humidity. He found an empirical equation including both the effects of temperature and relative humidity (Eqn. 4).
'n1~ = n 0 - 1 . 2 5 " 1 0 - 4 A t - 2 . 1 "10-7A t2-1.1"lO-3Arh (4)
winere At, and Arh are the changes in temperature and relative humidity from the standard condition. Kouchiwa's original formula (Eqn. 4) has been plotted for two temperatures 20~ and 40%3 over the range of relative humidity in Fig. 5. The transition temperatures of PMMA are relatively low at 266 K for glass, 399 K for rubber, while the range for melting is at 433-473 K (Pethrick, 1991). The effects on refractiw', index induced by changes in temperature and relative humidity for Fresnel lenses in actual applications (:an be neglected, as the effects are small, unless very high accuracy in imaging devices is required. The lens may serve a.,~outer cover of the collector, and is air cooled.
5
COLOR B E H A V I O U R OF T H E 0 2 / 1 2 - L E N S
Before we describe the pattern of wavelength dependent refraction at the nonimaging Fresnel lens, we shall calculate lmw the receiver is illuminated when light is incident from different directions. All light entering the lens aperture within t.lne acceptance half angle pairs is refracted towards the absorber. This is true for each of the prisms constituting the lens: one may imagine an upside--down pyramid of
Figure 6: Top view of examplatory prism and absorber of the nonimaging Fresnel lens with acceptance half angle pairs +0 = 20 and +~, = 12 ~ Incident light fills out an upside-down pyramid on any point of the lens, is refracted twice at the prism's faces, and shown intersecting the absorber level, where rays form a curved band of light due to perpendicular refraction. Oversized prism, for yellow light, refractive index n = 1.49. Rays entering the lens, their refraction at the front and back faces of the prism and their intersection with the plane of the absorber have been pictured in Fig. 6. In this top view of an examplatory prism of the 0212-lens, it is clearly shown that the band of light incident on the absorber level is curw:d, and some rays incident at perpendicular angles smaller that the perpendicular design angle miss the receiver. Fig. 6 has been drawn using data for yellow light, with the refractive index of 1.49 used throughout the design of the lens. The curvature of the concentration band becomes stronger with increasing design acceptance half angle pair ~b. This is the reason why +, cannot, be increased beyond vMues reasonably comparing to the cross-sectionM acceptance half angle 0. A large +, is desirable from the viewpoint of capturing solar rays during one-axis tracking, where d' should theoretically be +23.45 ~ In the light of the previous discussion of the lens' concentration ratio C = 1/sin 0, the character of ~, az restricting the ability to concentrate in
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this 2D-collector becomes evident (see Leutz al., 1999). So does the need for ray tracing when determining the actual concentration ratio. The incompleteness of illumination of the absorber can be fixed by employing a secondary concentrator, e.g. a Compound Parabolic Concentrator. The acceptance i, alf angle of tiffs concentrator depends on the position of the outermost prism of the lens P ( x , 9 ) , and the extent of the absorber d, enlarged by the maximum extent of the curved band in cross sectional x-direction, or even for bands formed by light of extreme wavelengths, if desired. The absorl~er of the truncated lens becomes the first aperture of the secondary concentrator, whose acceptance half angle must exceed
#cPC,min--tan-Z( x+dmax)y
(5)
For the truncated prototype of the 0212-1ens, this procedure results in a C P C of a minimum acceptance half angle pair of around 55 ~ which in turn yields a geometrical concentration ratio of 1.2. This illustrates that additional concentration after the optimum lens is hardly possible, although l.he incomplete illumination of the absorber can be corrected. The nonimaging Fresnel lens is approaching the ideal concentrator for ~b---0, neglecting losses.
"Fable 2: Polymethylmetacrylate: Refractive indices (Oshida 1961}, transmittance (Fresnel Technologies, 1995). Solar spectral terrestrial irradiance (Amakawa and Kuwano, 1994}, and cumulative solar spectral energy (Wiebcit and ltenderson, 1979) for three wavelengths in the near ultraviolet (uv}, yellow (D), and near infrared (it) light.
Wavelength, 10 -6 m
Refractive. index, Transmittance, % Solar irradiance, W / m ~ (';umulalive energy, %
uv 0.35
D 0.6
ir 1.5
1.515 0.80 483.6 < 2
1.49 0.94 1395.0 25
1.48 0.30 182.0 87
Adding color to these considerations allows for more accurate desciptions of the refraction at the prism. Following the data used in Fig. 4, values for refractive index, solar irradiance, transmittance and cumulative energy corresponding to ultraviolet, yellow, and infrared radiation are found, and listed in Tab. 2. Often, values describing
the solar spectrum are found as bordering the visible (400700 nm), or describing the working range of a particular photovoltaic semiconductor. The response range of crystalline silicone c-Si is 300-1200 nm, that of amorphous silicone a Si 300-900 nm, while crystalline InGaP cells respond to 300-650 nm, GaAs cells to 300-880 nm, and Ge cells (with less conversion efficiency) to 300-1880 nm, respectively. Photovoltaic cells may be stacked to increase their spectral response range. For a detailed discussion see Rumyantsev (1997), or King et al. (1997). The infrared part of the solar spectrum is of no great concern a~ PMMA remains transmittant, and the refractive index is almost constant for large wavelengths (not so in the far infrared). The ultraviolet part of solar irradiance is, while not being absorbed in P M M A with uv-enhanced transmittance (dotted line in Fig. 4 (middle)), refracted increasingly stronger than visible wavelengths.
>, 25 -|
r
\\ 10
I
~. ~". .
." .'. ..._ .] .
5
",SeCtionx 9
.... ---
ye0k~w 0.6 infrared l . 5 [10%nl
~
Figure 7: 0212-lens. Combinations of rays incident at 0 = - 2 ~ and r = - 1 2 , 0 , + 1 2 ~ are drawn with refractions at symmetrical prisms, and intersections with the absorber level. Ultraviolet, yellow, and infrared rays roughly describing the extent of the solar spectrum, with corresponding refractive indices of polymethylmetacrylate. In the three-dimensional Fig. 7 rays incident at combinations of 0 = - 2 ~ and r = - 1 2 , 0 , + 1 2 ~ are drawn with refractions a.,~ well as intersections with the absorber level. Edge rays (those incident at design angles, and calculated with design refractive index) are ifitting the edge of the absorber, while ultraviolet, or infrared rays are generally in greater danger to miss the absorber, although for the case of 0 = - 2 ~ and ~ = 0 ~ from the left prism, the strong refraction of the ultraviolet ray let it reach the receiver while the design ray (yellow) misses. The wavelength-dependent refractive power of a prism reaches a minimum for minimum deviation prisms. The angle of deviation is defined as the angle between the ray incident on the prism and the ray exiting the prism after two refractions. Although a prism can have only one angle of minimum deviation, the idea of 'reversible' prisms (Leutz et al., 1999) used in the novel nonimaging Fresnel lens creates conditions for reduced dispersion.
ISES Solar World Congress 1999, Volume III
Lorenzo (1981) evaluated chromatic aberrations in solar energy systems using Fresnel lenses. He found that lenses with acceptance half angles 0 < 5~ may lead t.o the refracted ray being spread/ wider than the width of the absorber. (,onsiderations included the essentially nonimaging h:ns of Lorenzo and Luque (1981), and mention the possibility to correct chromatic aberrations. This can be done by arranging each prism individually, like in aspherical lens design. The absorber of an imaging design may be placed at what is called the 'cirle of least, confusion' {CLC). Boise Pearson and Watson (1998) calculate the absorber position for this case explicitly, and credit llecht (1990) with the definition of the CLC. The CLC is located where the refracted rays of the longest design wavelength from the right side of ttle lens, and the refracted ray of the shortesl. design wavelength from the left side of the lens (or vice versa) are intersecting (where 'right' and 'left' are the sides right and left of the optical axis of the system defined in a cross sectional view). Using the CLC makes sense for actual imaging design where the focal area exceeds the ideal point, anti an equivalent is useful in nonimaging design (st.e Eqn. 5.). 1]owever, rays missing the absorber are only a minor problem for photovoltaics, whereas inhomogeneous illumination due to shading or color separation is known to influence the electrical current and output of the phot.ovoltaic cell.
Empha.,~is must be put. on this second effect of color aberration, and the behaviour of the 2D lens, where rays are incident within a pair of cross-sectional acceptance half angles +0 from both sides of the symmetrical lens, strongly influenced by the perpendicular acceptance half angle ~,, a.s was seen in Figs. 6, and 7. Presenting the rays in the latter in a cross sectional projection, Fig. 8 is obtained. The yellow rays from both sides hit the edge of the receiver only when *~'i,, = ~)destgn. If the perpendicular incidence is not equal to the design angle, colors are mixing. Sit,(:(: the usual case of operation of the nonimaging Frt'snel lens is collecting solar rays incident anywhere within the acceptance half angh: pairs, mixing of refracted, and color separated rays can be assumed. In fact, the concentrated sunlight on the absorber appears white in an experiment. it l~cks the colors lining the focus characteristic to imaging Fresnel lenses that may be observed in a similar experiment conducted with convelttional lenses, where the color aberrations increase with the rate of incidence deviating from th(: paraxial centerline of the optical system.
6
C()NCLUSIONS
A novel nonimaging Fresnel lens ha.,~ i)een presented in some detail, including preliminary tests of the lens under the full moon, which yield accurate visual results concerning the verification of the acceptance half angle design. "l'h(: h:ns is manufactured a.s flat sheet lens, which is bent
365
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ultraviolet yellow infrared
Absorber
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x
Figure 8: 0212-lens. Combinations of extreme rays incident at 0 = - 2 ~ anti r = - 1 2 , 0 , + 1 2 ~ are drawn with refractions at symmetrical prisms, and intersections with the absorber h:vel. Ultraviolet, yellow, and infrared rays are mixing for incidence other than that at design angles. Cross-sectional projection.
into shape. Both the design and manufacturing characteristics were found to be fulfilled highly satisfactory. The lens material, polymet.hylnletacrylate, is characterized according to temperature and humidity induced changes in its refractive index, which are found to be insignificant. ('hanges of the refractive index for waw;lengths of the solar spectrum are more relevant for practical lens design, and are examined in detail. Color aberration is not a major problem with the nonimaging 2D Fresnel lens concentrator. Not color induced inhomogeneous illumination, but incomplete illumination of the absorber must be regarded as being of prime importance, since colors separated by refraction at the prisms are usually mixed at absorber lew:i, when the perpendicular incidence on the 2D-lens is taken into acount. It is not n(:t:essary to develop a color corrected design approach for this type of nonimaging Fresnel lens. The incomplete illumination of the absorber may call for the use of a s(:condary concentrator, which only marginally increases the geometrical concentration ratio of the system, but ensures complete illumination of the absorber. The nonimaging lens is thought to be a suitable concentrator for pholovollai(: and solar thermal applications.
7
REFERENCES
K. Amakawa. Y. Kuwano (1994) Solar Energy Engineering--Photovoltaics; Advapced Electronics Series I-3, Tokyr E.C. Boes, A. Luque tor Technology;
{1992) Photovoltaic Concentrain: T.B. Johansson, H. Kelly,
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A. K. N. Reddy, R. H. Williams (eds.): Renewable Energy, Sources for Fuel and Electricity. Wa.ghington. I). C,.
Proceedings of the 26th IEEE Photovoltaic ,Specialists Confi'rencc, 29 September-3 October, Anaheim, CA, also at http://www.sandia.gov/pv/ieee.html
.I. Boise Pearson, M. D. Watson (1998) Analytical Study of the Relationship Between Absorber Cavity and Solar Fresn(:l Concentrator; Proceedings of th~ International Solar Energy Conference; A SME, 351-356, 14-17 June, Albuquerque, N M
T. Kouchiwa (1985) Design of a Plastic Lens for Copiers;
M. (3ollares-Percira (1979) High Temperature Solar Collect.or with Optimal Concentration: Non Fo(:using l'resnel i~ens With Secondary Concentrator: Solar Energy 23, 409-420 Fresnel Technologies, Inc. (1995) Fresn(:l Lenses; brochure available at http://www.Jrcsnt:llech.com/ html/products.html, or Fresnel Technologi(:s, Inc.. 101 West Morningside Drive, Fort Worth, "l'exa~s 76110, lISA M.A. Green (1992) CrystMline- and PolycrystallineSilicon Solar Cells; in: T.B. Johansson. H. Kelly, A. K. N. Reddy, R. H. Williams (eds.): Renewable Energy, Sources for Fuel and Electricity, Wa.,d~ington, 1). C.
Proceedings o] the SPIE 1985 International Lens Design (.'on]erence; Volume 554, 419-424, 10-13 June, (.'herry Hill, N J E. M. Kritchman, A.A. Friesem, G. Yekutieli (1979) Efficient Fresnel Lens for Solar Concentration; Solar Energy 22, 119-123 R. Leutz, A. Suzuki, A. Akisawa, T. Kashiwagi (1999) Design of a Nonimaging Fresnel Lens for Solar Concentrators; 5"olar Energy 65, 6, 379-388 (;. 1,6f ((:d.) (1992) Active Solar Systems, PrcIace, Cambridge, MA E. Lorenzo {1981) Chromatic Aberration Effect on Solar Energy Systems Using Fresnel Lenses; Applied Optics 20, "2I, 3729-3732 E. l,orenzo, A. Luque (1981) Fresnel Lens Analysis for Solar Energy Applications; Applied Optics 20, 17, 29412945
V.A.(;rilikhes (1997)Transfer and Distribution of Radiant Energy in Concentration Systems; in: V.M. Andreev, V.A. Grilikhes, V.D. Rumyantsev: Pl,otovoltaic (;onversion of Concentrated Sunlight, (~hichester
O. E. Miller, J.H. McLeod, W.T. Sherwood (1951) Thin Sheet Pia.~tic Fresnel Lenses of High Aperture; Journal of the Optical Society of America 41, 11,807-815
W. (;ra.sse, tt. P. Hertlein, C.-J. Winter (1991) Thermal Solar Power Plants Experience; in: (_:.-J.Winter, R. L. Sizmann, L.L. Vant-Hull (eds.) Solar Power Plants, Berlin
1. Oshida (1961) Step L(:nscs and Step Prisms for Utilization of Solar Energy; New Sources of Energy, Proceedings of the C:onfirence, United Nations, Vol. 4, S/22, 598-603.21-31 August, Rome
S. llarmon (1977) Solar---Optical Analyses of Ma.ssProd,ted Plastic Circular Fresnel L(:ns. Technical note, Solar Energy 19, 105- 108
R. A. Pethrick (ed.) (1991) Polymer Yearbook 8, Chur
E. Ile(:ht (1990) Optics, Reading, MA R. ll. Hildebrand (1983) Focal Plane Optics in FarInfrared amd Submilim(:t(:r Astronomy; Proceedings of
the SPIE- The International Socitty for Optical Engim.ering, Volume ~ 1 , International (~'onfirenct on Nonimaging Concentrators, 40-50. 25-26 August. San l)icgo, CA
M.J.O'Neill (1978) Solar (;oncentrator and Energy Collect.ion System: United States Patent 4069812
V. D. Rumyantsev (1997) I, umine~ent Phenomena in Concentrator Solar Cells; in: V.M. Andreev, V.A.(;rilikhes. V.D. Rumyantsev: Photovoltaic Conversion of Concentrated Sunlight, Chichester
G. Sala. J.C. Arboiro, A. Luque, J.C. Zamorano, J. C. Mifiano, (;. Dramsch, T. Bruton, D. Cunningham (no year) The EUCLIDES Prototype: An Effficient Parabolic Trough for PV Concentration; http://www. users, glo bal net. co. uk/" bloo tl / trackers/eu el. h tm
R. W..lans (1979} Acryclic Polymers for Optical Applications; Proceedings of the Society of Photo Optical In-
Schott (1992) SCHOTT Computer Glaskatalog 1.0, Schott Glaswerke M ainz
strumcntation Engineers (5'PIE), Volume 204, Physical properties of Optical Materials, 1-8, 27-28 August,
R.R.Shannon (1997) The Art and Science of Optical Design, Cambridge
San l)iego, CA F. A..lenkins, lt.E. White (1981) Fundamentals of Oplit:s. 4., international edn., Singapore S. Kaneff (1996) Solar Thermal Power--A llistori(:al. Technological, and Economic Overview; Proceedings of tht
:14th Annual Conference., Australia and New Zealand Solar Energy Society, 294-306, Darwin, ST I). L. King, J. A. Kratochvil, W. E. Boyson (1997) Measuring Solar Sprct:tral and Angle-of-Incidence Effects on Photovoltaic Modules and Solar Irradiance Sensors;
W.T. Welford, R. Winston (1989) High Collection Nonimaging Optics, San Diego .I.A. Wiebelt, J.B. Henderson (1979) Selected Ordinates for Total Solar Radiation Property Evaluation from Spectral Data; Transactions of the ASME, Journal of Heat Transfer 101, 101-107
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THERMO-MECHANICAL DESIGN OF A LARGE COMPOUND PARABOLIC CONCENTRATOR FOR 500 KWt SOLAR CENTRAL RECEIVER SYSTEM Gideon Miron, Shmuel Weis, Ido Anteby, Barak Ostreich, Ephi Taragan ROTEM Industries Ltd., P.O.Box 9046 Beer-Sheva 84190, ISRAEL, Tel: +972-7-6567496 Fax: +972-7-6554502, Email: [email protected] Abstract - A large Compound Parabolic Concentrator (CPC) was required as part of a 1 MW solar plant test facility. A unique modular design that meets the operating requirements is introduced. High solar radiation fluxes are expected on the reflecting surfaces inside the concentrator. The design involves structural and thermal analysis to establish acceptable temperatures and stresses within the reflecting surfaces. Experiments were carried out to evaluate missing data on the adhesive that was used. Temperatures of less than 180~ and stresses of less than 20 Mpa are expected on the hotter parts of the concentrator. The CPC was assembled on site and will go into operation within the next months.
1. INTRODUCTION A 1 MWt Solar Combined Cycle Electricity Generation Plant test facility is now at the last stage of construction at the Weizmann Institute of Science (WIS) in Rehovot, Israel. The project is a joint venture of Rotem Industries, Ormat, the Boeing Company and WlS as an activity within the CONSOLAR Israeli consortium. This test facility will prove the engineering concept for further larger commercial power plants. A prototype 50 KW receiver was run successfully in the Solar Tower (Kami et al., 1998, Weis et al., 1996). The plant is based on the concept of a tower reflector, which directs the light beam from the heliostat field down to an array of secondary concentrators and high temperature receivers. The present facility utilizes an existing heliostat field and includes a new Tower Reflector - located on the existing WIS Solar Tower, a secondary central Compound Parabolic Concentrator (CPC) (Miron,1998), high temperature volumetric receiver peripheral lower temperature CPC's and receivers, and electric power generating turbine and subsystems. Although the solar components are capable of producing 1MW of thermal energy, the plant will operate at 500 KW due to the limitations imposed by the existing heliostat field. The receiver is a high temperature, pressurized device with an inlet aperture diameter of 460 mm. It contains a large conical quartz window and ceramic bed to transfer and absorb the energy to high pressure (up to 22 Bar) - high temperature (1200-1300~ air. A high temperature turbine/generator receives the air to produce electricity, while the outlet air enters a recuperator and heats the inlet air to the receiver. The description of the receiver is not part of this article.
2. GENERAL DESCRIPTION 2.1 Secondary CPC The secondary CPC is attached to the top of the receiver (Fig.l) and accepts radiation from the Tower Reflector. Of 750 kW, which enters the central CPC, 660 kW enters the receiver and around 520 kW are absorbed in the receiver. Loses are based on assuming 2% optical loses, 90% reflection (dirty condition), re-radiation and convection loses from the receiver, conduction and cooling of critical components. The CPC is 5 meters high with inlet aperture diameter of 2.2 meters and outlet of 460 mm. For practical engineering reasons, it approximates the theoretical parabolic CPC's longitudinal form with ten flat
segments ranging in height from 200 to 900 mm and a decagon shaped cross section. Many ray tracing calculations were carried out at WIS to determine the geometric profile, taking into account optical efficiency as well as practical engineering aspects.
Figure 1: CPC and High Temperature Receiver- section view Strict and complex requirements were specified to enable a suitable and economical design. Special attention was given to the thermo-mechanical design of the reflecting surfaces due to expected high radiation fluxes. Radiation fluxes are peaking around 1500 kW/m2 at the bottom sections and adequate cooling is required. Stresses in the glass mirrors due to thermal loading and expansion could lead to cracks and braking. The CPC structure (fig.2) is rigid and self supported while still being light and easy to manufacture and assemble. Only three to four people in a matter of few days assembled the main structure on site. It carries the reflecting panels, maintain them at low temperature with minimal thermal stresses and allow for thermal expansion. The structure is modular and a single reflecting panel can be replaced without dismantling the whole structure. Tight dimensional tolerances are kept with minimal misalignment between adjacent reflecting panels. Each
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reflecting panel consists of a back coated glass mirror glued to an aluminum plate, which is in turn assembled on the frame structure.
3.1 General description As the light is being reflected inside and along the CPC, it is being concentrated to higher radiation fluxes and the cooling requirements are more intensive. At the upper sections, only moderate cooling is required. This is done by attachment of commercially available flat water-cooled heat exchangers (not shown in picture). Assuming an unclean state of the glass surface, absorbed heat load from returned thermal radiation (reradiation) in the lower parts is between 16 to 60 KW/m 2, depending on glass type, and additional 100 KW/m 2 from direct solar radiation. With sufficient cooling water on the backside of the lower parts, a temperature gradient between the glass surface and the backside is created. Due to the temperature difference and different thermal expansion of the materials, thermal stresses occur. It was therefor important to evaluate the thermal stresses as well as the temperatures in order to prevent glass failure or panel/glass separation. Two different glass types were checked in the analysis: Type 1 is quartz with conductivity of 1.45 W/m~ thermal expansion of 8xl0 7 1/~ and Poisson ratio of 0.17. Type 2 is a commercially available float glass with thermal conductivity of 0.95 W/m-~ thermal expansion of 8.5x10 "~ 1/~ and Poisson ratio 0.23. 3.1 Analysis A Finite Element Analysis (FEA) was performed on two typical sections of the bottom part of the CPC, designated CPC10 and CPC11 (Fig. 3 ). The sections include the glass, the adhesive and the supporting plate which in one case (CPC 1 l) is made of copper with rectangular cooling duct, and in the other (CPC10) of aluminium with circular cooling passes. Since the mechanical properties of the silicon adhesive are not fully defined, a broad range of values was used in the analysis. Later tests have confirmed the order of magnitude assumed.
Figure 2: Assembled CPC, external view showing the modular structure and supports. 2.2 Thermal design Thermal and structural calculations, including a FEA modeling, were carried out for the reflecting panels at the specified radiation fluxes. Different material specifications were considered and tested to verify unknown material properties and a suitable combination was selected. Experiments were done to establish the shear failure mode between panel/ adhesive/coating/glass. 2.3 Pre-Fabrication mechanical model A partial full-scale prototype of the CPC, comprising a set of 9 modules was manufactured and assembled for concept evaluation and testing. The manufacakring procedure and assembly were tested and approved. Plate and glass cutting and bonding were evaluated and tested for accuracy and ease of assembly. Several options were considered and one was chosen. As a consequence of the evaluation, slight modifications were incorporated into the final design.
3. Thermo-Mechanical Analysis
Figure 3: FEA model cross sections ofa. CPC10 and b.CPC11 The heat transfer coefficient on the waterside was calculated separately using Dittus-Boelter and Petukhov-Popov correlations. A high value of close to 10000 W/m2-~ was obtained.
ISES Solar World Congress 1999, Volume III
3.2 Results and discussion Temperature and stress results for the two CPC sections are presented in tables 1 and 2 below: Glass
type
Adhesive thickness
(mm) 0.05 0.05 0.05 0.1 0.1 0.1 0.05 0.05 0.05 0.1 0.1 0.1
Adhesive Module of elasticity
Max. stress in glass
Max. Temp.in adhesive
(MPa)
(oc)
10 1 0.1 10 1 0.1 10 1 0.1 10 1 0.1
20.8 5.1 0.6 16.6 3.1 0.4 11.9 16.2 18.1 7.1 15.5 17.9
94.1 94.1 94.1 133.7 133.7 133.7 114.4 114.4 114.4 168.8 168.8 168.8
Table 1. Results for bottom CPC 11 section Glass
type and location 1 top 1 top 1 top 1 top 1 botm. 1 botm. 1 botm. 1 botm. 2top 2 top 2 top ] 2 top 2 botm. 2 botm. 2 botm. 2 botm. _
_
_
_
Adhesive thickness
(mm) 0.05 0.05 0.1 0.1 0.05 0.05 0.1 0.1 0.05 0.05 0.1 0.1 0.05 0.05 0.1 0.1
Adhesive Module of elasticity
Max. stress in glass (MPa)
Max. Temp.in adhesive
1
5.214 1.236 3.341 0.929 3.393 1.17 2.33 0.92 2.004 1.336 0.818 1.156 0.924 1.398 0.523 1.123
56.8 56.8 72.2 72.2 98.6 98.6 145.5 145.5 52.6 52.6 65.5 65.5 95.1 95.1 139.6 139.6
0.1 1 0.1 1 0.1 1 0.1 1 0.1 1 0.1 1 0.1 1 0.1
(~
369
the metal has a higher thermal expansion coefficient, the temperature gradient serves to reduce stresses in the glass. At the same time as the adhesive becomes thicker it creates a more flexible connection between the mirror and the metal plate. As a general rule, the thicker the adhesive is, the lower the stresses are. The temperature is then higher and must be within an acceptable range. A temperature of 180~ was set as a maximum design limit. It was interesting to note that, although with quartz mirrors the stresses are smaller as the module of elasticity decreases, it acts in the reverse for most float glass cases except one. Further attention should be given to the influence of combined parameters in order to develop an optimized solution. Experiments that were done to establish the module of elasticity of the adhesive came out with a wide range of results. Part of it is due to the complex way in which the adhesive stretches and separates from the aluminium. A module of elasticity of between 0.01 to 1 is now thought to be an acceptable number. The influence of temperature and possible degradation is still to be studied, although sample panels were subject to temperatures as high as 200~ for long periods of time without noticeable damage. For all cases, the max stress and temperature are below the design temperature and the material allowable stress. Glass allowable stress is not fully defined but acceptable figures are between 20 to 70 Mpa. For the lower range of module of elasticity, quartz produces lower stresses than with float glass. Since the adhesive properties are not fully defined and local load and stress phenomena could be anticipated an adhesive of 0.05 mm was finally selected. Quartz mirrors were considered for the hotter CPC 11 section. 4. SUMMARY
1
i
Table 2. Results for bottom CPC10 section Results of CPC10 are shown for top and bottom sections separately. The bottom section receives higher heat flux of solar and re-radiation. Since the model is two dimensional at typical horizontal cross sections, the results are assumed to be on the conservative side. In the actual case, heat will flow from bottom to upper sections. It is clear that the adhesive acts in two different ways. On one hand it is a thermal barrier which results higher glass temperatures. As the thickness increases the temperature of the glass increases. Too high temperature will lead to possible adhesive failure and mirror separation. On the other hand since
A unique design was carried out for what is believed to be the largest CPC of its kind. The mechanical design involved a modular structure with reflecting panels which are exposed to varying heat loads. The bottom most sections are subject to heat load of up to 160 KW/m2 which has to be removed. This load leads to high temperature and thermal stresses on the glass mirrors and adhesive. The analysis has shown that with suitable selection of materials, acceptable temperatures and stresses within the allowable limits are achieved. The CPC is already installed at WIS and further measurements will be carried out as it starts operating as part of the complete solar test facility. REFERENCES
Karni, J., Kribus A., Rubin R., and Doron P., (1998). The Porcupine: a novel high-flux absorber for volumetric solar receivem J. Solar Energy Engineering 120, 85-95, Weis S., Maimon Y., Sagie D., Taragan E., Danino M., (1996). Solar Receiver System Compound Parabolic Concentrator, In Proceedings of the 26th Israeli Conference on Mech. Eng., May, Technion, Haifa, Israel pp. 455-457. Miron G., Weis S., Anteby I., Taragan E. and Sagie D.,(1998). Secondary Concentrator for a Commercial Solar Receiver System - Design and Evaluation, In Proceedings of the 27th Israel Conference on Mechanical Engineering, 19-20 May, Technion City, Haifa, Israel.
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SIMULATION AND ANALYSIS OF LOW CONCENTRATION PV MODULES M,ich,ae! Munschauer and Klemens Heumann Institute of Measurement and Control Engineering, TU Berlin, Einsteinufer 19, 10587 Berlin, Germany, phone ++49-30-314 22281, fax ++49-30-314 25526, e-mail [email protected]
Abstract - This paper desribes the methodology and the results of the investigation of low concentration PV modules with regard to energy yield and econmics. The study deals with PV elements that possess low geometric concentration factors and which are situated at locations that are predominantly exposed to diffuse radiation. The mirrors consist of fiat aluminium plates and fiat glass plates respectively. The study links experimental results to computer simulations. The simulations vary the material and the size of the mirrors, the reflector pitch angle, the orientation of the concentrating units and the rated power of the inverter in order to find out the optimal configuration of the PV plant. The simulation results show that low concentrating PV elements represent an attractive alternative to conventional PV plants even at locations that are predominantly exposed to diffuse radiation. Especially two-axes tracked PV plants allow a significant decrease of the energy costs. The measuring results show that the inhomogeneous irradiance on the surface of the PV modules results in significant energy losses due to the serial interconnection of the solar cells.
1. INTRODUCTION Concentrating the sunlight on solar cells by means of cheap refleeting or refracting materials earl significantly reduce the costs of solar energy. It is well known that the employment of coneentrating photovoltaie elements allows a considerable cost reduction at locations that are predominantly exposed to the direct solar radiation. Luque et al. (1995) estimated the energy costs at Madrid (Spain) to 0.07 SU$/KWh for PV plants that use high concentration PV modules and to 0.088 US$/kWh for PV plants that use low concentration PV modules. The energy costs of the conventional PV plants was estimated to 0.253 US$/kWh. Sehumm et al. (1994) investigated the potential of concentrating photovoltaic elements the geometric concentration factors of which are in the range of 11 to 500. They estimated the energy costs for the sunlit location of Widderstall (Southern Germany) to 0.24 US$/kWh bis 0.32 US$/kWh. If the geometric concentration factor decreases, the abilitity of the concentrating PV element to use the diffuse portion of the solar radiation increases. Low concentration PV plants perform better at locations that are predominantly exposed to diffuse radiation. The previous studies consider the diffuse radiation only in a simplified way using sky radiance distribution models. Some authors like Perers and Karlsson (1993), Bollentin and Wilk (1995) or R6rmelid (1996) presuppose an isotropie radiance distribution. Rauh et al. (1996) for instance use the Haymodel to calculate the radiance distribution. The present study is exclusively based on the measured radiance distribution. The objective of this study is: (I) the measurement of the sky radiance distribution for a long period of time and (II) the simulation of low concentration photovoltaic plants with regard to energy yield and econmics. The measurements are temporarily limited to a period of one year and they apply for the location of Berlin. A continuation of the measurements for several years will allow knowledgeable statements concerning the employment of concentrating PV elements. The following section introduces the measuring equipment that is used in this study. Alter that, the influence of the inhomoge-
neous irradiance distribution on the surface of the PV module is discussed. A summary of the the computer models is given before the results of the computer simulations are presented. 2. MEASURING EQUIPMENT
2.1 Measuring equipmentfor the sky radiance distribution The groundwork of all simulations is the meteorological database that contains informations about the sky radiance distribution, the total solar radiation on 289 differently tilted surfaces and the ambient temperature. The data originate in the measuring device that is shown in figure 1.
Fig. 1. Measuring equipment for the sky radiance distribution The supporting construction consists of a vertical metal plate. The plate carries 15 tubes which are anm~ged one above the other. Each tube contains a photodiode that measures the radiance of a sky fragment. Additionally, 8 tubes carry photodiodes besides the fore aperture. These diodes facilitate the measuring of the total solar radiation on differently tilted surfaces. A stepping motor turns the construction at constant angular velocity around 360 ~ During this rotation a measuring circuit records the currents of the photodiodes. Thereby the radiance values of 547, evenly distributed sky fragments and the total global radiation on 289 differently tilted surfaces are measured.
ISES Solar World Congress 1999, Volume III
One rotation needs 80 seconds and is repeated every 160 seconds. Digital temperature sensors measure the ambient temperature and the temperature of the photodiodes. A microcontroller drives the stepping motor, reads the temperatures and controls the measuring circuit. The measuring device is in operation since September 1997. 2.2 Measuring equipmentfor PV components For the measuring of the solar cells and PV modules, a measuring equipment was developed and put into operation. This equipment measures the I-V-curves of a conventional solar cell, of a conventional solar module, of a solar cell that is fitted with aluminium mirrors as concentrating devices and of a solar module that is fitted with glass mirrors as concentrating devices. The temperatures of the PV elements are also measured. Figure 2 shows the equipment. The measuring results allow the analysis of the performance of the concentrating devices as well as the adaption of the computer models of the concentrating devices.
Fig. 2 Measuring equipment for PV components The conventional solar cell represents a polycristallin solar cell that is commonly used in commercial PV modules. The cell measures 0.1 m x 0.1 m. Under standard test conditions, the open circuit voltage ist 0.65 V and the short circuit current is 3.05 A. The PV module measures 0.995 m x 0.45 m. The module is composed of 36 polycristallin solar cells that are connected in series. Two bypass-diodes, one per nine solar cells, prevent a voltage breakdown that could occur under inhomogeneous irradiation conditions. Under standard test conditions, the open circuit voltage of the module amounts to 21.1 V, the short circuit current amounts to 3.1 A and the maximum power amounts to 50 W. The solar cell that is fitted with the aluminium mirrors is of the same type and shows the same characteristics as the conventional solar cell. The mirrors are made of an high reflecting aluminium material and measure 0.1 m x 1.0 m. Both mirrors form a V-trough. The solar cell is located in the middle of the trough. The surplus length causes an irradiance on the surface of the solar cell that is typical for long V-troughs. The reflector pitch angles are 60 ~. The geometric concentration factor of such an optical system amounts to C = 2. The module that is fitted with the glass mirrors is of the same type and shows the same characteristics as the conventional PV module. The mirrors measure 0.995 m x 0.45 m and are made of glass with a silvered backside. This kind of mirrors is commonly used for residential purposeses and it is therefore very
371
inexpensive. The mirrors also form a V-trough with the PV module in its center. The reflector pitch angles are 60 ~, and the geometric concentration factor also amounts to C = 2. In contrast to the preceding optical system, the present optical system is of the same length as the PV element. The solar cells, the PV modules and the concentrating devices are fixed on a supporting structure that faces south and that is tilted by 45 ~. A microcontroller montors the equipment. The controller measures the I-V-curves of the solar cells and the PV modules as well as the temperatures of the PV elements. 3. ENERGY LOSSES CAUSED BY INHOMOGENEOUS IRRADIANCE
The measuring results show that the relative energy surplus of the concentrating solar cell in comparison to the energy yield of the conventional solar cell differs clearly from the relative energy surplus of the concentrating PV module in comparison to the energy yield of the conventional PV module. This is discussed in more detail. On the one hand, the mirror concentrates the beam of light that impinges on the concentrating unit at a low angle of incidence. This increases the irradiation on the surface of the PV element. On the other hand, the mirror obstructs the beam of light that impinges on the concentrating unit sideways. This decreases the irradiance on the surface of the PV element. It depends on the alignment of the concentrating unit and on the radiance distribution whether the mirrors increase or decrease the energy yield. If the PV element is a PV module the cells of which are connected in series, another effect may occur that results into a loss of energy. If the direct sunbeams do not impinge perpendicularly on the concentrating unit, the images of the mirrors on the surface of the PV element will be moved in the direction opposite to the incident beams. This causes remarkable differences in the irradiance on the solar cells of the module. As this cells are connected in series, the cell which is exposed to the lowest irradiance controls the current of the complete module. Neither this cell nor the other solar cells give out their maximum power. This can result in significant mismatching losses. Table 1 shows the relative energy surplus of the concentrating solar cell and the concentrating PV module for different weather conditions. Theses values refer to the energy yield of the conventional PV elements. Solar cell
PV modul
8,6%
13,8%
Cloudy day 8,9%
-4,4%
Sunny day
Table 1. Energy surplus of the concentrating solar cell and the concentrating PV module for different weather conditions The concentrating solar cell shows an energy surplus that is almost independent from the weather conditions. This means that the aluminim mirrors concentrate the direct and the diffuse solar radiation well. The loss of irradiation that is due to the shading effect of the mirrors is overcompensated by the additional irradiation that is due to the concentration.
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The concentrating PV module shows an higher relative energy surplus on the sunny day. This is explained by the higher reflectance of the glass mirrors. On a cloudy day, the shading of the diffus radiation along with the mismatch losses that are caused by the inhomogeneous irradiance on the surface of the module are placed in the foreground. Instead of an energy surplus, the concentrating PV modul shows a decrease of the energy yield that amounts to 4.4 % for this cloudy day. These results refer to two exemplary days and cannot be representative for a general statement. Therefore the period of time for this study is prolonged to half a year. Figure 3 shows the relative energy surplus of the concentrating solar cell and the concentrating PV module for the period from October 1998 to March 1999. During the winter months, the concentrating solar cell shows a significantly higher relative energy surplus in comparison to the concentrating PV module. In January, the concentrating PV module even yields less energy than the conventionial PV module. The relative energy surplus of 13.8 % on the examplary sunny day is not representative for the concentrating PV module. It is the result of an high solar altitude angle that effects a low angle of incidence of the direct solar radiation on the surface of the PV module. Neither the relative energy surplus nor the ratio between the relative energy surpluses of the two concentrating devices are constant. A more detailed study reveals that the relative energy surplus does not depend on the total solar radiation that is measured in the plane of the PV elements.
G=
2, ~2 Le(a,0z) cos0 cOS0z00z 0 a 9 a=0 0z=0
(1)
The angle 0 indicates the angular distance between the normal to this surface and the incident radiation. If the radiance is given by the database, the total radiation on a tilted surface can be calculated by G=
547 i=l Lic~
547 ~"~M = ~'~M i=lLiCos0i "
(2)
The angle 0i indicates the angular distance between the normal to this surface and the incident radiation that originates in the ith sky fragment. The quanitity Li represents the radiance of the i-th sky fragment. The solid angle f~M is defined by 2__~ = 0,0115 Sr"
(3)
~M = 547
4.2 Model of the solar cell The photocurrent IpH of the solar cell is given by IpH,STC IpH(E,O,'~) = 1000 W/m'-Xo cos(0)x'(0)( 1 + a i ( O - 25~
) E"
(4) IpH,STC represents the photocurrent under standard test conditions, E describes the irradiance, x0 is the maximum transmittance and x'(0) indicates the relative transmittance that depends on the angle 0 of the incident radiation. The temperature coefficient ai models the linear dependance of the photocurrent on the cell temperature 0. If the database is applied, the modified formula reads
Ips(O,/~) =
IpI/'STC
lOOOW/m,'ronu(l+ai(O-25~
.
(5)
i~LicosOi ({1,~)1:'( Oi )
Fig. 3. Relative energy surplus of the solar cell and the PV module for the period from October 1998 to March 1999 This study indicates that special attention should be paid to the intereonnection of the solar cells in concentrating PV elements. A homogeneous irradiance on the solar cells that are connected in series shoud be strived. 4. C O M P U T E R
MODELS
The following section summarises the models that are used in the following simulations.
4.1 Model of the incident radiation If the radiance distribution is given as a continuous function Le(~0z) that depends on the azimuth angle a and the zenith angle 0z, the total radiation on a tilted surface can be calculated
by
The 1-V-curve of the solar cell is calculated by means of the two diodes equivalent circuit. The temperature of the cell is calculated by an algorithm that considers the thermal energy balance. 4.3 Model of the mirrors The reflection of the incident light rays originating in the 547 sky fragments is calculated by means of a ray-tracing-methode that considers the reflectance of the aluminium and glass mirrors. 4.4 Model of the inverter The efficiency of the inverter depends on the input power. The efficiency is given by _
TI(P')=
l+v~ + I ( 'l +' v7~-) 2 + -P.........~.. Po 2r, p. 4r~po r,p~
(6)
The parameters are the scaled input power Pc, the scaled loss voltage Uv and the scaled loss resistance rv. The scaled input power is the ratio of the input power Pe to the rated power Ps of the inverter.
ISES Solar World Congress 1999, Volume III
5. SIMULATION OF LOW CONCENTRATION PV PLANTS
5.1 Simulated P V plants The following investigations consider four fictitious photovoltaic plants. Each of them shall consist of 796 solar cells. The cells shall be of the same type as the two solar cells which are used in the measuring equipment of section 2.2. Under standard test conditions and without the concentrating devices, each plant gives out a power of 1041.96 W. It is assumend that each of the 796 solar cells within the plant is exposed to the same irradiance so that there are no mismatch losses caused by any inhomogeneous irradiance. Furthermore it is assumed that the power loss that is caused by other mismatch losses, by the contact resistance and by the wire resistance amounts to 4 %. This means that the rated power ot each plant under standard test conditions is 1000.28 W. The first plant considered is a plant that consists of conventional solar cells without any concentrating devices. The solar cells of the second plant are placed in a V-trough that can be rotated in azimuthal direction by the angle a and that can be tilted in longitudinal direction by the angle 13. If this trough faces south, the longitudinal axis shows in north-south direction. Therefore this trough is also referred to as north/south trough. Furthermore the trough is charcacterised by the mirror width w and by the reflector pitch angle e. Figure 4 shows the north/south trough along with one solar cell and the parameters ~ 1~,wande.
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The solar cells of the fourth PV plant are enclosed by the mirrors at all sides. The concentrating unit can be rotated in azimuthal direction by the angle a and can be tilted sideways by the angle [3. This unit is also referred to as threedimensional trough. The trough is also charcacterised by the mirror width w and by the reflector pitch angle e. Figure 6 shows the threedimensional trough.
Fig. 6. Threedimensional V-trough The simulations consider fixed oriented PV plants, PV plants that are tracked about a single axis and PV plants that are trakked about two axes. The alignment of the fixed oriented PV plants is characterised by the angles a and [3. The PV plants that are tracked about a single axis are rotated about an inclined north-south axis which is characterised by the inclination angle [3. The rotation angle o~ is adjusted in such a way that the angle of incidence of the beam radiation is minimal. Figure 7 indicates the inclination angle and the rotation angle co of a system that is tracked about a single axis.
Fig. 4. North/south V-trough The solar cells of the third plant are placed in a V-trough that can be rotated in azimuthal direction by the angle a and that can be tilted in lateral direction by the angle ]3. If this trough faces south, the longitudinal axis shows in east-west direction. Therefore this trough is also referred to as east/west trough. This device is also characterised by the mirror width w and by the reflector pitch angle e. Figure 5 shows the east/west trough along with one solar cell.
Fig. 7. Single-axis tracked system The PV plants that are tracked about two axes are always adjusted in such a way that the angle of incidence of the beam radiation is minimal. The simulations extend from October 1997 to September 1998 and use the database and the computer models mentioned above.
Fig. 5. East/west V-trough
5.2 Maximum energy yield The objective of the first series of simulations is to identify the parameters that allow the maximum yearly de-energy yield Wj. The investigations systematically vary the mode of orientation, the width w of the mirrors and the reflector pitch angle e of the
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four PV plants. The considered reflector pitch angles are E = 30 ~ E = 45 ~ and e = 60 ~ The width of the mirrors is choosen to w = 0.1 m, w = 0.2 m, w = 0.3 m or w = 0.4 m. The mirrors shall be made of aluminium. Table 2 shows the results. Fixed oriented
,,
Conventional solar cells North/ south V-trough
a = 170 ~ I~ = 36~ Wj = 967.2 kWh (100%) a = 170% ~ = 36 ~ w = 0 . 1 m, e = 30 ~ Wj = 1000.8 . kWh (103.5%) East/ a = 167.1 o, west [~ = 48 ~ V-trough w = 0.4 m, = 60 ~ Wj = 1309.9 i kWh (135.4~ Threea = 171.4 ~ 13= dimen42 ~ sional w = 0.4 m, V-trough ~ = 60 ~ Wj = 1188.7
Single axis tracking
~=30~ Wj = 1228 k W h (100%)
Two axes tracking W j = 1277.4 k W h (100%)
= 40 ~ w=0.4m, e = 60 ~ W j = 1757.7 k w h (143.1%)
w = 0.4m, 8 = 60 ~ Wj = 1837.8 k W h (143.9%)
13= 50 ~ w=0.4m, e = 60 ~ Ws = 1850.2 kWh (150.7%)
w = 0.4m, e = 60 ~ Wj = 1901.3 k W h (148.8%)
13= 50 ~ w=0.4m, E = 60 ~ Wj = 2228.0 kWh (181.4%)
w = 0.4m, 8 = 60 ~ Wj = 2323.9 k W h (181.9%)
i kWh (122.9%) Table 2. Optimal orientation, optimal width of the mirrors, optimal reflector pitch angles and the corresponding energy yield for different modes of orientation (size of the solar cell: 0 . 1 m x 0 . 1 m ) Almost every concentrating PV plant gives out the maximum power if the reflector pitch angle amounts to E = 60 ~ and if the width of the mirrors amounts to 0.4 m. The only exception is the fixed oriented conventional PV plant. A more detailed analysis reveals that the energy yield Wj of this plant is scarcely affected by these two parameters. The optimal azimuth angle a reaches from a = 167.1~ to a = 171.4 ~ the optimal inclination angle 13reaches from I~ = 30 ~ to
that reaches from 26.9% in the case of the the conventional plant to 87.5% in the case of the threedimensional V-troughs. The huge difference between the conventional plant and the plant which is fitted with threedimensional V-troughs indicates that the tracking about a single axis enables particularly the exploitation of the direct solar radiation. The differences between the plant that is fitted with east/west V-troughs and the plant that is fitted with north/south V-troughs decreases significantly as soon as the tracking about a single axis is applied. The change from the single-axis tracking to the two-axes tracking gives no significant rise to the energy yield. The relative energy surplus between these two tracking modes reaches from 2.7% to 4.3%. This is explained by the fact that the highest energy yield is attained during the summer months in which the orientations of the single-axis tracked plants and the two-axes tracked plants are very similar. 5.3 Minimum energy costs The objective of the second series of simulations is to identify the parameters that allow the minimum energy cost. The investigations are the continuation of the first series of simulations. As a new component, an inverter is added. The optimal orientations which were elaborated in the preceding simulations are taken over. The investigations confine themselves to the conventional plant, to the plant that is fitted with east/west Vtroughs and to the plant that is fitted with threedimensional Vtroughs. All three modes of orientation are considered. The variable parameter is the rated power of the inverter. The width w amounts 0.1 m or 0.4 m, the mirrors consist of glass or of aluminium, and the reflector pitch angle amounts to 8 - 60 ~ For each configuration the maximum yearly energy yield Wj,AC on the at-side is calculated. For each optimised configuration the energy costs PE are calculated on the basis of the capital value method. The assumed costs are summarised in the following table. They come from different sources (Luque et al. (1995), Winje and Witt (1992), Uhlig and Wagner (1998)). Solar cells
I~ = 50o.
The distortion in eastern direction can be explained by the fact that on many days a cloudy afternoon follows the sunny morning. Turning the PV plants in eastern directions enables them to use more efficiently the higher irradiation during the morning. Among the fixed oriented PV plants, the plant that is fitted with north/south V-troughs yields only a very limited energy surplus of 3.5 % in comparison to the conventional PV plant. This is due to the shading effects of the mirrors which obstruct the direct solar radiation in the morning and in the aRernoon. The plant that is fitted with threedimensional V-troughs performs better and yields a relative energy surplus of 22.9%, the plant that is fitted with east/west V-troughs performs best and yields a relative energy surplus of 35.4%. These troughs allow the concentration of the direct solar radiation continuously from the morning to the evening. In comparison to the fixed oriented plants, the plants that are tracked about a single axis allow a considerable energy surplus
....
5.85 US$/Wp
Aluminium mirror
36.6 US$/m 2
Glass mirror
47.90 US$/m 2
Supporting structure (fixed oriented)
45.20 US$/m 2
Supporting structure (single axis tracked)
74.50 US$/m 2
Supporting structure (two-axes tracked)
95.75 US$/m 2
Assembly
18.60 US$/m 2
.,
Inverter
0,75 USS/W
Operation and maintanance
'
117 US$/year
Lifetime of PV unit
20 years
"Lifetime of inverter
10 years
Write-off time
10 years
Rise of energy costs
0.8%
Inflation
0.8%
Interest rate
10% .
Rate of taxation
.
.
.
.
30%
Table 3. Survey of the paramaters used for the capital value method
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375
Fixed oriented
Single axis tracking
Two axes tracking
Conventional solar cells
PE = 0.99 US$/kWh (100%)
PE = 0.79 US$/kWh (100%)
PE = 0.78 US$/kWh (100%)
East/west V-trough, glass, w = 0.1 m
PE = 0.95 USS/kWh (96.2%)
PE = 0.71US$/kWh (89.4%)
PE = 0.63 US$/kWh (80.7%)
East/west V-trough, glass, w = 0.4 m
PE = 1.12 US$/kWh (113.4%)
PE = 0.85 US$/kWh (106.3%)
PE = 0.87 US$/kWh (110.4%)
East/west V-trough, aluminium, w--0.1 m
PE = 0.97 US$/kWh (97.8%)
PE = 0.73 US$/kWh (92.5%)
PE = 0.65 US$/kWh (83.7%)
East/west V-trough, aluminium, w = 0.4 m
PE = 1.13 US$/kWh (114.2%)
PE = 0.86 US$/kWh (108.0%)
PE = 0.88 US$/kWh (112.1%)
Threedimensional V-trough, glass, w = 0.1 m
PE = 1.30 US$/kWh (131.7%)
PE = 0.79 US$/kWh (100.7%)
PE = 0.71 US$/kWh (90.1%)
Threedimensional V-trough, glass, w = 0.4 m
PE = 3.57 US$/kWh (361.5%)
PE = 1.99 US$/kWh (251.4%)
PE = 2.09 US$/kWh (266.4%)
Threedimensional V-trough, aluminium, w = 0.1 m
PE = 1.30 US$/kWh (131.3%)
PE = 0.81 US$/kWh (102.4%)
PE = 0.73 US$/kWh (92.6%)
Threedimensional V-trough, aluminium, w = 0.4 m
PE = 3.44 US$/kWh (348.1%)
PE = 1.96 US$/kWh (247.0%)
PE = 2.06 US$/kWh (263.1%)
Table 4. Energy costs for different parameters and PV plants The widening of the mirrors increases the energy costs of all PV plants. The plants that are fitted with threedimensional Vtroughs show extreme high energy costs due to the enormous amount of reflecting material that is required. From an economic point of view, the widening of the mirrors is not advisible. In the following discussion, only the concentrating devices the mirrors of which are 0.1 m wide are considered. The choice of the mirror material scarcely affects the energy costs. The glass mirrors are more expensive, but due to their higher quality these costs are compensated by the higher energy yield. Among the fixed oriented PV plants, only the system that is fitted with east/west V-troughs originates energy costs that are lesser than the energy costs of the conventional plant. The single-axis tracked concentrating plants show an higher cost advantage in comparison to the single-axis tracked conventional plant than the fixed oriented concentrating plants in comparison to the fixed oriented conventional plant. The change from the single-axis tracking to the two-axes tracking gives rise to a further cost advantage of the concentrating plants in comparison to the conventional plant. Among the tracked PV plants, the systems that are fitted with east/west V-troughs should be preferred to the threedimensional V-troughs. The two-axes tracked PV plant that is fitted with east/west aluminium troughs produces 16.4% more energy than the corresponding single-axis tracked plant, and this energy is 10.9% cheaper. The two-axes tracked PV plant that is fitted with threedimensional aluminium troughs produces 18.8% more energy than the corresponding single-axes tracked plant, and this energy is 10.5% cheaper. From the energetic and the economical point of view, the two-axes tracking should be preferred to the single-axis tracking. The most economical configuration is represented by the two-axes tracked system that is fitted with east/west aluminium troughs. The corresponding energy costs
amount to 0.65 US$/kWh in the case of aluminium mirrors and 0.63 US$/kWh in the case of glass mirrors. 6. CONCLUSIONS The simulations show that low concentration PV plants are an attractive alternative to conventional PV plants even at locations that are predominantly exposed to diffuse radiation. The widening of the mirrors increases the energy yield slightly and the energy costs significantly. Glass mirrors are more expensive, but due to their higher quality these costs are compensated by the higher energy yield. The optimal azimuth angle a of the fixed oriented PV plants reaches from a = 167,1 ~ to a = 171,4 ~ the optimal inclination angle ~ of the fixed oriented and of the single-axis tracked plants reaches from [i = 30 ~ to ]3 = 50 ~ The optimal reflector pitch angle amounts to e = 60 ~ Among the fixed oriented PV plants, the plant that is fitted with north/south V-troughs yields significantly less energy than the plant that is fitted with east/west V-troughs. The single-axis tracked PV plants yield more energy at lower costs in comparison to the fixed oriented plants. The two-axis tracked PV plants yield even more energy at even lower costs in comparison to the single-axis tracked plants. Among the tracked systems, the plants that are fitted with east/west V-troughs yield less energy, but they yield the energy at lower costs than the plants that are fitted with threedimensional V-troughs. REFERENCES Luque A., Sala G., Araujo G.L. and Bruton T. (1995). Cost Reducing Potential of Photovoltaic Concentration. International Journal of Solar Energy 17, 179-198
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Schumm G., Mohring H.-D. and Knaupp, W. (1994). Nachftihrung mad Konzentration zur Steigertmg des Energieertrages von PV-Systemen. Tagungsband 9. Internationales Sonnenforum, Stuttgart, pp. 503-510 Klotz F. (1997). Photovoltaikanlagen mit passiver Nachffihnmg mad V-Trog Konzentratoren. Themen 96/97, Forschungsverbund Sonnenenergie, K61n, 1997, pp. 54-60 Perers B. and Karlsson B. (1993). External Reflectors for Large Solar Collector Arrays, Simulation Model and Experimental Results. Solar Energy 5. 327-337 Bollentin J.-W. and Wilk R.-D. (1995). Modeling the Solar Irradiation on Flat Plate Collectors Augmented with Planar Reflectors. Solar Energy 5. 343-354
Rfnnelid M. (1996). Static Concentrators for Photovoltaic Modules at High Latitudes. In Proceedings of EuroSun 96, Freiburg, Germany, pp. 853-857 Rauh H.U., Pruschek R. and Weidele Th. (1996). Comparison of Concentrating and Non Concentrating Tracking PV Systems. In Proceedings of EuroSun 96, Freiburg,Germany, pp. 830-835 D. and Witt D. (1992). Energieberatung/Energiemanagement, Bd. 2: Energiewirtschaft. Springer
Winje
Verlag Berlin Heidelberg New York Tokio Uhlig A. and Wagner H. (1998). Versorgtmg von Umsetzerstationen: NetzanschluB oder Solaranlage? Tagungsband 11. Internationales $onnenforum, K61n, pp. 295-299
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PRACTICAL DESIGN CONSIDERATIONS FOR SECONDARY CONCENTRATORS AT HIGH TEMPERATURES Joseph J. O'Gallaflher and Roland Winston Enrico Fermi Institute, University of Chicago, 5640 S. Ellis Avenue, Chicago, Illinois 60637, USA, Phone (773)702-7757, FAX (773)702-6317, e-mail: [email protected]
Abstract- The initial optical quality of many solar dish concentrators is often found to fall well below design goals and to
deteriorate even further with time. This expectation should be taken into account in designing such systems and argues for the use of a secondary or terminal concentrator. The use of nonimaging secondary concentrators in two-stage solar thermal dish systems has been under study for some time and the optical advantages of this approach are well understood. However, practical questions having to do with the thermal behavior of any secondary and its possible effects on the performance of cavity type receivers have only recently begun to be investigated. An experimental demonstration of a "trumpet" type nonimaging secondary concentrator was carried out with a cavity receiver operating at 660C in combination with the Cummins Power Generation CPG-460 7.5 kWe concentrator system. Lessons learned from this and previous experiments are reviewed. There is no evidence of direct heat loss from the hot receiver to the cooled trumpet. The tests alleviated any operational concerns about the effectiveness of active water cooling and have shown that secondaries can be operated successfully at high temperatures without significant problems.
1. INTRODUCTION It has been over 25 years since the formulation of a useful expression for the well known "thermodynamic limit" governing the allowed geometrical concentration of optical systems (see Jenkins, O'Gallagher, and Winston, 1997). This corresponds to the maximum concentration permitted by physical concentration laws. In the case of solar concentrating systems, this limit depends not only on the angular sun size, but also on all sources of optical broadening of the solar image, such as random slope errors on the primary reflecting surface, system alignment errors and tracking errors. Moreover the practical limit for single stage designs typically falls short of the thermodynamic limit by at least another factor of four. If one tries to exceed these limits by making the target area smaller, the consequences will always be a reduction in geometric throughput from intercept losses.
Basic Two Stage Concentrator Configuration
Fig. 1. Schematic Illustration of two-stage dish trumpet concentrator ( not to scale). The secondary serves to increase the geometric intercept factor and/or achievable concentration for a given set of primary optical tolerances. If the optical quality of a primary does not achieve design goals, the solar image will be enlarged, often exceeding the
receiver aperture size, so that "spillage" of concentrated radiation occurs. Properly designed nonimaging secondary concentrators have the potential to collect this spillage and to increase the geometric concentration of the resulting two-stage system (see Fig. 1), so that it approaches the physical limit. In addition to the throughput losses, the optical broadening described above also usually produces associated damage to the edges of the target aperture itself. To alleviate this potential damage, it has sometimes been suggested that the receiver incorporate an actively cooled aperture plate. It was in the context of just such a situation that an opportunity arose to conduct an experimental test of a "trumpet type" nonimaging secondary concentrator (Winston, and Welford, 1979, O'Gallagher and Winston, 1986) in an operational system. In particular, a lightweight primary, employing circular stretched membrane facets, was under development by the Cummins Power Generation Company, a subsidiary of the Cummins Engine Company. This system, the CPG-460 7.5 kWe concentrator system (Bean, and Diver, 1993), offered the promise of providing a low cost dish-stirling solar electric generating module. Because of deterioration of the receiver aperture plate under operating conditions, a cooled aperture plate was being considered for incorporation into the design. Use of a cooled nonirnaging secondary was an obvious variation and it was decided to conduct a small scale experiment. The use of a nonimaging secondary in combination with a focusing primary (See Fig. 1) permits in principle either the recovery of significant intercept losses while maintaining a fixed geometric concentration or substantial increases in geometric concentration while keeping intercept losses negligible (O'Gallagher and Winston, 1987,1988). This can be done without requiring any improvement in the optical quality of the primary.. The optical advantages of using such concentrators in two-stage solar thermal dish systems have been apparent for a long time. However, practical questions having to do with the thermal behavior of the secondary and its possible effects on the performance of cavity type receivers could only be investigated by experiment with a cavity receiver operating at
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378
high temperature. In this paper we summarize what we feel are the major operating constraints for the use of two stage systems employing secondaries based on our long term experience to date with particular emphasis on the lessons learned from a high temperature operational experiment.
2. THE EXPERIMENT
In the CPG-460 system 24 circular stretched-membrane mirror facets, each 1.52 m in diameter, comprise the primary (Bean, and Diver, 1993). The facets are mounted on a light weight geodesic space frame support structure. The mirror facets are formed from edge supported aluminized polymer-film membranes with the focus for each maintained by a slight vacuum in the cavity between the front and rear membranes of each facet. They are arranged in a partially filled hexagonal close-packed configuration. The facets are totally contained within a circular projected aperture of radius R = 4.8 rrL All the facets are aimed a common focus on the aperture axis corresponding to a focal length F = 5.4 m. This configuration has an effective focal length to diameter ratio (f= F/D) of 0.56 and corresponding rim angle ~ = 48 ~ Trumpet Test- December 12,1995
600
.....
Trt~pet Throat
........
TrmptHb~*100
100 ~
--: . . . . . . . . -:
;.~q,~.,
maintained at the company's headquarters. These were the first and only such tests ever performed with a hot receiver. The optical quality of the facets on this dish was known to have undergone some deterioration due to stretching and/or sagging after prolonged exposure to moisture. Cooling for the trumpet was provided by a separate "openloop" circulating system. The inlet and discharge temperatures for the trumpet cooling were measured by immersion thermocouples. These, in combination with periodic "bucket and stop watch" measurements of the water flow rate, provided a direct measure of the heat load absorbed by the mnnpet. Finally, a single thermocouple was pressed tightly against the back of the "throat" of the trumpet exit aperture at a point where there is a space between the spiral windings of the thin,pet cooling coils. Two days of measurements with no mLmpet were taken with a 7.0 in. (17.8 cm.) diameter aperture and the performance baseline with this larger aperture was measured. Two more days of tests with the mnnpet ( 6.0 in exit aperture) mounted in place were then taken. No baseline measurements with a receiver and aperture directly comparable to the test trumpet exit diameter were taken. Scheduling constraints limited the total trumpet tests to these two days and the highly variable insolation and operational conditions limited the quantitative conclusions that could be drawn, as has already been reported (O'Gallagher, Winston, Diver, and Mahoney, 1996, O'Gallagher, Winston, 1997). In this paper we concentrate on the practical implications of the experiments for future applications of two-stage and secondary concentrators. Plotted in Fig. 2 are time profiles of three quantities for the operational tests with the trumpet in place on December 12, 1995. These are: i) the operating temperature of the cavity receiver, ii) the temperature measured at the exit of the watercooled trumpet throat, and iii) the heat absorbed and dissipated in the tnunpet itself.
0
3. RESULTS l"..e of I)ay
Fig. 2. Measured temperatures and power dissipated in the trumpet during this test of a two-stage system show the effectiveness of the secondary cooling and allow limits to be set on the heat transfer from the receiver to the secondary. A prototype "trumpet" type nonimaging secondary concentrator was designed and fabricated for use with this dish (O'Gallagher, Winston, Diver, and Mahoney, 1995). The final design was selected to achieve the relatively conservative objectives of reducing the receiver aperture diameter from 7.0 in. (17.8 crn) to 6.0 in. (15.2 cm), while at the same time, providing a modest increase in alignment tolerances. The resulting manpet design selected was a hyperboloid of revolution with an asymptotic angle t~ = 50 ~ and a "virtual target" diameter of 7.8 in. (19.9 cm) corresponding to a secondary geometric concentration ratio of 1.7 X. The test units were fabricated from polished copper spinnings, overcoated with vapor deposited aluminum and aluminum-oxide layers and were water cooled. The tests were performed in Abiline, Texas in early December 1995 on one of four Cummins CPG 460 dishes
This one day of on-sun testing with the tnnnpet in place achieved all of our qualitative objectives. In particular: 1) R is clear that there are no fundamental operational problems in operating a water cooled secondary in the immediate vicinity of a very hot (660~
cavity receiver. The
trumpet throat temperature remains less than 100~ throughout the tests. 2) There is no appreciable direct heat loss from the hot receiver to the cooled tnm~pet. That is, these tests have shown that the thermal isolation of the munpet from the hot receiver is very effective. Although the heat absorbed by the mxmpet is much larger than expected (see discussion below), it is not correlated with the temperature differences between the receiver and the thin,pet (in fact it appears to be anticorrelated). That is, there is no evidence of significant direct heat loss from the hot receiver to the cooled tnnnpet. The heat load absorbed by the mmlpet is consistent with being caused by partial optical absorption from a large amount of reflected spillage. 4.0 LESSONS LEARNED. We have noted that the optical quality of the facets on this dish is known to have undergone some deterioration due to stretching and/or sagging after prolonged exposure to moisture.
ISES Solar World Congress 1999, Volume III
We have obtained a quantitative measure of the level of this deterioration from a study of the base line data and the correlation between the insolation and output power (O'Gallagher, Winston, Diver, and Mahoney, 1995). The relationship between the thermal power delivered and the direct insolation under conditions of thermal equilibrium can be represented as a simple linear function of the insolation as follows.
Q,
out = p F A I - Q,
(1)
loss
Here, Q, out is the delivered thermal power, Q, loss is the (constant) thermal loss under the particular equilibrium conditions, p is the primary reflectivity, F is the optical intercept factor corresponding to the particular cavity aperture, A is the net collecting area of the primary, and I is the direct insolation. A regression analysis of the power delivered through a 7.0 inch aperture versus insolation thus can provide a measurement of pFA from the slope and the intercept corresponds to Q,
loss-
Such an analysis yielded pFA
=
0.0249 E r r o r ! a n d E r r o r ! l o s s = - 2.7 kw. With A = 41.6m2 and p = 0.86 (measured on the day of the tests) we found F = 0.69. Because of non-idealities in the actual receiver, this should be regarded as an upper limit on the actual optical intercept. This intercept value above is dramatically less than the expected intercept factor based on the optical quality originally used when designing the trumpet for this experiment. At the time the secondary was originally designed, in the summer of 1993, it was found that the measured focal plane distribution for newly fabricated and aligned facets could be well approximated by a circular gaussian of the form P( r ) = Po e x p [-
r
2
]
(2)
2ro 2
Here Po is the peak power per unit area at the center of the distribution and P( r ) is the power per unit area at a radial distance r away from the center (2) and ro is the characteristc root-mean-square (rms) radius of a given actual distribution. For purposes of designing the secondary, such primary focal plane distributions can simulated by a comprehensive MonteCarlo ray trace model whose parameters can be varied to obtain the best fit to the observed radial distribution. This model was developed earlier (O'Gallagher and Winston, 1987, 1988) and modified to accommodate the faceted primary. The initial optical quality of Cummins primary was well characterized by slope and specularity errors of 2.1 mr and 1.5 mr respectively and a gaussian sun of rms angular subtense 2.73 mr. This yielded a total effective rms angular spread of G = 5.2 mr and corresponded to a radial scale of ro = 3.5 cm. This distribution was an excellent match to the initial focal plane distribution measured in early 1993 for the new, well aligned facets. The optical design for the mmapet used in these experiments was based on these parameters and made no allowance for the subsequent deterioration.
379
For a given characteristic radial scale t o , one can calculate the intercept factor for a given aperture radius R. In particular, it is easy to show that R2 2 ]" (3) 2r o An aperture diameter of 7.0 inches (17.8 cm) corresponds to a radius R = 8.9 cm and the upper limit value of F = 0.69 calculated previously (above and Ref. 7) can be used in Eqn. 3 to solve for a lower limit on the effective present day characteristic scale. Such a calculation yields ro = 5.81 cm. This is to be compared with the original value observed in 1993 of ro = 3.5 cm for which the trumpet was designed. Thus, it appears that this characteristic scale has increased significantly due to the deterioration of the optical quality of the facets between the original focal plane mapping and the time of these first detailed mmapet measurements in December 1995. r = 1- e x p [-
TABLE 1 CUMM]NS CPG-460 DISH CALCULATED OPTICAL PERFORMANCE WITHOUT AND WITH TRUMPET SECONDARY
Percent Diameter Intercept Intercept ImproveOf Factor Factor Cooled ment W/O With Aperture Trumpet , Trumpet
6.0
ino
0.92
0.97
+5.3%
6.0 in. (actual**)
0.48
0.63
+32.0%
inD 7.0 (design quality*)
0.96
(design quality*)
7.0 in. (actual**)
0.55
0.99
0.72
+ i 3.3%
+29.0%
*design quality; t~slp = 2.1 mr: **actual; aslp --- 5 mr An alternative method of evaluating the effect of the optical deterioration can be seen in Figure 3, which shows the heat absorbed by the trumpet as predicted by detailed raytracing. This takes account of multiple reflections of both accepted and rejected rays and of course depends on the optical quality of the primary. The results are shown as a function of primary slope error for two different values of munpet surface reflectivity. The trumpet heating measured during operation was typically near 4 kWth as seen in Fig. 2. This is much larger than the value of 0.5 to 1 kWth predicted from ray trace calculations for
380
ISES Solar World Congress 1999, Volume III
the characteristic rms slope error of 2.1 milliradians that characterized the primary mirror facets when new. In fact we can use this observation to make an independent estimate of the primary slope error at the time of these measurements. The trumpet surface reflectivity was measured to be between 84% and 88% so that, as we see from Fig. 4, a heat load of 4 kWth would correspond to a primary slope error of close to 5 milliradians. This in turn corresponds to a characteristic radial scale of ro = 7.0 em and is completely consistent with the optical quality and intercept factor based on the results discussed above. This indicates that the actual optical quality of the primary used in these tests is very far from the conditions for which the trumpet was optimized. To emphasize the dramatic effect of the optical deterioration of the primary facets on the system performance, in Table 1 we summarize the intercept factors, calculated from our optical model, for the two different physical apertures relevant to our experiment under the two widely different optical condition, (design and actual). 5.0 SUMMARY AND CONCLUSIONS.
Our experience on this and previous experirnents shows repeatedly that the optical quality of any primary can be expected to fall well below design goals and to deteriorate further with time. This expectation should be taken into account in planning future experiments and developing new concentrating systems. A very important aspect of designing any kind of solar thermal electric system has to do with rational evaluation of cost and performance trade-offs. Many different approaches to the design of point focus dish concentrators for electricity generation have been investigated. These include faceted primaries, stressed membrane primaries and secondary concentrators among others. Often when evaluating such concepts, a major emphasis is put on developing a quantitative understanding of the technical performance, perhaps optimizing some standardized measure of conversion efficiency, such a optical quality, without regard to its cost. However, although there are clearly economic motivations in considering these approaches, there is a tendency to be much less quantitative in attaining an understanding of the cost trade-offs involved in optimizing the system. Often performance goals that are unattainable in practical economic systems are set and then used to design other parts of the system. As one of us has previously noted (O'Gallagher, 1994), he practice of maximizing the efficiency of a solar thermal system with respect to some design parameter may not yield the most cost-effective configuration. That is, designs which allow the use of inexpensive materials and construction techniques may not (and probably will not) approach the performance of the most efficient systems one could build. Despite the self-evident nature of these statements, one common approach has been simply to determine those parameter values required for maximum or near maximum efficiency and to select the corresponding designs as baseline or reference configurations. Our recommended methodology for the rational optimization of performance versus cost is based on the constraint that at the optimum, the relative incremental performance gains with respect to a particular performance parameter should balance the incremental costs associated with improvements in that parameter. Under this constraint it was shown that, as long as the cost of the secondary remains small, and unless all costs are virtually independent of optical errors, a two-stage thermal
system, so optimized, must always be cost effective relative to the corresponding single-stage system. When applied to oneand two-stage systems with and without optimally designed secondary concentrators, these models indicates that potential reductions in the cost of delivered energy of at least 10% to 20% and perhaps much more, are possible with secondaries. These gains in turn are likely to far outweigh the cost of the secondary. However a retrofit design strategy for secondaries is problematic. The only realistic approach strategy should include all components including a secondary, from the beginning and realistic technical goals allowing for inevitable performance short falls and deterioration should be adopted. Then the entire system should be optimized. Our recent tests have accomplished all of our operational objectives. In particular: 1) We have shown that there are no fundamental operational problems in operating a water cooled secondary in the immediate vicinity of a very hot (660~ cavity receiver. 2) We have shown that there is no appreciable direct heat loss from the hot receiver to the cooled mmapet. However, due to poor match between optical quality for which tnmapet was designed and the actual dish on which these first experiments have been carried out, the performance benefits associate with the trumpet were not accurately measurable from these tests. Careful attention must be paid to keeping the baseline concentrator and test conditions identical during the various phases of the test as they were not during these very first tests. These experiments have made a good beginning on understanding that much remains to be done to achieve our quantitative goals. However, the experience we have has generated further confidence in the approach so that future tests are being planned.
ACKNOWLEDGMENT: This work is supported by the U.S. Department of Energy under Contract DE-ACO4-94-AL85000 and under Grant DEFG02-87ER-13726.
LIST OF REFERENCES Bean, J. R. and Diver, R. B. (1993) Performance of the CPG 7.5-kW e Dish-Stirling System, Proceedings of the 28th IECEC, Atlanta, GA, Paper No. 93JEC-034
Jenkins, D., J. O'Gallagher, and R. Winston (1997), Attaining and using extremely high intensities of solar energy with nonimaging concentrators, Advances in Solar Energy, 11, IC Boer, Ed., American Solar Energy Society, Boulder, CO,). O'Gallagher, J., (1995), Evaluation of Performance and Cost Trade-Offs in the Optimization of Two-Stage Solar Dish Electric Systems, Proceedings of Solcom- I, The International Conference on the Comparative Assessments of Solar Power Technologies, Jerusalem, Israel, February 1994. O'Gallagher, J. J. and Winston, R. (1986) Test of a "Trumpet" Secondary Concentrator with a Paraboloidal Dish Primary, Solar Energy, 36, 37-44 O'Gallagher, J. and Winston, R. (1987) Performance and Cost Benefits Associated with Nonimaging Secondary Concentrators
ISES Solar World Congress 1999, Volume III
Used in Point-Focus Dish Solar Thermal Applications. Solar Energy Research Institute Report, SERI/STR-253-3113DE8801104. O'Gallagher, J. and Winston, R., (1988) Performance Model for Two-Stage Optical Concentrators for Solar Thermal Applications, Solar Energy, 41, 319. O'Gallagher, J. and R. Winston, (1997) Development and test of a practical mmapet secondary concentrator for cavity receivers at high temperatures, Proceedings of the ISES 1997 Solar World Congress, Volume 2 (Solar Thermal) pp. 222-234, Taejon, Korea. August 1997.
I
8
"
6
"
4
"
2
"
I
I
Predicted Heat Load versus System Optical Quality
/
381
O'Gallagher, J., Winston, R., Diver, R., and Mahoney, A. R. (1995) Improved prospects and New Concepts for Secondary Concentrators in Solar Thermal Electric Systems, Proceedings of the 1995 ASES Annual Conference, Minneapolis MN. O'Gallagher, J., R. Winston, R. B. Diver, and A. R. Mahoney, (1996) Experimental Demonstration of a Trumpet Secondary Concentrator for the Cummins Power Generation (CPG) 7.5 kWe Dish-Stirling System, Proceedings of the 1996 ASES Annual Conference, Asheville NC Winston, R., and W. T. Welford, (1979) Geometrical Vector Flux and Some New Nonimaging Concentrators, Journal of the Optical Society of America, 69, 532-536
I
I
f
f
"
Rho2 = 0.9 Rho2 = 0.8
/ 10 m O
el-
0
I 0
2
I 4
I 6
I
I
8
10
12
Characteristic Slope Error (mr)
Fig. 3. The expected heat absorbed by the trumpet as calculated by detailed ray tracing is strongly dependent on the optical quality of the primary as well as the reflectance of the secondary surface. See text.
ISES Solar World Congress 1999, Volume III
382
COMPARISON OF PREDICTED AND MEASURED PERFORMANCE OF AN INTEGRATED COMPOUND PARABOLIC CONCENTRATOR (ICPC) Roland Winston and Joseph J. O'Gallaflher Enrico Fermi Institute, University of Chicago, 5640 S. Ellis Avenue, Chicago, Illinois 60637, USA, Phone (773)702-7757, FAX (773)702-6317, e-mail: j-ogallagher@,uchicago.edu
Julius Muschaweck Solar Enterprises International, c/o Richardson Electronics, 40W267 Kesslinger Road, LaFox IL 60147, Phone (630) 208-2577 and
A. Rod Mahoney and Veme Dudley Sandia National Laboratories, Albuquerque, NM
Abstract - The Integrated Compound Parabolic Concentrator (ICPC) combines vacuum insulation, a spectrally selective absorber and nonimaging stationary concentration into a single unit. A wide variety of configurations have been investigated and developed. A particularly favorable optical design corresponds to the unit concentration limit for a fin CPC solution which is then coupled to a practical, thin, wedge-shaped absorber. Prototype collector modules using tubes with two different fin orientations (horizontal and vertical) have been fabricated and tested. Comprehensive measurements of the optical characteristics of the reflector and absorber have been used together with a detailed ray trace analysis to predict the optical performance characteristics of these designs. The observed performance agrees well with the predicted performance. 1. INTRODUCTION It was recognized by Garrison (1979) more than twenty years ago that the concept of combining nonimaging concentration with a selective absorber inside a glass tubular vacuum envelope offered the most promising path to providing a solar collector that can be both a high temperature collector and at the same time nontracking (completely stationary). Much effort went into the demonstration of the feasibility of this approach in the early 1980's when an experimental version using a 1.67X Compound Parabolic Concentrator (CPC) achieved an operating efficiency close to 50% while operating at 270C (Snail, O'Gallagher, and Winston, 1984)). In subsequent years this basic concept, now referred to as an Integrated CPC (or ICPC) has been implemented in a variety of sizes, shapes and configurations ( See O'Gallagher, Winston, Schertz, and Bellows, 1988, O'Gallagher, Winston, Duff and Bellows, 1989, and O'Gallagher, Winston, Cooke, and Duff, 1992) as work was carded out to develop a lower cost, manufacmrable version of such a collector. Two years ago, a particularly favorable optical design for an ICPC was selected as the basis for a production version of an ICPC (Winston, O'Gallagher, Duff, and Cavallaro, 1997)) to be used in a cooling demonstration project in Sacramento, CA. More than 300 collector tubes of two slightly different geometries were fabricated, and, in collaboration with the National Renewable Energy Laboratory and Sandia National Laboratory, a prototype module for each geometry was tested at Sandia's test facility in Albuquerque, NM. Preliminary results from those initial performance measurements have already been reported (Winston, O'Gallagher, Mahoney, Dudley, and Hoffman, 1998). The cooling project, employing an array of 316 of these collector tubes to drive a 20-ton commercial double effect chiller on an office building in Sacramento, CA, has been underway since March 1998 and preliminary system performance results from this demonstration are reported in
another paper at this conference (Winston, O'Gallagher, Duff, Henkel, Muschaweck, Christiansen, and Bergquam, 1999). Subsequently, comprehensive measurements of the optical and thermal characteristics of the reflector and absorber materials used in these collectors were carded out in the laboratories of the University of Chicago. These measured performance parameters have provided the basis for a comprehensive ray trace analysis and calculation of predicted optical performance characteristics. In this paper we compare this predicted behavior with the observed performance previously reported. 2. C O L L E C T O R DESIGN The efforts to develop a manufacturable design led to the evolution of a simple low concentration version of the ICPC which provides an elegant solution to several potentially expensive or difficult to implement features of previous concepts. The idealized optical design of this configuration corresponds to the unit concentration limit for a vertical fin CPC solution which is then coupled to a practical thin wedgeshaped absorber as shown in Figure l a. This "vertical fin" design is extremely simple, yet very effective. The concentric heat transfer tube provides rotational symmetry about the long axis so that operations on an automated glass lathe are greatly simplified. Also a relatively low fluid inventory per unit aperture is maintained by use of the "ice-cream cone" shaped absorber. Finally and most importantly, this configuration doubles the effective concentration relative to the usual flat horizontal fin absorber evacuated tube configuration (which loses heat from both sides). This, in combination with a low emissivity selective coating, is sufficient to reduce the thermal losses at operating temperatures between 150~ and 250~ to the levels associated with previous, more expensive ICPC's. The near unit concentration ratio also allows a nearly full sky angular acceptance so that collection of diffuse radiation makes
ISES Solar World Congress 1999, Volume III
the thermal efficiency comparable to or better than a tracking parabolic trough at these temperatures. As for manufacturability, this design does not require that a specially shaped concentrator profile be incorporated either by a metal insert or reshaping of the glass tube. The concentrator is simply the silvered surface of the inside bottom half circular cylinder of the glass tube. The individual tubes are 125 mm in diameter and 2.7 meters long. The "vertical fin" configuration of Figure l a has the advantage of being symmetric, but the disadvantage that, at normal incidence, almost all of the light must be reflected onto the absorber since the large surface area of the absorber fin lies in the shadow of the absorber tube. An alternative asymmetric configuration that has the same effective geometric concentration ( and hence the same thermal loss characteristics) but a lower average number of reflections at normal incidence(and hence a higher expected optical efficiency) is the "horizontal fin" shown in Figure lb. Here, normal incidence sunlight, falling directly on the top absorber surface without reflection, fills more than half the aperture. Tubes with this horizontal orientation retain all the other optical, thermal, and manufacturability advantages of the vertical fin orientation and it was felt that the lower average number of reflections might lead to better overall performance, so approximately half of the tubes were produced in each orientation and modules of each configuration were tested at Sandia. As will be seen, both the ray trace analysis and corresponding measurements confirm that the horizontal orientation results in significantly better overall performance.
measured optical performance. Analysis and discussion of the thermal parameters will be postponed to a later report. Comprehensive measurements of the angular and wavelength dependence of all of these material parameters were carried out using a Beckman DK-2A Integrating Sphere Spectrophotometer. These properties have provided the input data for a detailed ray-trace analysis of the optical behavior of the two configurations. A Monte-Carlo type simulation of the optical performance of each collector configuration was generated by tracing a very large number of randomly chosen rays through the outer glass envelope, off the reflector surface ( for as many times as it hits it) and onto the absorber surface. At each surface, the angle of a given ray's incidence is used to find the fractional energy in the transmitted and reflected rays using the measured angular dependence of the appropriate quantity. In this way, the optical efficiency (the resulting fraction of absorbed to incident energy) is calculated as a function of the angle of incidence in the planes perpendicular and parallel to the long-axis of the tube. For the analysis in this report we restrict ourselves to results in the transverse (perpendicular) plane in terms of the angle, relative to the normal (See Figure 1). The results of the ray trace calculations are shown in Figures 2, 4, 7, and 8.
1.0'
0.9.' 0.7: 0.6~ :-~ 0.5: == -,=~ 0.4 o.
I I
I
o 0.3
I I
0"~,,
i
~-
" q)
Low Concentration Integrated Compound Parabolic Concentrator (ICPC)
if---
383
S
0.2
I
/
II~-
/
~ s
S
I
Azimuthal angular response of ICPC with vertical( )
or horizontal(
) fin
as calculated by raytracing
0.1 I I I I i -1.0 -0.8 -0.6 -0.4 -02. 0.0 0.2 0.4 0.6 0.8 1.0
0.0
Azimuth direction sine bottom half of circular ~ , inside surface is silvered /
/ / ~ ~ --
/ "leo-Cream Gone" Absorber (wilt~ concentric heat transfer tube)
a) symmetric configuration (vertical fin)
b) asymmetric oonfiguration (horizontal fin)
Fig.1. Schematic cross-sectional profiles of the two ICPC configurations analyzed in this paper. The angular response characteristics are presented in terms of the incidence angle, 0, defined to be positive in the clockwise direction with respect to vertical as indicated.
3. RAY TRACE RESULTS. The optical performance of these collectors depends on the reflectance, p, of the silver surface, the absorptance, ~ of the absorber surface, as well as the transmittance x of the outer glass. In this paper we will discuss only the predicted and
Fig. 2. Optical efficiency as a function of sin0 calculated from detailed raytracing for the two fin orientations in Fig. 1. 4. DISCUSSION AND COMPARISON WITH OBSERVED PERFORMANCE. Prototype collector modules using tubes with the two different fin orientations were tested at Sandia National Laboratory during 1997. Preliminary results from the initial performance measurements on these modules have already been reported (7). We observed an instantaneous operating efficiency at normal incidence of 0.68 and 0.734 respectively for the vertical (Fig. 1a) and horizontal (Fig. lb) fin orientations. Fig. 3 shows the measured operating efficiency at normal incidence for the module with the fin oriented parallel to the aperture (Module #2). Note that a typical operating temperature for absorption cooling is about 130K above ambient where the normal incidence efficiency for this non-tracking collector is close to 60% of total insolation. We also measured the angular incidence behavior of the optical efficiency at ambient temperature on Sandia's two-axis tracking (AZTRAK) platform.
384
ISES Solar World Congress 1999, Volume III
The observed dependence on transverse incidence angle is asymmetric because the absorber fin orientation is asymmetric. The angular response in elevation (parallel to the long axis of the tube) was also measured. In addition, we measured similar performance characteristics of Module #1 (with the fin oriented perpendicular to the aperture). The ray trace results for the two orientations are compared directly with one another in Fig. 2. In this representation, optical efficiency is presented as function of the sine of the incidence angle 0. This is informative because the efficiency is scaled between 0 and 100% at all angles, and the same area under the graph represents the same energy for isotropic irradiance. It is interesting to compare the results for vertical and horizontal fins. The predicted difference of from 69% to 76% normal incidence efficiency is substantial. About 2.5% of this is most likely due to a 3 millimeter wide low-absorbing weld strip down the center of the absorber. We attribute the rest of the difference to better average incidence angles onto the absorber. Note that the ray trace results reproduce well the observed difference in normal incidence efficiency of a little more than 5 percentage points, although the absolute efficiencies are about 2-3 % below the predictions. The horizontal fin response is slightly asymmetric, flatter and broader and significantly higher in efficiency across the angular region near normal incidence. The vertical fin response is symmetric but shows a dip at normal incidence due both to the extra reflection loss and to the effects of the angular incidence distributions on the reflector and the absorber. The horizontal fin is quite superior up to 0.7=sin(45deg). At higher incidence angles, the raytracing neglects shading by adjacent tubes and thus overestimates the efficiency. Therefore the horizontal fin is definitely the best choice. The Incidence Angle Modifier (IAM) calculated from raytracing for the horizontal fin orientation is shown in its usual representation (normalized to unity at normal incidence) and plotted versus the transverse angle in degrees) in Fig. 4. The measured angular incidence behavior of the optical efficiency for the asymmetric (horizontal) orientation is reproduced in Figs. 5 and 6. The transverse incidence angle is observed to be asymmetric as expected because the absorber fin orientation is asymmetric. The increase in IAM above cosine for azimuth angles up to 60 ~ is well reproduced by the raytracing calculations shown in Figs. 7 and 8, and also the overall form of the curve. However, the measurement shows more asymmetry than the raytracing does. Also note that the measurement approaches 0 around 80 o while the raytracing remains finite to 90 o. Both these effects are thought to be due to neglecting the effects of adjacent tubes in the raytracing. 5. CONCLUSIONS These initial comparisons of performance measurements with detailed ray trace predictions based on measured material parameters have shown that we understand these collector's optical performance quite well. The degree of agreement gives confidence that we can predict the performance of other designs as well. Together with the measured thermal performance (we measured an instantaneous thermal efficiency of close to 60% of total insolation at about 130K above ambient) this analysis shows that this ICPC with it's full sky coverage ( no tracking) is well matched to the operating temperature requirements of high efficiency (double-effect) solar cooling.
ACKNOWLEDGMENT: This work was supported by the U.S. Department of Energy under Contracts NREL/DOE-XO-211232-1 and YAO-6-16309-01, and under Grant DEFG0287ER-13726. We thank Mary Jane Hale (NREL) for her active participation and support. We are particularly appreciative of the excellent work done by Ms. Sarah Yanes in carrying out all the spectrophotometer measurements of reflectance, transmittance and absorptance. Ms. Yanes was supported by the National Science Foundation under a Research Experiences for Undergraduates (REU) Site Grant. REFERENCES
Garrison, J. D., "Optimization of a fixed solar thermal collector," Solar Energy, ~ 93, (1979). Snail, K.A., J. J. O'Gallagher, and R. Winston, "A stationary evacuated collector with integrated concentrator," Solar Energy, 33, 44 1, (1984). O'Gallagher, J., R. Winston, W. Schertz, and A. Bellows, Systems and Applications Development for Integrated Evacuated CPC Collectors, Proceedings. of the 1988 Annual Meeting, American Solar Energy Society, June 20-24, Cambridge, Massachusetts, p. 469, 1988 O'Gallagher, J., R. Winston, W. Duff and A. Bellows, Development of Evacuated Integrated CPC Solar Collectors. Proceedings. of the ISES Solar World Congress, Kobe, Japan, 1989. O'Gallagher, J. J., R. Winston, D. Cooke, and W. Duff, The New Integrated CPC, Proceeding. of the Annual Conference of the American Solar Energy Society, 220-224, Cocoa Beach, FL., 1992. Winston, R., J. O'Gallagher, W. S. Duff, and A. Cavallaro, The Integrated Compound Parabolic Concentrator, Proceedings of the 1997 ASES Annual Conference,, pp. 4144, Washington, D.C., April (1997) Winston, R., J. O'Gallagher, A. R. Mahoney, V. E. Dudley, and R. Hoffman, Initial Performance Measurements from a Low Concentration Version of an Integrated Compound Parabolic Concentrator (ICPC), Proceedings of the 1998 ASES Annual Conference., pp. 369-374, Albuquerque, June, (1998)
Winston, R., J. O'Gallagher, W. Duff, T, Henkel, J. Muschaweck, R. Christiansen, and J. Bergquam, Demonstration of a new type of ICPC in a double-effect absorption cooling system, Proceedings of the 1999 ISES Solar World Congress, Jerusalem, Israel ( Paper # 266, this conference).
ISES Solar World Congress 1999, Volume III
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125
150
Above
Ambient(K)
175
200
Fig. 3: Thermal performance measured at normal incidence angle. The optical efficiency exceeds 73% of total insolation.
1.1
.
.
.
.
.
I
I
i
i
i
m
1.0 0.9 t.__
._~ 0.8 "0
o
0.7
a)
0.6
<"
0.5
(i)
o t-.
0.4
ICPC with horizontal fin Transverse angular response as calculated by raytracing
-o 0.3 o==.
0
0.2 0.1 .
0.0 -90
.
. -60
.
!
-30
.
,
I
0
,
,
I
.
30
,
60
90
Transverse incidence angle [~ Fig. 4. The Incidence Angle Modifier (IAM) calculated from detailed ray tracing for the horizontal fin orientation (Fig. lb). The curve is normalized to 1.0 at normal incidence so it exceeds unity (somewhat asymmetrically) at off-normal incidence.
ISES Solar World Congress 1999, Volume III
386
ICPC Collector Module #2; Positive Azimuthal Orientation ,v,
K = Cos(la) + 1.23E-O3(la) + 1.703E-O4(la) 2- 4.667E-O6(la) 3 + 2.573E-O8(la) 4
r-
1.1
0
1.0
Q. 00
0.9
~"
,"
T,I
I
-I--7.-----J:
"-_............ ~ ....
-'.-~
i
E
.
J
I
....................... T ............................. !..............T ..........."! ...............
L_
0.8 O~ E
0.7
o.8 r
0.5
"0 0
0.4
r
0.3
"0
0.2
L_
q)
ii ii iiii iiIiiii iiiii iIiiiiiiiiii
- ................ i .............. i ............... i ............... i .............. i ................ i . . . . . . . . . . . .
iii ii iil
'.... ..... ~.o,~.
.............
0.1 I
o.o
0
I
10
20
I
30
I -
40
50
60
70
80
90
Incident angle (deg)
Fig. 5. Measured Incidence Angle Modifier multiplied by cosine 0 for the horizontal fin orientation for positive angles of incidence (See Figure 1). Note that the response exceeds cosine by a significant amount.
ICPC Collector Module #2;Negative Azimuthal Orientation
(D
K = Cos(la)
tO
Q.
1.1
~
.
t
l
I I
1.0
+ 1.993E-03(la)
I I
r r
I I
I I
I
I
2
I I
! !
I
I
....... i .............................. :............... t .............. i .............. 1............... I i
0.9
- 3.68E-05(la)
!
I
/
!
. . . . . . . . . . . . . . . . . . . . . . . . . . : ............... : .............. T .............. I...............
-f ............ T .....' .....:1:, ......... ~
2
,
,
,
z
' ...............
~ ..............
' ...............
I
I ...............
0.8
-1 ..............
T ..............
i ..............
T ..........
~
..... ' ...............
0.7
0 r
(D "0 r0
0.6
9
0.5
"'"
i
]
"~"
0.4 0.3
=.n=
"0
0.2 0.1 0.0
0
i
~
10
20
~ 30
40
~
~
~
50
60
70
80
90
Incident angle (deg) Fig. 6. Measured Incidence Angle Modifier multiplied by cosine 0 for the horizontal fin orientation for negative angles of incidence (See Figure 1). Note the asymmetry with respect to positive angles (See Fig. 4) in qualitative agreement with the ray trace results (Figs. 2 and 4).
ISES Solar World Congress 1999, Volume III
387
1.1"
1.00.9
cc sine.---~
L_
.~_ 0.8 tim "O
o 0.7 0.6 0.5 -
<' -
o 0.4 -
c"
(D -o O .--.
,IL
I C P C with horizontal fin
.
0.3
-
-~ 0.2 -
o.1
o.o 0
Transverse angular response as calculated by ragtracing
I
I
I
I
I
10
20
30
40
50
60
70
80
90
Transverse incidence angle [~
Fig. 7. Calculated Incidence Angle Modifier multiplied by cosine O determined by detailed ray trace for the horizontal fin orientation for positive angles of incidence.
1.1" .
1.0-
"
~
0.9
ccsine..--~' ~
. t..
._~
0.8 =
"0
o 0.7
9 0.6
\
C:~
<" 0.5
o 0.4~ c-
I C P C with horizontal fin
"o 0 . 3 ~
Azimuthal angular response
qD
.._
O
0.2~ 0.1
as calculated by raytracing
I
0.0 0
-10
-20
I
I
I
-30
-40
-50
-60
-70
-80
-90
Transverse incidence angle [~
Fig. 8. Calculated Incidence Angle Modifier multiplied by cosine 0 determined by detailed ray trace for the horizontal fin orientation for negative angles of incidence. The direction of the asymmetry with respect to positive angles observed in Figs 5 and 6 is predicted but it is not predicted to be as pronounced as is observed. See text.
388
ISES Solar World Congress 1999, Volume III
DOUBLE-TAILORED IMAGING CONCENTRATORS Harald Ries Optics & Energy Consulting Landsberger Str. 476, D-81241 Miinchen, Germany Jeffrey M. Gordon
Ben-Gurion Universit yof the Negev, Beersheva, Israel Department of Energy and Environmental Physics, Blaustein Institute for Desert Research and The P earlstoneCenter for Aeronautical Engineering Studies Department d Me hanical Engineering A b s t r a c t - We present a new approach in optical design whereby tw o-stage axisymmetric reflectors are tailored with a completely imaging strategy, and can closely approach the thermodynamic limit to radiation concentration at near-maximum collection efficiency. Practical virtues include: a) an inherent large gap between the receiver and the second-stage mirror; b) an upward-facing receiver; c) the possibility of compact units (large rim angles), i.e., low ratios of total depth to total width; and d) no chromatic aberration. We describe how one can tailor both the primary and secondary mirrors so as to insure that spherical aberration is eliminated in all orders, and circular coma is canceled up to first order in the angle subtended by the radiation source. An illustrativ e solution that attains about 93% of the thermodynamic limit to concertration is presented for a far-field source, as is common in solar energy and infrared detection applications. Doubletailored imaging concentrators are similar in principle to complementary Cassegrain concentrators that comprise a paraboloidal primary mirror and the inner concave surface of a hyperboloid secondary reflector, but have monotonic contours that are substantially different with far superior flux concentration.
INTRODUCTION
This does not, ho w ev erpreclude strictly imaging design strategies for high-flux applications. T raditional imaging devices are designed to correct ubiqMaximum concentration of radiation at high col- uitous errors in the sense of a series expansion up lection efficiency is important in solar energy uti- to a certain order. The usual parameter used is the lization and infrared detection. It is commonly as- distance from the optical axis under which a ray serted that imaging optical systems are not w ell enters the system. In this contribution w e use a suited to the task [1]. A familiar example is a high- different parameter namely the angular size of the end camera objective for a 50 mm lens marked source itself 20. Hence the devices are particularly fl.4, which produces an image a factor of 8 less w ell suited to small sources and large apertures. bright than the source itself. Namely, this con- In the far-field approximation, the thermodynamic v en tionalimaging system only reaches one-eighth limit for flux concentration is (sin(8)) -2. of the thermodynamic limit to concentration which corresponds to the brightness of the object itself [2]. Indeed it can be shown that it is impossible to perfectly image a finite part of the plane [1].
ISES Solar World Congress 1999, Volume III
CASSEGRAIN
389
Of 2 secondary
An optical imaging device kno wnas Cassegrain consists of a parabola and a section of hyperbola Qfl which share one focus in common. Objects at infinite distance are first imaged by theparabolic reflector and then, in a second stage, by the outer side of the hyperbola as sho wnin Fig. 1. This classic Cassegrain is typically used as telescope. Its disadv a ~ g e as a concentrator is its inherently long focal length. The first (parabolic) stage acts as a concentrator, while the second stage in the classic Cassegrain acts as a deconcentrator, because rays Figure 1: Classic Cassegrain reflector design conare reflected on the outer, con v exside of the hy- sists of a primary paraboloidal concentrating reperbola. The classic Cassegrain design is not suit- flector and the outer convex part of a hyperboloid able for high concentration because, seen from the as second stage. P araboloid and lyperboloid share absorber, the hyperbolic second stage can subtend one common focus, f2. The radiation parallel to only a limited angle. the optical axis is collected at the second focus of In a previous paper we suggested using the inner the hyperboloid, fl. concave part of the closer side of the hyperbola, a design w ecalled Complementary Cassegrain [3] shown in Fig. 2. In the complementary Cassegrain, Q f2 both stages act as concentrators. This is essential if one strives for high concentration. Although secondary this design does allo w for high concentration it is plagued b y imaging errors. In this paper, w edeQfl scribe how to modify the shape of the primary and secondary reflectors, so as to correct these errors in the limit of a small source.
3
DOUBLE TAILORED IMAGING REFLECTORS
Assume for simplicity a point source. If we wish to design a reflector which concentrates all radiation from this source into an image point the simple solution is an ellipsoidal reflector with foci at the source and the image. If the source is infinitely far a w a y the ellipsoid becomes a paraboloid. If the rays undergo tw o reflections before rea~ing the image, we gain one degree of freedom which may be used to correct errors. T a k i n g ad~ntage of the small source size we expand the errors in terms of the angular size of the source. We call an optical device ideal for a small source if errors are corrected up to the first order in the angular size of the source. The leading error term is therefore quadratic.
Figure 2: Complementary Cassegrain reflector design consists of a primary paraboloidal concentrating reflector and the inner concave part of a hyperboloid, which acts as a second concentrating stage. Similar to the classic Cassegrain, paraboloid and h yperboloid share one common focus, ]2. The radiation parallel to the optical axis is collected in the second focus of the hyperboloid, fl.
Rays which deviate from a point source by a certain angle will reach the nominal focal plane at a distance from the focus which is proportional to that angle to first order. Correcting errors up to first order then means that this proportionality
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constant is independent of the incidence position on the input aperture of the system, as shown in Fig. 3.
move it along a ray s u c h that the size of the focus is constant for all rays. In effect w e replaced the common focus f2 of hyperboloid and paraboloid in our complementary Cassegrain design b y a caustic which is appropriately defined to conserve total magnification. The argument as outlined holds for the sagittal plane which includes the optical axis. A detailed calculation shows that a similar condition for the tangential plane actually has the identical solution.
secondary
point imaging
,,
Lo
rber
system r
focus Figure 3: Schematic of a general point imaging system. Input rays parallel to a given direction are focused to a point. The output direction at the focus depends on the input position.
:
.
Figure 4: Schematic of a rotational point imaging system. The output direction r at the focus depends on the input position r. F ermat'principle requires the sum of the strings L0, L1, L2 to be constant
Since primary and secondary reflector can only be designed simultaneously, we call this procedure double tailoring [4, 5]. The device w e describe The solution for this problem was formulated here is imaging because all rays undergo precisely by Abb ~ for an axisymmetric system illustrated in t w o reflections in W o co~in uous reflective surfaces. Fig. 4: the sine of the exit angle r in the image Therefore the o~erall map of incoming rays into plane must be proportional to the distance r of the outgoing rays is also contin uous in contrast to nonray from the axis in the entrance plane. A different imaging devices. Nevertheless, the surfaces dew ayto understand the Abb~ sine condition is to scribed here are not conic sections or compound look at the magnification of the system. The pri- parts thereof. mary reflector produces a primary image the size of which is proportional to the distance of the priGOVERNING EQUAmary along the ray. The secondary which is re- 4 quired to image this intermediate point into the TIONS AND SOLUTIONS final focus has a magnification given b y the ratio of the distances of intermediate image and final fo- In an axisymmetric system, the incident cone of cus. Since the intermediate image is free w emay rays from the source produces an elliptical focal
ISES Solar World Congress 1999, Volume III
spot. Let r denote the radial position at the entrance aperture of the primary mirror, and let r denote the angle made by a point on the secondary reflector with the optical axis, as illustrated in Fig. 4. Each axisymmetric imaging design is characterized b y a particular function r - r From the condition of imaging, and from phase space conservation in geometrical optics, one can derive the lengths of the axes in the tangential and sagittal planes, ]t and ]8 , respectively: ]~(r)-
sin(r
(1)
r
0 sin(r h(r)
=
Or
"
(2)
T o w a r d eliminating aberrations, evrequire that the focal spot be circular and of constant radius, i.e., that the sagittal and tangential axes be equal and independent of r. This requirement is satisfied b y a function of the form
-0.2 -0.4
-1 .o.
-1.0
-0.5
0.0
0.5
1.0
Figure 5: Double-tailored primary and secondary imaging reflectors. The absorber is barely visible because the figure is drawn to scale. Note that the secondary subtends the entire half angle visible from the absorber. This implies that, except for shading losses which in this example amount to 4% of the incident radiation, the device reaches the theoretical upper limit to concentration.
where K is a constant [6]. sin2(r = (K/rma=) 2 is the fraction of the thermodynamic limit for which one designs the system, where rma= is the radius of the primary aperture, and r is the largest angle incident on the absorber. The actual flux concentration is reduced by various losses, detailed in Section 5. T o calculate the coordinates of the primary and secondary mirrors, we use Fermat's method of constan t optical path length. The optical path lengths Lo, L1 and L2 are illustrated in Fig. 4. L0 starts at an arbitrarily chosen reference plane. L0, L1 and L2 are required to satisfy
0.05
Lo + L i + L2 - c o n s t a n t .
ab.sorber
-0.6 -0.8.
0.10-
- Kr
i ,,"e~"=~.,,., secondaryreflector
0.0
(3)
sin(r
391
secondaryreflector
0.00'
sorber
-0.05,
-0.10
-~.2
-~.1
0:0
0:1
0:2
Figure 6: Double-tailored secondary imaging reflector with u p w a r d facing absorber. Enlarged part of previous figure. Absorbeand secondary reflector are drawn to scale for a circular source of half-angle 0.01 rad and a primary reflector of unit radius
(4)
We use Snell's La wat both the primary and secondary reflectors, which provides t w o differential equations. With Equ. (3) and Equ. (4) we have 4 equations in 4 unknowns which are the coordinates of the cross-sections of the primary and secondary contours. This set of equations is readily solved n umerically once boundary conditions are pro vided, namely, the rim positions of the primary and secondary, as well as the focus and the optical axis. In Fig. 5 w e show the con tours of a t w o stage imaging concentrator tailored according to the
abo ve prescription:Fig. 6 zooms into the secondary and absorber. We ha v echosen an extreme case where secondary and primary are near to each other, leading to a compact device. In principle the starting point of primary and secondary are arbitrary .In order to demonstrate the ability of the device to approach the theoretical upper limit to concentration we have chosen the secondary rim in the absorber plane. Thereby we force the maximum exit angle to be perpendicular to the axis.
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ISES Solar World Congress 1999, Volume III
If we compare the tailored device described here to a Complementary Cassegrain based the same curv atureof primary and secondary at the apex, then as one moves out w ard~w ayfrom the axis, the tailored primary shows slightly less curvature than that of the paraboloid of the Complementary Cassegrain, whereas the secondary curves slightly stronger than its hyperboloidal analog. In fact for extreme designs of Cma= = lr/2 (interesting only from an academic point of view) the outer sections of the primary become hyperbolically curved, while the secondary becomes elliptical. In Fig. 5 and Fig. 6 w eshow an extreme design where the secondary extends all the way to the focal plane. This illustrates that it is indeed possible to reach the thermodynamic limit with a purely imaging design.
030 O25
02o ._~ 0.15
~
0.10
0.0s 00o o.oooo
" o.&~
"
o.o~,o
"
o.o~,~
o.o~o " o.o~
Sin2(e) Figure 7: Graph of ray rejection against sin2(0) for the double tailored imaging concentrator noted in the text. The open circles represent raytrace results. The solid line is the best linear fit.
RAYTRACE RESULTS We account for geometric losses only, i.e., perfect mirrors are assumed. F our classes of losses must be quantified: (a) shading of the primary by the secondary; (b) bloc king of the secondary b y the absorber; (c) rays that miss the secondary; and (d) rays that do not miss the secondary but miss the absorber. F or small0, shading is simply the ratio of the area of the secondary to that of the primary. As an illustrative example, we present computer raytrace results for a double-tailored imaging concen trator that is nearly identical to that shown in Fig. 5 and Fig. 6. The primary reflector has unit radius. The far-field source had 0 - 0.01 rad. The concentrator w as designedfor Cma= = 84-3 ~ ] = 1.016 and H = 1.073, with f and H respectively indicating the heigtt of the absorber and of the apex of the secondary abo v e thatof the primary. Blocking turns out to be negligible, shading is 0.04, and ray rejection is 0.017. Hence the flux concentration achieved isaround 0.93 of the thermodynamic limit, and the collection efficiency is about 0.94. In Fig. 7 w e summarize the results for assorted source sizes tow ard testing the ~idit y of the small source approximation. Note that collection and concentration losse~ncrease roughly proportional to sin2 (0), as expected. Suc h performance is comparable to the best nonimaging designs dev eloped to date, and clearly
leaves little room for imprcv emert in terms of flux concentration and collection efficiency. One challenge lies in designing ever more compact units. T o w ard thaend, one can trade off shading losses against an oversized absorber, i.e., designing for a maximum exit angle below 7r/2, while trying not to compromise flux concentration or collection efficiency.
6
DISCUSSION
The limits on flux concentration with conventional imaging strategies are w ell documented [1, 2, 3]. Restoring the brightness of the source usually engenders a substantial sacrifice in collection efficiency. Alternatively, high collection efficiency comes at the expense of marked reductions in flux concentration. Nonimaging optics has produced a spectrum of high-flux high-efficiency designs [1, 2, 7, S]. We ha v e solv ed thproblemf determining the mirror contours in tw o-stagestrictly imaging designs, such that spherical aberration is canceled in all orders, and circular coma is eliminated up to first order in the size of the radiation source. An illustrative example for small far-field sources, as is common in solar and infrared collection, reveals designs capable of nearly attaining the thermodynamic limit to concentration.
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393
It is tempting to design for reduced shading losses [8] H. Ries and A. Rabl, "The edge ray principle of simply b y creating a secondary that is far smaller nonimaging optics," J. Opt. Soc. Am. A 11(10), than the primary. The bottleneck to this approach pp. 2627-2632, 1994. Also included in Selected stems from the fact that so much concentration Papers in Nonimaging Optics, SPIE Milestone then derives from the primary mirror, with an enSeries Vol. MS 106, Roland Winston, editor, larged image of the source in the secondary reflecSPIE Optical Engineering Press, Bellingham, tor, that the small- source approximation grows inWash., 1995. valid for the secondary reflector. In solar concentrators and infrared detection systems, introducing a sizable gap between the receiver and the second-stage mirror is often mandated. Upw ard-facingreceiv ers can be advantageous. The types of concentrators delineated here satisfy both requirements. In addition, purely reflectiv e systems do not incur hromatic aberration. ACKNOWLEDGMENTS We wish to thank Daniel F euermannfor valuable discussions and his help in raytracing.
REFERENCES [1] W. T. Welford and R. Winston, High Collection Non-Imaging Optics, Academic Press, New York, 1989. [2] A. Rabl, "Comparison of solar concentrators," Solar Energy 18, pp. 93-110, 1976. [3] D. F euermann, J. M. Gordon, and H. Ries, "High-flux solar concentration with imaging designs," Solar Energy 65(2), pp. 83-89, 1999. [4] J. M. Gordon and H. Ries, "Tailored edge-ry concentrators as ideal second stages for fresnel reflectors," Appl. Opt. 32, pp. 2243-2251, 1993. [5] H. Ries and R. Winston, "T ailorededge ray reflectors for illumination," J. Opt. So c. Am. A 11, pp. 1260-1264, 1994. Also included in Selected P apersin Nonimaging Optics, SPIE Milestone Series V ol. MS 106, Roland Winston, editor, SPIE Optical Engineering Press, Bellingham, Wash., 1995. [6] M. Born and E. Wolf, Principles of Optics, Pergamon Press, New ~Srk, second ed., 1964. [7] P . Benitez and J. C. Mi ~nano, "Ultrahighnumerical-aperture imaging concentrator," J. Opt. Soc. Am. A 14(8), pp. 1988-1997, 1997.
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394
DEVELOPMENT AND TEST OF AN EQUIPMENT TO REPLACE BROKEN GLASS ENVELOPS OF RECEIVER TUBES IN PARABOLIC TROUGH COLLECTORS K.-J. Riffelmann, M. B6hmer, T. Fend, R. Pitz-Paal, C. Spitta, Javier L6on* Solare Energietechnik, Deutsches Zentrum fiir Luft- und Raumfahrt e.V. (DLR), Linder H6he, D-51147 K61n, Germany, Phone: +49 2203 601 2415, Fax: +49 2203 66900, E-Mail: [email protected] *Plataforma Solar de Almeria, Carretem de Series s/n, Tabemas (Almeria, Spain)
Abstract: In 1984 the first large scale parabolic trough power plant was built up in the Californian desert. Cylindrical parabolic mirrors concentrate solar radiation on to a black absorber tube transforming radiation into heat. A thermal oil flowing through the tube transports gained energy to a heat exchanger, which is connectedto a steam generating circuit. To minimise thermal losses the absorber tube is enveloped by an evacuated glass tube. In practice glass envelopes often are damaged because of e.g. thermal stresses, which cause vacuum loss or fracture of the glass tubes. Consequently, thermal losses cause a significant decrease of the collector efficiency. It is not economic to directly replace destroyed modules, because many of them are welded together to built one loop. For repair the hole loop has to be shut down. Replacement is carried out only when a larger number of modules in one loop are destroyed. In this paper, an equipment is presented, which may be installed prior to replacement or repair of the whole loop to reduce thermal losses of broken receivers. This equipmehtwas tested on the DISS-reference loop at the Plataforma Solar de Almeria in southern Spain.
Introduction
From 1984 to 1991 nine solar power plants (SEGS ~ I IX) with an electrical capacity between 30 and 80 MW each were built up in the Californian desert. The total installed capacity is 354 MW, more than 7000 GWh clean solar electricity were produced until now. The overall average efficiency in one year (solar to electric power) is about 14 %, while more than 20 % peak efficiency can be reached (Geyer et aL, 1998). Main reason for the difference between average annual efficiency and peak efficiency is due to the tracking mode: The parabolic troughs are installed horizontally in north-south direction, tracking the sun by turning around the north-south axis. Therefore solar radiation is normally not perpendicular to the aperture area, which would give best performance. Nevertheless this one axis tracking mode is the most economic way to collect sunlight in parabolic troughs all over the year. Other possible reasons decreasing the annual average efficiency are imperfect tracking, pollution or breakage of some mirrors, damage of the receiver tubes (HCE's2). The HCE consists of a black absorber tube with a special selective coating absorbing well the solar specmma but having low emission in the infrared range to
1 SEGS: Solar Electricity Generating Systems 2 HCE: Heat collecting Element
minimize thermal radiation loss. The absorber tube is enveloped by a glass tube to prevent convective heat losses. In its original state the volume between glass and absorber tube is evacuated to minimize thermal losses caused by heat conduction and natural convection. But in practice the glass envelopes of some HCE's in one loop may be damaged because of e. g. thermal stresses, which cause vacuum loss or fracture of the glass tube. It is not economic to replace directly damaged modules, because many HCE's of about 4 meter length are welded together. Typically 14 to 16 collectors of 48 m length each are connected to build one loop. For replacement the whole loop has to be shut down. Idea of the developed equipment presented here is to replace broken glass envelopes or glass tubes that have lost the vacuum until repair of the whole loop. The equipment can easily be mounted into the field and can be reused after repair. In a minor scale the equipment was tested in 1996 in the ARDISS project (Ajona and Gonzalez, 1996), a parabolic trough collector with a length of 4.7 meters and an aperture width of 2.6 meters. First tests presented by Binner et al. (1996) showed a significant increase of the thermal efficiency. Now the equipment was tested on a LS-3 testcollector on the Platafonna Solar de Almeda, the collector type used in the SEGS plants in California.
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ISES Solar World Congress 1999, Volume III
Theoretical performance Aim of the developed equipment is to increase the efficiency by reducing thermal losses of the absorber tube as compared to a not longer evacuated or damaged receiver. Thermal losses effects of the different states of the HCE are described in this chapter.
Qo=
G
ea
+
Ag
A~(T~
-T~)
-1
where G: Stefan-Boltzmann Constant ~: Emissivity of the absorber eg: Absoprtivity of the glass envelop A: Surface of the absorber (index a) resp. glass envelop (index g) T: Absolute temperature of the absorber surface (index a) resp. inner side of glass envelop (index g) 1.2
Heat loss C~1 for HCE without vacuum
In this case additionally to the heat loss by radiation (~0 a convective heat loss is observed, that can be described by:
Q~ = t~o + oqAa (r~ - rg ) where t~: convective heat coefficient depending on air properties and geometry of the system Figure 1: Thermal and optical losses at the heat collecting element In general the solar radiation is transformed into heat by absorption on the black surface of the absorber tube, where the temperature is at maximum. The useful part of the absorbed energy passes the thin wall by heat conduction and is picked up by the thermal oil, which flows through the tube in a turbulent way. Another part of the absorbed energy is transferred from the absorber surface to the glass envelop by different mechanisms described below, depending on the state of the HCE. This part passes the glass wall and leads to convective heat losses from the glass envelope to the environment. 1.1
Heat loss (~0for the original HCE with vacuum In this case the only transport mechanism of heat from the absorber surface (index a) to the glass envelop (index g) is effected by radiation:
1.3
Heat loss (~2for HCE with completely damaged glass envelop Heat loss is effected by concevtive heat transfer and heat loss due to radiation:
(~2 = a2Aa(ra
-ro)+eacrAa(r~-r4)
where 0~2." convective heat coefficient depending on air properties, wind speed and direction (including natural convection what is important only for negligible wind speed) To: ambient temperature 1.4
Heat loss 0 a for the developed equipment
The equipment described in more detail in the next paragraph consists of two parts: One third of the circumference of the absorber tube, the part in the opposite of the parabolic mirrors that is not radiated by concentrated sunlight, is insulated by a ceramic insulation material:
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ISES Solar World Congress 1999, Volume III
g,f
--+-- ~-1 e. Ag (eg
with
9:
insulated part of circumference
Tg,f: Temperature of the free part of the
glass half pipe Tg, i: Temperature of the insulated part of
the glass half pipe heat resistant coefficient of the ceramic insulation as a function of material conductivity and geometry Based on these equations for the heat loss a simulation program was written. For the input data listed in Table 1 the simulation gives the results given in Table 2: k:
Figure 2- Scheme of the repair solution equipment The heat loss consists of two parts with different transfer mechanism: convective heat loss and radiation heat loss at the free part, conductive heat loss at the insulated part of the circumference:
Table 1: Input data for simulation Direct normal irradiance Ambient temperature Oil inlet temperature Wind speed Optical efficiency Reynolds number for oil flow
900 W/m 2 25 ~ 300 ~ 3 m/s 71.5 % 37500
Table 2: Calculated thermal efficiency based on the equations above with input data from table 1 Od~inal HCE 67,2 % HCE without vacuum 59,7 % Repair solution equipment 62,2 % Bare absorber tube 37,1%
2 Construction of the equipment The equipment for one HCE consists of two pairs of glass half pipes. Due to manufacture reasons each half pipe was only 1.85 m in length for the tested prototype, in series production the half pipes should have 2.00 m each. The resting free space of 0.3 meter per HCE of 4 meter length, which is 7.5 %, was insulated. For evaluation this part was not taken into account, which means that
an equilibrium of solar radiation input and thermal losses was supposed. The glass half pipes were fixed on half rings, the latter directly mounted on the absorber tube and made from aluminum. To reduce heat losses in the part opposite to the parabolic mirror receiving no concentrated solar radiation, a ceramic insolation was inserted. Due to this additional insulation the equipment should have lower thermal losses as compared to the HCE without vacuum.
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Figure 3: Photograph of the repair equipment, mounted on the test facility at the Plataforma Solar de A!meria possible changes of the collector (e. g. pollution of the mirrors, tracking modifications).
3 Description of the test facility The equipment was tested on the DISS reference loop, half of a LS-3 collector (48 m length) on the Plataforma Solar de Almeria, in October 1998. For a detailed description of the collector see Fend et al. (1999) in this volume. Firstly the thermal efficiency of the collector in its original state, that means: with vacuum in the HCE's, was measured. Secondly a little hole was drilled in the first six HCE's, which is the first half of the collector, to destroy the vacuum. The efficiency of this constellation was measured during one day. Thirdly the glass envelope of the six destroyed HCE's was removed completely and the repair solution equipment was mounted. Efficiency was measured during one week at different weather conditions. The test series was completed by measurements of the bare absorber tube after dismounting the equipmem. Note that during all tests the second half of the collector has been in its original state to observe
4
Test results
The thermal efficiency r/was calculated for both parts of the collector from measured temperature difference and volume flow with respect to the measured irradiance on aperture:
rl= with
cppVAT IA
Cp: heat capacity of the oil at average collector temperature p: density of the oil as a function of oil temperature Q" measured volume flow of the oil AT: measured temperature difference of the oil between inlet and outlet of the regarded collector part I: measured direct normal irradiance A: area of apemu'e of the regarded collector part.
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Figure 4: Measured data during one experiment as an example for efficiency calculation
As an example, the measured data for one day testing are presented as a function of the incident angle of the sun on the aperture in Figure 4. At this day (October 6, 1998) clouds coming up shortly after solar noon (which corresponds to an incident angle of zero degree) led to incorrect calculation of the efficiency due to thermal inertia of the system and the test was stopped.
The following Figure 5 presents the efficiency calculated from these measured data around solar noon (+/- 10 minutes, cloudless sky) of all tests done in this test campaign. Beside the points the mean collector temperature minus ambient temperature and wind speed are given.
609
290~ 956Wlm =
55 r-.-i
232~
E 'r-
9023 W/m = ".
990 ~'lm = 286~
266~
45
291~
&
V
o r-.
983 V rim"
931 W/m =
9 311 ~
40
-
962 tg/m"
1240~ 996 ~ Im =
3190C, 946 W/m"
0~ 50
o~ . to B
291~
317~
9 3120C, 945 W/m = 899 Vqrlm=
9 9 9 9
933 W'lm~_ 3120C, 8 ~2 W/m =
original HCE HCE without vacuum repair solution equipment bare absorber tube
35 9 319~
998 u rm"
30 280~
749 W/m =
25 0
1
2
3
4
5
6
7
wind speed [m/s]
Figure 5: Test results, thermal efficiency as a function of wind speed, temperature level and irradiance
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It can be seen from Figure 5 that the HCE without vacuum has a better performance than the developed equipment. Main reason for this unexpected result was a little gap existing between the glass half pipes due to imperfect manufacture. This gap resulted in convective heat losses. Another reason may be the limited length of the glass half pipes as described above. For the evaluation an equilibrium of the insulated areas between incoming solar radiation and thermal losses was supposed. In fact thermal losses seem to be greater. For low wind speed (< 1 m/s) the bare absorber tube has surprisingly good efficiency due to the minimised optical losses of the receiver (no reflection and absorption losses in the glass tube!).
5
Conclusion
An equipment for the replacement of broken glass envelops of HCE used in the SEGS plants in California has been developed, constructed and tested on the DISS reference loop on the Plataforma Solar de Almeria. Theoretical calculations show that the equipment should give better efficiency than a HCE that had lost its vacuum between absorber tube and glass envelop. In the experiments the thermal efficiency was lower. Main reason for this unexpected result was a little gap resting between the two glass half pipes due to imperfect manufacture. An improved construction was manufactured during the last month and will be tested in September 1999.
References
Ajona J., Gonzalez L., Evaluation plan for the ARDISS at the PSA, 2"d dratt, IER-CmMAT, Madrid, 1996 Binner
P.,
B6hmer
M.,
Hennecke
K.,
Development and Test of Advanced Parabolic Trough Receivers, in: Solar Thermal Concentrating Technologies, Proceedings of the 8th International Symposium, October, 6-11, 1996, Krln, Germany, Vol. 2, Verlag C. F. Miiller, Heidelberg (1997), p. 715
399
Geyer M., Holl/inder A., Aringhoff R., Nava P.,
H6lfte des weltweit produzierten Solarstroms, Sonnenenergie, 3, Deutsche Gesellschaft ftir Sonnenenergie e. V. (ed.), Solar Promotion GmbH-Verlag, Miinchen (1998), p. 33
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COOLING OF PV MODULES EQUIPPED WITH LOW-CONCENTRATING CPC REFLECTORS
Mats R6nnelid Solar Energy Research Centre, EKOS, Dalarna University, S-781 88 Bofl~nge, Sweden, Phone +46 23 778712, Fax +46 23 778701, email [email protected]
Bengt Perers, Bj6rn Karlsson and Peter Krohn Vattenfall Utveclding AB, S-814 26 )idvkarleby, Sweden, Phone: 46-26-83500, Fax: 46-26-83670, email: [email protected]
Abstract- A series of measurements on the performance of solar cell string modules with low concentrating CPC reflectors with a concentration factor C "4X have been carried out. To minimise the reduction in efficiency due to high cell temperatures, the modules were cooled. Four different way of cooling were tested: 1) The thermal mass of the module was increased, 2) passive air cooling was used by introducing a small air gap between the module and the reflector, 3) the PV cells were cooled by a large cooling fin, 4) the module was actively cooled by circulating cold water on the rear side. The best performance was given with the actively cooled PV module, which gave 2,2 times the output of a reference module while the output from the module with a cooling fin the power production was increased by a factor 1,8. Voltage drop due to high serie resistance and high operating temperatures limits the performance of the PV-CPC combination. PV-modules with lower serie resistance are required for good performance. Active cooling is interesting due to the possibility of co-generation of thermal and electrical energy. This increases the energy output and decreases the costs. Simulations based on climate data from Stockholm, latitude 59.4~ show that there is a good possibility for simultaneous production of heat at temperatures of around 60 ~ and electricity with a relatively high efficiency.
1. INTRODUCTION The current high cost of solar cells makes the use of reflectors for increasing the irradiation on to the PV modules cost effective since the cost of reflector materials is considerably lower than the cost for the modules. The use of booster reflectors is a technology which is of interest since a single booster reflector can increase the annual output from a standard module with 20-25% (R6nnelid et al., 1997). With more advanced concentrator designs, such as compound parabolic concentrators (CPCs), the annual output can be considerably higher (R6nnelid eL al., 1998). The efficiency of a PV module decreases with cell temperature, and for silicon solar cells this decrease is 0.4 - 0.5 %/~ (Green, 1992). A n increased irradiance on the module due to the use of concentrators implies an increased cell temperatRre and therefore decreased efficiency, unless the PV module is cooled. Investigation of techniques for efficient cooling of modules equipped with concentrators is therefore required. The high current density in the cells during
concentration requires that modules with very low serie resistance is used for limiting the internal voltage drop. The combined effects of high temperatures and high current gives an undesirable low fiU-factor FF. 2. EXPERIMENTAL SET-UP In this study, symmetrical CPC concentrators with a concentration ratio C = 4X are used. The CPCs have an acceptance half angle of 10 degrees, tnmcated to a height of 45 cn~ The modules used are string modules consisting of one row with 10 serial coupled X-Si cells, each cell with a size of 10 * 10 cm. The modules were specially designed by G~illivare Photovoltaics (GPV) in Sweden. To avoid large errors due to uneven illumination of different cells in the module during morning and evening hours, the reflectors were 2 m wide, which means that they are extended 0.5 m on each side of the PV module. A single PV module without concentrator was used as reference. The experiment was performed during the period July 1998 - June 1999. The experimental setup is shown in fig. 1.
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Four different ways of cooling the modules were tested: Prototype 1) The thermal mass of the module was increased by having a small tank with 4.6 kg water in close contact with the back of the module. Since the increased thermal mass delays the heating up of the module during the day, the maximum module temperature occurs during after noon hours. This is intended to increase the daily performance compared with a module with lower thermal mass. Prototype 2) Passive air-cooling was used by introducing a 2-cm air gap between the module and the reflector. This facilitates natural air circulation across the front surface of the module. Prototype 3) The PV cells were cooled by a large cooling fin. The cooling fin, originally manufactured for cooling electronic equipment, was glued on the back of the module with a heat conducting glue.
Figure lb. Experimental set-up of CPC concentrators with PV modules for testing different methods for cooling the modules. Photo from Vattenfall Utveckling AB, .~lvkarleby Laboratory, 12 June
1999, 9.00. 3. E X P E R I M E N T A L RESULTS
Prototype 4) The module was actively cooled. An identical water tank as for prototype 1 was used with the cold water continously circulated through the tank. A logger was used to monitor W-characteristics and module temperature every 10th minute. From these data, the maximum efficiency of the module was calculated.
Figure l a. Experimental set-up of CPC concentrators with PV modules for testing different methods for cooling the modules. Photo from Vattenfall Utveckling AB, ~dvkarleby Laboratory, 11 June 1999, 12.00. The anodised aluminium reflector has a concentration factor of C=4 and a half angel acceptance 0 = 10 ~
The different cooling techniques were compared by summarising the total electrical energy generated during the measurement period. This was compared with the energy generated by the reference module without concentrator during the same period. The result is summarised in table 1. The temperature shown in table 1 is the module mean temperature, which is the measured temperature, weighted by the generated power during the measurement period. It should be noted that all four modules were not monitored simultanously; therefore the energy generated by the reference module is different for the different cases. The result is also illustrated in figure 2, which shows the output from the concentrator modules as a function of the output from the reference module august 11. The temperature of the thermal mass module was 80 ~, of the air cooled module 60 ~, of the reference 40 ~ and of the actively cooled module 15 ~ around noon that day. The cooling fin was tested after the september 1 during relatively low ambient temperatures and resulted in a temperature around 20 ~ above the temperature of the reference module. The temperature of the module for active cooling reached 100 ~ when the water cirqulation was stopped during hours of high intensity. This resulted in a large voltage drop and a very low efficiency.
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......... -- -- ..... ----------
Table 1. Summarized result from measurement of
cooled concentrator modules during the period July 19 - September 30, 1998. I , , , I , , ,
30 I-
Concentrator Reference module module Energy Tm~n Energy Tm~
Energy No. of ratio measurement
[kWh] [oc]
[kWh] loci
day Thermal mass 4.27 47.2 2.90 Air cooling 2.85 53.2 1.95 Cooling fin 2.58 38.8 1.41 Active cooling 3.32 18.7 1.51
,
300
_e O
E
250 200-
,
,
I I
n + o o
,
,
,
32.7
I I
,
1.47
33.5
1.46
37
32.2
1.82
29
36.7
2.20
21
.
.
I I
,
,
,
I I
cooling fin air cooling active cooling thermal mass
.
,
,
I I
,
,
,
o~ ~ Q
..-,:,
100~
50--
0
0
......... 0
.....-'"
I,.
I
,,I,,,
20 40 60 80 100 Output from reference [Wh/day]
I , ,
,
t
I
i
J~P"F'~ ~ : : ~ c : , ' - : : ; P - - ~ ".\ ,
t
t
t
I
::I
....
"
.I...I...I 8
10 12 14 Time [hour]
16
18
Figure 3. Output for some different concentrator modules during a clear day. Data from August 11, 1998.
~ N 150-
E .I~
,
20 T
6
o
t-
i , , ,
/
o
O
O tO
I , , ,
r, N
25 T
"5 10
60
,
Thermal mass Active cooling Air cooling Reference
120
Figure 2. Input/output diagram showing the daily output from the concentrator modules vs. t h e output from the reference module. All monitored days are included in the diagram. Table 1 and figure 2 show that the output from the actively cooled module is highest and the module with a cooling fin works relatively well. The output for the two modules with large thermal mass and aircooling is considerable lower.
The output from the concentrator modules is low compared to the relatively large concentration ratio of the CPCs. The acceptance angle is relatively narrow, thus giving undesirable reduction of the diffuse radiation in the near specular direction. The small acceptance angle also caused the collected radiation to be collected at incidence angles near the acceptance angle at morning hours or aitemoon hours, as visible in figure lb. This brings the reflected radiation to be concentrated in regions at the edges on the module near the reflectors. The concentration of radiation on a small area of the module implies large local electrical currents which reduce the output from the module due to the serie resistance and high temperatures, which can be seen as a local minimum in the output during the hours round noon (figure 3). Therefore, slightly larger acceptance angles is suggested. It was concluded that the water content was too small in the concentrator module with a thermal mass since the temperature of the water was increased by 40~ during less than two hours during a day with full sunshine. However, if the thermal mass shouls 2-5 times larger, the temperature rise would be considerably slower, which would increase the output compared to that measured during clear days. The module with air-cooling cannot be essentially improved. The performance of the module with a cooling fin shows that there is a possibility of achieving rather low module temperatures with passive cooling. However, in this experiment a commercial cooling fin was used, and an
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economically more reliable alternative would be to use the existing reflectors as cooling fins. Besides the saving the costs for a cooling fin such a construction should simplify the construction.
3.1 Hybrid P V-CPC thermal systems Active cooling of PV systems is interesting due to the possibility of co-generation of thermal and electrical energy. In the experiment performed, the mean temperature of the module was too low, about 19~ to be of interest for solar heating application. However, with slightly higher temperatures, the heat can be of pratical interest, for example for pool heating, for low-temperature heating systems or for pre-heating of domestic hot water. There are some important reasons why cogeneration of electricity and thermal energy are of interest. Today, the cost of PV-produced electricity is very high, in the order of 0.50-1.50 S/kWh (Bevan 1997). This can be compared, for example, to 0.04-0.10 S/kWh for land based wind power in the same report. Long distances from the electric grid this cost for PV electricity is still competitive in a large number of niche applications. It is now approaching the costs for power production with small diesel generators, if costs for operation and fuel transport are included. This can lead to a large expansion of the PV market. But to reach levels that will influence the global energy balance, grid connected PV systems are required. Unfortunately the present PV technology is far too expensive to compete in the grid without heavy subsidies. Even if the PV modules are free the inevitable costs for design, installation, power management equipment, wiring, and support structure will only allow the costs to be roughly halved (Leng 1996). Most of this is due to the comparatively low net energy output per area unit of PV cell area. Radical technology changes are therefore needed to make PV cost effective in grid applications. With cogeneration of heat and electricity, the total efficiency of the system will increase, and thereby the generation of heat can, to a larger or smaller extent contribute to the installation cost, giving better economy for the electrical fraction. Another reason for co-generation is that there is an obvious risk that there may be a lack of available space for solar applications. Today several schemes for large building integrated PV installations are under discussion. In the European Community, 500
403
000 roof-integrated grid-connected systems should be installed before the year 2010. Similar large-scale plans are discussed for Japan and USA. If these and other plans for building-integrated PV-systems will be realised, the systems may occupy a larger part of the available roof (or facade) and it may be difficult to find space for parallel solar thermal installations on these buildings, unless heat and electricity are produced simultaneously by the same installation. The use of concentration devices, similar to those previously discussed, for co-generation of heat and electricity has important advantages. With concentration, the heat losses can be kept at a low level, which is important for high thermal efficiency. Although it can be assumed that PV module prices will fall in the future, concentrating devices reduce the module area and thus reduce the cost for PV modules. For efficient cooling of the PV modules it is also very important that there is a good thermal contact between the total module area and the cooling fluid, otherwise there may be "islands" on the PV module with high local temperatures which destroy the overall effect of cooling. Efficient cooling of a module is simple if the module is small, and with reduced module area due to concentration, the piping cost for the thermal part of the equipment is reduced. The experiments, which started by investigating different cooling techniques for PV modules equipped with concentrators, will continue with measurements on a hybrid PV thermal concentrating system during the summer of 1999. The geometry will be similar to that shown in figure 1, but with an anti-reflection glazing on top of the concentrator to protect the reflector and PV module and to reduce the heat losses. The modules will be cooled by water circulating in 10 mm aluminium plates tightly glued to the back of the modules, which is expected to give a homogenous temperature on the module area. The amount of electricity needed to circulate the cooling fluid is estimated to correspond to approximately 2 % of the annually produced electricity from the PV module (Karlsson et. al., 1998). Figure 3 shows a simulation of the expected annual performance of such a system, based on Stockholm's (lat. 59.4~ climate. The module parameters used are C = 4X, tgl~g = 0.9 rreflector= 0.9, a~bsoeo~r= 0.9, U L -" 3 [W m E KI]. The PV module efficiency is estimated to 10% at 1000 W irradiation and 25~ cell temperature. Although these parameters may be somewhat different in a real installation, the results
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indicate that there are good prospects for producing useful temperatures of the cooling fluid with only a slightly reduced performance of the electrical fraction of the PV/thermal hybrid system. 4. CONCLUSIONS 9Different methods of cooling PV modules equipped with CPC-reflectors were investigated. Table 1 and figure 2 show that the output from the actively cooled module is largest and the module with a cooling fin has the second best result. The output for the two modules with large thermal mass and air-cooling is considerable lower. 9 Hybrid PV/thermal systems are of interest for reducing the cost of PV generated electricity. To a large extent this has to do with the low efficiency of conventional PV systems, which implies large areadepending costs. For good performance it is important that there is a good thermal contact between the total module area and the cooling fluid.
REFERENCES Bevan, G. " Key Issues in Developing Renewables". IEA 1997 ISBN 92-64-16009-4 (6versikt 6ver alia f'rmybara energik/illor och nuvarande och framtida priser) Green, M. A. , (1992)Solar C e l l s - Operating Principles, Technology and System Applications, University of New South Wales. Karlsson B., Krohn, P. and RSnnelid M., Solar cells with reflectors - result 1998 (Solceller med reflektorer - Resultat 1998). VattenfaU Utveckling AB, Ny Energi- & Miljfteknik. Rapport nr UR 98:21. In Swedish. (1998). Leng, G., Dinard-Bailey, L., Tamizhmani, G., Usher, E. et. al. "Overview of the Worldwide Photovoltaic Industry" EDRL 96-41-A1 (TR). CANMET, Varennes, Quebec, June 1996. Rfnnelid, M., Perers, B., Krohn, P. and Karlsson, B. Booster reflectors for photovoltaic modules at high latitudes. Proc. North Sun 97, Helsingfors, Finland (eds. P. Konttinen och P. D. Lurid), 555-562. (1997). Rfnnelid, M., Perers, B., Krohn, P., Spante L. and Karlsson, B. Experimental performance of a string module in a CPC reflector cavity, Proc. 2nd World Conference on Photovoltaic Solar Energy Conversion, Vienna, Austria, 2264-2267 (1998).
Acknowledgements This work has been carried out in co-operation with Vattenfall Utveclding AB within the framework of the national solar electricity programme SOLEL 9799 financed by NUTEK, the Swedish utility industry and NAPS Sweden AB. Figure 3. Expected output from a Hybrid PV/thermal solar collector. Collector parameters are defined in the text. Climatic data from Stockholm, lat. 59.4~
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A SOLAR BOWL IN INDIA Gilles Guigan and John Harper Center for Scientific Research, Auroshilpam, Auroville, 605101, Tamil Nadu, India Tel: +91.413.622.168, Fax: +91.413.622.057, email: [email protected]
Sylvie Rousseau Solar Consultant, 111 rue de la Station, Bruxelles, 1150, Belgium, Tel: +32.2.763.27.04, email: [email protected] Abstract- This paper presents a solar concentrating device, a <~solar bowl >~of 15 meter diameter, presently under construction in India. This system is used for community scale steam cooking. The project intends to assess concretely the interest of the solar bowl technology in India. Therefore, the bowl has to be an operational tool, as cheap and simple as possible to be realised and operated in an Indian environment. The design of the bowl, was adapted to these prerequisites. The size of the bowl was defined in order to cover the full steam requirements to cook 1000 Indian meals twice a day, but a diesel fired boiler is included in the loop to complement the solar steam production in case of a bad wheather. The main innovation is the realisation of the reflective surface, made of small flat mirrors glued on a cement shell. The paper describes extensively the problems, their solutions and the procedures defined to set up properly the 10 000 mirrors. The solar boiler contains a thermal fluid heated up to 250~ and sent to a heat storage tank. Steam is produced in a heat exchanger placed inside the storage tank. The boiler design is described. The boiler supporting arm and its tracking system are presently under installation and not detailed here. The paper concludes with some hnts about steam cooking.
The <<Solar Kitchen >>that is being built in Auroville
1. INTRODUCTION This paper describes a solar bowl that is being built in Auroville, near Pondichery in India. This bowl project has been initiated and realised by the CSR (Center for Scientific Resarch), an Indian research institute based in Auroville (Tamil Nadu) working essentially on renewable energies and appropriate technologies. This project has been financed by the
Government of India under the authority of the Ministry of Non conventional Energy Sources.
2. CHOICE OF THE SOLAR BOWL TECHNOLOGY The solar bowl is one of the wellknown thermal solar concentrating technology, but it has not been developed as much
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as the solar troughs, solar dishes and solar tower. A solar bowl is a fixed spherical cap covered with mirrors, it concentrates the sun rays on a radial line moving with the hour and date of the year. A boiler is hanged inside the bowl and tracks the sun in order to always be situated on this focus line.
A few bowls have been constructed in the 80's: - a small bowl of 3.5 m aperture diameter in Auroville, - a 20 m diameter bowl in Texas (Reicher, 1982), - a 10 m diameter bowl, called Pericles (Authier, 1982), built by a team from the CNRS (Centre National de la Recherche Scientifique) in Marseilles (France). It was later dismantled and rebuilt in Recife (Brasil) where it worked satisfactorily for 2 years. These two last bowls were extensively studied and tested. The conclusion was that the bowl technology could be advantageous in developing countries, especially because the low cost of civil work would allow the construction of a cheap spherical cap (Lodhi and O'Hair, 1982). But the proof was still to be made, the CSR wanted to try it out.
3. DESIGN AND SIZING OF THE SOLAR B O W L 3.1. Choice of the application At that time, the CSR was involved in the construction of a collective kitchen using steam for cooking. We decided to merge the two projects and to install a solar bowl on top of the kitchen building to provide the steam. Therefore, we had the opportunity of an application using directly the heat produced and functioning in actual conditions. This bowl had to be cheap and realised only with Indian equipment in order to be replicable afterwards, but it also had to be an operational tool, easily controlled and reliable with a reduced and simple maintenance. This implied solutions both simple and largely automated ... which was a real challenge for the equipment and for the staff as well. 3.2. Designing and sizing the solar bowl Other existing collective kitchen using steam cooking could give us the amount of steam used to cook one I ndian meal. Our collective kitchen is supposed to cook 1000 meals twice a day. We wanted to be able to provide the totality of the required steam with the solar system on clear days, but we also decided to complete the production of steam with a conventional boiler in case of a bad weather. People should be able to eat everyday, even during the monsoon! We designed a bowl of 15 meter diameter, 120 degrees aperture angle. It should supply the cooking pots with 600 kg of steam at 110~ 1.2 bars. This evaluation is based on the results of the few bowls previously built around the world, in correlation with
their size and the solar radiation available on the site. The bowl was nicely integrated in the roof of the kitchen, it is tilted of 12~ (latitude of Auroville).
Geometrical characteristics of the bowl concentrating surface - Aperture diameter: 15 m - Aperture angle: 120 ~ - Aperture area: 176.6 m 2 - Curve radius: 8.66 m - Tilt towards the South: 12 ~
4. REALISATION OF THE REFLECTIVE SURFACE In order to have an operational tool, we used as much as possible proven solutions; but we had to innovate to realize the reflective surface. In the other bowls the reflective surface was made of curved mirrors fixed on a metallic frame. In our case, we chose a different solution in order to get a minimum price in a rural Indian environment. The reflective surface is made of small flat mirrors fixed on a spherical cap built with civil work. This solution had not been applied anywhere before, so this is the more innovative part of our project. We had to define and experiment original solutions for the equipment and procedures to be used on the site to realise a reflective surface of a high quality. The main research work was to transform theoretical solutions into concrete realisations. Therefore we made a lot of tests beforehand but our work benefited essentially of the CSR experience in realising appropriate technology projects in rural India.
4.1. Making of the shell A structure made of compressed earth bricks stabilised with cement supports a spherical shell cap made of prefabricated ferrocement elements. Then this shell is covered with a cement plaster and small flat mirrors are glued on it. 4.2. Mirror size We tested different size of mirrors, and finally chose a base of 15 cm x 15 era. This leads to a reasonable number of 10 500 mirrors and an acceptable angular error due to the fact that the mirrors are flat instead of being curved. In fact, they are not squares but trapezoids of decreasing surface from near-squares of 225 cm 2 at the edge of the bowl down to triangles of 140 cm 2 at the bottom. 4.3. Mirror protection We had to undergo a lot of experiments to find out how to avoid the damaging of the silvering by the glue or the ions
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released by the cement. We finally decided to protect each mirror back by gluing a glass sheet of the same size.
Mirrors 10 500 mirrors Trapezoids of 140 to 225 cm 2 (from squares of 15 cm side at the edge to triangles of 10 cm side at the bottom) Float glass mirrors 3 mm thick glued to a 2 mm thick glass sheet Mirror protection
So, two stages of preparatory work are necessary before setting up the mirrors on the cement shell: 1) cutting the glass and mirror sheets in precisely defined trapezoids 2) gluing together with araldite a mirror and its glass protection 4.4 Setting up the mirrors Then these couples glass+mirrors are set up one by one on the cement shell. We had to define a procedure in order to place them precisely in their right position and orientation.
The correct position of the mirror on the shell is defined with the help of cardboard lines glued provisorily on the cement The orientation is correct when the center of the mirror plane faces exactly the virtual center of the bowl spherical surface. This is obtained with the help of a device fixed on a pivot situated at the virtual center of the sphere. This device is composed of a small laser pointer and a circular white target on either side of the sphere center. The target center is symmetrical to the laser emitted dot. The laser beam is pointed towards the center of a mirror. If the mirror is properly orientated, the reflected laser beam hits the center of the target. The following five steps have to be applied to set up each mirror: 1) positioning the laser beam 2) putting 6 spots of silicone glue on the back of the couple mirror+glass 3) gluing the couple mirror+glass in its exact place on the cement surface 4) checking the reflected laser spot on the target 5) pressing lightly on the mirror in order to shit~ this spot to the center of the target. We applied this procedure 10 500 times, we got an average value of 3 angular minutes for the precision of the mirror orientation.
407
A team of 4 persons was necessary to set up 20 mirrors an hour. Since it is too hot to work inside the bowl during daytime, a trained team could set 100 mirrors everyday from 5 pm to 10 pm. For 10500 mirrors, we needed 105 working days.., it actually took almost six months. It is obvious that such a solution would be a nonsense in a developed country, but it is indeed quite adapted technically, economically and socially to rural India.
5. THE HEATING SYSTEM The heating system is composed of two loops : - a primary loop containing a thermal fluid called Therminol 66, including the solar boiler and a heat storage tank, able to store one hour of peak steam production, a secondary loop containing water. Steam is produced in a heat exchanger situated inside the heat storage tank and then transported to the utilities. This loop includes also a diesel-fired boiler to complement or replace the solar boiler steam production in case of clouds, especially during the monsoon period. -
The solar boiler is supposed to heat a variable flow of Therminol 66 from 150~ to 250~ Its design derives closely from the boiler used in the Pericles bowl. It is composed of a cylindrical low concentration part topped by a high concentration conical part. A strap of 3 small parallel pipes is wrapped around a supportive sleeve. The shape and dimensions of the boiler have been optimized with the aid of a statistical method applied to the reflected sun rays. The diameter and number of small pipes are determined by the fluid temperatures and flow rates inside the boiler. These calculations are based on the studies made by the Pericles team: an optical simulation of the collector (Authier, 1979) and a modelisation of the thermal fluid flow (Pouliquen and Authier, 1979). The boiler (and its fluid feeding pipes) is attached on an arm hold by a pivot fixed at the virtual center of the spherical surface of the bowl. The whole is supported by a strong monopod standing inside the bowl, away from the boiler trajectory. The pivot allows the tracking of the sun. The arm is pivoting in the North-South and East-West directions to follow the sun movement during the day throughout the year. The tracking system will be driven by a step by step motor and controlled by a computerised system. 6. SOME HINTS ABOUT STEAM C O O K I N G Cooking with steam is fast, clean, safe, healthy and more energy efficient than any burner heating only the bottom of the pot. Two types of containers are used:
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- double-jacketed vats where steam condenses between the two jackets and is then sent back to the exchanger to be recycled. They are used to cook vegetables, lentils or noodles. - steam ovens where the steam is set free inside the oven and not recycled after use. These are preferably used to cook idlis (a popular Tamil breakfast) or rice in large cylindrical vessels containing the rice and its cooking water.
7. P R E S E N T
STATE OF THE WORK
The collective kitchen is functioning since December 1997 with steam entirely provided by the diesel-fired boiler. The reflective surface of the bowl is completed. The monopod and the solar boiler have been manufactured, they are now ready to be assembled. The tracking system is in its progrzmming and testing phase. The bowl should be operative in October 1999. REFERENCES
Reicher J.D. (1982) The Crosbyton Solar Power Project : Fixed Spherical Mirror / Tracking Receiver, Department of Electrical Engineering, Texas Tech University, Lubbock, Texas. Authier B. (1982) Le collecteur sph6rique fixe Mini-Pericles.
Revue Internationale d'H~liotechnique, 1er semestre 1982. Lodhi M.A.K. and O'Hair E.A. (1982) The solar bowl technology transfer to developing nations: a case study of Pakistan. Texas Tech University, Lubbock, Texas. Authier B. ( 1 9 7 9 ) Optical simulation for a fixed spherical solar collector. Applied Optics, Vol 18. N ~18. Pouliquen D. and Authier B. (1979) Mod61e d'6coulement fluidique dam les chaudi~res ~ veine h61ico'idale, Revue de Physique Appliqu~e, Janvier 1979, Tome 14, pp 91-95.
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THE DEVELOPMENT AND TESTING OF SMALL CONCENTRATING PV SYSTEMS George R. Whitfield, Roger W. Bentley, Clive K. Weatherby and Alison Hunt Dept. of Cybernetics, Univ. of Reading, Whiteknights, PO Box 225, Reading, RG6 6AY, U.K. Tel + 44 118 931 8223, Fax + 44 118 931 8220, e-mail [email protected]
Hans-Dieter Mohring, Fritz H. Klotz and Peter Keuber ZSW, Hessbruehlstr. 2 l c, D-70565 Stuttgart, Germany. Tel + 49 711 7870 272, Fax + 49 711 7870 230, e-mail [email protected]
Juan Carlos Mifiano and Elisa Alarte-Garvi I.E.S., Universidad Politecnica de Madrid, E.T.S.I. Telecomunicacion, Ciudad Universitaria, E-28040, Madrid, Spain. Tel + 3491 336 7222, Fax + 3491 544 6341, e-mail [email protected] Abstract - Spreadsheets have been used to compare some 90 possible small PV concentrator designs that might be
suitable for use at remote sites. They have apertures of about 2m 2, use BP Solar LBG cells, and employ small aperture modules to reduce heat sinking and construction costs. Designs include fixed V-troughs and CPC's, single axis tracked cylindrical lens and mirror systems, and 2-axis tracked spherical-symmetry systems. Performance and volume production costs were estimated. Four promising systems were constructed as prototypes. A- Point-focus Fresnel, 2-axis tracking; Cg = 32x; and 69x with secondaries. B - Line-focus mirror parabolic trough, 1-axis tracking, Cg = 20x. C - SMTS ('Single mirror two-stage'), 1-axis tracking, Cg = 30x. F - Multiple line-focus mirror parabolic trough, E-W 1/day manual tracking, Cg = 6x. The prototypes were tested at Reading, and three for up to a year's field trial at ZSW's test site, Widderstall, in Germany. The best module efficiencies, normalised to 25~ and excluding the end losses of linear systems, were: 12.5%, 13.2%, 13.6% and 14.3% for collectors A, B, C, and F, respectively. The tests have shown the practicality and robustness of the designs, and the performances of collectors B, C and F are only 10% below the estimates in the spreadsheet calculations. The best of the collectors have costs in the region of 1.5 to 1.8 $/Wp, yielding energy costs at a good site (excluding BOS and overheads) of between 5 and 7 cents/kWh. A conventional PV array costs 4.3 $/Wp, and 18 cents/kWh.
1. INTRODUCTION Photovoltaic systems have advantages as sources of small amounts of electrical power in remote areas, but conventional solar panels are expensive. Since lenses and mirrors in volume cost only about 1/20 as much as solar cells, it should be possible to reduce the cost of PV electricity by using them to concentrate the sunlight from a large area onto a small area of solar cells. Many concentrators have been developed in the past, but they have usually been not much cheaper than conventional solar panels, because concentrator solar cells have cost much more than one-sun cells and the optical and tracking systems have been expensive. Recent developments, such as BP Solar's Laser Buried Grid cells (Mason, Bruton and Heasman, 1995), have made it possible to manufacture solar cells little different in design and cost from one-sun cells that can be used at concentration ratios up to 40x. Using these cells with optical and tracking systems that are no better than required there is considerable scope for cost reduction. This JOULE III project built on the progress made in a previous JOULE II project, EUCLIDES (Sala et al., 1997), led by BP Solar and joint with UPM, ZSW and Reading University, which developed a PV concentrator for large grid-connected systems. A 480 kWp demonstration plant, based on the EUCLIDES work, has been built in Tenerife (Sala et al., 1998). In the present project the objective instead has been to reduce the cost of PV electricity by developing small concentrating systems of about 2 ~ aperture designed for use in remote areas. Specifically, the aim has been to produce a small number of prototypes of such systems sufficiently near practical
production to be of interest to industrial companies. We have examined a wide range of possible concentrators using BP Solar LBG cells. For each system, performance and cost have been estimated on common basis, assuming large-scale production. No such comparison of a wide range of systems been made for many years. 2. THE COMPARATIVE ANALYSIS OF A WIDE RANGE OF POSSIBLE CONCENTRATING SYSTEMS Insolation data were assembled for three representative sites, and a database of materials costs was generated. Using these data, some 90 possible small PV concentrator designs suitable for use at remote sites were then compared (Whitfield et al., 1997, 1998). The systems envisaged have apertures of 2 r~, use BP Solar LBG cells, and are composed of smaller aperture modules to reduce heat sinking and construction difficulties. Designs considered included: - fixed V-trough and CPC systems, - cylindrical single-axis tracked systems up to 30x, - spherical-symmetry 2-axis tracked systems up to 69x, - secondary optics, and a variety of manual and automatic tracking strategies. The performance of each system was calculated on a spreadsheet, taking due account of its dimensions, the properties of the optical components used, the efficiency of the solar cells, and the energy capture of the tracking strategy chosen. Each system was then given outline mechanical design, bearing in mind ease of mass production and long service fife. Using this mechanical design, each system's material cost was calculated, and with advice from a consulting engineer, the manufacturing
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cost, in mass production, was estimated. The total cost was combined with the estimated performance to give the resulting cost per peak watt, and per kWh. Typical results are given in Table 1, which shows the cost per peak watt and per kWh for the best of the collectors considered and a few others. The main conclusions from the analysis are:1. Concentrating collectors earl be much cheaper than conventional planar collectors, by a factor between 2 and 3. 2. To obtain this improvement, they must be made in large numbers. 3. There is a wide range of good designs. The best have relatively high concentration ratios and imaging primary optical systems. 4. Mirrors are usually more cost-effective than lenses.
5. Secondary optical elements often improve the performance. 6. Automatic tracking systems are better than manual tracking, in spite of their extra cost. 3. SELECTION OF COLLECTORS FOR PROTOTYPES
From the information in section 2, the project group then unanimously (!) selected six collectors, A to F in Table 1, for more detailed analysis; of these six, four collectors were built as prototypes. Selection criteria included not only low cost, but also innovative optics (SMTS) (Alarte, Benitez and Mifiano, 1998), and innovative tracking (once per day movement). Calculations were made to optimise the parameters of each
Table 1. Results for the best collectors analysed, and a few others. The last line is a conventional planar array Letter for Prototype A A B C B
D
D E
C G F F
Secondary Optics
Mounting
Cone Ratio
Cost $AVp
Point focus Fresnel lens Point focus Fresnel lens Weatherby's Cylindrical Paraboloid SMTS Collector, Plastic, 0.6m Cylindrical Paraboloid: Multiple offset Cylindrical Paraboloid Cylindrical Paraboloid Cylindrical Paraboloid: Multiple offset Cylindrical Paraboloid Linear Fresnel lens Cylindrical Paraboloid Curved TIR Lens Cylindrical Paraboloid
Pt-focus solid CPC No No
Gimbals Gimbals Polar
69 36 19
1.46 1.48 1.62
12.1 12.2 14.0
6.2 6.3 7.2
5.4 5.4 6.2
Yes No
2-axis Polar
30 20
1.78 1.78
14.7 15.4
7.6 7.9
6.6 6.8
No Point-Focus CPC Mirror CPC
2-axis Polar Polar
20 65 27
1.95 1.78 1.88
16.1 16.2 16.3
8.3 8.3 8.4
7.2 7.2 7.2
Solid CPC Solid CPC No No Mirror CPC
Polar Gimbals Polar Polar 2-axis
37 37 20 28 25
1.90 2.02 1.95 1.97 2.06
16.4 16.7 16.8 17.0 17.0
8.4 8.6 8.6 8.7 8.8
7.3 7.4 7.5 7.6 7.6
SMTS Collector, Alum, 0.3 m SMTS Collector, Alum., 0.3m: Alfilm Cylindrical Paraboloid
Yes Yes
Polar Polar
30 30
2.00 2.01
17.3 17.3
8.9 8.9
7.7 7.7
Oil filled CPC
Polar
37
2.04
17.6
9.1
7.8
Point-Focus CPC No No No
Chinese Polar Polar Polar
50 15 30 2
1.72 2.18 2.39 4.31
19.5 18.8 21.6 23.7
9.6 9.7 10.6 13.9
8.3 8.4 9.2 12.7
Mirror CPC
E-W axis 1/day E-W axis 1/day Fixed at latitude
8
2.52
31.4
15.8
13.6
6
2.64
32.8
16.6
14.2
1
4.31
30.5
19.6
18.1
Primary Optics
Cylindrical Paraboloid Curved Fresnel lens SMTS Collector, Plastic, 0.6m V-trough, Screen Printed Single Crystal Cells Cylindrical paraboloid: multiple offset Cylindrical paraboloid: multiple offset Flat, Screen Printed Single Crystal Cells
No No
Cost Cents/kWh Wid. Man. Aim.
Note. The costs given in the table are for cells, optical systems, mountings and trackers only; balance of system costs are omitted as they are similar for all types of collector. The cost in $/Wp is based on direct beam radiation at 850 W/m2. The cost in cents/kWh is site-specific; the three columns are for Widderstall, near Stuttgart, a relatively cloudy site, Manfredonia in Italy, and Almeria in Southern Spain, a particularly good site.
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collector, particularly the concentration ratio and aperture, to maximise the performance/cost ratio. In the detailed design, attention was paid to ease of large scale manufacture; for example cylindrical parabolic mirrors could be pressed from aluminium sheet or moulded in plastic, and lenses could be made by a rolling process. But the press tools and moulds were too expensive for these techniques to be used for our prototypes. So the prototypes were built with supporting ribs or glass fibre members to maintain the optical shapes. Care was also taken to use techniques and materials that would ensure a long working life and freedom from corrosion.
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the cost of production systems. So commercial systems were purchased from the USA. Collector A uses a 2-axis active tracking system from Wattsun, with a sensing head on the inner gimbal driving right ascension and declination servos. Collectors B and C use an open loop microprocessor-controlled system, developed initially by Maish of Sandia (Maish, 1991), and marketed by Enhancement Electronics; this drives a Wattsun servo, rotating the collector about a polar axis. The systems were reliable and weatherproof, but were stronger, heavier and more expensive than was required for small systems. The microprocessor-controlled systems were difficult to set up, but ran well once this had been done.
3.1 The choice of mirror surface For the reflector surface of the three mirror collectors (13, C and 10 we essentially had the choice between four materials; where the costs shown are our assumed very large order prices. 3M's ECP305+ silver film, exterior grade; p ~=- 0.94; $ 25/m z 3M's SA-85P aluminium, exterior grade; p ~-- 0.85; $ 7/m2 3M's SS-95P silver film, interior grade; p ~=- 0.94; $ 9/m 2 Anocoil anodised aluminium; p ~=- 0.84; $ 20/m 2 Note: The first three minors are reflective film, so the cost of substrate and laying down must be added; this will be typically $12/m 2. In collectors B and F the parabolic troughs are covered with glass or acrylic sheet, so the reflective surface need not be especially weatherable. 3M's ECP305+ was difficult to obtain, as 3M have ceased manufacture, but UPM, via Mifiano, kindly supplied a roll for use on collector C, where the reflector is in the open air. A comparison of ECP305+, SA-85P and SS-95P was carried out, using the cost/performance spreadsheets for collector B; for this comparison, the reflectivities were reduced to 0.90, 0.81 and 0.88 respectively, to allow for mirror profile errors. The annual energy cost, in c/kWh, was highest for ECP305+, marginally less for SS-85P, and about 7% less for SS-95P. Anocoil anodised aluminium was not include in this analysis, as it was more expensive than SA-85P, which has a higher performance. Therefore for collectors B and F we used the SS-95P interior grade film. This is only supplied pre-glued to 0.5 mm thick aluminium sheet. This had the advantage that we did not have to do the laminating ourselves, but 0.5 mm is thinner than we would have liked for heat-sinking. 3.2 Heat sinking Heat sinks for the cell strings were designed to hold the cells at temperatures similar to those of conventional flat arrays, typically 30-40 ~ above ambient. This requires an effective cooling area approximately twice the projected area of the array; for a flat array this is just the front and rear surfaces; for collectors A, B and F, there were suitable aluminium surfaces forming the structure or mirror surface of the collector. Studies carried out in the previous (EUCLIDES) project (Whitfield et al., 1997) and thermal tests on small mock-up collectors showed that adequate cooling would be obtained with the 0.5 mm aluminium sheet if the aperture was not above about 15 cm. So this aperture was chosen for collectors B and F. 3.3 Trackers Time did not permit the development of custom-built trackers for each collector, although they will be necessary to reduce
3.4 Solar cells The solar cells used for the project were BP Solar 'Saturn' laser buried grid cells (Mason, Bruton and Heasman, 1995). These cells, developed from work by Prof. Green of the University of New South Wales, Australia, can be made for little more than the cost of conventional screen-printed one-sun cells, but have inherently higher efficiency and lower series resistance; this makes them particularly suitable for moderate concentration ratio concentrators such as ours. The cells were cut to length 50 mm and widths appropriate for each collector, retaining a full length bus bar down one side. They were connected in strings with conventional tabbing strips, using thicker than usual material to keep the resistance down. The strings were mounted on machined aluminium carriers, using a thermally conducting electrically insulating tape, and encapsulated in transparent silicone, with glass or Tefzel covers. This technique was satisfactory, but was rather slow and messy. 3.5 End losses of cylindrical collectors. Collectors with cylindrical optics, such as linear Fresnel lenses or cylindrical paraboloid mirrors, concentrate the light into a line focus; the solar cell string is mounted along this line, and the collector is rotated about an axis parallel to this line to keep the focal line on the cell string. Thus a single set of bearings and a single tracker suffice.
Figure 1. Shading loss: The cell string has to be shorter than the collector If, as is common, the chosen axis points at the pole, then tracking is perfect at the equinox, when the sun's declination is zero; the sun's rays are perpendicular to the axis of the collector, and the cell string can be as long as the minor. But at any other date, when the declination is not zero, the focal line is displaced towards or away from the pole. The worst case is at the solstice, when the declination ~5reaches its maximum value of 23.5 ~ Then in a typical case, Figure 1, a length 2h tan8 of the cell string is not illuminated. At the other equinox, an equal length is shaded at the other end of the cell string. Since one shaded cell blocks the whole output current of the string, it is usual to reduce the cell string length to d - 2 x 2h tan~, so that the whole of the cell string is always illuminated. The output of the collector is reduced proportionately. This end loss has been
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computed for each collector and included in the spreadsheet analyses. With long narrow collectors the loss is small, but with wider apertures it becomes significant. For collector B, a narrow offset paraboloid, the correction for end losses is a factor of 89%, but for collector C, the wide SMTS collector, it is 65%. This end loss can be eliminated by using a gimbal mounting and two-axis tracking, but this adds cost and mechanical complication. But a partial solution is to use a full length cell string and bypass diodes, Figure 2. Suppose that there are n cells in the centre part of the string that is never shaded, and m cells at each end that are shaded at the solstice. Then a full string is n + 2m cells, and the correction for end losses if the shaded cells are omitted is n/(n+2m). ~
~
t
The collector of this project differs from previous similar systems primarily in the use of only moderate concentration levels, and hence the ability to use the commercially-priced essentially one-sun technology LBG cells. It has the potential to incorporate a secondary optical element, to raise the concentration ratio, or to increase tolerance to tracking errors. Half the collector was fitted with secondary concentrators, and the other half was not. This collector is an easy device to manufacture in volume. The acrylic lenses are made by 3M's calendaring process, already used in volume for street signs, while the aluminium sheet housings can be pressed to shape. The cells are relatively large, so either manual or automated tabbing, encapsulation and mounting are straightforward. 4.2 Collector B, Multiple offset cylindrical paraboloid, one
axis tracking
Figure 2. The longer cell string with bypass diodes If the shaded cells are fitted in the collector, then at the equinox, all the cells are illuminated and there is no end loss. In the worst case, at a solstice, the m cells at one end only are shaded, and the output is that from n + m cells. One bypass diode will be turned on, and the voltage drop across it will be about the same as the output voltage of one cell. So the output at a solstice will be that from n + m - 1 cells. Away from the solstice, fewer than m cells will be shaded and the output will increase. So the end loss will be less than half than that when the shaded cells are omitted; for the SMTS collector the correction factor is increased from 65% to about 83%. (A full calculation of the loss factor depends on the distribution of sunlight through the year, and has not been attempted). The cost of the extra 2m solar cells is small, and the cost of the diodes is negligible. The cost per peak watt of collector C, the SMTS collector, is reduced from 2.39 $/Wp to about 2.06 $/Wp. Of course the voltage of the peak power point changes as cells are shaded, and so a true peak power tracker and appropriate array to load converter are needed in the system; but they should be provided anyway for an optimal design. 4. DISCUSSION OF SPECIFIC COLLECTORS
4.1 Collector A, Point focus Fresnel lens, with two axis tracking This collector (Hunt, 1998) is a two-axis tracked, point focus system using flat acrylic Fresnel lenses. As such, it is similar to many previous systems, including: the original Martin Marietta Soleras collectors (Salim and Eugenio, 1990), various Sandia designs (Chiang and Quintana, 1990), and collectors from Alpha Solarco (Carroll, Schmidt and Bailor, 1990) and Midway. The advantages of this design include:- maximum beam insolation collection due to two-axis tracking potential for simple mass-produced optics, the use of the housing as heat sink; while the main disadvantages include:- the increased cost of the second axis tracking, the fact that flatFresnel lenses offer reduced efficiency at practical f-numbers.
This collector (Weatherby and Bentley, 1998) was designed for polar axis tracking. The designed concentration ratio was 20x. A single module consisted of two cylindrical parabolic mirrors, each with an aperture of 150mm x 1.5m, formed from thin (0.5mm)sheet aluminium alloy laminated with SS-95P reflective material. The mirrors were arranged back to back so that the two 7.5ram wide strings of cells, located near to the focal lines, were on the outer walls of the module, to facilitate the cooling of the cells. The cooling was further enhanced with a fin extending normal to the sun. A simple glass or acrylic cover was used to protect the optics and add strength to the module. The cell strings of two modules were connected in series forming one integrated double module of about 28Vor Advantages of this collector are the simple modular construction and the ability to mass-produce the housings by an inexpensive stamping method. The cooling arrangement removes the need for bulky expensive heat sinks whilst keeping the cells at temperatures no higher than those of conventional flat panels. To avoid the cost of the large press tools necessary for mass production, a series of fibre-glass ribs were moulded and used to retain the mirrors in their parabolic form. One disadvantage of this technique is that the ribs obstruct the convective flow of cooling air. It is expected that mass-produced modules, with no fibs, will give better cooling and even better performance than the prototype. 4.3 Collector E Multiple offset cylindrical paraboloid, moved manually once per day Of the four prototype collectors, this one (also Weatherby and Bentley, 1998) is designed to be the simplest and most robust, as no automatic tracker is required. It consists of parabolic troughs that run E-W, so that as the sun moves approximately along the axial plane of the parabola the focal line moves mainly E-W along the line of cells. The collector tilt is designed to be manually re-aligned every day, or every few days, to accommodate the changing declination of the sun. Even so, except at the equinox, the sun moves along a small circle, and so moves a little in a N-S direction off the great circle defined by the axial plane of the parabola. The width of the cells therefore has to be greater than the width of the focal line, to allow for this N-S movement. The chosen geometrical concentration ratio is 6, giving an acceptance angle of about 7~, which allows the sunlight to fall on the cell string for about 8 hours on most days.
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The design of the collector envisages a very simple process when in mass-production: - the full 2m 2 reflective sheet is stamped to profile, - the cell strings are attached, - the glass cover is bonded in place. For the prototype, we could not afford the tool for a stamping, and a rather elaborate fabrication of the parabolas, supported by a lattice of ribs and cross-members, was used instead.
can be allowed for in subsequent analysis. Collector A (section 4.1 above), provided experience with a two-axis tracker. 4.5 Collector D, Cylindrical paraboloid, with secondary
concentrators and single axis tracking This was another promising design, slightly more costly than A, B and C, but cheaper than F and G. It was not selected for manufacture in this project because we had already chosen two better mirror systems, and we wanted to include a system with manual movement instead of automatic tracking. 4.6 Collector F~ Linear Fresnel lens, with solid CPC secondaries and two axis tracking This collector, although more costly than some of the others, has the advantage of being simple and totally enclosed. Some work was done on the manufacture of prototype linear lenses, by machining them from acrylic sheet, but it was not possible to polish the milled surfaces to a sufficiently good finish in a reasonable time. Some lenses are commercially available, but they are designed for other purposes and are not of the correct dimensions for our collector. Special lenses could be made to our design by 3M, but the cost of a prototype was prohibitive. So this design was abandoned.
Figure 3. Cross section and workings of the SMTS collector 4.4 Collector C, Single mirror two stage (SMTS) collector, single axis tracking The SMTS collector, Figure 3, (Alarte, Benitez and Mifiano, 1998) presents some characteristics that make it a good candidate for achieving the required cost reduction. In particular: a) Since the cell strings and the heat sinks are at the mirror edges, the heat sinks can form part of the mechanical structure of the collector, reducing the mass and cost of other structural material. b) There is only one mirror per concentrator, in such an arrangement that two concentrators share their mirror so one mirror works as a first stage for one concentrator and as a second stage for the other. Therefore it gets a high acceptance angle (around 90% of the theoretical maximum corresponding to its concentration) yet maintaining the structural simplicity of a parabolic trough collector (whose acceptance angle reaches no more than 50% of the theoretical maximum). This high concentration-acceptance angle product makes the collector less sensitive to tracking or manufacturing errors, which reduces again the manufacturing cost. The spreadsheet analysis showed that it would be most costeffective to use a two-axis tracking system; however for mechanical convenience a single-axis polar mounting was chosen for the prototype. The loss, due mainly to end losses,
4.7 Collector G, V-trough, with single axis tracking This collector was already being investigated by ZSW, as a cheap alternative to fiat plate systems. Although not so cheap as the best collectors, it is much nearer to practical production, as the solar cell arrays can be conventional planar arrays and the optical design and construction of the mirrors is not critical. The tracker was to be thermo/hydraulic, which does not require an external power supply, and promises to be cheaper in production than an electronic system. But the prototype, being built under another programme, was not ready in time for this project. 5. THE PROTOTYPES BUILT The characteristics of the prototypes are summarised in Table 2, and photographs of them are shown in Figures 4 to 7. 6. TESTING All the prototype collectors were built at Reading. During construction, the solar cells were all individually tested on a computer controlled laboratory tester, to reject faulty ones (in fact, none) and to select matched sets for each string. The variation between cells was small, and perhaps due to measuring errors; it was probably unnecessary to select at all. The peak efficiency of the cells used was about 18 %, corrected to 25~ at an insolation of 15 kW/m 2, failing to about 17 % at
Table 2. The prototype collectors
Code A2 A1 B C F
Primary Optics Point-focus Fresnel Point-focus Fresnel Parabolic mirror SMTS Parabolic mirror
2ndry
Conc.
Optics No Yes No Yes No
Ratio 32x 69x 20x 30x 6x
Module Width (In) 0.225 0.225 0.150 0.300 0.150
Aperture
(mz) 0.81 0.81 1.80 2.40 1.80
Axes .Tracking 2-axis: cont. 2-axis:cont. Polar: cont. Polar: cont. E-W: 1/day
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Figure 4. Collector B. Multiple cylindrical semi-parabolic mirrors, tracking about a polar axis.
Figure 5. Collector C. The single-mirror two-stage collector, tracking about a polar axis.
Figure 6. Collector F. Multiple cylindrical semi-parabolic mirrors, moved manually every day about an E-W axis
4 and 3 7 kW/m2. The cell strings were tested at 1-sun, to check for faulty connections, before being fitted to the collectors. As soon as they were completed, the collectors were tested for a short period, measuring spot performance and I-V curves. The results are summarised in Table 3, 'Measured Efficiency', 'Best Module' and 'Collector'. Three of the collectors, A, B, and F were then sent to ZSW's test site at Widderstall for long term testing. A computer-based data logger was used to record, at one minute intervals, output current, voltage and power at the peak power point, cell temperature, tracking accuracy, and a range of meteorological data, notably direct and global insolation, ambient temperature, and wind strength and direction. From time to time I-V curves were taken. Typical results for insolation and output power for a clear warm summer day (8th August 1998), are shown in Figures 8 and 9. Beam radiation was above 800 W/m2 for more than 8 hours. The daily insolation sums were: diffuse horizontal: 859 Wh/m2, direct normal: 10180 Wh/m2, and total normal: 11393 Wh/m"~. Ambient temperature was between 20~ and 29~ and wind velocity between 1 m/s and 2.5 m/s while the beam radiation exceeded 800 W/m2. The 2-axis tracked collector, A2 was in operation for about 12 hours; the end switch stopped further tracking in the evening. The fixed collector F was active for approximately 8 hours. The polar axis tracked collector B generated power for more than 12 hours. The total generated energies for this day were 725 Wh for Collector F, 505 Wh for Collector A2 and 1239 Wh for Collector B. The output power of collectors A2 and F were lower than they were on a winter day, 25 th March, mainly due to the high summer temperature. Collectors B and F, tested at Widderstall proved to be reliable practical units. They withstood a year of weather with no significant deterioration. Heat sinking has proved satisfactory, with cell temperatures 30~ to 45~ above ambient, similar to conventional flat panels. The test results are summarised in Table 3. The best module efficiencies, normalised to 25~ andexcluding the end losses of linear systems, were: 12.5%, 13.2%, 13.6% and 14.3% for collectors A2, B, C, and F, respectively. Full systems performed somewhat worse than their best modules, due to manufacturing variations. There were two serious problems revealed by the tests:1. Collector A, using point focus Fresnel lenses had a concentration ratio of 32x without secondaries, and 69x with secondaries, too high for the cells used, which were optimised for 15 suns, causing a loss of efficiency at the higher concentration. The cure is to use cells optimised for a higher concentration ratio, or to redesign the optical system for a lower concentration ratio. 2. Collector C, the single mirror two stage concentrator was made of fibreglass (GRP) mouldings, coated with a silvered plastic reflecting film. The intensity of sunlight on the secondary part of the mirror was sufficiently great to burn the GRP wherever there was a flaw in the mirror. So, although the optical performance was very good, the collector would not be reliable in use. The cure is to build the mirror from aluminium, to spread any local heating over a reasonable surface area.
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Figure 8. Insolation on a clear summer day, 8th August 1998.
Figure 7. Collector A. Multiple circular Fresnel lenses, mounted in gimbals, with closed-loop active tracking about both axes. 7. C O N C L U S I O N S Several designs of small concentrator systems can be significantly cheaper than conventional planar arrays, reducing cost/watt and cost/kWh by a factor of 2 or 3. To achieve such reduced costs, the concentrators should be designed to use minimum amounts of material, and be manufactured in such a way, and in sufficient quantity, as to keep down the manufacturing cost.
Figure 9. The output of collectors A2, B and F on 8th August 1998.
Table 3. S u m m a r y o f costs and performances o f the prototype collectors
Code A2 AI B C F
Optics Pt-focus Fresnel Pt-focus Fresnel I Parabolic mirror I SMTS I Parabolic mirror
Conc~ Ratio 32x 69x 20x 30x 6x
From Spreadsheets Effcy. (%)* $/Wp I c/kWh* 16.0 1.48 [ 5.4 15.5 1.46 [ 5.4 15.1 1.62 I 6.2 15.1 2.39 I 9.2 15.4 2.64 I 14.2
Measured Efficiencies (%)* Best Mod. 12.5 9.8 1'3.2 1'3.6 14.3
I
Coll. I ZSW spot 12.3 [ 8.0-11.0 9.4 [ 2.5 -5.0 [10.619.4-11.0 . 12.2 I 13.4
. I 9-13
ZSW lon~* 7.0 - 8.5 2.2- 3.2 7.7 - 8.4 6.9 - 9.2
* Notes: - The spreadsheet c/kWh is for a good site (Almeria), and is a comparative number: it includes cells, optics, housing, structure, tracking, and manufacture, but does not include, e.g., power conditioning, land or overheads. The comparable figure for a flat array is 18.1 c/kWhr. - The spreadsheet efficiency is a spot (i.e. instantaneous) value. The annual efficiency assumed is sometimes lower due to tracking and other losses. The figures are normalised for cells at 25~ and do not include the 'end losses' of the linear systems (as these reflect each system's physical shape, rather than its cell and optical performance). - The measured efficiencies are: - Best Mod.: the better, or best, of the prototype modules, - Coll.: the complete prototype collector, - ZSW spot: typical spot efficiencies from the ZSW field test results. These three efficiencies are also for cells at 25~ and exclude 'end losses', and thus can be compared with the original 'spreadsheet' estimate. - The final efficiency ('ZSW long') is calculated from the total power output from each collector during its period of testing at the ZSW test site, divided by the total of the beam radiation normal to that collector's aperture over the same period. It has not been adjusted back to 25~ nor are end losses subtracted, so is not comparable with the figures given previously. However, this figure can be combined with the annual energy incident on each of the collector types to indicate that system's typical long-term output under field conditions.
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Prototypes of four designs of concentrator were built by hand. Design and construction, though taking longer than anticipated, presented few problems. Three were tested for some months at ZSW's test site at Widderstall. Overall, the prototypes have behaved roughly as we expected; they have proved to be robust and reliable, and capable of operating for long periods in the field. The performance figures from the tests generally support the estimates entered in the spreadsheet calculations. For collectors B, C and F, best prototype module outputs ranged from 7% to 13% below the expected performance figures. Collector A, where the cells were run well above their design concentration, gave a performance (without and with secondaries) of 22% and 37% below expectation. On the basis of the spreadsheet data, the best of the collectors have costs in the region of 1.5 to 1.8 $/Wp, yielding energy costs at a good site (excluding BOS and overheads) of between 5 and 7 cents/kWh. The corresponding figures for a fixed planar PV array are 4.3 $/Wp, and 18 cents/kWh.
REFERENCES Matte E., Benitez B. and Mifiano J. C. (1998). Design, Construction and Measurement of a Single-Mirror Two-Stage (SMTS) Photovoltaie Concentrator. In Proc. 2nd World Confr. on P V Solar Energy Conversion, Vienna, EC, Luxembourg, pp 2245-2247. Carroll D, Sehmidt E. and Bailor B. (1990). Production of the Alpha Solarco proof-of-concept array. In Proceedings of 21st IEEE PVSC Vol 2, May, Kissimimee, Florida, USA, pp 1136 -1141. Chiang C. and Quintana M. (1990). Sandia's concept-90 photovoltaic concentrator module. In Proceedings of 21a 1EEE PVSC Vol 2, May, Kissimimee, Florida, USA, pp 887891. Hunt A. C. (1998). Design and Manufacture of a Point-Focus Fresnel Lens Concentrator for a Stand-Alone PV System. In Proc. 2nd World Conf. on P V Solar Energy Conversion, Vienna, EC, Luxembourg, pp 2185-2188. Maish A.B. (1991). The Solartrak solar array tracking controller. Sandia Report, SANDgO--1471. UC-275. Mason N. B., Bruton T. M. and Heasman K. C. (1995). Optimisation of low-cost concentrator solar cells. In Proc. 13th European PV Solar Energy Confr., Nice, H.S. Stephens, Bedford, pp 2110-2112. Sala G. et al. (1997) Description and performance of the EUCLIDES concentrator. In Proc. 14th European PV Solar Energy Confr, Barcelona, H.S. Stephens, Bedford, pp 352 -355. Sala G. et al. (1998). 480 kW peak EUCLIDES concentrator power plant using parabolic troughs, In Proc. 2ad Worm Conf. on PVSolar Energy Conversion, Vienna, EC, Luxembourg, p p 1963-1968. Salim A. and Eugenio N. (1990). A comprehensive report on the performance of the longest operating 350kW concentrator photovoltaie power system. Solar Cells 29, pp 1-24. Weatherby C. K. and Bentley R. W. (1998). Further development and field test results of two low-material-cost parabolic-trough concentrators. In Proc. 2nd Worm Conf. on P V Solar Energy Conversion, Vienna, EC, Luxembourg, pp 2189-2192.
Whitfield G. R. et al. (1997). The development of optical concentrators for small PV systems. In Proc. 14th European PVSolar Energy Confr., Barcelona, H.S. Stephens, Bedford, pp 336-339. Whitfield G. R. et aL (1998). Development and testing of optical concentrators for small PV systems. In Proc. 2nd World Conf. on P V Solar Energy Conversion, Vienna, EC, Luxembourg, pp 2181-2184.
Acknowledgement The support of the EC for this work, under contract JOR3CT96-0101, is gratefully acknowledged.
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XXI. Active Cooling, Refrigeration and Dehumification
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THERMODYNAMIC DESIGN OF A SOLAR REFRIGERATOR TO PRESERVE SEA PRODUCTS
Hector D. Arias-Varela, Wilfredo Soto-Gomez and Oscar Castillo-Lopez Metal-mechanical Department, Institute of Technology of Tijuana, Calzada del Tecnologico s/n, Tijuana, Baja California., Mexico 22370, Phone/Fax (66) 833519, E-mail: [email protected]
Roberto Best-Brown Centro de Investigacion en Energia, Universidad Nacional Autonoma de Mexico, Apartado Postal No. 34, Temixco, Morelos, Mexico 62580, E-mail: [email protected] Abstract - This project was developed to determine a means of providing refrigeration to communities lacking conventional energy sources. The design of an absorption refrigeration system operating with solar energy was carried out. The refrigerator is used to conserve sea food. The system is adapted to an industrial size cold-storage room. A maximum of 200 kg of fish in ice may be introduced to this room daily, up to a total capacity of 2 tons. The lowest temperature the evaporator reaches is -10 C, the high and low system pressures are 13.4 atm and 2.87 arm respectively. The refrigerant-absorbent mixture is ammonia and water, where the refrigerant is ammonia. The design of this system requires six effective solar hours to generate the refrigerant needed by the refrigerator to work eighteen hours daily. Evacuated tube solar collectors are used. Only solar energy is used to operate the system. To compare the cost effectiveness of this solar refrigerator with a vapor compressor refrigerator of the same capacity, the following was considered: the vapor compression refrigerator requires electricity generated by internal combustion plant. The period of comparison is twenty five years with a MARR of 4.5%. Initially, solar energy refrigeration is more monetarily expensive, but less expensive ecologically than conventional refrigeration. However, at twenty three years of operation they become the same monetarily. Beyond twenty three years, conventional refrigeration is more expensive. I. INTRODUCTION Santa Clam Ca. Recreation Center, cooling space 2508 m 3, flat plate collectors copper/copper, collection area 650 m 2, and a refrigeration unit working by absorption with lithium-bromidewater with a capacity of 25 tons (USA). -
Most conventional refrigeration systems operate with electricity, however, there are regions where it is difficult or not cost efficient to provide electric service. In addition, the cost of generating electricity is high, both economically and ecologically. Therefore, a project was developed to determine a means of providing refrigeration to communities lacking conventional energy sources. The design of an absorption refrigeration system operating with solar energy was carried out. The refrigerator is used to conserve sea food. Solar refrigeration by absorption is considered to be an alternative to substitute conventional refrigeration equipment and a way to save electricity or make refrigeration possible in areas without electricity. While research on solar cooling has been carried, a design that operates efficiently and is economically within reach of its users has not been mass produced. The majority of research on solar cooling has focused on lithium bromide and water where water is the refrigerant. This limits the operation of the system to temperatures at or above 5 C. Moreover, even the systems designed for high temperature have low efficiency and little competitiveness in the market (Hacuz, 1982). A list of different experimental prototypes related with solar refrigeration around the world are listed as follows: - University of Ohita Building, cooling space 1860 m 3, flat plate collectors copper/copper, collection area 513 m 2, and a refrigeration unit working by absorption with lithium-bromidewater with a capacity of 30 tons (Japan). - Solar House Hirakata, cooling space 118.5 m 3, evacuated tube collectors copper-aluminum, area of collection 46.6 m z, and a refrigeration unit working by absorption with lithium-bromidewater with a capacity of 2 tons (Japan).
- Brisbane Solar House, cooling space 123 m 3, flat plate collectors copper/copper, collection area 6 m 2, and a refrigeration unit working by absorption with lithium-bromide-water with a capacity of 2 tons (Australia). - Carrier Corporation Phoenix Arizona, cooling space 3700 m 3, linear lenses, collection area 133.8 m2, and a refrigeration unit working by absorption with lithium-bromide-water with a capacity of 13.5 tons (USA) (Huacuz 1992), (Bogart, 1992) (Mongomery, 1991) and (instituto de Refrigeracion y Aire Acondicionado, 1989). This paper concentrates on the thermodynamic design and analysis of a solar refrigeration system. The project is unique since the system is adapted to an industrial size cold-storage room and developed to operate under the following conditions: temperatures less than 0 C in the evaporator, 18 hours of operation per day, storage of refrigerant, approximately 6 effective solar hours of solar collection with evacuated tubes and the use of solar energy as the only source of power to operate the system. The components of the solar refrigerator are shown in Figure 1. The elements of the system located in the high pressure region are the following: - Evacuated Tube Solar Collectors. The function of this component is to heat the thermal oil used by the generator.
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- Separator. This is a more sophisticated heat exchanger that ensures approximately one hundred percent of ammonia separation from water.
- Absorber. This element contains enough water to absorb the mixture of water-ammonia or ammonia that comes from the mixer. After the absorption takes place, the solution with a high concentration of ammonia is send to the tank of strong solution. - Tank of Strong Solution. This tank keeps the ammonia-water solution. This solution is released as required and pumped to the high pressure region.
- Condenser. This heat exchanger uses water in order to condense the ammonia that will be used as the refrigerant. After the refrigerant is condensed it is sent to the refrigerant tank.
- Pump. This component pumps the solution to the generator w (in the high pressure region). The pump is powered by electricity which is generated with photo-voltaic panels.
- Generator. This element is a heat exchanger whose main pm]x)se is to heat the ammonia-water mixture in order to separate the water from the ammonia.
2. THERMODYNAMIC ANALYSIS OF THE SYSTEM 1. S0I~ COI,LXCT01~S
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A thermodynamic analysis of the system was carried out quantitatively. Calculations were made to detemfine the minimum characteristics of the system that are needed to guarantee its operation.
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First, the balance of energy was calculated. The dimensions and characteristics of the refrigerator are shown in Figure 2. The walls are constructed of brick and concrete and the interior lined styrofoam sheets for insulation. The thickness of the insulation is l0 em. The roof consists of a concrete with gravel with a thickness of l0 em and styrofoam insulation of the same thickness. The construction of the floor is similar to the roof. The difference is that the insulation is placed between the concrete floor and the ground. The height of the refrigerator is 2.3 m. An interior temperattae of 1 C must be maintained. The exterior atmospheric temperature is considered to be 40 C, which is the average high temperature in the region. The total heat flow for the building was calculated to be 1.5 kw. 4.1m
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- Refrigerant Tank. This tank keeps the refrigerant that is released as required. ARer this tank, there is a pre-cooler that causes the temperature of the ammonia (at liquid state) increase.
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- Pre-cooler. This heat exchanger uses the ammonia that returns from the evaporator at high temperature in order to elevate the temperature of the refrigerant that comes from the refrigerant tank and passes to the expansion valve. - Expansion Valve. This valve regulates the refrigerant flow and lowers the pressure of the system before the ammonia enters the evaporator.
To
= 40 C
Figure 2. Dimensions of storage room. - Evaporator. Through this element, the refrigerant that absorbs the surrounding heat circulates so that the space around the evaporator becomes colder. The components of the system located in the low pressure region are the following: - Mixer. This component mixes the ammonia that comes from the Pre-Cooler with the water that has been separated from the ammonia in the high pressure region. Then, the mixture is absorbed by the Absorber.
Next, the refrigeration load was calculated. The system=s total refrigeration load originates from the following sources of heat: heat transmission between the difference in interior and exterior temperatures, load of the products which are intrOduced at room temperature and that must be cooled to the temperature of the refrigerator, air infiltration, supplementary loads caused by electric lights, motors, tools, and even people. The energy needed to reduce the temperature of the fish from 35 C to 1 C was calculated based on the conservation of 200kg of fish per day. The product load was found to be equal to 6.9 kW-h. The load caused by air infiltration was considered to be within the range of
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design safety factor along with the supplementary load since the volume of the refrigerator is relatively small and the number of times the doors will be opened is minimal. In addition, the refrigerator does not have much supplementary equipment. The total load of the refrigerator was found to be 2.0kW.
calculations necessary for the fabrication of these parts were carried out. Table 1. Energy Balance Results. 1
:
:
Enrgy Op. Time Energy In Energy Out Type KVV h KW-h KW-h QGE 14.02 6 84.12 QEV 2.00 18 36.00 v~ol o.os s o.3~ 3222 -6~-6- ~ s.3z 6 --29.52 ....QRE 4.92 6 --46.62 QAB 2.59 18 --1 2.00 QSD 2.00 6 --1 20.36 Total i 1 20.43
The following characteristics were also calculated:
. . .
. . .
1) Total mass of ammonia, 100kg 2) Concentration of ammonia and water, 0.42 3) Condensation pressure at 34 C, 12.94 atm 4) Total mass of the solution, 961.29 kg 5) Energy required by the generator, 14.02 kW 6) Rectification energy, 4.927 kW 7) Condensation energy, 5.37 kW 8) Absorption energy, 2.5867 kW 9) Energy transferred by the weak solution, 2.00 kW 10) Work by the pump, 0.0523 kW
_ . _
The thermodynamic properties of the system are indicated and explained in Figure 3. Based on the energy balance, the system is considered to respect the First Law of Thermodynamics. The energy types entering the system, inputs, are the heat of generation (Q~3E), the heat of evaporation (QEv) and the work of the pump (WBO). The energy types going out the system, outputs, are the heat of condensation (Qco), the heat of rectification (QRE), the heat of absorption (Q~), and the heat transfered by the weak solution while in the weak solution tank (Qws). The energy balance results are shown in Table 1 where it is observed that input of energy, 120.43 KW-H, is almost the same as the output, 120.36 KW-H. "
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The Solar Collectors. After analyzing the efficiency and cost of evacuated tube solar collectors manufactured by various companies the decision was made to use Suntube which is designed to provide optimal efficiency for the collection of solar energy at a minimum cost (Nipon Electric Glass, 1990). It allows a 30% reduction in the required collection area. In addition it has a long service life due to its closed construction which effectively prevents problems such as corrosion and leaks. The solar collector tube may be installed horizontally which not only permits the use of virtually any part of the surface, but also assures that the collector can be integrated into the existing building. Due to its e ~ i a l layer it has a high rate of absorption (0.91 minimum) and a low rate of emissions (0.12 maximum). Figure 4 show a diagram with which the efficiency can be determined as a function of irradiation, air temperature, and the average operating temperature of the fluid. The area of the solar collectors required by the system was determined based on the quantity of energy demanded and the mass flow of required heat exchange fluid. The energy demand of the generator is 16 KW ( considering a safety factor of 15%). The quantity of heat exchange fluid required can be calculated based on the proposed entrance and exit temperatures of the oil in the generator. The oil used is Marlotherml. The average specific heat is calculated to be 0.825 kcal/kg C. For 16 KW, a result of 1667.87 kg/h is obtained. Since the average density of Marlotherml (110-120 C) is 912 kg/m2, the volume required is calculated to be 1819.5 l/h.
18 18 18 18 18 18 s
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For the thennic and mechanical analysis, the main components of the system are the following: Solar Collectors, Generator, Separator-Rectifier, Condenser, Condensation Tank, Cooler, Evaporator, Absorber, Strong solution tank, Mixer, and Weak Solution Tank. The other components of the system are considered to be complementary equipment.
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Figure 3. Results of thermodynamic analysis. 3. THERMIC AND MECHANICAL ANALYSIS This section examines the thermic and mechanical characteristics of the elements of the system. For the selection of these elements, existing commercial equipments is considered first. For elements that are not available commercially, design
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422
Length without fins (cm): Total length (m): Fins per inch (2.54 cm): Height of fins (cm): Type of fins: Tube material: Number of Tubes:
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Figure 4. Solar collectors efficiency. To determine the efficiency of the solar collectors using Figure 4, it is necessary to know the temperature difference and the solar radiation. The temperature difference was calculated to be 83 C. The average solar radiation in W/m2 for the region studied was obtained l~om a manual of the Institute of Engineering of the UNAM ACalculo de la Radiation Solar Instantanea en la Republica Mexicana`@ (Fernandez, 1990). For an average of six hours of effective solar radiation, the value of 597.7 kcal/m2h. In this case the efficiency of the solar collectors is E = 0.4. The energy absorbed is 278 W/m2. This result in a total absorption area of 57.55 m2 to satisfy the demand of 16 KW, or 32 collectors arranged in four rows of eight. The use of oil in the solar collectors causes a drop in pressure which was calculated to be 67.87 lb/pig 2 This drop in pressure requires a 3/4hp motor. The Generator. A high pressure generator with carbon harden steel tubes and fins was selected because of the use of a solution containing ammonia. The Separator-Rectifier. Due to the difficulty of finding a separator-rectifier that is able to cool the vapor to allow all of the water to condense and produce a vapor that is pure ammonia, the decision was made to design it and order its construction by an industrial shop. The separator consists of a distillation column of carbon hardened steel. The nucleus of the column has a series of washers the facilitate the separation of the vapor. The Condenser. A heat exchanger of armor plating and tubes was chosen since a heat exchanger of double tubes required a length of nearly 30 meters. The Refrigerant Tank. The condensation storage tank must be able to store 170 1 of ammonia along with a addition 15% of the gases that are not able to be condensed or 25.51. The Precooler. The precooler of the liquid refrigerant uses 1.53 m long coiled tubing with a diameter of 5.5 mm in a carbon hardened steel tube 0.5 meters long with a diameter of 6.5cm. The Evaporator. An characteristics was chosen:
evaporator with the following
Finned tubing: Fins welded with high l~equency resistance in a helicoidal form Exterior diameter (cm): 3.34 Wall thickness (cm): 0.33 Length with fins (m): 4.89
10.16 4.99 5 1.90 Solid I-IF Carbon hardened steel 3
The absorber. The strong solution storage tank must guarantee the storage of 1,137.5 1 of a solution of ammonia and water at a pressure of 3 arm and a temperature of 34 C in addition to 10 % of the non condensable gases (113.75 1) for a total of 1,252.25 1. It is a carbon hardened horizontal cylinder 1.22 rots long. The mixer. A maximum of 200 kg of fish in ice may be introduced to this room daily, up to a total capacity of 2 tons. The lowest temperature the evaporator reaches is -10 C, the high and low system pressures are 13.4 arm and 2.87 arm respectively. The refrigerant-absorbent mixture is ammonia and water, where the refrigerant is ammonia. The design of this system requires six effective solar hours to generate the refrigerant needed by the refrigerator to work eighteen hours daily. Evacuated tube solar collectors are used. A photovoltaic system is also required to power the principle electrical components, the motors of the hydraulic system, each with a different capacity. The equipment is as followed: a pump for oil, 3/4 h.p., 750 W, operating 6 hours per day; a high capacity pump for water, 500 W, operating 6 hours pers day; a pump for NH3-H20, 1/4 h.p., 250 W, operating 6 hours per day; and a low capacity pump for water, 1/8 h.p., 125 W, operating 12 hours per day. The total amount of energy required is the sum of the potentials multiplied by the operating hours, 10,500 W. To detemdne the energy total that must be provided, it is also necessary to consider the energy that the components of the photovoltaic system consume. This self-consumption energy can be calculated with the energy efficiency of the elements of the system. The values are as follows: cmrent invertor 95%, charge controller 95%, lead acid batteries 90%. The total energy that must be furnished by the photovoltaic system is 12, 927.054 W. The peak potential, in KWp, is proportional to the average total energy required per day, over the average solar radiation per day, expressing the solar radiation in KWh/day. Therefore, the peak potential is 2,5854 KWp. The selected electric equipment works within the range of 115-125 volts. For calculations a nominal voltage of 125 volts is used. Therefore, the current demanded by the system is 20.68 amp. The proposed batteries are lead acid 12.5 volts DC, the recommended charging voltage is 14.5 volts. In addition, if the nominal voltage required is 125 volts, it is necessary to connect in a series 10 batteries of 12.5 volts. moreover, to charge 10 batteries in a series 145 volts are require& If the batteries are 12.5 volts/200 amp-h and if a battery bank with 3- day autonomy is desired, the bank has to be three parallel lines with a series of 10 batteries in each line. The proposed battery is made by SIEMEN with a potential of 53 W, 17 volts and 3 amps. The photovoltaic arrangement needed to provide 145 volts and 20.68 amps consists of seven parallel line with nine modules in a series in each one of them. Overall, 63 modules are necessary to supply the required energy. 4. COST ANALYSIS In this section a cost analysis which compares the solar absorption refrigerator designed for this project for which the
ISES Solar World Congress 1999, Volume III
only source of energy is solar with a conventional vapor compression refrigerator which uses a gasoline-powered electric generator was carried out. To compare the cost effectiveness of this solar refrigerator with a vapor compressor refrigerator of the same capacity, the following was considered: the vapor compression refrigerator requires electricity generated by internal combustion plant. This uses fuel whose price increases at an annual inflation index of 9.1% (average in Mexico for the last 10 years). The period of comparison is twenty five years with a minimal attractive rate of return (MARR) of 4.5% per year. All monetary values are stated in U.S. dollars since it is a more stable currency. The approximate cost of the solar refrigeration system was obtained using the prices of the components of the system: 32 solar collection modules, $17,280.00; oil pump, $540.00; generator $3,270.00; separator-rectifier, $1,166.67; condenser, $1,855.26; condensation storage tank, $933.63; pre-cooler, $318.34; evaporator $2,112.18; absorber, $1,985.53; strong solution storage, $2,235.60; mixer $2,485.00; weak solution storage $2,005.60; expansion valve (refrigerant), $135; expansion valve (weak solution), $65.00; pump (solution) $375; high capacity pump (water), $440.00, low capacity pump (water), $214.00; cooling tower, $2,471.73; photovoltaic system, $55,113.34. The total cost is $95,001.88 U.S. Dollars. The cost of the equipment necessary for the connections is not included, neither is the cost of the fluids. The price of the equipment does not include installation. The cost of the solar energy is $0.00 dollars. The cost of a conventional system with the same characteristics as the solar refrigeration system is $7,991.44 U.S. dollars. However, the replacement of equipment has to be considered. The gasoline powered generator which its cost is $960.00 has to be replaced every five years. The compressor unit and other components which their costs add up to $4,000.00 have to be replaced every 10 years. The cost of the gasoline required to operate the generator must also be calculated. It is estimated that it will consume 1.5 liters per hour of operation and run 9 hours a day for 25 years. Initially, solar energy refrigeration is more monetarily expensive, but less expensive ecologically than conventional refrigeration. However, at 23 years of operation they become the same monetarily. Beyond 23 years, conventional refrigeration is more expensive as shown in Figure 5.
423
Cost Break-Even Point
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5. CONCLUSIONS The construction of a solar refrigerator to conserve sea products is technologically feasible and it can be an alternative to provide refrigeration to rural communities which do not have electric service. Although the initial cost of the solar refrigerator is relatively high and it makes this alternative less attractive economically, it could be an acceptable solution since the cost of the conventional and solar refiigerators become the same at approximately 23 years of operation with a MARR of 4.5%. In addition, the costs of the solar refrigerator can be reduced significantly if this equipment is mass produced. Many of its components can be improved and standardized in the future which will lower the costs of the solar refrigerator. REFERENCES
Bogart Marcel (1992) Ammonia Absorption Refn'geration in Industn'alprocesses, 1st edn, pp. 5-15. GULF, New York. Fernandez Zayas Jose (1990) ACalculo de la Radiacion Solar Instantanea en la Republica Mexicana@ Instituto de Ingenieria UNAM Serie No. 472 Huacuz Jorge (1992) AEstudios de Refrigeration Solar@ Boletin del Instituto de Investigaciones Electricas, Vol. 6 Issue No. 2. lnstituto de Refi'igeracion de Aire Acondicionado (1989) Manual de Refrigeracion y aire Acondicionado, 2nd edn, pp. 213-218. Prentice Hall, Mexico D.F. Montgomery Richard (1991) Energia Solar. Seleccion del Equipo, Instalacion y Aprobechamiento, 1st edn, pp.86-90. LIMUSA, Mexico D.F. Nipon Electric Glass (1990) ATechnical Data for NEG Evacuated Tube Solar Collector and Solar Collector Module LD-2800/DP62800@ Nipon Electric Glass Co., Ltd.
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DEMONSTRATION OF A NEW TYPE OF ICPC IN A DOUBLE-EFFECT ABSORPTION COOLING SYSTEM
Roland Winston and Joseph J. O'Gallagher Enrico Fermi Institute, University of Chicago 5640 South Ellis Avenue, Chicago IL 60637 USA Telephone: 1-773-702-7756, FAX: 1-773-702-7756 E-mail: [email protected]
William S. Duff Dept. of Mechanical Engineering, Colorado State University Fort Collins CO 80523 USA Telephone: 1-970-493-1321, FAX: 1-970-495-0657 E-mail: [email protected], edu
Tom Henkel and Julius Muschaweck Solar Enterprises International, c/o Richardson Electronics, 40W267 Kesshnger Road, LaFox IL 60147, USA Telephone: 1-630-208-2577, E-mail: [email protected]
Rich Christiansen Ohio State University
Jim Bergquam Sacramento State University
Abstract- Recently two new technologies, a new ICPC solar collector design and the solar operation of a double effect chiller, were commercially demonstrated for the first time. A new family of ICPC designs was developed which allows a simple manufacturing approach to be used and solves many of the operational problems of previous ICPC designs. A low concentration member of this family of designs that requires no tracking was used in the demonstration. For the demonstration, an off-the-shelf 20 ton double effect direct fired absorption chiller was modified to operate with hot water. The new ICPC design and double effect chiller was able to produce cooling energy for the building using a collector field that was about half the size of that required for a more conventional collector and chiller.
1. INTRODUCTION Two new technologies, a new ICPC solar collector and the solar operation of a double effect (2E) chiller, were commercially demonstrated for the first time. In early 1998 the collector and a 2E chiller were installed in an office building in Sacramento, California. This paper describes the demonstration project and reports on component and system performance. Research on Integrated Compound Parabolic Concentrator (ICPC) evacuated solar collectors has been going on for more than twenty years. See (Garrison, 1979) and (Snail et al, 1984). With a good selective surface, a non-tracking ICPC evacuated solar collector can provide highly efficient daily collection of solar energy at temperatures up to 250C. A new family of ICPC designs was recently developed by researchers at the University of Chicago and Colorado State University for Solar Enterprises International. These designs allow a relatively simple manufacturing approach to be used and solve many of the operational problems of previous ICPC designs. A low concentration member of this family of designs that requires no tracking was used in the demonstration. The 336 tube ICPC collector away was fabricated for the demonstration by Solar Enterprises International. This design and the fabrication approaches are described in (Duff et al, 1997) and (Winston et al, 1997). Double effect absorption chillers have been commercially available for almost ten years. These 2E absorption chillers
require substantially higher operating temperatures (around 155C) than single effect (1E) chillers (around 80C). However, 2E chillers produce nearly twice as much cooling for the same energy input. Moreover, ICPC collectors operate as efficiently at these higher temperatures as do more conventional collectors at lower temperatures. Thus, with the advent of the new ICPC design, it became possible to produce cooling with a 2E chiller using a collector field that was about half the size of that required for a 1E chiller with the same cooling output. Various commercially available 2E chillers are either steam, hot water or direct fired. Steam and direct fired 2E absorption chillers are available in sizes ranging from 20 tons to 600 tom and above. However, at the time the project was initiated hot water fired 2E absorption chillers were only available in 280 ton and larger sizes. Thus, since the solar collector array produced heated water, an off-the-shelf 20 ton 2E direct fired absorption chiller had to be modified to operate with hot water. 2. ICPC SOLAR COLLECTORS The demonstration solar collector array consists of three banks of eight modules. Each module has 14 ICPC evacuated collector tube solar collectors. The evacuated tubes in the 106 m 2 (1150 It 2) aperture area array were oriented lengthwise north to south and sloped at 15 degrees to optimize solar energy collection during the air-conditioning season. Each of the 336 125 mm (5 inch) diameter 2.78 meter (9 foot, 1.5 inch) long soda-lime glass evacuated solar collector tubes
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contains a thin wedge shaped fin absorber and matched CPC reflector running the length of the tube. The inside bottom half of the glass tube is silvered to form the CPC reflector. A 12 mm tube that is plugged at one end is bonded to the absorber. A small feeder tube is placed inside the 12 mm tube to allow fluid to flow into and out of the evacuated tube. The module manifolds are a concentric tube-inside-tube design as well. Two absorber orientations were produced, one with a vertical absorber fin and one with a horizontal fin. The north bank of the collector array consists of all horizontal fin tubes and the middle bank consists of all vertical fin evacuated tubes. The south bank contains a mixture of fin orientations. A crosssection of the two collector tubes illustrating the two orientations is shown in figure 1.
Prior to the demonstration, fourteen tube modules of each absorber orientation were tested on the Sandia National Laboratories two-axis tracking (AZTRAK) platform. This platform allows a non-tracking solar collector to be positioned at any angle between zero and 90 degrees in any orientation. See (Winston et al, 1997).
3. COOLING SYSTEM A McQuay 20 ton LiBr and water 2E chiller was modified for firing by solar hot water in the 145 to 165C range. See (CEC Final Report, 1998). Auxiliary energy was provided to the chiller generator or to storage from a 70 KW (240,000BTU/hour) hot water boiler. Depending on the state of the system at the time, auxiliary energy was used to boost the temperature of the hot water delivered to the generator or to increase the storage temperature. The collector loop and chiller generator loop were interfaced by a pressurized and insulated 3785 liter (1000 gallon) horizontal cylindrical storage tank. 4.
Fig. 1. New ICPC Design Used in the Sacramento Demonstration Showing both Vertical and Horizontal Fin Orientations
Fig. 2. Sacramento Demonstration Collection Efficiencies and Array Temperature for August 14, 1998.
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DATA ACQUISITION SYSTEM
The primary data acquisition system (DAS) was provided by the National Renewable Energy Laboratory (NREL). The DAS consisted of thermocouples, thermopiles, flowmeters and a data recorder which could be remotely accessed and downloaded via modem. The DAS was put into full operation in August 1998.
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5.
RESULTS
5.1 Collection Efficiencies Performance of the ICPC array has fulfilled the expectations of the Sandia testing. Conversion efficiencies in excess of 60 percent at insolation levels of 750 to 850 W/m2 and collector temperatures of 120 to 130C were attained. As can be seen in figure 2, results for a temperature range of 115 to 130C show close agreement with the Sandia tests (Winston et al 1997). Operating results are not yet sufficient to statistically substantiate superofity of one or the other absorber fin orientations. 5.2 Daily Performance The system was designed to allow the chiller to operate on solar energy most of the time. Several circumstances led to more input energy being required by the chiller and less solar energy being delivered to the chiller than had been originally planned: 1. The desired operation of the modified chiller was not attained until late in the cooling season. Consequently, before acceptable chiller operation was achieved, the chiller operated at a lower than design COP. Thus, more input energy was often required for a given cooling load. 2. Because perfecting the performance of the chiller took longer than expected, the automated valves and heat rejection system intended for summer 1998 operation were not installed. This resulted in manual operation of solar energy system, including daily coveting and uncovering the collector array, which often reduced the amount of solar energy collected. 3. Collection efficiency varied with flow rate to the collector array. At the lower flow rates that were used much of the time, collection efficiency was reduced. In figure 2 the flow to the collectors was reduced after 11:00 where a reduction in performance, as compared to the Sandia predictions, can be observed. 4. A leak developed in the storage tank manhole cover gasket. This leak was not fixed until after cooling operation had ceased and the tank could be drained and inspected. The leak caused a loss of pressure in the tank overnight. The subsequent vaporizing of the stored fluid produced greater storage energy losses and lower system start-up temperatures the next morning. In spite of these limitations, exceptional system performance was often achieved. Even on some days when these limitations were severe, the daily performance of the system was exceptional: 1. Daily energy collected when operating at around 150C often reached nearly 50 percent of the total solar energy falling on the collectors. 2. Daily energy delivered to the storage tank when operating at and above 150C often reached nearly 40 percent of the total solar energy falling on the collectors. During the peak insolation months next summer and
under automatic control, a high value of this figure of merit is expected to be achieved routinely. 3. The chiller was brought up to the design target late in the air-conditioning season. Consequently, it had an opportunity to operate efficiently only well below maximum design load. Under these circumstances, it has produced daily COPs near and above 1.2. 4. Despite a generator energy draw that was greater than design, the chiller often satisfied the building cooling load while operating entirely or almost entirely on solar energy. Automated operation of the system has already been partially implemented for 1999 and the collectors no longer need to be covered and uncovered. Longer-term fully automated operation of the system is planned for the summer of 1999. As compared to other solar air-conditioning options, current performance of the system has been exceptional. The building air-conditioning load has been consistently met with a solar collector field that is about half the size of that required by more conventional collectors using 1E chillers. Even better performance is expected in 1999.
5.
ACKNOWLEDGEMENTS.
Mary Jane Hale, Pramod Kulkami, Prab Sethi, Don Osbom, Cliff Murley and Ranji George have all been especially supportive in this project. The California Energy Commission provided the primary funding for the research. Matching funding were provided by the National Renewable Energy Laboratory, Sacramento Municipal Utility District, South Coast Air Quality District, Sun Utility Network, Bergquam Energy and Thermal Energy Systems Specialists. REFERENCES Garrison, J. D. (1979) Optimization of Fixed Solar Thermal Collectors, Solar Energy, v23. Snail, J. J., O'Gallagher and R. Winston (1984) A Stationary Evacuated Collector with Integrated Concentrator, Solar Energy, v33. Duff, William S, R. Duquette, Roland Winston and Joseph O'Gallagher (April, 1997) Development, Fabrication and Testing of a New Design for the Integrated Parabolic Evacuated Collector, Proceedings of the ASES/ASME Solar Energy Forum, Washington D. C. Winston, R, J. J. O'Gallagher, William S. Duff and Alberto Cavallaro (April, 1997) The Integrated Compound Parabolic Concentrator: From Development to Demonstration,
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Proceedings of the ASES/ASME Solar Energy Forum, Washington D. C. "Demonstration of ICPC Solar Energy Collectors for Space Heating and Cooling Using a 2E Absorption Water Chiller" (December 1998) California Energy Commission Project 50095-018 Final Report by Bergquam Energy.
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INDIRECT EVAPORATIVE COOLING THROUGH A CONCRETE CEILING Baruch Givoni, Sukanya Nutalaya UCLA, Dept. Of Architecture and Urban Design, 405 Hilgard Ave. Los Angeles, USA Tel: +1-310-8252067, Fax: +1-310-8258959, Email: [email protected]
Keywords: Hot Humid Climate, Passive Cooling, Indirect Evaporative Cooling, Cooling Towers, Ceiling Heat Exchanger. ABSTRACT - Two experimental cells were used in this study. One cell remained without any cooling (control). The other cell (test cell) was cooled by the system reported in this paper. The cooling system consisted of a pond of cooled by water circulating within a concrete ceiling, acting as a heat exchanger. The water was cooled by a shower at the top of an open shaft, re-circulated by a pump. The use of the ceiling as a heat exchanger facilitates convective down flow of the cooled air and maximizes the radiant exposure of the inhabitants to the cooled ceiling surface. The performance of the cooling system was tested under two levels of heat capacity: first in the original configuration (described below) and later with added internal mass (concrete blocks). In hot climates the indoor average temperature may serve as the best indicator of the thermal performance of a passive building. On the basis of the experimental data a formula has also been developed, expressing the indoor maximum temperature of the cooled cell as a function of the outdoor average air temperature. INTRODUCTION One of the passive cooling options available in a hot humid climate is indirect evaporative cooling, which lowers the temperature without raising the indoor absolute humidity. The cooling system described below provides a design solution of this option. When fine drops of water, having a very large surface area, are sprayed vertically downward like a shower from the top of an open shaft, the falling water entrain a large volume of air, creating an inertial air flow down the shaft. The water is collected in a small pond at the bottom of the shaft and is pumped back to the shower head. The evaporation from the free drops cools the water, as well as the air in the shaft, to a level close to the ambient Wet Bulb Temperature (WBT). As any evaporation from the droplets takes place in the free air stream, any type of water, brackish or even sea water if available, could be used for cooling in this system. In fact, no difference was found in the cooling performance of the system when tested with either fresh or sea water. This system enables both direct and indirect evaporative cooling of buildings. Direct evaporative cooling utilizes the humid cool stream created by the falling drops, which is directed into a building. Direct evaporative cooling is applicable mainly in hot dry regions (Givoni & A1 Hemiddi 1995). The cooled water in the pond can be used for indirect evaporative cooling, by circulating the water, in a closed loop, through a water-to-air heat exchanger inside the cooled space. The best position of the heat exchanger is above the cooled space, in order to facilitate convective down flow of the cooled air and to maximize the radiant exposure of the inhabitants to the cool surfaces of the heat exchanger. Indirect evaporative cooling is applicable mainly in hot humid regions.
Such cooling system with specialized heat exchanger (a radiator) was tested and reported previously (Givoni et al. 1995). However, a ceiling cooled by water flow in embedded tubes can also serve as an architecturally integrated cooling element. The performance of such a system is reported in this paper. The fact that the shower system provides the same cooling performance with sea water as with fresh water has profound implication for its applicability in desert coastal regions, where fresh water is in short supply while sea water availability is unlimited. In this respect this system can be applied in places where other evaporative cooling systems, which need high quality water, are not applicable. In hot climates the indoor average temperature may serve as the best indicator of the thermal performance of a passive building. Therefore, in the analysis of the experimental data, the main emphasis is on the indoor maxima which can be obtained with this cooling system. On the basis of the experimental data a formula has been developed, expressing the indoor maximum temperature of the cooled cell as a function of the outdoor average air temperature.
The Experimental Setup The experimental setup, at UCLA, consisted of two similar test cells (available from a previous study), each 1.2xl.2x3.6 meters. One cell was used as a control (without any cooling) and the second cell (Test-cell) was cooled. The walls and floors of the two cells were identical: insulated sandwiches of plywood. The roof of the control cell was (the original) insulated sandwich, while the roof of the cooled cell was modified. It was built of concrete, insulated to provide the same U value as the roof of the control cell. Budget limitations did
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not enable to construct a concrete roof also for the control cell, so the thermal masses of the two cells were different. However, the Heat Loss Coefficients (UA) of the two cells, which are not affected by the difference between the heat capacities of the cells, were measured and were found to be very similar, 11.3 and 11.0 W/C, for the Control and the Test cells, respectively. Consequently, it was possible to calculate, from the differences of the indoor average temperatures of the two cells, the average diurnal cooling performance of the system. In the second stage of the study the same amount of additional mass, in the form of 169 small concrete bricks (5xlOx20 cm), was added to each cell. In the third stage, the two cells were ventilated during the night hours while the cooling system operated continuously. Thus, the cooling performance of the
Figure 1: Embedded plastic tube circulating cooled water Figure 2 shows the data acquisition system and the measuring points.
Figure 2: Data Acquisition System and Measuring Points
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system was measured under three configurations of the cells: First in the original construction, then with the added mass, and later with night ventilation. The shower tower was 3.6 meter-high. The tower was made of 5 cm. (2") rigid (polystyrene) insulation and contained 4 heads of nozzles on the top of shower tower. The pond (123x123x76 crn) was built of polystyrene sheets. The inside of the pond was a water-proved plywood, with two layers of plastic sheets inside forming the actual pond. The construction of the concrete roof with the embedded plastic tubes is illustrated in figure 1. Concrete bricks, 5 cm thick, were laid over a thin steel plate. Then, plastic tube circulating the cooled water was installed, and a second (5 cm) layer of concrete was poured, to create a 10 cm thick roof.
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Overall Performance of the System with Floating Condition Figure 3 (A to C) shows samples of the diurnal temperatures during the cooling in the original configuration (3a), with the added mass (3b) and with the night ventilation (3c).
Figure 3: Diurnal Temperature Patterns
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Table 1 shows the averages of the maximum, minimum and average temperatures, measured in the different configurations of the study. In this paper only the average temperatures will be discussed in details. Table 1: Summery of Grand Average of Maximum, Minimum and Average Temperature Conditions A)
Control- No Added Mass
AVG.
DBT
Control
Cooled
16.7
19.0
19.2
MAX. 21.8 29.0 22.6 MIN. 10.1 10.1 14.8 23.5 23.7 20.0 B) C o n t r o l - With Added AVG. Mass MAX. 26.1 29.2 25.4 MIN. 15.2 18.6 21.7 AVG. 19.4 21.8 19.4 C) Cooling- No Added Mass MAX. 24.2 31.5 22.7 MIN. 15.0 15.1 17.0 25.8 22.7 22.6 D) C o o l i n g - With Added AVG. Mass MAX. 28.8 31.2 24.5 MIN. 17.7 21.5 20.7 27.4 24.6 AVG. 27.0 E) Cooling With Mass and Night Ventilation 34.0 27.0 MAX. 34.6 22.4 22.3 MIN. 20.9 Note - The cooling effect is calculated by (Control - Cooled) temperature drop
Average Temperatures Under floating conditions the average of the "test" cell was higher by 0.2 ~ then the control cell. With cooling in the original configuration it was lower by 2.4 o C. With the added mass the cooled cell's average was lower by 3.1 C and with night ventilation the temperature reduction was 2.8 C.
Maximum Temperatures As the two cells had different masses, their maxima, as well as their minima, can not be compared directly. However, the differences between the DBT and the cooled cell maxima in the different configurations can be compared. In the floating conditions without the added mass the maximum of the "test" cell was higher by 0.6 ~ then the outdoors' maximum. With added mass, but without cooling, the indoor maximum was lower than the outdoors by 0.7 C, reflecting the suppression of the indoor swing by the mass. With cooling in the original configuration the maximum of the cooled cell was lower by 1.5 ~ than the control. With the added mass the difference was 4.3 ~ and with night ventilation it was 7.6 ~
Minimum Temperatures Under floating conditions without the added mass the minimum of the "test" cell was higher by 4.7 ~ then the outdoors' minimum. With added mass, but without cooling, the indoor maximum elevation above the DBT was 6.5 ~ again reflecting the suppression of the indoor swing by the mass. With cooling
Effect
Pond
Pond Dep.
18.0 20.0 16.3 20.3 22.5 18.1 21.8 23.8 19.6
-1.4 -4.2 +1.3 -2.3 -6.3 +0.4 -5.2 -10.8 +1.3
-2.4 -8.8 +1.9 -3.1 -6.7 +0.8 -2.8 -7.0 +0.1
in the original configuration the minimum of the cooled cell was 2 ~ higher than the DBT, and with the added mass the elevation was 3 ~ With night ventilation the indoor minimmn elevation was 1.4 ~
The Cooling Effectiveness of the System Without cooling (floating conditions) the indoor averages of the two cells were almost identical and about 2.5 ~ above the outdoors' average, as a result of solar energy absorption in the envelope of the cells. With cooling, the indoor temperature of the cooled-cell was lower than that of the control. The cooled pond is the source of the cooling and thus, the lowest attainable average indoor average temperature would be that of the cooled water (Tpond) average. Consequently, the overall thermal effectiveness of the cooling system was calculated by the formula:
Effectiveness = (Tcontrol - Tcooled) / (Tcontrol -
Tpond) The experimental thermal effectiveness of the system, in the configurations without night ventilation, it was about 0.6. The average cooling energy provided by the system can be calculated from the differences between the average indoor temperature of the cells and the average DBT, and the cell's BLC:
Cooling- (Tcontrol- Tcooled) * BLC; (Whr/Day)
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Formula Predicting the Maximum Temperature of the Cooled Cell
The indoor maximum in a given configuration is almost parallel to the pattern of the DBT average. Its relationship to the outdoors' maximum is weaker.
Figure 4 shows the daily indoor maximum and the DBT maximum, average and minimum temperatures, during the various periods with the different configurations. It illustrated the following relationships between the patterns of the indoor maximum and the DBT:
b) The day to day changes in the indoor maximum, especially in the configuration with the added mass, are smaller than the corresponding changes of the outdoors' average.
40
35
30
25 U
15
<
A A
- Original
9
<
Condition
B -
B
,,,~-
< C
Added Mass
C
Mass and Ventilation -
Time (Daily) I~Max.
Indoor mGrandAVG.DBT
~
A V G . DBT ~ M A X .
DBT - - - - - - G r a n d M A •
DBT - ~ - - G r a n d M I N
•
M IN.DBT I
Figure 4: Daily Indoor Maximum and DBT Maximum, Average and Minimum These observations form the basis for the following formula for predicting the indoor maximum of the cell cooled by this system: Indoor M a x . GrDBTavg)
GrDBTavg
+
DelT
+
k*0DBTavg-
V~rhere: DBTavg = daily average of the DBT. GrDBTavg = Period's Average of the DBT average; DelT = Elevation of the period's average of the indoor maximum; K = Ratio of the day to day changes of the indoor maximum to the corresponding changes in the DBT average;
The values of DelT and k for the differentconfigurationsare:
Configuration Without added mass With added mass With mass and fight ventilation
DelT(~ 3.2 1.9
k 1.0 0.7
0.6
0.8
Figure 5 shows the measured and the calculated (by the above formula) maximum temperatures, during the periods with the different configurations, with good agreement between them. Figure 6 shows the correlation between the measured and the computed maxima, with the data of the different configurations marked with different symbols. The overall correlation coefficient is 0.968.
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Correlation Coefficient = 0.968 - - A - Original Conditions: Den" = 3.24 C, k = 1.0 C - - B- Added Mass: Deit = 1.94 C, k = 0.7 C - - C- M a = and Ventilation: DelT = 0.61 C, k = 0.80 C
29
.
_a 181 ~ ~
"re
27
era U
_
25
~C)24
X
X
j
Computed MAX. Indoor - Grand DBT + DelT + (It*lDBT-Grand I)BT)) A
9
' 4 ' ' 1 ' ' 10 '13 '16 '19 '22 '25 ' ; ' 3 T
B
== 4..._
C ~
'34 '37 '40 '43 '46 '49 '52 '55 '58 '6T '64~
T~e (Days) "Measured
~
A
23
:
~
4
-'-
21 2O
. 2O
l a coo~
22
.
. 22
,, c o o ~ t , M ~
. 23
. 24
. ~,~ 25
.
. ~ 26
. 27
~ 28
~ 26
30
x CoolingWilh Mau and NightVent m ComputadMAX.lndoorI
M a x . I n d o o e C o m p u t e d M A X . IndoOr
Figure 5: Measured and Calculated Indoor Maximum Temperatures
Figure 6: Correlation between Measured and Computed Indoor Maxima
Scope and Limitations of the Formula
REFERENCES The specific values of the constants associated with the different configurations of the test cell are applicable, of course, only to test cells of similar design. However, the general form of the predictive formula can be used as a model for a predictive formula for the performance of real buildings. It should be noted that predictions of a similar formula were found to have high correlation with measurements in test buildings built of "real" materials, 5xSx2.Sm, in an extensive monitoring program in Pala, Southern California (Givoni 1994, 1998).
* 1995 "Applicability of a "Shower" Passive Cooling Tower in a Hot Dry Climate" (with N. A1-Hemiddi). Proceedings, ASES Congress 1995, Minneapolis. pp.143-148. * B. Givoni, 1994. Passive and Low Energy Cooling of Buildings. Van Nostrand Reihold. N.Y. * B. Givoni, S.Yajima and S. Nutalaya: 1997 Indirect Evaporative Cooling for Hot Humid Climate by the "Shower" Cooling Tower. With S Yajima & S. Nutalaya. ISES World Congress. S. Korea. * B. Givoni 1998 "Effectiveness of Mass and Night Ventilation in Lowering indoor daytime temperatures". Energy and Buildings. Vol. 28, No. 1. pp. 25-32.
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EXPERIMENTAL STUDIES ON A HYBRID DRYER Gauhar A. Mastekbayeva, Chandika P. Bhatta, M. Augustus Leon and S. Kumar Energy Program, Asian Institute of Technology, P.O. Box 4, Klong Luang, Pathumthani 12120, Thailand Tel: +66 2 524 5439, Fax: +66 2 524 5439, E-mail: [email protected] Abstract- A solar-biomass hybrid tunnel dryer has been designed and fabricated. A biomass stove-heat exchangerchimney using briquetted ricehusk as fuel, complements the solar tunnel dryer and thus extends the working time of the dryer. Experiments have been conducted to test the performance of the dryer, and chilli and mushroom have been dried. During the load test conducted for chilli, 19.5kg of ripe, fresh chilli, with a initial moisture content of 76% (wet bulb) was dried to a final m.c of 6.6% (w.b) within 12 hours. Similarly, the moisture content of 21kg of fresh harvested mushroom was reduced from 91.4% to 9.8% during 12 hours of drying. The results indicate that for both the products, drying is faster, and is within 12 hours in normal sunny weather, against 2-3 days in 'solar-only' operation of a tunnel dryer and 3-5 days in open sun drying. This paper evaluates the performance of the hybrid tunnel dryer against 'solaronly' operation of the same dryer and open sun drying. Efficiency of the dryer during its two mode of operation has been estimated and compared with other similar dryers.
1. INTRODUCTION Drying is an essential process in the preservation of agricultural products. Various drying techniques are employed to dry different food products. Each technique has its own advantages and limitations. Industrial drying offers quality drying whereas its high cost limits its use. Open sun drying suffers from quality considerations though it enjoys cost advantage. Choosing the right drying system is thus important in the process of drying agricultural products. Especially, in the tropical regions, where some crops have to be dried during rainy season, special care must be taken in choosing the drying system. Studies comparing traditional sun drying and other solar drying techniques show that the use of solar dryer leads to a considerable reduction of the drying time and to a significant improvement of the product quality in terms of color, texture and taste. Besides, the contamination by insects and microorganisms can be prevented. Solar tunnel dryers are a class of solar dryers that have been successfully tested under field conditions in about 30 countries under different climatic conditions, drying numerous agricultural commodities ranging from fruits, vegetables, root crops, oil crops, medicinal plants to fish and even meat (Grupp et al., 1985; Lutz et al., 1987; Sodha and Chandra, 1994; Esper et al., 1994; Schrimer et al., 1996). A solar tunnel dryer was first introduced by the Institute of Agricultural Engineering in the Tropics and Subtropics of University of Hohenheim (Germany), for use in the tropical region. The dryer proved to be successful in drying a variety of agricultural products including tropical fruits and vegetables, due to its economic viability (Esper et al., 1996). This dryer was however designed with a capacity that is suitable for use in single farm and small cooperatives. In an effort to adapt its design for small rural farmers, a scaled down version of the tunnel dryer was fabricated (Mastekbayeva et al., 1998). To reduce its dependence on solar radiation for operation and to improve the quality of
drying, a biomass stove - heat exchanger system was incorporated in this dryer, thus converting it to a hybrid dryer. The biomass stove - heat exchanger system was designed mainly to complement the solar operation of the dryer, and to sustain the drying process even during cloudy weather. However, it can also be used to extend the period of drying beyond sunshine hours, and perhaps during night as well, while drying high value addition crops. This paper evaluates the performance of the hybrid tunnel dryer against 'solar-only' operation of the same dryer and open sun drying, for comparison. The experimental set-up, data and its analysis, and results are described. 2. DESIGN OF THE HYBRID DRYER AND EXPERIMENTAL SET-UP The prototype solar-biomass hybrid dryer consists of a flat plate solar collector and a drying tunnel, fabricated as a single unit. The dryer is 1.8m wide, with a collector length of 4m and a dryer length of 4.25m. Glass wool insulation, with a thickness of 4cm was used to reduce the heat losses from the bottom of the collector (absorber). A 0.2mm thick UV stabilized polyethylene sheet was used as glazing. A cross flow shell and tube type heat exchanger was incorporated at the ambient air inlet to the collector. Five fans, each of 14W capacity, were used to force ambient air into the dryer, through the heat exchanger 'shell'. The fans had an air handling capacity of 130m3/hour each. The 'tubes' of the heat exchanger were connected to a biomass stove at one end, and a chimney at the other. The design of the biomass system was based on the following considerations: (i)
The heating will be indirect, i.e, flue gas from the biomass stove and the drying air would not be mixed. This will
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protect the product being dried from contamination by the smoke, soot and ash of the flue gas. (ii) The temperature of air heated by the heat exchanger, entering the dryer, would be in the range of 65-70~ This is based on the allowable maximum drying temperature for most of the tropical fruits and vegetables. (iii) Temperature control of the drying air would be possible, by controlling the combustion in the stove, by opening or closing the primary air supply gate in the stove. (iv) Biomass operation could be carried out for extended periods of time, unattended. The stove was designed to operate continuously for about one hour for a single fuel loading, with briquetted rice husk as fuel.
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inner diameter 44mm were used. Tubes were arranged in a staggered manner, as shown in figure 1. The biomass stove-heat exchanger unit, with attached chimney is shown in figure 2. The heat exchanger was insulated with 100mm thick rockwool and clad with lmm thick aluminium sheet, to reduce thermal losses. The biomass stove was insulated with tastable refractory mortar, along the inside walls.
For the above considerations, a channel stove of rectangular shape was chosen, and the main dimensions of the stove were designed. The thermal energy required to supply hot air (by the heat exchanger) at 70~ for drying was found to be 12.7 kW. From the calorific value of briquettes (18.8MJ/kg), the rate of fuel consumption was estimated as 2.44 k g ~ . A mild steel stove was fabricated with a width of 0.3 m, length of 0.275 m, and height of 0.4 m. A grate punched with 44 holes of 1.5 cm diameter was used to increase the efficiency and quality of combustion (Mastekbayeva, 1998). The stove used a high chimney to produce hot flue gas in natural draught. The chimney was designed for a flue gas flow rate of about 139 m3/hr. A rectangular shape was chosen for the chimney design with a cross section of 0.275 mx 0.16 m and the height of chimney was calculated to be about of 1 m.
Figure 2: Biomass stove-heat exchanger unit, before insulation The completed hybrid dryer is shown in figure 3.
A cross-flow shell and tube heat exchanger was chosen, with heated air in the shell side. Noting that the maximum permissible temperature for fruits and vegetable drying is about 60~ the temperature at the outlet of heat exchanger was to be not less than this value. During the combustion process, as volatiles bum at about 600~ (Baldwin, 1987), this was taken as the design inlet temperature of flue gas to the heat exchanger. The flue gas outlet temperature was assumed to be about 300~ A rectangular shape was chosen for the shell side to connect the heat exchanger with the solar tunnel dryer, with the following dimensions: length: 1.72m, width: 0.6m and height: 016m. To provide an estimated total required heat transfer area of 1.74m2, eight galvanised iron (GI) pipes with outer diameter 50mm and
Figure 3: Solar-biomass hybrid dryer 3. INSTRUMENTATION AND MEASUREMENTS
Figure 1: Arrangement of tubes in the heat exchanger
Air probe sensors (thermocouples) were used to measure the temperature of air at different points. The K-type thermocouples (JB 10) were calibrated before fixing them in the tunnel dryer. Surface probe sensors were used to measure the surface temperature of the collector plate, and air probe sensors, to measure the dry bulb and wet bulb temperatures at various points inside the collector and dryer parts of the tunnel. The locations of the sensors are given in figures 4 and 5.
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_
.
Chimney
~
Experiments were conducted to evaluate the performance of the dryer at no-load and at full-load conditions. Two separate full load experiments were conducted, one for chilli and the other for 'earlobe' mushroom.
~--]
4. NO-LOAD EXPERIMENTS
I I --
iom s
I
No load tests were conducted with five AC-operated fans (14W each) at the heat exchanger/dryer inlet to know the temperature gradient between the stove side and the chimney side of the tunnel dryer, as well as the temperature profile along the tunnel dryer.
T,, m iont
stove
Fig. 4: Location of temperature sensors at the biomass stove - heat exchanger - chimney
Chimnev side
iA
Collector part
J
i B
Dryer part
Ci
Three kg of rice husk briquettes was burned in the biomass stove during one hour of operation. Figure 6 shows the temperature profile across the tunnel, at the inlet of the dryer (exit of the collector). The temperature profile along the width of the dryer inlet was observed. The maximum temperature reached 57.9 ~ at the stove side, while at the chimney side, it was 38.5~ At the middle of the dryer inlet, the temperature was 54.5 ~ The average temperature rise at the chimney side, middle, and stove side were 4.75~ 14.93~ and 18.32~ respectively.
T15 'i...; :
i T17
*
T10 T12
Tll i T14 ~
T18
T16 * i T19 i
70
:
A
T]3 *...
6O
5o 4o
i
_
$
~.lv
t
~
Stove side
3o
* Dry bulb (air probe) Wet bulb (air probe) Surface probe
: T13 - T14 -" T15
E 20 I0 0
I
.....
25
Section A: Collector inlet Section B: Dryer inlet/Collector exit Section C: Dryer exit
40
55
70
T i m e (minutes)
Figure 6: Temperature profile at dryer inlet (Location of temperature sensors are shown in figure 5)
Figure 5: Location of surface and air probe, dry bulb and wet bulb temperature sensors in the hybrid dryer
The airflow rate inside the tunnel was measured with a vane type anemometer. Average airflow was 632 m3/hour with five fans
The uneven temperature dislribution across the tunnel would affect the dryer performance, due to non-uniform drying of the product. Faster drying will occur at the stove side, while drying will be slow at the chimney side. This could result in over drying at the stove side, and under drying at the chimney side. To overcome this, mixing fans were used to mix the hot air so that the air temperature is uniform across the tunnel, before it enters the
working Coiomass operation), and 385 m3/hour with three fans
dryer.
The temperature data were recorded in a 21X Campbell data logger, at five minutes intervals. Global solar radiation data was measured using a pyranometer.
working (solar operation). Humidity of air inside the tunnel dryer was estimated using the wet bulb temperature data recorded. The moisture content of the product during the drying process was determined using oven method, by taking samples periodically.
Six kg of rice husk briquettes were used in the biomass stove, which burnt for about two hours, and temperature at various point of the dryer was observed. The experiment was performed during 7:30 to 9:30 a.m. Two DC fans were used at the collector inlet, just after the heat exchanger, to facilitate effective mixing of air. The fans have an air handling capacity of 80m3/hour each.
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The maximum temperature recorded at the stove side of the tunnel was 50.9 ~ while that at the other end (chimney side) was 45.4~ (figure 7). At the middle of the dryer inlet, the temperature was 48.5 ~ The average temperature rise at the dryer inlet across the tunnel at the stove side, middle, and at the chimney side were 13.0~ 15.1~ and 17.8~ respectively.
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of dried chilli were produced in Thailand in 1996 (Kumar and Wattanapong, 1997). Chilli is dried as whole, like vegetables, without any chemical pretreatment. The full load capacity of the solar tunnel dryer is about 80 kg of raw chilli per batch. For the experiment, the dryer was loaded with 19.5kg of chilli as ripe fruits, by spreading them inside in a single layer. They can be spread directly on the dryer, in contact with the metal sheet that carries heat from the absorber to the dryer. The product can thus utilise the heat absorbed by the collector (during solar operation) more efficiently. The drying was started at 6:00 a.m., using the biomass stove. Six kg of rice-husk briquettes were loaded in the stove, and the fuel was ignited. The heat exchanger fans, mixing fans, and the data logger were switched on. The biomass operation was continued until 9:00 am. After 9:00 a.m., and until about 5:00 p.m., the system was utilized as a solar dryer. The biomass stove was started again, and the drying continued for one more hour.
Figure 7: Temperature profile across dryer inlet with mixing fans (Location of temperature sensors are shown in figure 5) The difference in temperature rise at the dryer inlet, across the tunnel reduced considerably from 13.57~ to 4.73~ without and with the mixing fans respectively.
Figure 8 shows the drying curves for chilli dried in the hybrid dryer, and compared with open sun drying. The final moisture content of 6.6 % (wet bulb) was achieved within 12 hours with hybrid drying, while open sun drying took about 5 days.
5. VEGETABLE DRYING Full-load experiments were conducted to study the capability of the dryer to operate beyond the sunshine hours. The dryer was operated from 7:00 a.rrL until 9:00 a.m. by biomass, when solar radiation was not enough for drying. It was then operated as a solar dryer from 9:00 a.m. until 5:00 p.m. Biomass operation was restored again, atier 5:00 p.m. The drying was stopped when the products reached their final moisture content. The products to be dried were spread on aluminium trays of 0.85m x 1.0m size, fabricated with angles and mesh. Control samples were dried simultaneously in open sun under the same weather conditions, for comparison. Rice husk briquettes were burned in the stove for the biomass operation of the drier. Airflow rate, fuel (biomass) consumption, and the current and voltage across the fans were measured during the experiment. The samples from the dryer were taken at twohour intervals, to measure the moisture content. During bio-mass operation, moisture content of the product to be dried was measured at one hour interval. Open sun drying sample was weighed every two hours on the first day of drying and every onehour the next day, to measure the moisture content. 5.1 Chilli Chilli, both fresh and dried, is a popular ingredient in food, and has high levels of protein and vitamins (Thanvi and Pande, 1987). It is estimated that 332, 079 tones of fresh chilli and 1,171 tones
Figure 8: Drying curves for chilli with hybrid drying compared to open sun drying 5.2 'Ear lobe' Mushroom Mushroom is an important agricultural food product used in soups and other dishes. In Thailand mushroom is used locally, and is also exported in dried form. The export of dried mushroom from Thailand was 25,173 kg in 1994 (DoEP, 1994).
Mushrooms are dried at a temperature of 60-70~ to a final moisture content 8-13%. Fresh mushrooms are often boiled before drying, for good color and aroma. Various types of mushrooms that can be dried are: Escherichia coli, Salmonella and Staphylococcus aureus. 'Ear lobe' mushroom is usually used in dried form.
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Ear-lobe mushroom is a high moisture content product, and has an initial moisture content of 90-92 % (w.b) when fresh. Market price of fresh ear-lobe mushroom was 15 Baht/kg while the dried ones cost 190 Baht/kg ]. For the load test, 21 kg of fresh ear-lobe mushroom was spread on the trays in a single layer, and the trays were placed inside the dryer. The fuel (ricehusk briquettes) in the biomass stove was ignited, and the experiment was started at 7:00 a.m. The biomass operation was stopped at 9:00 a.m., and the system was then utilized as a solar dryer until about 5:00 p.m. The biomass stove was started again, and the drying continued until 7:00 p.m. The experiment was repeated with the same quantity (21kg) of ear-lobe mushrooms, but with 'solar only' operation of the same dryer, on a day having similar weather conditions. It took two days of drying for the mushroom to attain almost the same final moisture content. Control samples dried in open sun took three days to attain almost the same final moisture content. The drying curves are presented in figure 9, for comparison.
100 90
tim"
80 i=
70
--
60
I
~~
I
I
I
-- Solar Drying
--,~--Open sun drying Hybrid Drying
,0
X
20
The drying efficiency of the hybrid tunnel dryer shows how effectively the input energy to the dryer (biomass energy and solar radiation) is used in drying the product. The system drying efficiency is defined as the energy used to evaporate the moisture in the product divided by the energy input to the dryer, or n, = ( ~ )
/ [(za, + P/) + (r~ * c . v ) ]
where W is the weight of water evaporated from the product, L is the latent heat of evaporation of water, mb is the mass of biomass fuel used in the stove, and C.V is the lower calorific value (LCV) of the biomass fuel. The LCV of rice husk briquettes was estimated as 11.69 MJ/kg. The total electrical energy (Pf) required to run the three AC fans (during solar operation) is estimated at about 0.36kWh (1.3 MJ) for one day of operation (8.5 hours), which is approximately 0.5% of the daily total solar energy incident on the system, and is therefore negligible. The total solar radiation for the first day was 17.3 MJ/m 2. The average drying efficiency of the solar-biomass hybrid tunnel dryer is thus estimated to be about 8.8% during chilli drying (Table 1).
~ 5o
8 E
The dryer used five numbers of 14W/AC fans in its heat exchanger and additional two fans of 4.7W/DC for mixing air inside the tunnel. The electrical power requirement during biomass operation amounts to 79.4W. During solar operation, however, only three fans were used in the heat exchanger, and the mixing fans were switched off. The power consumption thus reduced to 42W during solar only operation.
Table 1: Performance of the hybrid tunnel dryer during its solar and biomass operation, while drying chilli
10 0
,,
0
5
10
15
20
25
30
35
40
45
50
55
60
Time (hours)
Figure 9: Drying curves for 'ear-lobe' mushroom dried in the hybrid tunnel dryer compared to solar drying and open sun drying
Drying time (hours) 6:00-9:00 9:00-17:00 17:00-18:00
6. DRYING
Drying efficiency Total energy input to the Solar only Biomass Hybrid % % % System (M J) 105.2 256.9 35.1
-
17.8
6.1
-
-
3.6
8.8
EFFICIENCY
Drying efficiency of the tunnel dryer has been estimated by considering solar energy input to the collector and dryer parts (Sodha et al., 1987). The following parameters were considered for estimating the drying efficiency for both cases: a) Total radiation incident on the collector (considering the total duration of solar drying), b) Total electrical energy input to the fans, c) Total heat input by the biomass stove (during biomass operation) and d) Initial and final weight of the product and thus the amount of water evaporated. 1US$=37 Baht; March 99
The total capacity of the solar tunnel dryer is estimated to be about 80 kg of fresh chilli, while only 19.5 kg was loaded for the experimentation. Therefore, the drying area was not used fully, which explains the low efficiency of the drying system. This was also observed from the relative humidity values of the air leaving the dryer, which were fairly low. This indicates that the drying potential of the air was not fully utilised. Similar efficiency figures for drying of ear-lobe mushroom have been given in table 2. The total solar radiation for the day was 13.13 MJ/m 2.
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Table 2: Performance of the hybrid tunnel dryer during its solar and biomass operation, while drying ear-lobe mushroom Drying time (hours) 7:00-9:00 9:00-17:00 17:00-19:00
Drying efficiency Total energy input to the Solar only Biomass Hybrid % % % System (M J) 70.14 195.0 70.14
-
16.8
17.2
-
-
0.75
14.4
It may be noted that the drying efficiency is much higher while drying ear-lobe mushroom than while drying chilli. This is due to the much softer skin of mushroom (compared to chilli), which allows for easy diffusion of moisture through it to the drying air. The thickness of mushroom also contributed to faster drying, as the average thickness was only about 2mm. In general, the drying efficiency was reduced considerably during the final stages of drying. This is due to the much higher mass transfer resistance offered by the product after it is dried to a certain extent, which means that diffusion of moisture from inside the product to the surface of the product becomes much more difficult.
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Table 3" Summary of chilli drying with different solar dryers Type of dryer
1. Solar Tunnel Dryer (AC) 0Ylastekbayeva, 1998) 2. Solar Tunnel Dryer (DC/PV) (Mastekbayeva, 1998) 3. Hybrid Tunnel Dryer (AC) (Present study) 3. Low-Cost Solar Agricultural Dryer (Thanvi and Pande, 1987, India) 4. New Solar Dryer (Tiris and Dincer, 1994, Turkey)
Moisture Drying Average Quantity content(%), time temp. ofproduet Initial Final rise (~ dried (kg) 75.2
6.3 3 days
17.5
i 20 (25%
of full capacity) 74.9
6.8
2 days
17.5
20 (25% offull capacity)
76
6.6
12 ! hours
20
19.5 (25%
86.3
4.1 i 9 days
89.2 7.22
2.5 days
of full capacity) 28.2
10-15 (full capacity)
16
no data
7. RESULTS AND DISCUSSIONS The results indicate that drying of chilli and ear-lobe mushroom could be completed within 12 hours in a normal sunny weather (or even cloudy or rainy weather, when drying could be continued with biomass operation), against 2-5 days in 'solar-only operation' of a runnel dryer. Solar-biomass hybrid tunnel dryers seem to be an attractive and reliable alternative to open sun drying and solar tunnel dryers in the tropical climates of Asia.
The moisture content of exit air from the dryer indicates large untapped drying potential of the exit air. If this drying potential is used more efficiently by increasing the average humidity of the air leaving the dryer, apparently, more quantity of the product could be dried. It is estimated that about 40kg of mushroom can be dried with an average relative humidity at exit of 70%, for a total drying period of 12 hours.
The experimental investigations on drying chilli have been compared with different designs of solar dryers. The moisture content of raw chilli is usually in the range of 75-90%, while the dried chilli contains about 4-7% of moisture (w.b.). The drying time varied from 12 hours to 9 days depending on the weather conditions and dryer design, against 5-18 days required for open sun drying. The results of observations from various studies are summarized in Table 3.
8. CONCLUSION
The quality of the dried products in hybrid drying notably improved compared to open sun drying as well as solar drying due to the fact that drying was uninterrupted until the final moisture content was attained. This eliminated possible moisture reabsorption and mould growth during overnight storage of the product during open sun drying and solar drying. In certain products, cracks tend to develop within the product due to thermal stresses resulting from alternate heating and cooling of the product during day and night, with solar and open sun drying. This is also largely minimised with hybrid drying. It may be noted that the improvement in quality of ear-lobe mushroom in terms of taste and food value was distinctly recognised.
This paper describes the design and experimentation of a solarbiomass hybrid dryer. Separate experiments were carried out with two agricultural products chilli and ear-lobe mushroom in the hybrid tunnel dryer and the performance of the dryer was compared to 'solar only' operation of the same dryer, and to open sun drying. Considerable reduction in drying time is the major advantage reported with this hybrid dryer. While overcoming the limitations of solar drying during cloudy days, the solar-biomass hybrid dryer also enables drying during nighttime. The facilitating of continuous year-round operation of the dryer and the 60-80% reduction in drying time in comparison with open sun drying and solar drying, increases the utilisation of the dryer, and improves the financial viability of the tunnel dryer considerably. The financial support by the Swedish International Development Co-operation Agency (Sida) for this study in the framework of the project "Renewable Energy Acknowledgement:
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Technologies in Asia - A Regional Research and Dissemination Programme" is gratefully acknowledged. REFERENCES
Baldwin, S. F., (1987). Biomass Stoves: Engineering Design, Development, and Dissemination, Arlington: Volunteers in Technical Assistance, USA. DoEP, (1994), Exports by Commodity, Dec Department of Export Promotion, Bangkok.
1994. VI:
Esper, A., Hensel, O., and Muhlbauer, W. (1994), PV-Driven Solar Tunnel Dryer. Agricultural Engineering Conference, Bangkok, Dec. 6-9, 1994. Esper. A., Muhlbauer, W., Rakwchian, W., Janjai, S., and Smithabhindu, R., (1996) Introduction of Solar Tunnel Dryer for Drying Tropical Fruits in Thailand, Paper presented in the International Seminar on Financing and Commercialisation of Solar Energy Activities in South and East Asia, Kunming, China, August 24-31, 1996. Kumar S., and Wattanapong R., Evaluation of Solar Drying of Fruits and Vegetables in Thailand, Report submitted to ADEME, France, 1997. Lutz, K., Muhlbauer, W., Muller, J., and Reisinger, G. (1987). Development of Multi-Purpose Solar Crop Dryer for Arid Zones, Solar and Wind Technology, 4:417- 428. Mastekbayeva, Gauhar A., Performance Enhancement of AIT Solar Tunnel Dryer, Master Thesis, (August 1998), ET-98-1, Asian Institute of Technology, Bangkok, Thailand. Mastekbayeva, Gauhar A., M. Augustus Leon, and S. Kumar, (1988), Performance evaluation of a Solar Tunnel Dryer for Chilli Drying, Paper Presented at the 'ASEAN Seminar & Workshop on Drying Technology', 3-5 June, 1988, Phitsanulok, Thailand. Sehirmer, P., Janjai, S., Esper, A., Smitabhindu, R., and Muhlbauer, W. (1996), Experimental Investigation of the Performance of the Solar Tunnel Dryer for Drying Bananas, Renewable Energy, 7, 2:119-129. Sodha, M. S., Bansal, N. K., Kumar, K., Bansal, P. K., and Malik, M. S. S., (1987). Solar Crop Drying. Volume 11. CRC Press. Boca Raton, Florida. Sodha, M. S., and Ram Chandra, (1994), Solar Drying Systems and their Testing Procedures: A Review. Energy Conservation and Management. 35, 3: 219-267.
Tiris, M., and Dincer, I., 1994. Experimental Testing of a New Solar Dryer. International Journal of Energy Research vol. 18: p. 483-490
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COMBINED SOLAR HEATING AND RADIATIVE COOLING SYSTEM M. Meir, H. Stor~ts* and J. Rekstad Department of Physics, University of Oslo, P.O. Box 1048 Blindern, N-0316 Oslo, Norway Tel.: +47 22 85 64 69, FAX: +47 22 85 64 22, [email protected] Abstract - The radiative cooling potential of flat-plate coolers of polymer plastics has been investigated. The cooling radiators are part of a complete heating and cooling system with water as heat transfer medium. This system consists of solar collectors and radiative coolers with optimal thermal coupling to a hydronic floor system. The radiator's stagnation temperature of 5-7 K under ambient air temperature during night-time is demonstrated. The cooling power at ambient temperature found under favourable climatic conditions is ca. 40 W/m2. The system seems to be able to cover the need for cooling at locations where the night temperature decreases to 20~ with moderate relative air humidity.
1. I N T R O D U C T I O N A study of low cost polymer plastics as radiative cooling surface was carried out. The application was inspired by a solar heating concept [1] in which a collector of advanced thermoplastic materials is an important design factor. This concept includes a combined solar DHW- and space heating system, low temperature heating in terms of hydronic floor heating, large collector area and a large heat store. The collector was slightly modified in order to utilise it as a radiative cooler. Further the basic components of the heat distribution system are re-designed so that solar heating and radiative cooling can <<simultaneously, be applied for indoor climatisation of buildings. Hydronic floor systems were investigated for transferring heating and cooling to the building mass. The performance of floor cooling systems has been studied e.g. by [2], [3], [4]. The present system should cover the heating demand during the heating season and provide cooling during the summer months. There are world-wide many locations with heating demand in the winter time which require also cooling in the summer season. This cooling possibility represents an added value to the system with limited extra costs.
2. RADIATIVE COOLING Radiative cooling occurs when the energy transport in terms of electromagnetic radiation from a body is larger then the absorbed radiative energy emitted from the surroundings and from the sky. The equilibrium temperature of the body may be several degrees lower then the ambient temperature due to the spectral properties of the atmosphere. The spectral distribution of a black body with a temperature of T ~ 290 K has its maximum according to Wien's displacement law at " 10 lain. Ca. 30% of the radiation lays in the spectral region between 8 and 13 pro, where the atmosphere can almost be considered transparent. The "depth" of the atmospheric window depends on weather conditions, daytime and geographical location. The net cooling power of a black body surface under
such conditions may reach 60-120 W m "2 at ambient temperature [5]. If the other energy flows to the surface are small, the surface of the radiator will cool down to a temperature T ~ below the ambient air temperature Tmb until thermal equilibrium is established. For surfaces with backing insulation the temperature depression ? T= (T~mb-T,d)Can be under favourable conditions up to 5-7 K during night-time. The principle of "active" radiative cooling has found large applications in the so-called "roof-pond" concept, where water reservoirs are placed on flat roofs, cooled down during night time. These systems reveal a rather high cooling effect, but due to the exposure to the solar irradiation during day-time up to one third of the gained cooling is lost by conduction and convection through the top [6].
3. SYSTEM DESCRIPTION 3.1 Solar collector and radiator
The solar collector is a flat plate collector with water as heat transfer fluid. A simplified version of the collector without transparent cover sheet has been studied as radiative cooler. Figure (1) shows a graphic of the flat plate cooler. It consists of an extruded twin-wall sheet in polycarbonate and moulded end-cups to both ends. The potential of using thermoplastic materials as a cooling surface is given by the radiative properties. The emmisivity of polycarbonate in the thermal part of the spectrum from 3-30 Brn is close to the black-body emmisivity.
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replacing conventional building materials, the radiator area can be large and adapted to the cooling demand inside the building. The squaremeter price of such <<energy roofs>> lays in the range of conventional building and facade materials.
3.2 System design and operation
The present system is composed of the solar collectors, a large and unpressurized heat store, a heat distribution system in the building, a controller unit and an auxiliary heating source (fig. 2). There exits direct coupling between
Fig. 1. A twin-wall sheet o f thermoplastic material as radiative flat-plate cooler. The polycarbonate twin-wall sheets used as radiators have a channel cross section of 10 mm x 10 mm with a wall thickness of 1.0 mm. By filling the twin-wall sheet's channel structure with light weight clay granulate [7], a capillary effect is achieved which ensures optimum thermal coupling between the radiator's surface and the liquid circulating inside. The heat carrier is water which is pumped through the centre channel of the twin-wall sheet to the top of the radiator. The upper end-cup distributes the liquid to the single channels and, driven by gravity, the liquid trickles through the radiator and transfers heat. The water is collected in the lower end-cup and circulates back to the store. An minimum tilting angle of approx. 10~ is necessary for the drain-back function. A control unit stops the circulation when no cooling effect is obtained or freezing occurs. The liquid drains to the store and the radiators are filled with air during standstill. The radiator/collector loop is not pressurised. An important aspect of the concept is the radiator's/collector's design as a facade or-roof integrated building element available with fixed width and in different standard lengths. Due to the low specific costs, the savings obtained by building integration and by
Fig. 2.
Principle for a combined solar heating system with a collector made ofpolymer plastics.
the collector loop, the heat store and the floor system. By avoiding heat exchangers, the thermal losses are minimized, the operating temperature level is lowered [8] and the system efficiency for heating or cooling improved. Figure (3) shows the modifications necessary for a combined solar heating and radiative cooling system. A favourable solution can be to cover the south facing roof with solar collectors operating during daytime, while the radiative coolers are integrated in the north facing roof and charge the cooling store during night-time in the summer season (northern hemisphere). The collector and radiator loop are connected through the heat store to the same hydronic distribution system which is characterised by large heating/cooling surfaces, such as floors or ceilings.
3.3 Heat- and cooling reservoir
As illustrated in figure (3) a smaller section of the total store volume is used for solar DHW-preheating. The DHW-unit can be a tank-in-tank or a mantel tank solution. The larger section is a buffer store which serves as a cooling reservoir during the months which require indoor cooling and as a heat store during the
Fig. 3. An extension of the system in fig(2): principle for combined solar heating and radiative cooling system
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heating season. The buffer stores are stainless steel tanks which are open to atmospheric pressure and directly coupled to the radiator/collector loop and the hydronic "heat" distribution system in the building. In addition the store volume represents the drain-back reservoir for the liquid in the solar heating/radiative cooling system during system standstill. Compared to the roof pond concept this building integrated design avoids heating of the cooling store during day-time. The reservoir is removed from the roof, is smaller in size, easier to insulate, not in need of building reinforcement, or evtl. moving mechanics which cover the pond during day-time. Fig. 4. Hydronic floor system in concrete slab construction
4. RADIANT FLOOR COOLING In many countries hydronic radiant floor systems are widely used to heat buildings, but few are also used for cooling purposes. Floor heating finds application in residential, commercial and industrial buildings, especially in high ceiling spaces, floor systems have proved to be efficient [9]. More than half of the thermal energy emitted from the floor is radiant heat, which is exchanged with the surrounding walls, the ceiling and the occupants in the room and establishes an uniform thermal environment [10]. The convective heat from a floor system is directly delivered to the occupied zone at floor level, where the occupants are. Due to the high radiant heat output and the fact that the occupants are close to the heat source, the same distribution system could also be used for cooling. The heat transfer coefficient of the convective contribution is however much lower for the temperature range relevant for cooling. To evaluate the performance of radiant floor cooling it is important to consider comfort, cooling capacity, control and design. A limiting factor for living and working areas is the minimum floor temperature which should be secured for comfort reasons. The comfort temperature lays between 19-26~ depending on clothing, activity level, etc.. Another limiting factor is the dewpoint temperature in the room. In order to avoid supply temperatures which are too low and present a condensation risk, the difference between supply- and return temperature to the floor circuit should be ca. 3-5 K with a supply temperature in the range of 18-20~ The maximum cooling capacity is for most spaces about 50 Wm z for a floor system (typical in the order of 10 Wm -z cooling power per degree temperature difference between indoor and floor surface). Normal cooling demands correspond to a floor temperature 2-3 K under the acceptable indoor temperature (2225~ The cooling capacity, however, depends also on the floor construction, the distance between tubes, water flow rate and floor covering. Figure (4) shows an example of a typical floor system's construction. The floor pipes are integrated in a concrete slab layer placed on layers of sound- and thermal insulation supported by the bearing underfloor.
Quantitative experiments on dimensioning and design of floor systems are not presented in this study. Examples are e.g. given in [3] (floor cooling) and [14] (floor heating). Due to the limited cooling capacity of a hydronic floor system, the room temperature might, dependent on the climate, not always be effectively controlled. Applying radiative floor cooling is of advantage when the system is also used for heating in the winter. An important aspect of the concept is that it provides a soundless and hidden indoor climatisation system with no discomfort due to draft.
5. E X P E R I M E N T S The solar heating system mentioned is a commercial product, the development of the fiat plate coolers is in the prototype phase. The cooling effect of such cooling radiators were experimentally investigated with modified solar collectors, twin-wall sheets of a polyphenylen resin [ 11], and prototypes with polycarbonate twinwall sheets. Figure (5) shows schematically the experimental setup with the measuring equipment used in the experiments. The experimental data were recorded by a battery powered datalogger TT 128 [ 12]. The datalogger receives readings from the thermosensors, measuring the temperature of the liquid in the calorimeter at three different levels, the ambient air temperature and the wind speed. The read-out time is 10 min. for all temperature signals. The temperature range of the thermosensors is from -30~ to +90~ with an accuracy at 0~ of :L-0.2~ The average temperature of the radiator was determined by the temperature sensors place close to the inlet and outlet to the radiator loop in the calorimeter and by considering the radiator's heat capacity.
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store was a calorimeter of 11 litres volume. Due to the small system size stagnation effects can be neglected. The tilting angle of the radiating surface was 21 o. As a reference for the max. radiative cooling effect the temperature depression of a 10 x 10 cm black painted aluminium plate was simultaneously measured (not shown in the figure). Measuring data of the sky occupancy and for determining the dew point temperature were provided by the Norwegian Meteorological Institute (DNMI) located ca. 500 m from the experimental site. One of the experimental recordings with clear night sky conditions is shown in figure (6). The data from 31.08.95 show a stagnation temperature of the liquid in the calorimeter (Tw~t~r) which lays ca. 4 K under the ambient temperature (T~r). This corresponds to a cooling power in the range of 40 W/m9-.
5.2 Experiments with larger radiator size Another series of experiments were carried out in the Sun-Lab at the University of Oslo, a small test house built after to
Fig. 5. Experimental set-up with measuring equipmentfor determining the radiative cooling power 5.1 Small-scale studies First small scale experiments were carried out with an 0.6 m 2 polycarbonate twin-wall sheet as radiator with backing insulation of 50 mm mineralwool and without cover sheet [ 13]. The cooling
Norwegian building traditions. Radiative flat-plate coolers of polymer plastics and with backing insulation are mounted on the tilted roof. The cooling loop is connected to a pressureless reservoir which stores the cooled heat carder during daytime. The cooling potential was investigated with an active radiator area
Fig. 6. Experiment from 31.08.95: Shown are temperature recordings for the ambient air temperature, radiator surface, painted aluminium reference plate and water temperature in the calorimeter. The temperature o f the water in the calorimeter (Twater)lays ca. 4 K under the ambient air temperature (T~r). The vertical lines represent time o f sunset and sunrise. The maximum wind speed during this experiment was 2 ms -1.
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Fig. 7. Cooling power o f a polymer twin-wall sheet with inner water circulation as a function of the temperature = (Trad- Tamh)between the mean radiator temperature Tradand the ambient temperature Tam~
of 5,3 m 2 and with a tilting angle of 30 ~ The reservoir for the heat cartier was a calorimeter of 285 litres. The cooling loop was directly connected to the heat store. The relative air humidity and sky occupancy were hourly measured by DNMI. The measurements were carded out during night-time, approx, between 22:00 and 5:00 local time, in May/June '99. Figure (7) illustrates the dependency of the cooling power of a polymer twin-wall sheet with inner water circulation as a function of the temperature difference ? T = (T~-Tamb) between the mean radiator temperature T ~ and the ambient temperature T~b in W m 2
The curve fit represents the maximum cooling power with clear night sky and with a wind speed on the radiator's surface of less then 0,7 ms-1. The experimental data were corrected for the heat contribution of the circulation pump in the cooling loop. The series of scattered points show the cooling power with less favourable atmospheric conditions. Figure (7) shows a cooling power of ~ 40-50 Wm -2 at ? T = (T~-Ta~b) = 0 K temperature difference.
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difference ? T
6. D I S C U S S I O N A N D S U M M A R Y The experiment reported above show that radiative cooling from polymer flat plate coolers can under suitable climatic conditions
be an <<energy free>> alternative to conventional air conditioning systems. The use of thermoplastics as a material for flat plate coolers opens new perspectives when it comes to material production, manufacturing, specific weight, squaremeter price, design of system size and integration possibilities. Combined with a solar heating system utilising the same infrastructure, it represents an added value with limited extra costs 9 A night temperature of ca. 20~ or lower in absence of significant relative humidity is necessary in order to gain the necessary cooling capacity of the store.
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REFERENCES [1] J. Rekstad et al. (1999). Effective solar energy utilization - more dependent on system design than solar collector efficiency, Proceedings ISES Solar World Congress 1999, Jerusalem [2] B.W. Olesen (1997). Possibilities and limitations of radiant floor cooling, ASHRAE transactions, vol. 103, no. 1, 42-48. [3] W. Kast, H. Klan, J. Rosenberg (1994). Leistung von Heiz-und Kfihlflfichen, HLH Bd. 45, Nr. 6-Juni, 278-281. [4] J.-P. Iosardi (1989). Le refraichissement par le plancher, PROMOCLIM, 20, n ~ 5, 9/1989. [5] T.S. Eriksson and C.G. Granqvist (1982). Appl. Opt. 21, 4381. [6] M. Martin (1998). Radiative Cooling, M.I.T. Press, Cambridge, MA. [7] LECA| Light Weight Aggregate, 12.001E (1996) [8] J. Rekstad et al. (1999). Kombinerte soloppvarmingsanlegg for varmtvann og romoppvarming, Report 05-99, SolarNor AS, Erl. Skjalgssons gt. 19A, N-0267 Oslo. [9] J. Rekstad, S.Bjerke, F. Ingebretsen (1981). Rapport om solenergifors~ks veal forsokshus i I.~renskog, Report series University of Oslo, Rep. 81.05. [10] J. Rekstad (1998). New technology for cost-effective and energy saving heating systems, SolarNor AS, Erl. Skjalgssonsgt. 19B, 0267 Oslo. [11] SOLNOR NL6 solar collector, Technical data sheet, rev.3/99, SolarNor AS [12] B. Bjerke (1993). Datalogger for energimAlingcr, Fysisk institutt, Univcrsitetet i Oslo, Rapport 93-26, ISSN-0332-5571. [ 13] H. Storts (1997). Om kj~ling av bygninger ved utnyttelse av infrared st~ing mot nattehimmelen, master thesis, Dept. of Physics, University of Oslo.
ISES Solar World Congress 1999, Volume III
HYBRID SOLAR/GAS COOLING EJECTOR UNIT FOR A HOSPITAL IN MEXICO Jorqe L. Wolpert, Minh V. Nguyen and Saffa B. Riffat Institute of Building Technology School of the Built Environment University of Nottingham University Park, Nottingham, NG7 2RD KEY WORDS:
ejector cycle, air conditioning, solar cooling
Abstract: A solar powered air conditioning system with a 13kW cooling capacity is described in this paper. The system was designed to operate in a hospital in Manzanillo, Mexico, which is currently under construction. The solar cooling unit consists of a commercial solar parabolic concentrator which drives an ejector cycle cooling system. During periods of low solar radiation a gas burner can provide heat to drive the system. Heat is rejected from the system to ground water. Preliminary experimental work has been carried out on a prototype solar powered ejector refrigeration system installed at an office building in Loughborough, UK. A description of this prototype is also given in this paper. Cost-benefit analysis for the hospital unit in comparison to a conventional vapour-compression system shows a payback time of five years. Computer simulations indicate that a COP of up to 0.62 can be achieved by the system under typical operating conditions encountered in Manzanillo.
1.- INTRODUCTION Conventional vapor compression air-conditioning and refrigeration systems consume large quantities of electrical energy and use refrigerants with significant ozone depletion potential or global warming potential. This ejector air conditioning system uses solar thermal energy for operation. In addition the refrigerant used is water and so the system is environmentally sound. There is a growing need for new technologies that are safe for the environment and which use alternative energy sources to the burning of fossil fuels. Ejector systems have been proposed in the past as alternative cooling devices (A1-Khalidy, 97; Da-Wen, S and Eames, I, 95). As ejectors have no moving parts, their construction is simple and their maintenance requirements are low (Petrenko, V e t al, 94). This makes them an economic alternative to conventional vapour compression air-conditioning sytems. 2. DESCRIPTION OF THE SYSTEM
Figure 1 illustrates the cooling system designed for installation at the hospital in Manzanillo. The major components are
the solar collector and gas burner (not shown), boiler, ejector, evaporator, condenser and feed pump. The solar collector is a concentrating parabolic trough type. This type of collector allows high boiler temperatures to be achieved which enhance the efficiency of the ejector system (Sokolov M and Hersgal, D, 1993). Solar concentrators have been used in the past to run solar thermal power plants (Haddock, C and McKee, J S C, 1991). The integration of solar collectors with a natural gas-fired boiler will ensure that the system can operate during periods of low solar irradianee. A heat exchange medium (oil) is used to transfer heat from the collector/gas burner to the boiler. The boiler is a compact brazed plate (CBE) type heat exchanger which transfers heat from the solar heating loop to the refrigerant. Potentially the solar collector could be used to heat the refrigerant directly. This method was not chosen as the integrity of the cooling system could be compromised. The refrigerant boils at a high pressure and temperature in the boiler. The resulting vapour is exhausted through the ejector compressor. A typical ejector cross section is shown in figure 2. Inside the ejector this primary stream is accelerated to a high velocity (typically
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Figure 1 Ejector cycle layout roach no. >2). This jet passes across the inlet from the evaporator and entrains fluid. This causes a low pressure inside the evaporator resulting in the refrigerant boiling at a low temperature. The evaporator is a direct expansion unit in which heat for the boiling process is extracted directly from the air, thereby cooling the air down. The refrigerant vapour flows into the ejector and is compressed. The mixed stream from the boiler and evaporator is exhausted from the ejector to the condenser. The condenser is a CBE type heat exchanger which cools the refrigerant stream using ground water. The refrigerant condenses on cooling. Condensate is passed to the boiler using a liquid feed pump and to the evaporator using a metering valve. A unit of this design, with a 13kW cooling capacity, is currently being fabricated for installation at an operating hospital in Mexico. A unit that is similar in operation to this has been installed and tested at Beacon Energy, Loughborough, UK.
constant pressure constantarea mixing chamber "rni~xingch ~ r
diffuser
primary nozzle
'/ ~i~et
Figure 2 Ejector cross section
3.- SOLAR COOLING SYSTEM FOR UK An ejector solar cooling system was designed for an office building in Loughborough, UK. It uses heat from an evacuated tube heat-pipe solar collector array to drive the system. Concentrating solar collectors offer the best performance at high collection temperatures if high levels of direct solar radiation are available. However, this type of collector will not operate when solar radiation is diffuse.
The solar collector array installed at this office buiding was to contribute to building heating during the winter time when solar radiation is generally diffuse. During the summer months when cooling is required but the solar radiation intensity is low, heat can be provided by a gas burner. The cooling system described in this section was designed to operate with a low pressure steam ejector so that the condenser-boiler pressure differential was small and the feed pump between the condenser and the boiler could be dispensed with. Return of condensate to the boiler was accomplished using a gravity feed. The unit was installed in a single storey building with the condenser and ejector assembly installed in the roof space. The boiler was located in a shallow pit so as to retain the correct vertical spacing between these components for gravity pumping to occur succesfully. Condenser cooling water was sourced from an existing pond and was pumped to the condenser via a strainer and flow meter using a small submersible pump. The system used a direct expansion evaporator which was installed in the existing air handling system. Figure 3
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shows the layout of the cooling system installed in Loughborough, UK.
4. SOLAR E J E C T O R UNIT F O R MANZANILLO HOSPITAL
The design parameters fior this system are given in Table 1. During operation this system runs at a slightly off-design condition also shown in Table 1. This system is able to provide up tp 6.5kW of cooling and operate with a COP of up to 0.3.
Manzanillo is a city located on the Pacific coast of Mexico at a latitude of 19.35 ~ Manzanillo a high solar irradiance throughout the year (1845 kWh / m 2 per annum). Tables 2 and 3 show the geographical data for the hospital site and the solar radiation and temperature chart for Manzanillo, respectively. The hospital is currently under construction and it forms part of a new hospital construction program under taken by the Instituto Mexicano del Seguro Social (IMSS) for 1999-2001. The area within the hospital where the systen will be installed is physiotherapy. This has a total area of 275 m 2. Figures 4 and 5 show the climatic site plan for the hospital and the plan of the area that will be serviced by the cooling unit respectively. The area where the system will be installed corresponds to climatic zone 3. These climatic zones are identified according to their solar radiation gain. Areas with a high solar gain have a high climatic zone designation. Calculations for the cooling load were made with reference to the architectural design of this particular space, including its building materials, so that an accurate assessment could be made.
Operation
Design Boiler temp 90~ Cond. Temp 30~ Evap. Temp 9~ Capacity 7 kW COP 0.32 Table 1. UK system conditions motorised ball valve
design condenser
liquid
85~ 30~ 8~ 6.5 kW 0.30
ejector
condenser
A concentrating solar collector was chosen for this location as solar radiation is generally direct and there was no requirement for building heating. A concentrating solar collector can operate at a much higher temperature than a conventional flat plate collector or an evacuated tube collector. Collector efficiency is dependant on the intensity of the solar radiation. For a given heat requirement, a higher solar radiation intensity will require a smaller absorber plate area. The losses from this smaller absorber will be less than for a large absorber area. By focusing the radiation from a large area onto the absorber plate, the radiation intensity can be enhanced. The solar collector considered for the Manzanillo project is a concentrating non-tracking parabolic trough solar collector which results in a considerably reduced cost compared to the cost of evacuated tubes such as the ones installed at the office building in Loughborough or any tracking concentrating devices.
evaporator solenoid ground
steam
droplet
design boiler liquid
one-way valve
boiler boiler pit
~I
..........................................................................................................................
Figure 3 Cooling system layout for UK
maintains
Use of a concentrating collector allows a high boiler temperature to be used. This results in increased performance. Due to the high condenser-boiler pressure differential a feed pump must be included to return condensate to the boiler. The refrigerant used is water and the heat exchange medium between the solar collector/gas burner and the boiler is oil. Figure 6 shows the cooling system and details the refrigerant conditions around the cycle.
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Altitude Latitude Climatic Zone Situation Time Zone Azimuth angle (solar collector) Inclination angle (solar collector) Solar concentrator facing
50 m 19.35 ~ 503 City -6G 0~ 0~ South
Table 2 Geographical Data for Hospital Site
Global Irradiation (kWh/ha) Temperature
Jan 141
Feb 151
Mar 184
Apr 182
May 173
Jun 163
Jul 159
Aug 153
Sep 132
Oct 141
Nov 130
Dec 134
Year 1845
24.2
23.7
23.1
23.7
25.4
27.6
28.7
28.7
28.1
28.1
26.5
25.4
26.1
(~ Table 3 Solar Radiation and Temperature for Manzanillo
Figure 4 Climatic zone plan for Manzanillo Hospital
Figure 5 Physiotherapy chambers plan for the ManzaniUo Hospital
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ejector live steam droplet CBE evaporator water/glycol secondary refrigerant 13kW @ 10C
CBE boiler oil heated 21.02kW ~ 160C
metering
valve
CBE condenser ground water cooled 34.02kW @ 30C
feed pump 0.48kg/min dP6.18bar 4.9W minimum duty
Figure 6 Cooling system for Manzanillo
The solar collector heats oil to 200~ This oil then passes to the boiler CBE where it is used to raise steam. Steam leaving the boiler passes through a droplet separator to ensure that it is dry and then to the ejector. This steam will typically be dry saturated at a temperature of 160~ and a pressure of 6.2bara. Water vapour from the evaporator is entrained causing the remaining liquid to boil. The evaporator operating condition is typically 10~ 0.012bara. The mixed streams exit the ejector and are cooled by ground water in the condenser. Refrigerant is returned to the evaporator through an expansion valve or to the boiler using a feed pump. Note that the feed pump power requirement is very small. Table 4 outlines the design parameters for this system.
Boiler
160~ 6.2bara 0.48kg/min 21.02kW Condenser 30~ 0.043bara 0.81 kg/min 34.03kW Evaporator 10~ 0.012bara 0.33kg/min 13kW COP 0.62 Table 4. Manzanillo system conditions
5.- DESCRIPTION OF SIMULATIONS Data for the meteorological calculations regarding solar radiation was taken from, the Meteonorm Software, version 1997. Hevacomp was used to estimate the cooling load considering the architectural design and building materials for the hospital in Manzanillo. A computer programme developed at Nottingham University was used to predict ejector performance and overall cooling cycle efficiency. 6.- COST BENEFIT ANALYSIS
The system described is a bespoke product and that factor makes it considerably more expensive than a mass produced unit. The total investment for the solar cooling unit for the IMSS Manzanillo Hospital is approximately s A conventional vapour compression system with a similar cooling capacity would cost approximately s in Mexico. For the cost effectiveness analysis, consideration was made for the costs of packaging, transport and import taxes of the solar cooling unit into Mexico and to Manzanillo. The comparative analysis was made using the net present value method. This method is a comparison between the investment made at present, using the present value of money considering interest rates and inflation over a period of time. In this case two lifetime periods were considered for analysis. The lifetime considered for the ejector unit was of 30 years and that assumed for the vapour compression system was of 15 years. The lifetime of a vapour compression system is relatively short due to high rates o f wear in the reciprocating compressor. The lifetime of the ejector refrigeration unit is long as it contains few active parts. Table 5 shows the economical considerations for the net present value analysis. Time span considered Infaltion interest rate energy cost per kWh
30 y~ars 7% ,ready 11% yearly l l p Stg.
Table 5 Economic considerations Figure 7 shows the results obtain from the net present value analysis.
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Haddok, C. and McKee (1991), Solar energy collection, concentration, and thermal conversion - a review. Journal of Energy Sorces, UK, Volume 13, Pp. 461482. Kroll, E. (1947) The design of jet pumps. Chemical engineering progress, Vol. 1, No. 2, pp. 21-24.
Figure 7 Economic Analysis
8. CONCLUSIONS The solar powered ejector system described is an environmentally safe alternative to conventional vapour compression air conditioning systems. A payback time of 5 years for the ejector cooling system was calculated. After this period, the vapour compression system will continue to consume over s worth of electrical power annually at today's tarifs and consequently continue to contribute to the burming of fossil fuels with its associated environmental effects. Solar powered cooling has the advantage that performance increases with solar radiation intensity, the period when cooling is generally required. Solar radiation intensity at the site chosen for installation of the proposed solar ejector cooling unit is high. Viability of the unit has been demonstrated at an office building in Loughborogh where a COP of 0.3 was achieved. A COP of 0.62 has been predicted for the unit to be installed in Manzanillo. AKNOWLEDGEMENTS The authors wish to acknowledge the IMSS authorities for considering the system for the Regional Manzanillo Hospital. REFERENCES
A1-Khalidy, N. A. Performance of a solar refrigerant ejector refrigerating machine. ASHRAE Transactions: Research 4016 Sun, Da-Wen and Eames, I. Recent developments in the design theories and applications of ejectors- a review. Journal of the Institute of Energy, June 1995, no. 68, pp. 65-79. Sokolov, M and Hershgal, D. (1993) Solar powered compression enhanced ejector air conditioner. Solar Energy, vol. 51, no. 3, pp. 183-194. Petrenko, V., Bulavin, I. V. and Samofatov, Y. A. Investigation of the methods increasing the efficiency of solar ejector cooling and refrigeration systems. (1994), Technical report for the Odessa State Academy of Refrigeration,Odessa, 270100, Ukraine.
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The Freezing Process of Water inside a Vertical Cylinder with a Finned Tube Yim Changsoon and Seo Taebeom Department of Mechanical Engineering, Inha University, 253 Yonghyun-Dong, Nam-Ku, Inchon-city, Korea, 82-32-860-7312, 8232-860-7327, cs~m~dragon.inha.ac.kr, [email protected] Koh Jaeyoon Graduate school, Department of Mechanical Engineering, Inha University, 253 Yonghyun-Dong, Nam-Ku, Inchon-city, Korea, 8232-872-8228, [email protected] Keyword: Phase change material, Thermal energy storage, Latent heat Abstract - In order to increase the effective conductivity of the EC.M during the freezing process, annularly finned tubes are widely used for the ice-on-coil type thermal storage system. In the present study, its heat transfer characteristics are experimentally investigated. The heat transfer rates of the annular finned tube are compared to those of smooth tube in order to demonstrate the increasing effect of the heat transfer rate.
NOMENCLATURE C: Specific heat [kJ/kg. ] E~i: Sensible heat energy from ice E~: Sensible heat energy from water Etot: Total energy stored in the PCM E~,~t: Total supplied energy at tube E-: Latent heat energy [El] H: Height of cylinder M: Mass R: Radius of cylinder R* = RtCRt: Ratio of radii of fin and tube T: Temperature V: Volume [m3] Y: Vertical direction coordinate
[kJ] [kJ] [kJ] [kJ] [m] [kg] [m] [ ] [m]
Greek symbols v.r Density v: Latent heat of fusion 10 Effectiveness of fins
[kg/ms] [kJ/kg] [%]
Subscripts i :Ice w : Water ini : Initial f : Freezing point 9I N T R O D U C T I O N
The thermal resistance of the ice-on-coil type for the ice storage system increases as thickness of the frozen layer on the heat transfer surface grows, because the
thermal conductivity of ice is 2.22 W/ml~ which is relatively low compared to that of water. Since the quantity of ice packed in the system decrease with increasing the thermal resistance of the ice-on-coil type, the techniques to increase the heat transfer rate to the P.C.M (Phase change material) are widely investigated. One of these is that PCM is mixed with the metallic fragments which has high conductivity. The other is to increase the heat transfer area using several types of fins. It is difficult to use the PCM mixed with the metallic fragments for the direct contacting melting system because of the maintenance. Therefore, the finned tubes are suitable for general applications. For the freezing process of P.C.M. in the ice storage system, free convection and conduction become dominant in water and the ice layer, respectively. Bathlet et al. (1981) investigated the effect of fin during the freezing and melting process. Ismail et al. (1999) studied the performance of the PCM cylindrical storage tank using analytic and empirical methods. Kim et al. (1991) studied the effect of the longitudinal type fins with the paraffin wax as the P.C.M. and showed that heat transfer rate increased as the number of fins increased. Kim et al. (1992) used copper net in order to increase the effective thermal conductivity of the PCM. In the present study, a vertical cylinder type annularly finned tube was adopted and the heat transfer characteristics around the tube during the freezing process was investigated. In addition, the shape of solidliquid phase, the temperature distribution in the ice storage system, the storage capacity and the characteristics of heat transfer during the freezing process were studied based on the experimental results.
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2. E X P E R I M E N T
2. Experimental setup The experimental apparatus consists of the test section, the isothermal air supplying system, the brine chiller,the temperature control unit, the circulating pump of working fluid,the flow meter, the temperature measuring device and the data logger. The temperature of the surrounding of the test section is maintained as its initial temperature, while low temperature fluid flows through the annularly finned tube. In order to prevent the heat loss from the test section, the temperature of the surrounding of test section is kept to the same as that of the inner wall of the test section. (Fig. I) The schematic diagram of the test section is shown in Fig.2. It consists of a Pyrex cylinder (inner diameter: 100ram, height: 250ann, thickness: 5ram), a copper tube (thickness: Imm, outer diameter: 20mm), 10 fins (with the outer diameter of 40, 50, 60ram respectively, inside diameter: 20ram and pitch: 20ram) which are attached by soldering. For measurement of inner temperature of the storage tank, 26 thermocouples (T-type) are used. Measured temperatures are recorded on a PC by the data acquisition system ( Y O K O G A W A HR2500E). A camcorder is used to observe the freezing phenomenon.
2.2 Experimental Process The distilledwater, of which initialtemperature is 1 and 7, is enclosed in the verticaltest section. Then, the working fluid flows upward and downward through the tube which is attached to the annular fins. The inlet temperature is -I0 an d the mass flow rate is I0 liter/min.And the ratios of the fin to the tube are 2.0, 2.5 and 3.0. Freezing process is executed for 150 minutes. The table.1 shows the experimental conditions.
3. R E S U L T S and D I S C U S S I O N
3 . . Solidification shape Fig. 3 shows the freezing shapes with time around the annularly finned tube for the several different radius ratios.The initialtemperature for is Fig. 3 is I . Fig. 4 shows the phase interfaces of a smooth tube and the data were taken from the prior studies that wcrc conducted under the same conditions. During the process of the smooth tube, the shape of phase interface shows the differentshapes from those of annuarly finned tubes due to natural convection caused by density inversion. During the fzeezing process of annularly finned tube, convection does not influence the
shape of the phase interface and phase interface grows only through the surface of the fin and the tube. As the cooling process is continued, the shape of the phase becomes less curved and finallyitbecomes flat. The experiments show that the phase interface in the inlet region is faster than that of outlet region. As the ratios of fin radius increase, the ice storage capacity increases. This result shows that the heat transferrate in inlet region is large, so that heat conducted in outlet region becomes less. As the diameter of fin increases, this pattern of the phase change becomes clear. If the working fluid flows upward through the tube, the freezing process of water, the density inversion occurs in the sensible cooling region whose temperature is around 4. Therefore, the temperature at the bottom of the cylinder becomes low and the water spreads out to the entire test section transferringheat to the upper region. If the working fluid flows downward, on the other hand, the cold P.C.M in the upper region does not moves to the lower region so that it causes the temperature stagnation. Fig. 5 shows the comparisons of the phase interfacial shapes for the different fin radius ratios. The initial temperature is I an d the fluid flows downward. Although the quantity of ice increases with the ratio of the fin radius, the average thickness of the solidification lager becomes thin. Especially, the solidification thickness at the middle between the two consecutive fins becomes large as the fin radius increase.
3.2. Temperature distribution on fin surface Fig. 6 shows the average surface temperatures at the fin base and tip for the different of fin radius. When the working fluid flows upward, the mean temperatures at the fin base and fin tip becomes high with increases of the ratio of the fin radius because the heat transfer rate between the fin and the P.C.M is large. The average temperature difference between the base and the tip of the fin are shown in Fig. 7. Because of the differences of temperature are caused by the heat transfer on the fin surfaces,the increasing ratio of fin radii makes the heat transferring large and show the large temperature differences.In the beginning of the freezing process, the temperature differences of the fin base and tip are small, but they increase suddenly. This results show that the heat transfer of the fins is small in the beginning process (thistime, heat transfer on ice tube is dominant). But as time goes the heat resistance increases the heat transferring according to the growth of solidificationlayer and decreases the heat transferringon the fin surface. As the diameter ratio of fin is small, a conversion period of temperature decreasing is fast.
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3.3 Freezing Energy
follows;
Etot (total energy stored in the PCM) can be evaluated as the sum of the sensible heat stored in the PCM (unfrozen), the sensible heat stored in the solidified layer and the latent heat. This gives
(1) The water froze more rapidly at the tube inlet than at the tube outlet, and the freezing rate decreased because the thermal resistance increased with an increase of the ice thickness.
E SW = % C ESi =u E__ = u E
tot
=E
W (Tini - T) u bV
{ C w (Tini - Tf ) + C i (Tf - Ti )} u bY 1
bV
SW
+E
+E Si -u
Fig. 8 shows the variation of stored energy with the change of the fin radius ratio. The data of the smooth tube were taken from the prior study that was conducted under the same conditions. The quantity of the stored cool thermal energy increases when the fin radius ratio is large. The increasing rate of it decreases slowly as the fin radius ratio increases. Fig. 9 shows that the average heat flux on the heat transfer surface of the finned tube decreases as the fin radius ratio increases. Also, the average heat flux of the finned tube is much smaller than that of the smooth tube.
(2) During the freezing processes of the EC.M., free convection in the water and conduction in solidified layer occurred respectively. The total stored cool thermal energy for the upward flow of the brine was higher than that for the downward flow. (3) The heat transfer rate from the fin was small at the beginning of the process, but it increased rapidly as the ice layer on the tube was developed. And then, the heat transfer rate decreased slowly as the ice layers on the fin surface grew. Therefore, it was found that the maximum point of the heat transfer rate existed. If the ratio of R* was small, the maximum point appeared sooner. (4) The heat transfer rate in the annularly finned tube increased as the ratio of R* increased, but the increasing rate gradually decreased and the fin efficiency became too small. Therefore, the fin size was very important for the system design. The result of the study can be applied to the optimal design for the ice storage system.
References 3.4. Efficiency The efficiency of energy storage in the ice storage tank defined as follows;
u
Erm E input
Bathlet, A. G, Viskanta, R. (1981) Heat Transfer and Interface motion During melting and Solidification around a Finned Heat Source/Sink, Transactions of ASME, Vol. 103, November, pp. 720-726. Dorgan, C.E., Elleson, J.S. (1994) Design guide for cool thermal storage, ASHRAE. Huh, K. (1995) A study of Heat Transfer during freezing process of water in a vertical cylinder, INHA University.
It is obvious that the quantity of the stored energy increases as the fin radius. It is shown in Fig. 10 that the quantity of the energy storage efficiency decreases rapidly as time goes. Especially, the efficiency for the large fin radius ratio is relatively low, and the decreasing rate increases with time. Therefore, the configuration of the finned tube for the energy storage system should be carefully designed with considering both the capacity required and the efficiency of the system.
Kim, T., Kirn, K. (1992) Experiment on latent heat storage system using heat siphon, KSES, Vol.12, No.3, pp. 28-46.
4. Conclusions
Kim, S., Ro, S. (1986) Effective conduction rate of EC.M mixed with fragment of metal, KSME, Vol.10, No.6.
The freezing process of the annularly finned tube of which the radius ratios of the fin to the tube is 2.0, 2.5 and 3.0 in the ice storage tank is investigated in this study. The results of the study are summarized as
Ismail, K.A.R., Goncalves, M.M (1999) Thermal performance of a PCM storage unit, Energy Conversion & Management 40, pp.115-138. Kim, K., Yoo J., Kim, K. (1991) A study of energy storage system using E C.M, KSES, Vol.11, No. 1, pp.69-77.
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Fig. 1 Schematic Diagram of experimental apparatus
Fig. 2 The test section
Fig. 3 Variation of the solid-liquid interface with time( Tmi=l
) (.. 9downward, v~. upward )
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Fig. 4 Variation of the solid-liquid interface with time (smooth tube)
Fig. 5 Variation of the solid-liquid interface with time (Ti~i=l , downward)
Fig. 6 Average temperatures at the fin surfaces. (Upward, T=i=7 )
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R*=3.0
R*=2.5
y/H=0.05 4
,-~
9
y/H=0.95
0
o
R*=2.0
9 y/H=0.05
A
y/H=0.05
y/H=0.95
o
y/H=0.95
r
0
r~
0 1.4
A 9
9
All, 0
v
9
&
2
9
0 [-
9
0
.
9
,
30
9
.
60
9
,
90
9
,
120
150
Time [Min.]
Fig. 7 Temperature differences between the tip and the base of the tube. (upward, T=i=7 )
Downward
250
-
[]
R*=3.0
A
R*=2.5
O
R*=2.0
Upward []
R*=3.0
9 R*=2.5 9 R*=2.0 II
9 smooth fin
[]
200
150
~)
B & e
100
5O
310
"
6"0
9n0
"
1:20
"
150
ri=r [M~.]
Fig. 8 Stored cool thermal energy. (T=i=l
)
Downward 12000 -
Upward
[]
R*=3.0
[]
R*=3.0
[]
R*=2.5
O
R*=2.5
9 R*=2.0
A
9000
r ,.~ v-q r O
R*=2.0 Smooth fin
& 0 []
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9
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Fig. 9 Total heat flux. (T~=I
I
150
)
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250
3.0
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200
461
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i
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Table 1 Experimental conditions. Temp. of working fluid
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ratio (R*)
Flow direction of working fluid
Initial temp. of water
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7
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2.0
-10
2.5
3.0
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THE "ClCLOPS" SYSTEM: OPTIMISED MANAGEMENT OF MIDDLE-SIZED HYBRID WIND-PV-DIESEL PLANTS
Ermen Llobet, Joan Sol/z, Joan Pitarch, Josep Prats ECOT~CNIA sccl, c/Amistat 23, Barcelona, Catalunya, 08005, Spain Tel. 34 93 225 7600, Fax. 34 93 221 0939, E-mail: [email protected]
A b s t r a c t - The "Ciclops" System is electronic equipment that has been developed in order to carry out the energy conditioning from energy sources to batteries and from these to the consumption in middle-sized hybrid stand-alone wind-PV-diesel plants. The paper describes the developed system and mentions its major applications. According to its global concept and the set of technical solutions developed, the "Ciclops" is also described in terms of a technology itself. The main technical features and the results of the equipment testing, as well as some specific technical aspects are presented and discussed. A short economic approach is also included in the presentation. It is concluded that the Ciclops equipment is an advanced product that can give a reliable electric supply to a wide range of applications in remote areas, benefiting from the "hybrid" approach. The technology has a high level of performance and can be developed, in the future, in other power ranges, with other power sources, applications and complementary performances.
1. INTRODUCTION The "Ciclops" System is electronic equipment that has been developed in order to carry out the energy conditioning fi'om energy sources to batteries and from these to the consumption in middle-sized hybrid stand-alone wind-PV-diesel plants. It aims to manage the system, at the same time, in an optimised way. The Ciclops consists of a set of source adapters (wind, PV, diesel AC/DC converters), a DC/AC converter and a management unit. The units are linked at the level of a 120 V DC bus and have a power capacity in the range of 10 kW. The first phase of development has been supported by the Joule Programme of the European Commission and has allowed commissioning two prototype plants. With the acquired experience a redesign of the system has been carried out in order to improve some features and make it able to be commercialised as a standard and modular product. The present paper aims to describe the system, its applications and the main characteristics and features of the developed technology, as well as some testing results. Some conclusions of an economic analysis are presented, too. 2. LAYOUT OF THE CICLOPS SYSTEM
Fig. 1 shows a general layout of a stand-alone hybrid plant based on the Ciclops equipment. To each source generator is coupled a charger which transforms the input voltage - in DC or AC, depending on the case - into DC battery voltage. The chargers can operate on the maximum power point of the generators or at a given current preset value. The DC/AC output inverter supplies 10 kW of power at three-phase sine-wave voltage, 220/380 V. All of the units are linked by a communications network via a serial port. A plant management unit (which is not shown in the Figure), may be implemented as a physically separated module or included in the same hardware of one of the other devices; it acts as a 'master' and commands the global management of the system: start/stop of the Diesel set and of the inverter, operation at preset current of the generators and, optionally, management of extra loads. A typical installation with the CICLOPS consists of: 9 An ECOTI~CNIA 7/10 wind turbine of 10 kW. 9 A photovoltaic solar array (2, 5 or 10 kW). 9 A 15 kW diesel generator. 9 A 120-volt, 70 kWh stationary battery. 9 The CICLOPS (wind, solar, d i e s e l - up to 10 kW chargers, 10 kW inverter and system manager). 3. APPLICATIONS All of the professional applications for stand-alone electric supply are possible, inside the mentioned power ranges.
Fig. 1. System layout.
Small centres of population. Agrarian exploitations. Mountain shelters. Rural tourism centres. Campsites. Small workshops or factories. Nature classrooms. Fish hatcheries. Telecommunication installations. Environmental surveillance installations.
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The design and the size of the plants will depend, firstly, on the consumption needs of each site, secondly on the local energy conditions and thirdly on a foreseen energy balance (percentage of diesel contribution). The analysis of the complementarity of the renewable sources will determine the most suitable combination in each case. Typical consumption levels for which the system is suitable are between 20 and 100 kWh/day. The Ciclops system aims to be commercially available after a demonstration period that will take place in several plants in the next months.
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sources and other cases beyond the classic stand-alone application. 4.2. Power switching The power conditioning is managed by means of PWM driven power transistors. The power configuration of the chargers is based on a step-down switching topology; a passive rectifying bridge is coupled for the AC sources. The inverter consists of a three phase inverting bridge and a low frequency transformer. Fig. 3 and 4 show the power layout of the step down source adapters and the inverter, respectively.
4. Ti:iE CICLOPS TECHNOLOGY
4.1. The global concept The Ciclops system has been developed in a first generation and implemented in two prototype plants. Even working rather correctly, a serious redesign of the equipment has been undertaken in order to improve some features and allowing the industrialisation of the product.
Fig. 3. Power layout of the step-down configuration used for the source adapters.
Fig. 4. Power layout of the inverter.
Fig. 2. Prototype pl~tinTenerifel ............................................ As a global system, the Ciclops aims to be a modular, integral, full automatic and highly reliable supplying solution. As modular it combines the principles of standardisation and adaptability to the different applications. As integral the equipment has been developed according to the global point of view of the hybrid plant. The full automatisation allows an operation non-depending from the user's criterion; this, together with the combined availability of the sources and the reliability of the technology lead to a highly reliable supply. At the same time the Ciclops aims to be a global technology, in the sense that it gives a set of technical solutions, in different aspects, that can be applied to other power ranges, other power
Both Mosfet and IGBT' technologies have been used, mainly depending on the working voltage conditions. For voltages up to 200 V the Mosfet technology allows reducing the switching losses, while the IGBT' is more suitable for higher voltages. Thus, Mosfet transistors have been used for the inverter and IGBT are being used for the step down chargers, which input DC voltage range is typically 200 to 500 V. The inverter, finally, uses a 50 Hz low losses transformer. 4.3. Control hardware and software Each unit is controlled by means of a micro-controller that manages the PWM drivers, the measurements, the calculations, the output commands, the communications and the functional control. The micro-controller selected for the control purposes is the Hitachi SH7000, a powerful, 32 bit component with a RISC processor and peripherals integrated. The switching frequency is of 15 kHz, a compromise between control accuracy and reduction of switching losses (and eventually noise).
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The functional control of the renewable source chargers aims to track the Maximal Power Point under free charge conditions. The diesel charger operates under current command conditions (which is usually the corresponding to the charger nominal power). By battery full charge conditions the source adapters operate at limited charge current conditions, usually at fixed charge voltage conditions. The control of the inverter aims to keep a high quality sine wave at the transformer output side. Among others, the inverter is able to supply reactive power without limits and to respond to non-linear loads avoiding the transformer saturation. See results from testing, below.
Source adapters 9 efficiency at nominal conditions 94% and 90% at load factor of 10%; 9 range of ambient operational temperatures: 0 to 40 ~ 9 measurement and control accuracy: 2% of full-scale values.
4.4. Interfacing and internal communications Each equipment incorporates a full user's interfacing, based on keyboard, display, signalling LED's and switching. A complete set of information about instantaneous, accumulated and status variables is available by means of a deployable menu; some parameters may also be programmed. All the micro-controllers are linked to the main control unit by means of a serial link supported by optical fibre. The configuration is of the master-slave type. 4.5. System management A master unit is able to manage the control of the whole hybrid plant. Its main goal is to control the state-of-charge of the battery, which is the energy buffer of the plant. In view of this, it controls the backup energy source, that is the start/stop of the diesel generator. In the event of energy surpluses it orders the source adapters charge current limitation and eventually can manage the connection of extra-loads (pumping, water heating, etc.). In case of emergency it can order the inverter switchingoff for avoiding the battery full discharge. Periodically it manages a battery equalisation charge. The master unit will manage, when available, the external communication for two main purposes: monitoring of the plant and operation surveillance. This last point is capital in order to improve the availability of the plant and make easier and cheaper the maintenance tasks. Each unit has a control soWvvare, that can, per default, cover the main tasks of the system control, in the event that the master unit is unavailable.
Fig. 5. Wind-diesel prototype plant in Catalonia. Inverter
9 supply: three phase 220/380 V, 50 Hz 9 efficiency between 92 and 95% for load factors beyond 10% 9 self-consumption at no-load conditions: <1% of nominal power (83 W) 9 wave form: sine wave with maximum THD of 5% 9 maximum phase imbalance: 100% 9 input voltage range: 100 to 150 V. 9 range of ambient operational temperatures: 0 to 40 *C 9 measurement and control accuracy: 2% of full-scale values 9 minimum cos phi: 0 9 overload capacity: 170% during 5 minutes at 25"C 9 output voltage stability: +/- 2% 9 output frequency stability: +/- 0,5 Hz 9 full capacity for supplying non-linear loads
4.6. The whole hybrid plant The specificity of each application- due to the varying energy conditions and load p a t t e r n - lead to the need for the modularity. The PV charger can be of 2, 5 to 10 kW. More than one unit is acceptable for each kind of generator, as well as for the inverter. The sizing of each plant, according to the local resources, the consumption pattern and a specific energy balance criterion is to be made. A sizing software tool has developed, which allows evaluating the annual energy balance by means of data time series related to the resources and the consumption. The Fig. 5 shows a wind-diesel prototype plant installed in Catalonia. 5. TECHNICAL FEATURES AND TESTING RESULTS The main technical features of the source adapters and the inverter of the last generation are the following:
Fig. 6. Voltage and current at nominal power for the Cielops 10 kW inverter.
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Figure 6 shows the voltage and current curves for a resistive load at nominal power (in one phase). It can be seen the quality of the resulting sine waves. Figure 7 and 8 show the results of applying non-linear loads; it can be seen, respectively, the current and voltage waves at the output and input of the low frequency transformer. The first one responses to the load demand and the second one is formed by the control algorithm in order to fit the demand without saturating the transformer. (This means, in fact that the output current has a DC component which is supplied by the inverter but is not present in the primary of the transformer). Even in this extreme case the voltage wave maintains a good shape.
Fig. 9. Output current and voltage waves of the Ciclops inverter with an overload of 170%.
Finally, Fig. 10 shows the efficiency curve of the inverter. It can be seen that the efficiency arise up to 95% for a 3,4 kW load, but is kept in a range between 92 and 95% for a power factor range between 15% and 100%. The selfconsumption of the inverter has been found to be of 83 W, which is lower than 1% of the nominal power.
Fig. 7. Current and voltage at the inverter output side generated by a non-linear load.
Fig. 10. Efficiency curve of the Ciclops inverter.
Fig. 8. Current and voltage waves stated in the transformer primary side, corresponding to the previous non-linear load.
6. LESSONS REDESIGN
LEARNT
FROM
PROTOTYPES
AND
6.1. Conclusions taken from the first prototypes Figure 9 shows the voltage and current waves in the event of an overload of 170%. It can be seen that the quality of the sine is fully kept, even in those extreme conditions.
Figure 11 shows an energy balance of a wind-diesel prototype plant. Under a minimum of good wind conditions the diesel is rarely used, in this case. The single line in this figure shows the battery voltage at different moments, always in no load conditions, which represents, in a way, the relative state-ofcharge of the battery.
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power supply source, as well as a good design of a low losses transformer have led to very good efficiency results. On the other hand a high accuracy on designing the PC boards - by using multilayers boards and minimising the current loops -, the cabling and the connections have led the inverter to a very low level of electromagnetic emissions. 6.3. Other remarkable aspects to be discussed The conclusions taken from the previous prototypes and the redesign process have led to other technical aspects worthy to taken into account: It has already mentioned the discussion about the use of Mosfet or IGBT technologies; nowadays the second one has more capacity for managing higher power levels, but with the available devices, able to work at relatively high voltages, the switching losses are in some cases too high; at the same time there is another interesting point about the use of modular or discrete components; the first is easier and cheaper to be mounted, but with discrete components there is, at the moment, a greater range of available devices among which the best one can be chosen. In any case this is a market with a very fast evolution and the last word cannot be said. The work voltages are another "hot point". Higher the voltages are, lower the currents are and lower the switching losses are, too; this will also lead to decrease the size and the cost of the filters; but at the same time the higher voltages cause additional problems with the most commonly available capacitors, that are not able to support too high voltages; this lead to using capacitors in parallel which means more cost, more space and less reliability; finally increasing the voltage level of the batteries can lead to increase the chock danger for the people in contact with them. At the level of applications of the Ciclops system it is difficult to guaranty that only the technical staff will be near the batteries. -
Fig. 11. Day-to-day energy balance in the prototype plant of Mas d'en Mestre (Catalonia), measured at the DC bus.
The main conclusions of these first prototypes have been: inverter: high level of reliability; selfconsumption too high (2,'/% of full power) and efficiency below what was expected (80 to 90%). This was related to the power configuration, as it will be discussed below. There was some cut-off probably due to electromagnetic interference Chargers: good efficiency; some problems with high voltages, especially for the wind charger, causing some capacitors breaking down. System management: to order a full charge by the diesel when the battery is empty leads the diesel to operate for a long time and the user tends to stop it; this alters the automatic management strategy The communications among the equipment by means of a standard serial 485 port were highly affected by switching emissions. Those problems were eliminated with the optical fibre. The physical design of the boards, the cabling and the topology of the whole equipment had a high influence in the emission levels with several negative consequences in the unit operation. The diesel generator is usually already existing at place and it can be a source of problems if it doesn't run properly. The user most times thinks that the problems come from the Ciclops The sizing of the plant suitable to the real needs is fundamental. A high increase in the consumption pattern due to a change of the activity, in one of the plants, has led the plant to be rather useless, after some time. The equipment has to be located in suitably cabinets that can be more user' friendly, easier to transport and install and more compact. -
-
-
-
-
6.2. Redesign of the inverter The previous configuration of the inverter consisted of two stages: a DC/DC converter stage and an inverter stage; between them there was a high frequency transformer. The configuration was rather innovative and tried to avoid the losses in the low frequency transformer; in fact the losses were too high due to the high switching losses in the double switching power train; the main conclusion has been that this configuration was suited for higher power and voltage ranges. The use of Mosfet transistors, instead of IGBT's, a better calculation of the inductances and a better design of the internal
7. ECONOMIC APPROACH The profitability of the stand-alone systems has to be analysed in terms of the comparative cost of such solution with the one of the extension of the grid. It has to be pointed out that the Ciclops electric supply solution cannot be compared with the supply by means of just a diesel generator, due to the high difference in the level of service given. The unit cost of the generated energy is highly related to the local level of resources. The cheapest generation is the one with wind-diesel plants with average wind speeds beyond 6 m/s, for instance. Figure 12 shows a comparison between the energy unit cost for several cases of stand-alone hybrid plants and a cost considered standard (at least in Spain) for the grid extension, depending on the distance to the connection point.
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8. CONCLUSIONS
b
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Fig. 12. Comparative cost analysis between several typical cases of Ciclops stand-alone systems and the conventional grid extension. Some of the economic conclusions that can be derived from the mentioned figure are the following: the cost per kilowatt-hour of wind power is lower than that of both solar and diesel for annual average wind speeds of 4.5 m/s or higher. For consumptions of up to 20 kWh/day, the CICLOPS hybrid wind-solar-diesel system offers a cost per kilowatthour lower than that of the extension of the grid for distances greater than 2 km; If the consumption reaches 50 kWh/day, the distance to the grid at which the stand-alone system is competitive, is of between 4 and 6 km. In a well windy site, with an average annual wind speed of 6m/s (22 km/h), the supply of 100 kWh daily is economical if the distance to the grid is greater than only 2 Km. -
-
The following conclusions may be pointed out: The Ciclops system is an advanced product that it is already commercially available It is able to solve a wide range of applications for the professional electrification of remote sites in a middlesized power range. Good performances of the equipment and accurate solutions for the design and implementation of the whole hybrid plant support the previous afftrmation. The global concept of the developed product (hybrid point of view, modularity, integration, automatisation and supply reliability) together with the wide set of technical solutions given in the different aspects (power train, control hardware and software, user interfacing, communications, optimised system management and procedures for a suitable design of the plants) allows talking about the Ciclops as a particular technology The Ciclops technology may allow future developments in order to enlarge the number of complementary performances and the system power range, to couple other power sources or even to extend the technology to other applications (local or weak grids, hybrid grid-connection, etc.). The last generation of the Ciclops, and particularly the inverter has a high level of technical features The development done has led to several technical conclusions about the best way to go over some typical problems; however some subjects as the kind of switching devices to be used are in a permanent evolution due to the fast development of the components market. Although the unit cost of the generated energy is widely depending on the local energy resources, the consumption pattern and the plant sizing, it can be said that in most of the typical "Ciclops" applications the system is concurrent with the extension of the conventional grid when the distance to this is in a range between 2 and 6 km. -
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SOLAR DISTRICT HEATING WITH A COMBINED PIT AND DUCT STORAGE IN THE UNDERGROUND Manfred Reuss and Jens P. Mueller Bayerische Landesanstalt fuer Landtechnik, Technical University Munich, Voettinger Str. 36, D-85354 Freising, Germany, Fax ++49-8161-714048, email: [email protected] - This paper describes a new concept of a seasonal underground thermal energy storage. This combination of pit and duct storage combines the technical and economic advantages of both systems. Constructive details of a not insulated water pit made of concrete located in the center of the duct store were analyzed. TRNSYS modules describing these storage types were modified for simulation of a solar district heating for a new development of 30 low energy houses. This installation of 800 m" solar collector, hybrid storage, heat pump and district heating system was designed to save at least 50% of primary energy for space heating and domestic hot water. A detailed cost analysis was carried out for selection of the most promising variant.
Abstract
1. INTRODUCTION Seasonal thermal energy storage in a temperature range up to 90~ is a basic requirement for implementation of solar district heating in Mid Europe and for more rational use of conventional energy sources. For this long-term storage of high amounts of thermal energy underground heat storage is favorable from the technical and economical point of view. These storage techniques are classified in three categories/1/: 9 storage medium water (convective): rock cavern storage, pit or underground tank storage 9 storage medium soil (conductive): duct or borehole storage 9 storage medium ground (mixed, convective, conductive): aquifer storage, gravel water pit. Convective storages, especially pits or tanks are typically used for shorter periods. They have the big advantage that the storage medium can also be used as heat transport fluid. Thus high heat transfer power at low temperature differences is gained which is of importance e.g. in combination with solar energy with extreme variations in charging power. Additionally water as storage medium has a high heat capacity, it is cheap, easy available and environmentally beneficial. The major disadvantage of this storage type are the relatively high construction costs. Conductive storages (ducts or boreholes) in soil or rock have an additional heat exchange process from the fluid to the underground. This makes the system less flexible. The big advantage of this storage type are much lower construction costs compared to water tanks or pits. The new storage under consideration is a combination of both types. It consists of a central cylindrical water tank surrounded by several rings of ducts. It will be called hybrid storage within this paper. This combination of the two systems promises to combine the operational advantages of water tanks or pits with the economical ones of the duct storage. Due to the different characteristics of both storage systems the combination can use either a conventional oil or gas boiler but also a heat pump as supplementary heating system. This storage type will be investigated as an important component in a solar district heating system. The major objectives are to optimize the construction with respect to function and costs as
well as the development of a simulation model which is implemented in TRNSYS /2/ to carry out system simulation. Such a simulation model is an important tool for parameter studies and system optimization. It was used to carry out a feasibility study about a solar district heating system with seasonal underground storage for a new developing area with 30 houses in a village near Freising about 40 km north of Munich. In the design phase modeling will assist planning and later on in the monitoring phase it will be used for evaluation. 2. COMPONENTS OF SOLAR DISTRICT HEATING The general layout of the solar district heating consists of the 9 solar collector array 9 long-term storage 9 piping of the district heating system 9 oil/gas boiler or the heat pump for supplementary heating and consumers
Depending on the local situation the solar collectors can either be installed as a central army or as decentralized smaller systems on the different buildings. Central systems using solar roof collectors show typically a better economy. The solar collector delivers the energy directly to the district heating or to the storage. The selection of storage type depends on the geological and hydrogeological conditions of the site as well as on the size of storage capacity required. For technical and economic reasons borehole storage systems need always a buffer storage for matching big variations in power delivered by the solar collectors. Otherwise the ducts have to be designed for peak power which is not cost effective. Here only the combined pit and borehole or hybrid storage is considered. In many locations in southern Germany aquifer storage systems are not applicable due to the hydrogeological situation, legal restrictions or the required size of system. Borehole or pit storages can be an option. The layout of the district heating itself depends on the system concept and can consist of 2, 3 or 4 pipes. The major influencing factors for selection are situation of site, location of the solar system with respect to the storage, the heating center and the buildings as well as the costs.
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Typically only a significant part of the heat demand (base load) for domestic hot water and space heating is covered by solar (e.g. 50%) supplementary heating is required. Conventional fuels like oil or gas are used to cover the rest (peak load). In such cases the operational temperature range of the storage is restricted. The minimum is determined by return temperature of the district heating net and the maximum by the solar system. Heat pumps allow a higher operational temperature range of the storage which results in smaller size and less costs. Additionally the district heating system can be designed for lower temperatures with less heat losses. In this case the heating system in each house has also to be appropriate for low temperatures like floor or wall heating. For smaller housing areas heat pumps systems are more favorable for technical and economic reasons than conventional supplementary heating.
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the liner to tighten the pit and the insulation material to reduce thermal losses. Duct storage systems have construction costs of 90.-- - 160.-- DM/m3-water equivalent (48.-- - 85.-- US$/m3) with some potential for digression. Buffer storages required for power matching in solar district heating plants with borehole storage have a size of 50 - 200 m 3 and cost at least 900.-- DM/m 3 (475.-US$). The hybrid storage should combine the technical advantages of both systems with much lower costs compared to pits. A general layout of this store is shown in fig. 2.
Fig. 1 shows as an example the layout of a solar district heating plant with solar collectors on different buildings, an underground thermal energy storage (duct store), the district heating net and decentralized heat pumps in each house.
Fig. 2: Layout of the hybrid storage In the center of the hybrid storage the pit built out of concrete is located surrounded by several rings of ducts. They heat up the underground and thus the thermal losses of the water pit which has only a thermal insulation layer on the top are reduced. The duct storage will have a significant radial temperature gradient. It is also an active thermal insulation for the pit. The thermal losses of the duct storage can partly be covered by solar energy which could not be delivered to the pit because of too high temperatures in the water. The operation period of the collector array is enlarged and thermal losses are covered by otherwise not usable solar energy. Furthermore, in installations using a heat pump for discharging the storage, part of the losses can be recovered. Fig. l:Layout of a solar district heating system The objective of the research work is the development of this new storage type and its integration in a solar district heating system which allows to cover at least 50 % of the demand for space heating and domestic hot water by solar, the rest is taken from the backup system.
3. HYBRID STORAGE 3.1 General Considerations In several projects a variety of underground storages of different types for different applications were investigated. Pit stores or water tanks have the big advantage of high heat capacity and good heat transfer properties. An analysis of several projects/3/shows a significant digression in construction costs with size of the storage. A 600 m 3 pit in Rottweil costs about 700.-- DM/m 3 (370.-US$/m3), the 4500 m 3 storage in Hamburg amounts to about 400.DM/m 3 (210.-- US$/m 3) and the largest one with 12500 m 3 in Friedrichshafen costs 250.-- DM/m 3 (132.-- US$/m3). A detailed analysis of the cost structure shows a significant contribution of
3.2 Constructive Considerations In a first step of investigation typical operational conditions of the pit storage were analyzed with respect to temperature induced stress to the construction and the materials, especially the concrete. As the pit is located in a warm environment the temperature gradient in the concrete wall is rather moderate which results in less mechanical stress for the construction material. For a typical system layout with 2500 m 2 solar collector array, 5000 m 3 storage (cylinder 11.5 m radius, 12 m high) for 110 buildings, monthly, daily and hourly water temperatures in pit were calculated with TRNSYS using the multi-flow stratified storage model XST.
A separate finite differences model was developed and used to analyze the temperature profile in the concrete walls. It was assumed that the storage has an infinite thermal insulation layer 20 cm thick on top and no insulation to the side and the bottom. The surrounding underground is assumed to be similar to unsaturated sand with a thermal conductivity of 1.5 W/mK and a volumetric heat capacity of 1500 MJ/m 3. The discretization of the cross section under consideration is shown in fig. 3.
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Fig. 3: Discretization of the storage and the surrounding area
Fig. 4: Temperature distribution in the upper comer (concrete wall and ceiling covered with a thermal insulation layer) of the pit storage with ATmax = 14.2 K. The temperature distribution in- and outside the pit is shown in fig. 4. It has to be recognized that the distances are not equidistant but have the size shown in fig. 3. Due to the border conditions the highest temperature gradients occur in the upper comer when the top water layers reach temperatures of 90~ Under these conditions the temperature gradients in the upper comer found are about 14 K, in the middle of the wall they are much smaller, 6 K only, and in the lower comer about 9.5 IC This result shows that from the point of view of mechanical stress
in the concrete it is possible to build such a storage without thermal insulation on the side and the bottom walls. The thermal induced stress on the material will be rather low and thus no special construction and type of concrete is required. Furthermore no additional ~acks in the concrete resulting in a significant loss of water are expected. For a 5000 m 3tank (cylinder 11.5 m radius, 12 m high) the water losses estimated will be in the order of 50 - 100 mVa. If drinking water is used as storage medium for environmental concerns these losses are no severe problem. The costs for refill over the life expectancy are several orders of magnitude smaller than the costs for a liner. This theoretical investigation shows that from the construction point of view this hybrid storage concept is feasible. In a next step the thermal behavior of such a store has to be analyzed.
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3.3 ThermalPerformance Modeling For modeling of a solar district heating plant TRNSYS is an appropriate tool. The program package includes modules describing a duct or borehole storage (DST) and a water pit (XST) but both modules are thermally not coupled. For more exact modeling of the thermal behavior of the combined store it was necessary to modify the TRNSYS module DST in the revision of D. PAHUD /4/. Instead of the XST model the Multiport Store Model (MPS) /5/turned out to be more appropriate for this purpose because the coupling of models was easier. Since the TRNSYS version of DST considers cylindrical storage shapes only, within a first step the consideration of annular shapes had to be realized. For this the global mesh generation had to be adapted. In the modified version the mesh generation is similar to that in the original one, however the outer radius and the bottom of the pit storage are arranged as additional mesh boundaries. In order to obtain more accurate results, the mesh approaching these boundaries is arranged with decreasing grid spacing (fig.5). The existing specifications and features of the original version are not affected.
Fig. 5: Mesh arrangement of modified DST model in cross section of the hybrid storage The heat flows between both storage parts due to a temperature difference are handled as boundary condition in the modified DST version. The DST model receives the heat flow data from the pit storage model which are assigned to the corresponding boundary cells (fig.6). In a first attempt, these heat flows are spatially weighted for the top-, mantle- and bottom area of the pit storage, so totally only three values are received. As mentioned above, in DST these mean values are splitted again and allocated to the corresponding boundary cells. After calculation of the new global temperature, the actual boundary temperatures between duct- and pit storage, defined in-between the boundary cells of each storage part are calculated. The obtained boundary temperatures are spatially weighted again for each area and transferred to the pit storage model, where the new heat flows due to these boundary temperatures are calculated.
Fig. 6: Transfer of heat flows and calculation of boundary temperatures The pit storage is arranged with its appropriate size, dimensions and thermal properties. For the calculation of the heat flows through the top and the bottom of the pit storage, two heat exchangers are to be arranged, one each in the top- and bottom layer of water. The heat transfer capacity rate of each heat exchanger corresponds to the heat transmissivity of each storage wall. With the corresponding boundary temperature as heat exchanger inlet fluid temperature, the internal energy change of each heat exchanger is obtained as corresponding heat flow between the two storage parts, provided the mass flow rate through the heat exchangers is sufficiently high. Since the mantle area is in contact with different layers of water inside the pit, the calculation of the heat flow through this area of the storage becomes more complicated. However, this earl be avoided easily by arranging the mean boundary temperature at the outer radius of the pit as ambient temperature. The heat flow through the side of the pit then corresponds to the heat loss rate to the ambient, which is calculated in MPS as well. For this, the heat loss capacity rates of the top and the bottom of the storage have to be set equal zero. This simulation tool has been tested for some special cases with very good agreement. So it can be assumed that it describes the storage with acceptable accuracy. A validation against monitoring data is planed at the first plant realized. 4. SYSTEM ANALYSIS OF A SOLAR DISTRICT HEATING In parallel to the model and component development a whole system was studied. A community near Freising asked for a feasibility study/6/ofa solar district heating for a new developing area with 20 single-family and 5 twin-houses which will be built within the next three years. A scheme of this development is shown in fig. 7. The major objective was to design a system for this settlement which saves at least 50% of primary energy compared to a conventional oil based heating system. The demand for the case study is specified by of low energy building standard. Each house was assumed with 230 m 2 (singlefamily) respectively 320 m 2 (twin-house) living space. The actual valid German building standard WSchV'95 (Warmeschutzverordnung '95) lays down an annual heating energy consumption for such single-family houses of 75 kWh/am 2 and for the twin-houses 70 kWh/am 2. The heat demand for domestic hot water was as-
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sumed to be 845 kWNa-person (40 l/d of 60 ~ which is about 102 MWNa for the whole settlement. In order to reduce the energy demand the houses have to be designed with an insulation standard approx. 25% better than laid down in WSchV'95. The total heat demand of the whole settlement is for space heating 385 MWNa, for domestic hot water 102 MWNa and thus in total 487 MWh/a.
6.
7.
central solar system, seasonal hybrid storage, supplementary heating with a central heat pump, district heating system with 2 pipes central solar system, seasonal hybrid storage, supplementary heating with remote heat pumps in each house, low temperature district heating system with 2 pipes
Fig. 8: Scheme of the central solar system with seasonal duct storage and supplementary heating with a central oil boiler
Fig. 7: Scheme of development For simulation test reference year weather data of the region (TRY9) are used. The geology at the site was investigated by a test drilling down to 50 m depth. The geological profile shows a series of clay layers with thin fine sand layers down to about 30m, sand and fine gravel till 37 m and the clay to 50 m. No ground water was found in this test drilling. Thus the location is favorable for and underground storage - duct, pit or hybrid store. In the feasibility study several system variants were analyzed from the energetic and economic point of view. The major ones are: 1. conventional heating of the buildings with remote oil boilers in each building as reference case 2. central solar system, seasonal duct storage, supplementary heating with a central oil boiler, district heating system with 2 pipes (fig. 8) 3. central solar system, seasonal duct storage, supplementary heating with a central heat pump, district heating system with 2 pipes (fig. 9) 4. central solar system, seasonal duct storage, supplementary heating with remote heat pumps in each house, low temperature district heating system with 2 pipes 5. central solar system, seasonal hybrid storage, supplementary heating with a central oil boiler, district heating system with 2 pipes (fig. 10)
Fig. 9: Scheme of the central solar system with seasonal duct storage and supplementary heating with a central heat pump Each variant was also simulated with remote solar systems on each house which requires a district heating system with 3 pipes. Such a district heating is about 30 % more expensive than one with 2 pipes. All systems were designed in a way that they cover the demand of the settlement specified above with a saving of primary energy of 50 - 60%. So each system concept has different size of components like collector area and storage size. The selection of variant was done acxording to economic measures like total investment, operational costs and thermal energy prize.
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A solar collector array of 800 m 2 delivers the gained heat either to the pit (high priority) or to the duct store (low priority) with respect to the temperature levels in the two storage parts. There is also an option of direct delivery to the district heating. This occurs mainly in spring, summer and autumn for domestic hot water but also space heating. In case of heat demand first the pit as shortterm storage is discharged. The duct store can either be discharged directly if the temperature is high enough or via a heat second heat pump.
Fig. 10: Scheme of the central solar system with seasonal hybrid storage and supplementary heating with a central oil boiler The results of all simulations cannot be presented here in detail except the most promising layout which is a central solar system with an hybrid storage and two central heat pumps as shown in fig. 11.
Fig. 11: Scheme of the central solar system with seasonal hybrid storage and supplementary heating with two central heat pumps
Fig. 12: Annual energy balance
The characteristics of components are the following: 800 m 2 solar roof collector in 2 arrays on a sports building 500 m 3 pit store, 8.9 m diameter, 8 m deep 9350 m 3 duct store made of 110 boreholes 25 m deep heat pump to the district heating 200 kW for 0~ evaporation and 50 ~ condensation temperature heat pump for discharging of ducts 60 kW for 0~ evaporation and 50 ~ condensation temperature district heating net (approx. 500 m) The big advantage of using a heat pump is this system is the increase of the operational temperature range of the storage system. Additionally the duct storage discharging heat pump is mainly operated during night at a special electricity tariff heating up the pit. So the temperature range between the duct store and the district heating net can be split in two parts and thus a two stage heat pump process is realized resulting in a higher COP. As a result of the system simulation the annual heat balance is given in fig. 12. The 800 m 2 solar collector deliver 415 MWNa useful heat 543 MWh/a are losses. Thus the average collector efficiency is 43%. 279 MWNa are stored in the pit and 139 MWNa in the duct storage. The heat losses of the pit to the duct store amount to 18 MWNa. The pit storage runs about 9 full storage cycles (10 90~ whereas the duct store runs only one. About 109 MWNa of electricity are necessary to operate the heat pumps which corresponds to 285 MWh/a of primary energy. The COP of the storage discharging heat pump amounts to 5.9 that of the district heating one is 4.7. These high values are due to the rather low temperature differences they have to cover. Compared to remote oil boilers in each house the primary energy savings amount to 57 % which results in a reduction of CO2-emission of 121 t/a or 60%.
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5. COST ANALYSIS As the community is very interested in realizing a solar district heating even in that small scale the cost analysis was a very important point. Therefore already during system modeling a cost estimation was carried out for selection of system. Each component of the system was designed by modeling and the costs evaluated on basis of offers. For comparison the cost for a conventional heating system (oil boiler) in each house was determined. The components are: Component heating center (building) district heating net solar system seasonal storage heat pumps system control planning of the system total
Costs DM 58,000 280,000 450,600 332,000 91,000 179,500 190,000 1,581,100
The costs for conventional remote heating systems in each house are 810,000 DM. Therefore the extra costs are 771,100 DM or 25,700 DM for each house. According to the VD12067 guidelines the investment is capitalized with an interest rate of 5% over a 20 years period. Combined with the operation costs a energy prize (full cost basis) of 200.- DM/MWh for conventional systems and 282.- DM/MWh for the solar district heating was determined. 6. C O N C L U S I O N S Seasonal thermal energy storage is an important issue in solar district heating. Different types of underground storage systems are feasible like water pits, duct, aquifer or rock cavern stores under different geological conditions. A new storage concept, a combination of pit and duct storage, is investigated because of its technical and economical advantages. In a first step technical problems of the pit were analyzed which can be caused by thermal stress in the walls. The most endangered parts, the upper and lower comers of the walls, will see a maximum temperature gradient of about 14 K in the concrete. Thus expensive thermal insulation can be avoided. Additionally no cracks due to thermal stress are expected which means only a small leakage and so it is planned to do without liner.
System simulations were carried out for 20 single family and 5 twin houses build in low energy standard (25% lower than WSchV'95) with solar district heating. A solar system of 800 m 2, a hybrid storage of 500 m 3pit with 9350 m 3 duct store and supple-
mentary heating with heat pumps give a solar fraction of about 57 % for space heating and domestic hot water. Such a system allows annual savings of 121 t/a of CO2 compared to conventional oil heating. The cost analysis as an important factor for the feasibility shows a total investment of 1,581,100.-- DM which means about 771,100.-- DM extra costs compared to the conventional system. Therefore the prize for thermal energy was determined to be 282.DM/MWh compared to 200.- DM/MWh for conventional heating. With the proposed system concept even smaller developments can be equipped with solar district heating for reasonable additional costs. This is important because new developments in rural communities but also in bigger cities tend to smaller size nowadays. 7. ACKNOWLEDGMENT This research was carried out in close cooperation of: Bayerisehes Zentrum far Angewandte Energieforschung (ZAE), Bayerisehe Landesanstalt fiir Landtechnik der TU-Miinchen, Institut fiir Allgemeine und Angewandte Geologie (IAAG) der Universit~t Miinchen and the Company Dywidag. This joint research project (SOLEG) is funded by: Bayerische Forschungsstifhmg. REFERENCES
/1/ G. Bakema, A.L. Snijders, B. Nordell: Underground thermal energy storage, state of the art 1994. ISBN 90802769-1-x, Amhem, The Netherlands, 1995 T R N S Y S - A Transient System Simulation Program. Solar Energy Laboratory, University of Wisconsin- Madison, W153706 USA
/2/
/3/ V. Lottner, E. Hahne: Status of seasonal thermal energy storage in Germany. In Proceedings of Megastock 1997, Vol. 2, pp. 93 l, Sapporo, Japan 1997 /4/ D. Pahud, G. Hellstrfm, L. Mazzarella: Heat Storage in Ground, Duct Ground Heat Storage Model for TRNSYS (TRNVDST) (Type 141) - User manual for the October 1996 version, Lausanne, Switzerland 1997 /5/ H. Driick, T. Pauschinger: MULTIPORT Store- Model for TRNSYS (Type 140) version 1.90, March 1997, Stuttgart, Germany 1997 /6/ M. ReuB, J./driller, P. Rogmann: Solare Nahw~rmeversorgung im Neubaugebiet "Am Sportplatz" der Gemeinde Attenkitchen, Machbarkeitsstudie im Auftrag der Gemeinde Attenkirchen, Freising, Germany 1998
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SOLAR HEATING WITH HEAT PUMP AND ICE S T O R A G E Anton B. Schaap and Jos. M. Warmerdam Ecofys, P.O. Box 8408, 3503 RK Utrecht, The Netherlands, tel. +3130-2808300, fax +3130-2808301, [email protected]
Eobert E. Gramsbergen Gramsbergen, Tus 7, 5507 MG Veldhoven, The Netherlands, tel. +3140-2301500, fax +3140-2301499, Gramsbergen. [email protected]
Abstract - In the Netherlands the SPF of an electrical heat pump system should be higher than 2.5 in order to reduce the CO2 emission in comparison with a natural gas heating system. A system with electrical heat pump, covered solar collectors, ice storage and sewage heat recovery was designed and tested. This system is interesting because it promises low costs per ton CO2 saved per year at high CO2 savings. The SPF can be as high as 5 (including the direct delivery of solar heat to the house in spring, summer and autumn). The system, as it was installed and monitored in Veldhoven The Netherlands, works, but the performance is not yet good enough. The sewage heat recovery performs well, but the ice storage performs less than calculated. The collectors perform at 80 % of the calculated value at temperatures around 0 ~ supply temperature. The consumed auxiliary pumping power is still to high.
1
INTRODUCTION
The application of electric heat pumps for domestic hot water and space heating can deliver a substantial reduction of the CO2 emission allocated to this heating purposes. However the efficiency of the heat pump strongly depends on the available heat source and so does the CO2 reduction. The economy of the system depends on the efficiency of the system and on the local tariffs for electricity and natural gas (for the Netherlands). This means that the heat source of the electrical heat pump is of decisive importance for the feasibility of the heat pump system in comparison with a natural gas heating system for CO2 reduction as well as cost effectiveness. 2
DESCRIPTION OF THE DUTCH SITUATION
The Dutch situation of space heating and hot water differs in a number of ways from the situation in the surrounding countries. The main characteristics of the Dutch situation are: 9 The country is densely populated so the costs for a natural gas grid connection are low (about euro 500 to euro 700). Almost all houses are connected to the natural gas grid. 9 The emission of CO2 per kWh electricity is relatively high (about 0.56 kg/kWh) because of the utilization of mainly coal (about 40 %) and natural gas (about 50 %) for electricity generation. 9 The electricity price for small consumers is high in comparison with the natural gas price (about euro 0.03 per kWh for natural gas compared to about euro 0.12 per kWh for electricity). 9 Low temperature heating systems (floor heating or air heating) are not common. Almost all systems comprise of wall mounted radiators. 9 New houses are (by law) energy efficient, and will even be more so in the near future when the building requirement (The "EPN" = energy efficiency standard) is improved. For a standard 100 m 2 single family house (built in a row) this will mean about 19.5 GJ per annum (54 kWh per m 2) for space heating in the year 2000. 9 High efficiency ventilation heat recovery heat exchangers are now on the market (efficiency up to 90%).
9
9
9
9
3
The sales of heat pumps are very low. Heat pumps have until this date only been installed in demonstration projects. Houses are built merely as housing schemes of more than one house. The houses are planned and realized by semi governmental organizations or by housing scheme developers. The houses are relatively small (a living space of about 100 m 2 and a content of about 250 m3). There is only a very small space for the heating system. The gardens of the houses (if any) are small. Houses with more than 200 m 2 of total area are rare.
NATURAL GAS HEATING SYSTEM
The houses are mostly heated with high efficiency natural gas central heating systems combining space heating and hot water (mostly without storage). The efficiency for space heating is above 85 % and for hot water production around 65 %. Because of the use of natural gas the CO2 emission is relatively low. The space dimensions of the system are about 400 mm deep 500 mm width and 800 mm high. The system is wall mounted (about 1 m above the floor) so it requires no floor space. The total investment for the complete heating system is about euro 3500 (including heat delivery system). The capacity is high (around 25 kW) in order to produce about 6 liter of hot water per minute. The heating system is mostly controlled by a cheap (about euro 150) but accurate and comfortable self adapting controller mounted in the living room (with several programmable functions; for instance week patterns for room air temperature) combined with thermostatic valves on the radiators in the other rooms. 4
INTRODUCTION OF THE HEAT PUMP
The introduction of electric (compression) heat pump systems is sponsored by the government and by the utilities. The goal is to reduce the CO2 emission for space heating and hot water in the domestic sector. Because of the specific Dutch situation the heat pumps have to comply with several demands:
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9
9 9 9
9
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5
Because of the high CO2 production per kWh electricity the heat pumps must have a high SPF (Seasonal Performance Factor = average COP over one year and COP = Coefficient Of Performance) in order to realize a significant CO2 reduction in comparison to the natural gas heating system. At a SPF of 2.5 the CO2 production is equal to that of a natural gas heating system. Heat pumps with ambient air as heat source can hardly meet this requirement (Schaap, 1997). The system must be cheap and very compact. The systems must have a high comfort level and the control system must be very flexible. Because of the high investment, the heat pumps can only be introduced in new building schemes that will not be connected to the natural gas grid. The houses must be very energy efficient, because the price of the heat pump and of the heat source are almost proportional to the installed capacity. There has to be a special low electricity price per kWh for the heat pump system (about euro 0.06 per kWh). The low price could be justified when the utilities have the right to interrupt the electricity deliverance on peak hours.
POSSIBLE HEAT SOURCES
There are numerous ways to extract low temperature heat from the surroundings of a building. The most important heat sources are: the ambient air, the ground, the ground water, the surface water, the sewage water and the solar irradiation. The last two are being discussed in the next chapter. For a compression heat pump the efficiency is related to the temperature of the heat source; the higher the temperature the higher the COP.
5.1
Ambient Air
As stated above the temperature of the ambient air is to low during wintertime to achieve a high SPF. The yearly CO2 reduction is not high enough, in comparison to a natural gas burner, to justify the higher exploitation costs of the heat pump.
5.2
Ground water
Ground water is the best source of heat around the house in wintertime. Under about 3 m of depth, the temperature is steady around 10 ~ all year round. Almost anywhere in the Netherlands there are aquifers, from which ground water can be extracted. The water has to be injected in the aquifer again, when used for heating or cooling purposes. Such a system is being used on several locations in the Netherlands for the direct cooling of buildings and production processes. The problem is that this system is to complex for a single family house. It can however be used to feed a network of about 200 houses. Such a system is planned for a part of the new housing scheme "Broekpolder" in Beverwijk, the Netherlands (Klaver, 1997).
5.3
Verticalground heat exchanger
Second best option to extract heat from the ground is the vertical ground heat exchanger. In this case a closed loop is inserted vertically to a certain debt. The heat has to flow to the closed loop by ground conduction mainly. This makes the
achievable temperature 5 to 10 ~ lower than by using ground water. A drawback of vertical heat exchangers is that they can't easily be removed after the life of the system. There is however a possibility to combine the vertical heat exchanger with the supporting pillars of the houses. In some areas these pillars are rammed to a depth of 20 m to reach the first stable sand layer. If used by all houses in a neighborhood the ground has to be regenerated with summer heat. The average source temperature (just above 0 ~ is somewhat higher than with ambient air as a heat source, but the COe reduction is still not very impressive (see chapter 7).
5.4
Horizontal ground heat exchanger
A horizontal ground heat exchanger consists of a number of closed loops inserted horizontally at a depth of about 1 to 2 m. Because of the small average area per house in the Netherlands a horizontal ground heat exchanger can seldom be used.
5.5
Surfacewater
Surface water is a rather difficult heat source, because it freezes at 0 ~ This makes shallow waterways unsuited. Only deep lakes and rivers can be used. These are only at a small number of locations available. DESCRIPTION OF THE SOLAR HEAT PUMP WITH ICE STORAGE A heat pump system which uses the solar irradiation and the sewage water as a heat source was developed and tested. The system had to fulfill two important goals: 9 The system can be utilized for any single (very energy efficient) house anywhere in the country, with no dependence on the structure of the ground nor on the available ground space. 9 The system had to have a very high SPF in order to reach a considerable CO2 reduction.
6.1
Descriptionof the system
Starting point is the heat demand of a very energy efficient single family house. For space heating the heat demand amounts to about 12 GJ/year (= 33 kWh/m2) and for domestic hot water about 8 GJ/year (floor area typically about 100 m2). The solar heating system with heat pump and ice storage is designed to meet de energy demand of such a house (see fig. 1).
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the ice storage. It would operate at a source temperature around 0 ~ all winter. In the summer the combination of daily storage, hot water storage and solar collectors delivers directly hot water to the house. In spring and fall the solar collectors can also bypass the heat pump for the deliverance of (low temperature) space heating to the house, if a high enough temperature is available. For this bypass, a heat exchanger is installed parallel to the heat pump. Because of the high temperatures from the solar collectors in spring, fall and especially in summer, the SPF can be very high. It must be possible to reach a SPF up to 5 (including the direct solar delivery in summertime; Schaap, 1997).
6.2
Fig. 1. The heat pump system with solar collector, ice storage and sewage heat recovery. The solar thermal collectors are being used as a heat source for the (electric) heat pump (3.5 kW thermal capacity). However when the ambient temperatures are low (below 0 ~ and the solar irradiance is low, then the contribution of the solar collectors (10 m z single glazed) is insufficient. During these periods the heat is withdrawn from an "ice storage" with a content of 3 m 3. The uninsulated ice storage (maximum temperature about 15 ~ can be buried under or next to the house (see fig. 2).
Rain water storage (cistern)
The ice storage is also used as a rainwater storage. The r a i n water is used for non-drinking water purposes in the house (flushing the toilet, washing clothes etc.). In this way part of the investment of the ice storage can be allocated to the rain water system. In winter only part of the capacity can be used for the rain water system. In summer the full capacity can be used. In this period the ice storage is not used for the heating system. For rain water storage's the (sometimes dry) summer is the important period.
6.3
System of the Fachhochschule Lfibeck
In Germany a comparable system was designed and tested in the early eighties by the Fachhochschule Liibeck (see fig. 3). The system had two separate collectors. Single glass fiat plate collectors as a heat source for the heat pump and evacuated tube collectors for the direct deliverance of heat to the DHW. The high temperature storage is placed in the seasonal storage (ice storage) to collect the heat losses of the high temperature storage (Weik, 1984).
Fig. 2. Ice storage with heat recovery sewage tank The water in this storage can be transformed into ice, by circulating a antifreeze mixture through the heat exchanger, that runs all through the storage. In this way also the very high latent heat of the water is used to enlarge the thermal capacity. A temporary storage of about 300 liter for the sewage (toilet excluded) is placed in the ice storage. The ice storage recovers the heat from the sewage in wintertime (about 5 GJ/winter). Between solar collectors and heat pump a daily storage is placed (see fig. 1). This storage acts like a domestic hot water system in summer. During the heating season the daily storage improves the COP by storing the somewhat higher temperatures coming from the solar collectors. Without the daily storage the heat pump is always connected to
Fig. 3. System with ice storage and solar collectors of the Fachhochschule Liibeck The system was installed at a research house of the Fachhochschule. The heat demand was 43.2 GJ for space heating and domestic hot water. The differences with the new system are: 9 There is no regeneration of heat from the sewage water, therefore the ice storage has to be bigger. 9 There is a big storage at high temperature, therefore high heat losses occurred.
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9
There is no daily storage at the evaporator side, therefore the heat pump feels the temperature of the ice storage all winter. 9 There are several separate circuits, therefore a lot of heat exchangers had to be used. This system reached a SPF of about 3.3 in the season 82/83 (all year round so including the direct solar delivery in summertime). This seems not good enough to compete with less complicated systems. COMPARISON BETWEEN DIFFERENT HEAT SOURCES In order to determine the cost effectiveness of the new system several calculation were made. First a simulation program was written, with which the combinations of heat pump and different heat sources could be calculated. After that cost figures were gathered and a cost calculation was made for the different combinations and compared to a reference system (a natural gas burner). The costs were estimated on the basis of the production of a reasonable amount of systems per year; otherwise the reference system (that is produced in huge quantities) was to much in favor. The system combinations were judged on the basis of exploitation costs, primary energy use and costs per ton CO2 saved per year.
7.1
Simulation program
Because of the highly dynamic nature of a system with solar collectors a detailed calculation is necessary. The simulation program enables to calculate the system parameters every ten minutes. For the weather data a local Test Reference Year is used. The hourly data are interpolated for the ten minute values. The simulation system can simulate different heat sources like a vertical heat exchanger, a horizontal heat exchanger, an ambient air heat exchanger, an ice storage with sewage heat recovery and different types of solar collectors (Sehaap 1997). Fig. 4 gives the yearly temperature in the daily storage and the ice storage for the system with solar collectors, ice storage and sewage heat recovery (system 2).
1 Reference system (natural gas burner) The reference system is a high efficiency natural gas burner for combined space heating and domestic hot water. This is also the auxiliary burner in case of the systems without heat pump. The yearly efficiency is 85 % (higher calorific value) on space heating and 70 % on domestic hot water. 2 Heat pump, ice storage and covered collectors The system with heat pump, ice storage and covered collectors has 10 m2 single glazed covered collectors, a daily storage of 300 liters, an ice storage of 3 m 3 and a regeneration tank of sewage water (300 liter). The heat pump has a nominal thermal capacity of 3.5 kW. 3 Heat pump, ice storage and uncovered collectors This system is equal to system 2, but in stead of single glazed covered collectors, 10 m e of uncovered collectors are used (swimming pool collectors) 4 Heat pump, and uncovered collectors This system has only uncovered collectors as a heat source for the heat pump (10 m2). This means that for extreme cold weather an electric resistance heater has to be added. 5 Heat pump, ice storage and ambient air heat exchanger As system 2, only in stead of the collectors an air to water heat exchanger is being used. The heat exchanger extracts heat from the ambient air. 6 Heat pump and vertical ground heat exchanger A vertical ground heat exchanger is being used as a heat source for the heat pump (one single loop to a depth of 40
m). 7 Heat pump, vertical ground heat exchanger and covered collectors As system 6, but 10 m 2 of single glazed solar collectors and a daily storage are added. The solar collectors can lift the source temperature in winter, can regenerate the ground and can supply domestic hot water directly outside the heating season. 8 Heat pump, vertical ground heat exchanger and uncovered collectors As system 7, only with uncovered collectors, in stead of covered collectors (10 m2). 9 Heat pump and horizontal ground heat exchanger System with a horizontal ground heat exchanger as a heat source for the heat pump (tubes inserted over an area of 50 m 2 at a depth of 1 m).
Fig. 4. Temperature of the daily storage and of the ice storage over a year (daily average values)
7.2
Different systems to compare
A comparison was made between a number of different heat sources including the system with ice storage. Also some solar systems without heat pump were added. The systems are described below. The comparison is based on investment, exploitation costs, primary energy use and costs per ton CO2 saved per year.
10 Heat pump and aquifer system System with an aquifer as the source of ground water from which the heat is extracted. The aquifer has to be regenerated in summer (with ambient air). The aquifer serves about 200 houses. The heat pumps are placed in the houses, to avoid distribution losses (so the distribution network can be uninsulated; Klaver 1997) 11 Heat pump, aquifer and covered collectors As system 10, only with ~ v e r e a collectors to regenerate the aquifer. In this way the aquifer can be brought on a slightly
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higher temperature level, to increase the COP of the heat pump. The collectors can also deliver DHW directly to the houses. The supply line of the distribution network has to be insulated.
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pensive than the reference system). The system with ice storage and covered collectors is about 8500 euro (system 2). So this is one of the more expensive heat pump systems. The most expensive is the system with a seasonal storage and without heat pump (system 14 = about 16000 euro).
12 Reference system with solar DHW system Reference system with a standard solar DHW system (2.8 m E of solar collectors and a storage of 100 liter). The energy production of such a solar system is about 3.6 GJ/year. 13 Reference system with solar heating and DHW system Reference system with an enlarged solar system for space heating and DHW. The system has 10 m E of collectors and 300 liter of storage. 14 Solar system with seasonal storage For this system the energy demand for space heating is reduced from 12 to 8 GJ/year, by installing triple glass with krypton filling and by enlarging the insulation thickness of walls, floor and roof from 200 to 300 mm PS foam. The solar system consists of 17 m 2 double covered high temperature collectors (one layer of glass and one layer of teflon foil) and a seasonal storage for a single family house. The solar heat is stored at temperatures up to 90 ~ and delivered directly to the house (no heat pump). The seasonal storage has a content of 30 m 3 of water and has an insulation thickness of 300 mm (Schaap, 1995). The reference system (natural gas burner) serves as auxiliary heater. Internationally a seasonal storage is mostly designed to serve a number of houses (Fisch, 1998). The heat is distributed by a district heating network. The construction mostly consists of a concrete (Germany) or steel (Sweden) insulated tank of about 100 to 10.000 m 3. In the Netherlands however there are two problems with these constructions (Schaap, 1995): ,, A huge concrete or steel tank will not be accepted in a neighborhood. The storage has to be placed underground. 9 The water table in the ground is mostly very high (about 0.5 to 2.0 m under ground level) and the ground is very unstable. The storage has to be watertight not only at the inside but also at the outside, or the insulation material must be water-resistant and watertight (for example foam glass). These factorsmake the construction of such a tank expensive and complicated. W h e n the seasonal storage can be integrated in the construction of the house, part of the costs can be shared (for instance the support pillars).
7.3
Costsand C02 comparison
The calculations are based on the Dutch electricity and natural gas costs for small consumers; respectively Euro 0.12 per kWh and Euro 0.29 per m 3 of natural gas. For space heating the heat demand of the very energy efficient house amounts to 12 GJ/year (= 33 kWh/m ~) and for domestic hot water to 8 GJ/year (floor area typically about
1O0 mE). As we can see in fig. 5 the reference system (system 1; natural gas burner) is the cheapest system Somewhat more expensive is the reference system with a standard solar DHW system (system 12). The cheapest heat pump systems (system 4, 6 and 9) are about 6000 euro (about 2500 euro more ex-
Fig. 5. Investment for the different systems. The primary energy consumption is calculated with the simulation program and given in fig. 6. As we can see the reference system (system 1) has the highest primary energy consumption. The system with seasonal storage (system 14) has the lowest primary energy consumption. System 2 (the system with ice storage) comes out third best together with the system with an aquifer as a heat source and solar collectors (system 11). The system with a vertical ground heat exchanger and solar collectors (system 7) is somewhat less good. This is because there is more solar energy needed to regenerate the ground in comparison with system 2, while there is no sewage heat recovery.
Fig. 6. Primary energy consumption of the different systems Fig. 7 gives the exploitation costs, the C O 2 emission and the costs per ton CO2 saved per year. The reference system has the lowest exploitation costs, so there is no alternative system that is cost effective. From this figure we can see that the system with ice storage is attractive on the basis of the costs per ton CO2 saved. The investment is rather high, but the CO2 emission is low, so the costs per ton CO2 saved are relatively low.
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ISES Solar World Congress 1999, Volume III
Fig. 7. Comparison of the systems on the basis of the exploitation costs, The CO2 emission and the costs per ton CO2 saved per year. In most eases uncovered collectors are being used as a heat source for heat pumps. From the simulations it was coneluded that in combination with an ice storage with sewage heat recovery, spectrally selective covered collectors gave a higher SPF. The costs per avoided ton of emission of CO2 are also lower. It appeared that covered collectors have two advantages in this combination: 9 In summer covered collectors earl deliver domestic hot water directly to the house and in spring and autumn part of the heating demand, 9 In winter the ice storage with sewage heat recovery restrains the heat source temperature to just below 0 ~ With outside temperatures below 0 ~ the contribution of uncovered collectors is lower than that of covered collectors. 8
DEMONSTRATION AT THE HOUSE GRAMSBERGEN IN VELDHOVEN
8.2
Descriptionand heat demand of the house
The house is built as a row of two houses. The house consists of a living part and an office. The total floor area is 163 m 2. The total heat loss area is 334 m 2 (see fig. 8). The house is heated with floor heating (84 m 2) and wall heating (49 m 2) don stairs and in the bathroom and enlarged radiators at the rest of the second floor. Because of the high heat transfer the supply temperature of the heating circuit could be kept under 30 ~
OF
With financial support from NOVEM the system is built by Gramsbergen in a new housing scheme in Veldhoven in the Netherlands. This location was chosen because of the experience of Gramsbergen with the implementation of renewable energy systems and the availability of his house under construetion. However the energy efficiency of the house couldn't be influenced very much.
8.1
tion from the different appliances was measured. The power consumption of the heat pump was logged on a separate channel. The system was controlled by a personal computer on which control and measuring software was running. This enabled us to chance the control settings very easily. Also all the control settings were logged. All measurements were logged to disk every minute during the heating season 98/99. However we had to use almost all winter to eliminate the faults in the system and to optimize the system. Only from 17 of February the system was running without interruption. Therefore it is not possible to present overall data from the whole heating season. The data that are presented have been extracted form a number of days on which the system was running well. The results presented are preliminary and not yet fully complete, because the analysis of all the data will take some more time.
Fig. 8. The house of Gramsbergen at Veldhoven the Netherlands. On the roof the solar collectors and in front of the office the underground ice storage after installation.
Descriptionof the monitoring system
The system was monitored extensively by Ecofys during the last heating season 98/99. The heat flows in the different circuits were measured as well as the heat demand of the domestic hot water system. The temperature was measured on 16 locations, under which the room temperature, the ambient temperature, the collector temperature and the temperature in the daily storage. The temperature in the ice storage and in the sewage tank was measured at three different heights (bottom, middle and top). This enabled us to study the thermal stratification in these tanks. In the ice storage also the water level was measured, to determine the amount of ice in the storage. During the monitoring the ice storage was not used as a rain water storage. The electrical power consump-
Double glazing is applied with a heat loss factor of 1.8 W/m2IC The heat loss factor for the walls is 0.4 and for the roof 0.3 W/m2K. The house has a calculated space heating demand of 32 GJ/year (54 kWh/m2 year) for a reference winter, so it isn't a very energy efficient house. From the measurements it appeared however that the energy demand was far higher than calculated (see fig. 9).
ISES Solar World Congress 1999, Volume III
'~ - ~ " ~ 4
reached would be 4.2 over a whole year (including direct delivery of solar heat in summer). When the amount of ice is limited to 4000 kg (the capacity of the ice storage) then the SPF drops to 3.6. The lower SPF is caused by the fact that part of the winter the electrical resistance auxiliary heater has to be used to limit the ice formation. However with a heat demand for space heating of 60 GJ the SPF sinks further to 2.3. To improve the system on this specific location an ambient air heat exchanger will be placed to enlarge the amount of source heat that can be drawn into the system.
M eas ured
"
.
G. 2
Calculated
-5
~
e
0
,
5
481
10
Tam b l e n t [oC]
Fig. 9. Heat demand of the house (measured and calculated) as a function of the ambient temperature. The points depict average daily measured values. The measured values are almost twice as high as the calculated values. The reason for this could be: 9 The ventilation heat recovery (efficiency 90 %) was not yet installed. According to the calculations this heightens the heat demand with 40 %. 9 A newly built house contains a lot of moisture in the walls and floors. The evaporation of this moisture extracts heat from the home. 9 At certain points the insulation from the house is insufficient. For instance there is wall heating in the wall between the two houses. This wall however consists of two separate blades with an air gap in between. The air gap is not insulated where it meets the files on the roof of the house.
Ice storage Figure 10 gives the relationship between the length of the heat exchanger tube and the resulting temperature difference when extracting 3500 W through an ice layer. The calculation was made for the situation with a heat exchanger tube with an outside diameter of 20 mm and a cylindrical ice cover with a diameter of 150 mm. The heat conduction of ice is rather high (2 W/m K) in comparison with water (0.6 W/m K). I""l
35
0
.o. 30 q) r
= 25
~- 20 12 9
L._
15
Q
E
5
i-
0
Q
I
0
100
i
i
200
300
Length [m]
Before the next winter we plan to improve the energy efficiency of the house by installing ventilation heat recovery and by improving the insulation of the air gap between the two houses. When the measured heat demand for space heating is calculated over a whole heating season we get around 60 GJ (= 100 kWh/m 2) in stead of the calculated 32 GJ/year.
8.3
Dimensioningand performance of the system
For a very energy efficient single family house (typically a heat demand of around 12 GJ/year = 33 kWh/m 2 year at 100 m 2 floor area) the idea is to choose the dimensions of the system in such a way that the ice storage gets completely filled with ice only during an extreme winter. During normal winters there is hardly any ice formation in the ice storage (see fig. 2). The SPF calculated with the simulation program is around 5 (Schaap, 1997). However for the house of Gramsbergen we have chosen smaller dimensions for the system (relatively smaller collector (16 m 2) and ice storage (6 m3), for three reasons (Schaap, 1998): 9 By the building commission on aesthetics we were not allowed to place a larger collector area. 9 There was not more space in front of the house to install the ice storage. 9 It is important to the development of the system with ice storage to study the ice formation process. The yearly simulation showed that with a space heating demand of 32 GJ/year (and 8 GJ for domestic hot water), 8900 kg of ice would be formed during a normal winter. The SPF
Fig. 10. Temperature difference when extracting 3500 W through an ice layer for the situation with a diameter of 150 mm ice cover around the tube (outside diameter 20 mm) We can see that the temperature difference is about 5 ~ at a length of 200 m. A diameter of 150 mm of ice implies that there is about 3.5 m 3 of ice in the storage at a tube length of 200 m. Because of the logarithmic relationship between the heat conduction through the ice mantle and the thickness of the ice mantle, the temperature difference doesn't rise very strongly with rising ice mantle diameter (until the ice mantles of the different tubes grows together). From the measurement data it could be extracted that at a temperature difference of 5 ~ between ice storage and water glycol mixture (average between supply and return) a heat transfer of 2500 W was reached, without ice around the tubes. When the storage was filled with ice (4 m3), the heat transfer is lowered to around 1500 W. It seems that not only the thickness of the ice mantle forms an important resistance (conduction) but also the heat transfer by convection from the water to the ice mantle, because the expansion coefficient of water just above 0 ~ is very low. Because of the lower heat transfer in the ice storage, the working temperatures at night at the evaporator side of the heat pump were between - 5 and - lO~ Figure 11 shows the ice forming process during the winter 98/99. Part of the data is missing. The ice formation was monitored by measuring the water level in the ice storage and
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482
calculating the amount of ice from the density difference between ice and water. As we can see around mid February the ice formation is about 4000 kg. This appeared to be the maximum value because the ice formation reached the sewage tank, and we can't allow the sewage to freeze. The content of the tank is about 6 m 3, so 2 m 3 cannot be frozen. In the design of the ice storage we had left a space free of heat exchanger tubes so that every part of the ice storage was accessible for maintenance. 500o
,"V
,--,4OOO 13D
evaporator and condenser heat exchanger. The water-glycol mixture has especially at below zero temperatures a high viscosity. Very width pipes are needed to allow for high mass flows at low pumping power. We had all pipes (except for one connection on the daily storage) laid out somewhat bigger than standard, but this appeared not enough. The collectors are of the serpentine flat plate type. There are three collectors in parallel. The inner diameter of the serpentine is only 8 mm (so called low flow collector). The consequence of this is that we had to install pumps with an electric power of 180 W each (in stead of the around 50 W standard pumps estimated).
f
,._=
oo 3 0 0 0 0
==2000 el
~= 1000
11-.dec 21-dec 31-dec lO-jan 20-jan 30-jan 9-feb 19-feb Date
fig. 11. Ice formation in the ice storage during winter 98/99. When the ice storage is full with ice, the source heat for the heat pump has to be delivered by the electrical auxiliary heating in the daily storage. Collector With the simulation program the measured heat delivery by the collectors was compared to the simulated yield. This was done for three separate days: 9 12 Feb. 1999: Measured yield 78 % of simulated value. 9 12 March 1999: Measured yield 8 1 % of simulated value. 9 30 April 1999: Measured yield 95 % of the simulated value. On 12 Feb. the supply temperature to the collectors was around 0 ~ At 12 of march the supply temperature was around 10 ~ and at 30 April around 50 ~ So the higher the supply temperature the higher the measured yield in relation to the calculated yield. R seems that the viscosity of the water glycol mixture is of influence on the performance of the collectors. The viscosity of the mixture was not a parameter in the simulation program.
Heat pump The measured COP of the heat pump was about 10 % lower than specified by the supplier. Because of the high heat demand of the house, the ice storage was during part of the winter filled with ice. The source temperature of the heat pump was under that circumstances around -10 ~ This was also the lowest working temperature of the heat pump. Several times the heat pump was switched off by a safety switch because of a to low pressure on the evaporator side of the heat pump. Auxiliary_ Energ?/ Using a mixture of water and 30 % propylene glycol as the heat transfer medium brings greater consequences than anticipated. For the good functioning of the heat pump a high mass flow is needed (or a small temperature difference) over
Sewage heat recovery_ The sewage heat recovery is performing well (see fig. 12). The figure shows what happens on a typical morning when the hot water enters the sewage tank in the ice storage. A strong stratification is created because the hot water is entering the cylindrical tank horizontally. The ice storage is at a temperature of 0 oC at all heights and filled with about 4 m 3 of ice. We can see that the heat from the sewage is transferred to the ice storage within about 8 hours. The average time over which the sewage water stays in the sewage tank is more than 24 hours, so almost all heat above 0 ~ is recovered from the hot sewage water as well as from the cold sewage water. 35
~, 25 O
~ 20 ~ ID D.
15
E
~. lO
|
,
,
,
0
5
10
15
Time [hours]
Fig. 12. Temperature in the sewage tank (top, middle and bottom). Date; 12-2-99 and starting time 6.00 hour. We designed the sewage tank with the idea that dirt might stick to the wall of the tank and might obstruct the heat transfer. However this doesn't happen. The dirt is deposited at the bottom of the tank. After about half a year the sewage tank was cleaned and about 2 liters of dirt were collected from the bottom. EFFICIENCY OF A SYSTEM FOR AN ENERGY EFFICIENT HOUSE With the simulation program we can calculate what the SPF would be of the system as it is performing in Veldhoven, but installed at a very energy efficient house (a heat demand of 12 GJ/year - 33 kWh/m2 year at 100 m 2 floor area and 8 GJ/year for hot water). Dimensions of the system; 10 m 2 of covered solar collectors and an ice storage of 3 m 3, with 200 m of heat transfer tube and a heat pump with a thermal capacity of 3.5 kW. The following corrections have to be made on the simulation as a result of the monitoring: 9 Heat pump performs 10 % less.
ISES Solar World Congress 1999, Volume III
9 9
Collectors perform around 20 % less. Auxiliary electricity will be around 100 W for each pump in stead of 50 W. With these settings the simulated SPF decreases from 5.1 to 3.8 (including direct delivery of solar heat to the house). This drop cannot be compensated by enlarging the collector area nor the ice storage, because it consists mainly of direct electrical consumption (heat pump and especially the auxiliary electricity). Necessary improvements of the system: 9 Collector with lower pressure drop and higher heat transfer to water-glycol. 9 Lay-out with lower pressure drop in the different circuits (collector, ice storage and evaporator circuit) 9 Consult the manufacturer of the heat pump in order to clarify the difference between the measured and specified COP. I0 9
9
9
9
CONCLUSIONS In the Netherlands the SPF of an electrical heat pump system should be higher than 2.5 in order to reduce the CO2 emission in comparison with a natural gas heating system. The system with electrical heat pump, covered solar collectors, ice storage and sewage heat recovery is interesting because it promises low costs per ton CO2 saved per year, at high CO2 savings. The system, as it was installed in Veldhoven, works, but the performance is not yet good enough: 9 The ice storage performs less than calculated. 9 The sewage heat recovery performs well. 9 The collectors perform at 80 % of the calculated value at temperatures around 0 ~ 9 Auxiliary pumping power is to high. Necessary improvements to be made: 9 Collector with lower pressure drop and higher heat transfer to water-glycol at low temperatures. 9 Lay out of the system with lower pressure drop 9 Ice storage with relatively longer heat transfer tube.
REFERENCES
Weik H., Plagge J. (1984),Voll-Wiirmeversorgung von Wohngebauden dutch exergetische und Anergetische Nutzung der Sonnenenergie, 5. ISF, Berlin. Schaap A.B., Veltkamp W.B. (1995), Design of an Energy Matched Single Family House, ISES Solar World Congress Harare Zimbabwe. Schaap A.B. (1997), Haalbaarheid van een zonneverwarmingssysteem met warmtepomp en maandopslag, Ecofys, Utrecht, The Netherlands, E1084. Klaver P.J., Zegers F.T.S., Schaap A.B. (1997) Toepassing van warmtepompen en lage-temperatuur warmtedistributie voor energiezuinige woningen in de Broekpolder, Ecofys, Utrecht, The Netherlands, E 514.
483
Schaap A.B. (1998), Dimensionering Zonnewarmtepomp voor de woning van Gramsbergen, Ecofys, Utrecht, The Netherlands, E 1130. M.N. Fisch, M. Guigas, J.O. Dalenb~ick (1998), A review of Large-Scale Solar Heating Systems in Europe, Solar Energy vol.63, No. 6.
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484
AN ANALYSIS OF PHASE CHANGE HEAT TRANSFER IN A SOLAR THERMAL ENERGY STORE
Anica Trp, B e m a r d Frankovic a n d Kristian Lenic
Faculty of Engineering University ofRijeka, Vukovarska 58, HR-51000 Rijeka, Croatia, Europe, Phone: ++ 385 51 651 514, 675 801, Fax: ++ 385 51 675 801, E-mail: [email protected] A b s t r a c t - In this paper, a numerical analysis of heat transfer in the latent heat store has been performed. Heat transfer in the latent heat store is the conjugate problem of the phase change of the phase change material and the transient forced convection between the heat transfer fluid and the wall. The differential equations of flow and heat transfer, with initial and boundary conditions, have been discretized by control volume approach and then solved using an iterative procedure. The set of algebraic equations has been solved by FORTRAN software. Velocity and temperature distributions of heat transfer fluid as well as temperature distributions of the wall and PCM inside the latent heat store have been obtained.
1. INTRODUCTION Effective energy management is today one of the most actual problems. Because of that, interest in development of utilizing renewable energy sources has been growing. In this field solar energy has special significance but the radiated solar energy and the energy requirements are oiten at variance with time. To avoid this negative effect of periodic nature of solar energy, heat accumulation in heat storage tanks has been used. Latent heat storage system with solid-liquid phase change has some advantages like high storage density and short temperature interval of heat transfer. Latent heat store analyzed in this paper is shell and tube heat exchanger with the phase change material (PCM) on the shell side. A representative element of the latent heat store is shown in Fig. 1. PCM
2. MATHEMATICAL MODEL OF FLOW AND HEAT TRANSFER IN THE LATENT HEAT STORE
w
HTF
heat transfer is performed primarily by natural convection, and during solidification heat transfer is controlled by heat conduction. Some mathematical models that represent the numerical description of the heat transfer in the latent heat store use enthalpy method to describe heat transfer inside the PCM (Belleci and Conti, 1993) and some other use the temperature transforming model (Cao and Faghri, 1991). In this paper, a numerical analysis of transient phase change heat transfer problem combined with conjugate forced convection, has been performed. Mathematical model that describes flow and heat transfer of the heat transfer fluid, phase change material and wall has been formulated by studying heat transfer phenomena in the control volume of the latent heat store. To describe heat transfer inside the PCM the enthalpy method has been used. Differential equations, with initial and boundary conditions, have been solved numerically by FORTRAN software using an iterative procedure.
-~ o
v
T~
L
Fig. 1. Representative element of the latent heat store A heat transfer fluid (HTF) flows through the inner polypropylene tube and exchanges heat with the PCM on the shell side. During the sunlight, i.e. active phase, hot fluid heats the PCM, the PCM melts and the heat is stored. During the eclipse phase, the PCM solidifies and the stored heat has been delivered to the cold fluid. Heat transfer in such latent heat store is the conjugate problem of the phase change of the PCM and the transient forced convective heat transfer between the heat transfer fluid and the adjacent wall. During melting of the PCM,
The phase change heat transfer and fluid-wall convective heat transfer is an unsteady 2D problem. To establish a convenient mathematical model of flow and heat transfer, the following assumptions have been introduced: the heat transfer fluid is incompressible and viscous heating has been neglected, natural convection in the liquid phase of the PCM has been ignored, thermal losses and conduction through the outer wall of the store have been ignored, flow of the heat transfer fluid is laminar, the initial temperature of the latent heat store is uniform and the PCM is in the solid phase, inlet velocity and inlet temperature of the working fluid are constant. The continuity, momentum and energy equations governing the flow and heat transfer in latent heat store for the heat transfer fluid, the wall and the phase change material are 9
HTF
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ISES Solar World Congress 1999, Volume III
igwx --+
1 0(r-Wr) . . . .
/)x
r
~r
0
(1)
9
boundary conditions
inlet plane c)wx
c)wx + Wx 9
~gr
Owx + Wr 9
~gx
~
x=0
1 /)p = . . . .
p ~x
+
0
i
Wx=Win,
(2) -
+v
"t"> 0
-t- ~ - ~
0x 2
r
Or
Tf=T m
bTw = 0
r~ < r < rw
r-
Wr=0,
Ox
-~-r) t)Tp = 0 rw < r < r~
t~W r
~W r ~Wr b Wx 9 + Wr "--
~r
/)x
Or
1
OP+
p
Or
outlet plane
Wr]
+D-
r
~gx
x=L
(3) 0
2
~)Wx --=0,
i
~x
Ox
=0
OT.
ri < r < r + Wx 9
c)r
9
c)x
+ Wr 9
c)r
[o f+lr'gr'-
(4)
J
--=0
w
Ox
r.
~)x
fluid- wall interface
Wall
0<x
Pw "Cw
r-~
+ ~
(5)
8x 2
r = ri
w x -w
w a l l - PCM interface
--/~'P "
.
~)'t"
~
""~r)
outer wall
Tp7
+
' '3x -'T-
(6)
where H is the enthalpy related to the temperature with eguation T = A- H + B where the coefficients A and B are 1
A = ~ ,
B=0
for
Hp < Ps "Cs "Tin
for
0<
P s "Cs Hp
A=0,
B=T m
--Ps PL
A ~
~
"Cs "Tin
<1
"q
1 , B=T m.ll-ps'cs ] q
PL'CL
PL for
CL
CL
H p - Ps " Cs "Tin
>1
PL "q
The initial and boundary conditions are following: 9 0
initial conditions
0
"t"= 0
O<x
Wx = w r = 0
O<x
Tf - T w "- Tp - Tilfit
0<x
2f ~-''~'w
Or
r = rw
Xw
O r - -
PCM
~
r -0,
Or
O<x
=o
Or
r = ro
br
=0.
3. NUMERICAL SOLUTION OF M A T H E M A T I C A L MODEL
Differential equations, with initial and boundary conditions, have been discretized using control volume method. The resulting algebraic eguations are solved simultaneously using Gauss-Seidel iterative procedure. Due to nonlinearity of the problem iterations are needed during each time step. Convergence criteria is set on 0.01% for all variables of the system. System of discretization equations has been solved using FORTRAN software. As a result of calculation, velocity and temperature distributons of heat transfer fluid and temperature distributions of the wall and PCM inside the latent heat store have been obtained.
4. NUMERICAL RESULTS Described numerical analysis has been applied on the latent heat store with water as the HTF and calcium chloride hexahydrate CaC12 6H20 as PCM. Thermophysical properties of CaC12 6H20 are following: melting temperature Tm= 29.9 ~ latent heat q = 187 kJ/kg
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ISES Solar World Congress 1999, Volume III
thermal conductivity solid phase liquid phase specific heat solid phase liquid phase density solid phase liquid phase
temperature field has not reached fully developed state. Temperature field is changing as the melting interface progresses. Because of small Prandtl numbers, thermal conductivity is large, and heat transfer to the PCM is large. Fig. 2 shows the radial temperature distribution in cross section of the tube x/L = 0.4 for different time periods. Regions of the HTF, wall and the PCM are indicated in diagram. Dimensionless temperature is defined as
2 = 1.09 W/mK 2 = 0.53 W/mK c = 1.4 kJ/kgK c = 2.2 kJ/kgK p = 1710 kg/m 3 p = 1530 kg/m 3
r-/'m~
= ~
,
(7)
T-m The system is initially at the temperature 25 ~ that is less then the melting temperature of the PCM. Inlet velocity of the water is 1 m/s and the inlet temperature is 50 ~ Obtained numerical results show that steady state of the fluid velocity inside the tube has been reached quickly, while the
and dimensionless time is defined as Wm 9T
= ~ .
(8)
Di/2
[]
i.
wall
p=10
........
p=20 p=30
.....
p=40 J
p=70
- - p = 1 5 0 =
. . . . . . . . . . . 7 M. . . . . .
p=200
0.2 '!
v
~
1.o
1.'5
2.0
-r
-v
2.5
3.0
!
3.5
r/(DI2)
Fig. 2. Radial temperature distribution at x/L = 0.4 for different time periods
From the figure can be seen that the PCM temperature in specified cross section of the tube increases with time until the melting temperature is achieved. Then, melting of the PCM starts around the wall surface and the width of melting surface in the specified cross section becomes larger for longer time periods. During melting of the PCM the heat is delivered from the water and stored in the PCM. It can be also seen that heat transfer is high because of small Prandtl numbers of the water. Temperature distributions of heat transfer fluid, wall and the PCM, in analyzed latent heat store, for different time periods are shown in Fig. 3.
It can be seen from the figure that at dimensionless time period ~: =200, melting surface of PCM has reached outer surface of the latent heat store in cross sections near to inlet surface of the heat transfer fluid, while in other cross sections some of the PCM remains unmelted. It can be concluded that heat transfer from the heat transfer fluid to PCM is fast because of large thermal conductivity of the water.
ISES Solar World Congress 1999, Volume III
Fig. 3. Temperature distributions in latent heat store in different time periods
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ISES Solar World Congress 1999, Volume Ill
Fig. 4. Melting surfaces of PCM for different time periods
Melting surfaces of PCM for different time periods are shown in Fig. 4. From the figure it can be seen that melting of the PCM starts on the w a l l - PCM interface at the inlet surface of the heat storage tank. Melting front is then moving along the wall-PCM interface to the outlet surface of the heat store. Because of large thermal conductivity of the water, melting front reaches the outer surface of the tank in a cross sections of the store near the inlet surface before if= 200 i.e. for a relatively short time period.
5. CONCLUSIONS Phase change heat transfer in solar thermal latent heat store has been analyzed in this paper. The results of numerical analysis show that fluid velocity field reaches steady state quickly, while temperature field changes with progressing of the melting interface. Heat transfer from the heat transfer fluid to the PCM is fast due to small Prandtl numbers of the HTF. Because of that melting surface reaches the outer surface of latent heat store in the cross sections near the inlet surface of the HTF after relatively short time. Exact temperature distribution in latent heat store obtained in the analysis is necessary for calculation of thermal efficiency of the heat storage tank. It is the basis for design optimization of the latent heat store.
ISES Solar World Congress 1999, Volume III
NOMENCLATURE c D H L p q
J/kgK m J/m 3 m Pa J/kg
r
m m
T w
K m/s
x
m
specific heat diameter of the tube enthalpy per unit volume length of tube pressure latent heat of the PCM radius of the tube radial coordinate temperature fluid velocity axial coordinate
Greek symbols 2
W/mK
D
m% kg/m 3
P 2"
S
d
thermal conductivity kinematic viscosity dimensionless time density time dimensionless temperature
Subscripts f i in init L m o p r s w x
heat transfer fluid inside radius of the tube inlet initial liquid phase of the PCM melting outer surface of the latent heat store PCM radial coordinate solid phase of the PCM wall axial coordinate
REFERENCES
[1] Bellecci C. and Conti M. (1993). Transient behaviour analysis of a latent heat thermal storage module. Int. J. Heat Mass Transfer 36, 3851-3857. [2] Bellecci C. and Conti M. (1993). Latent heat thermal storage for solar dynamic power generation, J. Solar Energy 51, 169-173. [3] Cao Y. and Faghri A. (1991). Performance characteristics of a thermal energy storage module: a transient PCM/forced convection conjugate analysis. Int. J. Heat Mass Transfer 34, 93-101. [4] Duffle J. A. and Beckman W. A. (1991) Solar Engineering of Thermal Processes, John Wiley & Sons Inc., New York. [5] Esen M., Durmus Aydin and Durmus Ayla (1998). Geometric design of solar-aided latent heat store depending on various parameters and phase change materials, J. Solar Energy 62, 19-28. [6] Hasan A. (1994). Phase change material energy storage system employing palmitic acid, J. Solar Energy 52,143154.
489
[7] Minkowycz W. J., Sparrow E. M., Schneider G. E. and Pletcher R. H. (1988) Handbook of Numerical Heat Transfer, John Wiley & Sons Inc., New York. [8] Patankar S. V. (1980) Numerical Heat Transfer and Fluid Flow, McGraw-Hill, New York. [9] Voller V. and Cross M. (1981). Accurate solutions of moving boundary problems using the enthalpy method. Int. J. Heat Mass Transfer 24, 545-556. [10] Watanabe T. and Kanzawa A. (1995). Second law optimization of a latent heat storage system with PCMs having different melting points. J. Heat Recovery Systems 15, 641-653. [11] Zhang Y. and Faghri A. (1996). Semi-analytical solution of thermal energy storage system with conjugate laminar forced convection. Int. J. Heat Mass Transfer 39, 717-724.
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ISES Solar World Congress 1999, Volume III
M O D E L L I N G OF T W O - L A Y E R S T R A T I F I E D S T O R E S
Jacob van Berkel Entry Technology, Spoorbaanweg 15, 3911 CA Rhenen, The Netherlands tel: +31 (0)317 620470, fax: +31 (0)317 620493, emaih [email protected] C a m i l o C . M . R i n d t a n d A n t o n A. v a n S t e e n h o v e n Eindhoven University of Technology, P.O.Box 513, 5600 MB Eincthoven, The Netherlands tel: § (0)40 2472978, fax: +31 (0)40 2433445, emaih [email protected] A b s t r a c t - A two-layer stratified storage process is studied analytically, numerically and experimentally. Detailed insight into the mixing phenomena is provided by the numerical and experimental model. It appears that the predominant mixing mechanism inside two-layer stores is shear induced entrainment of thermocline fluid. An analytical model is developed, which is especially suited to quickly obtain indicative CTE-performance data under practical operation conditions. With the presented numerical model more accurate F O M - p e r f o r m a n c e data can be obtained. As a demonstration how the numerical code can serve as an optimisation tool, the laboratory store as used in the experiments is optimised with respect to its inlet and outlet height.
1. I N T R O D U C T I O N In 1973, thermal stratification was identified as a means to increase the efficiency of thermal energy storage (Fisher, 1975). Undoubtedly thermal stratification must have been present in stores before, however, its potential to increase the efficiency of, for example, a solar system was not recognised. Due to the separation of hot and cold water inside the store, the solar collector is supplied with relatively cold water. As a result the collector pump switch-on criterion (collector output temperature higher than collector input temperature) is more frequently satisfied which means that the solar system is longer in operation. In addition, hot water supplied by the collector is kept at a relatively high temperature which means that auxiliary heating is less frequently needed (which also results in a higher solar system yield). Many heat and mass transfer phenomena are known to have an adverse effect on thermal stratification. Noteworthy are natural convection flows caused by thermal conduction in tank wall material and inserts (e.g. heat exchanger), see Hermansson (1993). The present paper is focused on distortion of thermal stratification by the inlet and outlet flows which, for short term stores, is the dominating stratification detrimental mechanism. Though in practice the stratification in solar energy stores is continuously variable (due to the variable solar irradiation) here the attention is focused on two-layer hot and cold water stores because of their reduced complexity and higher relevancy with other heat stores (co-generation and chilled water). Like in a previous paper (van Berkel eL al., 1996), the twolayer store is studied in two modes; one in which the thermocline is preserved in the store by on-time flow reversal (multiple-sweep mode) and one in which the thermocline is formed, swept through the store and discharged through the outlet (once-through mode). The first mode is of more practical significance as in this mode strong variable outlet temperatures are avoided, which may otherwise lead to system malperformance. The second mode is of more the-
oretical significance. As this mode only comprises a single stroke, it takes less time to complete the analysis (both experimentally and numerically). Depending on the store mode (multiple-sweep or oncethrough) the performance of a stratified store is expressed in two performance numbers. For the multiple-sweep mode the Cycle Thermal Efficiency (CTE) is used:
rT2'"" (Thot,o,,t-Tcold,,,,)dt C T E --- J 89 -~f~oT.o. (T~o~,~. - T.ozd,~. )dr
(1)
in which the first half period (0 --. 1/2 T) the store is charged with hot water at a temperature Tho~,i,~ and the second half period (1/2 T --* T) the store is discharged. When discharging the store is fed with cold water at a temperature Tco~a,~,,. If the volume of water which has been charged into a store equals the volume which can be extracted, this equation relates the heat contents charged to a store (denominator) to the heat contents which can be extracted from the store (numerator). For the oncethrough mode the Figure of Merit ( F O M ) , see W i l d i n (1990), is applied: FOM
-
r
f[P'" (Thot,out -- Tcold,lf,)df M , to,.~ (Taot,i,,~ - T~old,~,,)
(2)
with Tp~, - M o ~ o , ~ / ~ . At the beTnning of the charging process the store is at a temperature Tsot,i~. In addition to these practical measures the local entropy production rate is used to express the local mixing rate in a store:
,~ = ~(7)
~
(3)
which expresses the production of entropy due to diffusion of thermal energy (~). In a water store entropy production associated with the diffusion of impulse shows to be negligible, see van Berkel (1997). As for short term stores mixing depends on a balance between the stratification promoting force (buoyancy) and
ISES Solar World Congress 1999, Volume III
stratification detrimental force (inertia) the store performance is predominantly expressed as a function of the (overall) Richardson number: Rill =
gApH 2 p 'u~,~
(4)
Alternatively the Richardson number can be based on the momentary position of the thermocline h. Beside the Richardson number the dynamics in the storage vessel is also a function of the inlet Reynolds number defined as: Re, =
~in hin
(5)
/2
In the present research, a two-layer store is studied analytically, numerically and experimentally. Goal of the research is to develop tools with which the design of twolayer stores can be o p t i m i ~ d regarding the above mentioned performance numbers. To this end an analytical model for the multiple-sweep store and a numerical model for the once-through store is developed. The results obtained with these models are compared to experimental results for the laboratory store. Finally an optimisation is carried out. 2. R E S E A R C H
TOOLS
2.1 Analytical model for the multiple.sweep store As pointed out earlier, under practical (multiple-sweep) conditions the store always features a two-layer stratification. As a result the store can be considered onedimensionally which enables development of a relatively simple analytical model. It is assumed that the thermodine oscillates between levels h , , = and H - h=i= (thereby not entering the thermocline exclusion zones with thickness h,~i=), see figure 1. An important assumption is that entrainment is a one-way process. Fluid is transported over the thermocline from the non-turbulent (outlet) layer (for the cold water charge cycle the upper layer, see figure 1) to the turbulent (inlet) layer, not vice versa. This assumption, which is supported experimentally by means of fluid dye colouring tests, implies that during the upward stroke, the temperature of the upper layer does not change. The same applies for the bottom-layer temperature during the downward stroke.
Fig. 1: Cold water charge cycle. Tbot is the bottom-layer temperature at the beginning and at the end of cold water chartp-ug. %or is defined similarly for the upper layer. The entrainment velocity can be coupled to the thermodine velocity by: dh q~ d-t = A- + u~ (6)
491
which states that the true (Lagrangian) thermocline velocity (dh/dt) is the combined effect of fluid displacement q~,,/A and entrainment of fluid over the thermocline u~. The entrainment velocity can be assessed by evaluating the equation of mechanical energy, which provides the link between the cause of entrainment (kinetic energy) and its effect (change of potential energy). In integral form it states that the change rate of kinetic and potential energy within a system is equal to the kinetic and potential energy fluxes over its boundary, the mechanical work acting on the boundary plus the dissipation of mechanical energy within the system (Kundu, 1990). To simplify this equation, the following assumptions are made: 1. Inlet and outlet areas are equally sized and small compared with the tank height. 2. Inlet and outlet velocities are block-profiled, hence u~,~ = Uo,,t = u. Besides it is assumed t h a t no shear stresses are present at the inlet and outlet. Hence, fluid stress is caused by pressure alone. 3. Density variations are accounted for only in the potential energy and pressure terms. 4. Flow separation (from the vertical tank wall) occurs at the inlet. As a result the fluid pressure at the inlet equals the static pressure at t h a t height in the store. 5. Flow separation does not occur in the converging outlet flow. As a result the fluid pressure at the outlet equals the local static pressure inside the store at t h a t height, minus the dynamic pressure head. 6. In first approximation, it is assumed that only a fraction 9/ of the jet kinetic energy influx is converted into change of potential energy. The remaining part (1 - ~ / ) is dissipated into heat and neglected. The flow separation assumptions 4 and 5 are supported by experimental observation. Adopting these assumptions and taking z = 0 as the reference level for potential energy, the equation of mechanical energy for a system as shown in figure 1 (cold water charge cycle) becomes, see van Berkel (1997):
r
p ~t 2
- p) + ~. A ( p - p,o~) - n - ~ T r
(~)
This equation states that the kinetic energy available for mixing (right hand side) is used for thermocline entrainment (second term left hand side) and mixing of the inlet flow with the lower layer contents (first term left hand side). The additional equation is the conservation of mass equation. During the upward stroke, the temperature of the bottom layer may change due to the inflow of cold fluid and entrainment of hot fluid. Assuming a uniform bottom layer temperature, conservation of mass during the propagation of the thermocline from h to h + dh yields (see van Berkel, 1997):
d(h p ) = ( ~ p , . + ~ ~o~)dt
(8)
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In summary, the evolution of the bottom layer density p, the interface ]mini the entrainment velocity u~ for the cold water charge cycle is given by the system of equations 6, 7 and 8. In addition to these equations, the periodic condition holds, which means that at the end of a cycle the temperatures are the same as in the beginning. Besides, assuming that in the upward stroke the temperature of the upper layer doestm~t change, and versa, means t h a t the upper and lower layer temperatures, when the thermocline is/at position = h , ~ , are equal to the corresponding values when the thermocline is at h -- H - h ~ . For solving the coupled set of equations 6, 7 and 8, a numerical procedure is applied. Each time step the relevant parameters (p, h a n d u~) are calculated after which (Euler forward) time stepping proceeds, until the final thermocline level is reached. The accuracy of the solution is checked by back substitution of the solution into the system of equations. E
E E
E
E
E I
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III
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III
"II
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III
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III
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this stage the constant supply of cold water starts to dominate over thermocline entrainment, causing a decrease in bottom-layer temperature. W h e n the thermocline reaches the upper extreme level, the temperature is back at its initial value (thereby satisfying the periodic conditions). During the downward hot water charge stroke, the temperature of the bottom layer does not change and the system returns to its initial conditions from which point the charge process may start again. As a result of the mixing process, the temperature of the lower layer during downward motion is higher than the inlet temperature. Water can not be discharged with the same temperature it has been charged. The CTE is therefore lower than unity. 2.2 QF ~h]ddie omdhrough To gain more insight in the store thermo-hydrodynamics, two store geometries have been studied using a Computational Fluid Dynamics (CFD) model. The most advanced simulations have ]~en done in 3 on the interaction process of the jet, formed by the inlet flow, and the thermocline. The results of these simulations will be presented elsewhere. More practical simulations comprise 2s of a real store (like the one shown in figure 1), the results of which will be presented in this paper. With regard to the simulations, common assumptions are made for the physics involved: Boussinesq approximation, no viscous dissipation of mechanical energy, no radiative heat transfer. The equations to be solved then read :
I
Ou
V-u : 0
(9)
+ V . (u u) = - ~ V p + ~{/~ T -a T ~) + vV2u
(10)
1
0T Fig. 2: Time evolution of (a) bottom-layer temperature, (b) thermocline level and (c) entrainment velocity for cold water charging. The dashed line depicts the level the thermocline would have in case of no entrainment. A = 1 m 2, H ----- 1 m, ~ -- 1/3600.~n1, cold and warm water inlet temperatures 10 ~ and 20 ~ respectivel]~ Ri = 10.6. Quantitative analytical model results will be given later in comparison with experimental results. Here the results are discussed in a qualitative sense. A typical result is presented in figure 2 where the time evolution of the bottom-layer temperature T, the thermoclin&lm&l the entrainment velocity ue are given for the cold water charge cycle. The dimensions of the store and relevant process parameters are given in the caption of the figure. Using the analytical model, it is found that in the initial stage of the cold water charge cycle, when the thermocline is close to the inlet, the entrainment rate is maximal, causing a sharp increase of bottom- layer temperature. Soon after, however, the entrainment velocity levels-off because a larger fraction of the jet kinetic energy available for mixing is required to mix the cold inlet flow with the b o t t o m layer. W h e n the thermocline moves away from the inlet, the entrainment velocity decreases further. In
+ V - ( u T) -- a V 2 T
(11)
As it was anticipated that all fluid motion length scales could be resolved, no turbulence model was incorporated in the numerical program. In essence the solution m e t h o d is a pressure-correction method. It projects an intermediate velocity field u* obtained from the Navier-Stokes equation (excluding pressure), to a divergence-free velocity field u ~+1 by evaluating the pressure gradient term in the separate c o r r e c t ~ j e c t i o n step. In this step, a resulting Poisson equation for pressure is solved directly using a finite difference approximation on an equidistant staggered grid (Schumann and Sweet, l ~ n FJ~mrd~ie equations read: U* -- U n-1
2 At
= - V - ( u u ) " + g ( f l T-T=,,)"+vV2u ~'-1 (12) U n~l
-- U*
2 At
=--Vpn+l
U ~+1 -- U*
- V 2 p ~+I -- V . (
2 At
(13) U*
) -- - - V . ( ~ )
(14)
Tn+ 1 _ T n --- - V - ( u T) ~ + a V 2 T ~ (15) At For the advection and buoyancy components of the momentum equation, time stepping is performed according to the time central, second-order accurate, neutrally stable
ISES Solar World Congress 1999, Volume III
leap-frog scheme. As a time central t r e a t m e n t is unconditionally unstable for diffusion, this term is treated with an Euler forward scheme over 2 At. To attain second- order accuracy for the treatment of buoyancy, this term is evaluated at time level midpoint n. Splitting between odd and even time levels is avoided by means of a time filter. More information on the discretisation of advection and the implementation of the boundary conditions can be found in van Berkel et al. (1996). All simulations are started with a zero velocity field, and a uniform temperature field and step-wise increase of the inlet velocity. The simulations presented in this paper are performed on a 80 • 160 grid, for 10 000 time steps of 0.04 s. This simulation requires about one hour to run on a Pentium 300 MHz computer, which roughly corresponds with 3 . 1 0 -5 seconds per cell, per time step. 2.3 Experiments To provide overall insight in the storage tank flow patterns for representative operating conditions, visualisation experiments are performed in a laboratory size transparent storage tank. Besides, the experimental set-up is also used for validation of the numerical 2D-model for the oncethrough system and the analytical 1D-model for the multiple sweep store. The two-layer store set-up is configured such as to enable multiple sweep and once-through thermocline operation, see figure 3. The size of the actual store volume is (width x depth x height = 400 x 400 • 800 ram). Two inlet/outlet chambers are applied for proper introduction and withdrawal. Water enters the inlet chamber from one side via a tubular inlet distribution manifold to assure an even distribution of the flow. Hot water is prepared and stored in a separate large (150 1) head tank. During testing, hot water is pumped to and from the test tank by means of a centrifugal pump in the hot water line. Without the storage function of the head tank, hot water should be prepared instantaneously, which would require a high power heating rate (50 kW maximum). By occasionally heating the head tank, the hot water contents could be kept at a constant temperature.
Fig. 3. Two-layer stratified store set-up. To avoid usage of a cooling device in the cold water circuit, cold water is drawn directly from the tap. It is fed
493
to the test tank, via the small (50 1) cold water head tank, which serves as an overflow. The overflow imposes a constant pressure in the test tank and thus protects the test tank from high static pressure heads. The flow rate is measured with a calibrated water capacity meter having a measurement accuracy of 0.5- 10 -3 m 3. In addition to the inlet and outlet temperature sensors, the tank contents is monitored by 9 sensors, equidistantly placed at the centerline of the tank. The temperatures are recorded with 0.5 and 1 m m Thermocoax TK105/25/Dtype thermocouples, having estimated response times less than 0.1 s and 1 s, respectively. Tests showed typical values for the random error of 0.05 ~ The temperature data were recorded with an 80386 CPU Personal Computer/data-logger, typically at a sample rate of 10 Hz and either first grabbed into RAM-memory or directly recorded to hard-disk. During the multiple-sweep mode, the thermocline is kept in the tank by on-time flow reversing. The tank mode (cold water charge/hot water charge) is controlled on the basis of temperature signals provided by two thermocouples, placed 114 mm from the tank top and the tank bottom. As soon as the top thermocouples records a temperature lower than the top switch temperature, the flow is reversed manually from cold water charge to hot water charge. Similarly, the hot water charge flow is reversed when the bottom switch thermocouple records a temperature higher than the b o t t o m switch temperature. The accuracy of the charge/discharge times is estimated to be 1 s. All tests are started with a uniform initial temperature distribution in the store. It is found that after 6 cycles the main store parameters (charge/discharge times, thermocline thickness, inlet and outlet temperatures) do not change notably for the subsequent cycles. Flow visualisation is performed mainly by fluorescent dye colouring. A 10- 15 m m thick vertical light sheet is created by illuminating the domain from both sides through slotted masks, attached to the side walls, thereby creating a 2D-cut of the flow field. The domain contents is prepared by adding premixed fluorescent dye in a mass ratio of 1 : 2000. To preserve the dye in the test set-up, it is added to the hot water contents. Visualisation results will be given for cold water charge tests. The colour pattern then visualises the entrainment process in the bottom, transparent turbulent layer. All tests are performed for a flow rate of 0.32- 10 -3 m3s -1, corresponding with an inlet Reynolds number of Re~ = 615. The Richardson number is set via the toplayer temperature. It may be noted that in case of the multiple-sweep mode for Rill = 3.3 and 9.8 the switch temperatures are equal to the mean layer temperature. For higher R/H-numbers, however, this strategy causes partial withdrawal of the thermocline fluid as the thermocline grows thicker than the 114 m m distance between the thermocouples and the tank bottoms. To overcome this problem, for these cases the switch temperatures are set closer to the warm and cold water temperatures.
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3. R E S U L T S 3.1 Flow phenomena for the once-through system A global impression of the flow pattern inside the laboratory two-layer storage tank (for Rill -- 9.8 and Re~ -- 615) is provided by two fluorescent dye snap shots, taken during cold water charging, see figures 4 and 5. In figure 4 the inlet jet enters the store bottom fight. After it has crossed the tank bottom, it collides with the vertical tank wall opposite to the inlet. Then, the jet is deflected upwards and finally collides with the thermocline, at which stage entrainment of top-layer fluid occurs. Marked by the arrows are overturning motions near the thermocline zone, which are likely attributed to the Kelvin-Helmholtz shear-layer instability. A rough estimate obtained from figure 4 yields for the first Kelvin-Helmholtz eddy diameter a value of 3040 m m and for the initial distance between two subsequent eddies a value of 90 mm.
which are drawn into the bottom layer, thereby contributing to the entrainment process. The span wise vortices are most likely a result of the inhomogeneous vertical momentuna of the inlet jet when it collides with the thermocline, causing local return-flow and span wise vortices to occur. Further insight in the stratified store phenomena is provided by the 2D numerical simulation, see figure 6. The local entropy production rate distribution (figure 6a), reveals that mixing (in the thermodynamic sense) mainly takes place in the thermocline. Furthermore, it becomes clear that entrainment mainly takes place by withdrawal of a thermocline fluid filament into the b o t t o m layer. OccasionaUy overturning motions occur (marked by the arrow), quite similarly as observed during the experiments. The vorticity plots (figure 6b) reveal the presence of many whirls in the bottom layer, originating from the inlet wall jet. It furthermore appears that the jet detaches from the wall approximately hMfway the tank bottom. As this effect is not observed during the experiments, it is hypothesised that jet detachment in the numerical results is caused by large-scale sub-thermocline circulation connected to the standing thermocline wave (which is found in the numerical results only). The cause for the numerical artificial standing wave motion is most likely the 2D inverse energy cascaAe for kinetic energy.
Fig. 4. Front view of fluorescent dye image during cold water charging. Top-layer temperature 23 ~ inlet temperature 11 ~ inlet height 20 m m and inlet velocity 0.04 ms -1. The inlet is positioned bottom left. The total domain height is 800 r The thermocline level is 600 ram. Ri~ = 6, Rei = 615, Ri~ = 9.8.
Fig. 5. Side view of fluorescent dye image during cold water charging. Image is taken 10 mm from the right wall. The total domain width is 400 ram. The conditions are the same as outlined in figure 4.
Fig. 6. Calculated flow pattern, halfway the cold water charge process for Rill -- 9.8, (a) entropy production rate (white - 0, b l a ~ --- 1 W m - S K - 1 ) , (b) vorticity (white --2, black - 2 s - l ) .
Thermocline vortices are also found in the side view plane, see figure 5. Fluid filaments (cusps) are visible
In the development of the analytical model it is assumed that entrainment is a one-way process, i.e. fluid is
ISES Solar World Congress 1999, Volume III
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transported over the thermocline from the non-turbulent (outlet) layer to the turbulent (inlet) layer and not vi~ versa. From both the visualisation experiments and the calculations it may be concluded that this assumption holds, at least for the Richardson number and the inlet Reynolds number under consideration.
I
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3.2 Multiple sip ml ysig ~erimatal ndml y~cal mod ~ u l ts The evolution of temperatures inside the store and the inlet and outlet temperatures during a multiple-sweep test are shown in figul~ 7 for Ri = 9.8. Shown is the eleventh cycle which, as analysis of previous cycles revealed, represents stabilised conditions. During the cycle, the thermocline first moves downwards and then upwards. In the narrow bands between the vertical lines, the pump is switched off and the store throughput is zero. A first thing to notice is that (within the 1 s inaccuracy due to imprecise timing of valve operation) the time interval for charging is almost equal to the time interval for discharging (277 and 275 s, respectively). Under the assumption of constant flow rate, this indicates that the water volume charged equals the volume discharged, a finding which was predicted by the I]:lalytical model. visible in figure 7 is that, due to entrainment, warm water which has been charged into the store, is discharged with a slightly lower temperature. Similarly, cold water is discharged with a slightly higher temperature. The difference in charge and discharge temperatures is used to determine the store Cycle Thermal Efficiency for several operation conditions, see figure 8a.
Fig. 7. Oscillating thermocline test for Rill = 9.8. The grey lines represent the internal thermocouples, positioned 114, 171, 228, 342, 399, 456, 570, 627, 684 mm from the bottom. The solid black lines denote the inlet and outlet temperatures.
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I
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I
Fig. 8. Cycle Thermal Efficiency (a) and conversion efficiencyq(b) as a function of the tank Richardson number RiH. The measured values for the CTE can be used to estimate the main parameter in the analytical model which is the efficieafiff conversion of jet kinetic energy flux into change in potential energy (entrainment). When the store geometry and the operation conditions are incorporated in the analytical model, the conve~dan I~tet~ulated for which the analytical model predicts the actually measured CTE, see figure 8b. A tentative explanation for the observed increase of ~/with Rill is that the stronger the effect of buoyancy (higher Rill), the more turbulent motions in and near the thermocline are suppressed, the less jet kinetic energy dissipates and the larger the fraction which remains for entrainment. This explanation is speculative and a better understanding of the functional behaviour of the efficiency remains for further research. On the basis of figure 8, it is cq~cluded that Ri - 15 is an apt design condition. The CTE sharply drops for lower R/H-values whereas it increases just slightly for higher Rl~-values. 3.30nchPouglml ysig ~ , m t a l mt nuneiml mod ~ u l ts With respect to numerical simulation, a choice has to be made with regard to the temporal and spatial resolutions. In figure 9a the convergence of the FOM-error is shown, defined as the difference between the FOM computed for a specif~ ~ the FOM computed with the finest grid, see figure 9a. Despite the anomalous result for the 2.5 mm grid spacing, a general convergence with grid spacing can be recognised. However, despite the apparent convergence in the FOM, no convergence is found in the store dynamics, represented by the store outlet temperature, see figure 9b. Most remarkable is that a fluctuating outlet temperature is found for the numerical simulation (which intensifies for finer grids), while it is absent for the experiment. Closer analysis reveals that the fluctuating outlet temperature signal is caused by the standing wave motion of the thermocline (with a wavelength equal to half the tank width). As the outlet temperature fluctuation is not observed in the corresponding experiments (nor is the thermocline standing wave), the thermocline oscillation is possibly a result Efit~t~e energy cascade. Further insight in the accuracy can be gained by comparison of the store simulation results with store experimental
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results. For a fixed ~eometry ( ~, = ho,,t = 20 mm), constant (11 ~ inlet temperature and various initial storage temperatures (19 ~ 23 ~ 27 ~ and 43 ~ the computed and the experimental FOM values are presented in figure 10. In addition the measured results are given, corrected for mixing in the inlet and outlet chambers (see van Berkel, 1997). A first thing to notice is that the magnitude of the FOM is about 0.02 - 0.03 lower than the measured CTE (see figure 7). This effect may be explained by the large mixing rate encountered during the density current stage (which is only present in the FOM tests).
discrepancy is largest ~ t ~ ~ s . Like the CTEtest, the present FOM-tests indicate optimal store conditions around an overall Richardson number of 10.
0.98 0.96
To.~ II
,.o o.92
II"
0.9
/
0.88
II
0. O86
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10
20
R~H[-]
30
4O
50
||
Fig. 10. M e a ~ ~ t ( o m solid line) and computed (o, solid line) FOMs for the labbratory store ( i~ = 0.02 n~ u ~ = 0.04 m s -1). To account for mixing in the inlet and outlet chambers, see van Berkel (1997), the measured results a r ~ p g r a ~ l p t b ~ solid line).
I I1"
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(b)
--10
15
~5
.~ 2.5 1.25
~',
- - Measui ement 1:?1()0
320
340
Tune[s]
360
380
4O0
Fig. 9. The error in F O M (a) and the outlet temperatures (b), as a function of grid spacing for the laboratory store base case. Initial temperature 23 ~ cold water inlet temperature 11 ~ and flow rate 0.032m as-1 (Riu = 10, Re, = 615). Though the levels of the measured and simulated FOM agree reasonably well, the measurements result in a flatter profile. The discrepancy is ascribed to the artificial thermocline standing wave motion, due to which cold fluid is discharged too early (when the wave crest has passed the outlet). As the wave-like thermocline motion is most intense at small temperature differences, the FOM-
3.4 $~e o#miston It must be stressed here that both the CTE-tests and the FOM-tests outlined in the previous paragraphs are performed for a fixed geometry and variable store temperatures. In practice however, the store temperatures are pre-determiued and a suitable store geometry (in particular the relatise ~ the inlet and outlet) must be chosen. As a demonstration how the numerical code can serve as an optimisation tool, the laboratory store is optimised with respect to its inlet and outlet height (which are assumed to be identical). The optimization routine involves computation of the FOM-value for 8 different inlet and outlet heights while holding the initial store temperature, the inlet temperature and flow rate constant at 23 ~ 11 ~ and 0.32- 10-31~-1 respectively. All computations are carried out with a 80 • 160 grid and 0.04 s time step. The result is shown in figure 11. The FOM simulation results indicate that a maximum for the FOM is attained at an overall Richardson number of 65. A comparison between figure 10 (obtained for variable temperature) and figure 11 (obtained for variable inlet velocity) shows an agreement for small Richardson numbers (when the outlet height is so small that withdrawal of cold fluid does not occur long before the thermodine reaches the tank top). This indicates that under these ~nditions the Ri is a consistent parameter (only its value matters, not whether it has been changed by varying the temperature difference or inlet velocity). For larger outlet openings (larger R / u in figure 11) early withdrawal occurs, which causes the discrepancy with figure 10.
ISES Solar World Congress 1999, Volume III
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discretisation, the present numerical model is restricted to rectangular stores. With respect to the accuracy, the present study has shown that though the dynamics of the store is not fully represented, for optimal laboratory store c o n d i t i o ~ i m u l a t i o n on an 80 • 160 grid is sufficiently accurate. With respect to computational speed, it can be stated that the simulations can be performed on an ordinary Personal Computer (a single run typically then takes in the order of one hour). This makes it possible to use the numerical code as a future engineering tool for optimisation of a stratified thermal store. To summarise the final conclusions of this study.
:-
-
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Fig. 11. Figure of Merit depending on the height of the inlet and outlet, represented as a function of the corresponding PdR (Ti~, = 11 ~ T~,~t = 23 ~ U4,L 9hi, -- 80- 10 -3 ~Tt2s-1). For interpretation of the functional behaviour of the
FOM with the r s l ~ d m i inlet and outlet, it must be reminded that for the once-through store mode, mixing takes place during the initial density current intrusion stage, the thermocline phase and the withdrawal phase. Optimal store conditions are found when the sum of the mixing contributions of the three phases are minimal. For a small inlet and outlet, a high velocity jet is formed which induces a high mixing rate during the thermocline phase. On the other hand, when the inlet and outlet openings are large, a high mixing rate is encountered especially during the withdrawal phase when thermocline fluid is withdrawn long before the thermocline reached the tank top. 4. CQIISC
497
SIONS
It must be noted that in the present study the optimal store conditions are determined for the laboratory store under specific conditions. As a consequence no general design rules (for other stores) can be given. Apart from the question whether or not simple but accurate design rules truly exist, it can be argued that they loose importance as analytical and numerical models may offer a higher accuracy, at acceptable costs. After further validation and development, these models can be applied to predict the performance of new stores. The analytical and numerical models may thereby serve different purposes. The analytical model is especially suited to quickly obtain indicative CTE-performance data. As the model is tentative (its prime parameter, the jet kinetic energy conversion efficiency in generally unknown) the model results are indicative too. A major advantage of the model is that it does not strongly depend upon the store geometry. The numerical model provides more insight and more accurate FOMperformance data. However, due to the type of spatial
9 Store mixing is found to be a two-stage process. Firstly, thermocline entrainment takes place, mainly as a result of shear exerted by the deflected inlet jet. Shear layers form, which, after detachment from the thermocline, may become unstable due to Kelvin-Helmiiiit waves. In span wise direction, cusp entrainment takes place due to vortices which are a result of inhomogeneous jet penetration and back flow. Secondly, once drawn into the bottom layer, stretching and folding takes place. Due to the increased interface area and decreased normal width, ~ i o n of heat effectively mixes thermal energy. Entropy is produced mainly in the thermocline which, as a result of entrainment, is kept thin. 9 The analytical model r ~ u l t s (for -- 1), overestimate the mixing rate and underestimate the CTE. It is concluded ~ l s t increase (roughly proportionally) with the Richardson number. A tentative explanation for the behaviour is that turbulent motions are more suppressed for stronger stratification (higherd~ due to which jet kinetic energy is more efficiently converted into change of potential energy (entrainment). 9 The experimental visualisation results correspond well with the numerical simulation results, though the simulated flow pattern is more dynamical. The experimental and numerical simulation results show a deviation (in terms of FOM) of less than 0.04. The mismatch in store dynamics is ascribed to the inverse cascade for kinetic energy which forms a fundamental d r a w ~ f d a t i o n s . 9 The experimental and numerical analyses show to be complementary. Where the experiment is a true realisation of the thermocline entrainment process, the numerical simulation provides the temporal and spatial resolution for detailed analysis. Moreover the experiments facilitated validation of the numerical as well as the analytical model. The store model results leave some ambiguity with respect to the behaviour of the jet conversion efficiency as a function of Rill. To come to a physical understanding of this behaviour, the conversion efficiency should be determined experimentally with a more accurate model, for a wide range of operation conditions. To eliminate the/~ artificial thermocline wave excitation, short-term thermally stratified store simulations should be performed
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in 3D. The computational work can be kept within limits by considering only a part of the tank. For the laboratory store, this would typically require a grid of width • height • depth = 80 • 160 • 20 cells, which can be solved within hours on a Gigailop (super) computer. In addition, the numerical code should be made fit to cope with complex geometries, possibly by the use of boundary fitted co-ordinates. With the numerical code, the store performance can be evaluated for many parameter settings (tank aspect ratio, inlet and outlet height). From the results, more insight can be gained in the functional behaviour of the store Figure Of Merit on parameters like the overall Richardson number and the inlet Reynolds number.
REFERENCES Fischer, L.S., van Koppen C.J.W. and Mennink M.D. (1975) The thermodynamics and some practical aspects of thermally stratified heat storage in water. Report WPS3.75.11.R247, Eindhoven University of Technology, The Netherlands. Hermansson R (1993) Short Term Water Heat Storage, an experimental and numerical investigation of phenomena that affect the degree of thermal stratification. Doctoral-
theses, 127D, Lulea University of Technology, Sweden. Kundu P.F. (1990) Fluid Mechanics. London.
Academic Press,
Schumann U and Sweet It. (1976) A direct method for the solution of Poisson's equation with Neumann boundary conditions on a staggered grid of arbitrary size, J. of Computational Physics 20, pp. 17- 182. Van Berkel J, van Steenhoven A.A. and Rindt C.C.M. (1996) Thermocline entrainment in stratified energy stores. In Proceedings of the Eurotherm seminar # ~9 : Physical models for thermal energy stores, van Steenhoven A.A. and van Helden W.G.J. (Eds), pp. 63-73, Eindhoven, The Netherlands. Van Berkel J. (1997) Thermocline entrainment is Stratified Energy Stores. PhD-thesis, Eindhoven University of Technology, The Netherlands. Wildin M.W (1990) Diffuser Design for Naturally Stratified Thermal Storage. ASHRAE transactions A T 90-13-~, pp. 1094-1102.
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Full Table of Contents
Keynotes Financing of Private Renewable Energy Projects; Hurdles and Opportunities L. Y. Bronicki .............................................................................................................................................................................................. I-3 Solar Energy in the Built Environment: the Building as a System Plus the Systems in the Building E. de Oliveira Femandes ........................................................................................................................................................................... I-5 Angular Selectivity of Seasonal Sun Protection Devices A. Goetzberger, H. Wirth ........................................................................................................................................................................... I-9 Recent Developments in Photocatalytic Detoxification and Disinfection Processes of Water and Air D. Y. Goswami ......................................................................................................................................................................................... 1-16 solar - Powered Systems for Cooling, Dehumidification and Air - Conditioning G. Grossman ............................................................................................................................................................................................ 1-21 Regulatory and Institutional Measures by the State to Enhance the Deployment of Renewable Energies - the German Experience P.-G. Gutermuth ...................................................................................................................................................................................... 1-29 Building Integration of a solar Energy Systems A. G. Hesl~es ........................................................................................................................................................................................... 1-36 Packaged Solar Water Heating Technology,Twenty Years of Progress G. L. Morrison, B. D. Wood ..................................................................................................................................................................... 1-42 Design Tools for Bio-Climatic and Passive Solar Buildings E. Shaviv .................................................................................................................................................................................................. 1-53 The Solar Chemistry Program of the International Energy Agency's Implementing Agreement Solarpaces A. Steinfeld, V. Anikeev, J. Blanco, M. Epstein, K. -H. Funken, J. L~cld, A. Lussi, J. Murray ..............................................................1-64 High Temperature Solar Energy Conversion Systems A. Yogev, U. Fisher, A. Erez, J. Blackmon ............................................................................................................................................. 1-71 "Energy Towers" Producing Electricity and Desalinated Water Without a Collector D. Zaslavsky ............................................................................................................................................................................................ 1-79
Photovoltaic Cells and Modules ITO/InP Photovoltaic Devices H. Aharoni ................................................................................................................................................................................................ 1-95 Spray-Deposited SnO2-nSi Solar Cells E. Bobeico .............................................................................................................................................................................................. 1-109 Simulation and Test of Peltier Elements in Connection with Photovoltaic Cells B. Cancino, P. Roth, A. A. AIvarado ..................................................................................................................................................... 1-113 Preparation and Characterization of Sb-Se Thin Films by Electrodeposited Technique for Photovoltaic Application A. M. Femandez, M. G. Merino ............................................................................................................................................................. 1-120 Investigation of the Back Contact of Cadmium Telluride Solar Cells R. Gottschalg, B. Elsworl~, D. G. Infield, M. J. Keamy ........................................................................................................................ 1-124 Comprehensive Approach for the Estimation of Outdoor Performance of Amorphous Silicon Photovoltaic Devices R. Gottschalg, G. Perentzis, D. G. Infield, g. J. Kearny ...................................................................................................................... 1-129 Heat Flow Analysis in Solar Cell Modules
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S. Martin, D. R. Harris, W. Y. Saman .................................................................................................................................................... 1-134 Electrodeposited CdTe Based Photovoltaic Structures on Metallic Substrates X. Mathew, P. J. Sebastian, J. Pantoja, A. P. Grifd, M. E. Calixto, J. C. McClure, V. P. Singh ......................................................... 1-142 Experimental Investigation of Transient Processes and Developing of Equivalent Diagram of a Solar Cell Panel M. A. Slonim, A. A. Slonim .................................................................................................................................................................... 1-149 Development of Single Junction Cell Amorphous Silicon Solar Photovoltaic Modules with Improved Resistance to Degradation N. B. Udi ................................................................................................................................................................................................. 1-154 Study of Thin Film Photovoltaic Cells of CdS/CdTe and CdS/Cu_xS P. Yianioulis, D. Patrikios ...................................................................................................................................................................... 1-160 24,7% Efficient Perl Silicon Solar Cells and Other High Efficiency Solar Cell and Module Research at the University of New South Wales J. Zhao, A. Wang, M. A. Green ............................................................................................................................................................. 1-165
Photovoltaic Electricity and Systems Aeration of Fish-Ponds by Photovoltaic Power J. Appelbaum, D. Mozes, I. Segal, M. Barak, M. Reuss, P. Roth........................................................................................................ 1-175 Solar-Photo-Voltaic/ThermaI-Cogeneration Collector B. J. Huang, T. H. Lin, W. C. Hung, F. S. Sun ...................................................................................................................................... 1-181 Solar Energy Resources and Their Application Perspectives in Georgia (Using Semiconductive Photovoltaic Cells) N. P. Kekelidze, T. V. Jakhutashvili, E. G. Chachkhiani, G. E. Chachkhiani ...................................................................................... 1-185 Applications of Dispersed Generation Systems in the Utility Network K. Y. Khouzam ....................................................................................................................................................................................... 1-192 Design of Grid-Connected Inverters L. Lori, P. Redi, M. Ruzinsky ................................................................................................................................................................. 1-200 Leveraging the Value of Photovoltaics in Urban Areas through Their Use in Traffic; Lighting and Exterior Shelter D. A. Nezer ............................................................................................................................................................................................ PV System Connected to a Grid for Home Applications S. Panyakeow, S. Sopitpan ................................................................................................................................................................... On the Performance of Nine-Year-Old Solar Home Systems and Street Light Systems in Sukatani Village in Indonesia A. Reinders, A. S. Pramusito ................................................................................................................................................................ Demonstrating the Superior Performance of Thin-Film Amorphous Silicon for Building-Integrated Photovoltaic Systems in Warm Climates R. R0ther ................................................................................................................................................................................................ Performance of Grid-Connected Photovoltaic Plants F. Scapino, A. Abete, L. Ferrads, F. Spertino ....................................................................................................................................... Study on Islanding of Dispersed Photovoltaic Power Systems Connected to Utility Power Grid O. Tsukamoto, T. Okayasu, K. Yamagishi ........................................................................................................................................... Cies Islands Stand-Alone Photovoltaic Plant: Evaluation and First Results M. Vazquez ............................................................................................................................................................................................
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Solar Thermal Electricity An Overview and Operation Optimization of the Kramer Junction solar Electric Generating Systems R. G. Cable, S. D. Frier ......................................................................................................................................................................... 1-241 Controller Design for Injection Mode Driven Direct Solar Steam Generating Parabolic Trough Collectors M. Eck, M. Eberl .................................................................................................................................................................................... 1-247 A Multistage Solar Receivers: The Route to High Temperature A. Kdbus, P. Doron, R. Rubin, J. Kami, R. Reuven, S. Duchan, E. Taragan ...................................................................................... 1-258 The TROF (Tower Reflector with Optical Fibers): a New Degree of Freedom for Solar Energy Systems A. Kribus, O. Zik, J. Karni ...................................................................................................................................................................... 1-266 Transition Strategies for Solar Thermal Power Generation D. R. Mills, C. J. Dey ............................................................................................................................................................................. 1-272 TRNSYS Software Application for Solar Thermal Power Plants Simulation and Comparative Analysis O. Popel, S. Frid, E. Shpilrain, R. Pitz-Paal, K. Hennecke .................................................................................................................. 1-280 Distributed Power From Solar Tower Systems: A Mius Approach M. Romero, M. J. Marcos, F. M. T~llez, M. Blanco, V. Femandez, F. Baonza, S. Berger ................................................................. 1-286 Simulation of Dynamic Behaviour of a Solar Reactor-Receiver as a Function of Solar Concentrated Radiation Profile H. Romero-Paredes, E. Torijano, A. V~.quez, A. Tortes, J. J. Ambriz ............................................................................................... 1-296 Potential Efficiencies of Solar-Operated Gas Turbine and Combined Cycle, Using the Reflective Tower Optics A. Segal, M. Epstein .............................................................................................................................................................................. 1-302 Project Diso (Direct Solar) Update on Project Status and Future Planning E. Zarza, K. Hennecke, O. Goebel ....................................................................................................................................................... 1-307
Wind Power Systems, Solar-Wind Hybrids and Electrical Storage Lessons Learned From the Xcalak Village Hybrid System: A Seven Year Retrospective R. E. Foster, R. C. Orozco, A.-R.-P. Rubio .......................................................................................................................................... The Analysis of Wind Data and Wind Energy Potential in Bandirma, Turkey D. Inan, C. Dundar ................................................................................................................................................................................. Phoebus - an Autonomous Supply System with Renewable Energy C. Meurer, H. Barthels, W. A. Brocke, B. Emonts, H. G. Groehn ........................................................................................................ Optimization of the Combination of Power Units in Isolated Grids F. Sp~te, J. Plettner-Marliani, U. Mades, J. Tzschoppe, H.-J. Haubrich .............................................................................................
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Solar Radiation Measurement and Analysis Interannual Variability of Meteorological Parameters in Temperate Climates R. Aguiar, J. Boland ............................................................................................................................................................................... 1-353 Distribution of Solar Irradiation in Brazil Derived From Geostationary Satellite Data S. Colle, S. L. de Abreu, P. Couto, S. Mantelli, E. B. Pereira, E. Raschke, R. Stuhlmann ................................................................. 1-362 Determination of the Spectral Emittance in the Visible Range at High Temperatures Supported by Laser Heating S. Eckhoff, I. Alxneit, M. Schubnell, H.-R. Tschudi .............................................................................................................................. 1-372 Distribution of the Ultraviolet solar Radiation in the Sky of San Luis (Argentina) A. J. Fasulo, M. T. Deluigi, E. Crino ...................................................................................................................................................... 1-376
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Calculation of Solar Radiation on Inclined Surfaces in Turkey M. GQnes................................................................................................................................................................................................ 1-380 Estimation of Direct Solar Irradiance From Global Irradiance by Means of Signal (Wavelet) Analysis M. Higashi, S. Rukugawa ...................................................................................................................................................................... 1-386 A Comparison of Spectral Total Atmospheric Transmission Between Summer and Winter in Athens, Greece H. D. Kambezidis, A. D. Adamopoulos, D. Zevgolis ............................................................................................................................ 1-396 Variability of Atmospheric Turbidity in Athens, Greece H. D. Kambezidis, A. K. Fotiadi, B. D. Katsoulis .................................................................................................................................. 1-400 The Meteorological Radiation Model H. D. Kambezidis, A. D. Adamopoulos, N. K. Sakellariou, H. G. Pavlopoulos, R. Aguiar, J. Bilbao, A. de Miguel, E. Negro ...................................................................................................................................................................................................... 1-406 Short-Term Forecasting of Solar Radiation Based on Satellite Data - an Application of Neural Networks and Markov Random Fields E. Lorenz, A. Hammer, D. Heinemann, B. L0ckehe ............................................................................................................................. 1-411 Characterization and Inter-comparison of the Global and Beam Radiation Measured at Three Sites in the Southern Region of Israel by Statistical Analysis V. Lyubansky, A. lanetz, I. Seter, A. I. Kudish, E. G. Evseev .............................................................................................................. The Helioclim Project: From Satellite Images to Solar Radiation Maps C. Rigollier, L. Wald ............................................................................................................................................................................... A Climatological Database of the Linke Turbidity Factor C. Rigollier, L. Wald, J. Angles, L. M6nard, O. Bauer .......................................................................................................................... Comparison of Several Parameterized Models for Global Insolation under Cloudy Skies M. A. Rubio, F. J. Batlles, G. Lopez, L. Alados-Arboledas .................................................................................................................. Solar Radiation Modelling in a Complex Enclosure A. Trombe, L. Serres, M. Moisson ........................................................................................................................................................
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Solar Collector Optical Materials Experimental Evaluation of Selective Surfaces in a High Vacuum W. S. Duff, D. Hodgson ......................................................................................................................................................................... 1-451 Pearl Luster Pigments as Overheating Protection in Transparently Insulated Solar Facades O. F. Gross, A. Beck, S. Weismann, J. Fricke, E. Steudel, C. Schank ............................................................................................... 1-453 A New Heat Reflective Polycarbonate Sheet with Spectral Selectivity G. Hakim ................................................................................................................................................................................................ 1-461 Diffusive Properties of Dry and Wet Glass and Plastics J. G. Pieters, I. V. Pollet ........................................................................................................................................................................ 1-462 Solar Radiation Transmittances of Dry and Wet Plastic Films J. G. Pieters, I. V. Pollet ........................................................................................................................................................................ 1-470 Double-Tailored Microstmcturss H. Ries, J. Muschaweck ........................................................................................................................................................................ 1-477 Temperature Dependence of Thermal Conductivity of Advanced Insulators K. Tajiri, T. Nishio, S. Tanemura ........................................................................................................................................................... 1-482 Optical Properties and Radiative Cooling Power of White Paints S. Tanemura, M. Tazawa, P. Jing, T. Miki, K. Yoshimura, K. Igarashi, M. Ohishi, K. Shimono, M. Adachi ...................................... 1.485 Solar Optical and Infrared Radiative Properties of Transparent Polymer Films G. Wallner, H. Schobermayr, R. W. Lang, W. J. Platzer...................................................................................................................... 1-489 Systematic Study of Electrochromic Devices for Optical Applications P. Yianoulis, S. Papaefthimiou, G. Lefthedotis ..................................................................................................................................... 1-495 Simulation of a Test Cell Dynamic Behavior for the Evaluation of Glazing Thermal Properties P. Yianoulis, G. Lefthedotis ................................................................................................................................................................... 1-504
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Hydrogen, Chemical Energy Storage and Fuels Theoretical Analysis and Experimental Results of a 1KW Chemsynthesis Reactor for a Solar Thermochemical Energy Storage System H. Kreetz, K. Lovegrove ........................................................................................................................................................................ A Solar Driven Ammonia Based Thermochemical Energy Storage System K. Lovegrove, A. Luzzi, H. Kreetz ......................................................................................................................................................... Solar Production of Aluminum by Direct Reduction of Ore to AI-Si Alloy J. P. Murray ............................................................................................................................................................................................ The Production of Zinc by Thermal Dissociation of Zinc Oxide - Solar Chemical Reactor Design A. Steinfeld, P. Haueter, S. Moeller, R. Palumbo ................................................................................................................................. Solar-Assisted Syngas-Driven Power System C. Sugarmen, M. Epstein, I. Spiewak, U. Fisher, R. Tamme, R. Buck ................................................................................................ Heat Recovery Experiments with Concentration Gradient Catalyst Layer in a Solar Chemical Heat Pump T. Takashima, T. Doi, Y. Ando, T. Tanaka ...........................................................................................................................................
1-515 1-523 1-531 1-539 1-544 1-549
Biomass Energy Conversion Two-Stage Gasification of Wood with Preheated Air Supply: A Promising Technique for Producing Gas of Low Tar Content S. C. Bhattacharya, A. Durra ................................................................................................................................................................. 1-557 Anaerobic Digestion System Installation of Cattle Manure in Two Farms in Puebla, Mexico F. Munoz, L. Lopez ................................................................................................................................................................................ 1-562 Municipal Solid Waste Evaluation as a Source of Energy in Mexico City F. Munoz, M. Arciga ............................................................................................................................................................................... 1-566
Solar Energy Systems for Buildings and Solar Architecture Bioclimatic Designs for the New University of Cyprus Campus. 1st Competition: Facilities for Science and Technology D. W. Aitken, A. Kyprianou ....................................................................................................................................................................... Numerical Model of a Building with Transparent Insulation A. K. Athienitis, H. Ramadan .................................................................................................................................................................. Roof Integrated Heating and Cooling System M. Belusko, W. Saman ........................................................................................................................................................................... The Analytic Solution of the Differential Equations Describing Heat Flow in Houses J. Boland ................................................................................................................................................................................................. On the Use of the Solar Collection Envelope for Determining the Building Shape I. G. Capeluto .......................................................................................................................................................................................... Elaboration on the Design and Operation Principles of a Heavy Duty Universal Sunlight Heliodon Assembled From Precision Machining Tools K. P. Cheung, S. L. Chung ..................................................................................................................................................................... Solar Architecture in Turkey: State-of-the-Art F. N. Demirbilek, D. I. Eryildiz ................................................................................................................................................................ Windows in the Attic: Thermophysical Problems of Inclined Windows A. Donath ................................................................................................................................................................................................
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Thermal Performance of Building Integrated Ventilated PV Facades U. Eicker, V. Fux, D. Infield, L. Mei, K. Vollmer ..................................................................................................................................... 11-55 Building Design in Tropical Climates Elaboration of the ECODOM Standard in the French Tropical Islands F. Garde, H. Boyer, R. Celaire, L. Seauve ............................................................................................................................................ 11-59 Passive Cooling System for Remote Locations S. Grignaffini, G. Galli, F. Gugliermetti ................................................................................................................................................... 11-66 The Israeli Insulation Standard for Offices S. Hassid, D. Feuermann, A. Roitgur, D. Sergovich ............................................................................................................................. 11-71 B.A.M.A. (Energy Conserving Buildings) Project: Passive Solar Energy in Popular Residential Apartment Buldings in Israel S. Hassid, M. Poreh, D. Wegner ............................................................................................................................................................ 11-75 Sustainability and the Use of Solar Energy: Life Cycle Analyses of a Norwegian Solar Dwelling A. G. Hestnes, B. N. Winther .................................................................................................................................................................. 11-78 Early Results on the Effectiveness of Natural Ventilation at Verbena Height - a High Rise, High Density Housing Development in Hong Kong H.-M. Ho, Z.-Y. Liao, M.-S. Tse, R. Tsang ......................................................................................................................................... 11-83 Combined Photovoltaic and Solar Thermal Systems for Facade Integration and Building Insulation S. Krauter, G. Aradjo, S. Schroer, M. J. Salhi, C. Triebel, R. Lemoine, R. Hanitsch ........................................................................... 11-88 An Imaginative Environment- Responsive Laboratory Building in the Harsh Climate of Botswana E. S Leus, B. Marland ............................................................................................................................................................................. 11-92 The New Headquarters for Botswana Technology Centre: Innovative Technologies in a Hot- Dry Southem African Climate E. S. Leus, B. Marland ............................................................................................................................................................................ 11-97 Encapsulated Venetian Blind: A New Numerical Model L. Mazzarella, M. Motta ........................................................................................................................................................................ 11-101 Understanding the Potential of Ventilated PV Facades L. Mei, D. G. Infield, U. Eicker, V. Fux ................................................................................................................................................. 11-110 Modelling Solar Energy Input in Greenhouses J. G. Pieters, J. M. Deltour ................................................................................................................................................................... 11-117 Comparative Assessment of the Thermal Behavior of a Planted Roof vs. a Bare Roof in Thsssaloniki F. A. Psomas, E. A. Eumorfopoulou, N. K. Tsakiris ............................................................................................................................ 11-126 Facade Integrated Solar Collectors G. Rockendorf, S. Janssen .................................................................................................................................................................. 11-134 Bioclimatic Designs for the New University of Cyprus Campus. 2nd Competition: Face A, Student Housing D. Serghides, C. Chrysanthou, E. Papachristou ................................................................................................................................. 11-141 Photovoltaica (PV) Modelling for Cities: a GIS-Building Integrated PV (BIPV) Simulations Approach M. Snow, P. Jones, D. K. Prasacl......................................................................................................................................................... 11-147 The Solar -Campus JQIich - Actual Status F. Slate, M. Melil3, K. Backes .............................................................................................................................................................. 11-156 Research and Development on the First AC BIPV Installation in Canada L. Stamenic, E. Smiley, K. Colbow, J. Jones ....................................................................................................................................... 11-165 The Influence of a Planted Roof on the Passive Cooling of Buildings T. G. Theodosiou, D. G. Aravantinos, D. G. Tourtoura ....................................................................................................................... 11-169 Modelling the Thermal Effects of Semitransparent PV - Modules W. G. J. van Helden, B. J. de Boer, J. L. Balenzategui ....................................................................................................................... 11-178 Passive Solar Office Building: Results of the First Heating Period R. Wagner, A. Spieler, K. Vajen, S. Beisel .......................................................................................................................................... 11-183 PASYS- a Knowledge Based CAD System for Determining the Passive Systems for Heating and Cooling A. Yezioro .............................................................................................................................................................................................. 11-189
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Daylighting Heat Transfer through a Duovent Glass with Chemically Deposited Solar Control Coating G. Alvarez, J. J. Flores, C. Cortina ...................................................................................................................................................... A Novel Ventilated Reversible Glazing System E. Erell, Y. Etzion .................................................................................................................................................................................. Comparative Analysis of Daylighting Systems Investigating Illumination and Structure L. I. Filetoth ........................................................................................................................................................................................... Numerical Simulation and Scale Model Measurements of Daylighting Systems in an Existent Building F. Fillipetti, M. Paroncini, B. Calcagni .................................................................................................................................................. Sun Protection System Based on CPC's with Total Internal Reflection A. Goetzberger, C. B0hler, H. Wirth ..................................................................................................................................................... Energy Savings Related with the Natural and Artificial Light in the Underground Car Parking Areas S. Grignaffini, F. Gugliermetti ............................................................................................................................................................... Comparison of Experimental Measurements and Numerical Simulation in an Atrium Building P. Zazzini, M. Paroncini, B. Calcagni, A. Manni ..................................................................................................................................
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Education and Information Exchange Use of the Electronic Book "Building Thermal Analysis" in Passive Solar Design and Education A. K. Athienitis ....................................................................................................................................................................................... Multimedia Library of Renewable Energies M. Castro, A. Colmenar, A. Vara, J. A. Rodriguez, J. Carpio, J. Peire............................................................................................... Photovoltaic Water Pumping Systems Installer Training: a Partnership Experience Between the University and Sao Francisco Hydroelectric Power Plant E. M. de Souza Barbosa, C. J. C. Salviano, A. M. Carvalho, M. F. Lyra............................................................................................ The Centre for the Application of Renewable Energies (C.A.R.E.) D. Herold, M. Hoffmann, V. Horstmann, A. Neskakis, J. Plettner-Marliani ........................................................................................ Developing a Web-Based Learning Environment for Building Energy Efficiency and Solar Design in Hong Kong S. C. M. Hui, K. P. Cheung ................................................................................................................................................................... Integration of Communication and Development in the "Alta Valle Di Susa" Project for Solar Energy F. Jarach, G. Del Tin ............................................................................................................................................................................ Set-Up of a Laboratory for Research and Education in Solar Energy in Rio De Janeiro S. Krauter, R. Stephan, L. Batos, R. Hanitsch ..................................................................................................................................... A Web Based Course for Learning Solar Thermal Processes S Kumar, R. A. Attalage ...................................................................................................................................................................... The School Physics Program of the Finnish Physical Society A. Lampinen, R. Hemberg, M. A. Paalanen ........................................................................................................................................ Using the World Wide Web for Tertiary Level Renewable Energy Education - the Potential, the Practice and the Possible Problems C. Lund, P. Jennings ............................................................................................................................................................................ Raps in a Virtual World - a Web Based Remote Area Power Supply System C. Lund, N. Wilmot, T. Pryor, G. Cole .................................................................................................................................................. "Solar Matters". A Comprehensive School Unit K. G. Sheinkopf, B. M. Sheinkopf ........................................................................................................................................................ A Solar Energy Education Program K. G. Sheinkopf .....................................................................................................................................................................................
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Environmental and Social Impacts of Energy Systems Life Cycle Assessments of Solar Collectors in Denmark T. D. Jacobsen, H. Wenzel .............................................................................................................................................. ..................... 11-333 CO2-Mitigation by Solar Conversion of Hydrocarbons J. Ortner, K.-H. Funken, F. Ploetz ....................................................................................................................................................... 11.340 Cumulative Energy Demand of Wind Energy and Solar Water Heating Systems H. -J. Wagner, D. G0rzenich, E. Pick. .................................................................................................................................................. 11-345 Socio-Economic Impact Assessment of Solar vs. Grid-Electrified Rural Households in Namibia N. Wamukonya ..................................................................................................................................................................................... 11-351
Indirect Solar Resource Evaluations A Method for Establishing a Solar Power Network for Emergency Integrated Cost Effectively in a CliP (Combined Heat and Power) Network G. Georgiades, T. Chikahisa, Y. Hishinuma ........................................................................................................................................ 11.361 Daylight and Solar Irradiance Data Derived From Satellite Observations - the Satellight Project D. Heinemann, A. Hammer, A. Westerhellweg, H. G. Bayer, C. Reise .............................................................................................. 11.368 The Use of Solar Energy: Considerations for Calculations of Greenhouse Gas Reduction by Photovoltaics S. Krauter .............................................................................................................................................................................................. 11-375
Policy and Programs Brundtland Solar City Network T. Esbensen .......................................................................................................................................................................................... 11-379 Hungarian UNESCO Solar Participation Program L. Imre, I. Farkas ................................................................................................................................................................................... 11-382 Solar Energy in Social Housing in the UK R. Oldach, A. Wheldon, A. Francis ...................................................................................................................................................... 11-386 Barriers for Introducing Photovoltaics in Central Europe: Case of Poland S. M. Pietruszko .................................................................................................................................................................................... Renewable Energy Sources Utilization in Russia O. S. Popel, E. Shpilrain, S. Frid, V. Dobrokhotov, N. Koshkin .......................................................................................................... The New European Solar Radation Atlas: a Tool for Designers, Engineers and Architects K. Scharmer, B. Bourges ...................................................................................................................................................................... The Kyoto Mechanisms and the Prospect for Renewable Energy Technologies N. Wohlgemuth, F. Missfeldt ................................................................................................................................................................ Mechanisms to Support Alternative Technologies for Electricity Generation in the European Union N. Wohlgemuth ..................................................................................................................................................................................... Employment Effects of Greenhouse Gas Reduction Strategies A. Ziegelrnann, M. Mohr, H. Unger ......................................................................................................................................................
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Solar Means Business Compound Parabolic Concentrator Technology Development to Commercial Solar Detoxification Applications J. Blanco, S. Malato, P. Fernandez, A. Vidal, A. Morales, P. Trincado, J. C. Oliveira, C. Minero, M. Musci, C. Casalle, M. Brunotte, S. Tratzky, N. Dischinger, K. H. Funken, C. SaWer, M. Vincent, M. Collares-Pereira, J. F. Mendes, C. M. Rangel .......................................................................................................................................................................... Marketing and Selling Solar Energy Equipment T. Book .................................................................................................................................................................................................. An Enabling Technology Opens the Way to Large Scale Use of Solar Energy L. Y. Bronicki ......................................................................................................................................................................................... Transportation of Electricity Production L. Y. Bronicki, I. Dostrovsky, U. Fisher ................................................................................................................................................ The Market for Solar Energy in the Caribbean I. Haraksingh ......................................................................................................................................................................................... A Guide for Financial Feasible Large-Scale Solar Thermal IPP's R. Kistner, M. Geyer, R. Hanitsch, H. W. Price ................................................................................................................................... Solar Energy: Time to Get Commercial A. Mor, U. Halperson ............................................................................................................................................................................ The Design and Development of a Suitable Universal Means of Terminating, Interconnecting and Packaging Photovoltaic Panels for Present and Future Applications P Timbrell, M. Instance ........................................................................................................................................................................ Strategic Analysis of the Integration of a Biomass Power Plant in Spain M. Varela, Y. Lechbn, R. Sdez .............................................................................................................................................................
11-427 11-437 11-444 11-447 11-451 11-455 11-463
11-468 11-472
Developing Countries Solar Energy - a True Option for Rural Electrification in Kenya Hindered by Unfavourable Policy M. M. Agumba ....................................................................................................................................................................................... 11-481 A Study on Improved Institutional Biomass Stoves S. C. Bhattacharya, A. H. Md. M. R. Siddique, M. A. Leon, H.-L. Pham, C. P. Mahandari .............................................................. 11-484 Dissemination of Renewable Energy in Developing Countries: Experiences of a Regional Project in Asia S. C. Bhattacharya, S. Kumar .............................................................................................................................................................. 11-489 Using Photovoltaica for Agricultural Processing Activities in Upper Mustang (Nepal) D. Blamont, P. Amado .......................................................................................................................................................................... 11-495 Social-Technical Assessment of Photovoltaic Systems Installed in the First Region of Chile B. Cancino, P. Roth, E. Galvez, A. Bonneschky ................................................................................................................................. 11-503 Evacuated Tubular Collector Water Pasteurization Systems W. S. Duff, D. Hodgson ........................................................................................................................................................................ 11-509 Hybrid System Heat Pump - Solar Air Heater for the Drying of Agricultural Products W. Soto Gornez, H. D. Arias-Varela, P. Melin, J. A. Ortega-Herrera, R. Best-Brown ........................................................................ 11-512
Special Topics Introduction of Dish Stirling Systems in Morocco. Project Proposal for a Moroccan - German Co-Operation V. H&ussermann, W. Schiel ................................................................................................................................................................. 11-517 Small Scale Photovoltaic R. O. Desalination - Experience in Gran Canaria D. Herold, V. Horsvnann, A. Neskakis, J. Plettner-Marliani, R. Calero, G. Piemavieja ..................................................................... 11-527
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Flat Plate and Non-Concentrating Solar Collectors Hybrid Solar Collectors for Microclimate Forming System G. J. Besler, D. Kwiecien ......................................................................................................................................................................... 111-3 Testing of a Flat Plate Collector with Selective and Nonselective Absorbers That Are Otherwise Identical W. S. Duff, D. Hodgson ........................................................................................................................................................................... 111-4 Comparison Between a Simple Solar Collector Accumulator and a Conventional Accumulator A. J. FasuIo, J. Follari ............................................................................................................................................................................ II1-11 Solar Air Collectors - Investigations on Several Series-Produced Collectors H. Fechner, O. Bucek ............................................................................................................................................................................ 111-17 An Empirical Heat Transfer Equation for the Transpired Solar Collectors, Including No-Wind Conditions K. G. T. Hollands, G. W. E. van Decker. ............................................................................................................................................... 111-23 A CFD Heat Transfer Analyses of the Transpired Solar Collector under No-Wind Conditions K. G. T. Hollands, S. J. Arulanandam, E. Brundrett ............................................................................................................................. 111-29 Analysis of Thermal Performance on an Air-Type Solar Collector with 2- Glass Using Carbon Fiber Sheet as Collecting Material X.-m. Jiang, H. Baba, K. Kanayarna, N. Endoh ................................................................................................................................... 111-35 Research and Development of Solar Collectors Fabricated From Polymeric Material A. I. Kudish, E. G. Evseev, M. Romrnel, M. K6hl, G. Walter, T. Leukefeld .........................................................................................111-40 Study of a Mixed (Water Or Air) Solar Collector S. Lalot ................................................................................................................................................................................................... 111-50 Uncertainty in Solar Collector Testing Results E. Mathioulakis, K. Voropoulos, V. Belessiotis ..................................................................................................................................... 111-56 Optimized Finned Absorber Geometries for Solar Air Heating Collectors K. Pottier, C. M. Sippel, A. Beck, J. Fricke ........................................................................................................................................... 111-62 Inclination Dependency of Flat Plate Collector Heat Losses G. Rockendorf, B. Bartelsen, M. Kiermasch ......................................................................................................................................... 111-72 PV-Hybrid and Thermo-Electric-Coliectors G. Rockendorf, R. Sillmann, L. Podlowski, B. Litzenburger .................................................................................................................111-76 Elastomer-MetaI-Absorber- Development and Application G. Rockendorf, B. Bartelsen, N. Vennernann, R. Tepe, K. Lorenz, G. Purkarthofer ..........................................................................111-83 Solar Absorber System for Preheating Feeding Water District Heating Nets K. Vajen, M. Kr~cner, R. Orths, E. K. Boronbaev, A. Paizuldaeva .......................................................................................................111-90 Statistical Analysis of Solar Collector Test Results in View of Future Certification K. VoropouIos, E. Mathioulakis, V. Belessiotis ..................................................................................................................................... 111-92 Thermal and Electrical Yield of a Combipanel H. A. Zondag, D. W. de Vries, A. A. van Steenhoven, W. G. J. van Helden, R. J. C. van Zolingen ..................................................111-96 A Comparative Investigation of Radiation Heat Transfer in Transparent Insulation with Differemt Reflection Models Y. Zvirin, B. Aronov .............................................................................................................................................................................. 111-192
Solar Hot Water and Thermal Energy Supply Thermal Destratification in Small Standard Solar Tanks Due to Mixing During Tapping E. Andersen, S. Furbo ......................................................................................................................................................................... II1-111 Integrated Thermal Improvements for Greenhouse Cultivation in the Central Part of Argentina J. R. Barral, P. D. Galimberti, A. Barone, M. A. Lara ......................................................................................................................... 111-120
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In Situ Short -Term Test for Large Solar Thermal Systems N. Benz, T. Beikircher, M. Gut, P. Kronthaler, C. Oberdorf, W. SchSIkopf, H. DrOck.......................................................................111-126 Solar Process Heat with Non-Concentrating Collectors for Food Industry N. Benz, M. Gut, T. Beikircher, W. RuB .............................................................................................................................................. 111-131 Laboratory Testing of Integrated Collector Storage (ICS) Systems with Transparent Insulation Material M. Bosanac, J. E. Nielsen ................................................................................................................................................................... 111-137 Uncertainty in Economical Analysis of Solar Water Heating and Photovoltaic Systems S. Co,e, S. L. de Abreu, R. ROther..................................................................................................................................................... 111-141 Solar Pond as a Power Source for Desalination U. Fisher ............................................................................................................................................................................................... 111-150 Multistage Still J. Franco, L. R. Saravia, S. Esteban ................................................................................................................................................... 111-155 Development of a Smart solar Tank S. Furbo, E. Andersen ......................................................................................................................................................................... 111-160 Thermal Modelling and Performance Prediction of Drying Processes under Open-Sun-Drying H. P. Garg, R. Kumar .......................................................................................................................................................................... 111-170 Medium Scale solar Crop Dryers for Agricultural Products O. Headley, W. Hinds .......................................................................................................................................................................... 111-175 The Marstal Central Solar Heating Plant: Design and Evaluation A. Heller, J. Dahrn................................................................................................................................................................................ 111-180 A Combined Ejector Cooling and Hot Water Supply System Using Solar and Waste Heat Energy B. J. Huang, V. A. Petrenko ................................................................................................................................................................ 111-188 A Solar Still with Minimum Inclination and Coupled to an Outside Condenser D. Inan, A. El-Bahi ............................................................................................................................................................................... 111-191 Modelling of a Thermosyphonally Driven Discharge Unit of a Storage Tank U. Jordan, K. Vajen, B. Knopf, A. Spieler, F. Hilmer .......................................................................................................................... 111-197 Performance of Transparently Insulated Solar Passive Hot Water Systems N. D. Kaushika, K. S. Reddy ............................................................................................................................................................... 111-203 Thermodynamic Study of a Regenerative Water Distiller G. Koury Costa, N. Fraidenraich ......................................................................................................................................................... 111-211 The Performance and Analysis of a Multiple - Effect Solar Still Utilizing Solar and/or Waste Thermal Energy A. I. Kudish, E. G. Evseev, L. Horvath, G. Mink ................................................................................................................................. 111-216 Performance and Analysis of a Multiple Effect Solar Still Utilizing an Internal Multi - Tubular Heat Exchanger for Thermal Energy Recycle G. Mink, L. Horvarth, E. G. Evseev, A. I. Kudish................................................................................................................................ 111-226 Modelling Thermosyphon Solar Water Heaters with Mantle Collector-Loop Heat Exchangers G. L. Mordson, G. Rosengarten, M. Behnia ....................................................................................................................................... 111-236 Bridging the Gap: Research and Validation of the DST Performance Test Method for CEN and ISO Standards- Project Results D. Naron, M. Rolloos, M. J. Carvalho ................................................................................................................................................. 111-245 Research on a New Type of Heat Pipe Vacuum Tube Solar Water Heater N. Zhu, H. Zinian .................................................................................................................................................................................. 111-253 Solar Process Heat: Distillation, Drying, Agricultural and Industrial Uses B. Norton .............................................................................................................................................................................................. 111-256 Brackish Water Destillation with Plane Microporous Membranes Driven by Temperature Difference L. Odicino, J. Marchese, D. A. Perellb, G. Lesino .............................................................................................................................. 111-261 Effective Solar Energy Utilisation - More Dependent on System Design Than Solar Collector Efficiency J. Rekstad, L. Henden, A. G. Imenes, F. Ingebretsen, M. Meir, B. Bjerke, M. Peter ........................................................................111-265 Dynamical Model for Solar Still Validated From Optical and Thermal Parameters Measured Experimentally H. Romero-Paredes, E. Torijano, A. Vdzquez, A. Tortes, J. J. Arnbriz, E. Torijano Jr .....................................................................111-271 Characteristics of Vertical Mantle Heat Exchangers for Solar Water Heaters L. J. Shah, G. L. Mordson, M. Behnia ................................................................................................................................................. 111-276 A System for Solar Process Heat for Decentralised Applications in Developing Countries F. Sp~te, B. Hafner, K. Schwarzer ...................................................................................................................................................... 111-286
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Performance of a Cascade of Flat Plate Collectors T. Tomson ............................................................................................................................................................................................ 111-292 A Solar Absorption Air-Conditioning Plant Using Heat-Pipe Evacuated Tubular Collectors H. Zinian, Z. Ning ................................................................................................................................................................................. 111-297 Advanced Fuzzy Control of the Temperature in the Test Chamber B. Zupancic, I. Skrjanc, A. Krainer, B. Furlan ..................................................................................................................................... 111-304
Solar, Thermal and Photovoltaic Concentrating Collectors Design and Construction of a Line-Focus Parabolic Trough Solar Concentrator for Electricity Generation G. (3. Bakos, D. AdamopouIos, N. F. Tsagas, M. Soursos ................................................................................................................ 111-315 The Duct Selective Volumetric Receiver: Potential for Different Selectivity Strategies and Stability Issues • G. Casals, J. I. Ajona ...................................................................................................................................................................... 111-324 A Parabolic Dish Concentrator From a Telecommunication Antenna: Optical and Thermal Study of the Receiver C. A. Estrada, R. Dorantes, E. Rincon ................................................................................................................................................ 111-333 Efficiency Improvement of Parabolic Trough Collectors by Means of Additional End Reflectors T. H. Fend, J. Leon, P. Binner, R. Kemme, K. -J. Riffelrnann, R. Pitz-Paal ...................................................................................... 111-337 Experimental Performance of a PV V-Trough System N. Fraidenraich, E. M. de Souza Barbosa .......................................................................................................................................... 111-342 Performance Analyses of a Combined Photovoltaicrrhermal (PV/T) Collector with Integrated CPC Throughs H. P. Garg, R. S. Adhikad ................................................................................................................................................................... 111-349 An Astigmatic Corrected Target Aligned Solar Concentrator for Solid State Laser Pumping M. Lando, J. Kagan, B. Linyekin, L. Sverdalov, G. Pecheny, U. Achiam .......................................................................................... 111-354 Nonimaging Fresnel Lens Concentrators for Photovoltaic Applications Ft. Leutz, A. Suzuki, A. Akisawa, T. Kashiwagi .................................................................................................................................. 111-358 Thermo-Mechanical Design of a Large Compound Parabolic Concentrator for 500 KWt Solar Central Receiver System G. Miron, S. Weis, I. Anteby, B. Ostreich, E. Taragan ....................................................................................................................... 111-367 Simulation and Analysis of the Performance of Low Concentration PV Modules M. Munschauer, K. Heumann ............................................................................................................................................................. 111-370 Practical Design Considerations for Secondary Concentrators at High Temperatures J. O'Gallagher, R. Winston .................................................................................................................................................................. 111-377 Comparison of Predicted and Measured Performance of an Integrated Compound Parabolic Concentrator (ICPC) J. O'Gallagher, FT.Winston, J. Muschaweck, A. Ft. Mahoney, V. Dudley .......................................................................................... 111-382 Double-Tailored Imaging Concentrators H. Flies, J. M. Gordon .......................................................................................................................................................................... 111-388 Development and Test of an Equipment to Replace Broken Glass Envelops of Receiver Tubes in Parabolic Trough Collectors K. -J. Riffelmann, M. B6hmer, T. Fend, R. Pitz-Paal, C. Spitta, J. Leon ........................................................................................... 111-394 Cooling of PV Modules Equipped with Low Concentrating CPC Reflectors M. R6nnelid, B. Kadsson, P. Krohn, B. Perers ................................................................................................................................... 111-400 A Solar Bowl in India S. Rousseau, G. Guigan, J. Harper .................................................................................................................................................... 111-405 The Development and Testing of Small Concentrating PV Systems G. R. Whiffield, R. W. Bentley, C. K. Weatherby, A. Hunt, H. -D. Mohdng, F. H. Klotz, P. Keuber, J. C. Minano, E. Alarte-Garvi .......................................................................................................................................................................................... 111-409
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Active Cooling, Refrigeration and Dehumidification Thermodynamic Design of a Solar Refrigerator to Conserve Sea Products H. O. Arias-Varela, W. Soto Gomez, O. Castillo-Lopez, R. Best-Brown ...........................................................................................111-419 Demonstration of a New Type of ICPC in a Double Effect Absorption Cooling System W. S. Duff, R. Winston, J. J. O'Gallagher, T. Henkel, J. Muschaweck, R. Chdstiansen, J. Bergquam ...........................................111-424 Indirect Evaporative Cooling through a Concrete Ceiling B. Givoni, S. Nutalaya ......................................................................................................................................................................... 111-428 Experimental Studies on a Hybrid Dryer S. Kurnar, G. A. Mastekbayeva, P. C. Bhatta, M. A. Leon................................................................................................................. 111-434 Combined Solar Heating and Radiative Cooling System M. Meir, H. Storas, J. Rekstad ............................................................................................................................................................ 111-441 Hybrid Solar/Gas Cooling Ejector Unit for a Hospital in Mexico J. L. Wolpert, M. V. Nguyen, S. B. Riffat............................................................................................................................................. 111-447
Thermal Storage The Freezing Process of Water Inside a Vertical Cylinder with a Finned Tube Y. Changsoon, S. Taebeorn, K. Jaeyoon ........................................................................................................................................... 111-455 The Ciclops System: Optimised Management of Middle-Sized-Hybrid Wind-PV-Diesel Plants E. Uobet, J. Sold, J. Pitarch, J. Prats.................................................................................................................................................. 111-462 Solar District Heating with a Combined Pit and Duct Storage in the Underground M. Reuss, J. P. Mueller ....................................................................................................................................................................... 111-468 Solar Heating with Heat Pump and Ice Storage A. B. Schaap, J. M. Warmerdam, E. E. Gramsbergen ....................................................................................................................... 111-475 An Analysis of Phase Change Heat Transfer in a Solar Thermal Energy Store A. Trp, B. Frankovic, K. Lenic ............................................................................................................................................................. 111-484 Modelling of Two - Layer Stratified Stores J. van Berkel, C. C. M. Rindt, A. A. van Steenhoven ......................................................................................................................... 111-490
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Index of Authors
A A. Abete Performance of Grid-Connected Photovoltaic Plants ........................................................................................................................... 1-223 U. Achiam An Astigmatic Corrected Target Aligned Solar Concentrator for Solid State Laser Pumping ......................................................... 111-354 M. Adachi Optical Properties and Radiative Cooling Power of White Paints ....................................................................................................... 1-485 A. D. Adamopoulos A Comparison of Spectral Total Atmospheric Transmission Between Summer and Winter in Athens, Greece ............................... 1-396 The Meteorological Radiation Model .................................................................................................................................................... 1-406 D. Adamopoulos Design and Construction of a Line-Focus Parabolic Trough Solar Concentrator for Electricity Generation .................................... 111-315 R. S. Adhikari Performance Analyses of a Combined PhotovoltaicJThermal (PV/T) Collector with Integrated CPC Throughs ............................. 111-349 R. Aguiar Interannual Variability of Meteorological Parameters in Temperate Climates .................................................................................... 1-353 The Meteorological Radiation Model .................................................................................................................................................... 1-406 M. M. Agumba Solar Energy - a True Option for Rural Electrification in Kenya Hindered by Unfavourable Policy ................................................... 11-481 H. Aharoni ITO/InP Photovoltaic Devices .................................................................................................................................................................
1-95 D. W. ARken Bioclimatic Designs for the New University of Cyprus Campus. 1st Competition: Facilities for Science and Technology ............................................................................................................................................................................................... 11-3 J. I. Ajona The Duct Selective Volumetric Receiver: Potential for Different Selectivity Strategies and Stability Issues ................................... 111-324 A. Akisawa Nonimaging Fresnel Lens Concentrators for Photovoltaic Applications ............................................................................................ 111-358 L. Alados-Arboledas Comparison of Several Parameterized Models for Global Insolation under Cloudy Skies ................................................................. 1-435 E. Alarte-Garvi The Development and Testing of Small Concentrating PV Systems ................................................................................................ 111-409 A. A. Alvarado Simulation and Test of Peltier Elements in Connection with Photovoltaic Cells ................................................................................. 1-113 G. Alvarez Heat Transfer through a Duovent Glass with Chemically Deposited Solar Control Coating ............................................................. 11-199 I. Alxneit Determination of the Spectral Emittance in the Visible Range at High Temperatures Supported by Laser Heating ........................ 1-372 P. Amado Using Photovoltaics for Agricultural Processing Activities in Upper Mustang (Nepal) ....................................................................... 11-495 J. J. Ambriz Dynamical Model for Solar Still Validated From Optical and Thermal Parameters Measured Experimentally ................................ 111-271 Simulation of Dynamic Behaviour of a Solar Reactor-Receiver as a Function of Solar Concentrated Radiation Profile ..................................................................................................................................................................................................... 1-296 E. Andersen Development of a Smart Solar Tank ................................................................................................................................................... 111-160 Thermal Destratification in Small Standard Solar Tanks Due to Mixing Dudng Tapping ................................................................. II1-111
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Y. Ando Heat Recovery Experiments with Concentration Gradient Catalyst Layer in a Solar Chemical Heat Pump ..................................... 1-549
J. Angles A Climatological Database of the Linke Turbidity Factor ..................................................................................................................... 1-432 V. Anikeev The Solar Chemistry Program of the International Energy Agency's Implementing Agreement Solarpaces ...................................... 1-64 I. Anteby Thermo-Mechanical Design of a Large Compound Parabolic Concentrator for 500 KWt Solar Central Receiver System .................................................................................................................................................................................................
111-367
J. Appelbaum Aeration of Fish-Ponds by Photovoltaic Power .................................................................................................................................... 1-175 G. Aradjo Combined Photovoltaic and Solar Thermal Systems for Facade Integration and Building Insulation ................................................ 11-88
D. G. Aravantinos The Influence of a Planted Roof on the Passive Cooling of Buildings ............................................................................................... 11-169
M. Arciga Municipal Solid Waste Evaluation as a Source of Energy in Mexico City ........................................................................................... 1-566
H. D. Arias-Varela Hybrid System Heat Pump - Solar Air Heater for the Drying of Agricultural Products ....................................................................... 11-512 Thermodynamic Design of a Solar Refrigerator to Conserve Sea Products ..................................................................................... 111-419 B. Aronov A Comparative Investigation of Radiation Heat Transfer in Transparent Insulation with Differernt Reflection Models ................... 111-102 S. J. Arulanandam A CFD Heat Transfer Analyses of the Transpired Solar Collector under No-Wind Conditions .......................................................... 111-29
A. K. Athienitis Numerical Model of a Building with Transparent Insulation .................................................................................................................. 11-10 Use of the Electronic Book "Building Thermal Analysis" in Passive Solar Design and Education .................................................... 11-245
R. A. Attalage A Web Based Course for Learning Solar Thermal Processes ............................................................................................................ 11-293
B H. Baba Analysis of Thermal Performance on an Air-Type Solar Collector with 2- Glass Using Carbon Fiber Sheet as Collecting Material .................................................................................................................................................................................
111-35
IC Backes The Solar-Campus JOlich - Actual Status ........................................................................................................................................... 11-156 G. C, Bakos Design and Construction of a Line-Focus Parabolic Trough Solar Concentrator for Electricity Generation .................................... 111-315
J. L. Balenzategui Modelling the Thermal Effects of Semitransparent PV - Modules ...................................................................................................... 11-178
F. Baonza Distributed Power From Solar Tower Systems: A Mius Approach ...................................................................................................... 1-286 M. Barak Aeration of Fish-Ponds by Photovoltaic Power .................................................................................................................................... 1-175
A. Barone Integrated Thermal Improvements for Greenhouse Cultivation in the Central Part of Argentina ..................................................... 111-120
J. R. Barral Integrated Thermal Improvements for Greenhouse Cultivation in the Central Part of Argentina ..................................................... 111-120
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B. Bartelsen Elastomer-Metai-Absorber - Development and Application .................................................................................................................
111-83
Inclination Dependency of Flat Plate Collector Heat Losses ............................................................................................................... H. Barthels
111-72
Phoebus - an Autonomous Supply System with Renewable Energy .................................................................................................. F. J. Batlles
1-336
Compadson of Several Parameterized Models for Global Insolation under Cloudy Skies ................................................................. 1-435 L. Batos Set-Up of a Laboratory for Research and Education in Solar Energy in Rio De Janeiro .................................................................. 11-288 O. Bauer A Climatological Database of the Linke Turbidity Factor ..................................................................................................................... A. Beck
1-432
Optimized Finned Absorber Geometries for Solar Air Heating Collectors .......................................................................................... 111-62 Pearl Luster Pigments as Overheating Protection in Transparently Insulated Solar Facades ........................................................... 1-453 M. Behnia Characteristics of Vertical Mantle Heat Exchangers for Solar Water Heaters .................................................................................. 111-276 Modelling Thermosyphon Solar Water Heaters with Mantle Collector-Loop Heat Exchangers ....................................................... 111-236 T. Beikircher In Situ Short -Term Test for Large Solar Thermal Systems ..............................................................................................................
111-126
Solar Process Heat with Non-Concentrating Collectors for Food Industry ....................................................................................... II1-131 S. Beisel Passive Solar Office Building: Results of the First Heating Period ..................................................................................................... V. Belessiotis
11-183
Statistical Analysis of Solar Collector Test Results in View of Future Certification ............................................................................ 111-92 Uncertainty in Solar Collector Testing Results ..................................................................................................................................... 111-56 M. Belusko Roof Integrated Heating and Cooling System ....................................................................................................................................... R. W. Bentley
11-18
The Development and Testing of Small Concentrating PV Systems ................................................................................................ 111-409 N. Benz In Situ Short -Term Test for Large Solar Thermal Systems ..............................................................................................................
111-126
Solar Process Heat with Non-Concentrating Collectors for Food Industry ....................................................................................... II1-131 S. Berger Distributed Power From Solar Tower Systems: A Mius Approach ...................................................................................................... 1-286 J. Bergquam Demonstration of a New Type of ICPC in a Double Effect Absorption Cooling System ................................................................... 111-424 G. J. Besler Hybrid Solar Collectors for Microclimate Forming System ..................................................................................................................... R. Best-Brown
111-3
Hybrid System Heat Pump - Solar Air Heater for the Drying of Agricultural Products ....................................................................... 11-512 Thermodynamic Design of a Solar Refrigerator to Conserve Sea Products ..................................................................................... 111-419 H. G. Bayer Daylight and Solar Irradiance Data Derived From Satellite Observations - the Satellight Project .................................................... 11-368 S. C. Bhattacharya Dissemination of Renewable Energy in Developing Countries: Experiences of a Regional Project in Asia ..................................... 11-489 A Study on Improved Institutional Biomass Stoves ........ ..................................................................................................................... 11-484 Two-Stage Gasification of Wood with Preheated Air Supply: A Promising Technique for Producing Gas of Low Tar Content ................................................................................................................................................................................................... P. C. Bhatta
1-557
Experimental Studies on a Hybrid Dryer. ............................................................................................................................................ J. Bilbao
111-434
The Meteorological Radiation Model .................................................................................................................................................... P. Binner
1-406
Efficiency Improvement of Parabolic Trough Collectors by Means of Additional End Reflectors .................................................... 111-337
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B. Bjerke Effective Solar Energy Utilisation - More Dependent on System Design Than Solar Collector Efficiency ...................................... 111-265 J. Blackmon High Temperature Solar Energy Conversion Systems ..........................................................................................................................
1-71
D. Blamont Using Photovoltaics for Agricultural Processing Activities in Upper Mustang (Nepal) ....................................................................... 11-495 J. Blanco Compound Parabolic Concentrator Technology Development to Commercial Solar Detoxification Applications ............................ 11-427 The Solar Chemistry Program of the International Energy Agency's Implementing Agreement Solarpaces ...................................... 1-64 M. Blanco Distributed Power From Solar Tower Systems: A Mius Approach ...................................................................................................... 1-286 E. Bobeico Spray-Deposited SnO2-nSi Solar Cells ................................................................................................................................................ 1-109 M. B6hmer Development and Test of an Equipment to Replace Broken Glass Envelops of Receiver Tubes in Parabolic Trough Collectors ............................................................................................................................................................................................. 111-394 J. Boland The Analytic Solution of the Differential Equations Describing Heat Flow in Houses .......................................................................... 11-27 Interannual Variability of Meteorological Parameters in Temperate Climates .................................................................................... 1-353 A. Bonneschky Social-Technical Assessment of Photovoltaic Systems Installed in the First Region of Chile .......................................................... 11-503 T. Book Marketing and Selling Solar Energy Equipment ..................................................................................................................................
11-437
E. K. Boronbaev Solar Absorber System for Preheating Feeding Water District Heating Nets ..................................................................................... 111-90
M. Bosanac Laboratory Testing of Integrated Collector Storage (ICS) Systems with Transparent Insulation Material ....................................... 111-137 B. Bourges The New European Solar Radation Atlas: a Tool for Designers, Engineers and Architects ............................................................. 11-400 H. Boyer Building Design in Tropical Climates Elaboration of the ECODOM Standard in the French Tropical Islands .................................... 11-59 W. A. Brocke Phoebus - an Autonomous Supply System with Renewable Energy .................................................................................................. 1-336
L. Y. Bronicki An Enabling Technology Opens the Way to Large Scale Use of Solar Energy ................................................................................. 11-444 Financing of Private Renewable Energy Projects; Hurdles and Opportunities ....................................................................................... I-3 Transportation of Electricity Production ............................................................................................................................................... 11-447 E. Brundrett A CFD Heat Transfer Analyses of the Transpired Solar Collector under No-Wind Conditions .......................................................... 111-29
M. Brunotte Compound Parabolic Concentrator Technology Development to Commercial Solar Detoxification Applications ............................ 11-427 O. Bucek Solar Air Collectors - Investigations on Several Series-Produced Collectors ..................................................................................... 111-17 R. Buck Solar-Assisted Syngas-Driven Power System ...................................................................................................................................... 1-544 C. B0hler Sun Protection System Based on CPC's with Total Internal Reflection ............................................................................................. 11-226
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516
C R. G. Cable An Overview and Operation Optimization of the Kramer Junction Solar Electric Generating Systems ............................................. 1-241 B. Calcagni Comparison of Experimental Measurements and Numerical Simulation in an Atrium Building ........................................................ 11-237 Numerical Simulation and Scale Model Measurements of Daylighting Systems in an Existent Building ......................................... 11-218 R. Calero Small Scale Photovoltaic R. O. Desalination - Experience in Gran Canaria ...................................................................................... 11-527
M. E. Calixto Electrodeposited CdTe Based Photovoltaic Structures on Metallic Substrates .................................................................................. 1-142 B. Cancino Simulation and Test of Peltier Elements in Connection with Photovoltaic Cells ................................................................................. 1-113 Social-Technical Assessment of Photovoltaic Systems Installed in the First Region of Chile .......................................................... 11-503 I. G. Capeluto On the Use of the Solar Collection Envelope for Determining the Building Shape .............................................................................. 11-33 J. Carpio Multimedia Library of Renewable Energies ......................................................................................................................................... 11-254 A. M. Carvalho Photovoltaic Water Pumping Systems Installer Training: a Partnership Experience Between the University and Sao Francisco Hydroelectric Power Plant ................................................................................................................................................... 11-260 M. J. Carvalho Bridging the Gap: Research and Validation of the DST Performance Test Method for CEN and ISO Standards Project Results ..................................................................................................................................................................................... 111-245 C. Casalle Compound Parabolic Concentrator Technology Development to Commercial Solar Detoxification Applications ............................ 11-427 X. G. Casals The Duct Selective Volumetric Receiver: Potential for Different Selectivity Strategies and Stability Issues ................................... 111-324 O. Castillo-Lopez Thermodynamic Design of a Solar Refrigerator to Conserve Sea Products ..................................................................................... 111-419 M. Castro Multimedia Library of Renewable Energies ......................................................................................................................................... 11-254 R. Celaire Building Design in Tropical Climates Elaboration of the ECODOM Standard in the French Tropical Islands .................................... 11-59
E. G. Chachkhiani Solar Energy Resources and Their Application Perspectives in Georgia (Using Semiconductive Photovoltaic Cells) ..................... 1-185 G. E. Chachkhiani Solar Energy Resources and Their Application Perspectives in Georgia (Using Semiconductive Photovoltaic Cells) ..................... 1-185 Y. Changsoon The Freezing Process of Water Inside a Vertical Cylinder with a Finned Tube ................................................................................ 111-455
K. P. Cheung Developing a Web-Based Leaming Environment for Building Energy Efficiency and Solar Design in Hong Kong ......................... 11-278 Elaboration on the Design and Operation Principles of a Heavy Duty Universal Sunlight Heliodon Assembled From Precision Machining Tools ..................................................................................................................................................................... T. Chikahisa A Method for Establishing a Solar Power Network for Emergency Integrated Cost Effectively in a CHP (Combined Heat and Power) Network ....................................................................................................................................................................
11-38
11-361
R. Christiansen Demonstration of a New Type of ICPC in a Double Effect Absorption Cooling System ................................................................... 111-424 C. Chrysanthou Bioclimatic Designs for the New University of Cyprus Campus. 2nd Competition: Face A, Student Housing ................................. 11-141
ISES Solar World Congress 1999, Volume III
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S. L. Chung Elaboration on the Design and Operation Principles of a Heavy Duty Universal Sunlight Heliodon Assembled From Precision Machining Tools .....................................................................................................................................................................
11-38
K. Colbow Research and Development on the First AC BIPV Installation in Canada ......................................................................................... 11-165 G. Cole Raps in a Virtual World - a Web Based Remote Area Power Supply System ................................................................................... 11-315
M. Collares-Pereira Compound Parabolic Concentrator Technology Development to Commercial Solar Detoxification Applications ............................ 11427 S. Colle Distribution of Solar Irradiation in Brazil Derived From Geostationary Satellite Data ......................................................................... 1-362 Uncertainty in Economical Analysis of Solar Water Heating and Photovoltaic Systems .................................................................. 111-141 A. Colmenar Multimedia Library of Renewable Energies ......................................................................................................................................... 11-254 C. Cortina Heat Transfer through a Duovent Glass with Chemically Deposited Solar Control Coating ............................................................. 11-199
P. Couto Distribution of Solar Irradiation in Brazil Derived From Geostationary Satellite Data ......................................................................... 1-362 E. Crino Distribution of the Ultraviolet Solar Radiation in the Sky of San Luis (Argentina) .............................................................................. 1-376
D J. Dahm The Marstal Central Solar Heating Plant: Design and Evaluation ..................................................................................................... 111-180 S. L. de Abreu Distribution of Solar Irradiation in Brazil Derived From Geostationary Satellite Data ......................................................................... 1-362 Uncertainty in Economical Analysis of Solar Water Heating and Photovoltaic Systems .................................................................. 111-141
B. J. de Boer Modelling the Thermal Effects of Semitransparent PV - Modules ...................................................................................................... 11-178 A. de Miguel The Meteorological Radiation Model .................................................................................................................................................... 1-406 E. de Oliveira Fernandes Solar Energy in the Built Environment: the Building as a System Plus the Systems in the Building ..................................................... I-5
E. M. de Souza Barbosa Experimental Performance of a PV V-Trough System ...................................................................................................................... 111-342 Photovoltaic Water Pumping Systems Installer Training: a Partnership Experience Between the University and Sao Francisco Hydroelectric Power Plant ................................................................................................................................................... 11-260 D. W. de Vries Thermal and Electrical Yield of a Combipanel ...................................................................................................................................... 111-96 G. Del Tin Integration of Communication and Development in the "Alta Valle Di Susa" Project for Solar Energy ............................................ 11-284 J. M. DeRour Modelling Solar Energy Input in Greenhouses .................................................................................................................................... I1-117 M. T. Deluigi Distribution of the Ultraviolet Solar Radiation in the Sky of San Luis (Argentina) .............................................................................. 1-376 F. N. Demirbilek Solar Architecture in Turkey: State-of-the-Art. ....................................................................................................................................... 1141 C.J. Dey Transition Strategies for Solar Thermal Power Generation ................................................................................................................. 1-272
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ISES Solar World Congress1999, Volume III
N. Dischinger Compound Parabolic Concentrator Technology Development to Commercial Solar Detoxification Applications ............................ 11-427 V. Dobrokhotov Renewable Energy Sources Utilization in Russia ...............................................................................................................................
11-395
T. Doi Heat Recovery Experiments with Concentration Gradient Catalyst Layer in a Solar Chemical Heat Pump ..................................... 1-549 A. Donath Windows in the Attic: Thermophysical Problems of Inclined Windows ................................................................................................
11-48 R. Dorantes A Parabolic Dish Concentrator From a Telecommunication Antenna: Optical and Thermal Study of the Receiver ....................... 111-333 P. Doron A Multistage Solar Receivers: The Route to High Temperature ..........................................................................................................
1-258
I. Dostrovsky Transportation of Electricity Production ...............................................................................................................................................
11-447
H. Dr0ck In Situ Short -Term Test for Large Solar Thermal Systems ..............................................................................................................
111-126 S. Duchan A Multistage Solar Receivers: The Route to High Temperature .......................................................................................................... 1-258 V. Dudley Comparison of Predicted and Measured Performance of an Integrated Compound Parabolic Concentrator (ICPC) .................... 111-382 W. S. Duff Demonstration of a New Type of ICPC in a Double Effect Absorption Cooling System ................ ................................................... 111-424 Evacuated Tubular Collector Water Pasteurization Systems ............................................................................................................. 11-509 Experimental Evaluation of Selective Surfaces in a High Vacuum ...................................................................................................... 1-451 Testing of a Flat Plate Collector with Selective and Nonselective Absorbers That Are Otherwise Identical ....................................... 111-4 C. Dundar The Analysis of Wind Data and Wind Energy Potential in Bandirma, Turkey ..................................................................................... 1-329
A. Dutta Two-Stage Gasification of Wood with Preheated Air Supply: A Promising Technique for Producing Gas of Low Tar Content ...................................................................................................................................................................................................
1-557
E M. Eberl Controller Design for Injection Mode Driven Direct Solar Steam Generating Parabolic Trough Collectors ....................................... 1-247 S. Eckhoff Determination of the Spectral Emittance in the Visible Range at High Temperatures Supported by Laser Heating ........................ 1-372
M. Eck Controller Design for Injection Mode Ddven Direct Solar Steam Generating Parabolic Trough Collectors ....................................... 1-247 U. Eicker Thermal Performance of Building Integrated Ventilated PV Facades .................................................................................................. Understanding the Potential of Ventilated PV Facades ......................................................................................................................
11-55 11-110
A. EI-Bahi A Solar Still with Minimum Inclination and Coupled to an Outside Condenser ................................................................................. 111-191 B. Elsworth Investigation of the Back Contact of Cadmium Telluride Solar Cells .................................................................................................. 1-124 B. Emonts Phoebus - an Autonomous Supply System with Renewable Energy .................................................................................................. 1-336 N. Endoh Analysis of Thermal Performance on an Air-Type Solar Collector with 2- Glass Using Carbon Fiber Sheet as Collecting Material ................................................................................................................................................................................. 111-35
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M. Epstein Potential Efficiencies of Solar-Operated Gas Turbine and Combined Cycle, Using the Reflective Tower Optics ............................ 1-302 Solar-Assisted Syngas-Driven Power System ...................................................................................................................................... 1-544 The Solar Chemistry Program of the International Energy Agency's Implementing Agreement Solarpaces ...................................... 1-64 E. Erell A Novel Ventilated Reversible Glazing System ...................................................................................................................................
11-205
A. Erez High Temperature Solar Energy Conversion Systems ..........................................................................................................................
1-71
D. I. Eryildiz Solar Architecture in Turkey: State-of-the-Art. .......................................................................................................................................
11-41
T. Esbensen Brundtiand Solar City Network ............................................................................................................................................................. 11-379 S. Esteban Multistage Still ...................................................................................................................................................................................... 111-155 C. A. Estrada A Parabolic Dish Concentrator From a Telecommunication Antenna: Optical and Thermal Study of the Receiver ....................... 111-333 Y. Etzion A Novel Ventilated Reversible Glazing System ................................................................................................................................... 11-205 E. A. Eumorfopoulou Comparative Assessment of the Thermal Behavior of a Planted Roof vs. a Bare Roof in Thessaloniki .......................................... 11-126 E. G. Evseev Characterization and Inter-Compadson of the Global and Beam Radiation Measured at Three Sites in the Southem Region of Israel by Statistical Analysis .................................................................................................................................................
1-419
The Performance and Analysis of a Multiple - Effect Solar Still Utilizing Solar and/or Waste Thermal Energy .............................. 111-216 Performance and Analysis of a Multiple Effect Solar Still Utilizing an Internal Multi - Tubular Heat Exchanger for Thermal Energy Recycle .....................................................................................................................................................................
111-226
Research and Development of Solar Collectors Fabricated From Polymeric Material .......................................................................
111-40
P.G.M. Eykens Diffusive Properties of Dry and Wet Glass and Plastics ......................................................................................................................
1-462
I. Farkas Hungarian UNESCO Solar Participation Program .............................................................................................................................. 11-382 A. J. Fasulo Comparison Between a Simple Solar Collector Accumulator and a Conventional Accumulator ....................................................... II1-11 Distribution of the Ultraviolet Solar Radiation in the Sky of San Luis (Argentina) .............................................................................. 1-376 H. Fechner Solar Air Collectors - Investigations on Several Series-Produced Collectors ..................................................................................... 111-17 T. H. Fend Development and Test of an Equipment to Replace Broken Glass Envelops of Receiver Tubes in Parabolic Trough Collectors ............................................................................................................................................................................................. 111-394 Efficiency Improvement of Parabolic Trough Collectors by Means of Additional End Reflectors .................................................... 111-337 A. M. Fernandez Preparation and Characterization of Sb-Se Thin Films by Electrodeposited Technique for Photovoltaic Application ...................... 1-120 P. Fernandez Compound Parabolic Concentrator Technology Development to Commercial Solar Detoxification Applications ............................ 11-427 V. Fernandez Distributed Power From Solar Tower Systems: A Mius Approach ...................................................................................................... 1-286 L. Ferraris Performance of Grid-Connected Photovoltaic Plants ........................................................................................................................... 1-223
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ISES Solar World Congress 1999, Volume III
D. Feuermann The Israeli Insulation Standard for Offices ............................................................................................................................................. 11-71 L. I. Filetoth Comparative Analysis of Daylighting Systems Investigating Illumination and Structure ................................................................... 11-213 F. Fillipetti Numerical Simulation and Scale Model Measurements of Daylighting Systems in an Existent Building ......................................... 11-218 U. Fisher High Temperature Solar Energy Conversion Systems .......................................................................................................................... 1-71 Solar-Assisted Syngas-Driven Power System...................................................................................................................................... 1-544 Solar Pond as a Power Source for Desalination ................................................................................................................................ 111-150 Transportation of Electricity Production ............................................................................................................................................... 11-447 J. J. Flores Heat Transfer through a Duovent Glass with Chemically Deposited Solar Control Coating ............................................................. 11-199 J. Follari Comparison Between a Simple Solar Collector Accumulator and a Conventional Accumulator ....................................................... II1-11 R. E. Foster Lessons Leamed From the Xcalak Village Hybrid System: A Seven Year Retrospective ................................................................. 1-319 A. K. Fotiadi Variability of Atmospheric Turbidity in Athens, Greece ........................................................................................................................ 1-400 N. Fraidenraich Experimental Performance of a PV V-Trough System ...................................................................................................................... 111-,342 Thermodynamic Study of a Regenerative Water Distiller .................................................................................................................. 111-211 A. Francis Solar Energy in Social Housing in the UK ........................................................................................................................................... 11-386 J. Franc<) Multistage Still ......................................................................................................................................................................................
111-155
B. Frankovic An Analysis of Phase Change Heat Transfer in a Solar Thermal Energy Store ............................................................................... 111-484 J. Fricke Optimized Finned Absorber Geometries for Solar Air Heating Collectors .......................................................................................... 111-62 Pearl Luster Pigments as Overheating Protection in Transparently Insulated Solar Facades ........................................................... 1-453 S. Frid Renewable Energy Sources Utilization in Russia ............................................................................................................................... 11-395 TRNSYS Software Application for Solar Thermal Power Plants Simulation and Comparative Analysis ........................................... 1-280 S. D. Frier An Overview and Operation Optimization of the Kramer Junction Solar Electric Generating Systems ............................................. 1-241 K.-H. Funken CO2-Mitigation by Solar Conversion of Hydrocarbons ....................................................................................................................... 11-340 Compound Parabolic Concentrator Technology Development to Commercial Solar Detoxification Applications ............................ 11-427 The Solar Chemistry Program of the Intemational Energy Agency's Implementing Agreement Solarpaces ...................................... 1-64 S. Furbo Development of a Smart Solar Tank ................................................................................................................................................... 111-160 Thermal Destratification in Small Standard Solar Tanks Due to Mixing Dudng Tapping ................................................................. II1-111 B. Furlsn Advanced Fuzzy Control of the Temperature in the Test Chamber .................................................................................................. 111-304 V. Fux Thermal Performance of Building Integrated Ventilated PV Facades .................................................................................................. 11-55 Understanding the Potential of Ventilated PV Facades ...................................................................................................................... 11-110
ISES Solar World Congress 1999, Volume III
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G P. D. Galimberti Integrated Thermal Improvements for Greenhouse Cultivation in the Central Part of Argentina ..................................................... 111-120
G. Galli Passive Cooling System for Remote Locations ....................................................................................................................................
11-66
E. Galvez Social-Technical Assessment of Photovoltaic Systems Installed in the First Region of Chile .......................................................... 11-503
F. Garde Building Design in Tropical Climates Elaboration of the ECODOM Standard in the French Tropical Islands .................................... 11-59
H. P. Garg Performance Analyses of a Combined Photovoltaic/Thermal (PV/T) Collector with Integrated CPC Throughs ............................. 111-349 Thermal Modelling and Performance Prediction of Drying Processes under Open-Sun-Drying ...................................................... 111-170
G. Georgiades A Method for Establishing a Solar Power Network for Emergency Integrated Cost Effectively in a CHP (Combined Heat and Power) Network ....................................................................................................................................................................
11-361
M. Geyer A Guide for Financial Feasible Large-Scale Solar Thermal I PP's ......................................................................................................
11-455
B. Givoni Indirect Evaporative Cooling through a Concrete Ceiling ..................................................................................................................
111-428
O. Goebel Project Diss (Direct Solar) Update on Project Status and Future Planning ........................................................................................ 1-307 A. Goetzberger Angular Selectivity of Seasonal Sun Protection Devices .........................................................................................................................
I-9
Sun Protection System Based on CPC's with Total Intemal Reflection ............................................................................................. 11-226 J. M. Gordon Double-Tailored Imaging Concentrators .............................................................................................................................................
111-388
D. Y. Goswami Recent Developments in Photocatalytic Detoxification and Disinfection Processes of Water and Air ................................................ 1-16 R. Gottschalg Comprehensive Approach for the Estimation of Outdoor Performance of Amorphous Silicon Photovoltaic Devices ....................... 1-129 Investigation of the Back Contact of Cadmium Telluride Solar Cells ..................................................................................................
1-124
E. E. Gramsbergen Solar Heating with Heat Pump and Ice Storage .................................................................................................................................
111-475
M. A. Green 24,7% Efficient Perl Silicon Solar Cells and Other High Efficiency Solar Cell and Module Research at the University of New South Wales ..............................................................................................................................................................................
1-165
A. P. Grifd Electrodeposited CdTe Based Photovoltaic Structures on Metallic Substrates .................................................................................. 1-142
S. Grignaffini Energy Savings Related with the Natural and Artificial Light in the Underground Car Parking Areas .............................................. 11-231 Passive Cooling System for Remote Locations ....................................................................................................................................
11-66
H. G. Groehn Phoebus - an Autonomous Supply System with Renewable Energy ..................................................................................................
1-336
G. Grossman Solar - Powered Systems for Cooling, Dehumidification and Air - Conditioning .................................................................................. 1-21 O. F. Gross Pearl Luster Pigments as Overheating Protection in Transparently Insulated Solar Facades ........................................................... 1-453
F. Gugliermetti Energy Savings Related with the Natural and Artificial Light in the Underground Car Parking Areas .............................................. 11-231 Passive Cooling System for Remote Locations ....................................................................................................................................
11-66
G. Guigan A Solar Bowl in India ...........................................................................................................................................................................
111-405
522
ISES Solar World Congress 1999, Volume III
M. GGnes Calculation of Solar Radiation on Inclined Surfaces in Turkey ............................................................................................................
1-380
D. GGr'zenich Cumulative Energy Demand of Wind Energy and Solar Water Heating Systems .............................................................................
11-345
P.-G. Gutermuth Regulatory and Institutional Measures by the State to Enhance the Deployment of Renewable Energies - the German Experience .................................................................................................................................................................................
1-29
M. Gut In Situ Short -Term Test for Large Solar Thermal Systems ..............................................................................................................
111-126
Solar Process Heat with Non-Concentrating Collectors for Food Industry .......................................................................................
111-131
H B. Hafner A System for Solar Process Heat for Decentralised Applications in Developing Countries ............................................................. 111-286
G. Hakim A New Heat Reflective Polycarbonate Sheet with Spectral Selectivity ...............................................................................................
1-461
U. Halperson Solar Energy: Time to Get Commercial ...............................................................................................................................................
11-463
A. Hammer Daylight and Solar Irradiance Data Derived From Satellite Observations - the Satellight Project .................................................... 11-368 Short-Term Forecasting of Solar Radiation Based on Satellite Data - an Application of Neural Networks and Markov Random Fields .......................................................................................................................................................................................
1-411
R. HanRsch Combined Photovoltaic and Solar Thermal Systems for Facade Integration and Building Insulation ................................................ 11-88 A Guide for Financial Feasible Large-Scale Solar Thermal IPP's ......................................................................................................
11-455
Set-Up of a Laboratory for Research and Education in Solar Energy in Rio De Janeiro .................................................................. 11-288 I. I-laraksingh The Market for Solar Energy in the Caribbean ....................................................................................................................................
11-451
J. Harper A Solar Bowl in India ...........................................................................................................................................................................
111-405
D. R. Harris Heat Flow Analysis in Solar Cell Modules ............................................................................................................................................
1-134
S. Hassid B.A.M.A. (Energy Conserving Buildings) Project: Passive Solar Energy in Popular Residential Apartment BuIdings in Israel ....................................................................................................................................................................................................
11-75
The Israeli Insulation Standard for Offices .............................................................................................................................................
11-71
H.-J. Haubrich Optimization of the Combination of Power Units in Isolated Grids ......................................................................................................
1-342
P. Haueter The Production of Zinc by Thermal Dissociation of Zinc Oxide - Solar Chemical Reactor Design .................................................... 1-539
V. Hliussermann Introduction of Dish Stirling Systems in Morocco. Project Proposal for a Moroccan - German Co-Operation ................................. 11-517
O. Headley Medium Scale Solar Crop Dryers for Agricultural Products ...............................................................................................................
111-175
D. Heinemann Daylight and Solar Irradiance Data Derived From Satellite Observations - the Satellight Project .................................................... 11-368 Short-Term Forecasting of Solar Radiation Based on Satellite Data - an Application of Neural Networks and Markov Random Fields .......................................................................................................................................................................................
1-411
W. G. J. van I-leiden Thermal and Electrical Yield of a Combipanel ......................................................................................................................................
111-96
ISES Solar World Congress 1999, Volume III
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A. Heller The Marstal Central Solar Heating Plant: Design and Evaluation ..................................................................................................... 111-180 L. Henden Effective Solar Energy Utilisation - More Dependent on System Design Than Solar Collector Efficiency ...................................... 111-265 T. Henkel Demonstration of a New Type of ICPC in a Double Effect Absorption Cooling System ................................................................... 111424 K. Hennecke Project Diss (Direct Solar) Update on Project Status and Future Planning ........................................................................................ 1-307 TRNSYS Software Application for Solar Thermal Power Plants Simulation and Comparative Analysis ........................................... 1-280 R. Hernberg The School Physics Program of the Finnish Physical Society.., ......................................................................................................... 11-298 D. Herold The Centre for the Application of Renewable Energies (C.A.R.E.) .................................................................................................... 11-269 Small Scale Photovoltaic R. O. Desalination - Experience in Gran Canaria ...................................................................................... 11-527 A. G. Hestnes Building Integration of a Solar Energy Systems ..................................................................................................................................... 1-36 Sustainability and the Use of Solar Energy: Life Cycle Analyses of a Norwegian Solar Dwelling ...................................................... 11-78 K. Heumann Simulation and Analysis of the Performance of Low Concentration PV Modules ............................................................................. 111-370 M. Higashi Estimation of Direct Solar Irradiance From Global Irradiance by Means of Signal (Wavelet) Analysis ............................................. 1-386 F. Hilmer Modelling of a Thermosyphonally Driven Discharge Unit of a Storage Tank .................................................................................... 111-197 W. Hinds Medium Scale Solar Crop Dryers for Agricultural Products ................................................................................................................ 111-175 Y. Hishinuma A Method for Establishing a Solar Power Network for Emergency Integrated Cost Effectively in a CHP (Combined Heat and Power) Network .................................................................................................................................................................... 11-361 D, Hodgson Evacuated Tubular Collector Water Pasteurization Systems ............................................................................................................. 11-509 Experimental Evaluation of Selective Surfaces in a High Vacuum ...................................................................................................... 1451 Testing of a Flat Plate Collector with Selective and Nonselective Absorbers That Are Otherwise Identical ....................................... 1114 M. Hoffmann The Centre for the Application of Renewable Energies (C.A.R.E.) .................................................................................................... 11-269 K. G. T. Hollands A CFD Heat Transfer Analyses of the Transpired Solar Collector under No-Wind Conditions .......................................................... 111-29 An Empirical Heat Transfer Equation for the Transpired Solar Collectors, Including No-Wind Conditions ....................................... 111-23 V. Horstmann The Centre for the Application of Renewable Energies (C.A.R.E.) .................................................................................................... 11-269 Small Scale Photovoltaic R. O. Desalination - Experience in Gran Canaria ...................................................................................... 11-527 L. Horvath The Performance and Analysis of a Multiple - Effect Solar Still Utilizing Solar and/or Waste Thermal Energy .............................. 111-216 Performance and Analysis of a Multiple Effect Solar Still Utilizing an Internal Multi - Tubular Heat Exchanger for Thermal Energy Recycle ...................................................................................................................................................................... 111-226 H.-M. Ho Eady Results on the Effectiveness of Natural Ventilation at Verbena Height - a High Rise, High Density Housing Development in Hong Kong ................................................................................................................................................................... B. J, Huang
11-83
A Combined Ejector Cooling and Hot Water Supply System Using Solar and Waste Heat Energy ................................................ 111-188 Solar-Photo-Voltaic/ThermaI-Cogeneration Collector .......................................................................................................................... 1-181 S. C. M. Hui Developing a Web-Based Leaming Environment for Building Energy Efficiency and Solar Design in Hong Kong ......................... 11-278 W. C. Hung Solar- Photo-Voltaic/l'hermaI-Cogeneration Collector .......................................................................................................................... I-181
ISES Solar World Congress 1999, Volume III
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A. Hunt The Development and Testing of Small Concentrating PV Systems ................................................................................................
111-409
A. lanetz Characterization and Inter-Comparison of the Global and Beam Radiation Measured at Three Sites in the Southem Region of Israel by Statistical Analysis .................................................................................................................................................
1-419
K. Igarashi Optical Properties and Radiative Cooling Power of White Paints .......................................................................................................
1-485
A. G. Imenes Effective Solar Energy Utilisation - More Dependent on System Design Than Solar Collector Efficiency ...................................... 111-265 L. Imre Hungarian UNESCO Solar Participation Program ..............................................................................................................................
11-382
D. Inan The Analysis of Wind Data and Wind Energy Potential in Bandirma, Turkey ..................................................................................... 1-329 A Solar Still with Minimum Inclination and Coupled to an Outside Condenser ................................................................................. 111-191
D. G. Infield Comprehensive Approach for the Estimation of Outdoor Performance of Amorphous Silicon Photovoltaic Devices ....................... 1-129 Investigation of the Back Contact of Cadmium Telluride Solar Cells .................................................................................................. Thermal Performance of Building Integrated Ventilated PV Facades .................................................................................................. Understanding the Potential of Ventilated PV Facades ......................................................................................................................
1-124 11-55 11-110
F. Ingebretsen Effective Solar Energy Utilisation - More Dependent on System Design Than Solar Collector Efficiency ...................................... 111-265
M. Instance The Design and Development of a Suitable Universal Means of Terminating, Interconnecting and Packaging Photovoltaic Panels for Present and Future Applications ...................................................................................................................
11-468
T. D. Jacobsen Life Cycle Assessments of Solar Collectors in Denmark ....................................................................................................................
11-333
K. Jaeyoon The Freezing Process of Water Inside a Vertical Cylinder with a Finned Tube ................................................................................ 111-455 T. V. Jakhutashvili Solar Energy Resources and Their Application Perspectives in Georgia (Using Semiconductive Photovoltaic Cells) ..................... 1-185 S. Janseen Facade Integrated Solar Collectors .....................................................................................................................................................
11-134
F. Jarach Integration of Communication and Development in the "Alta Valle Di Susa" Project for Solar Energy ............................................ 11-284
P. Jennings Using the World Wide Web for Tertiary Level Renewable Energy Education - the Potential, the Practice and the Possible Problems ................................................................................................................................................................................
II-305
X. -M. Jiang Analysis of Thermal Performance on an Air-Type Solar Collector with 2- Glass Using Carbon Fiber Sheet as Collecting Material .................................................................................................................................................................................
111-35
P. Jing Optical Properties and Radiative Cooling Power of White Paints .......................................................................................................
1-485
ISES Solar World Congress 1999, Volume III
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J. Jones Research and Development on the First AC BIPV Installation in Canada ......................................................................................... 11-165 P. Jones Photovoltaics (PV) Modelling for Cities: a GIS-Building Integrated PV (BIPV) Simulations Approach ............................................. 11-147 U. Jordan Modelling of a Thermosyphonally Driven Discharge Unit of a Storage Tank .................................................................................... 111-197
K J. Kagan An Astigmatic Corrected Target Aligned Solar Concentrator for Solid State Laser Pumping ......................................................... 111.354 H. D. Karnbezidis A Comparison of Spectral Total Atmospheric Transmission Between Summer and Winter in Athens, Greece ............................... 1.396 The Meteorological Radiation Model ....................................................................................................................................................
1-406
Variability of Atmospheric Turbidity in Athens, Greece ........................................................................................................................
1-400
K. Kanayama Analysis of Thermal Performance on an Air-Type Solar Collector with 2- Glass Using Carbon Fiber Sheet as Collecting Material .................................................................................................................................................................................
111-35 B. Karlsson Cooling of PV Modules Equipped with Low Concentrating CPC Reflectors ..................................................................................... 111400
J. Karni A Multistage Solar Receivers: The Route to High Temperature .......................................................................................................... 1-258 The TROF (Tower Reflector with Optical Fibers): a New Degree of Freedom for Solar Energy Systems ........................................ 1-266 T. Kashiwagi Nonimaging Fresnel Lens Concentrators for Photovoltaic Applications ............................................................................................ 111.358 B. D. Katsoulis Variability of Atmospheric Turbidity in Athens, Greece ........................................................................................................................ 1-400 N. D. Kaushika Performance of Transparently Insulated Solar Passive Hot Water Systems .................................................................................... 111-203 M. J. Kearny Comprehensive Approach for the Estimation of Outdoor Performance of Amorphous Silicon Photovoltaic Devices ....................... 1-129 Investigation of the Back Contact of Cadmium Telluride Solar Cells .................................................................................................. 1-124
N. P. Kekelidze Solar Energy Resources and Their Application Perspectives in Georgia (Using Semiconductive Photovoltaic Cells) ..................... 1-185 R. Kemme Efficiency Improvement of Parabolic Trough Collectors by Means of Additional End Reflectors .................................................... 111-337
P. Keuber The Development and Testing of Small Concentrating PV Systems ................................................................................................ 111-409 K. Y. Khouzam Applications of Dispersed Generation Systems in the Utility Network ................................................................................................ 1-192 M. Kiermasch Inclination Dependency of Flat Plate Collector Heat Losses ............................................................................................................... 111-72 R, Kistner A Guide for Financial Feasible Large-Scale Solar Thermal IPP's ...................................................................................................... 11-455
F. H. Klotz The Development and Testing of Small Concentrating PV Systems ................................................................................................ 111-409 B. Knopf Modelling of a Thermosyphonally Driven Discharge Unit of a Storage Tank .................................................................................... 111-197 M. K6hl Research and Development of Solar Collectors Fabricated From Polymeric Material ....................................................................... 111-40
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ISES Solar World Congress 1999, Volume III
N. Koshkin Renewable Energy Sources Utilization in Russia ............................................................................................................................... 11-395 G. Koury Costa Thermodynamic Study of a Regenerative Water Distiller .................................................................................................................. 111-211 A. Krainer Advanced Fuzzy Control of the Temperature in the Test Chamber .................................................................................................. 111-304 M. Krimer Solar Absorber System for Preheating Feeding Water District Heating Nets ..................................................................................... 111-90 S. Krauter Combined Photovoltaic and Solar Thermal Systems for Facade Integration and Building Insulation ................................................ 11-88 Set-Up of a Laboratory for Research and Education in Solar Energy in Rio De Janeiro .................................................................. 11-288 The Use of Solar Energy: Considerations for Calculations of Greenhouse Gas Reduction by Photovoltaics .................................. 11-375 H. Kreetz A Solar Driven Ammonia Based Thermochemical Energy Storage System ....................................................................................... 1-523 Theoretical Analysis and Experimental Results of a 1KW Chemsynthesis Reactor for a Solar Thermochemical Energy Storage System .........................................................................................................................................................................
1-515
A. Kribus A Multistage Solar Receivers: The Route to High Temperature .......................................................................................................... 1-258 The TROF (Tower Reflector with Optical Fibers): a New Degree of Freedom for Solar Energy Systems ........................................ 1-266 P. Krohn Cooling of PV Modules Equipped with Low Concentrating CPC Reflectors ..................................................................................... 111-400 P. Kronthaler In Situ Short -Term Test for Large Solar Thermal Systems .............................................................................................................. 111-126 A. I. Kudish Characterization and Inter-Comparison of the Global and Beam Radiation Measured at Three Sites in the Southem Region of Israel by Statistical Analysis ................................................................................................................................................. 1-419 The Performance and Analysis of a Multiple - Effect Solar Still Utilizing Solar and/or Waste Thermal Energy .............................. 111-216 Performance and Analysis of a Multiple Effect Solar Still Utilizing an Intemal Multi - Tubular Heat Exchanger for Thermal Energy Recycle ..................................................................................................................................................................... 111-226 Research and Development of Solar Collectors Fabricated From Polymeric Material ....................................................................... 111-40
R. Kumar Thermal Modelling and Performance Prediction of Drying Processes under Open-Sun-Drying ...................................................... II1-170 S. Kumar Dissemination of Renewable Energy in Developing Countries: Experiences of a Regional Project in Asia. .................................... 11-489 Experimental Studies on a Hybrid Dryer............................................................................................................................................. 111-434 A Web Based Course for Leaming Solar Thermal Processes ............................................................................................................ 11-293
D. Kwiecien Hybrid Solar Collectors for Microclimate Forming System ..................................................................................................................... 111-3 A. Kyprianou Bioclimatic Designs for the New University of Cyprus Campus. 1st Competition: Facilities for Science and Technology ............................................................................................................................................................................................... 11-3
L S. Lalot Study of a Mixed (Water Or Air) Solar Collector ................................................................................................................................... 111-50 A. Lampinen The School Physics Program of the Finnish Physical Society ............................................................................................................ 11-298
M. Lando An AsUgmatic Corrected Target Aligned Solar Concentrator for Solid State Laser Pumping ......................................................... 111-354
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R. W. Lang Solar Optical and Infrared Radiative Properties of Transparent Polymer Films ................................................................................. 1-489 M. A. Lara Integrated Thermal Improvements for Greenhouse Cultivation in the Central Part of Argentina ..................................................... 111-120 Y. Lech6n Strategic Analysis of the Integration of a Biomass Power Plant in Spain ........................................................................................... 11-472 J. L~d6 The Solar Chemistry Program of the International Energy Agency's Implementing Agreement Solarpaces ...................................... 1-64
G. Leftheriotis Simulation of a Test Cell Dynamic Behavior for the Evaluation of Glazing Thermal Properties ........................................................ 1-504 Systematic Study of Electrochromic Devices for Optical Applications ................................................................................................ 1-495
R. Lemoine Combined Photovoltaic and Solar Thermal Systems for Facade Integration and Building Insulation ................................................ 11-88 K. Lenic An Analysis of Phase Change Heat Transfer in a Solar Thermal Energy Store ............................................................................... 111-484 J. Leon Development and Test of an Equipment to Replace Broken Glass Envelops of Receiver Tubes in Parabolic Trough Collectors .............................................................................................................................................................................................
111-394
Efficiency Improvement of Parabolic Trough Collectors by Means of Additional End Reflectors .................................................... 111-337 M. A. Leon Experimental Studies on a Hybrid Dryer. ............................................................................................................................................
111-434
A Study on Improved Institutional Biomass Stoves .............................................................................................................................
11-484
G. Lesino Brackish Water Destillation with Plane Microporous Membranes Driven by Temperature Difference ............................................ 111-261 T. Leukefeld Research and Development of Solar Collectors Fabricated From Polymeric Material ....................................................................... 111-40 E. S. Leus An Imaginative Environment- Responsive Laboratory Building in the Harsh Climate of Botswana .................................................... 11-92 The New Headquarters for Botswana Technology Centre: Innovative Technologies in a Hot- Dry Southem African Climate .................................................................................................................................................................................................... 11-97 R. Leutz Nonimaging Fresnel Lens Concentrators for Photovoltaic Applications ............................................................................................ 111-358 Z.-Y. Liao Eady Results on the Effectiveness of Natural Ventilation at Verbena Height - a High Rise, High Density Housing Development in Hong Kong ................................................................................................................................................................... 11-83 B. Linyekin An Astigmatic Corrected Target Aligned Solar Concentrator for Solid State Laser Pumping ......................................................... 111-354 T. H. Lin Solar-Photo-Voltaic/ThermaI-Cogeneration Collector .......................................................................................................................... B. Litzenburger PV-Hybdd and Thermo-Electric-Collectors ...........................................................................................................................................
1-181 111-76
E. Llobet The Ciclops System: Optimised Management of Middle-Sized-Hybrid Wind-PV-Diesel Plants ...................................................... 111-462 G. Lopez Comparison of Several Parameterized Models for Global Insolation under Cloudy Skies ................................................................. 1-435 L. Lopez Anaerobic Digestion System Installation of Cattle Manure in Two Farms in Puebla, Mexico ............................................................ 1-562 E. Lorenz Short-Term Forecasting of Solar Radiation Based on Satellite Data - an Application of Neural Networks and Markov Random Fields .......................................................................................................................................................................................
1-411
K. Lorenz Elastomer-Metal-Absorber- Development and Application .................................................................................................................
111-83
528
ISES Solar World Congress 1999, Volume III
L. Lori Design of Grid-Connected Inverters .....................................................................................................................................................
1-200
K. Lovegrove A Solar Driven Ammonia Based Thermochemical Energy Storage System ....................................................................................... 1-523 Theoretical Analysis and Experimental Results of a 1KW Chemsynthesis Reactor for a Solar Thermochemical Energy Storage System .........................................................................................................................................................................
1-515
B. LGckehe Short-Term Forecasting of Solar Radiation Based on Satellite Data - an Application of Neural Networks and Markov Random Fields .......................................................................................................................................................................................
1-411
C. Lund Raps in a Virtual World - a Web Based Remote Area Power Supply System ................................................................................... 11-315 Using the World Wide Web for Tertiary Level Renewable Energy Education - the Potential, the Practice and the Possible Problems ................................................................................................................................................................................
11-305
A. Luzzi The Solar Chemistry Program of the Intemational Energy Agency's Implementing Agreement Solarpaces ...................................... 1-64 A Solar Driven Ammonia Based Thermochemical Energy Storage System ....................................................................................... 1-523 M. F. Lyra Photovoltaic Water Pumping Systems Installer Training: a Partnership Experience Between the University and Sao Francisco Hydroelectric Power Plant ...................................................................................................................................................
11-260
V. Lyubanaky Characterization and Inter-Comparison of the Global and Beam Radiation Measured at Three Sites in the Southem Region of Israel by Statistical Analysis .................................................................................................................................................
1-419
M U. Mades Optimization of the Combination of Power Units in Isolated Grids ......................................................................................................
1-342
C. P. Mahandari A Study on Improved Institutional Biomass Stoves .............................................................................................................................
11-484
A. R. Mahoney Comparison of Predicted and Measured Performance of an Integrated Compound Parabolic Concentrator (ICPC) .................... 111-382 S. Malato Compound Parabolic Concentrator Technology Development to Commercial Solar Detoxification Applications ............................ 11-427
A. Manni Comparison of Experimental Measurements and Numerical Simulation in an Atrium Building ........................................................ 11-237 S. Mantelli Distribution of Solar Irradiation in Brazil Derived From Geostationary Satellite Data ......................................................................... 1-362
J. Marchese Brackish Water Destillation with Plane Microporous Membranes Driven by Temperature Difference ............................................ 111-261 M. J. Marcos Distributed Power From Solar Tower Systems: A Mius Approach ...................................................................................................... 1-286 B. Marland An Imaginative Environment- Responsive Laboratory Building in the Harsh Climate of Botswana .................................................... 11-92 The New Headquarters for Botswana Technology Centre: Innovative Technologies in a Hot- Dry Southem African Climate ....................................................................................................................................................................................................
11-97
S. Martin Heat Flow Analysis in Solar Cell Modules ............................................................................................................................................
b134
G. A. Mastekbayeva Experimental Studies on a Hybrid Dryer............................................................................................................................................. 111-434 X. Mathew Electrodeposited CdTe Based Photovoltaic Structures on Metallic Substrates .................................................................................. b142
ISES Solar World Congress 1999, Volume III
529
E. Mathioulakis Statistical Analysis of Solar Collector Test Results in View of Future Certification ............................................................................
111-92
Uncertainty in Solar Collector Testing Results .....................................................................................................................................
111-56
L. Mazzarella Encapsulated Venetian Blind: A New Numerical Model .....................................................................................................................
I1-101
J. C. McClure Electrodeposited CdTe Based Photovoltaic Structures on Metallic Substrates ..................................................................................
1-142
M. Meir Combined Solar Heating and Radiative Cooling System ................................................................................................................... 111-441 Effective Solar Energy Utilisation - More Dependent on System Design Than Solar Collector Efficiency ...................................... 111-265 L. Mei Thermal Performance of Building Integrated Ventilated PV Facades .................................................................................................. Understanding the Potential of Ventilated PV Facades ......................................................................................................................
11-55 11-110
P. Melin Hybrid System Heat Pump - Solar Air Heater for the Drying of Agricultural Products ....................................................................... 11-512 M. MeliB The Solar -Campus JOlich - Actual Status ...........................................................................................................................................
11-156
L. M6nard A Climatological Database of the Linke Turbidity Factor .....................................................................................................................
1-432
J. F, Mendes Compound Parabolic Concentrator Technology Development to Commercial Solar Detoxification Applications ............................ 11-427 M. G. Merino Preparation and Characterization of Sb-Se Thin Films by Electrodeposlted Technique for Photovoltaic Application ...................... 1-120 C. Meurer Phoebus - an Autonomous Supply System with Renewable Energy ..................................................................................................
1-336
T. Miki Optical Properties and Radiative Cooling Power of White Paints .......................................................................................................
1-485
D. R. Mills Transition Strategies for Solar Thermal Power Generation .................................................................................................................
1-272
J. C. Minano The Development and Testing of Small Concentrating PV Systems ................................................................................................
111-409
C. Minero Compound Parabolic Concentrator Technology Development to Commercial Solar Detoxification Applications ............................ 11-427
G. Mink The Performance and Analysis of a Multiple - Effect Solar Still Utilizing Solar and/or Waste Thermal Energy .............................. 111-216 Performance and Analysis of a Multiple Effect Solar Still Utilizing an Internal Multi - Tubular Heat Exchanger for Thermal Energy Recycle .....................................................................................................................................................................
111-226
G. Miron Thermo-Mechanical Design of a Large Compound Parabolic Concentrator for 500 KWt Solar Central Receiver System .................................................................................................................................................................................................
111-367
F. Missfeldt The Kyoto Mechanisms and the Prospect for Renewable Energy Technologies ..............................................................................
11-407
S. Moeller The Production of Zinc by Thermal Dissociation of Zinc Oxide - Solar Chemical Reactor Design .................................................... 1-539
H.-D. Mohring The Development and Testing of Small Concentrating PV Systems ................................................................................................
111-409
M. Mohr Employment Effects of Greenhouse Gas Reduction Strategies .........................................................................................................
11-421
M. Moisson Solar Radiation Modelling in a Complex Enclosure .............................................................................................................................
1-440
A. Morales Compound Parabolic Concentrator Technology Development to Commercial Solar Detoxification Applications ............................ 11-427 G. L. Morrison Characteristics of Vertical Mantle Heat Exchangers for Solar Water Heaters ..................................................................................
111-276
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Modelling Thermosyphon Solar Water Heaters with Mantle Collector-Loop Heat Exchangers ....................................................... 111-236 Packaged Solar Water Heating Technology,Twenty Years of Progress .............................................................................................. 1-42 A. Mor Solar Energy: Time to Get Commercial ............................................................................................................................................... 11-463 M. Motta Encapsulated Venetian Blind: A New Numerical Model ..................................................................................................................... 11-101 D. Mozes Aeration of Fish-Ponds by Photovoitaic Power .................................................................................................................................... 1-175 J. P. Mueller Solar District Heating with a Combined Pit and Duct Storage in the Underground .......................................................................... 111-468 F. Munoz Anaerobic Digestion System Installation of Cattle Manure in Two Farms in Puebla, Mexico ............................................................ 1-562 Municipal Solid Waste Evaluation as a Source of Energy in Mexico City ........................................................................................... 1-566
M. Munschauer Simulation and Analysis of the Performance of Low Concentration PV Modules ............................................................................. 111-370 J. P. Murray The Solar Chemistry Program of the International Energy Agency's Implementing Agreement Solarpaces ...................................... 1-64 Solar Production of Aluminum by Direct Reduction of Ore to AI-Si Alloy ........................................................................................... 1-531 J. Muschaweck Comparison of Predicted and Measured Performance of an Integrated Compound Parabolic Concentrator (ICPC) .................... 111-382 Demonstration of a New Type of ICPC in a Double Effect Absorption Cooling System ................................................................... 111-424 Double-Tailored Microstructures ........................................................................................................................................................... 1-477 M. Musci Compound Parabolic Concentrator Technology Development to Commercial Solar Detoxification Applications ............................ 11-427
N D. Naron Bridging the Gap: Research and Validation of the DST Performance Test Method for CEN and ISO Standards Project Results .....................................................................................................................................................................................
111-245
E. Negro The Meteorological Radiation Model ....................................................................................................................................................
1-406 A. Neskakis The Centre for the Application of Renewable Energies (C.A.R.E.) .................................................................................................... 11-269 Small Scale Photovoltaic R. O. Desalination- Experience in Gran Canada ...................................................................................... 11-527 D. A. Nezer Leveraging the Value of Photovoltaics in Urban Areas through Their Use in Traffic; Lighting and Extedor Shelter ......................... 1-204 M. V. Nguyen Hybrid Solar/Gas Cooling Ejector Unit for a Hospital in Mexico ........................................................................................................ 111-447 J. E. Nielsen Laboratory Testing of Integrated Collector Storage (ICS) Systems with Transparent Insulation Material ....................................... 111-137 N. Zhu Research on a New Type of Heat Pipe Vacuum Tube Solar Water Heater ...................................................................................... 111-253 Z. Ning A Solar Absorption Air-Conditioning Plant Using Heat-Pipe Evacuated Tubular Collectors ............................................................ 111-297
T. Nishio Temperature Dependence of Thermal Conductivity of Advanced Insulators ...................................................................................... 1-482 B. Norton Solar Process Heat: DistillaUon, Drying, Agricultural and Industrial Uses ......................................................................................... 111-256
S. Nutalaya Indirect Evaporative Cooling through a Concrete Ceiling ..................................................................................................................
111-428
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ISES Solar World Congress 1999, Volume III
O J. J. O'Gallagher Comparison of Predicted and Measured Performance of an Integrated Compound Parabolic Concentrator (ICPC) .................... 111-382 Demonstration of a New Type of ICPC in a Double Effect Absorption Cooling System ................................................................... 111-424 Practical Design Considerations for Secondary Concentrators at High Temperatures .................................................................... 111-377 C. Oberdorf In Situ Short -Term Test for Large Solar Thermal Systems ..............................................................................................................
111-126
L. Odicino Brackish Water Destillation with Plane Microporous Membranes Driven by Temperature Difference ............................................ 111-261 M. Ohishi Optical Properties and Radiative Cooling Power of White Paints .......................................................................................................
1-485
T. Okayasu Study on Islanding of Dispersed Photovoltaic Power Systems Connected to Utility Power Grid ....................................................... 1-228 R. Oldach Solar Energy in Social Housing in the UK ...........................................................................................................................................
11-386
J. C. Oliveira Compound Parabolic Concentrator Technology Development to Commercial Solar Detoxification Applications ............................ 11-427 R. C. Orozco Lessons Learned From the Xcalak Village Hybrid System: A Seven Year Retrospective ................................................................. 1-319 J. A. Ortega-Herrera Hybrid System Heat Pump - Solar Air Heater for the Drying of Agricultural Products ....................................................................... 11-512 R. Orths Solar Absorber System for Preheating Feeding Water District Heating Nets ..................................................................................... 111-90 J. Ortner CO2-Mitigation by Solar Conversion of Hydrocarbons .......................................................................................................................
11-340
B. Ostreich Thermo-Mechanical Design of a Large Compound Parabolic Concentrator for 500 KWt Solar Central Receiver System ...............................................................................................................................................................
.................................. 111-367
P M. A. Paalanen The School Physics Program of the Finnish Physical Society ............................................................................................................
11-298
A. Paizuldaeva Solar Absorber System for Preheating Feeding Water District Heating Nets ..................................................................................... 111-90 R. Palumbo The Production of Zinc by Thermal Dissociation of Zinc Oxide - Solar Chemical Reactor Design .................................................... 1-539
J. Pantoja Electrodeposited CdTe Based Photovoltaic Structures on Metallic Substrates .................................................................................. 1-142 S. Panyakeow PV System Connected to a Grid for Home Applications ......................................................................................................................
1-206
E. Papachristou Bioclimatic Designs for the New University of Cyprus Campus. 2nd Competition: Face A, Student Housing ................................. 11-141
S. Papaefthimiou Systematic Study of Electrochromic Devices for Optical Applications ................................................................................................ 1-495
M. Paroncini Comparison of Experimental Measurements and Numerical Simulation in an Atrium Building ........................................................ 11-237
532
ISES Solar World Congress 1999, Volume III
Numerical Simulation and Scale Model Measurements of Daylighting Systems in an Existent Building ......................................... 11-218
D. Patrikios Study of Thin Film Photovoltaic Cells of CdS/CdTe and CdS/Cu_xS ................................................................................................. 1-160 H. G. Pavlopoulos The Meteorological Radiation Model ....................................................................................................................................................
1406
G. Pecheny An Astigrnatic Corrected Target Aligned Solar Concentrator for Solid State Laser Pumping ......................................................... 111-354
J. Peire Multimedia Library of Renewable Energies ......................................................................................................................................... 11-254 E. B. Pereira Distribution of Solar Irradiation in Brazil Derived From Geostationary Satellite Data ......................................................................... 1-362 D. A. Perell6 Brackish Water DestillaUon with Plane Microporous Membranes Driven by Temperature Difference ............................................ 111-261 G. Perentzis Comprehensive Approach for the Estimation of Outdoor Performance of Amorphous Silicon Photovoltaic Devices ....................... 1-129 B. Perers Cooling of PV Modules Equipped with Low Concentrating CPC Reflectors ..................................................................................... 111-400 M. Peter Effective Solar Energy Utilisation - More Dependent on System Design Than Solar Collector Efficiency ...................................... 111-265 V. A. Petrenko A Combined Ejector Cooling and Hot Water Supply System Using Solar and Waste Heat Energy ................................................ 111-188 H.-L. Pham A Study on Improved Institutional Biomass Stoves .............................................................................................................................
11-484
E. Pick Cumulative Energy Demand of Wind Energy and Solar Water Heating Systems ............................................................................. 11-345 G. Piernavieja Small Scale Photovoltaic R. O. Desalination - Experience in Gran Canaria ...................................................................................... 11-527 J. G. Pieters Modelling Solar Energy Input in Greenhouses .................................................................................................................................... I1-117 Solar Radiation Transmittances of Dry and Wet Plastic Films ............................................................................................................ 1470 Diffusive Properties of Dry and Wet Glass and Plastics ...................................................................................................................... 1462 S. M. Pietruszko Barders for Introducing Photovoltaics in Central Europe: Case of Poland ......................................................................................... 11-391 J. Pitarch The Ciclops System: Optimised Management of Middle-Sized-Hybrid Wind-PV-Diesel Plants ...................................................... 111-462 R. Pitz-Paal Development and Test of an Equipment to Replace Broken Glass Envelops of Receiver Tubes in Parabolic Trough Collectors ............................................................................................................................................................................................. 111-394 Efficiency Improvement of Parabolic Trough Collectors by Means of Additional End Reflectors .................................................... 111-337 TRNSYS Software Application for Solar Thermal Power Plants Simulation and Comparative Analysis ........................................... 1-280
W. J. Platzer Solar Optical and Infrared Radiative Properties of Transparent Polymer Films ................................................................................. 1-489 J. Plettner-Marliani The Centre for the Application of Renewable Energies (C.A.R.E.) .................................................................................................... 11-269 Optimization of the Combination of Power Units in Isolated Grids ...................................................................................................... 1-342 Small Scale Photovoltaic R. O. Desalination - Experience in Gran Canada ...................................................................................... 11-527 F. Ploetz CO2-Mitigation by Solar Conversion of Hydrocarbons .......................................................................................................................
11-340
L. Podloweki PV-Hybrid and Thermo-Electdc-collectors ...........................................................................................................................................
111-76
I. V. Pollet Solar Radiation Transmittances of Dry and Wet Plastic Films ............................................................................................................ Diffusive Properties of Dry and Wet Glass and Plastics ......................................................................................................................
1-462
1470
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O. Popel TRNSYS Software Application for Solar Thermal Power Plants Simulation and Comparative Analysis ........................................... 1-280 O. S. Popel Renewable Energy Sources Utilization in Russia ............................................................................................................................... 11-395 M. Poreh B.A.M.A. (Energy Conserving Buildings) Project: Passive Solar Energy in Popular Residential Apartment Buldings in Israel ....................................................................................................................................................................................................
11-75
K. Pottier Optimized Finned Absorber Geometries for Solar Air Heating Collectors .......................................................................................... 111-62
A. S. Pramusito On the Performance of Nine-Year-Old Solar Home Systems and Street Light Systems in Sukatani Village in Indonesia ................................................................................................................................................................................................
1-212
D. K. Prasad Photovoltaics (PV) Modelling for Cities: a GIS-Building Integrated PV (BIPV) Simulations Approach ............................................. 11-147
J. Prats The Ciclops System: Optimised Management of Middle-Sized-Hybrid Wind-PV-Diesel Plants ...................................................... 111-462 H. W. Price A Guide for Financial Feasible Large-Scale Solar Thermal IPP's ...................................................................................................... 11-455 T. Pryor Raps in a Virtual World - a Web Based Remote Area Power Supply System ................................................................................... 11-315
F. A. Psomas Comparative Assessment of the Thermal Behavior of a Planted Roof vs. a Bare Roof in Thessaloniki .......................................... 11-126 G. Purkarthofer Elastomer-MetaI-Absorber - Development and Application ................................................................................................................. 111-83
R H. Ramadan Numerical Model of a Building with Transparent Insulation .................................................................................................................. I1-10 C. M. Rangel Compound Parabolic Concentrator Technology Development to Commercial Solar Detoxification Applications ............................ 11-427 E. Raschke Distribution of Solar Irradiation in Brazil Derived From Geostationary Satellite Data ......................................................................... 1-362
K. S. Reddy Performance of Transparently Insulated Solar Passive Hot Water Systems .................................................................................... 111-203 P. Redi Design of Grid-Connected Inverters ..................................................................................................................................................... 1-200 A. Reinders On the Performance of Nine-Year-Old Solar Home Systems and Street Light Systems in Sukatani Village in Indonesia ................................................................................................................................................................................................ 1-212 C. Reise Daylight and Solar Irradiance Data Derived From Satellite Observations - the Satellight Project .................................................... 11-368 J. Rekstad Combined Solar Heating and Radiative Cooling System ................................................................................................................... 111-441 Effective Solar Energy Utilisation - More Dependent on System Design Than Solar Collector Efficiency ...................................... 111-265 M. Reuss Aeration of Fish-Ponds by Photovoltaic Power .................................................................................................................................... 1-175 Solar District Heating with a Combined Pit and Duct Storage in the Underground .......................................................................... 111-468
R. Reuven A Multistage Solar Receivers: The Route to High Temperature .......................................................................................................... 1-258
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ISES Solar World Congress 1999, Volume III
H. Ries Double-Tailored Imaging Concentrators ............................................................................................................................................. 111-388 Double-Tailored Microstructures ........................................................................................................................................................... 1477 S. B. Riffat Hybrid Solar/Gas Cooling Ejector Unit for a Hospital in Mexico ........................................................................................................ 111-447 K.-J. Riffelmann Development and Test of an Equipment to Replace Broken Glass Envelops of Receiver Tubes in Parabolic Trough Collectors ................................................................................ ,............................................................................................................ 111-394 Efficiency Improvement of Parabolic Trough Collectors by Means of Additional End Reflectors .................................................... 111-337 C. Rigollier A Climatological Database of the Linke Turbidity Factor ..................................................................................................................... 1-432 The Helioclim Project: From Satellite Images to Solar Radiation Maps .............................................................................................. 1-427 E. Rincon A Parabolic Dish Concentrator From a Telecommunication Antenna: Optical and Thermal Study of the Receiver ....................... 111-333 C. C. M. Rindt Modelling of Two - Layer Stratified Stores .......................................................................................................................................... 111490 G. Rockendorf Elastomer-MetaI-Absorber- Development and Application ................................................................................................................. 111-83 Facade Integrated Solar Collectors ..................................................................................................................................................... 11-134 Inclination Dependency of Flat Plate Collector Heat Losses ............................................................................................................... 111-72 PV-Hybrid and Thermo-Electric-Collectors ........................................................................................................................................... 111-76 J. A. Rodriguez Multimedia Library of Renewable Energies ......................................................................................................................................... 11-254 A. Roitgur The Israeli Insulation Standard for Offices ............................................................................................................................................. 11-71 M, Rolloos Bridging the Gap: Research and Validation of the DST Performance Test Method for CEN and ISO Standards Project Results ..................................................................................................................................................................................... 111-245 H. Romero-Paredea Dynamical Model for Solar Still Validated From Optical and Thermal Parameters Measured Experimentally ................................ 111-271 Simulation of Dynamic Behaviour of a Solar Reactor-Receiver as a Function of Solar Concentrated Radiation Profile ..................................................................................................................................................................................................... 1-296 M. Romero Distributed Power From Solar Tower Systems: A Mius Approach ...................................................................................................... 1-286 M. Rommel Research and Development of Solar Collectors Fabricated From Polymeric Material ....................................................................... 111-40 M. R6nnelid Cooling of PV Modules Equipped with Low Concentrating CPC Reflectors ..................................................................................... 111-400 G. Rosengarten Modelling Thermosyphon Solar Water Heaters with Mantle Collector-Loop Heat Exchangers ....................................................... 111-236 P. Roth Aeration of Fish-Ponds by Photovoltaic Power .................................................................................................................................... 1-175 Simulation and Test of Peltier Elements in Connection with Photovoltaic Cells ................................................................................. 1-113 Social-Technical Assessment of Photovoltaic Systems Installed in the First Region of Chile .......................................................... 11-503 S. Rousseau A Solar Bowl in India ...........................................................................................................................................................................
111-405
R. Rubin A Multistage Solar Receivers: The Route to High Temperature .......................................................................................................... 1-258 A. Rubio Lessons Leamed From the Xcalak Village Hybrid System: A Seven Year Retrospective ................................................................. 1-319 M. A. Rubio Comparison of Several Parameterized Models for Global Insolation under Cloudy Skies ................................................................. 1-435 S. Rukugawa Estimation of Direct Solar Irradiance From Global Irradiance by Means of Signal (Wavelet) Analysis ............................................. 1-386
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W. RuB Solar Process Heat with Non-Concentrating Collectors for Food Industry ....................................................................................... 111-131 R. R0ther Uncertainty in Economical Analysis of Solar Water Heating and Photovoltaic Systems .................................................................. 111-141 Demonstrating the Superior Performance of Thin-Film Amorphous Silicon for Building-Integrated Photovoltaic Systems in Warm Climates ...................................................................................................................................................................
1-217
M. Ruzinsky Design of Grid-Connected Inverters .....................................................................................................................................................
1-200
S R. Sdez Strategic Analysis of the Integration of a Biomass Power Plant in Spain ........................................................................................... 11-472 N. K. Sakellariou The Meteorological Radiation Model ....................................................................................................................................................
1-406
M. J. Salhi Combined Photovoltaic and Solar Thermal Systems for Facade Integration and Building Insulation ................................................ 11-88 C. J. C. Salviano Photovoltaic Water Pumping Systems Installer Training: a Partnership Experience Between the University and Sao Francisco Hydroelectric Power Plant ...................................................................................................................................................
11-260
W. Y. Saman Heat Flow Analysis in Solar Cell Modules ............................................................................................................................................
I-134
Roof Integrated Heating and Cooling System .......................................................................................................................................
11-18
L. R. Saravia Multistage Still ......................................................................................................................................................................................
111-155
C. Sattler Compound Parabolic Concentrator Technology Development to Commercial Solar Detoxification Applications ............................ 11-427 F. Scapino Performance of Grid-Connected Photovoltaic Plants ...........................................................................................................................
1-223
A. B. Schaap Solar Heating with Heat Pump and Ice Storage .................................................................................................................................
111-475
C. Schank Pearl Luster Pigments as Overheating Protection in Transparently Insulated Solar Facades ........................................................... 1-453 K. Scharmer The New European Solar Radation Atlas: a Tool for Designers, Engineers and Architects ............................................................. 11-400 W. Schiel Introduction of Dish Stirling Systems in Morocco. Project Proposal for a Moroccan - German Co-Operation ................................. 11-517 H. Schobermayr Solar Optical and Infrared Radiative Properties of Transparent Polymer Films ................................................................................. 1-489 W. SchOlkopf In Situ Short -Term Test for Large Solar Thermal Systems .............................................................................................................. 111-126 S. Schroer Combined Photovoltaic and Solar Thermal Systems for Facade Integration and Building Insulation ................................................ 11-88 M. Schubnell Determination of the Spectral Emittance in the Visible Range at High Temperatures Supported by Laser Heating ........................ 1-372 K. Schwarzer A System for Solar Process Heat for Decentralised Applications in Developing Countries ............................................................. 111-286 L. Seauve Building Design in Tropical Climates Elaboration of the ECODOM Standard in the French Tropical Islands .................................... 11-59 P. J. Sebastian Electrodeposited CdTe Based Photovoltaic Structures on Metallic Substrates .................................................................................. 1-142
ISES Solar World Congress 1999, Volume III
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A. Segal Potential Efficiencies of Solar-Operated Gas Turbine and Combined Cycle, Using the Reflective Tower Optics ............................ 1-302 I. Segal Aeration of Fish-Ponds by Photovoltaic Power ....................................................................................................................................
1-175
D. Serghides Bioclimatic Designs for the New University of Cyprus Campus. 2nd Competition: Face A, Student Housing ................................. 11-141 D. Sergovich The Israeli Insulation Standard for Offices .............................................................................................................................................
11-71
L. Serres Solar Radiation Modelling in a Complex Enclosure .............................................................................................................................
1-440
I. Seter Characterization and Inter-Comparison of the Global and Beam Radiation Measured at Three Sites in the Southem Region of Israel by Statistical Analysis .................................................................................................................................................
1-419
L. J. Shah Characteristics of Vertical Mantle Heat Exchangers for Solar Water Heaters .................................................................................. 111-276 E. Shaviv Design Tools for Bio-Climatic and Passive Solar Buildings ...................................................................................................................
1-53
B. M. Sheinkopf Solar 9 Matters'. A Comprehensive School Unit ...................................................................................................................................
11-322
K. G. Sheinkopf A Solar Energy Education Program .....................................................................................................................................................
11-328
Solar 9 Matters". A Comprehensive School Unit...................................................................................................................................
11-322
K. Shimono Optical Proper'des and Radiative Cooling Power of White Paints .......................................................................................................
1-485
E. Shpilrain Renewable Energy Sources Utilization in Russia ............................................................................................................................... 11-395 TRNSYS Software Application for Solar Thermal Power Plants Simulation and Comparative Analysis ........................................... 1-280 A. H. Md. M. R. Siddique A Study on Improved Institutional Biomass Stoves .............................................................................................................................
11-484
R. Sillmann PV-Hybrid and Thermo-Electric-Collectors ...........................................................................................................................................
111-76
V, P. Singh Electrodeposited CdTe Based Photovoltaic Structures on Metallic Substrates .................................................................................. 1-142 C. M. Sippel Optimized Finned Absorber Geometries for Solar Air Heating Collectors ..........................................................................................
111-62
I. Skrjanc Advanced Fuzzy Control of the Temperature in the Test Chamber ..................................................................................................
111-304
A. A. Slonim Experimental Investigation of Transient Processes and Developing of Equivalent Diagram of a Solar Cell Panel .......................... 1-149 M. A. Slonim Experimental Investigation of Transient Processes and Developing of Equivalent Diagram of a Solar Cell Panel .......................... 1-149 E. Smiley Research and Development on the First AC BIPV Installation in Canada .........................................................................................
11-165
M. Snow Photovoltaics (PV) Modelling for Cities: a GIS-Building Integrated PV (BIPV) Simulations Approach ............................................. 11-147 J. Sold The Ciclops System: Optimised Management of Middle-Sized-Hybrid Wind-PV-Diesel Plants ...................................................... 111-462 S, SopRpan PV System Connected to a Grid for Home Applications ......................................................................................................................
1-206
W. Soto Gomez Thermodynamic Design of a Solar Refrigerator to Conserve Sea Products ..................................................................................... 111-419 Hybrid System Heat Pump - Solar Air Heater for the Drying of Agricultural Products ....................................................................... 11-512
M. Soursos Design and Construction of a Line-Focus Parabolic Trough Solar Concentrator for Electricity Generation .................................... 111-315
ISES Solar World Congress 1999, Volume III
537
F. Spiite Optimization of the Combination of Power Units in Isolated Grids ...................................................................................................... 1-342 The Solar -Campus JQlich - Actual Status ........................................................................................................................................... 11-156 A System for Solar Process Heat for Decentralised Applications in Developing Countries ............................................................. 111-286 F. Spertino Performance of Grid-Connected Photovoltaic Plants ........................................................................................................................... 1-223 A. Spieler Modelling of a Thermosyphonally Driven Discharge Unit of a Storage Tank .................................................................................... 111-197 Passive Solar Office Building: Results of the First Heating Period ..................................................................................................... 11-183 I. Spiewak Solar-Assisted Syngas-Driven Power System ...................................................................................................................................... 1-544 C. Spitta Development and Test of an Equipment to Replace Broken Glass Envelops of Receiver Tubes in Parabolic Trough Collectors .............................................................................................................................................................................................
111-394
L. Stamenic Research and Development on the First AC BIPV Installation in Canada ......................................................................................... 11-165 A. Steinfeld The Production of Zinc by Thermal Dissociation of Zinc Oxide - Solar Chemical Reactor Design .................................................... 1-539 The Solar Chemistry Program of the International Energy Agency's Implementing Agreement Solarpaces ...................................... 1-64 R. Stephan Set-Up of a Laboratory for Research and Education in Solar Energy in Rio De Janeiro .................................................................. 11-288 E. Steudel Pearl Luster Pigments as Overheating Protection in Transparently Insulated Solar Facades ........................................................... 1-453
H. Storas Combined Solar Heating and Radiative Cooling System ................................................................................................................... 111-441 R. Stuhlmann Distribution of Solar Irradiation in Brazil Derived From Geostationary Satellite Data ......................................................................... 1-362 C. Sugarmen Solar-Assisted Syngas-Driven Power System ...................................................................................................................................... 1-544 F. S. Sun Solar-Photo-Voltaic/ThermaI-Cogeneration Collector .......................................................................................................................... 1-181 A. Suzuki Nonimaging Fresnel Lens Concentrators for Photovoltaic Applications ............................................................................................ 111-358 L. Sverdalov An Astigmatic Corrected Target Aligned Solar Concentrator for Solid State Laser Pumping ......................................................... 111-354
T S. Taebeom The Freezing Process of Water Inside a Vertical Cylinder with a Finned Tube ................................................................................ 111-455 K. Tajiri Temperature Dependence of Thermal Conductivity of Advanced Insulators ...................................................................................... 1-482 T. Takashima Heat Recovery Experiments with Concentration Gradient Catalyst Layer in a Solar Chemical Heat Pump ..................................... 1-549 R. Tamme Solar-Assisted Syngas-Driven Power System ...................................................................................................................................... 1-544 T. Tanaka Heat Recovery Experiments with Concentration Gradient Catalyst Layer in a Solar Chemical Heat Pump ..................................... 1-549 S. Tanemura Optical Properties and Radiative Cooling Power of White Paints ....................................................................................................... 1-485 Temperature Dependence of Thermal Conductivity of Advanced Insulators ...................................................................................... 1482
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ISES Solar World Congress 1999, Volume III
E. Taragan A Multistage Solar Receivers: The Route to High Temperature .......................................................................................................... Thermo-Mechanical Design of a Large Compound Parabolic Concentrator for 500 KWt Solar Central Receiver System .................................................................................................................................................................................................
M, Tazawa Opbcal Properties and Radiative Cooling Power of White Paints .......................................................................................................
1-258 111-367 1-485
F. M. T011ez Distributed Power From Solar Tower Systems: A Mius Approach ......................................................................................................
1-286
R. Tepe Elastomer-MetaI-Absorber - Development and Application .................................................................................................................
111-83
T. G. Theodosiou The Influence of a Planted Roof on the Passive Cooling of Buildings ...............................................................................................
11-169
P. Timbrell The Design and Development of a Suitable Universal Means of Terminating, Interconnecting and Packaging Photovoltaic Panels for Present and Future Applications ...................................................................................................................
11-468
T. Tomson Performance of a Cascade of Flat Plate Collectors ...........................................................................................................................
111-292
E. Torijano Jr. Dynamical Model for Solar Still Validated From Optical and Thermal Parameters Measured Experimentally ................................ 111-271 E. Torijano Dynamical Model for Solar Still Validated From Optical and Thermal Parameters Measured Experimentally ................................ 111-271 Simulation of Dynamic Behaviour of a Solar Reactor-Receiver as a Function of Solar Concentrated Radiation Profile .....................................................................................................................................................................................................
1-296
A. Torres Dynamical Model for Solar Still Validated From Optical and Thermal Parameters Measured Experimentally ................................ 111-271 Simulation of Dynamic Behaviour of a Solar Reactor-Receiver as a Function of Solar Concentrated Radiation Profile .....................................................................................................................................................................................................
1-296
D. G. Tourtoura The Influence of a Planted Roof on the Passive Cooling of Buildings ...............................................................................................
11-169
S. Tratzky Compound Parabolic Concentrator Technology Development to Commercial Solar Detoxification Applications ............................ 11-427 C. Triebel Combined Photovoltaic and Solar Thermal Systems for Facade Integration and Building Insulation ................................................ 11-88 P. Trincado Compound Parabolic Concentrator Technology Development to Commercial Solar Detoxification Applications ............................ 11-427 A. Trombe Solar Radiation Modelling in a Complex Enclosure .............................................................................................................................
1-440
A. Trp An Analysis of Phase Change Heat Transfer in a Solar Thermal Energy Store ............................................................................... 111-484 N. K. Tsakiris Comparative Assessment of the Thermal Behavior of a Planted Roof vs. a Bare Roof in Thessaloniki .......................................... 11-126 R. T u n g Early Results on the Effectiveness of Natural Ventilation at Verbena Height - a High Rise, High Density Housing Development in Hong Kong ...................................................................................................................................................................
11-83
H.-R. Tschudi Determination of the Spectral Emittance in the Visible Range at High Temperatures Supported by Laser Heating ........................ 1-372
M. -S. Tse Early Results on the Effectiveness of Natural Ventilation at Verbena Height - a High Rise, High Density Housing Development in Hong Kong ................................................................................................................................................................... 11-83 N, F. Tsagas Design and Construction of a Line-Focus Parabolic Trough Solar Concentrator for Electricity Generation .................................... 111-315 O. Teukamoto Study on Islanding of Dispersed Photovoltaic Power Systems Connected to Utility Power Grid ....................................................... 1-228
ISES Solar World Congress 1999, Volume III
J. Tzschoppe Optimization of the Combination of Power Units in Isolated Grids ......................................................................................................
539
1-342
U H. Unger Employment Effects of Greenhouse Gas Reduction Strategies .........................................................................................................
11421
N. B. Urli Development of Single Junction Cell Amorphous Silicon Solar Photovoltaic Modules with Improved Resistance to Degradation ...........................................................................................................................................................................................
1-154
V K. Vajen Modelling of a Thermosyphonally Driven Discharge Unit of a Storage Tank ....................................................................................
111-197
Passive Solar Office Building: Results of the First Heating Period .....................................................................................................
11-183
Solar Absorber System for Preheating Feeding Water District Heating Nets .....................................................................................
111-90
J. van Berkel Modelling of Two - Layer Stratified Stores .......................................................................................................................................... G. W. E. van Decker
111490
An Empirical Heat Transfer Equation for the Transpired Solar Collectors, Including No-Wind Conditions ....................................... 111-23
W. G. J. van Helden Modelling the Thermal Effects of Semitransparent PV - Modules ...................................................................................................... A. A. van Steenhoven
11-178
Modelling of Two - Layer Stratified Stores ..........................................................................................................................................
111490
Thermal and Electrical Yield of a Combipanel ......................................................................................................................................
111-96
R. J. C. van Zolingen Thermal and Electrical Yield of a Combipanel ...................................................................................................................................... A. Vara Multimedia Library of Renewable Energies ......................................................................................................................................... M. Varela Strategic Analysis of the Integration of a Biomass Power Plant in Spain ........................................................................................... A. Vdzquez
111-96 11-254 11472
Dynamical Model for Solar Still Validated From Optical and Thermal Parameters Measured Experimentally ................................ 111-271 Simulation of Dynamic Behaviour of a Solar Reactor-Receiver as a Function of Solar Concentrated Radiation Profile .....................................................................................................................................................................................................
1-296 M. Vazquez Cies Islands Stand-Alone Photovoltaic Plant: Evaluation and First Results ....................................................................................... 1-234 N. Vennemann Elastomer-Metal-Absorber - Development and Application ................................................................................................................. 111-83 A. Vidal Compound Parabolic Concentrator Technology Development to Commercial Solar Detoxification Applications ............................ 11427 M. Vincent Compound Parabolic Concentrator Technology Development to Commercial Solar Detoxification Applications ............................ 11427 K. Vollmer Thermal Performance of Building Integrated Ventilated PV Facades ..................................................................................................
11-55
K. Voropoulos Statistical Analysis of Solar Collector Test Results in View of Future Certification ............................................................................
111-92
Uncertainty in Solar Collector Testing Results .....................................................................................................................................
111-56
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540
W H.-J. Wagner Cumulative Energy Demand of Wind Energy and Solar Water Heating Systems ............................................................................. 11-345 R. Wagner Passive Solar Office Building: Results of the First Heating Period ..................................................................................................... L. Wald A Climatological Database of the Linke Turbidity Factor ..................................................................................................................... The Helioclim Project: From Satellite Images to Solar Radiation Maps .............................................................................................. G. Wallner Solar Optical and Infrared Radiative Properties of Transparent Polymer Films .................................................................................
11-183 1-432 1-427 1-489
G. Walter Research and Development of Solar Collectors Fabricated From Polymeric Material ....................................................................... 111-40 N. Wamukonya Socio-Economic Impact Assessment of Solar vs. Grid-Electrified Rural Households in Namibia .................................................... 11-351 A. Wang 24,7% Efficient Perl Silicon Solar Cells and Other High Efficiency Solar Cell and Module Research at the University of New South Wales .............................................................................................................................................................................. 1-165 J. M. Warmerdam Solar Heating with Heat Pump and Ice Storage ................................................................................................................................ 111-475 C. K. Weatherby The Development and Testing of Small Concentrating PV Systems ................................................................................................ 111-409 D. Wegner B.A.M.A. (Energy Conserving Buildings) Project: Passive Solar Energy in Popular Residential Apartment Buldings in Israel ....................................................................................................................................................................................................
11-75
S. Weismann Pearl Luster Pigments as Overheating Protection in Transparently Insulated Solar Facades ........................................................... 1-453 S. Weis Thermo-Mechanical Design of a Large Compound Parabolic Concentrator for 500 KWt Solar Central Receiver System .................................................................................................................................................................................................
111-367
H. Wenzel Life Cycle Assessments of Solar Collectors in Denmark .................................................................................................................... 11-333 A. Westerhellweg Daylight and Solar Irradiance Data Derived From Satellite Observations - the Satellight Project .................................................... 11-368 A. Wheldon Solar Energy in Social Housing in the UK ........................................................................................................................................... 11-386 G. R. Whitfield The Development and Testing of Small Concentrating PV Systems ................................................................................................ 111-409 N. Wilmot Raps in a Virtual World - a Web Based Remote Area Power Supply System ................................................................................... 11-315 R. Winston Comparison of Predicted and Measured Performance of an Integrated Compound Parabolic Concentrator (ICPC) .................... 111-382 Demonstration of a New Type of ICPC in a Double Effect Absorption Cooling System ................................................................... 111-424 Practical Design Considerations for Secondary Concentrators at High Temperatures .................................................................... 111-377 B. N. Winther Sustainability and the Use of Solar Energy: Life Cycle Analyses of a Norwegian Solar Dwelling ...................................................... 11-78 H. Wirth Angular Selectivity of Seasonal Sun Protection Devices ......................................................................................................................... I-9 Sun Protection System Based on CPC's with Total Internal Reflection ............................................................................................. 11-226
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N. Wohlgemuth The Kyoto Mechanisms and the Prospect for Renewable Energy Technologies .............................................................................. 11-407 Mechanisms to Support Altemative Technologies for Electricity Generation in the European Union ............................................... 11-413 J. L. Wolpert Hybrid Solar/Gas Cooling Ejector Unit for a Hospital in Mexico ........................................................................................................ 111-447
B. D. Wood Packaged Solar Water Heating Technology,Twenty Years of Progress .............................................................................................. 1-42
Y K. Yamagishi Study on Islanding of Dispersed Photovoltaic Power Systems Connected to Utility Power Grid ....................................................... 1-228 A. Yezioro PASYS- a Knowledge Based CAD System for Determining the Passive Systems for Heating and Cooling ................................... 11-189
P. Yianioulis Study of Thin Film Photovoltaic Cells of CdS/CdTe and CdS/Cu_xS ................................................................................................. 1-160
P. Yianoulis Simulation of a Test Cell Dynamic Behavior for the Evaluation of Glazing Thermal Properties ........................................................ 1-504 Systematic Study of Electrochromic Devices for Optical Applications ................................................................................................ 1-495 A. Yogev High Temperature Solar Energy Conversion Systems .......................................................................................................................... 1-71 K. Yoshimura Optical Properties and Radiative Cooling Power of White Paints ....................................................................................................... 1-485
Z E. Zal7.a Project Diss (Direct Solar) Update on Project Status and Future Planning ........................................................................................ 1-307 D. Zaslavsky "Energy Towers" Producing Electricity and Desalinated Water Without a Collector ............................................................................ 1-79 P. Zazzini Comparison of Experimental Measurements and Numerical Simulation in an Atrium Building ........................................................ 11-237 D. Zevgolis A Comparison of Spectral Total Atmospheric Transmission Between Summer and Winter in Athens, Greece ............................... 1-396 J. Zhao 24,7% Efficient Perl Silicon Solar Cells and Other High Efficiency Solar Cell and Module Research at the University of New South Wales .............................................................................................................................................................................. 1-165
A. Ziegelmann Employment Effects of Greenhouse Gas Reduction Strategies ......................................................................................................... 11-421
O. Zik The TROF (Tower Reflector with Optical Fibers): a New Degree of Freedom for Solar Energy Systems ........................................ 1-266
H. Zinian Research on a New Type of Heat Pipe Vacuum Tube Solar Water Heater ...................................................................................... 111-253 A Solar Absorption Air-Conditioning Plant Using Heat-Pipe Evacuated Tubular Collectors ............................................................ 111-297 H. A. Zondag Thermal and Electrical Yield of a Combipanel ...................................................................................................................................... 111-96 B. Zupancir Advanced Fuzzy Control of the Temperature in the Test Chamber .................................................................................................. 111-304 Y. Zvirin A Comparative Investigation of Radiation Heat Transfer in Transparent Insulation with Differernt Reflection Models ................... 111-102
542
ISES Solar World Congress 1999, Volume III
Index of Papers
24,7% Efficient Perl Silicon Solar Cells and Other High Efficiency Solar Cell and Module Research at the University of New South Wales J. Zhao, A. Wang, M. A. Green ............................................................................................................................................................. 1-165
A Advanced Fuzzy Control of the Temperature in the Test Chamber B. Zupancic, I. Skrjanc, A. Krainer, B. Furlan ..................................................................................................................................... 111-304 Aeration of Fish-Ponds by Photovoltaic Power J. Appelbaum, D. Mozes, I. Segal, M. Barak, M. Reuss, P. Roth........................................................................................................ 1-175 Anaerobic Digestion System Installation of Cattle Manure in Two Farms in Puebia, Mexico F. Munoz, L. Lopez ................................................................................................................................................................................ 1-562 Analysis of Phase Change Heat Transfer in a Solar Thermal Energy Store, An A. Trp, B. Frankovic, K. Lenic ............................................................................................................................................................. 111-484 Analysis of Thermal Performance on an Air-Type Solar Collector with 2- Glass Using Carbon Fiber Sheet as Collecting Material • Jiang, H. Baba, K. Kanayarna, N. Endoh ................................................................................................................................... 111-35 Analysis of Wind Data and Wind Energy Potential in Bsndirma, Turkey, The D. Inan, C. Dundar ................................................................................................................................................................................. 1-329 Analytic Solution of the Differential Equations Describing Heat Flow in Houses, The J. Boland ................................................................................................................................................................................................. 11-27 Angular Selectivity of Seasonal Sun Protection Devices A. Goetzberger, H. Wirth ........................................................................................................................................................................... I-9 Applications of Dispersed Generation Systems in the Utility Network K. Y. Khouzam ....................................................................................................................................................................................... 1-192 Astigmatic Corrected Target Aligned Solar Concentrator for Solid State Laser Pumping, An M. Lando, J. Kagan, B. Linyekin, L. Sverdalov, G. Pecheny, U. Achiam .......................................................................................... 111-354
B Barriers for Introducing Photovoltaice in Central Europe: Case of Poland S. M. Pietruszko .................................................................................................................................................................................... 11-391 Bioclimatic Designs for the New University of Cyprus Campus. 1st Competition: Facilities for Science and Technology D. W. Aitken, A. Kyprianou ....................................................................................................................................................................... 11-3 Bioclimatic Designs for the New University of Cyprus Campus. 2nd Competition: Face A, Student Housing D. Serghides, C. Chrysanb~ou, E. Papachristou ................................................................................................................................. 11-141 Brackish Water Destiliatlon with Plane Microporous Membranes Driven by Temperature Difference L. Odicino, J. Marchese, D. A. PerelkS, G. Lesino .............................................................................................................................. 111-261
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543
Bridging the Gap: Research and Validation of the DST Performance Test Method for CEN and ISO Standards- Project Results D. Naron, M. Rolloos, M. J. Carvalho ................................................................................................................................................. 111-245 Brundtland Solar City Network T. Esbensen .......................................................................................................................................................................................... 11-379 Building Design in Tropical Climates Elaboration of the ECODOM Standard in the French Tropical Islands F. Garde, H. Boyer, R. Celaire, L. Seauve ............................................................................................................................................ 11-59 Building Integration of a Solar Energy Systems A. G. Hestnes ........................................................................................................................................................................................... 1-36 B.A.M.A. (Energy Conserving Buildings) Project: Passive Solar Energy in Popular Residential Apartment Buldings in Israel S. Hassid, M. Poreh, D. Wegner ............................................................................................................................................................ 11-75
C
Calculation of Solar Radiation on Inclined Surfaces in Turkey M. GOnes................................................................................................................................................................................................ 1-380 Centre for the Application of Renewable Energies (C.A.R.E.), The D. Herold, M. Hoffmann, V. Horstmann, A. Neskakis, J. Plettner-Marliani ........................................................................................ 11-269 CFD Heat Transfer Analyses of the Transpired Solar Collector under No-Wind Conditions, A K. G. T. Hollancls, S. J. Arulanandam, E. Brundrett ............................................................................................................................. 111-29 Characteristics of Vertical Mantle Heat Exchangers for Solar Water Heaters L. J. Shah, G. L. Morrison, M. Behnia ................................................................................................................................................. 111-276 Characterization and Inter-Comparison of the Global and BeamRadiation Measured at Three Sites in the Southern Region of Israel by Statistical Analysis V. Lyubansky, A. lanetz, I. Seter, A. I. Kudish, E. G. Evseev .............................................................................................................. 1-419 Ciclops System: Optimised Management of Middle-Sized-Hybrid Wind-PV-Diesel Plants, The E. Llobet, J. Sold, J. Pitarch, J. Prats .................................................................................................................................................. 111-462 Cies Islands Stand-Aione Photovoltaic Plant: Evaluation and First Results M. Vazquez ............................................................................................................................................................................................ 1-234 Climatological Database of the Linke Turbidity Factor, A C. Rigollier, L. Wald, J. Angles, L. M6nard, O. Bauer .......................................................................................................................... 1-432 CO2-Mitigation by Solar Conversion of Hydrocarbons J. Ortner, K.-H. Funken, F. Ploetz ....................................................................................................................................................... 11-340 Combined Ejector Cooling and Hot Water Supply System Using Solar and Waste Heat Energy, A B. J. Huang, V. A. Petrenko ................................................................................................................................................................ 111-188 Combined Photovoltaic and Solar Thermal Systems for Facade Integration and Building Insulation S. Krauter, G. Aradjo, S. Schroer, M. J. Salhi, C. Triebel, R. Lemoine, R. Hanitsch ........................................................................... 11-88 Combined Solar Heating and Radiative Cooling System M. Meir, H. Storas, J. Rekstad ............................................................................................................................................................ 111-441 Comparative Analysis of Daylighting Systems Investigating Illumination and Structure L. I. Filetoth ........................................................................................................................................................................................... 11-213 Comparative Assessment of the Thermal Behavior of a Planted Roof vs. a Bare Roof in Thesseloniki F. A. Psomas, E. A. Eumorfopoulou, N. K. Tsakiris ............................................................................................................................ 11-126 Comparative Investigation of Radiation Heat Transfer in Transparent Insulation with Differernt Reflection Models, A Y. Zvirin, B. Aronov .............................................................................................................................................................................. 111-102 Comparison Between a Simple Solar Collector Accumulator and a Conventional Accumulator A. J. Fasulo, J. Follari ............................................................................................................................................................................ II1-11 Comparison of Experimental Measurements and Numerical Simulation in an Atrium Building P. Zazzini, M. Paroncini, B. Calcagni, A. Manni .................................................................................................................................. 11-237
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ISES Solar World Congress 1999, Volume III
Comparison of Predicted and Measured Performance of an Integrated Compound Parabolic Concentrator (ICPC) J. O'Gallagher, R. Winston, J. Muschaweck, A. R. Mahoney, V. Dudley..........................................................................................111-382 Comparison of Several Parameterized Models for Global Insolation under Cloudy Skies M. A. Rubio, F. J. Batlles, G. Lopez, L. Alados-Arboledas ..................................................................................................................1-435 Comparison of Spectral Total Atmospheric Transmission Between Summer and Winter in Athens, Greece, A H. D. Kambezidis, A. D. Adamopoulos, D. Zevgolis ............................................................................................................................1-396 Compound Parabolic Concentrator Technology Development to Commercial Solar Detoxification Applications J. Blanco, S. Malato, P. Femandez, A. Vidal, A. Morales, P. Trincado, J. C. Oliveira, C. Minero, M. Musci, C. Casalle, M. Brunotte, S. Tratzky, N. Dischinger, K. H. Funken, C. Sattler, M. Vincent, M. Collares-Pereira, J. F. Mendes, C. M. Rangel.......................................................................................................................................................................... 11-427 Comprehensive Approach for the Estimation of Outdoor Performance of Amorphous Silicon Photovoltaic Devices R. Gottschalg, G. Perentzis, D. G. Infield, M. J. Kearny ......................................................................................................................1-129 Controller Design for Injection Mode Driven Direct Solar Steam Generating Parabolic Trough Collectors M. Eck, M. Eberl .................................................................................................................................................................................... 1-247 Cooling of PV Modules Equipped with Low Concentrating CPC Reflectors M. R6nnelid, B. Karlsson, P. Krohn, B. Peters ...................................................................................................................................111-400 Cumulative Energy Demand of Wind Energy and Solar Water Heating Systems H.-J. Wagner, D. GQrzenich, E. Pick...................................................................................................................................................11-345
D
Daylight and Solar Irradiance Data Derived From Satellite Observations - the Satellight Project D. Heinernann, A. Hammer, A. Westerhellweg, H. G. Beyer, C. Reise..............................................................................................11-368 Demonstrating the Superior Performance of Thin-Film Amorphous Silicon for Building-Integrated Photovoltaic Systems in Warm Climates R. R0ther................................................................................................................................................................................................ 1-217 Demonstration of a New Type of ICPC in a Double Effect Absorption Cooling System W. S. Duff, R. Winston, J. J. O'Gallagher, T. Henkel, J. Muschaweck, R. Christiansen, J. Bergquam ...........................................111-424 Design and ConsblJction of a Line-Focus Parabolic Trough Solar Concentrator for Electricity Generation G. C. Bakos, D. AdamopouIos, N. F. Tsagas, M. Soursos ................................................................................................................111-315 Design and Development of a Suitable Universal Means of Terminating, Interconnecting and Packaging Photovoltaic Panels for Present and Future Applications, The P. Timbrell, M. Instance........................................................................................................................................................................ 11-468 Design of Grid-Connected Inverters L. Loft, P. Redi, M. Ruzinsky................................................................................................................................................................. 1-200 Design Tools for Bio-Climatic and Passive Solar Buildings E. Shaviv.................................................................................................................................................................................................. 1-53 Determination of the Spectral Emittance in the Visible Range at High Temperatures Supported by Laser Heating S. Eckhoff, I. AIxneit, M. Schubnell, H.-R. Tschudi ..............................................................................................................................1-372 Developing a Web-Based Learning Environment for Building Energy Efficiency and Solar Design in Hong Kong S. C. M. Hui, K. P. Cheung................................................................................................................................................................... 11-278 Development and Test of an Equipment to Replace Broken Glass Envelops of Receiver Tubes in Parabolic Trough Collectors K. -J. Riffelmann, M. B6hmer, T. Fend, R. Pitz-Paal, C. Spitta, J. Leon ...........................................................................................111-394
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545
Development and Testing of Small Concentrating PV Systems, The G. R. Whitfield, R. W. Bentley, C. K. Weatherby, A. Hunt, H. -D. Mohring, F. H. Klotz, P. Keuber, J. C. Minano, E. Alarte-Garvi .......................................................................................................................................................................................... 111409 Development of a Smart Solar Tank S. Furbo, E. Andersen ......................................................................................................................................................................... 111-160 Development of Single Junction Cell Amorphous Silicon Solar Photovoltaic Modules with Improved Resistance to Degradation N. B. Urli ................................................................................................................................................................................................. 1-154 Dissemination of Renewable Energy in Developing Countries: Experiences of a Regional Project in Asia S. C. Bhattacharya, S. Kumar .............................................................................................................................................................. 11-489 Distributed Power From Solar Tower Systems: A Mius Approach M. Romero, M. J. Marcos, F. M. T~llez, M. Blanco, V. Femandez, F. Baonza, S. Berger ................................................................. 1-286 Distribution of Solar Irradiation in Brazil Derived From Geostationary Satellite Data S. Coils, S. L. de Abreu, P. Couto, S. Mantelli, E. B. Pereira, E. Raschke, R. Stuhlmann ................................................................. 1-362 Distribution of the Ultraviolet Solar Radiation in the Sky of San Luis (Argentina) A. J. Fasulo, M. T. Deluigi, E. Crino ...................................................................................................................................................... 1-376 Double-Tailored Imaging Concentrators H. Ries, J. M. Gordon ........................................................ ,................................................................................................................. 111-388 Double-Tailored Microstructures H. Ries, J. Muschaweck ........................................................................................................................................................................ 1477 Duct Selective Volumetric Receiver: Potential for Different Selectivity Strategies and Stability Issues, The X. G. Casals, J. I. Ajona ...................................................................................................................................................................... 111-324 Dynamical Model for Solar Still Validated From Optical and Thermal Parameters Measured Experimentally H. Romero-Paredes, E. Torijano, A. V~.quez, A. Torres, J. J. Ambriz, E. Torijano Jr ..................................................................... 111-271
E
Early Results on the Effectiveness of Natural Ventilation at Verbena Height - a High Rise, High Density Housing Development in Hong Kong H.-M. Ho, Z.-Y. Liao, M.-S. Tse, R. Tsang ......................................................................................................................................... 11-83 Effective Solar Energy Utilisation - More Dependent on System Design Than Solar Collector Efficiency J. Rekstad, L. Henden, A. G. Imenes, F. Ingebretsen, M. Meir, B. Bjerke, M. Peter ........................................................................ 111-265 Efficiency Improvement of Parabolic Trough Collectors by Means of Additional End Reflectors T. H. Fend, J. Leon, P. Binner, R. Kemme, K. -J. Riffelmann, R. Pitz-Paal ...................................................................................... 111-337 Elaboration on the Design and Operation Principles of a Heavy Duty Universal Sunlight Heliodon Assembled From Precision Machining Tools K. P. Cheung, S. L. Chung ..................................................................................................................................................................... 11-38 Elastomer-MetaI-Absorber- Development and Application G. Rockendorf, B. Bartelsen, N. Vennemann, R. Tepe, K. Lorenz, G. Purkarthofer .......................................................................... 111-83 Electrodeposited CdTe Based Photovoltaic Structures on Metallic Substrates X. Mathew, P. J. Sebastian, J. Pantoja, A. P. Grif6, M. E. Calixto, J. C. McClure, V. P. Singh ......................................................... 1-142 Empirical Heat Transfer Equation for the Transpired Solar Collectors, Including No-Wind Conditions, An K. G. T. Hollands, G. W. E. van Decker ................................................................................................................................................ 111-23 Employment Effects of Greenhouse Gas Reduction Strategies A. Ziegelmann, M. Mohr, H. Unger ...................................................................................................................................................... 11421 Enabling Technology Opens the Way to Large Scale Use of Solar Energy, An L. Y. Bronicki ......................................................................................................................................................................................... 11-444 Encapsulated Venetian Blind: A New Numerical Model L. Mazzarella,M. Motta ........................................................................................................................................................................ 11-101 Energy Savings Related with the Natural and Artificial Light in the Underground Car Parking Areas S. Grignaffini, F. Gugliermetti ............................................................................................................................................................... 11-231
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Energy Towers" Producing Electricity and Desalinated Water Without a Collector D. Zaslavsky ............................................................................................................................................................................................ 1-79 Estimation of Direct Solar Irradiance From Global Irradiance by Means of Signal (Wavelet) Analysis M. Higashi, S. Rukugawa ...................................................................................................................................................................... 1-386 Evacuated Tubular Collector Water Pasteurization Systems W. S. Duff, D. Hodgson ........................................................................................................................................................................ 11-509 Experimental Evaluation of Selective Surfaces in a High Vacuum W. S. Duff, D. Hodgson ......................................................................................................................................................................... 1-451 Experimental Investigation of Transient Processes and Developing of Equivalent Diagram of a Solar cell Panel M. A. Slonim, A. A. Slonim .................................................................................................................................................................... 1-149 Experimental Performance of a PV V-Trough System N. Fraidenraich, E. M. de Souza Barbosa .......................................................................................................................................... 111-342 Experimental Studies on a Hybrid Dryer S. Kumar, G. A. Mastekbayeva, P. C. Bhatta, M. A. Leon ................................................................................................................. 111-434
Facade Integrated Solar Collectors G. Rockendorf, S. Janssen .................................................................................................................................................................. 11-134 Financing of Private Renewable Energy Projects; Hurdles and Opportunities L. Y. Bronicki .............................................................................................................................................................................................. I-3 Freezing Process of Water Inside a Vertical Cylinder with a Finned Tube, The Y. Changsoon, S. Taebeom, K. Jaeyoon ........................................................................................................................................... 111-455
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Guide for Financial Feasible Large-Scale Solar Thermal IPP's, A R. Kistner, M. Geyer, R. Hanitsch, H. W. Price ................................................................................................................................... 11-455
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Heat Flow Analysis in Solar Cell Modules S. Martin, D. R. Hards, W. Y. Saman .................................................................................................................................................... 1-134 Heat Recovery Experiments with Concentration Gradient Catalyst Layer in a Solar Chemical Heat Pump T. Takashima, T. Doi, Y. Ando, T. Tanaka ........................................................................................................................................... 1-549 Heat Transfer through a Duovent Glass with Chemically Deposited Solar Control Coating G. Alvarez, J. J. Flores, C. Cortina ...................................................................................................................................................... 11-199 Helioclim Project: From Satellite Images to Solar Radiation Maps, The C. Rigollier, L. Wald ............................................................................................................................................................................... 1-427 High Temperature Solar Energy Conversion Systems A. Yogev, U. Fisher, A. Erez, J. Blackmon ............................................................................................................................................. 1-71 Hungarian UNESCO Solar Participation Program L. Imre, I. Farkas ................................................................................................................................................................................... 11-382 Hybrid Solar Collectors for Microclimate Forming System G. J. Besler, D. Kwiecien ......................................................................................................................................................................... 111-3
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Hybrid solar/Gas Cooling Ejector Unit for a Hospital in Mexico J. L. Wolpert, M. V. Nguyen, S. B. Riffat ............................................................................................................................................. 111-447 Hybrid System Heat Pump - solar Air Heater for the Drying of Agricultural Products W. Soto Gomez, H. D. Adas-Varela, P. Melin, J. A. Ortega-Herrera, R. Best-Brown ........................................................................ 11-512
Imaginative Environment- Responsive Laboratory Building in the Harsh Climate of Botswana, An E. S Leus, B. Marland ............................................................................................................................................................................. 11-92 In Situ Short -Term Test for Large Solar Thermal Systems N. Benz, T. Beikircher, M. Gut, P. Kronthaler, C. Oberdorf, W. Sch61kopf, H. Dr0ck ....................................................................... 111-126 Inclination Dependency of Flat Plate Collector Heat Losses G. Rockendorf, B. Bartelsen, M. Kiermasch ......................................................................................................................................... 111-72 Indirect Evaporative Cooling through a Concrete Ceiling B. Givoni, S. Nutalaya ......................................................................................................................................................................... 111-428 Influence of a Planted Roof on the Passive Cooling of Buildings, The T. G. Theodosiou, D. G. Aravantinos, D. G. Tourtoura ....................................................................................................................... 11-169 Integrated Thermal Improvements for Greenhouse Cultivation in the Central Part of Argentina J. R. Barral, P. D. Galimberti, A. Barone, M. A. Lara ......................................................................................................................... 111-120 Integration of Communication and Development in the "Alta Valle Di Susa" Project for Solar Energy F. Jarach, G. Del Tin ............................................................................................................................................................................ 11-284 Interannual Variability of Meteorological Parameters in Temperate Climates R, Aguiar, J. Boland ............................................................................................................................................................................... 1-353 Introduction of Dish Stirling Systems in Morocco. Project Proposal for a Moroccan - German Co-Operation V. H&ussermann, W. Schiel ................................................................................................................................................................. 11-517 Investigation of the Back Contact of Cadmium Telluride Solar Cells R. Gottschalg, B. Elsworth, D. G. Infield, M. J. Keamy ........................................................................................................................ 1-124 Israeli Insulation Standard for Offices, The S. Hassid, D. Feuermann, A. Roitgur, D. Sergovich ............................................................................................................................. 11-71 ITO/inP Photovoltaic Devices H. Aharoni ................................................................................................................................................................................................ 1-95
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Kyoto Mechanisms and the Prospect for Renewable Energy Technologies, The N. Wohlgemuth, F. Missfeldt ................................................................................................................................................................ 11-407
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Laboratory Testing of Integrated Collector Storage (ICS) Systems with Transparent Insulation Material M. Bosanac, J. E. Nielsen ................................................................................................................................................................... 111-137 Lessons Learned From the Xcalak Village Hybrid System: A Seven Year Retrospective R. E. Foster, R. C. Orozco, A.-R.-P. Rubio .......................................................................................................................................... 1-319 Leveraging the Value of Photovoltaics in Urban Areas through Their Use in Traffic; Lighting and Exterior Shelter D. A. Nezer ............................................................................................................................................................................................ 1-204
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Life Cycle Assessments of Solar Collectors in Denmark T. D. Jacobsen, H. Wenzel ................................................................................................................................................................... 11-333
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Market for Solar Energy in the Caribbean, The I. Hamksingh ......................................................................................................................................................................................... 11-451 Marketing and Selling Solar Energy Equipment T. Book .................................................................................................................................................................................................. 11-437 Marstsl Central Solar Heating Plant: Design and Evaluation, The A. Heller, J. Dahm ................................................................................................................................................................................ 111-180 Mechanisms to Support Alternative Technologies for Electricity Generation in the European Union N. Wohlgemuth ..................................................................................................................................................................................... 11-413 Medium Scale Solar Crop Dryers for Agricultural Products O. Headley, W. Hinds .......................................................................................................................................................................... 111-175 Meteorological Radiation Model, The H. D. Kambezidis, A. D. Adamopoulos, N. K. Sakellariou, H. G. Pavlopoulos, R. Aguiar, J. Bilbao, A. de Miguel, E. Negro...................................................................................................................................................................................................... 1-406 Method for Establishing a Solar Power Network for Emergency Integrated Cost Effectively in a CHP (Combined Heat and Power) Network, A G. Georgiades, T. Chikahisa, Y. Hishinuma ........................................................................................................................................ 11-361 Modelling of a Thermosyphonally Driven Discharge Unit of a Storage Tank U. Jordan, K. Vajen, B. Knopf, A. Spieler, F. Hilmer ......................................... ' ................................................................................ 111-197 Modelling of Two - Layer Stratified Stores J. van Berkel, C. C. M. Rindt, A. A. van Steenhoven ......................................................................................................................... 111-490 Modelling Solar Energy Input in Greenhouses J. G. Pieters, J. M. Deltour ................................................................................................................................................................... 11-117 Modelling the Thermal Effects of Semitransparent PV - Modules W. G. J. van Helden, B. J. de Boer, J. L. Balenzategui ....................................................................................................................... 11-178 Modelling Thermosyphon Solar Water Heaters with Mantle Collector-Loop Heat Exchangers G. L. Morrison, G. Rosengarten, M. Behnia ....................................................................................................................................... 111-236 Multimedia Library of Renewable Energies M. Castro, A. Colmenar, A. Vara, J. A. Rodriguez, J. Carpio, J. Peire ............................................................................................... 11-254 Multistage Solar Receivers: The Route to High Temperature, A A. Kribus, P. Doron, R. Rubin, J. Kami, R. Reuven, S. Duchan, E. Tamgan ...................................................................................... 1-258 Multistage Still J. Fmnco, L. R. Saravia, S. Esteban ................................................................................................................................................... 111-155 Municipal Solid Waste Evaluation as a Source of Energy in Mexico City F. Munoz, M. Amiga ............................................................... ................................................................................................................ 1-566
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New European Solar Radation Atlas: a Tool for Designers, Engineers and Architects, The K. Scharmer, B. Bourges ...................................................................................................................................................................... 11-400 New Headquarters for Botswana Technology Centre: Innovative Technologies in a Hot- Dry Southern African Climate, The E. S. Leus, B. Marland ............................................................................................................................................................................ 11-97
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New Heat Reflective Polycarbonate Sheet with Spectral Selectivity, A G. Hakim ................................................................................................................................................................................................ 1-461 Nonimaging Fresnel Lens Concentrators for Photovoltaic Applications R. Leutz, A. Suzuki, A. Akisawa, T. Kashiwagi .................................................................................................................................. 111-358 Novel Ventilated Reversible Glazing System, A E. Erell, Y. Etzion .................................................................................................................................................................................. 11-205 Numerical Model of a Building with Transparent Insulation A. K. Athienitis, H. Ramadan .................................................................................................................................................................. 11-10 Numerical Simulation and Scale Model Measurements of Daylighting Systems in an Existent Building F Fillipetti, M. Paroncini, B. Calcagni .................................................................................................................................................. 11-218
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On the Performance of Nine-Year-Old Solar Home Systems and Street Light Systems in Sukatani Village in Indonesia A. Reinders, A. S. Pramusito ................................................................................................................................................................ 1-212 On the Use of the Solar Collection Envelope for Determining the Building Shape I. G. Capeluto .......................................................................................................................................................................................... 11-33 Optical Properties and Radiative Cooling Power of White Paints S. Tanemura, M. Tazawa, P. Jing, T. Miki, K. Yoshimura, K. Igarashi, M Ohishi, K. Shimono, M. Adachi ...................................... 1-485 Optimization of the Combination of Power Units in Isolated Grids F. Sp&te, J. Plettner-Marliani, U. Mades, J. Tzschoppe, H.-J. Haubrich ............................................................................................. 1-342 Optimized Finned Absorber Geometries for Solar Air Heating Collectors K. Pottier, C. M. Sippel, A. Beck, J. Fricke ........................................................................................................................................... 111-62 Overview and Operation Optimization of the Kramer Junction Solar Electric Generating Systems, An R. G. Cable, S. D. Frier ......................................................................................................................................................................... 1-241
Packaged Solar Water Heating Technology,Twenty Years of Progress G. L. Morrison, B. D. Wood ..................................................................................................................................................................... 1-42 Parabolic Dish Concentrator From a Telecommunication Antenna: Optical and Thermal Study of the Receiver, A C. A. Estrada, R. Dorantes, E. Rincon ................................................................................................................................................ 111-333 Passive Cooling System for Remote Locations S. Grignaffini, G. Galli, F. Gugliermetti ................................................................................................................................................... 11-66 Passive Solar Office Building: Results of the First Heating Period R. Wagner, A. Spieler, K. Vajen, S. Beisel .......................................................................................................................................... 11-183 PASYS- a Knowledge Based CAD System for Determining the Passive Systems for Heating and Cooling A. Yezioro .............................................................................................................................................................................................. 11-189 Pearl Luster Pigments as Overheating Protection in Transparently Insulated Solar Facades O. F. Gross, A. Beck, S. Weismann, J. Fricke, E. Steudel, C. Schank ............................................................................................... 1-453 Performance Analyses of a Combined Photovoltaic/Thermal (Pvrr) Collector with Integrated CPC Throughs H. P. Garg, R. S. Adhikari ................................................................................................................................................................... 111-349 Performance and Analysis of a Multiple - Effect Solar Still Utilizing Solar and/or Waste Thermal Energy, The A I. Kudish, E. G. Evseev, L. Horvath, G. Mink ................................................................................................................................. 111-216
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Performance and Analysis of a Multiple Effect Solar Still Utilizing an Internal Multi - Tubular Heat Exchanger for Thermal Energy Recycle G. Mink, L. Horvarth, E. G. Evseev, A. I. Kudish ................................................................................................................................ 111-226 Performance of a Cascade of Flat Plate Collectors T. Tomson ............................................................................................................................................................................................ 111-292 Performance of Grid-Connected Photovoltaic Plants F. Scapino, A. Abete, L. Ferraris, F. Spertino ....................................................................................................................................... 1-223 Performance of Transparently Insulated Solar Passive Hot Water Systems N. D. Kaushika, K. S. Reddy ............................................................................................................................................................... 111-203 Phoebus - an Autonomous Supply System with Renewable Energy C. Meurer, H. Barthels, W. A. Brocke, B. Emonts, H. G. Groehn ........................................................................................................ 1-336 Photovoltaic Water Pumping Systems Installer Training: a Partnership Experience Between the University and Sao Francisco Hydroelectric Power Plant E. M. de Souza Barbosa, C. J. C. Salviano, A. M. Carvalho, M. F. Lyra. ........................................................................................... 11-260 Photovoltaics (PV) Modelling for CRies: a GIS-Building Integrated PV (BIPV) Simulations Approach M. Snow, P. Jones, D. K. Prasad ......................................................................................................................................................... 11-147 Potential Efficiencies of solar-Operated Gas Turbine and Combined Cycle, Using the Reflective Tower Optics A. Segal, M. Epstein .............................................................................................................................................................................. 1-302 Practical Design Considerations for Secondary Concentrators at High Temperatures J. O'Gallagher, R. Winston .................................................................................................................................................................. 111-377 Preparation and Characterization of Sb-Se Thin Films by Electrodeposited Technique for Photovoltaic Application A. M. Femandez, M. G. Merino ............................................................................................................................................................. 1-120 Production of Zinc by Thermal Dissociation of Zinc Oxide - Solar Chemical Reactor Design, The A. Steinfeld, P. Haueter, S. Moeller, R. Palumbo ................................................................................................................................. 1-539 Project Diss (Direct solar) Update on Project Status and Future Planning E. Zarza, K. Hennecke, O. Goebel ....................................................................................................................................................... 1-307 PV-Hybrid and Thermo-Electric-Collectors G. Rockendorf, R. Sillmann, L. Podlowski, B. Litzenburger ................................................................................................................. 111-76 PV System Connected to a Grid for Home Applications S. Panyakeow, S. Sopilpan ................................................................................................................................................................... 1-206
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Raps in a Virtual World - a Web Based Remote Area Power Supply System C. Lund, N. Wilmot, T. Pryor, G. Cole .................................................................................................................................................. 11-315 Recent Developments in Photocatalytic Detoxification and Disinfection Processes of Water and Air D. Y. Goswami ......................................................................................................................................................................................... 1-16 Regulatory and Institutional Measures by the State to Enhance the Deployment of Renewable Energies - the German Experience P.-G. Gutermuth ...................................................................................................................................................................................... 1-29 Renewable Energy Sources Utilization in Russia O. S. Popel, E. Shpilrain, S. Fdd, V. Dobrokhotov, N. Koshkin .......................................................................................................... 11-395 Research and Development of Solar Collectors Fabricated From Polymeric Material A. I. Kudish, E. G. Evseev, M. Rommel, M. K6hl, G. Walter, T. Leukefeld ......................................................................................... 111-40 Research and Development on the First AC BIPV Installation in Canada L. Stamenic, E. Smiley, K. Colbow, J. Jones ....................................................................................................................................... 11-165 Research on a New Type of Heat Pipe Vacuum Tube Solar Water Heater N. Zhu, H. Zinian .................................................................................................................................................................................. 111-253
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School Physics Program of the Finnish Physical Society, The A. Lampinen, R. Hemberg, M. A. Paalanen ........................................................................................................................................ 11-298 Set-Up of a Laboratory for Research and Education in Solar Energy in Rio De Janeiro S. Krauter, R. Stephan, L. Batos, R. Hanitsch ..................................................................................................................................... 11-288 Short-Term Forecasting of Solar Radiation Based on Satellite Data - an Application of Neural Networks and Markov Random Fields E. Lorenz, A. Hammer, D. Heinemann, B. L0ckehe ............................................................................................................................. 1-411 Simulation and Analysis of the Performance of Low Concentration PV Modules M. Munschauer, K. Heumann ............................................................................................................................................................. 111-370 Simulation and Test of Peltier Elements in Connection with Photovoltaic Calls B. Cancino, P. Roth, A. A. AIvarado ..................................................................................................................................................... 1-113 Simulation of a Test Cell Dynamic Behavior for the Evaluation of Glazing Thermal Properties P. Yianoulis, G. Leftheriotis ................................................................................................................................................................... 1-504 Simulation of Dynamic Behaviour of a Solar Reactor-Receiver as a Function of Solar Concentrated Radiation Profile H. Romero-Paredes, E. Torijano, A. Vazquez, A. Torres, J. J. Ambriz ............................................................................................... Small Scale Photovoltaic R. O. Desalination - Experience in Gran Canaria D. Herold, V. Horstmann, A. Neskakis, J. Plettner-Marliani, R. Calero, G. Piernavieja ..................................................................... Social-Technical Assessment of Photovoltaic Systems Installed in the First Region of Chile B. Cancino, P. Roth, E. Galvez, A. Bonneschky ................................................................................................................................. Socio-Economic Impact Assessment of Solar vs. Grid-Electrified Rural Households in Namibia N. Wamukonya ..................................................................................................................................................................................... Solar-Assisted Syngas-Driven Power System C. Sugarmen, M. Epstein, I. Spiewak, U. Fisher, R. Tamme, R. Buck ................................................................................................ Solar-Photo-Voltaic/Thermal-Cogeneration Collector
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B. J. Huang, T. H. Lin, W. C. Hung, F. S. Sun ...................................................................................................................................... 1-181 Solar - Powered Systems for Cooling, Dehumidification and Air - Conditioning G. Grossman ............................................................................................................................................................................................ 1-21 Solar -Campus JOlich - Actual Status, The F. SpSte, M. Melil3, K. Backes .............................................................................................................................................................. 11-156 Solar Absorber System for Preheating Feeding Water District Heating Nets K. Vajen, M. Kr&mer, R. Orths, E. K. Boronbaev, A. Paizuldaeva ....................................................................................................... 111-90 Solar Absorption Air-Conditioning Plant Using Heat-Pipe Evacuated Tubular Collectors, A H. Zinian, Z. Ning ................................................................................................................................................................................. 111-297 Solar Air Collectors - Investigations on Several Series-Produced Collectors H. Fechner, O. Bucek ............................................................................................................................................................................ 111-17 Solar Architecture in Turkey: State-of-the-Art F. N. Demirbilek, D. I. Eryildiz ................................................................................................................................................................ 11-41 Solar Bowl in India, A S. Rousseau, G. Guigan, J. Harper .................................................................................................................................................... 111-405 Solar Chemistry Program of the International Energy Agency's Implementing Agreement Solarpeces, The A. Steinfeld, V. Anikeev, J. Blanco, M. Epstein, K. -H. Funken, J. Ldd6, A. Lussi, J. Murray .............................................................. 1-64 Solar District Heating with a Combined Pit and Duct Storage in the Underground M. Reuss, J. P. Mueller ....................................................................................................................................................................... 111-468 Solar Driven Ammonia Based Thermochemical Energy Storage System, A K. Lovegrove, A. Luzzi, H. Kreetz ......................................................................................................................................................... 1-523
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Solar Energy - a True Option for Rural Electrification in Kenya Hindered by Unfavourable Policy M. M. Agurnba ....................................................................................................................................................................................... 11-481 Solar Energy Education Program, A K. G. Sheinkopf ..................................................................................................................................................................................... 11-328 Solar Energy in Social Housing in the UK R. Oldach, A. Wheldon, A. Francis ...................................................................................................................................................... 11-386 Solar Energy in the Built Environment: the Building as a System Plus the Systems in the Building E. de Oliveira Femandes ........................................................................................................................................................................... I-5 Solar Energy Resources and Their Application Perspectives in Georgia (Using Semiconductive Photovoltaic Cells) N. P. Kekelidze, T. V. Jakhutashvili, E. G. Chachkhiani, G. E. Chachkhiani ......................................................................................1-185 Solar Energy: Time to Get Commercial A. Mor, U. Halperson ............................................................................................................................................................................ 11463 Solar Heating with Heat Pump and Ice Storage A. B. Schaap, J. M. Warmerclam, E. E. Grarnsbergen ....................................................................................................................... 111-475 Solar Matters". A Comprehensive School Unit K. G. Sheinkopf, B. M. Sheinkopf ........................................................................................................................................................ 11-322 Solar Optical and Infrared Radiative Properties of Transparent Polymer Films G. Wallner, H. Schobermayr, R. W. Lang, W. J. Platzer ...................................................................................................................... 1-489 Solar Pond as a Power Source for Desalination U. Fisher ............................................................................................................................................................................................... 111-150 Solar Process Heat with Non-Concentrating Collectors for Food Industry N. Benz, M. Gut, T. Beikircher, W. RuB .............................................................................................................................................. 111-131 Solar Process Heat: Distillation, Drying, Agricultural and Industrial Uses B. Norton .............................................................................................................................................................................................. 111-256 Solar Production of Aluminum by Direct Reduction of Ore to AI-Si Alloy J. P. Murray ............................................................................................................................................................................................ 1-531 Solar Radiation Modelling in a Complex Enclosure A. Trombe, L. Serres, M. Moisson ........................................................................................................................................................ 1-440 Solar Radiation Transmittances of Dry and Wet Plastic Films J. G. Pieters, I. V. Pollet........................................................................................................................................................................ 1-470 Solar Radiation Transmittances of Dry and Wet Plastic Films J. G. Pieters, I. V. Pollet........................................................................................................................................................................ 1-462 Solar Still with Minimum Inclination and Coupled to an Outside Condenser, A D. Inan, A. EI-Bahi ............................................................................................................................................................................... 111-191 Spray-Deposited SnO2-nSi Solar Cells E. Bobeico .............................................................................................................................................................................................. 1-109 Statistical Analysis of Solar Collector Test Results in View of Future Certification K. Voropoulos, E. Mathioulakis, V. Belessiotis ..................................................................................................................................... 111-92 Strategic Analysis of the Integration of a Biomase Power Plant in Spain M. Varela, Y. Lechbn, R. Sdez ............................................................................................................................................................. 11-472 Study of a Mixed (Water Or Air) Solar Collector S. Lalot ................................................................................................................................................................................................... 111-50 Study of Thin Film Photovoltaic Cells of CdSlCdTe and CdS/Cu_xS P. Yianioulis, D. Patrikios ...................................................................................................................................................................... 1-160 Study on Improved Institutional Biomaes Stoves, A S. C. Bhattacharya, A. H. Md. M. R. Siddique, M. A. Leon, H. -L. Pham, C. P. Mahandari ..............................................................11-484 Study on Islanding of Dispersed Photovoltaic Power Systems Connected to Utility Power Grid O. Tsukamoto, T. Okayasu, K. Yamagishi ........................................................................................................................................... 1-228 Sun Protection System Based on CPC's with Total Internal Reflection A. Goetzberger, C. BOhler, H. Wirth ..................................................................................................................................................... 11-226 Sustainability and the Use of Solar Energy: Life Cycle Analyses of a Norwegian Solar Dwelling A. G. Heslnes, B. N. Winther .................................................................................................................................................................. 11-78
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System for Solar Process Heat for Decentralised Applications in Developing Countries, A F. SpAte, B. Hafner, K. Schwarzer ...................................................................................................................................................... 111-286 Systematic Study of Electrochromic Devices for Optical Applications P. Yianoulis, S. Papaefthimiou, G. Leftheriotis ..................................................................................................................................... 1495
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Temperature Dependence of Thermal Conductivity of Advanced Insulators K. Tajiri, T. Nishio, S. Tanemura ........................................................................................................................................................... 1-482 Testing of a Flat Plate Collector with Selective and Nonselective Absorbers That Are Otherwise Identical W. S. Duff, D. Hodgson ........................................................................................................................................................................... 1114 Theoretical Analysis and Experimental Results of a 1 KW Chemsynthesis Reactor for a Solar Thermochemical Energy Storage System H. Kreetz, K. Lovegrove ........................................................................................................................................................................ 1-515 Thermal and Electrical Yield of a Combipanel H. A. Zondag, D. W. de Vries, A. A. van Steenhoven, W. G. J. van Helden, R. J. C. van Zolingen .................................................. 111-96 Thermal Destratification in Small Standard Solar Tanks Due to Mixing During Tapping E. Andersen, S. Furbo ......................................................................................................................................................................... II1-111 Thermal Modelling and Performance Prediction of Drying Processes under Open-Sun-Drying H. P. Garg, R. Kurnar .......................................................................................................................................................................... 111-170 Thermal Performance of Building Integrated Ventilated PV Facades U. Eicker, V. Fux, D. Infield, L. Mei, K. Vollmer ..................................................................................................................................... 11-55 Thermo-Mechanical Design of a Large Compound Parabolic Concentrator for 500 KWt Solar Central Receiver System G. Miron, S. Weis, I. Anteby, B. Ostreich, E. Taragan ....................................................................................................................... 111-367 Thermodynamic Design of a Solar Refrigerator to Conserve Sea Products H. D. Arias-Varela, W. Soto Gomez, O. Castillo-Lopez, R. Best-Brown ........................................................................................... 111419 Thermodynamic Study of a Regenerative Water Distiller G. Koury Costa, N. Fraidenraich ......................................................................................................................................................... 111-211 Transition Strategies for Solar Thermal Power Generation D. R. Mills, C. J. Dey ............................................................................................................................................................................. 1-272 Transportation of Electricity Production L. Y. Bronicki, I. Dostrovsky, U. Fisher ................................................................................................................................................ 11-447 TRNSYS Software Application for Solar Thermal Power Plants Simulation and Comparative Analysis O. Popel, S. Frid, E. Shpilrain, R. Pitz-Paal, K. Hennecke .................................................................................................................. 1-280 TROF (Tower Reflector with Optical Fibers): a New Degree of Freedom for Solar Energy Systems, The A. Kribus, O. Zik, J. Karni ...................................................................................................................................................................... 1-266 Two-Stage Gasification of Wood with Preheated Air Supply: A Promising Technique for Producing Gas of Low Tar Content S. C. Bhattacharya, A. Dutta ................................................................................................................................................................. 1-557
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Uncertainty in Economical Analysis of Solar Water Heating and Photovoltaic Systems S. Colle, S. L. de Abreu, R. RQther..................................................................................................................................................... 111-141 Uncertainty in Solar Collector Testing ResuRs E. Mathioulakis, K. Voropoulos, V. Belessiotis ..................................................................................................................................... 111-56
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ISES Solar World Congress 1999, Volume III
Understanding the Potential of Ventilated PV Facades L. Mei, D. G. Infield, U. Eicker, V. Fux ................................................................................................................................................. Use of Solar Energy: Considerations for Calculations of Greenhouse Gas Reduction by Photovoltaics, The S. Krauter .............................................................................................................................................................................................. Use of the Electronic Book "Building Thermal Analysis" in Passive Solar Design and Education A. K. Athienitis ....................................................................................................................................................................................... Using Photovoltaics for Agricultural Processing Activities in Upper Mustang (Nepal) D. Blamont, P. Amado .......................................................................................................................................................................... Using the World Wide Web for Tertiary Level Renewable Energy Education - the Potential, the Practice and
11-110 11-375 11-245 11-495
the Possible Problems C. Lund, P. Jennings ............................................................................................................................................................................ 11-305
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Variability of Atmospheric Turbidity in Athens, Greece H. D. Kambezidis, A. K. Fotiadi, B. D. Katsoulis .................................................................................................................................. 1-400
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Web Based Course for Learning Solar Thermal Processes, A S. Kumar, R. A. Attalage ...................................................................................................................................................................... 11-293 Windows in the Attic: Thermophysical Problems of Inclined Windows A. Donath ................................................................................................................................................................................................ 11-48
