8th International Conference on Turbochargers and Turbocharging
Organising Committee Kian Banisoleiman (Chair) Steve Birnie Henry Tennant Ennio Codan Les Smith Andrew Stapleton Ricardo Martinez-Botas ]oerg Seume Chris Brace ] ohn MardeII
Lloyd's Register EMEA Borg-Warner Automotive Holset Turbochargers ABB Turbo Systems MIRA QinetiQ Imperial College London University of Hannover University of Bath Consultant
8th International Conference on Turbochargers and Turbocharging Institution of Mechanical Engineers Combustion Engines & Fuels Group
• IDGTE
CRC Press Boca Raton Boston New York Washington, DC
WOODHEAD PUBLISHING LIMITED Cambridge England
Published by Woodhead Publishing Limited, Abington Hall, Abington, Cambridge CB 1 6AH, England www.woodheadpublishing.com Published in North America by CRC Press LLC, 6000 Broken Sound Parkway NW, Suite 300, Boca Raton, FL 33487, USA First published 2006, Woodhead Publishing Limited and CRC Press LLC © 2006, Institution of Mechanical Engineers unless otherwise stated The authors have asserted their moral rights. This book and CD-ROM contain information obtained from authentic and highly regarded sources. Reprinted material is quoted with permission, and sources are indicated. Reasonable efforts have been made to publish reliable data and information, but the authors and the publishers cannot assume responsibility for the validity of all materials. Neither the authors nor the publishers, nor anyone else associated with this publication, shall be liable for any loss, damage or liability directly or indirectly caused or alleged to be caused by this book or CDROM. Neither this book, CD-ROM, nor any part thereof may be reproduced or transmitted in any form or by any means, electronic or mechanical, including photocopying, microfilming and recording, or by any information storage or retrieval system, without permission in writing from Woodhead Publishing Limited. The consent of Woodhead Publishing Limited does not extend to copying for general distribution, for promotion, for creating new works, or for resale. Specific permission must be obtained in writing from Woodhead Publishing Limited for such copying. Further terms and conditions concerning the CD-ROM are included on the CD-ROM. Trademark notice: Product or corporate names may be trademarks or registered trademarks, and are used only for identification and explanation, without intent to infringe. British Library Cataloguing in Publication Data A catalogue record for this book is available from the British Library. Library of Congress Cataloging in Publication Data A catalog record for this book is available from the Library of Congress. Woodhead Publishing ISBN-13: 978-1-84569-174-5 Woodhead Publishing ISBN-lO: 1-84569-174-1 CRC Press ISBN-lO: 0-8493-0777-5 CRC Press order number: WP0777 Printed by Antony Rowe Limited, Chippenham, Wilts, England
CONTENTS PART I: COMPRESSORS AND NOVEL INTAKE SYSTEMS 1 Prediction and measurement of turbocharger compressor wheel temperature A Yamagata, S Nagai, K Nakano, and T Kawakubo, Ishikawajima-Harima Heavy Industries Company Limited, Yokohama, Japan
3
2 Turbocharger compressor development for diesel passenger car applications H Chen and J F Yin, Honeywell Turbo Technologies Limited, Skelmersdale, UK
15
3 The reduction of turbocharger whoosh noise for diesel powertrains D Evans, Ford Motor Company Limited and A Ward,Ricardo UK Limited
29
4 The influence of installation parameters on turbocharged automotive engine performance G Capon, A Leong, and T Morris, Ford Motor Company Limited, UK
43
5 Using the centrifugal compressor as a cold air turbine M Miiller, S Sumser, P Fledersbacher, K Rofiler, and K Fieweger, DaimlerChrysler AG, Germany, and H-J Bauer, University of' Karlsruhe, Germany
55
6 Extending the knock limit of a turbocharged gasoline engine via turboexpansion J W G Turner, R J Pearson, and N Milovanovic, Lotus Engineering, UK, and D W Taitt, Loughborough University, UK
69
7 Turbo-cooling applied to light duty vehicle engines CD Whelan and R A Richards, WDL Limited, UK
81
PART II: IMPROVED MODELS FOR CYCLE SIMULATION 8 A one-dimensional model for variable and fixed geometry radial turbines for turbochargers J M Lujan, J R Serrano, C Cervell6, and F J Arnau, Universidad Politecnica de Valencia, Spain, and S Soltani, Renault, France
97
9 Analysis of turbocharger non-adiabatic performance S Shaaban and J R Seume, University of Hannover, Germany
119
10 Part-load performance prediction of turbocharged engines S Shaaban and J Seume, University of Hannover, Germany, R Berndt, Technical University Berlin, presently Ingenieurgesellschaft Auto und Verkehr IA V GmbH, Germany, H Pucher, Technical University Berlin, Germany and H J Linho.fj; Linhoff Engineering, Germany
131
v
PART III: ELECTRO BOOST SYSTEMS 11 Development of electrically assisted turbocharger for diesel engine Y Yamashita, S Ibaraki, and H Ogita, Mitsuhishi Heavy Industries Limited, Tokyo, Japan
147
12 The design and testing of an electrically assisted turbocharger for heavy duty diesel engines o Ryder, Holset Engineering Company Limited, Huddersjield, UK, H Sutter, ATE GmbH, Germany, and L Iaeger, Iveco Motorenforschung AG, Switzerland
157
PART IV: TURBINES 13 A numerical study of the performance characteristics of a radial turbine with varying inlet blade angle L Barr, S W T Spence, and A McNally, Queen's University Belfast, UK
169
14 Experimental study on the performance of a variable geometry mixed flow turbine for automotive turbocharger S Rajoo and R Martinez-Botas, Imperial College London, UK
183
15 Turbocharger turbine performance under steady and unsteady flow: test bed analysis and correlation criteria M Capobianco and S Marelli, University of Genoa, Italy
193
16 Flexible turbocharger turbine test rig MONA VI D Filsinger, G Fitzky, and B Phillipsen, ABB Turbo Systems Limited, Baden, Switzerland
207
17 Active control turbocharger for automotive application: an experimental evaluation A Pesiridis and R Martinez-Botas, Imperial College London, UK
223
PART V: MECHANICAL ASPECTS 18 Thermomechanical analysis of a turbocharger turbine wheel based on CRTcalculations and measurements T Heuer, B Engels, H Heger, and A Klein, BorgWarner Turbo Systems Engineering GmbH, Germany 235 19 Dynamics of mistuned radial turbine wheels X Sheng, D C Clay, and I Allport, Holset Engineering Company Limited, Huddersfield, UK
251
20 Improving analysis capability in order to reduce turbine RCF S T Kitson, D C Clay, D H Brown, R 0 Evans, D M Eastwood and P K Tootill, Holset Engineering Company Limited, Huddersfield, UK
261
VI
21 Axial load capacity of V-section band clamp joints K Shoghi, BorgWarner Turbo Systems, Bradford, UK, S Barrans and P Ramasamy, University of Huddersfield, UK
273
PART VI: ADDITIONAL PAPERS 22 Reliability trends, operating issues and acceptance criteria related to exhaust gas turbochargers used in the marine industry - a classification society view K Banisoleiman and N Rattenbury, Lloyd's Register, London, UK
289
23 A novel method of high efficiency pressure charging A 0 Dye, Epicam Limited, Linton, Cambridgeshire, UK
305
24 Turbine wheel design for Garrett advanced variable geometry turbines for commercial vehicle applications H Chen, Honeywell Turbo Technologies Limited, Skelmersdale, UK
3 17
25 Compact long-route exhaust gas recirculation mixer design and optimization J Yin, N Deschatrettes, 0 Han, and P Renaud, Honeywell Turbo Technologies Limited, Skelmersdale, UK
329
26 Transient performance prediction of the turbocharging system with the variable geometry turbochargers H Uchida, A Kashimoto, and Y Iwakiri, Toyota Central R&D Laboratories Incorporated, Aichi, Japan
341
27 Plain and full floating bearing simulations with rigid shaft dynamics I McLuckie, S Barrett, and B K Teo, Advanced Integrated Solutions Limited, Market Harborough, Leicestershire, UK
351
VII
Prediction and measurement of Turbocharger compressor wheel temperature A. Yamagata, S. Nagai, K. Nakano and T. Kawakubo Ishikawajima-Harima Heavy Industries Co., Ltd., Yokohama, JAPAN
ABSTRACT Conjugate heat transfer (CHT) analysis for a high-pressure ratio turbocharger compressor has been conducted to estimate the temperature distribution of a rotating impeller. CHT analysis have been performed with several thermal boundary conditions, which are temperature around a compressor housing, of a back plate of the impeller and of a turbine shaft. We revealed the effects of these conditions on the compressor impeller temperature. Using a radiation thermometer, temperature measurement of a rotating impeller also has been performed. It was utilized to verify the accuracy of CHT analysis for the temperature prediction of a rotating impeller. Subscripts blp Back plate of impeller c/h Compressor housing exit Compressor exit oil Lubricant oil ref Reference condition shaft Turbine shaft tip Impeller tip
Nomenclature Diameter Mu Rotation Mach number N Rotation speed Q Volumetric flow rate T Temperature Tt Total temperature Zb Number of blades It c Compressor total pressure ratio
1)
1. INTRODUCTION In recent years, the need of Diesel engines for passenger vehicles is increasing in the worldwide and the turbo charging for Diesel engines becomes very important due to the economical and environmental reasons. In order to improve the output power and the fuel consumption and also to suppress the emission of Diesel engines, a high efficiency and high-pressure ratio turbocharger is required. Since a single stage centrifugal compressor is used usually in a turbocharger for a passenger vehicle, high rotational speed is necessary to achieve high-pressure ratio at a turbocharger compressor. Therefore, due to the increase of rotational speed, the compressor wheel, impeller, is exposed to high centrifugal stress. High-pressure ratio also leads the increase of discharged air temperature, and this causes higher material temperature and consequently the impeller material strength decreases. In order to guarantee the lifetime of a turbocharger, it is important to know the impeller metal temperature correctly. However, there is much difficulty in a direct temperature measurement of rotating parts under high rotational speed, because thermal sensors set on
3
the impeller surface cannot bear the high centrifugal force. Therefore, analytical method that can predict the impeller metal temperature at high-pressure ratio accurately is desired. As a practical method to predict the impeller metal temperature, heat transfer analysis assuming a heat transfer coefficient on the impeller surface and fluid temperature near the wall as thermal boundary conditions has been applied. Mukherjee and Baker [1] have calculated the metal temperature distribution of a high-pressure ratio turbocharger compressor wheel using a heat transfer analysis. They applied a heat transfer coefficient derived from empirical method, and also have investigated thermal stress occurred at the impeller together with centrifugal load. In the aspect of taking account of heat transfer between solid walls and fluid simultaneously, conjugate heat transfer (CHT) analysis can be useful to obtain the metal temperature distribution with high accuracy. One of the main issues of CHT analysis is computational cost, i.e. limitation of memory size to run a calculation and computational time to obtain a converged solution. However, recent progress of computer hardware and parallel computing technology has realized a large CHT calculation. Bohn [2] has conducted a CHT analysis for the whole of a turbocharger, including a compressor, a turbine and a bearing housing. He described about the effect of heat transfer from a turbine to a compressor through a bearing on a compressor performance. Heuer [3] also has done CHT calculations of some twin-entry turbine housings with an integrated manifold, and avoided a thermal shock occurred at a turbine scroll by the modification of a turbine housing geometry. In the present study, in order to construct the temperature prediction method of a rotating impeller analytically, CHT analyses for a high-pressure ratio turbocharger compressor with high rotational speed have been conducted. This analysis includes a compressor housing and a back plate of the impeller. Further, the metal temperature of the rotating compressor impeller has been measured using a radiation thermometer and the accuracy of CHT analysis was verified.
2. METHODOLOGY 2.1 Compressor configuration RHG8V turbocharger with variable geometry turbine system, which has been developed in IHI for truck size Diesel engines, is chosen for this study. Figure 1 shows a cross sectional view of the RHG8V turbocharger, including the information of a selected computational domain and temperature measurement locations. This turbocharger compressor has a cavity inside a shroud casing, which is called "Casing Treatment". This is made to enhance the compressor operating range by the flow re-circulation through this cavity, and included in a computational model used in this study. The compressor design parameters are summarized in Table 1. Total pressure ratio of the compressor :rtc becomes 3.5 at 100% test rotation speed, and the discharged air temperature reaches over 200 deg. C. The computational domain is limited only to compressor side, which includes a compressor impeller, a back plate behind the impeller, a compressor housing and a shaft connected to a turbine. A bearing housing and a turbine are not included. At the interface between a compressor and a bearing or between a compressor housing and atmosphere, temperature or heat flux is given as a thermal boundary condition.
4
Temperature measurement of the rotating impeller is performed at the compressor inlet and the back surface of the impeller using an infrared thermo camera and a radiation thermometer, respectively, which are shown in Figure 2. Further, thermo couples are set in stationary parts, which are a compressor housing and a back plate, and stationary parts temperatures are measured and used for the verification and the boundary conditions of CRT analysis. Table 1. Specification of RRG8V turbocharger compressor. 7 +7
Number of blades
Zb
Impeller diameter
Dtip
Impeller tip Mach number (*)
MUtip
Total pressure ratio (*)
ltC
(mm)
92.00 1.40
- - - - - -------
~~
3.50
(*) at 100% test rotation speed. Radiation thenoometer
~----------.------:
! 0>":,,::::::::"1 l~-~-----1Iil&dP""
- --~::.-:--Ir\\~ 5{ ./:.
/1----0.
I
)~
(I
3"1.\JII J~'_J:t.i:~~~k ~ r':/: ~~HJ ~.~
\.~~ - ~~ , , , C:::::'_,L I
ColJ1lressor impeller
I
~ 'f~,~
::'
J:"
i
\
',~, ii,!
'- .. '-). 'I: N\..
----t-e' / \ L~/ -\ \
~--~
r
----,
!
~
11;
-
~\
\
~~--L ~
-T------------------+f------~--
--4~---_____c_
Infrared thenoo caImra
J
!~
________
" '"\ II (I ~ :,r)
! : \ :1/ : \i'~ ~ . -~"I!I -'
• Y ~~.~/r, I ~--L
11orr::=---=iDIC~HL
I
I
~
I· •
(\,'
\ I
•
'~/.
Thenoo couples in a compressor housing
'
",
I
in a back plate
I
-~...'.: . ±) _~
I
I
I
------------------------------, Figure 1. Overview of the computational domain and the measurement points.
5
Radiation
a) Radiation thermometer.
b) Infrared thermo camera.
Figure 2. Thermometers for the compressor wheel temperature measurement. 2.2 Computational model The computational model used in this study is shown in Figure 3. Since a compressor impeller is composed of 7 full-length blades and the same number of short blades, the computational domain is reduced to a 1/7 sector of the annulus to save the computational cost and the periodic boundary condition is applied circurnferentially. Here, though the compressor housing has a non-axisymmetric geometry, a volute region is simulated by the representative geometry of a volute section for the simplicity. At the outer surfaces of solid regions, which are a compressor housing, a back plate of the impeller and a shaft end to a turbine, thermal boundary conditions are applied, i.e. surface temperature or heat transfer coefficient and neighbor temperature are given, and heat fluxes through these boundaries are determined by the heat balance between the interior and the adjacent boundary. Computation is the steady-state calculation performed by FLUENT version 6.2. Computational grid is generated by GAMBIT, a pre-processor of FLUENT, and mainly consists of tetra mesh. Solid and fluid region have about 1.0 million and 1.5 million cells respectively and total number of grids becomes 2.5 million cells. As a flow and heat transfer solver, the segregated solver within FLUENT 6.2 is selected and the Spalart-Allmaras one-equation model determines the turbulent viscosity of the fluid. In order to control stability and convergence of the CHT calculation, the relaxation factor for each equation is adjusted appropriately. Calculation has been done by Intel Xeon multiprocessor machine and continued till mass flow rate, total pressure and total temperature at inlet and outlet boundary and the metal temperature became converged. About 2 days of computational time is required to obtain a converged solution. 2.3 Boundary conditions As flow boundary conditions, uniform total pressure, total temperature and flow angles are fixed at the inlet boundary and static pressure is fixed at the exit boundary. Thermal boundary conditions are summarized in Table 2. CRT analysis has been performed for following four typical conditions and the effect of each condition on the impeller metal temperature are revealed. 1) Outer surfaces are all adiabatic, i.e. there is no heat transfer through the outer surface of the computational domain.
6
2) In order to evaluate the effect of heat transfer through the rotor shaft on the impeller metal temperature, the thermal boundary condition at the shaft end is changed parametrically. At first, it is assumed that the shaft temperature at the compressor side is the same as the lubricant oil temperature. After that, the normalized shaft end temperature is increased to +0.17 and +0.34. Other surfaces are all adiabatic. 3) In order to evaluate the effect of heat transfer through the outer surface of the compressor housing, the thermal condition on this surface is changed. Assuming the heat transfer coefficient at the outer surface of a compressor housing at a fixed value, the neighbor atmosphere temperature is increased from the reference temperature to +0.17 and +0.34. Other surfaces are all adiabatic. 4) In order to evaluate the effect of heat transfer through the back plate surface, the back plate surface temperature is changed parametrically. First, the temperature distribution of the back plate obtained by the stationary part temperature measurement is applied to a solid region of the back plate. After that, the temperature distribution level is changed to both -0.17 and +0.17. Other surfaces are all adiabatic. - No. of grids : Solid 1.0 million : Fluid 1.5 million - Grid type : Tetrahedron
a) Side view of the computational domain
b) Impeller and shroud cavity.
Figure 3. Computational model of CRT analysis. Table 2. Summery of thermal boundary conditions. Case
(1)
(2)
(3)
(4)
LlT shaft
adiabatic
0, +0.17, +0.34
adiabatic
<=
LlTc/h
adiabatic
<=
0, +0.17, +0.34
adiabatic
LlTb/p
adiabatic
<=
<=--_._._----- -0.17,0, +0.17 --
*) Above temperatures are normalized by the reference value.
7
2.4 Temperature measuremeut In the usual turbocharger test, temperatures of compressor air, turbine gas and lubricant oil are measured at the inlet and the outlet of a compressor, a turbine and a bearing to calculate the total performance of a turbocharger. In this study, the additional measurement of stationary parts' temperature is performed at the locations shown in Figure 1 using the thermo couples set in the compressor material. These data are used to validate and calibrate the result of CHT analysis. In order to measure the metal temperature of a rotating impeller, the optical radiation thermometers, shown in Figure 2, are used in this study. The radiation thermometer has a large merit in the ability of no-contact temperature measurement on a solid surface, and has the accuracy within a few degrees if enough calibration is conducted. Since a radiation thermometer detects energy and frequency radiated from a black heated body, it is required that there is no obstacle between the measured surface and a thermometer, and the measured surface must be painted black. The temperature on a hub surface at compressor inlet is measured using the infrared thermo camera, which is set upstream of an open compressor inlet, shown in Figure 2-b. To obtain the back surface temperature, a radiation thermometer probe is inserted through the inclined hole that is drilled at a back plate. For both two thermometers, a calibration has been done at the stationary condition. It is confirmed that the measured temperature using a radiation thermometer and an infrared thermo camera is agreed with that by the contact temperature measurement.
3. RESULTS AND DISCUSSIONS 3.1 Results of CRT analysis Figure 4 shows the temperature distribution of solid regions obtained by CHT analysis at pressure ratio of 3.5. The value of temperature is expressed as a normalized difference from that of reference condition. In this calculation, adiabatic thermal boundary condition is applied, i.e. there is no heat flux through the outer surfaces of the computational domain except for flow inlet and outlet boundary. This result indicates that the highest temperature occurs at the volute and high radius of the back plate. Because of the deceleration of the outflow from the impeller, static pressure and temperature increase gradually at the diffuser, and static temperature of the compressed air reaches the maximum at the compressor exit. As a result of high temperature flow in the volute, heat flows from the fluid into the solid at high radius part, and the solid temperature is increased there. Further, heat conducts through a compressor housing and a back plate, therefore, temperature of stationary parts is greater than that of the impeller. Especially, temperature of the back plate is almost the same as the maximum temperature because the heat flux between a back plate and a bearing housing is not considered in the adiabatic condition, i.e. the cooling effect due to the lubricant oil flowing in a bearing housing is not taken into account. Heat flux distribution on the impeller surface at the same condition is shown in Figure 5. In this figure, positive value means heat inflow from the neighbor fluid and negative value means heat outflow. This result shows that positive heat flux is observed on the back surface of the impeller, i.e. heat flows from fluid into solid here. This positive heat flux at the back surface is mainly caused by the following two reasons. First is that the hot back
8
plate is faced with the impeller back surface, and second is a viscous heating due to the shear flow between a rotating disk and a stationary wall. On the other hand, negative heat flux is observed on the blade and hub surfaces, so the compressor main flow takes heat away from the impeller material. The metal temperature of the impeller is determined as the balance between these two heat fluxes in this calculation. 0.61
•• •• = = •• •• •• •
0.40
• ••• = = • ••• • ••
(T-T<ef)
.
(T - T ref)
IT'd
0.08
•
IT,e!
0.14
a) Whole of computational model
b) Compressor impeller.
Figure 4. Normalized temperature distribution obtained by CHT analysis.
(+)
(+)
Heat in
Heat in
•• ••• ••• •• • •••
••• ••
• ••• • ••
•• •
0.0
•• •
Heat out (-)
a) Blade and hub surfaces.
0.0 Heatout (-)
b) Back surface.
Figure 5. Heat flux distribution on the impeller surface. 3.2 Effects of thermal conditions Additional calculations, in which thermal boundary conditions around a compressor are varied parametrically, are performed as described at section 2.3. Figure 6 shows the change of the impeller metal temperature from that of adiabatic condition, when the shaft end temperature and the atmosphere temperature around a compressor housing are changed, respectively. The values of temperature are normalized using the reference value in the
9
same manner of Figure 4. In this figure, it can be found that a change of impeller temperature is very small, that is less than 0.01, even though the shaft temperature and the atmosphere temperature are changed from base to over 0.3. As a reason for the small effect of the shaft temperature, it is considered that heat transfer through the shaft is very small compared with that on a blade surface and a back surface because of its small conduction area, and this leads the small change of the impeller temperature. Also, though the atmosphere temperature affects the compressor housing temperature, the impeller faces a compressor housing by small area at the blade tip only, and this lead the same result when the shaft end temperature is varied. From these results, it can be said that the effect of the shaft and atmosphere temperature on the impeller metal temperature is small and heat transfer through the turbine shaft and the compressor housing can be ignored at the impeller metal temperature prediction. The change of the impeller temperature, when the normalized back plate temperature is varied from -0.17 to +0.17, is also shown in Figure 6. As described at section 2.3, the temperature distribution measured by thermo couples is applied on a surface of the back plate as a thermal boundary condition. The result shows that the effect of the back plate temperature is not small compared with that of a shaft end and a compressor housing, and normalized temperature changes from -0.025 to 0.027. This is mainly caused because the faced area with the impeller back surface is very large compared with a casing and a shaft. Therefore, in order to predict the impeller temperature correctly, it is very important to estimate the heat transfer between the impeller back surface and the back plate wall. 0.03
--c- dTshaft varied
-0.03 -0.30
-0.20
-0.10
0.00
0.10
----{s-
dTc/h varied
:::J
dTb/p varied
0.20
0.30
0040
0.50
Temperature change at thermal boundaries fl T / Tref Figure 6. Effects of thermal boundary conditions. 3.3 Temperature measurement The relation between the compressor discharged air temperature and the metal temperature of a rotating impeller is shown in Figure 7. They are measured at different compressor flow rate, from choke to surge, and different rotation speed, 80%, 95% and 100% test rotation speed. This figure includes the back plate temperature at the same radius of the impeller diameter, and all values of temperature are normalized by the reference value. This result shows that impeller metal temperature depends on the compressor
10
discharged air temperature strongly and it is found that all temperatures in a compressor show the linear variation with the compressor discharged air temperature. The lubricant oil temperature and turbine gas temperature are measured at the same time, which are also shown in Figure 7. The lubricant oil temperature is almost constant when the rotation speed and the compressor flow rate are changed, except for 80% rotation speed in which inflow lubricant oil is heated intentionally. In this figure, whether the lubricant oil temperature changes or not, it seems that the impeller metal temperature is affected by the discharged air temperature only. The normalized turbine gas temperature is varied with 0.5 at 95% and 100% rotation speed because the combustion gas is supplied to drive the turbine at these conditions. At the lower speed, 80% rotation speed, turbine gas temperature is constant because noncombustion compressed air is used to drive the turbine at this speed. From this result, whether the turbine gas temperature changes enormously or not, it seems that the discharged air temperature only affects the impeller metal temperature. As a reason of this result, it is considered that the lubricant oil flowing in the bearing housing shields the heat transfer from a turbine to a compressor. As a result of above considerations, it is concluded that the metal temperature of a rotating compressor impeller is determined mainly by its discharged air temperature increase due to compression of the suction air, and not affected by the lubricant oil and the turbine gas temperatures. This trend agrees with the result of the parametric CRT calculations described in the previous section. Comparison of the impeller metal temperature between measurement and calculation is shown in Figure 8. Although calculation predicts the impeller temperature slightly lower than measurement at both a back surface and a hub inlet of the impeller, it is confirmed that CRT analysis can predict the impeller metal temperature with accuracy within 0.03 of the normalized temperature. As a reason for the deviation between measurement and calculation, it is considered that the computational mesh near the wall adjacent to the fluid is relatively large to estimate heat transfer, and the wall function works automatically to determine the heat transfer between a fluid and a solid. In order to improve the accuracy of the temperature prediction further, a sufficient grid resolution is required near the wall to resolve the temperature boundary layer.
11
0.60 0.50 0.40
-0-
Back Plate
- 15.
ImpeDer (Back)
--L
ImpeDer (Hub)
- 0-
Turbine Gas
0.30 't
E-<~
--.
0.20
't
E-<~
0.10
E-<
0.00
'-"' Q)
£!
~ ~
~
2.40 2.00
-fr--- Lubricant Oil
1.60
ConDustion gas 1.20 0.80
Non-conDustion gas
croo--o
0.40
~
0.00 0.00
0.10
0.20
0.30
0.40
/tl\~
0.50
0.60
0.70
Compressor discharged air temperature ( Ttexil - Trd ) / Tref Figure 7. Temperature measurement results.
....
fo..t"
0.70
--. .... " E-<~
-
0.50
E-<
0.40
,
• Measurement
0.60
L
Calculation
'-"'
~
0.30
-:;j
.... 0.20
Q)
~
~
0.10 0.00
Compressor Discharged Air
ImpeDer Back Surface
ImpeDer Hub Surface
Figure 8. Comparison between measurement and cak:ulatbn.
12
4. CONCLUSIONS Conjugate heat transfer analysis for a high-pressure ratio turbocharger compressor was conducted to obtain the temperature distribution of a rotating impeller. Temperature measurement using a radiation thermometer was also performed to validate the computational results. The conclusions of this study are as follows. 1) Heat inflow occurs on the back surface of the impeller, which is mainly caused by the hot back plate faced with the impeller back surface and the viscous heating due to the shear flow between a rotating disk and a stationary wall. On the other hand, the compressor main flow takes heat away from the impeller material, and the impeller metal temperature is determined as the balance between these two heat fluxes. 2) Parametric CHT calculation and a temperature measurement reveal that the compressor metal temperatures vary proportionally to the compressor discharged air temperature. The compressor impeller metal temperature is mainly affected by the compressor discharged air temperature, and the effects of a lubricant oil flowing in a bearing housing and a turbine gas temperature are relatively small. 3) Although CHT analysis predicts the impeller temperature slightly lower than the measured value at both a back surface and a hub inlet of the impeller, it is confirmed that CHT analysis can predict the impeller metal temperature with enough accuracy to estimate the lifetime of a turbocharger.
ACKNOWLEDGEMENTS The authors would like to thank the Engineering Department of Vehicular Turbocharger Division in Ishikawajima-Harima Heavy Industries Co., Ltd., who gave the opportunity to conduct this research and to publish the paper, and also the Thermal and Fluid Machine Testing Group in Ishikawajima Inspection & Instrumentation Co., Ltd. for the contribution to the arrangement and operation of the measurement.
REFERENCES [1] Mukherjee, S. and Baker, D., "Thermal Design of High Pressure Ratio Turbocharger Compressor Wheels", SAE-2002-01-0162, 2002. [2] Bohn, D.E. et aI., "Conjugate Flow and Heat Transfer Analysis of a Turbocharger", ISABE-2003-1167,2003. [3] Heuer, T. et aI., "Thermomechanical Analysis of a Turbocharger Based on Conjugate Heat Transfer", GT2005-68059, 2005.
13
Turbocharger compressor development for diesel passenger car applications Hua Chen and lunFei Yin Honeywell Turbotechnologies Ltd., Skelmersdale, UK NOTATION
BPF EI Trim VGC
Blade Passing Frequency Exit to inlet annulus area ratio of compressor wheel times 100 Square of inlet to exit diameter ratio of compressor wheel times 100 Variable geometry compressor
SYSNOPSIS
This paper describes the development of a turbocharger compressor for modern diesel passenger car applications, The requirements of such applications are first discussed. The aerodynamic design of the compressor wheel is then described and the test results of the wheel are presented in non-ported and ported shroud vaneless diffuser housings as well as in ported shroud vane diffuser housings. The difficulty in achieving wide map width with high efficiency at high pressure is highlighted and measures to overcome this difficulty are discussed. INTRODUCTION
The introduction of variable geometry turbines for passenger diesel car turbochargers about a decade ago created a new challenge to turbocharger compressor designers. Instead of remaining constant once the peak torque of the engine has been reached, the compressor delivery pressure continues to increase with engine speed. This coupled with ever stringent vehicle emission regulations and the continuous demand for better driveability of passenger cars has pushed up considerably the pressure ratios of turbocharger compressors. Fig. 1 shows two new (2 litre) engine running lines given to Honeywell Turbotechnologies by customers in 2001 and in 2004. It can be seen that in three years required pressure ratios for engine peak torque point and engine rated power point have both risen by about 10%. With increased pressure ratios, compressor delivery air becomes hotter and this affects engine performance. Higher compressor efficiency is in turn required to reduce the delivery air temperature. To meet the needs of higher pressure ratios and higher choke flow, rotational speed of passenger car turbocharger compressors has been increased steadily and is now at the same level as the larger turbocharger compressors for commercial vehicles. Compressors tend to be less efficient because the friction loss is higher and less stable because of reduced tolerance to incidence at high speeds. This is clearly in conflict with the requirements of higher efficiency and more stable operation of compressors. A conventional flow stabilising measure such as high backsweep angle (Ref. 1) no longer works because it reduces the compressor pressure ratios. Ported shroud compressor housings, where a bypass channel is introduced in compressor housings between the compressor inlet and housing shroud at wheel inducer location to automatically recirculate the low momentum flow accumulated near the inducer shroud to the compressor inlet, can significantly improve the stability of compressors at high speeds 15
and have long been used in commercial diesel turbochargers for this purpose. However, an earlier trial of ported shroud compressor on passenger car turbocharger by Honeywell Turbotechnologies was not successful because of noise issue. Compared with commercial vehicles, the passenger cars have stricter noise requirements. The ported shroud compressor used in the trial increased the noise lever of so called bladepassing-frequency (BPF) noise, a noise with a frequency equal to impeller inducer blade number times impeller rotational frequency, to an unacceptable level. The compressor wheel described in this paper was designed between 2001 and 2002. It was specifically designed for reducing the BPF of ported shroud compressors. It was subsequently tested in various compressor housings including both ported and nonported shroud housings. Both vaneless diffuser housings and vane diffuser housings were also tested. In this paper, the aerodynamic design of the compressor wheel will be described and the test results of the compressor presented and discussed. COMPRESSOR WHEEL DESIGN
The compressor wheel has a tip diameter of 52mm and its major geometrical parameters are given in Table 1. Table 1 Summary of compressor wheel geometry
Tip diameter Number of full blade Number of splitter blade Inducer tip diameter Blade tip width Blade angle at inlet tip Blade angle at exit Axial length Trailing edge (TE)
52mm 11
none 38.56mm (55trim) 3.474mm (55EI) 64deg 48deg I9.5mm 4: 1 ellipse
This compressor wheel was designed with 11 full blades and no splitter blades. This was to reduce the BPF noise. Compressor wheel inducer blades generate blade loadings to convert shaft mechanical energy into air internal energy. Viewed by a stationary observer, the unbalance pressure force on the two surfaces of a blade creates a disturbance travelling at the speed of wheel rotation, hence generating the BPF noise. By increasing the inducer blades from 5 or 6 of a splitter bladed wheel to 11, the blade loading is reduced. In addition, as the frequency of this noise is proportional to the inducer blade number, the doubling of the inducer blades doubles the frequency of the noise, making it less perceptible to human ears. In non-ported shroud housings, the use of splitters often improves compressor map width, but our experiences show that in ported shroud housings, full bladed wheels can produce the same map widths as splitter bladed wheels. The compressor wheel was designed with very high inducer turning at the hub and less so at the shroud. This is to tum the low velocity flow at the hub as soon as it enters the wheel. At the shroud, a balance had to be struck between early turning flow for efficiency and not accelerating flow too much to subsequently cause strong shockwaves. 16
Figs. 2 & 3 show calculated relative Mach number loading near the suction and pressure surfaces at about 90% span for non-ported and ported cases respectively at 100% design speed and same mass flow rate (near diffuser stall when non-ported). Three shockwaves can be seen one at the leading edge, one further in the inducer and one at the trailing edge. The main shock wave in the inducer is very strong suggesting that the flow turning near the shroud is still large. By reducing the pressure gradient in the vicinity of inducer shockwave, the ported shroud greatly reduced strength of this shockwave. The CFD results also showed that the non-ported compressor would have a narrow operating range (not stable) at this speed. It however showed that the compressor could be stable with a reduced tip width at 90% design speed. TEST RESULTS Vaneless diffuser, non-ported housings
Effects of wheel tip width
The compressor was first tested in vaneless diffuser and non-ported housings. Fig. 4 and Fig. 5 show the compressor maps with SSEI and SOEI (about 9% narrower tip width) wheels respectively. The narrower tip width wheel produced a more stable map at high speeds, enabled the 2001 new engine running line to be fitted in. The compressor peak efficiency on the other hand was reduced with reduced tip width because of reduced diffusion in the wheel. Effects of wheel trailing edge shape
Two identical compressor wheels but one with 4: 1 elliptical trailing edge and the other with blunt trailing edge were tested. Resultant compressor maps are given in Figs. 6 & 7. By subtly changing the wheel backsweep angle and reducing the strength of wheel trailing edge shockwave (particularly near the hub where the blade thickness is greater) the elliptical trailing edge has increased compressor efficiency but also made it less stable. Effects of wheel meridional shroud shape
Two identical wheels but with different meridional shroud contours, one with a tighter shroud comer than the other (Fig. 8), were tested and results are given in Figs. 9 and 10. The tighter comer wheel is more stable at higher speeds but is less efficient. This study and studies of the tip width and the trailing edge shape show that the high speed stability of passenger car turbocharger compressors is greatly affected by a small change of compressor wheel geometry, and the compressor stability often comes with penalty in compressor efficiency. Vaneless diffuser, ported shroud housing
The compressor wheel was scaled to S6mm and tested in a ported shroud housing. The result is given in Fig. 11. Because of the stabilising effect of ported shroud, a larger diffuser gap (axial dimension of vaneless diffuser passage) can be used to improve compressor efficiency and a very good efficiency was achieved.
17
The compressor was also stable and had a good map width. In comparison with a 52mm state-of-art, non-ported and splitter bladed Garrett compressor whose map width is superimposed on the map in Fig. 11, the 56mm ported compressor has achieved a similar surge flow as the 52mm non-ported compressor, but with a significantly higher choke flow and pressure ratios (the 56mm map is 8% under-speed). To achieve stability at high speeds, the non-ported compressor sacrificed efficiency and its efficiency was substantially lower than the ported. Vane diffuser, ported shroud housings The compressor wheel was also tested in a Garrett variable geometry compressor (VGC) for passenger car turbochargers. The VGC used a swing vane diffuser to control its stability and also used ported shroud housing. The wheel gave the best results among all the Garrett compressor wheels. Fig. 12 shows the VGC map at a given diffuser position. Because of various clearances in the vane-swinging mechanism, the VGC has lower peak efficiency than the same compressor wheel tested in other type of housings. But these clearances and the ported shroud also produced a stable and wide map, despite the existence of a vane diffuser. Compressor noise on engine test Four different compressors were tested for compressor noise on a customer's engine during engine full load transient acceleration. Compressor exit duct noise levels recorded during the test are shown here in Figs. 13 to 15. In these figures, the x axis gives noise frequencies in kHz, the colours of the maps give the noise level labelled as db on the left-hand side y axis, while the right-hand side y axis gives the instantaneous engine speed in rpm. Fig. 13 is the result of a Garrett VGC with a 5 + 5 bladed compressor wheel in a nonported shroud housing. The inducer induced BPF noise is seen as a red curve visible from about 5000Hz and 1000 engine rpm and raising to the right. The exducer BPF noise is shown as a straight line with frequencies double those of the inducer BPF noise and is weaker than the inducer BPF noise. Fig. 14 shows the same VGC with the 11 fully bladed compressor wheel. The inducer BPF noise, shown as a green coloured curve at the bottom-right comer of the noise map, is much weaker than that of the 5 + 5 bladed wheel. Since both inducer and exducer now have the same blade number, the inducer BPF noise and the exducer BPF noise have the same frequencies and are showing on the map as a single curve. Fig. 15 is the same 11 full bladed wheel with a ported shroud housing. The BPF noise is again very weak. The overall noise level, in comparison with the non-ported 5 + 5 bladed compressor, is also better. CONCLUSIONS Modem passenger car applications require higher turbocharger compressor delivery pressures and wider compressor maps with high efficiency. Compressor tip speed has increased to meet the pressure requirement. This reduces the compressor efficiency and map width and makes compressor performance sensitive to small changes in geometry. It is still possible to meet applications' requirements with non-ported shroud compressor housings but this will be more and more difficult in future. Ported shroud housings greatly improve the compressor stability and make high pressure, wide compressor 18
maps possible. They do, however, increase compressor noise and hinder their application in passenger car turbochargers. The full bladed compressor wheel, as one described in this paper, can reduce the ported shroud compressor noise and in particularly the BPF noise. In ported shroud housings, it can also produce similar if not better performances both in efficiency and in map width to those wheels with splitter blades.
REFERENCES 1.
2. 3.
Hua Chen & William Connor, Turbocharger compressor developments for passenger car gasoline engine applications, i h International conference on turbochargers and turbo charging, IMechE, C602/0 16/2002, 14-15 May 2002. N.A. Cumpsty, Compressor aerodynamics, Addison Wesley Longman Ltd., 1998. David Japikse, Centrifugal compressor design and performance, Concepts ETI, Inc. 1996.
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27
THE REDUCTION OF TURBOCHARGER WHOOSH NOISE FOR DIESEL POWERTRAINS Dominic Evans Ford Motor Company Ltd.
Andrew Ward Ricardo UK Ltd. ABSTRACT
Diesel vehicle buyers in Europe are now able to select from a wide range of desirable products with excellent performance and levels of refinement competitive with gasoline engine products. However, the trend towards higher ratings and increased low speed torque appears to have resulted in increased levels of turbocharger generated flow or 'whoosh' noise. This paper, which is based on a previous publication [1], discusses the phenomenon of turbocharger whoosh and its impact on powertrain [2] and vehicle level NVH [3] including: • • • •
Possible mechanisms for whoosh noise occurrence The effect of turbocharger compressor selection on whoosh noise Mechanisms for the radiation of whoosh noise Potential countermeasures for reducing turbocharger whoosh
Case studies are drawn from a range of relevant powertrain NVH projects to illustrate the turbocharger whoosh noise development process. The paper includes an insight to the future development of predictive tools for the whoosh noise phenomenon. INTRODUCTION
Diesel vehicle buyers in Europe are now able to select from a wide range of desirable products with excellent performance and levels of refinement competitive with gasoline engine products. However, the trend towards higher ratings and increased low speed torque appears to have resulted in increased levels of turbocharger generated flow or 'whoosh' noise. Turbocharger generated flow or 'whoosh' noise is a broadband flow-type noise typically evident under full load acceleration on many turbocharged HSDI engines. It can lead to poor sound quality perception and driver annoyance when present to a sufficiently high level. Figure I shows 4-microphone average sound pressure levels (SPL) for 9 modem HSDI engines tested at full load with a best fit line removed to highlight the deviation from a linear SPL. Under further investigation, the rise in sound pressure level at low engine speeds has been attributed, at least in part, to turbocharger whoosh, thus demonstrating that whoosh noise is present in a significant number of HSm powertrains in the European market.
29
HSDI DIesel EngIne CCIn1*ison 4-IIc Av SPL@FulI Load - Deviation from Unear Best fit 4,------,-----,------,-----,------,-----,------,
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Figure 1 Comparison of 4-microphone average SPL for a range of modern HSDI engines - deviation from linear increase When present, whoosh noise can have a significant impact on powertrain level sound pressure. Figure 2 shows powertrain I-metre microphone SPL at full load during the early stages of the development process on a 4-cylinder HSDI engine. It can be seen that whoosh noise resulted in an increase in overaIllevel of circa 6dB(A) for the 4-microphone average - a significant issue.
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30
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Figure 3 Frequency content of whoosh noise at powertrain level for a full load speed sweep At a powertrain level, whoosh noise is in general broadband, typically from 750Hz to in excess of 10kHz, with the majority of the energy most apparent in the 1.5-3kHz range. For this example, the majority of the energy is present in the 1.6 to 2.6kHz range as can be seen in Figs 2 & 3. In vehicle, whoosh noise is most likely to be encountered by the customer during high load acceleration typically below 2500rev/min or during transient tip-out or tip-in manoeuvres. It is often most perceptible accelerating up hill in high gears where the engine speed remains in the speed range where whoosh is present for longer periods. However, it is not confined to full load conditions, for applications where whoosh is present at full load, it may also be audible at part load. For the example powertrain with data shown above, whoosh noise was present to an unacceptable level under driving conditions above 45% load between 1500 and 2500rev/min and during tip-in / tip-out manoeuvres where a rapid increases or decreases in engine load were applied.
MECHANISM FOR GENERATION AND RADIATION OF WHOOSH NOISE Whoosh noise excitation is due to the generation of high levels of turbulence within the turbocharger compressor. Figure 4 shows the dynamic pressure measured within the compressor outlet duct for an engine exhibiting significant turbo whoosh noise. It can be seen that the broadband features of the whoosh excitation are present within the in-duct measurement and subjective assessment of this data reveals the whoosh character. Most
31
Frequency [Hz]
Figure 4 In-duct measurement of whoosh noise
significantly this demonstrates that turbo whoosh excitation can be measured in duct allowing easier quantification of the level of whoosh with minimal contribution from other noise sources or alternatively during perfonnance and emissions development in a nonanechoic testcell. Figure 5 shows the source-path-receiver model for turbocharger whoosh noise. The primary path for radiation of the whoosh noise is breakout from the intake system ducts, most commonly the compressor inlet and outlet hoses with, in general, the largest contribution from the compressor outlet. This is confinned for a range of engines in Fig 6, which shows the effect of shielding sections of the intake system over the typical whoosh noise operating region. Structurebome Transfer
I. _ _••~
Duct shell raciates v.roosh noise
Vibration structurally transferred to duct
Insufficient duct affenuation allows Ykoosh noise to be radated
Duct resonances COLIlIe "";th excitation to raciate Ykoosh noise
Figure 5 Source-path-receiver model for turbocharger whoosh
32
Average 4-Microphone dB Delta For Intake Shielding Tests for lOOO-3000rev/min Full Load 3.00 , -_ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ __
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Figure 6 Effect of Intake Shielding on 4·Microphone Average SPL The generation of turbulence is commonly thought to be associated with operation close to the surge line on a compressor operating map (Fig 7). This line, usually measured on a steady state gas stand, denotes the point where the compressor is close to blade or stage stall where reverse flow or recirculation of the intake air occurs generating high levels of turbulence. Figure 7 also shows a full load engine operating line on the compressor map with the speed range where high levels of whoosh noise were encountered marked. It is clear that the speed range where whoosh noise was encountered is not the area where the operating line is closest to the surge line. Therefore, simple examination of the operating line versus the surge line is not a useful guide in predicting the level of whoosh noise or assessing the risk of whoosh noise occurring. However it is also known that operation within the region of the compressor map where, for a constant compressor speed, the pressure ratio across the compressor increases with increasing mass air flow results in unstable running and the generation of turbulence [4]. This phenomenon can be referred to as marginal surge or soft surge. This region, determined by inspection of the turbo speed contour lines, is shown on the compressor map in Fig 7a. The map and operating line for an alternative compressor design tested on the same engine is shown in Fig 7b. This alternative design gave significantly reduced levels of turbo whoosh. Again, there is little correlation between proximity to the surge line for the relevant speed range and the level of whoosh noise. It can be seen however, that the operating line for the compressor with reduced whoosh is further away from the marginal surge region showing some correlation with the measured whoosh noise results. Therefore,
33
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Figure 7) Compressor map showing marginal surge region for compressor a) exhibiting high levels of whoosh b) exhibiting low levels of whoosh consideration of the proximity of the operating line to the marginal surge region is a useful guide to assessing the risk of the occurrence of whoosh noise. This approach can be helpful in the early stages of a programme to support compressor selection. In this situation, the operating line would be predicted using a 1 dimensional CFD package such as W AVE [5] for each compressor under consideration. When examining compressor maps, it should be remembered that these are generally measured on a steady state flow bench without vehicle intake systems fitted. It is known that the geometry of the vehicle intake ducting can significantly influence the characteristics of the compressor map. It is therefore recommended that the compressor map be measured with representative intake system for the intended application on the flow bench at the earliest opportunity in a programme.
34
Dynamic Envelope
Corrected Mass Air Flow
Figure 8 WAVE predicted engine operating line showing dynamic content of pressure ratio and mass air flow FUTURE DIRECTION FOR INCREASING THE UNDERSTANDING OF THE WHOOSH MECHANISM Consideration of the proximity of the operating line to the marginal surge region as shown above does not take into account the dynamic properties of the pressure ratio and mass air flow that result from the turbomachinery supplying air to an engine that inducts air for certain periods within its cycle. It can be seen from a W AVE prediction of the dynamic pressure ratio and mass air flow for a 4 cylinder engine shown at certain speeds on the full load operating line (Fig 8), that the dynamic component of the mass air flow is significant compared to its mean value and to the width of a typical compressor map. It should be noted that W AVE includes the effects of conservation of momentum and assumes that the speed contour lines have a gradient of zero with respect to mass air flow to the left of the surge line. Even where margin to the surge line or marginal surge region is present, the dynamic content of the mass flow is sufficiently large to give instantaneous operation in this region potentially resulting in turbulence. Further studies are required to fully understand the link between measured whoosh noise and the dynamic content of pressure and mass flow. Dynamic considerations aside, consideration of the proximity of the mean operating line to the marginal surge region can only be at best an aid in selection of a compressor specification. Determining the marginal surge region for compressor maps where the turbo speed contours are relatively flat with mass airflow will be particularly sensitive to any measurement error encountered during the flow rig testing. More importantly, the approach
35
does not, for example, allow a prediction of the level of whoosh noise to be made. The trend in the European market is for diesel specific ratings to continue to increase with more complex turbocharging including the recent introduction of two-stage turbocharging and future possibilities for variable geometry compressors. Therefore, it will become more important to be able to fine-tune the tradeoffs between whoosh noise, surge and altitude performance and compressor efficiency. The capability to predict whoosh noise excitation to allow the comparison of different compressor designs and different operating regions within the map would aid this process greatly. It is clear that being able to predict the generation of turbulence in a turbomachine requires a 3 dimensional CFD approach and presents a very significant challenge in this area. Work is underway to attempt to understand and resolve some of the associated issues to allow application of these techniques to support the design process. This work includes the prediction of turbulence at the compressor inlet, generated for example by flow separation in the inlet system, as this is known to influence compressor performance. MINIMISING TURBOCHARGER WHOOSH
From consideration of the source-path-receiver model (Fig 5), it can be seen that there are two methods to reduce turbocharger whoosh, either by improving the transfer functions of the transmission paths or by reduction of the source excitation. Both approaches were taken in parallel in this case study. Modifications to the transmission path In order to reduce the whoosh noise through modifications to the transmission path it is necessary to firstly understand which paths contribute most strongly. In this case, this was
performed at powertrain level to understand the paths that had the greatest contribution to radiated noise. It should be remembered that in vehicle, parasitic structurebome paths may exist such as pressure pulsation induced vibration of the intercooler transmitted into the vehicle interior. These were not considered at this phase of the study. Figure 9, shows the baseline condition for the 4-microphone average SPL as measured in a powertrain testcell. It can be seen that lagging of the inlet ducts reduced the SPL by up to 4.5dB(A) effectively removing the peak at low speed. Subjective assessment of the recordings demonstrated a significant reduction or removal of the whoosh noise. This confirmed that the primary path for transmission of whoosh was excitation of and subsequent radiation from the intake components. Lagging of individual hoses in tum, demonstrated that the compressor outlet hose was responsible for the majority of the contribution to I m microphone SPL with some contribution from the compressor inlet hose (Fig 9). This allowed activities to focus foremost on the compressor outlet hose. Two approaches were taken to reduce radiated noise from the compressor outlet hose. Firstly, by attempting to improve the transmission loss through the duct walls by increasing the thickness of the duct material. Secondly, by fitting a broadband resonator as close as possible to the compressor outlet to reduce the downstream pressure pulsations.
36
-5-d~;A~ r
I
T
- - -
-
- -
T------1-------, -I
- - - - - -
I -1- -
- - -
- -
I
I
I
I
-r-::--::-~-::T~:-:::;;-~-~-~-~-~-$r~-~-~-~-~-'j'l
I -.- -
- -
- -
I - -I - -
- -
-
I - - - T - -
- - - - -
I I
-
- -
- -
-
T
----i-------~------,~~~--~----,I
I
-Ba<;eline
'
-- Compressor Out Lagged - Entire Intake Lagged
I
I
,
, ,
:
I
-------~------~-------1-------~-----
,
,
I
--
Compressor In Lagged
-----~-------1-------~-----
I
1000
1500
:
2000
2500
3000
Engine Speed [rev/min]
3500
4000
450
Figure 9 Effect of intake system lagging on 4-microphone SPL
The effect of applying the two countermeasures to the compressor outlet hose in tum is shown in Fig 10. Increasing the thickness of the hose wall by circa 50% gave a reduction in 4-microphone average SPL of 1.5dB(A). A larger reduction of up to 3dB(A) was achieved by fitting an acoustic resonator to the standard thickness duct which was designed to attenuate pressure pulsation based 1st order turbocharger whine in the 1.6 - 3.5kHz range. Further benefits are likely should a specific resonator be tuned to a broader frequency range more appropriate to the whoosh noise, for example I-10kHz.
I
I
I
I
I
I
------~-------T-------~------~-------T-------I------
,
- - - - - - _1- _
,
I
-
I
I
I
I
Repeat baseline
I
-----4-------i-------~------~----
,
I '
- 50% Increase in Comp Outlet Hose Thickness
I
I
I
I
------~-------f-------r------~----
I
tOOO
1500
2000
2500
3000
-
Broadband Resonator
_ _~------~----~
3500
4000
4500
Engine Speed [revlmin]
Figure 10 Effect of compressor outlet hose countermeasures on 4-mic average SPL
37
, I
I
1 5dB(A) :
I
I
I
I
------~--------~-------1--------~-------1--------:--------
;
I I
I I
-,
I
:
:
I I
I I
I
--"1-
I
,
I
I I
---1--------1--------
I
,
___ 1__ ,__ , _, _ .__ J.
,
,
,
,
_ _ _ _ _ _ ._1 __ , __
,
- Turbo 1 - Baseline
-------1--------~-------,----
,
T
,
I
'
I
I .~-
1000
1500
- Turbo 2 - Modified Compressor Trim
I
- Turbo 3 - Mod Diffuser & Trim
,
- Turbo 4 - Larger Compressor
I , ··--1---------1---T
2000
2500
3000
3500
4000
4500
Engine Speed [revlmin]
Fig 11 Effect Of Changes To Compressor Design On 4-Mic Average Spl At Full Load Reducing excitation at source For transient conditions such as tip-in / tip-out manoeuvres, a previous published study has demonstrated that EGR control and VGT vane position can influence the whoosh noise level [6]. To reduce the excitation for steady state conditions, several compressor variants were investigated with the potential to give similar engine performance while allowing full load operation to occur further from the marginal surge region. These compressors were then fitted in tum to the powertrain and assessed in the anechoic testcell. Engine performance, intake manifold pressure and mass airflow were monitored during testing to ensure that engine operation was largely unaffected by the compressor changes. The resulting effect on 4-microphone average SPL is shown in Fig II. The most effective changes were: • •
a reduction in the compressor trim accompanied by a diffuser geometry change a complete change to a larger compressor wheel with appropriate trim and other geometry to suit that wheel
In either case, reductions of up to 3dB(A) in 4-mic average SPL at full load were achieved.
38
Turbocharger 1
Turbocharger 2
..-
Turbocharger 3
Turbocharger 4 .J~~
:..:-w
'-'--'-- '''_L,-I
JJ~tH' I I I J
,
J.JJ.J_
,'ill', J" _1_1_
"
JJnJ ~-!-I I I
'"'1--T-"
-;-""'/-""..-.':;"'0,.1.-0-:: "
~ ,
~~ffl
,... + ... I
f ...; -1-,-1-1· 1'1 I I
r
I- + + I I
...
.-
Fig 12 Compressor maps for the baseline and for the 3 other variants assessed The associated compressor maps for these variants can be seen in Fig 12. It can be seen that there is a general trend in that the further away the operating line is from the marginal surge line the lower the sound pressure level in the whoosh noise speed range.
39
, ,
, ,
------------"-------------c--------------------------c-----------]
I
I
I
I
I
I
I
:
:
:
I
I
I
5dB(A); I
I
,
"
------------~-------------~--------------------------r ------------
, , , -
~
, ,
-
-
, ,
I- _
-e-- Turbo
,
,
------------~--------,
o
20
1 - Baseline ~ Turbo 2 - Modified Compressor Trim - Turbo 3 - Mod Diffuser & Trim - Turbo 4 - Larger Compressor
40 60 % of engine Full Load
80
10
Fig 13 Effect of compressor variant on 4-microphone SPL for a load sweep at 1750rev/min Interestingly, the benefit seen at full load with the compressor change was also seen for loads above 45% where whoosh was originally present, for example at 1750rev/min (Fig 13). Consideration of the operating line showing the load sweep at l750rev/min (Fig 13 Turbocharger 1) shows that this is some distance away from the marginal surge region or surge line. This demonstrates that benefits from changing compressor design can be seen even under operation away from these regions. As is often the case with NVH issues, whoosh noise is not the only concern when selecting a compressor variant. Optimisation and compromise are likely to be required versus other attributes. Firstly, attention should be paid to the margin to the choke line to avoid an excessive de-rate to prevent turbo overspeed at altitude. In addition, the effect of compressor efficiency on compressor outlet temperatures (a potential durability and compressor fouling concern) and fuel economy should be considered. For this case study, although there was some resulting compromise to these attributes, this was considered acceptable against the reduction in whoosh noise achieved. Finally, it should also be remembered, that engine calibration has an influence on the whoosh noise by controlling the point at which the engine operates on the map. Whoosh noise is reduced by reducing boost pressure although, for this case study, the resulting derate needed to avoid excessive smoke emissions prevented the realisation of significant whoosh reduction through calibration changes alone.
40
CONCLUSION
Turbocharger whoosh noise has been demonstrated to be present for a significant number of HSDI engines in the European market. It is a broadband phenomenon most often encountered at full load, low engine speed or during tip in / tip out manoeuvres. Turbocharger whoosh is associated with the turbulence generated in the compressor during operation within or close to the marginal surge region or surge line. Consideration of the proximity of operation to the marginal surge line has been shown to be a good guide to support compressor design selection. In conjunction with the consideration of other attributes, this allows the minimisation of whoosh noise excitation at the source. The main transmission path is excitation of and radiation from the intake ducting, most commonly the compressor outlet hose often with some contribution from the compressor inlet hose. Increasing hose thickness or mass has been shown to give significant reductions in the radiation of whoosh noise. Broadband resonators, commonly fitted as a countermeasure for turbocharger whine, can also provide similar significant benefits. Future studies are planned to understand the influence of dynamic mass flow and pressure ratio on whoosh noise and to investigate the feasibility of using 3D CFD to provide a means of predicting whoosh noise or aiding in compressor design. ACKNOWLEDGMENTS
The Authors would like to thank the Directors of Ford Motor Company Ltd and Ricardo pIc for the permission to publish this paper and to the Authors' many colleagues who have contributed to the studies discussed. REFERENCES l. Evans D, Ward A: Minimising Turbocharger Whoosh Noise for Diesel Powertrains,
2005 Noise and Vibration Conference, SAE 2005-01-2485 2. March J, Bennett C, Towalski C, Ward A: Achieving Diesel Vehicle Appeal, Part 2: Powertrain NVH Perspective, 2005 Noise and Vibration Conference, SAE 2005-012489. 3. March J, Strong G, Gregory S, Rediers B: Achieving Diesel Vehicle Appeal, Part I: Vehicle NVH Perspective, 2005 Noise and Vibration Conference, SAE 2005-01-2484. 4. Schieman J: Turbocharger Compressors - The Phenomenon of Surging, Turbo Magazine Issue I 1995. 5. Wren C S, Johnson 0, Gas Dynamics Simulation For The Design Of Intake And Exhaust Systems - Latest Techniques. SAE 951367. 6. Soh Kang-Young, Yoo Dong-Kyu, Kang Koo-Tae: Introduction of Noise Reduction Examples in Turbocharger System of Diesel Engines. FISITA, 2004, 30th Congress.
41
DEFINITIONS, ACRONYMS, ABBREVIATIONS CFD: Computational Fluid Dynamics EGR: Exhaust Gas Recirculation HSDI: High Speed Direct Injection NVH: Noise, Vibration and Harshness SPL: Sound pressure level VGT: Variable Geometry Turbo
42
The Influence of Installation Parameters on Turbocharged Automotive Engine Performance Geoff Capon, Alex Leong, Tim Morris Ford Motor Company Limited
ABSTRACT An automotive turbocharged engine configuration, compnsmg a turbocharger compressor, intercooler, EGR system, diesel engine and turbocharger turbine, are usually connected by systems of ducts and pipes. These must satisfy non-performance criteria such as clearance and crash without compromising aerodynamic and thermodynamic performance of the component parts of the air path system due to poor system design or installation. A study was made of the air inlet system design up to the turbocharger compressor inlet, and shows the relationship between the inlet flow field and the compressor wheel blade design. The study has shown that the turbocharger compressor performance can be influenced by inlet duct design and that CFD analysis can be used to characterise the gas flow behaviour within the duct. Flow bench test data is presented to support the CFD analysis and to validate the CFD process. Engine test data is also presented showing improved turbocharger compressor performance with the CFD optimised duct, thus further validating the methodology.
1.
INTRODUCTION
As the power and torque output of the passenger car diesel engine continues to increase with an attendant requirement to conform to continually restrictive emissions legislation, so the requirement for continually improving engine airflow becomes necessary. Recent technological advances that have contributed to these performance improvements embrace variable geometry turbines, improved compressor and turbine aerodynamic designs and improved intercooler thermal effectiveness. Air path system developments that have contributed to improved performance include intercooler bypass, EGR (exhaust gas recirculation) cooler bypass, two-stage and mixed sequential turbocharging, and inter-stage cooling. In parallel with these developments has been continual rapid advancement in control system technology together with strategies enabling fast and accurate control of engine air/fuel ratios over the wide engine operating envelope of speed, load, temperature and altitude. Initial development of these airpath system technologies and developments are often conducted in isolation from the application for they are destined, namely the vehicle. The simplest turbocharged engine configuration, comprising a compressor, intercooler, EGR system, diesel engine and turbine, is usually connected by systems of ducts and pipes and must also satisfy other non-performance criteria such as clearance and crash. In satisfying these criteria, attention must be given to the installation such that the advances achieved in aerodynamic and thermodynamic performance of the component parts of the airpath system are not compromised by poor installation. The following demonstrates the effect of air inlet system design up to the turbocharger compressor inlet, and will show the relationship between the inlet flow field and the compressor wheel blade design. CAE analysis is presented showing effects of poor inlet system matching together with design modification to overcome the shortfalls demonstrated. The analysis is supported by steady-state flow rig test data and engine dynamometer test results.
43
NOMENCLATURE
2. A; Alor
flow area of cell i
K Ill'
pressure coefficient total pressure drop between
r
inflow and outflow of model flow uniformity index
total flow area of cross-section FSD flow speed difference
3.
INLET SYSTEM INFLUENCE
3.1
Turbocharger matching
U
mean gas velocity of cross-section
gas flow velocity at cell i U; U max maximum velocity at cross-section Um;n minimum velocity at cross-section Vlan tangential velocity component axial velocity component Vax;al p flow density
Accurate turbocharger matching calculations and predictions usually rely on data for the turbocharger elements in the form of non-dimensional characteristics. In the case of the turbocharger compressor the characteristics provide relationships between pressure ratio, mass flow, rotational speed and adiabatic efficiency and are usually derived from steady-state gas stand tests. The compressor is invariably tested with an 'ideal' inlet system having at least ten diameters of straight ducting to the compressor inlet. Replacing this ideal inlet ducting with a realistic vehicle system comprising an inlet duct, an air filter assembly and a clean air duct to the compressor can significantly modify the compressor characteristics. Fig. 1 [1] illustrates an example of the change that can occur where a vehicle inlet system is fitted and a gas stand compressor characterisCompressor with 3.4 'ideal' inlet system tic obtained. The inlet system 3.2 has effectively shifted the 3 characteristic to the left to a 2.8 lower flow capacity. In this 0 instance the compressor surge ~ 2.6 IX: 2.4 margin is greatly increased, ~ ::::J 2.2 but any turbocharger speed !l! 2 and engine altitude operating ~ D.. margins have been greatly 1.8 eroded. Many contemporary 1.6 automotive diesel engine 1.4 turbochargers implement 1.2 variable geometry turbines with closed-loop control of 20 40 •• 80 6 100 I 120 1l 140 boost pressure, and this arCorrected NO Mass Flow rangement further exacerbates the effect of the inlet system on Fig, 1 Effect of vehicle inlet system on the turbocharger compressor compressor characteristric [1] characteristics. Note that a turbocharger match derived 14
44
from an 'idealised' compressor gas stand test can be significantly removed from an optimal match when a vehicle inlet system is included in the air path which provides non-ideal inlet conditions at the compressor impeller eye.
3.2
Compressor inlet conditions
The automotive engine turbocharger radial flow compressor impeller is invariably designed for the following flow conditions at the impeller eye:1. 2. 3. 4.
uniform pressure and velocity distribution across the complete inlet inlet velocity Mach number not exceeding 0.8 zero inlet pre-swirl (when no pre-swirl vanes are employed) ambient pressure and temperature
The design of the air path system up to the compressor inlet is influenced by other nonperformance constraints, and often results in bends and cross-sectional changes close to the compressor inlet which influence the first three items in the above list. The inlet system also has a pressure drop associated with it and this together with heat transfer to the inlet air influences ambient pressure and temperature.
3.3
Case study of compressor inlet duct
During development of a 2.2 litre four cylinder common rail direct injection turbocharged diesel engine for automotive application, the turbocharger performance, particularly the compressor, was found inferior to that predicted. The compressor clean side inlet air duct was identified as causing inferior flow to the compressor inlet eye and influencing the parameters listed in 3.2 above. This duct can be seen in Fig. 2. The turbocharger and turbocharger housing used for the study are standard GT-20 frame size production units supplied by Honeywell Engine Boosting System. Whilst it is recognised that the performance of the compressor can be significantly influenced by the ducting between the compressor outlet and engine inlet manifold, the current study focuses only upon the effect of the turbocharger compressor inlet duct. It is intended that the current work be extended to include one-dimensional and threedimensional flow analysis of the outlet ducting.
45
Fig. 2 Compressor inlet duct
4
CFD ANALYSIS
4.1
Introduction
The commercial CFD code, STAR-CD, [2] has been used in this study to perform the three-dimensional calculation of the flow inside the clean side air duct. The results obtained from this method have been compared and correlated with the results from steady state flow rig tests and dynamometer tests at fixed engine speeds.
4.2
Numerical model
STAR-CD is a finite-volume based CFD code. The flow domain is divided into numerous discrete control volumes of 'cells' and the code employs a 'boundary fitted' grid that allows the shape of flow boundary to be matched exactly to the physical boundary. In each control volume the three-dimensional momentum and mass conservation equations are solved by evaluating the temporal and spatial derivatives using flow variables within the cell and those of its neighbours. The result is a set of simultaneous non-linear equations that are then solved iteratively. The flow to be simulated is assumed to be steady state, isothermal, compressible and turbulent. The effects of turbulence on the flow are represented through the turbulent viscosity. For this study, the turbulent viscosity is calculated by using the 'High Reynolds number k-£ model' with wall function for closure. The first-order Upwind Differencing (UD) scheme is used for the momentum, energy and turbulence equations. The convergence criteria used for the steady state calculations is based on the default residual pf 0.001.
4.3
Computational grid
The geometry of the clean side air duct used in this study is shown in Fig. 2. The design is based on joining two 90° elbows in an off-planar configuration. To simplify the modelling requirements, all the components upstream of the clean side air duct, such as air filter and housing, are ignored. In addition, the Mass Air Flow sensor (which monitors the amount of Fig. 3 CFD model of inlet duct air mass flow passing through the system and is usually installed inside the clean side air duct) and the Crank Case Ventilation hose (which link the crank case housing with the inlet end of the compressor) are not modelled. To isolate the engine dynamic motion, the ducts are joined together by a flexible convoluted tube. In this design the convoluted tube is immediately upstream of the second elbow, and the
46
irregular surfaces could trigger flow separation downstream. To capture the effects of this feature, the true geometry of the convoluted tube was included as part of the model. An enlarged view of the convoluted tube section is shown in the insert in Fig. 3. In order to accurately simulate the physical phenomena at the entry of the compressor, the CFD model, as shown in Fig. 3, has included the geometrical profile of the compressor inlet duct. However, the impeller itself and other associated components are not included. Furthermore, to minimize the numerical instability that may be created by the high-speed flow at the outlet, the exit end was extended and the extension is equivalent to four times the diameter of the outlet section. The computational grid consisted of approximately 170,000 hexahedral cells. To ensure grid independence and improved accuracy of the results, additional calculations (not included here) with different grid densities were performed and compared with the reference mesh. No significant differences between the solutions were observed, so for the present study, the solution obtained with the reference grids was considered to be sufficiently accurate.
4.4
Boundary conditions
The CFD calculations were performed at steady state with a fixed mass flow rate of 640 kg/h and inlet gas temperature at 30° C. These values are based on an engine operating at 4000 rpm at full load. As shown in Fig. 3, the prescribed mass flow rate was applied as the inlet flow conditions whereas the constant pressure boundary condition was assigned as the outlet flow conditions. This arrangement reproduced the test conditions adopted in the cold flow bench measurements. All wall surfaces were treated as adiabatic walls with the 'no slip' option applied. The latter is required for modelling turbulent boundary layer (hydrodynamic and thermal) in turbulent flow calculations.
4.5
Characterisation parameters
To address the flow conditions required at the impeller eye (as noted in Section 3.2), the authors have developed a number of parameters that could be used to characterise the flow behaviour in the inlet duct. These flow parameters are applied to three attributes, namely pressure drop, flow uniformity (based on the flow velocity) and pre-swirl (based on the flow rotation). Some of these performance parameters can be measured directly in the flow test rig and could be used for validation. The definition of each parameter is given as follows. Unless specified, all these characterisation parameters are calculated at a 'plotting' section that coincides with the compressor eye (also known as the inducer plane).
4.5.1
Flow uniformity index (also known as GAMMA value)-This is an industry standard measure that integrates the velocity distribution profiles over the plotting section. A value of one indicates a totally uniform flow.
i.~(Ui _U)2 A r =1- ..:.i=:::.'_---==-_ __ 2·U· Arm In which 'j' is one of the 'N' cells in the cross-section
47
4.5.2
Flow Speed Difference-This quantity is a direct measure of the velocity fluctuation across the plotting section and is expressed as percentage, FSD = (U max- U min) U
The non-slip wall boundary condition implies that the flow velocity on the wall is zero. As a result, a small velocity magnitude in the sub-layer region could lead to large variations. To avoid this, the authors suggested that only the flow distribution within a circle which radius is 90% of the normal radius is considered. A value of zero indicates a totally uniform flow.
4.5.3
Pressure coefficient-In this study, this parameter is defined as, K=
M ~·p·U
4.5.4
2
Pre-Swirl Ratio--This is the ratio between the angular momentum and the axial momentum. It is simplified here as the ratio of the tangential velocity component to the axial velocity component, with its axis of rotation at the centre of the plotting plane. This quantity can be expressed as a mean value, e.g.
. . 1 Mean Pre - SWIrl RatIO = -
Vtan L-N i=N
i=! Vaxial
4.6 Flow conditioning devices As mentioned in the previous section, the original design of the compressor inlet duct consists of two 90° bends. A sectional view (with one half of the body removed) of the second elbow in the current design is shown in Fig. 4(a). Although there is no evidence that there are interactions between the bends, the off-planar configuration could create a strong swirl [3]. There is evidence that flow conditioning devices, such as the flow turning vanes proposed by Nakato & Kriitten [4] installed at a pump intake, and the guiding vanes tested by Kindl, et al [5] could improve the flow behaviour, but the effects of such devices on the swirl have not been properly discussed. In this study, the authors developed and tested two types of flow-conditioning devices. The results from the CFD calculations will be compared with the corresponding flow bench/engine dynamometer tests.
(a) Grillage In this design, eight solid plates are used to divide the cross-section of the duct into five parallel, equal sized channels. It is 25 mm in length, and as shown in Fig. 4(b), the device is resembled to a thick grillage. It is installed at the exit end of the duct. Although the plate used for the construction is only 2 mm in thickness, the reduction in flow area will cause an increase in pressure loss.
48
(b) Guiding vane
In this design, a curved plate (guiding rib) was inserted into the second elbow. As shown in Figs. 4(c) and 4(d), the curvature of the bend follows the mean flow path of the elbow and is designed to divide the cross section into two parallel channels. The thickness of the rib is 2 mm and it span across the width of the elbow. Two sizes have been investigated, namely: a. Short guiding vane: the length of the rib is approximately half the elbow's mean flow path and b. Long guiding vane: the length of the rib is approximately three-quarter the length of the elbow's mean flow path.
(b) Grillage
(a) Baseline
(c) Short guiding vane
(d) Long guiding vane
Fig. 4 Flow conditioning devices
5.
EXPERIMENTIAL FACILITIES
5.1 Cold flow bench test facility
The flow rig test facility used in this study was originally built for cylinder head flow and swirl measurement. One end of the test duct is connected to the air blower of the rig whilst the delivery end of the inlet duct is connected to a turbocharger compressor impeller housing that is taken from a production unit. Instead of passing through a
49
turbocharger impeller and volute, air is discharged into a dummy cylinder of the same diameter. The cylinder is 150 mm in length and its end is attached to an impulse flow meter. The impulse flow meter, described in [6], comprises a honeycomb flow straightener that is mounted on bearings axial to the cylinder; a precision arm that includes a strain gauge bridge assembly restrains the honeycomb. By measuring the force required to convert the rotational motion to axial motion, the meter effectively measures the total torque exerted by the swirling flow. The swirl ratio is the product of the measured torque value divided by the prevailing mass flow.
5.2
Dynamometer test facility
Engine testing was conducted using a high technology steady-state dynamometer test cell. In addition to measuring the usual performance parameters, the engine was instrumented for pressure and temperature throughout the complete air path system from air filter inlet to engine inlet manifold. At the compressor entry a manometric ring was employed to determine the pressure distribution at four points on the circumference. Engine testing was conducted at full-load from 1750 to 4500 rpm using the four variations of compressor inlet ducting.
6.
RESULTS AND DISCUSSIONS
6.1
Introduction
Calculations of the gas flow behaviour in the inlet ducts have been performed and evaluated by using the characterisation parameters defined in Section 4. The effects of the various flow-conditioning devices have been assessed. The CFD predicted data are compared with those measured on the flow rig. In order to assess the impact of these devices on the engine performance, the tests were repeated on an engine with the various devices installed.
6.2
CFD results
Calculations to study the gas flow behaviour in the inlet duct have been performed and the results compared with the same design but with different flow conditional devices attached. The results, expressed in various characterisation parameters, are listed in Table 1. Design Options Baseline With grillage With long guiding vane With short guiding vane
GAMMA
FSD*(%)
K
0.987 0.984 0.988
9.70 11.6 10.4
0.430 0.557 0.502
0.987
10.1
0.496
Mean preswirl ratio (%) 5.0 0.4 1.0 1.9
*Based on the area of a section whose radius is 90% of the original plotting section.
Table 1. CFD characterisation parameters.
50
As mentioned in Section 4, GAMMA is a measure of the flow unifonnity and in this case, the resulting values of all four cases are good and are close to the ideal value of one. However, it has to be emphasised that in this design, the compressor inlet duct is tapered and between the entry and the impeller eye the cross-sectional area is reduced by 45%. The reduction of cross-sectional flow area increases the flow speed. As a result, the variations of GAMMA values between different designs are much smaller. In other similar cases but with the straight inlet duct, the variations of GAMMA values between different designs are more significant. The alternative approach to assess the flow variation across the plotting section is to compare the flow speed difference values. As expected, the baseline case has the smallest fluctuation whereas the grillage case in which the cross-sectional flow area is partially blocked by the lattice structures has the largest fluctuation. In tenns of the pressure loss the general trend is about the same as observed in the flow velocity distribution. On the other hand, it has to be emphasised that because the inlet duct is part of an air intake system and only the fonner was considered in the calculation, these pressure values do not represent the true pressure distribution in the system. As expected, the baseline case has the highest mean swirl ratio. Installing the flow conditioning devices has considerably reduced the rotational motion in the duct. As shown in Table 1, most of the swirl in the grillage case has been removed. In the guiding vane cases, the long design has a greater effect in stopping the swirl than the short design. But unless the vane is stretched to the full length of the elbow, the guiding vane design is not as effective as the grillage in reducing swirl.
6.3
Cold flow test results
In additional to the swirl measurement, the cold flow test rig described in Section 5.1 can also be used to measure flow coefficient (ratio between the actual mass flow rate to the ideal mass flow rate). The data shown in Fig. 5 includes both swirl ratio and loss coefficient and shows flow rig test results of the four flow conditioning designs. The tests were conducted over a mass flow rates ranged from 200 to 650 kg/hr.
....
0 . 0 , - - - - - - - - - - - - - - - - - · - : : - : : - - - - , . . , . . - - . . , . . - - - - - r 0.155
i
'" 3i
-
'C ~1.0
-=~
. I: I : ": Grilage Swirl Ratio
Long Guide Swirl Ratio
•
*
•
0.150
Baseline Plain Duct SWirl Ratio
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.
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-7.0 ~---r--..,._-__.--_r_-___.--_r_-~._-_._--.---_+0.125 200 250 300 350 400 450 500 550 600 650 700 Air Ma8s Flow Kg/hr
Fig. 5 Flow & swirl rig test results
51
u
~
~.
_ • GrillageAowCoeff.
I
6.4
CFD vs. flow test
Fig. 6 compares the predicted swirl ratio with the corresponding measured values under the same test conditions. It appears that the predicted values are much higher than the measured values. On the other hand, both data sets have shown a similar trend, namely higher swirl in the plain duct case (baseline) and lower swirl with the grillage flow conditioning device installed. In addition, the flow coefficient values measured by the flow rig test (Fig. 5) have also shown similar trends to the figures given in Table 1.
5
o Measured
4
• Predicted 0
~
3
a:
~
2
01-"'"'-Baseline
short vane
long vane
Grillage
Oesgin Variations
Fig. 6 Comparison between measured and predicted swirl ratio (normalised by the corresponding baseline value)
6.S
Engine test data
The results from the engine full-load tests with the four variations of air inlet duct are shown in Fig. 7. The data is in the form of compressor outlet static pressure and temperature variation with compressor tip speed. In all tests the turbocharger was operating in closed loop control thus delivering consistent boost pressure for all tests. The 4000 rpm full-load point is highlighted for comparison with the CFD analysis. At the 4000 rpm operating points the desired level of compressor outlet pressure requires the highest impeller tip speed with the baseline design. The short and long guiding vanes give lower, but similar, impeller tip speeds for the same outlet pressure. Below an engine speed of 3000 rpm the outlet pressure characteristics converge and the variation with impeller tip speed diminishes. The characteristic obtained with the grillage flow conditioner shows a much improved performance with the impeller tip speed being significantly lower throughout the engine speed range. The temperature plots in Fig. 7 show that the compressor delivery temperature varies little between the
52
baseline, short and long guiding vanes, but reduces by approximately 14°C with the grillage device. This equates to an improvement in compressor adiabatic efficiency of some 7%, which is significant. The engine test data has demonstrated that the flow conditioning devices affect the performance of the turbocharger compressor by modifying the flow field presented to the compressor impeller eye, and that the grillage device provides the best flow field enabling the compressor to operate close to its design point and effectively improving adiabatic efficiency.
290
--+- Baseline - •. Short
_ 280 ..,.. Long
,'~
- •. Grillage
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Fig. 7 Engine test data for test ducts
7.
CONCLUSIONS • • • •
The performance of the turbocharger compressor can be influenced by the design of the inlet duct. CFD can be used to characterise the gas flow behaviour within the duct. It has been demonstrated that the CFD process is a valid method of optimising inlet duct design. The flow conditioning devices, such as grillage, have shown to give improved compressor performance when tested on the engine, thus further validating the analytical process.
53
ACKNOWLEDGEMENT
The authors would like to express their gratitude to Ford Motor Company for supporting this study and giving permission to publish this article.
REFERENCES:
[I] Effect of Air Inlet System on Compressor Operation. Ford Motor Co. Ltd. Internal Communication. July 2002 [2] Computational Dynamics Ltd., STAR-CD users Guide and Methodology (version 3.2),2004. [3] D.S. Miller, 'Internal Flow System', 2nd Edition, Flowmaster International Ltd. [4] T. Nakato, and M. Krtitten" 'Experimental Studies of Flow-Staightening Devices to Improve Approach-Flow Distributions at Pump Intake under Cross Flows', Iowa Institute of Hydraulic Research, The University of Iowa, Iowa City, Iowa, 1998. [5] H. Kindl, N. Schorn, , H. Schulte, J.R. Serrano, X., Margot, and J.C. Donayre, 'Influence of Various Compressor Inlet Designs on Compressor Performance', THIESEL 2004 Conference on Thermo- and Fluid Dynamics Processes in Diesel Engines, 2004. [6] Heywood, J.B. Internal Combustion Engine Fundamentals, McGraw-Hill Book Company, First Edition, 1988.
Copyright Assertion
© Ford Motor Company Ltd.
54
Using the Centrifugal Compressor as a Cold Air Turbine M. Milllera , S. Sumsera, P. Fledersbachera, K. Ro/3lera , K. Fiewegera , H.-J. Bauerb a DaimlerChrysler AG, Research Body and Powertrain b University ofKarlsruhe, Institute for Thermal Turbomachinery The exhaust gas turbocharger will continue to gain importance for future forced induction concepts. A major challenge for turbocharged spark ignition (SI) engines is to overcome their delayed response to sudden load increases from lower partial load and low engine speed conditions. This study presents one way to improve transient operating behavior by using the throttle enthalpy potential hidden in the SI spark ignition engine throttling process for A = 1 concepts. An innovative compressor concept should be able to maintain the exhaust gas turbocharger at significantly higher speeds during non-charged operating phases. For this purpose, the radial compressor impeller serves as a "cold air turbine" in the throttle engine operating points. This paper begins with a discussion of the effects of increases in turbocharger initial speed on transient engine response, using ID engine cycle simulations. Subsequently the thermodynamic background will be introduced and the technical concept of the DOT (Delay Optimized Turbocharger) will be shown. Test bench measurements outline the feasibility of the concept and demonstrate an initial turbocharger speed increase of approximately 50,000 rpm. Additionally, results of transient engine measurements will be discussed. Theoretical analysis and visualization of the flow during cold air turbine operation are shown through a 3D CFD simulation. The turbine speed characteristic will be demonstrated and first optimizations gained by the calculations will be highlighted. The experimental results clearly indicate that the speed bandwidth (speed spread) of a standard exhaust gas turbocharger can be drastically reduced. The simple prototype used in this study achieved a factor of 5 reduction. Hence, the target of a turbocharger operation at quasi-steady speed seems to be feasible. Such a device will significantly improve the response characteristics of turbocharged SI engines. INTRODUCTION Downsizing internal combustion (IC) engines is a promlsmg way to reduce fuel consumption. In the downsizing process, a naturally aspirated, larger displacement engine will be replaced by a turbocharged engine with a smaller displacement. The goal is to achieve better fuel economy in the relevant cycle range by shifting the operating point towards higher brake mean efficiency pressures. For spark ignition engines, potential fuel consumption savings in the 10 to 30 percent range are currently assumed [4]. Downsizing requires a powerful forced induction system. The exhaust gas turbocharger invented by Dr. Bilchi in 1905 is a particularly suitable charger variant. Its high efficiency, high achievable compression ratio, small space requirement, and relatively moderate cost make it a concept that is frequently used [4]. One difficulty this
55
turbocharging system has not yet satisfactorily solved, however, is the turbo lag of the engine during rapid load increases from lower partial load and low engine speeds [6]. If the load increases suddenly, large turbocharger speed ranges must be overcome until the necessary charge pressure is available. For a typical throttle controlled spark ignition engine with a displacement of 1.8 to 2.0 I, this will result in a turbocharger speed range of 10,000 to 200,000 rpm. The delay in boost pressure build-up is caused by inertia and bearing friction losses during rotor run-up, as well as by the low enthalpy provided to the turbine in these operating ranges. This investigation presents a concept that allows for maintaining the turbocharger speed at a significantly higher level during non-charged phases of operation. This enables the higher initial speed to accelerate boost pressure build-up during sudden load increases. DaimlerChrysler calls this new charger concept DOT (Delay Optimized Turbocharger). The goal is to operate the turbocharger at a quasi-stationary speed state. The energy necessary for the higher initial speed comes from using the enthalpy potential gained from the throttling principle, which is the basis of common spark ignition engines. To do so, the radial flow compressor of the turbocharger is operated as a cold air turbine. The necessary conceptual redesign of the classic turbocharger compressor and its effect on engine operation is presented below. BENEFIT OF INCREASED TURBOCHARGER INITIAL SPEED
Before concentrating on how to increase turbocharger initial speed, it was important to discover what effect a higher turbocharger speed would have on transient engine behavior. For this purpose, an engine model for a 4-cylinder in-line engine was set up as a full model in the GT-Power ID calculation program. The outputs from this established simulation software include not only detailed analyses of the load changes, but also of the transient exhaust gas turbocharger and engine run-up [8]. Acceleration from a lower part load at a constant vehicle speed of 20 km/h in second gear is considered the standard example. The vehicle weight is assumed to be 1640 kg. To simulate acceleration with the new DOT concept, exhaust gas turbocharger speeds of 60,000,90,000, 120,000, and 150,000 rpm were postulated. A calculation that included a standard waste gate exhaust gas turbocharger was used for comparative purposes. The resulting turbocharger speed and brake mean effective pressures are shown in Figure 1. It is obvious that immediately after a load demand, the brake mean effective pressure of the engine increased to a higher level. The initial amount of this increase depended on the initial speed of the exhaust gas turbocharger. This means that the increased initial speed of the turbocharger has a direct effect on the response characteristics of the engine. The DOT thus increases the amount of air provided by the charger system for combustion when acceleration starts. Also apparent was a drop in exhaust gas turbocharger speed, starting with initial speeds of more than 90,000 rpm. The reason for this is that the hot gas turbine cannot provide the stationary compression power necessary for the defined initial speed of the exhaust gas turbocharger. During a sudden load increase, the rotational energy stored in the exhaust gas turbocharger rotor thus supports the charge pressure build-up.
56
Therefore, in contrast to conventional turbochargers where inertia of the rotor is reduced to improve its responsiveness, a higher inertia may be propitious for the DOT concept. In general, it should be noted that the DOT causes a significantly higher change in brake mean effective pressures at the beginning of acceleration. The turbine is responsible for the following increase in brake mean effective pressure. This gradient can be improved by the variants presented in [9]. The DOT causes the engine response to approach the characteristics of a larger displacement in a naturally aspirated engine or supercharged engine. During load increases, more air is immediately provided to the engine. 2Srr=========~--,,--~-_,
240rr======,----r--...,----, Turbocharged Engine -0- DOT 60000
J200
1 =.
~
Turbocharged Engine -0- DOT 60000
;tgg~=
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160 I~~::;:::.;;:r:;::;;;::::t.~-~~~~-~
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~
!
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OJ 80 Io--o--of-o--<>----i>.;"
40
~~2--~-1--~0--~1--~-~
o
Time[s]
I
Time[s]
Figure 1: Spark ignition passenger car engine, simulated acceleration processes from a driving speed of 20 km/h in second gear
THERMODYNAMIC SYSTEM ASSESSMENT Using the enthalpy hidden in the throttling process of a SI engine is supported by a fundamental thermodynamic assessment of the overall system. The following calculation uses the results measured from a turbocharged SI passenger car engine with a displacement of 1.8 I to determine the enthalpy difference on the intake side between compressor entry PI and the pressure downstream throttle valve P2,s (Figure 2, left), and the exhaust gas side, between exhaust gas turbocharger turbine intake P3 and the turbine P4 (Figure 2, right). At high engine speeds and loads, a large isentropic enthalpy difference is available on the turbine side. This allows for high air compression in the exhaust gas turbocharger compressor. In contrast, the enthalpy difference in the left bottom part of the engine map (low engine speed and low brake mean effective pressure) is very small. This is caused by the ineffective back pressure behavior of the exhaust gas turbocharger turbine at small mass flow rates. Bearing in mind the isentropic energy potential of throttling on the air side (Figure 2, left) for the partial load operating points at the lower engine speed range, it is clear that some of this is operates at a much higher level than in the exhaust gas turbine (Figure 2, right). If this throttling energy could be used, significantly higher exhaust gas turbocharger speeds would be achieved at the partial load points.
57
The assessment in Figure 2 shows the maximum available energy potential; any losses or component efficiencies were intentionally ignored so it would not influence assessment of the theoretical potential. The efficiencies of the overall DOT system must be identified and optimized for the aspects relevant to engine operation. • Turbocharged SI Engine
a. w :; co
• Displacement 1.8 I
• " = 1 Concept ·4 Cylinder • Wastegate TC Engine Speed Isentrop Throttle Losses Cold Air Side
14
0
12
-
'C'
10
e'"
8
Isentrop Turbine Power Exhaust Gas Side 14
Enthalpy [Watt]
p", I p, [-]
12
~
a. w :; co
a.
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10
8 6
4
o 1000
1500
2000
2500
3000
3500
1000
4000
Engine Speed [rpm]
1500
2000
2500
3000
3500
4000
Engine Speed [rpm]
Figure 2: Engine map of a 1.8 I turbocharged spark ignition passenger car engine
The effects of using the enthalpy potential of throttling on the turbocharger speed are significant. While equation (I) yields the power balance at the turbocharger shaft in traditional operation, equation (2) applies to throttle-controlled operation. It is assumed that the compressor impeller will act as a cold air turbine.
+ PFriction
PTurbine
=
PTurbine
+ PColdAirTurbine
Pcompressor
(1)
= PFriction
(2)
Since the right side of the equation describes the "consumers," bearing friction is the only power consumer in cold air turbine operation. In throttle engine operation there are two power sources: the hot gas turbine and the compressor acting as a cold air turbine. The thermodynamic assessment in Figure 2 shows that the cold air side can provide 500 Watts (linear interpolation) of isentropic energy at engine speeds of 1,500 rpm and a break mean effective pressure of 2 bars. Assuming a cold air turbine efficiency of 20 percent, the cold air turbine should provide about 100 Watts of energy. This energy, and that of the hot gas turbine, will then be in balance with the bearing friction loss at the resulting turbocharger speed. When using traditional journal bearings, the power (shown in Figure 3) needs to be overcome. This should allow for maintaining the turbocharger speed at 50,000 rpm using cold air turbine power alone.
58
Hence, the real benefit of the cold air turbine concept is obvious. It should be remembered, however, that the compressor wheel operating at 50,000 rpm and the minimum possible mass flow (limit comes from the surge line) require a power input of 400 Watts. Therefore, even a power output of just 100 Watts, which corresponds to a very low efficiency of only 20 percent, would be of clear benefit. Through additional optimizations, such as increasing cold air turbine efficiency, use of low-friction bearing systems, and variable turbine geometries (to obtain more power from the hot gas turbine during low load conditions), it should be possible to reach significantly higher turbocharger speeds. 350
/
300 250 [
200 ~
~
.g
150
u.
c..
100
~
50
o
.....
~ o
-/
/
/
/
/
....
25,000
50,000
75,000
100,000
Turbocharger Speed [rpml
Figure 3: Typical curve for exhaust gas turbocharger bearing friction
TECHNICAL IMPLEMENTATION The above shows that it is quite possible to raise the turbocharger speed by using throttling losses. A simple design was used for technical implementation, [12] based on a classic gas exhaust turbocharger. Here use of the throttling energy occurs in the radial compressor impeller. To do this, the cold air side of the turbocharger had to be redesigned (Figure 4), so a movable plunger CD was integrated that allows for adjusting the cross section of the compressor intake opening. Additional guide vanes @) allow radial entry of the air to the compressor impeller. The flow cross section of these guide vanes can be varied with a matrix d:J, which is connected to the plunger. Adaptations to the compressor housing were necessary to house these components. During the charged operating phases of the engine, the air flows axially to the compressor impeller (sketch at top right of Figure 4). The retracted plunger creates a very large cross section of flow at the compressor intake. In this operating state, the guide vanes, which are arranged concentrically around the compressor impeller, are also open. This open position was initially chosen for ease of technical implementation. In addition, the opening also acts as a stabilizing map measure in the compressor surge limit range; which expands the usable compressor operating range [3]. The transition to turbine operation is accomplished by moving the plunger. First it moves into the conical stop C?) shown in Figure 4. When the plunger has reached the stop (sketch at bottom right of Figure 4), air can no longer enter the compressor impeller axially. Now air can only flow radially via guide vanes to the compressor impeller. At the maximum opening of the guide vanes, the impeller can still function as an extreme co-swirl compressor.
59
If the plunger is moved further toward the compressor impeller, the matrix starts to slide over the guide vanes. The cold air turbine thus practically becomes a variable axial slider turbine. This sliding motion reduces the opening cross section of the guide vanes and restricts the air flow, resulting in a high flow speed (which may be close to the speed of sound) at the exit of the guide vanes. The very high co-swirl of the flow drives the compressor impeller in the direction of rotation. The compressor acts as a cold air action turbine. This design makes it possible to control the entire system using a single actuator. The quantity control of the spark ignition engine is now achieved through the expansion process in the DOT guide vanes and the subsequent energy transfer in its rotor, and not in a classical throttling process in the throttle valve. Initial simulations were conducted to test this control strategy. They show that in principle it is possible to regulate the engine using such a component, and to replace the throttle valve. It should be mentioned, however, that moving the throttling point ahead of the compressor requires special sealing measures for the bearing housing. This is necessary to prevent oil leaking from the bearing to the intake system. This seal was established by adding a by-pass component behind the compressor impeller, as well as a labyrinth seal.
Q) Moveable Plunger
<2> Conical Stop
O>,@ Exchangeable guide vanes I Variable guide vanes with matrix
Figure 4: DOT concept to use the centrifugal compressor wheel as a cold-air turbine
EXPERIMENTAL VERIFICATION OF THE BASIC DOT CONCEPT Stationary flow test stand
For the basic functional testing of the cold-air turbine operation, a prototype was set up, using interchangeable guide vanes that represent different fixed plunger and matrix positions. The measurements were carried out on a stationary flow test stand [14]. First the conventional compressor map with fully opened plunger was measured. In a second step, a vacuum was created at the cold air side, at the compressor exit (on the P2 side), while the turbine did not receive any additional power. This set-up allowed
60
generating the cold-air turbine map. In this case, the hot gas turbine dissipates energy, in addition to bearing friction. While this operating mode allows demonstrating the basic function of the DOT concept in principle, the efficiency of the cold-air turbine cannot be accurately determined (see section "Cold-air turbine optimization through 3D flow simulation"). The compressor map can be expanded by adding the turbine operating points (Figure 5). Cold air turbine operation takes place below the 1.0 pressure ratio line. For these conditions turbine operation is illustrated as a VTG 1 turbine map according to Hiereth. [5] The lines connecting the speed points depict the flow rate behavior of the cold air turbine at a constant opening cross section of the guide vanes (similar to a constant VTG opening). Here a typical nozzle discharge characteristic becomes apparent. The measured map thereby underscores the basic functioning of the innovative DOT technique and the possibility of using a compressor wheel as a turbine.
i: ...!...
~ f-
I'
Compressor Map:
I
~-:~;:;:,:.:.::::. ·~l+i,-/-+l/-_._-./_7·_·_~I_=··_·=···_=·_F.__-._-.....
norm (2,OOC/981 mbar)
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i eA. 30mm' guide vanes, ax;alent'Y totally closed
------i .B __ 60mm 2 guide vanes, axial entry totally closed : .c. 90mm 2 guide vanes, axial entry totally closed i .-D" 90mm 2 guide vanes, axial entry opened O.5mm
f-~if'4AO-IB'+-'---'Dlj-i-;'E-+--+--+--+-:'1;~": :~;_;_: _;~:~: ::~:::_;_;:;_:_ ;~!~_~~;~;;.,~;~,~,;., om
0.04
0.06
0.08
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Reduced Mass Flow [kg/s[
Figure 5: Extended compressor map with cold-air turbine range
1 Variable Turbine Geometry
61
..i
Stationary engine test stand
For a more accurate quantification of the turbocharger speed increase, the DOT was attached to a 1.8 I 4-cylinder engine. In Figure 6 the basic turbocharger speed with a waste gate concept is compared to the speed line measured with the DOT concept. The varying load conditions measured with the DOT concept were established by changing the interchangeable guide vanes. Hence, the engine throttling process took place in the guide vanes. The measurement results of the DOT demonstrate a stationary increase at a turbocharger speed of about 50,000 rpm. In the test described here, a classic turbine with a waste gate valve was used. This kind of turbine does not greatly contribute to an increase in turbocharger speed because of the low back pressure behavior at the relevant operating points. By using an adapted VTO turbine, it is expected to achieve more power to boost the exhaust gas turbocharger's basic speed. In addition, the rotor bearings used were of a simple journal bearing type, from series production. The use of low-friction bearing technology permits a reduction in overall drive power. Thus, there is still considerable development potential for further speed increases. 120000
100000
n
! ~
1
Idealized line variable guide vanes
I
\
I
A
Speed increase about 50,000 rpm
/ a1 ~
20000
/f
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._0... - ...
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-®-Baseline: axial entry totally opened €) 15 mm2 guide vanes ~ 30 mm2 gu!de vanes t 60 mm2 gUide vanes @ 90 mm2 guide vanes
6
7
I
I
I
I
10
11
-;
b I 12
BMEP [bar]
Figure 6: Spark ignition passenger car engine with DOT (stationary test on engine test stand)
Transient engine response
The effect of the increased initial speed during the acceleration process has been explained in Figure 1. When there is a sudden load increase, the engine gets more air and will act like an engine having a greater displacement. Moreover, the system response is favorable at sudden load reductions, which frequently occur in vehicle engine operation. In Figure 7 a sudden load reduction in a DOT system with variable guide vanes is compared with the behavior of a throttle valve controlled system. One very important aspect here is that the turbocharger speed stays high in the phases where the load is reduced. This is because in the throttled engine operation point the compressor wheel does not consume energy from the system, but instead feeds energy to the system. Thus, even 1 s after engine load reduction, the
62
turbocharger speed is still as high as in full load mode. This almost enables it to reach the level of a quasi-stationary turbocharger operation. Compressor surge during slam load reduction frequently occurs in turbocharged engines and has to be mitigated by introducing a recirculation valve. The DOT concept, however, prevents the compressor from surging in this kind of operation (compare P2 curve on Figure 7). Therefore, it is possible to dispense with a compressor recirculation valve with a DOT system. Another advantage of the DOT concept is that the pressure P3 that drives the exhaust gas turbine remains at a higher level after a slam load reduction than with a standard turbocharger. This is because the turbine's flow characteristic varies with the rotational speed. The overall benefit of the above effects is that the turbocharger almost remains at the same turbocharger speed after an engine load reduction; even if the speed drops slightly, turbocharger re-acceleration will be improved because the torque generated by the exhaust gas turbine will be enhanced from the increased pressure ratio.
:::.j~~ E140
" ,
e-
\
~120
•
'-.
I
\ " ,
~100",
"-'"-"'<" .....
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o
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•
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r-
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-
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...."":- ~ ...:.:.:-
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-
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',._._._._._._,_ 0.25
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'.
iH~'l~".'~~~'\"'~;
OL-________________________~ ... -0.25
r....-..-.. -p,--=D-,-OT=--------,· 100
1400,-. _.
200 ___ \
...... _ . _ . _ . - . _ . _ . _ •• 0.0
0.25
0.5
0,75
•
1,0
Time[s]
Time [s]
Figure 7: Sudden load reduction with variable DOT compared to standard turbocharger (TC) system
Cold-air turbine optimization through 3D flow simulation To better understand the flow pattern in the compressor alternatively cold-air turbine wheel, respectively, 3D-CFD 2 flow simulations with StarCD are presently being conducted. Turbo system design optimization using CFD is becoming increasingly common [1], [2], [7], [11]. A virtual test stand of the DOT concept should be used to evaluate the efficiency of different designs. The 3D CFD flow simulation is used to help optimization of cold air turbine operation without affecting compressor operation. Therefore, the flow in the standard DOT setup (both in compressor as well as in turbine operation mode) was investigated. The computational grid is comprised of approximately 4.5 million cells. Based on the experience with the previous turbocharger simulations a frozen rotor approach has been selected.
2 Computational
Fluid Dynamics
63
The computational results shown in Figure 8 were performed for a rotational speed of 50,000 rpm, a pressure ratio of P2s/Pl = 0040, and guide vanes with opening width of 3 mm. These data correspond to an engine operating point of 2 bar break mean effective pressure and 1,500 rpm engine speed. Streamlines of the co-rotating coordinate system and the static pressure distribution on the compressor wheel are shown in Figure 8. It is clear that the air flow exiting the guide vanes hits the compressor blades with an extreme co-swirl before being deflected. The air then flows in the void in front of the compressor. There it is redirected and passes (close to the hub) for a second time through the compressor impeller towards the diffuser. The power uptake of the compressor impeller in cold air turbine operation of the DOT is thus caused by a pulse exchange of the swirl flow in the outer blade area. Because the output power of the cold air turbine is difficult to measure, the 3D CFD results are used to quantify its characteristics. To point out the benefit of the cold air turbine concept more than the DOT alone was calculated. In addition the power consumption of a centrifugal compressor operating with pre-throttled air is shown as a baseline in Figure 9. It should be kept in mind that compressor surge prevents throttling the air on the downstream side of the compressor at high rotational speeds and low mass flow rates. The calculations shown in Figure 9 support the experimental findings: for a certain range the pre-swirl introduced by the DOT guide vanes cause the compressor wheel to operate as a turbine generating power output. The intent of the design is to extend the positive power range to the highest turbocharger speed possible
64
Figure 8: 3D simulation of DOT speed distribution in cold air turbine operation at pressure ratio 0.4 and 50,000 rpm ·+··Throttle valve downstream com~ressor (measured d a t a i = l 200
--
"-3D-CFD Compressor operating with pre-throttled air (No pre-swirl) ...-sD-CFD standard DOT case
L_.,.....-...
100
....3D-CFD with reduced rest volume
-~=::---~--,-
I !
[
-100
i 0
0..
f
-200
-300
~O
-500
Figure 9:
Cold air turbine power according to 3D CFD calculations
65
[
SUMMARY AND OUTLOOK
This investigation shows that using the throttle enthalpy potential of throttle-controlled spark ignition engines is thermodynamically possible and technically feasible for achieving turbocharger speed increases in exhaust gas turbochargers. To achieve this, the DOT (Delay Optimized Turbocharger) concept is presented. Engine cycle simulations of ID show the effect of increased turbocharger initial speed on the acceleration process. Here a direct impact of the turbocharger speed at the beginning of the acceleration process and the break mean effective pressure built-up is detected. Initial results from a simplified prototype with interchangeable, fixed-width guide vanes are presented. Using this technical design, it was demonstrated that the compressor wheel can act as a cold-air turbine on the turbocharger test bench. In addition, results from a stationary engine test stand are presented. Here an increase in turbocharger speed of 50,000 rpm at an engine speed of 1,500 rpm and engine load of 2 bar brake mean effective pressure was achieved. Transient measurements of the variable DOT system were discussed, showing the high potential this concept has on transient engine operation. The flow curve is visualized using a 3D CFD simulation. These simulations were used to analyze the speed characteristic of the innovative cold air turbine. First optimizations based on 3D CFD calculations were outlined. The ID review shows that a further increase in rotor speed adds to a greater boost of the engine moment. Thus, in the next development phase, the goal is to achieve the highest possible initial speed of the DOT through efficiency optimization of the cold air turbine, adapted back pressure behavior of the hot gas turbine, and use of low-friction bearing technology. It should be possible to reduce the large speed range of the standard exhaust gas turbocharger drastically, by a factor of 20 (currently, a factor of 3 for DOT, in the future, a factor of 2), thus achieving the behavior of an exhaust gas turbocharger with quasi-stationary speed. This will enable fuel efficient downsizing concepts. ACKNOWLEDGEMENTS
The authors would like to thank all the colleagues at DaimlerChrysler who contributed to the progress of the DOT concept. Major aspects of this project were developed as part of the NICE project (New Integrated Combustion System for Future Passenger Car Engines). The project leader was N. Freisinger, DaimlerChrysler AG, in the Sixth European Research Program.
66
BIBLIOGRAPHY [1]
S. Arnold, D. Caita, K. Dullack, C. Judd, G. Thompson, "Development of an Ultra-High Pressure Ratio Turbocharger", 2005-01-1546, SAE World Congress, Detroit, USA, 2005.
[2]
H.-P. Dickmann, T. Wimmel, J. Szwedowicz, D. Filsinger, C. Roduner, "Unsteady Flow in a Turbocharger Centrifugal Compressor - 3D CFD Simulation and Numerical and Experimental Analysis ofImpeller Blade Vibration", ASME Turbo Expo, Reno-Tahoe, USA, 2005.
[3]
F. B. Fisher, "Application of Map Width Enhancement Devices to Turbocharger Compressor Stages" (SAE 88074),1988.
[4]
R. Golloch, "Downsizing bei Verbrennungsmotoren - Ein wirkungsvolles Konzept zur Kraftstoffverbrauchssenkung", Springer-Verlag, Berlin, Germany, 2005.
[5]
H. Hiereth, P. Prenninger, "Aufladung der Verbrennungskraftmaschine," Springer-Verlag, Berlin, Germany, 2003.
[6]
R. Kemmler, H.-G. Lehmann, J. Schommers, "Entwicklungstendenzen aufgeladener Ottomotoren", Seventh Supercharging Conference, Dresden, Germany, 2000.
[7]
M. Klaus, A. Schulz, S. Wittig, "Eigenfrequencies and Mode Shapes of Radial Turbine Blades: Single Blade Calculation vs. Entire Calculation", ISROMAC-lO, 2004.
[8]
T. Kuhn, D. Mowll, F. Wirbeleit, J. Willand, "Optimization of the valve lift strategy during the acceleration of a diesel engine using WAVE and DOE", Fifth Ricardo Software International User Conference, Detroit, USA, 2000.
[9]
H.-G. Lehmann, T. Gotz, V. Siedle, M. Bechtold, M. Bottcher, H.-K. Weining, C. Enderle, "Potenzialbetrachtung zum Instationarverhalten von hochaufgeladenen Ottomotoren", Ninth Supercharging Conference, Dresden, Germany, 2004.
[10] Z. Liu, D. Hill, "Issues Surrounding Multiple Frames of Reference Models for Turbo Compressor Application", Dresser-Rand, Olean, USA, 2004. [11] D. Palfreyman, R. Martinez-Botas, "Turbocharger Turbine Unsteady Computational Study", THIESEL Conference, Valencia, Spain, 2004. [12] "Brennkraftmaschine mit einem Abgasturbolader und Verfahren hierzu," Patent DE 199 55 508 C I, German Patent Agency, 1999.
[13] I. Wilke, H.-P. Kau, "CFD-Simulation von Verdichterstufen mit Casing Treatment", DGLR-JT 2000-102, DGLR lahrestagung, Leipzig, Germany, 2000. [14] J. Willand, F. Wirbeleit, G. Hertweck, S. Sumser, W. Frief3, P. Fledersbacher, 2000, "Vorgehensweise bei der Entwicklung innovativer Aufladesysteme", Seventh Supercharging Conference, Dresden, Germany, 2000.
© 2006 DaimlerChrysler AG. All rights reserved.
67
Extending the knock limit of a turbocharged gasoline engine via turboexpansion J.W.G. Turner, R.J. Pearson and N. Milovanovic Lotus Engineering D.W. Taitt Loughborough University ABSTRACT Turboexpansion is a concept which is aimed at reducing the fuel consumption of pressure-charged gasoline engines by providing over-cooled air to the engine in order to extended the knock limit (relative to a conventional charge-cooled system). An extension to the knock limit allows several possible shifts in the compromise adopted in pressure-charged gasoline engines, including increased specific output, an increase in trapped compression ratio or a reduction in the octane appetite of the engine. All are considered important in the drive to reduce the CO 2 output of passenger cars fitted with gasoline engines against the backdrop of the ACEA commitment of 140g/km of CO 2 per km by 2008. The present work reports on initial results from running a single-cylinder engine under conditions of varying temperature at fixed charge-air densities to assess the worth, in combustion terms, of pursuing reduced charge-air temperature to extend the knock limit. To enable this, a charge-air conditioning rig has been constructed based on three positive-displacement superchargers and heat exchangers. This device is capable of providing close control of the charge air required for this testing programme. It has been shown that reducing the charge air inlet temperature, at a fixed density, provides an extension to the knock limit of up to 3.5 0 of spark advance at high compression ratio for a temperature reduction from 40 to 20°C at 2000rpm.
1.
DOWNSIZING AND ENGINE KNOCK
The torque generated by an internal combustion engine at a given operating condition is dependent on the inlet charge air density. In the case of the spark ignition engine (particularly when operating on a relatively low octane fuel such as gasoline), for a given density level, reducing the temperature of the charge trapped in the cylinder will reduce the propensity of the engine to knock and thus enable the use of a higher compression ratio. An extension to the knock limit allows several possible shifts in the compromise adopted in pressure-charged gasoline engines, including increased specific output, an increase in trapped compression ratio, or a reduction in the octane appetite of the engine. All are considered important in the drive to reduce the CO 2 output of passenger cars fitted with gasoline engines against the backdrop of the ACEA commitment of 140g/km of CO 2 per km by 2008, itself a reaction to the Kyoto Protocol of 1997 [1]. An effective route to reducing the CO 2 output from motor vehicles is to 'downsize' the engine and run it at a higher specific load for any given road speed; this increase in the specific output of the engine is normally achieved by using some type of pressure charging system. The resulting shift in the operating point of the engine to a more efficient regime is chiefly a result of reduced pumping work coupled with a reduction in friction losses. Using a conventional approach of port-fuel injection and single-stage charging with intercooling, an 18% reduction in fuel consumption has been predicted
69
[2]. This can be extended considerably by employing a variable compression ratio (VCR) mechanism, Schwaderlapp et al predicting a further 9% improvement [2] and Drangel and Bergsten [3] reporting a 30% improvement in drive-cycle fuel economy over the basic turbocharged type of engine. Knock is a result of the uncontrolled autoignition of the end gas I. As well as creating soot [4], its audible manifestation is symptomatic of why it needs to be avoided: when it occurs extremely high pressure waves move at supersonic speeds across the combustion chamber, in various modes of vibration [5]. These compression waves can often ignite unburnt charge in the top ring land, causing severe local overpressures and hence characteristic damage patterns [6]. Furthermore, the high local gas velocities induced by knock result in extremely high rates of heat transfer [5]. The resulting catastrophic damage to the piston and cylinder head face that often results from these pressure and heat transfer effects if an engine is run in a knocking condition for long is easy to understand. The end gas is heatcd by compression from the advancing flame front as well as by convection from the burnt gas, and autoignites when the Livengood-Wu integral equation (1)
is satisfied. The parameter ritself is defined by an Arrhenius rate equation of the form (2)
where A, nand B are constants which vary for different fuels at different conditions in different engines. Charge-air cooling via a heat-exchanger device (some times known as an intercooler) is the most commonly-used palliative to the increased charge air temperature which results from the work donc on the air in the compression stage. However, there is a limit to what an intercooler can achieve since, for any given compressor discharge temperature, a simple intercooler cannot reduce the charge air temperature to the atmospheric (or 'sink') temperature. The intercooler effectiveness 2, E, is given by t: =
[lin - Tout]
.
(3)
[1jn - Tcoolant ]
The lowest value of the coolant temperature, Tcoolanh without refrigerating the coolant, is the ambient temperature. To reduce the air outlet temperature to ambient air temperature would require an intercooler of 100% effectiveness, i.e. one of infinite size or one in which the charge air flowed at an infinitesimal rate. Air-to-air intercoolers typically have an effectiveness of 70-75%, and air-to-water intercoolers can offer 'airside' effectivenesses of 80-90% under transient conditions}.
Here we define the end gas to be that portion of the unburnt charge furthest from the spark plug. This is more properly teaned the thermal ratio of the intercooler. :3 'Air-side effectiveness' is considered using ambient air temperature and not coolant temperature for equation (6) - this is done to provide a useable figure without the complication of the different specific heat capacity of the water and the air. 1
2
70
The effectiveness of any intercooling stage can be considered of great importance because a key parameter determining knock initiation and propagation is the charge temperature at the time of ignition; this is one of the main controls on avoiding knock in most engines, as it influences the total heat energy of the end gas after combustion has been initiated. If the charge temperature cannot be reduced at this point in the cycle, the ignition must be retarded. However, retarding the ignition results in an unfavourable shift in the heat release curve and therefore a reduction in engine efficiency and also results in higher gas temperatures at the point of exhaust valve opening (EVO) which can mean extra fuel is necessary to reduce piston, turbine inlet, or catalyst inlet temperature (in turn reducing combustion efficiency and directly increasing fuel consumption further). Some means of reducing the end-of-compression temperature directly and beyond the level attainable via intercooling might therefore provide a possible way out of this problem.
2.
TURBOEXPANSION
Turboexpansion is one possible means of reducing the inlet charge temperature below the level which can be achieved with a conventional heat exchanger arrangement (i.e. a charge cooler) by extracting work from the gas flow. This is arranged by lowering the end-of-compression temperature (EOCT) through a reduction of the start-ofcompression temperature (SOCT); for any given compression ratio, the EOCT will therefore be lower. Thus turboexpansion is a concept which is aimed at reducing the fuel consumption of pressure-charged gasoline engines by providing over-cooled air to the engine in order to extended the knock limit (relative to a conventional charge-cooled system). Previous work by the authors has proposed the use of the turboexpansion concept for down"sized gasoline engines and has characterised the cooling effectiveness of such a charging system. This concept is shown schematically in Figure I together with a temperature-entropy diagram for the intake air. Compressor
Turbine
T
X
PUpper PPlenum
AirFilw
EBP Turboexpander
PAlm
~,--~---,
Energy Recovery ----(Electrical, Mechanical
or Hydraulic etc.)
s
Fig. 1: Schematic of a possible turboexpansion charging system for a turbocharged engine (left) and associated temperature-entropy diagram (right) [note: EBP = Exhaust Back Pressure] Historically, there has been much interest in such systems, with one of the first patents filed by the diesel engine manufacturer Paxman in 1950, and a system was successfully employed on spark-ignited natural gas versions of heavy duty diesel engines by the Cooper-Bessemer Corporation in 1959 [7]. Following an initial modelling investigation [8], some initial dynamometer results gathered by the authors have also been reported [9] which were used successfully to validate a further model [10]. This work has in turn suggested that if the necessary system efficiencies and effectivenesses can be achieved, an increase in compression ratio in the order of 0.5-1 unit could be supported at a brake mean effective pressure of 19.7bar. This is slightly 71
less than one might expect from a turbocharged gasoline engine with direct injection (GDI) [11,12,13]. In the right-hand side of Figure 1 it is clear that an initial compression of the charge air is performed to an upper system pressure (represented by the line from' Atmospheric Conditions' to point X in the figure), at which state conventional intercooling is carried out (points X to Y). The charging system to this point is similar to what might be considered a 'conventional' pressure-charged one, albeit at a higher pressure; however, there is then a supplementary expansion through a turbine (points Y to Z) which, in addition to reducing the pressure of the air, also reduces its temperature further. Two points of interest in this approach are that firstly, because of the necessarily higher pressure ratio of the compressor, the intercooler operates with a larger temperature difference across it than in the conventional system (and so can reject a larger amount of heat energy), and secondly, that the turbine rejects power which must either be wasted or collected by the vehicle for useful ends. While lower charge air temperatures can theoretically lead to an extension of the knock limit, which would in turn allow the geometric compression ratio to be increased, there is also a further benefit in that for any given cylinder size at that higher compression ratio, the clearance volume will be smaller and hence the amount of combustion residuals left in the chamber to mix with the incoming charge will theoretically be less. The propensity for residuals to promote knock in engines is due to the thermal energy retained in the cylinder during the next cycle in general and the formation of inhomogeneous exothermic centres in particular. Therefore any increase in compression ratio would particularly be of benefit in PFI engines where the cylinder cannot be purged with only fresh air during the valve overlap phase. However, it must be remembered that the mass of residuals trapped is a function of pre-turbine pressure and temperature (i.e. density) as well as compression ratio. The exact form of the pressure-charging system used will therefore have a bearing on the amount of residuals retained in the cylinder as a result of the multiplication of the main exhaust system back pressure by the overall expansion ratio of the turbine stage (in the case of a conventional turbocharger-based system). The turbine stage can therefore have a significant bearing on the knock limit of the engine, and is one of the factors motivating the need for more efficient turbocharging systems with respect to balancing the pressures across the combustion chamber. The importance of this within an engine system has been shown in a comparison of cylinder-charge dilution methods by Duchaussoy et al [14]. The modelling work performed by the authors suggested real merit in pursuing the turboexpansion concept further [81. This led to the design of a system for test on a new research engine, and initial dynamometer results being gathered [9,15]. The coefficient of performance of such systems has been characterised and the likely limits of the operating envelope quantified. The initial dynamometer results were modelled with a good fit found using the knock model of Douaud and Eyzat [16]. These results were disappointing as far as the potential benefits from a turboexpansion system were concerned but they do suggest that, if the charge air temperature can be reduced sufficiently, then an increase between 0.5 and 1 ratio of compression relative to a simple intercooled PFI engine can be sustained. On a pressure-charged engine this is of interest because they tend to have to be configured as low compression devices to support full-load operation, for reasons of avoiding the knock limit as discussed in Section 1. This modelling result has led to the configuration of a significant test programme using a new single-cylinder test engine fed with externally-conditioned charge air and
72
both port- and direct-fuel injection. The remainder of this paper will be concerned with some preliminary test results from the engine in PFr format and a discussion of further work.
3.
TEST EQUIPMENT AND PROCEDURE
3.1
Engine
The engine used for this test work was a new type of single-cylinder engine designed by Lotus for use in the UK Engineering and Physical Science Research Council-funded collaborative project 'HOTFIRE' [17]. The engine has a bore and stroke of 88mm and 82.lmm respectively, giving a swept volume of 499.3cc, and the compression ratio for this work was 10: 1. It is fitted with the Lotus Active Valve Train (A VPM) fully variable valve train system, which has been reported extensively elsewhere [17,18]. The detail design of this engine will be reported in a later publication.
3.2
Charge-Air Conditioning Rig
Fundamental to this work was the ability to be able to vary charge air temperature and pressure over a relatively wide range. To that end a dedicated charge air conditioning (CAC) rig was conceived and built, protected with the ability to flow sufficient air for multi-cylinder operation to provide a more flexible tool for use on other projects. This rig has two stages of compression with two stages of intercooling, one each before and one after the compression stages. The first stage of compression consists of two Eaton M45 superchargers in parallel with a single M45 unit for the second stage. Cooling water for the intercoolers is cooled by an in-cell chiller unit with the capability of reducing the intercooler water to below 10°C in a 25°C ambient. The superchargers are all driven by a single toothed belt taking power from a 45kW variable-speed electric motor which with the throttling of the two first-stage superchargers controls the flow rate. Twin surge tanks are provided in parallel downstream of the second stage intercoolers to damp the pulsations in the air flow from the rig. A photograph of the CAC rig together with a schematic of its layout fitted a spill for finer control of air temperature is shown in Figure 2.
45kW variabl9-6peed
electric motor
Engine
Fig. 2: Photograph and schematic of Charge Air Conditioning rig (SC - supercharger; IC - intercooler)
73
3.3
Test Cell Equipment
3.3.1
General Measuring Equipment
The test cell utilises a Froude type AG30 eddy-current dynamometer with an automated data logging system. Air temperature into the test cell was controlled to 25°C. Fuel consumption was recorded using a gravimetric flow meter and the fuel used for this work was 95 RON unleaded gasoline. A Kistler type 6041 A water-cooled pressure transducer was mounted in the cylinder head; combustion data was recorded on an AVL Indiset combustion analyser.
3.3.2
Intake and Exhaust Systems
Silencing of the engine was achieved using a low back-pressure free-standing industrial silencer, with engine back-pressure being set by two orifice plates in the exhaust system itself. The diameter of these orifice plates was determined by engine modelling work carried out prior to the testing programme, and have been decided upon as though they would yield a back pressure of 125kPa (gauge) at the exhaust port at 6000rpm. This value was considered representative of a turbocharged engine. Additionally, volumes were put into the exhaust and intake systems to mimic the presence of runner and collector volumes in a multi-cylinder engine. The subject of matching single-cylinder engine performance to multi-cylinder units will be discussed further in the Section 5.2.
3.4
Test Procedure
In order to fully research the effect of intake air temperature at fixed charge-air density ratios a test matrix was drawn up. The variables for this matrix are shown in Table 1.
Table 1: Test variables defining the number of test points Variables Plenum Density [kg/m3] Plenum Temperature [0C]
1.2 - 1.4 - l.6 1.8 20 - 30 - 40 - 50
No. of sites 4 4
There are 16 potential test points for this investigation. However, a significant limitation was set using the criterion that no test point should be attempted if the ignition timing had to be retarded beyond top dead centre; this meant that some of the points in Table 2 were not achievable. Nevertheless, the test procedure is considered to be robust and will yield significant amounts of data; even the fact that a point is unachievable because of unacceptably late ignition timing is considered a significant finding. All of this data was gathered at 2000rpm and a fixed AFR of Lambda 0.9, the use of the separate CAC rig with no catalyst fitted to the engine ensuring that fuel enrichment (for component protection) was not needed and therefore a true assessment of the effect of intake air temperature could be obtained. At each test point the procedure was to obtain the knock-limited spark advance (KLSA) using the Indiset's in-built knock intensity indicator set to a limit of rate of cylinder pressure rise of 2bar per degree as a knock condition and to record data at this
74
condition. The ignition was then retarded in one or two degrees steps from KLSA for a further four points (giving KLSA, _1°, _2°, _4° and _6°). At each of these all data was logged.
4.
RESULTS
Test results will be presented with plenum charge air temperature as the control variable, for fixed charge-air density. For brevity, four graphs are included, showing knock-limited spark advance, uncorrected BMEp4, BSFC and finally exhaust back pressure. Missing points on for the higher two constant densities are a result of the test criterion that no point should be attempted if the spark advance during a test loop should have to be retarded beyond top dead centre. While interpreting these results it must be remembered that the engine is not driving the charging system: the power to drive it is supplied externally. 25
f--
r---
-r-
1--= r----
f.~
-.
.~
15
..
t---
f--
20
25
30
35
40
45
55
50
Plenum Temperature I [degC] 1~1.2 ~1.4 ....... 1.6 -+-1.81
Fig. 3: Knock-limited spark advance versus charge-air temperature for different charge-air densities (in kg/m3) 15 - - ' - .
-~
14 13
'"
1-
=--.
--
'----
-
-
~
t----
-
-:-. .~ ~--
-
.-
~
-.
..
9
-f15
20
25
30
35
40
45
50
55
Plenum Temperature I [degC] 1-0-1.2 -0-1.4 -+-1.6 -+-1.81
Fig. 4: Uncorrected BMEP versus charge-air temperature for different charge-air densities (in kg/m3 )
4 Uncorrected BMEP is exclusively used in this investigation because of the conditioned nature of the air being supplied to the engine; with sufficiently accurate control of temperature and pressure no correction factor is deemed necessary.
75
280
265 260
15
20
25
35
30
40
45
50
55
Plenum Temperature' [degC] 1-0-1.2 -0-1.4 ....... 1.6 -+-1.81
Fig.5: BSFC versus charge-air temperature for different charge-air densities (in kg/m3 )
10 15
25
20
45 30 35 40 Plenum Temperatura' [degC]
50
55
1-<>-1.2 -0-1.4 ...... 1.6 ~1.al
Fig. 6: Exhaust back pressure versus charge-air temperature for different chargeair densities (in kg/m3) 5.
DISCUSSION
5.1
Extension of the Knock Limit
The primary purpose of performing these tests was to investigate the theory that overcooling the charge air would allow an extension of the knock limit in a pressure-charged engine. Consequently various factors have been employed to gain a very poor base engine knock limit while using commercially available gasoline; these factors include employing a high compression ratio for a PFI engine and a relatively large bore diameter for a pressure charged engine. Immediately apparent from the results is that the knock limit of the engine has indeed been extended by the use of over-cooled air: this is shown in Figure 3 by the greater amounts of spark advance the engine can support at lower charge air temperatures for all of the charge air densities investigated. Specifically, for the highest two densities, an extra 3.5 0 of spark advance can be supported for a temperature reduction from 40 to 20°C at the 1.6kg/m3 density point, and 1.5 0 at 1.8kg/m3 for a temperature reduction from 30 to 20 0 • While these gains not massive, it must be remembered that the engine is port-fuel injected and has a relatively high compression ratio of 10: 1, and so these improvements are considered significant, representing as they do approximately 2 degrees of spark advance per 10 reduction in charge air temperature across all the data points in Figure 3. However, at first investigation, there appear to be no parallel 0
76
significant and consistent benefits in the form of increased BMEP and reduced BSFC apparent (Figures 4 and 5 respectively). The reasons for this are believed to include the fact that charge air density was the fixed variable, and that as the air temperature was increased, so the pressure had to be increased to compensate. This in turn means that the positive work (externally supplied) on the piston during the intake stroke increases. Simultaneously, as the air temperature is increased, when the fuel is injected, more fuel will evaporate in the intake port and less on the engine structure, and so have a greater charge-cooling effect, thus offsetting the effect of a reduction in combustion efficiency caused by the retarded ignition timing. There are also likely to be some gas-dynamic effects arising from attempting to force a single-cylinder engine to mimic a multi-cylinder one through the use of exhausts orifices; this will affect the cylinder charging and so vary the BMEP. The effect and compromises of this approach is discussed later in this section. The combined effect of these various phenomena is that for the two lowest densities the BMEP can be seen to increase slightly with the charge air temperature, which may seem counter to the normal characteristic of a 'conventional' pressure-charged engine in that as intercooler effectiveness is increased, so engine performance improves. Since the engine is not having to drive the charging system, the penalty that this would incur in BMEP is not present. In the case of a full engine, intercooler effectiveness can be traded against boost pressure, the reduction of which means that less work has to be done by the charging system in the form of a reduced pressure ratio across the compressor, and this is the biggest factor determining its power consumption. Furthermore, in such a 'conventional' engine, boost pressure, not charge density, is normally the control variable, and so, as the intercooler effectiveness is improved, for a fixed boost pressure, the positive work on the piston will not change. Instead, the mass air flow through the engine will increase and a power increase will ensue (assuming that the spark advance can be held constant or increased due to the greater charge-cooling effect of a higher intercooler effectiveness). Thus, from the strong trend in the results that the knock limit can be extended using over-cooled air it is reasonable to infer that this approach could be used to increase the compression ratio of the base engine for any given BMEP level, to the benefit of partload fuel consumption. Therefore, further investigation of turboexpansion charging systems is warranted from the perspective of improved vehicle fuel consumption.
5.2
Manifold Boundary Conditions in Single- and Multi-Cylinder Engines
The data presented above shows somewhat low levels of BMEP for the engine considering the density of air applied to the intake. This is considered likely to be due to several factors, one of which is the increased friction of a single-cylinder engine in relation to a multi-cylinder one. This is especially the case for this configuration which has seven engine-bearings (three on the crankshaft and four in the balancing system). Against this must be set the fact that because of the AVT system, the engine is not driving the valve gear directly. In order to evaluate the relative magnitude of these effects cylinder pressure analysis is planned as future work. Of more pertinent interest with regard to this investigation is the effect on combustion of residuals, as discussed in Section 1. In thc data presented here, the cycleaveraged pressures across the combustion chamber show a positive gradient from intake to exhaust, and so it could be assumed that the engine is well scavenged (Figures 6c, 7c, 8c and 9c). However, there are additional issues when attempting to configure a singlecylinder engine to give equivalent operating conditions to a multi-cylinder engine. In
77
the present work the authors have employed representative volumes to mimic the gasdynamic effects of multi-cylinder manifolds, and have set the mean back pressure to imitate a multi-cylinder engine. Doing this can lead to a greater pressure amplitude in the valve open period in the single-cylinder engine in order to maintain the same mean pressure as that generated by a multi-cylinder engine. The scavenging will thus be inhibited and a greater mass of residuals will be left in the single-cylinder combustion chamber than is the case with the multi-cylinder engine. It is intended to investigate this phenomenon further and, if necessary, to adjust the size of the hole diameter in the exhaust orifice plates to gain a better match with the operating conditions of a representative multi-cylinder engine. It is likely that this will improve the knock limit further and may heighten the sensitivity of the system to reducing the charge air temperature.
5.3
Performance of Charge-Air Conditioning Rig
With regards to the manner of presenting the results, the use of uncorrected BMEP is believed to be vindicated by the accuracy with which the charge-air conditioning rig could hold the demanded temperatures and pressures. This is shown by the fact that, across the entire range of points tested, analysis of the density in the plenum showed that the maximum deviation from the desired value was +0.7%, with 70% of test points showing a deviation of less than ±0.3%. This is considered a very good result that vindicates the use of observed data in this work.
6.
CONCLUSIONS
6.1
Over-cooling the charge air in the manner of a turboexpansion system has been shown significantly to be capable of significantly extending the knock limit. This can be used to either increase the mass flow through the engine (to the benefit of specific output and therefore 'downsize' factor) or to increase the compression ratio (to the benefit of thermodynamic efficiency). Both approaches can be used to improve vehicle fuel consumption.
6.2
The test single-cylinder engine was configured with volumes in the intake and exhaust systems to assist in mimicking a multi-cylinder engine, and mean backpressure was adjusted accordingly. However, further investigation is warranted to obtain a better match between engine types through a similar pressure-time history during the overlap period. It is believed that this will reinforce the improvement in knock limit further.
6.3
The charge air conditioning rig employed in this test can supply air to an engine up to an amount suitable for 450bhp, and has been shown to be capable of controlling air density at the far lower air flows reported here to within 0.7% accuracy.
7.
FURTHER WORK
The authors intend to conduct a comparison of GDI with PFI and the merits of a turboexpansion approach to both. This work will provide a bank of data to assist in determining charging system performance requirements for future engines. This work
78
will be supported by a multi-national fuel company in the supply of test fuel, so that fuel variation is removed from the analysis ofthe results of the forthcoming work.
8.
ACKNOWLEDGEMENTS
The authors wish to thank the directors of Group Lotus for permission to publish this paper. The work of all the designers who have worked on the single-cylinder engine and of Pete Brooks, Steve Gedge and James Young, Lotus Powertrain Technicians, is gratefully acknowledged. Also, Allan Saville contributed significantly to discussions regarding the manifold conditions used in the single-cylinder engine and outlined the potential difficulties.
9.
REFERENCES
1.
UN Kyoto Protocol website: http://unfccc.intiessential background Ifeeling the heatlitemsl2879.php Schwaderlapp, M., Habermann, K and Yapici, KI., "Variable Compression Ratio - A Design Solution for Fuel Economy Concepts". SAE paper number 2002-01-1103,2002. Drangel, H. and Bergsten, L., "The new Saab SVC Engine - An Interaction of Variable Compression Ratio, High Pressure Supercharging and Downsizing for Considerably Reduced Fuel Consumption". Aachen Colloquium, October 2000. Konig, G. and Sheppard e.G.W., "End Gas Autoignition and Knock in a Spark Ignition Engine". SAE paper number 902135, 1990. Blunsdon, e.A. and Dent, J.e., "The Simulation of Autoignition and Knock in a Spark Ignition Engine with Disk Geometry". SAE paper number 940524, 1994. Maly, R.R, Klein, R., Peters, N. and Konig, G., "Theoretical and Experimental Investigations of Knock Induced Surface Destructions". SAE paper number 900025, 1990. Crooks, W.R, "Combustion air conditioning boosts output 50 per cent". CIMAC Vol. 15, 1959, pp. 475-494. Turner, J., Pearson, R., Bassett, M., and Oscarsson, J., "Performance and Fuel Economy Enhancement of Pressure Charged SI Engines through Turboexpansion - An Initial Study". SAE paper no. 2003-01-0401, 2003 and also SAE 2003 Transactions, September 2004. Turner, J., Pearson, R., Bassett, M., Blundell, D., and Taitt, D., "The Turboexpansion Concept - Initial Dynamometer Results". SAE paper number 2005-01-1853,2005. Taitt, D.W., Garner, C.P., Swain, E, Blundell, D., Pearson, R.J. and Turner, J.W.G., "An Automotive Engine Charge-Air Intake Conditioner System: Analysis of Fuel Economy Benefits in a Gasoline Engine Application". Submitted to Proc. Instn Mech. Engrs Part D. Krebs, R, Boehme, J., Dornhoefer, R, Wurms, R., Friedmann, K, Helbig, J. and Hatz, W., "The New Audi 2.0T FSI Engine - The First Direct Injection Turbo-Gasoline-Engine from Audi". Vienna Motor Symposium, April 2004. Ranini, A. and Monnier, G., "Turbocharging a Gasoline Direct Injection Engine". SAE paper number 2001-01-0736, 200 I. Lang, 0., Geiger, J., Haberman, K and Wittler, M., "Boosting and Direct Injection - Synergies for Future Gasoline Engines". SAE paper number 200501-1144,2005.
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Duchaussoy, Y., Lefebvre, A. and Bonetto, R., "Dilution Interest on Turbocharged SI Engine Combustion". SAE paper number 2003-01-0629,2003. Turner, l.W.G., Kalafatis, A. and Atkins, c., "The Design of the NOMAD Advanced Concepts Research Engine". Accepted for publication at SAE Congress, 2006. Douaud, A. and Eyzat, P., "Four-Octane-Number Method for Predicting the Anti-Knock Behaviour of Fuels and Engines". SAE paper number 780080, 1978. Turner, l.W.G., Pitcher, G., Burke, P., Garner, c.P., Wigley, G., Stansfield, P., Nuglisch, H., Ladommatos, N., Patel, R. and Williams, P., "The HOTFIRE Homogeneous GDI and Fully Variable Valve Train Project - An Initial Report". Accepted for publication at SAE Congress, 2006. Turner, l, Kenchington, S. and Stretch, D., "Production AVT development: Lotus and Eaton's electrohydraulic closed-loop fully variable valve train system". 25th Vienna Motor Symposium, 2004.
© Lotus Engineering
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TURBO-COOLING APPLIED TO LIGHT DUTY VEHICLE ENGINES C D Whelan and R A Richards WDLltd, UK The concept of charge air-cooling by turbo-expansion applied to internal combustion engines was developed in the 1950s. It was successfully applied to natural gas fuelled power generation engines, enabling useful power increases and providing protection from detonation effects caused by varying gas properties. On light duty engines for passenger cars and light trucks, the application of turbocooling offers a range of thermodynamic opportunities. Any turbocharged engine can be turbo-cooled, although the potential benefits vary with both the combustion type and the application. The key principle is to achieve increased intake charge density in combination with reduced charge air temperature. This can offer benefits in respect of detonation and thermal loading, with control over combustion rates and in-cylinder pressures. On gasoline engines there is an increasing use of turbocharging. This opens up opportunities for turbo-cooling for improvements in performance, fuel economy and emissions. The majority oflight-duty diesel engines are turbocharged. Their maximum power potential is dictated by a combination of boost system and structural limits that can potentially be extended by turbo-cooling. This paper presents a study into the application of turbo-cooling to light duty engines. The basic concept and historical work are presented. 1D performance simulation and simple thermodynamic analyses are used to present a scenario for a gasoline light duty engine system. The potential for improvement in power, fuel economy and emissions are explored, both qualitatively and quantitatively. Concept designs and potential production applications of turbo-cooling are also reviewed, together with a roadmap for development and market introduction.
BACKGROUND The concept of charge air-cooling by turbo-expansion applied to internal combustion engines was developed in the 1950s and 60s [1,2]. It was successfully applied to natural gas fuelled power generation engines, enabling useful power increases and providing protection from detonation effects caused by varying gas properties. Little or no work in this field appears to have been performed during the 20-year period after these studies and applications. In the mid-1980s the author carried out experimental work [3] as part of a Formula 1 racing engine programme. The regulations then allowed unlimited boosting of a gasoline 1.5 litre 4-stroke engine. A detonation-limited combustion system, in combination with significant excess exhaust energy, encouraged investigation of the turbo-expansion concept. Rig tests were performed to establish feasibility, and engine designs prepared. The system was never vehicle tested. In the early 1990s a theoretical study [4] was made into the use of charge refrigeration for emissions improvements on medium and heavy-duty diesel and gasoline vehicle engines. The study concluded that the application was feasible, and the effects of various component efficiencies were explored. Again little work was performed between this study and 2003, when a modelling and design study, based on a mechanically driven positive displacement system, was published [5].
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In 2003, the authors considered that current regulatory, technical and market trends in light duty powertrain development warranted a further study into turbo-cooling systems. OPERATING PRINCIPLES The operating principle of turbo-expansion is well known; the flow path and main components are shown in Figure 1. The sequence of 'events' for the charge air is:
Compression by the compressor of the main turbocharger Cooling by the first charge air cooler (CAC1) Re-compression in the compressor of the turbo-expander Cooling by the second charge air cooler (CAC2) Expansion through the turbine of the turbo-expander
CACI
Engine
Figure 1: Turbo-Cooling Concept Layout and Gas Flowpath
Air expansion through the turbo-expander turbine provides the power to drive its compressor. The overall loss of enthalpy through this second stage of compression and expansion is compensated for by increased work in the first compressor. The net effect on the engine is increased exhaust pressure into the main turbine. In summary, for a given charge air flow and density, which generally defines the indicated engine power, the intake pressure and temperature can be lowered at the expense of high primary boost pressure and increased engine pressure gradient (and pumping loss). The benefits of turbo-expansion result from the various ways of exploiting reduced intake pressure and temperature. Before considering a modem application of turbo-cooling, it is important to understand the current technical and market trends that would support its introduction onto road vehicles.
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CURRENT TRENDS AND OPPORTUNITIES
On engines for passenger cars and light trucks, turbo-cooling offers a range of thermodynamic opportunities. Any turbocharged engine can be turbo-cooled, although the range of potential benefits varies with both the combustion type and the vehicle application. The key principle is to achieve increased intake charge density in combination with reduced charge air temperature. This can offer benefits relating to output, detonation, thermal loading, and engine-out emissions, linked with additional control over combustion rates and in-cylinder pressures. Some specific recent technical trends are worthy of review in the context of turbo-cooling. Downsizing
For passenger cars, drive-cycle based fuel economy can be improved by reducing engine capacity for a given power and vehicle weight. The drive-cycle power demand is fixed, so with a smaller engine the load (Pme) increases, resulting in improved specific fuel consumption. These gains are not always realised in 'real-world' driving conditions. European vehicle manufacturers are aggressively pursuing engine downsizing, in support of the ACEA C02 target reductions for 2008 and 2012. There are penalties resulting from engine downsizing that differ between gasoline and diesel engines. On diesel engines, higher engine loads significantly increase engine-out NOx emissions. There is also some deterioration in low speed performance 'feel' due to the transition from near normally aspirated torque to very high levels when the boosting system becomes effective. A potential solution to this problem is the application of electrically driven boost 'augmenters' (e-boosters) and these will be reviewed later in the context of turbo-cooling. On gasoline engines, the low speed torque transition problem is similar to that of diesels, but less severe. Additionally there are fuel economy issues resulting from 'realworld' driving. This is due to enriched air-fuel mixture at high engine loads, necessary to control thermal loads and combustion stability within acceptable levels. With downsized engines it is possible to regularly cruise the vehicle in the enriched operating zone of the engine, with a resulting deterioration in fuel consumption and range. There are also emissions issues, due to the deviation from a stoichiometric air-fuel ratio, leading to imbalance in the 3-way catalyst. The study will indicate how the increasing application of downsized engines would benefit from turbo-cooling. Specific Power
Whilst technically closely linked to downsizing, increases in specific power also result from other influences. For example, the selling price of a passenger car tends to be based more on absolute power than on engine capacity. This results in a wide range of powers for given engine sizes, and the need to increase the power for a given package size. It is much easier to develop a medium sized car with a medium sized engine of high power than it is to install a larger capacity engine. These trends, applicable to both gasoline and diesel engines, are driving the specific power of all engine sizes. Increases in specific power are subject to well-known constraints. For both engine types, intake pressure and temperature are key influences on maximum power. For gasoline engines this is due to their effects on combustion stability, thermal efficiency and thermal loading. On diesel engines they influence the mass and durability of the engine, via mechanical loading (max. cylinder pressure) and thermal stresses (high-cycle fatigue related to individual firing events). 83
Emissions and EGR Although emissions influences run through the subjects presented above, it is worth considering the special needs of medium and heavy-duty diesel engines. These engines must conform to emissions limits over a wide range of engine speeds and loads, resulting in the general need to operate with exhaust gas recirculation (EGR) at very high engine loads. This can lead to two very specific problems. On these engines the pressure gradient across the engine is broadly positive, as opposed to highly negative on small engines, resulting in the need for special devices to induce an EGR flow from exhaust to intake. Also, significant EGR cooling is required to contain the intake charge temperature to acceptable levels. This is compounded by the need to increase charge air boost pressures to compensate for the air displaced by the EGR; a vicious circle. Turbo-cooling potentially offers relief, both in terms of engine pressure gradient and intake temperature. High load EGR has also been considered for highly boosted gasoline engines, as a means of controlling combustion stability without fuel enrichment. This approach brings similar problems of controlling intake manifold temperature. The current approach is EGR cooling with engine coolant, which is inefficient, demanding of package space, and increases the size and mass of the vehicle cooling system. All these effects are counter-intuitive to the concept of light, efficient downsized engines and vehicles. Turbo-cooling potentially addresses these issues, in terms of engine pressure gradient, intake temperature and EGR conditions. Future Combustion Systems This paper does not aim to review the vast body of work performed on pre-mixed autoignition combustion systems. In the context of turbo-cooling, the key issue is intake temperature control. For combustion stability in HCCVCAI mode, and the transitions to/from 'normal' combustion operation at mediumlhigh loads, accurate and precise control of intake charge temperature is a 'must'. Turbo-cooling can play a significant and useful role in the future application of these combustion systems.
CASE STUDY As discussed above, there is a range of potential applications of turbo-cooling to vehicle engines. This paper will present just one of these; namely the downsized gasoline engine. Nevertheless, this simple study will clearly indicate some elements of turbocooling that may be applied to other engine types and vehicle applications. Approach The downsizing study approach is based on the following scenario: A turbocharged 1.6 litre engine replaces a normally aspirated 2.0 litre passenger car engine; an exercise performed by numerous manufacturers and investigators. The power is typically 110kW. Power of the 1.6 litre engine is increased to the practical limit of the base design, to provide a high performance vehicle and wider product range. The maximum power is typically 150kW. The second step above is the key element of this study. The effect of increasing the power of the conventional turbocharged engine is compared with a similar power increase using turbo-cooling.
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Modelling Technique The case study presented below has been perfonned using AMESim ID modelling software. This multi-physics modelling platfonn uses energy conservation and acoustics to link individual 'components' (standard and bespoke) to create system models. It includes all the elements necessary for thennodynamic simulation of a complete turbo-cooled powertrain, including: Gas flow path (properties and geometry) Turbo-machinery (scalable maps and non-dimensional parameters) Heat exchangers (maps or heat transfer functions) Engine model (mean-value or fully-defined) Control loops and functions Vehicle definition and mission profile operation, as required The circuit model developed for the turbo-cooling system analysis is shown in Figure 2. The TCS device modelled is a typical turbocharger, complete with variable area turbine. Production TCS systems would also feature a by-pass loop for engine operation without turbo-cooling; this has been omitted from this study to avoid complexity. Some simple control features have been incorporated; these include functions for CAC perfonnance matching and turbine area control according to various charge air inputs. A typical strategy is to modulate main turbine and TCS turbine areas to maintain constant intake pressure or temperature. This study uses the 'mean value' engine model, with fixed efficiency and heat rejection, but with varying pumping loss. This approach reduces complexity in order to concentrate on TCS operation. For specific applications the available full 'emptying and filling' engine model would be used.
Turbocharged Gasoline Engine with TCS MAIN TURBOCHARGER
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CAC1 TURBO-EXPANDER
._._._._._-_._._.-.
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Figure 2: AMESim Turbo-Cooling System Model
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Process The modelling process was to: Build a 'Baseline' model, using test data and information from real engines Use real turbo-machinery and charge air cooler performance and efficiency data Fix the engine speed (6000 rev/min) Set the nominal power of the engine as 11OkW; equivalent to a 2.0 litre normally aspirated engine Calculate the Baseline performance up to 150kW, by varying the turbine flow area Calculate the TCS engine performance over the same power range, by varying the main turbine flow area Control TCS turbine area to maintain constant engine air intake (manifold) temperature Results The results of the study are shown in Figures 3 to 8, and will be presented and discussed in sequence. Figure 3 shows the engine charge air intake (manifold) temperature for both engine configurations. For the Baseline engine, manifold temperature is near 50°C at 11OkW, rising to over 65°C at 150kW. A larger CAC would partially mitigate this, but with diminishing returns in package size. For the TCS build, the manifold temperature is constant at 30°C as boost pressure and power increase. With appropriate matching andlor use of a by-pass, rising or falling manifold temperature can be delivered, with increasing boost pressure. 70
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Engine intake and exhaust manifold pressures are shown in Figure 4. For a given power, the TCS intake pressure is lower and the exhaust pressure higher than the Baseline engine. The greater pressure difference across the TCS engine increases pumping work, and this will be discussed further.
86
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Figure 5: Main Turbine and Engine Pumping Power Figure 6 shows the pressure ratios of the Baseline and TCS main turbocharger. The performance of these components is based on production performance maps. Figure 6 clearly indicates the increased pressure ratios of the TCS main turbocharger compressor and turbine, but these do not reach extreme levels. Not shown in the figure, the TCS turbo-expander pressure ratios are approximately 1.1 and 1.4 for the compressor and turbine respectively.
87
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Engine Power [k \IV]
Figure 6: Main Turbocharger Pressure Ratios
In Figure 7, the temperature of the charge air throughout the complete TCS intake system is shown. The ambient temperature (Main Compressor Tin) used throughout the study is 27°C. Charge air from the main compressor (Main Compressor Tout) is cooled by CAC! before entering the turbo-expander compressor (Turbo-Expander Compressor Tin). The temperature rises during compression to (Turbo-Expander Compressor Tout), and the air is then cooled again by CAC2 before entering the turbo-expander turbine (Turbo-Expander Turbine Tin). Expansion through the turbine further reduces the temperature (Turbo-Expander Turbine Tout) before the charge air enters the engine. In this example, the charge air is cooled back to the initial temperature at entry to the main compressor. Because the turbo-expander power increases as mass flow rises, the cooling effect automatically increases as engine power rises. 2OO-r------.---.----.------.----.
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Figure 8: Compressor Power and CAC Heat Rejection
TCS system compressor powers and CAC heat rejection are presented in Figure 8. Heat rejection from CACI is significantly less than the heat input represented by the main compressor power. The temperature rise imparted by the secondary compression in the turbo-expander compressor enables more of this heat to be rejected in CAC2, as represented by the difference between CAC2 heat rejection and the turbo-expander compressor power. It is worth noting that, within the turbo-expander, the heat rejected from the air approaches twice that of the power absorbed, thus providing a 'coefficient of performance' of approximately 2: 1. Summary The results of the study are summarised in Table 1 below. The effect ofTCS operation is highlighted by comparing the Baseline engine at a 'production' power of 11 OkW (Baselinel), then increased to a maximum of 150kW (Baseline2). The Baseline2 and TCS engine performance at 150kW are then compared. There are two versions of the TCS modelling conditions. TCS2 uses the same equivalence ratio as the Baseline engine at 150kW. TCS3 has the equivalence ratio reduced to align the pre-turbine temperature with that of the Baseline2 engine. This is a simple way of illustrating the potential to reduce fuel enrichment with a TCS system, due to reduced thermal loading. It should be recognised that the calculated performance does not include any combustion system effects of reduced intake pressure and temperature, or any reoptimisation. The calculated specific fuel consumption is based on constant in-cylinder conditions but incorporating the effect of increased backpressure on pumping losses. Thus the predicted fuel consumptions are very similar for the 150kW baseline and TCS3 builds. In a separate study the combustion system investigation will be confirmed, with re-optimised compression ratio, air-fuel ratio and ignition timing.
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Table 1: Baseline and TCS Engine Performance
Baseline1 Engine Engine Speed Brake Power Brake Torque Equivalence Ratio BSFC Inlet Manifold Density Pumping Loss Inlet Manifold Temperature Exhaust Manifold Temperature Inlet Manifold Pressure Exhaust Manifold Pressure Main Turbocharger Compressor Pressure Ratio Turbine Pressure Ratio Turbine Power TCS Turbo-expander Compressor Pressure Ratio Turbine Pressure Ratio Turbine Power CAC1 Air Inlet Temperature Air Outlet Temperature Heat Rejection Effectiveness CAC2 Air Inlet Temperature Air Outlet Temperature Heat Rejection Effectiveness
(rev/min) (kW) (Nm)
(-) (g/kWh) (kg/m') (kW) (OC) (OC) (barA) (barA)
(- ) (-) (kW)
(- ) (- ) (kW) (OC) (OC) (kW) (- ) (OC) (OC) (kW) (- )
Baseline2
TCS2
TCS3
6,000 110 175 1.17 334 1.62 2.9 54 945 1.52 1.88
6,000 150 238 1.17 337 2.23 4.5 66 952 2.17 2.73
6,000 150 238 1.17 350 2.32 10.5 30 933 2.02 3.33
6,000 150 238 1.13 339 2.31 10.4 30 951 2.01 3.31
1.60 1.58 7.6
2.29 2.04 19.9
2.76 2.46 27.2
2.76 2.43 27.3
NA NA NA
NA NA NA
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1.11 1.42 3.4
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142 66 13.1 0.78
178 72 19.2 0.79
178 19.2 0.79
NA NA NA NA
NA NA NA NA
90 49 7.4 0.82
90 49 7.4 0.82
72
EXPERIMENTAL PROGRAMME Work has commenced on a rig-based experimental programme. The overall aims are to demonstrate the performance of the turbo-expander, and to generate input data for a dedicated expander design and overall engine performance matching. A standard production turbocharger was procured, and mounted into a special test rig. The main components of the rig comprise: Independent high boost air supply, with fully variable flow, pressure and temperature Air pre-cooler (air-cooled vehicle charge air cooler) Turbocharger, forming the turbo-expander unit Intercooler mounted between turbocharger compressor outlet and turbine inlet (aircooled vehicle charge air cooler) Outlet throttle, for overall flow control Figure 9 shows the rig installation with the major components as described above.
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Figure 9: Experimental Turbo-Expander Test Rig Results Sample results from the test programme are presented in Figure 10. In this experiment, the mass flow was varied at constant intake conditions. The mass flow range chosen was representative of automotive gasoline and diesel engines. The key result is the overall temperature drop due to the compression-cooling-expansion process, coupled with the net loss of pressure. A convenient way of assessing the overall perfonnance is the ratio of outlet to inlet density. This allows calculation of the increased intake pressure that would be required to maintain constant exit density and hence nominal engine power. The temperature drop through the system can clearly be seen, resulting in exit temperature of between 28°C and O°C over the flow range.
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APPLICATIONS, PRODUCTION AND INTEGRATION
This preliminary study has indicated potential production applications of a modem turbo-cooling system; some of these options are discussed below. Applications
The case study, of a downsized turbocharged gasoline engine, showed the potential for useful power increases with reduced intake manifold temperature. The next step for this type of application would be to investigate the effects of: High load EGR: an increasing area of investigation to improve combustion stability and reduce fuel enrichment. Improved fuel economy at high speed cruise: lower intake temperatures reduce thermal load and improve the detonation margin, thus reducing the fuel enrichment requirement. 'Low Pressure' EGR routing: in the extreme case of turbo-expansion, the postturbine pressure in a vehicle system is higher than the intake manifold pressure. This allows an EGR route from the turbine exit to the engine intake manifold, with consequential benefits from gas temperature reduction and increased turbine work. Using the basic thermodynamic principles of turbo-cooling, other engine applications worth investigating include: Light duty diesel: increased specific power due to more favourable intake air pressure/temperature trade-off, plus improved high load EGR operation for light commercial vehicles. Heavy duty diesel: high load EGR flows can be increased due to the reduced charge temperature and more favourable pressure gradient across the engine. HCCVCAI operation and combustion mode transition: precise and accurate modulation of charge air temperature by VG control of turbo-expander, with controlled bypass operation. Production
The system design principles are based on current turbo-machinery and charge air coolers. Recent developments in electrically powered 'boost-augmenters' (e-boosters) have shown the potential for a compact lightweight turbo-expander. These design principles would allow the construction of a turbo-expander with the following features: Low pressure ratio Lightweight materials (aluminium and plastic) Sealed bearings (rolling contact or air type) Integration
The turbo-expander forms part of an independent system that can be added onto an existing powertrain installation. Subject to charge air cooling requirements, location and packaging within the vehicle are flexible. There is little or no impact on the external arrangement of the base engine. Beyond this simple and cost effective concept design, future enhancements could be made, namely: Integration of the turbo-expander with the second charge air cooler, by direct mounting on the header tank of the cooler Inclusion of a bypass circuit to allow close control of intake conditions Integration with e-booster components and design
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Integrating a turbo-expander with an e-booster provides an interesting opportunity to improve both low speed torque and high-speed power with a single system. A turbine is added to the existing e-booster compressor-motor assembly. At low engine speed the e-boost compressor is electrically driven. At high speed and load the unit operates as the stand-alone TCS system. This concept would require a 'downstream' ebooster design, with additional sealing to contain the air pressure from the main turbocharger compressor. CONCLUSION The aim of this study has been to investigate the potential for turbo-expansion applied to meet the needs of modem vehicle engines. A range of potential benefits, applications and opportunities has been presented. The case study, on a downsized gasoline engine, has quantified differences in the boosting system characteristics between TCS and standard turbo charging as power levels increase. The case study confirms the basic feasibility ofTCS application to a modem engine; specifically the ability to reduce intake pressure and temperature at increasing engine power. For gasoline engine applications the exploitation of the reduced intake charge pressure and temperature depends on re-matching of the combustion system, particularly the compression ratio, air-fuel ratio and ignition timing. The thermodynamic and fluid flow results from the case study have also indicated potential for emissions and EGR flow/cooling opportunities at high engine loads on diesel and gasoline engines.
NEXT STEPS Building on the authors' experience and this initial study, a range of activities are planned to further investigate and develop the potential of a modem TCS system. These include: Completion of the preliminary rig test work Further gasoline engine modelling using specific engine geometry and a 'fullfeature' model Application studies on light and heavy duty diesel engines, concentrating on emissions reduction Motorsports applications for intake restricted gasoline and diesel engines Concept and detail design of a dedicated turbo-expander unit REFERENCES 1 W R Crooks, 'Combustion Air Conditioning Boosts Output 50 Percent', CIMAC AI5,1959. 2 M J Helmich, 'Development of Combustion Air Refrigeration System Enabling Reliable Operation at 220 psi bmep for a Large Four-Cycle Spark-Ignited Gas Engine' AS ME 66-DGEP-7, 1966. 3 C Whelan and N Hayes, 'Formula 1 intake refrigeration system, in combination with a turbo-compound system'. Cosworth Engineering ltd, internal report, 1984. 4 R C Meyer and S M Shahed, 'An Intake Charge Cooling System for Application to Diesel, Gasoline and Natural Gas Engines', SAE910420, 1991. 5 R W G Turner, R J Pearson, M D Bassctt and J Os cars son, 'Lotus: Performance and Fuel Economy Enhancement of Pressure Charged SI Engines through Turbo-expansion - An Initial Study', SAE2003-01-0411, 2003.
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A one-dimensional model for variable and fixed geometry radial turbines for turbochargers J M Lujan, J R Serrano, C Cervell6, F J Arnau CMT-Motores Termicos. Universidad Politecnica de Valencia, Spain. S Soltani Renault, France. ABSTRACT The paper presents a fluid dynamic model of fixed and variable geometry turbines. The aim of this model is to provide an efficient boundary condition to model turbocharged internal combustion engines with one-dimensional gas dynamic codes. The model is based on the measured characteristics of the turbine from its very conception. Nevertheless, it is capable of extrapolating operating conditions which differ from those included in the turbine maps, since the engines usually work within these zones. The presented model has been implemented in a one-dimensional gas dynamic code and has been used to calculate unsteady operating conditions for several turbines. The results obtained have been compared with on-engine measurements of instantaneous pressure evolutions upstream and downstream of the turbine. LIST OF SYMBOLS Abbreviations AFT FGT VGT
Angle Fixed Turbine Fixed Geometry Turbine Variable Geometry Turbine with moving stator blades
Latin symbols
a A c cp D Disp h M
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97
rexp
r
T TQ u
W
Expansion ratio Estatistic. Rate of variability explained by a correlation (%) Temperature (K) Torque (Nm) Blade tip velocity (mls) Work transfer rate (J/s)
Greek symbols
y a 17 co
Adiabatic exponent (cp/c y ) Stator blades angle e) Efficiency Angular velocity (rad/s)
Subscripts and superscripts
o 2
eff g
k limit
rot
5t T
Indicates stagnation conditions and inlet turbine conditions Indicates conditions between stator and rotor Indicates outlet turbine conditions Axial Effective Polytropic coefficient of the expansion in the rotor Polytropic coefficient of the expansion in the stator Value for wich the stator blades direct the flow tangential to the rotor Radial Rotor Indicates isentropic evolution Stator Indicates total conditions
INTRODUCTION
Turbocharging is a technique which increases the power of internal combustion engines and reduces specific consumption. However, the application of this technique poses coupling problems between the engine and the turbocharger. For example, at low engine speed with a small mass flow rate, a turbine with a high expansion ratio (i.e., with a small effective section of the exhaust gas passage) is necessary to supply the power needed to meet the compressor requirements. However, for high-speed engine operating points, a turbine with a larger effective area would be enough to supply the power required by the compressor. Therefore, a single turbine might fail to adapt correctly to all the working conditions of an engine. To solve this problem variable geometry turbines, capable of altering the effective area of the gas flow passage, can be used. A solution commonly used is to vary the angle of inclination of the stator guide blades, thus changing the effective flow area. Variable geometry turbines of this kind are referred to in this study as VGT. Another possibility is to change the width of the turbine stator gas flow passage by relocating the stator guide blades, with a constant angle, along an axis parallel to the rotor shaft. Variable geometry turbines of this second type are referred to in this study as angle fixed turbine (AFT). The uptake of variable geometry turbines by engineering firms has been a slow process, as it is difficult to manufacture them at a reasonable cost, and also due to
98
problems in achieving the required reliability of the variable geometry mechanism. However, many of the existing high speed direct injection diesel engines are equipped with variable geometry turbines. The reason lies in the wide operating range of such engines, in which this type of turbine allows, with the appropriate controls, to improve the transient response of the engine and to reduce the pumping losses in steady operation. Therefore, smoke emissions, NOx emissions (when combined with EGR) and specific fuel consumption are reduced in comparison with fixed geometry turbines. One-dimensional models are able to reproduce the global engine behaviour with reasonable computational costs; therefore, in this context the correct physical modelling of the variable geometry turbine provides a powerful tool for the design of the necessary matching between turbocharger and engine plus the required control strategies. On the hand the modelling of the turbine must take into account the fluid dynamic behaviour of the gas, that is, the boundary conditions to be set at the exhaust manifold end, so that the dynamic interaction between the cylinders and the turbine, as well as the flow evolution downstream of the turbine and along the rest of the exhaust system, are correctly computed. On the other hand the modelling of the turbine must take into account the energy conversion, that is, the production of mechanical energy from the gas expansion through the turbine stator and rotor. Presumably, the easiest way to attain the proposed objectives is the introduction of the turbine characteristic curves, as suggested by Benson (1) and found in the literature (2). However, these simulations require a wide range of previous measurements in order to characterise the turbine. Additionally, only quite recently turbine manufacturers have attempted to test them under pulsating flow conditions, as those found in real engine operation, and thus it will usually be necessary to relate the characteristic curves of the turbine under steady flow to its behaviour when coupled to the engine. This difference in turbine behaviour under steady flow and pulsating flow conditions has been studied by several authors over the last few years (3)-(5). Moreover, imposing the turbine characteristic curves always implies the need to interpolate and excludes any option to extrapolate, in addition to assuming a totally quasi-steady behaviour in the turbine, thus making it impossible to take into account mass accumulation in the turbine. Finally, the use of interpolation functions has the drawback of increasing the calculation time. The simplest model developed for a radial turbine was that proposed for a fixed geometry unit by Watson and Janota (6). Central to this model is the representation of the turbine as a nozzle located at the exhaust manifold outlet which reproduces the pressure drop across the turbine for a specific mass flow rate. In 1991, Payri et al. (7)(8) presented a model based on two nozzles in series, separated by an intermediate volume (figure 1). This geometry solves the problem of high expansion ratios and the calculation of the instantaneous pressure downstream of the turbine. Furthermore, it permits the accumulation of mass and the consideration of unsteady phenomena. The main hypotheses of this model are that the behaviour of the turbine is quasi-steady throughout the nozzles simulating the turbine; from both the fluid dynamics and the thermodynamic viewpoint, and that the pressure drop is constant across the stator and the rotor. The first hypothesis is common practice in the modelling of all the boundary conditions in the internal combustion engines. The second hypothesis facilitates the calculation of the effective areas of the nozzles representing the turbine stator and rotor. Moreover, it is quite representative of what occurs in a turbine with a reaction degree (R) of 0.5, as it is the case in turbines without guide blades in the stator, but with radial or straight blades, normally used in automotive applications (6).
99
Enthap hoo,h,.
00
Stator nozzle
10
-------- ----1-------------po
Rotor nozzle
-0 p=p1
Turbine inlet
c~.
"'2
Internal volume
Turbine outlet
p"p2
______________ :___l
Entro y
Figure 1: Thermodynamic evolution of the flow on a radial turbine and geometry diagram of the turbine model In order to calculate the effective areas of the nozzles, it is assumed that the mass flow through each one of them is the same as that passing through the turbine, as indicated in Eq, 1 and Eq, 2:
m.~ =Aeff·F(Pou,) PinO
F(POUI) = r.(POUl) Yrr. ~. [ 1_(POUI)(r-Xl r PinG
PinO
Eq.l
PinO
r
1
Eq.2
PinO
Here, Aelf is the effective area of the nozzle equivalent to the stator or the rotor, and "in" and "out" represent the inlet and outlet openings of the stator or the rotor, respectively. In spite of the flow inlet conditions being known, i.e. mass flow rate, pressure and temperature at the turbine inlet, it is necessary to define the pressure drop across the stator in order to determine the effective area of the nozzle. Subsequently, when the outlet conditions of the flow are known, the pressure drop produced in the rotor must also be defined in order to calculate the effective area of the nozzle representing the rotor. Consistent with the second hypothesis, the value of the intermediate pressure between the stator and the rotor (PI) is calculated by imposing the condition of an equal expansion ratio across the stator and the rotor. Only a few articles are found in the literature referring to the one-dimensional modelling of radial turbines with variable geometry. An example is the one presented by Kessel et al. (9) in which a turbine model was designed in order to obtain data to train a neural network aiming to simulate the behaviour of a variable geometry turbine. The model is based on a series of thermodynamic transitions which represent the processes occurring inside a radial turbine. In 1999, Nasser and Playfoot (10) presented a model for a radial turbine with moving blades. The mass flow rate through the VaT is calculated using the nozzle equation (Eq. 1) and by taking the cross section of the throat area of the stator blades as the effective area. As such, there is no distinction between the geometrical and the effective sections. 100
CO2
Figure 2: Velocity triangles of a radial turbine at the rotor inlet and outlet ESTABLISHING THE MODEL Before variable geometry turbines, it was commonplace to use the R = 0.5 hypothesis to model the behaviour of radial turbines used for turbo charging recipocating internal combustion engines, as these turbines were designed without guide blades in the stator and with radial blades in the rotor, thereby being consistent with that hypothesis (6). Thus, the model proposed by Payri et al. (8) would be a useful tool in modelling the behaviour of a variable geometry turbine if instead of assuming R = 0.5 an appropriate method could be found to calculate the pressure drop across the turbine stator and rotor, when the turbine operating conditions change. This paper proposes a method for calculating R and the pressure drops with the purpose to extend the validity of models based on two nozzles in series. In this way, once the pressure drop is established for each ideal nozzle, it is possible to calculate the effective area and to reproduce the pressure drops across the stator and rotor of the variable geometry turbine. A diagram of the velocity triangles of a radial turbine at the rotor inlet and outlet is shown in Figure 2. This figure also shows the nomenclature used for the angles. As it can be observed in the Appendix, from the definition of R and using the hypotheses that there is no swirl at the turbine outlet and that the radial component of the velocity at the rotor inlet is equal to the axial component at the outlet of the turbine (both common design hypotheses (6)), R can be expressed as shown in Eq. 3 (see Appendix): R =1- tan a, . co2 2
Eq.3
U,
Considering that the mass flow rate at the rotor outlet is a function of the geometric area at its exit (A 2 ) and of the gas conditions at the turbine outlet, Eq. 4 can be rewritten as: R = 1- tan a, . m'JiT2 Eq.4 2 ·u, .AzP2 If sufficiently precise measurements are available, Eq. 4 can be used to calculate R. Otherwise, it can be calculated as a function of the corrected variables supplied by the manufacturers, as follows. The velocity UI and the rotor exit area can be written as: _ 21'CND, _ u, -----1'CND, Eq.5 2 A 2
= 1'CDJ 4
101
Eq.6
Here, D\ and Dz are the external and internal diameters of the turbine rotor (see figure 1). Substituting equations 5 and 6 into Eq. 4 gives:
R-l tana1 m lR7; - - 27rD1N' 7rDi . P2
Eq.7
- - Poo 4 Poo
On the other hand, if cp is assumed to be a constant, the isentropic behaviour of the variable geometry turbine can be expressed in terms of temperatures as:
1Jr, =
Too - T20 Too -7;,
Eq.8
Taking into account the previous stated hypothesis of no swirl at the turbine outlet and assuming that Co""Cz, the following approximate expression can be established.
1Jrs =
Too - T20 To - 7; Too - 7;, "" Too -7;,
Eq.9
Now, solving for the exit gas temperature and considering that:
(r-Xl (_2 ) Poo P
T2s = Too
Eq.lO
the value of T2 is obtained as a function of the turbine entry conditions:
(r-Xlll =Too·f'(li..,;: ,1Jr,)
(
[
T2 = Too ;: -1Jrs 1-(li..) 00 Poo
Poo
Eq.ll
00
When taking into account Eq. 11 (obtained from Eq. 8), R can be expressed as:
.
lR'TooI'(ll, To ,1Jr,)
R = 1- tana1 .~. Poo Too Eq.12 27rD1N 7rDi P II 00 4 Poo The relationship between the turbine inlet temperature (To) and the inlet stagnation or total temperature (Too) can be obtained from the corrected variables. Their ratio is:
Too=1+y-l M2 To 2
Eq.13
0
The Mach number for the turbine inlet conditions can, in tum, be rewritten as: r+l
M _~_ mlRTo._1_ _ m' f9i(Too)2(r- l ) o - ao - PoAo ~tR1'o - Ao fr To
Eq.14
By substituting this value into Eq. 13, the following is obtained, that can be solved by an iterative process, giving an initial value of 1 to the temperatures ratio, which easily converge to a value lightly higher than 1:
Too 1'0
=1+ lR(Y-l)(m*)2(1'o0);:: 2y Ao 1'0
102
Eq.15
Now, introducing this result in Eq. 12 and rewriting it as a function of corrected variables yields: R =1- 2·tanal 9l. 1"(
2DD2 I 2
m* '/(..!2.. m*,j
N*
Poo
'
'"'o,r/r.
) Eq. 16
Therefore, once the measurable geometric parameters of the turbine and the turbine map for each position of the turbine are known, it is possible to calculate R for each opearing point. In the case of a VGT, it is not always possible to use the equations of the Appendix to calculate R. Indeed, in Figure 3 it can be observed that there will be a specific limit value (al:::::68°), for which the axis of each blade is tangential to the turbine rotor. For values of a] above this limit value, the flow exiting the stator is no longer directed towards the rotor, but to the intermediate volume between the stator and the rotor. Thus, the direction of the rotor inlet velocity is not dependent upon the angle of the stator blades, a], but upon the angle of the rotor blades, Pl. Therefore, the turbine can be viewed as one without guide blades in the stator, and assuming the presence of straight rotor blades and the remaining components of the design hypothesis, R can be assumed to be 0.5, as demonstrated in the Appendix. In order to apply the previous model, several tests were carried out on a VGT turbine (Table 1), and R was calculated from the measurements obtained using Eq. 16. The performance maps for the VGT were obtained using a specific test rig for turbochargers. This is fully described in (11). VGTopen
0.85 0.8 0.75
a:::
I
VGTclosed
0.85
~~~~~~~~~~~-q
0.7
°O~~
I
0.55 0.5 40
50
0.8 0.75
a:::
I I 60
0.7
°o~~
0.55 .0
.0
70
80
Q.....~Oo....O'~ ~ ?l~
"~
'~~O
U'b~
0.5 ~4~~6~111111!!18-~10~~1~2~1;""4-~16
DO
90
Corrected Mass Flow (kg/s~K1MPa)
Ue)
Figure 3: Different blade positions of a VGT and R obtained for the different measured points
103
Table 1: Characteristics of the turbochargers used to validate the model
Inlet Diameter (mm) Outlet Diameter (mm) Number of rotor blades Number of stator blades
ComE·l{m 39 38 6
VGT 33 38 9 11
Comp.AH 41 35 6
AFT 40 50 11 11
ComE·~m 95 63 7
FGT 58 66 12
0.42
0.61
0.42
0.62
0.77
0.83
AIR
0.75
.....
............. • ............ •..... OOmm ::1mm
0.75
~. mfSIo~="~i
i>2mm
.• ·i'c····· .................. ... .
0.50
•...'
~
,;,3mm
<>4mm :>6mm
0:::
0.50
A~ ~ ..... . ...............•.•
0.25
··················0·············· ............ .
'l$'~0;)
0:::
010mm
0.25
0.00
·0'
'@
000
+-~--.-~---,-~-.--~-r-~--.-~
o
2
4
6
8 10 AFT Disp (mm)
CD
0.00 -t---.------,,.-----,------,
o
5
10
15 20 m* (kg/s·...jKlMPa)
Figure 4: Diagram and view of an AFT. Variation of R versus the displacement of the turbine shaft and versus the turbine mass flow rate Since the angle of the blades in the VGT stator is directly related to the movement of the rack from the variable-geometry mechanism, the blades angle can be clearly determined once the rack position is established. An angle of the stator blades of 86° corresponds to the closed VGT and an angle of the stator blades of 42° corresponds to the open VGT. Measurements were carried out for seven constant positions of the VGT, and for each position several operating speeds and expansion ratios were tested. The results obtained are shown in Figure 3, where it can be observed that R increases when the VGT opens. In addition, R tends to 0.5 when the blades angle tends to be tangential to the rotor circumference (a\::;;68°). Additionally, an AFT, whose blades present a constant a\ angle (Figure 4) was tested coupled to an engine on a specific bench. This turbine consists of a moving rod to which the stator blades are attached and a vacuum pump which controls the position of the rod. The characteristics of this variable geometry turbine are shown in Table 1. In this case, for a given opening and turbine speed only one point was measured, and then R was calculated by using the average values of pressure and temperature measured on 104
engine tests. The results obtained are shown at the bottom of Figure 4. They clearly show that R is dependent on the motion of the turbine axis. It can be seen that, in contrast to what occurs in the VGT, R increases when the turbine closes. The cause of this behaviour can be the particular way in which the AFT reduces the stator effective section, so that it can be presumed that a lamination process is produced in the stator when the AFT is closed (AFT Disp = 0) and therefore R is increased. There is also a dependency with other turbine variables; for example, R generally decreases when the exhaust gas flow rate increases. The results plotted in Figure 3 show this trend for each position and each turbine speed, except in the case of VGT closed, where R = 0.5 has been imposed as hypothesis. It is interesting to notice that when the AFT is closed, R remains also virtually constant, since the flow rate in the AFT is almost constant for the different expansion ratios considered, (Figure 4). In Figure 4, it can be observed that R varies between 0.1 and 0.25 when the AFT is open (AFT Disp = 10). Therefore, the increament of relative speed should be reduced in the rotor and this would account for the very low efficiency that is usually observed for a completely open AFT. Calculation of pressure at the stator outlet
Once R has been obtained as a function of the corrected turbine variables and of some easily measurable geometric parameters, it is necessary to establish its relation with the intermediate pressure between turbine stator and rotor in order to apply the chosen model of two nozzles plus an intermediate volume. If R is defined according to equation A.l, and considering the gas to be a perfect gas, one may write:
J
Eq.17 1; = (1- R) . (T2 + R 1'a 1'a If the type of evolution occurring in the turbine was known, it would be possible, with the help of Eq 18, to relate the intermediate pressure to the temperature, and therefore, to R. Taking into account that in any thermodynamic evolution, a flow which traverses the turbine may start from certain initial conditions (Po, To) (the terminology established in figure 1 is still maintained) and may achieve certain final conditions (P2, T2), in such a way that PO>P2 and To> Tz. the polytropic exponent of the evolution can take values between 1, for the extreme case in which To=T2, and ~1.33, in the case of exhaust gases and isentropic evolution. Assuming that the process undergone by the gas in the turbine is adiabatic but irreversible, it is physically impossible for the polytropic exponent to take values below 1, since that would imply a temperature increase. Likewise, any value above 1.33 would imply a decrease of entropy in the final state when compared with the initial state. If the evolution experienced by the gas traversing the turbine was a polytropic evolution, with a constant polytropic index (n), it could be represented by:
P2 (T2 Po = 1'a
)%-1
n In(P2/ Po) => n-l = In(I;/1'a)
Eq.18
However, when the turbine is quite closed, the evolution of the gas inside the stator is much more irreversible (since when the passage area diminishes, the losses caused by friction increase) than the evolution inside the rotor. Therefore, it seems convenient to assume the hypothesis that the evolution across the stator and the evolution across the
105
rotor have different polytropic exponents. Thus, Equation 19 must be fulfilled for each evolution, assuming that i is constant throughout the process in the stator or in the rotor: Eq. 19 and for the total evolution from the turbine inlet to the turbine outlet, it can be established that: P2 Po
= P2
. PI PI Po
=
(7; J%-I .(7; JKI = (7; J%-I 7;
1'0
Eq.20
1'0
Since it is not possible to measure pressures and temperatures at point 1, which is located between the turbine stator and rotor, the polytropic exponents k and g cannot be directly obtained. However, it is possible to relate them to n, as shown in equation 21. k
g
( ~_~).lnT2 n-l g-l To
k-1
g-1
In[(1-R)~ +R]
-=-+
Eq.21
Now the problem is that there is only one equation, Eq. 21, with two unknown values, k and g. Therefore, it is necessary to consider an additional hypothesis. To do this, the behaviour of the VGT and the AFT under extreme open and closed positions must be analysed. Closed variable geometry turbine In the case of a VGT, the hypothesis that R is 0.5 when the turbine is closed has been used. In this case, the thermodynamic evolution should be similar to the k-g evolution shown in the left part of Figure 5, due to the fact that the high velocity of the flow at the stator will increase considerably the increment of entropy. If the objective is to calculate the pressure between stator and rotor (P1k-g in Figure 5), it can be confirmed that a satisfactory hypothesis would be to assume the evolution in the rotor to be isentropic (y) and to calculate the polytropic exponent of the stator k' as a function of n and y.
----- y ------ k'-r - - n --k-Q
h
Po
h
~n
8h,
8h.
P.k-g
8h
&1
!=In
Qy !=Ik·.., 8h
I ----------------I. ------2s1
s
s
Figure 5: h-s diagram for a closed VGT (left) and an AFT (right) 106
Thus, p1k'-y should be taken as the sought pressure value, which is much more appropriate thanplyor pIn, obtained in the case where the entire evolution was assumed to be isentropic or with a constant polytropic exponent, respectively, In addition, it has been verified that the more irreversible the evolution in the stator, the more accurate this assumption will be. That is, when the VGT is closed (Figure 5), and the passage area of the stator is very small, the irreversibilities caused by friction are very significant and the entropy increment must be higher in the stator than in the rotor. In the case of a variable geometry turbine of the AFT type, when it is closed and the actual passage area of the stator is very small, R increases in accordance with what was experimentally obtained in Figure 4. If, in tum, the hypothesis of flow lamination is assumed, it should cause a very high increase of entropy across the stator due to the increased flow friction. Therefore, the evolution found will be similar to that labelled in Figure 5 as k-g evolution. Once again, when considering that the final objective is to calculate the intermediate pressure p1k-g, in the right side of Figure 5, it can be observed that the differences between p h-g and p lk'_y are the least significant ones. Open variable geometry turbine In the case of a VGT, when it is open, R is higher than 0.5 (Figure 3), which implies a low level of expansion at the stator. In this case that the turbine directs the flow towards
the rotor, the thermodynamic evolution should be as observed in the left part of Figure 6 with little increase of entropy at the stator and, therefore, the differences in PI calculated using any of the hypotheses are very small, and the real solution is always between Plk'-y andply. In turn, in the right part of Figure 6 a diagram is presented to illustrate what occurs under those hypotheses which are stated at the beginning of the section, for the case of a variable geometry turbine of the AFT type when open. Here, as previously seen, (Figure 4), R is low (~0.20). It can be observed in Figure 6 that once R is fixed and a small increase in entropy occurs, due to the low velocity of the flow at the stator, the differences between any of the proposed evolutions and p1k-g are small and, in any case, the required pressure lies between Ply and Plk'-y' ------ '"1
--n
h
------ k'-y
--k-g
Po
h
Figure 6: h-s diagram for an open VGT (left) and an AFT (right)
107
Sumarizing, in order to simplify the model and since the calculation ofPI is the main objective, it will be assumed for all the cases that the evolution in the rotor is isentropic, due to the fact that for the case of open turbine the differences are not significant for the calculation ofPI. For the evolution at the stator, a polytropic exponent k given by Eq. 21 where g=rand n is calculated using Eq. 18 will be assumed. Thus, the calculation of the intermediate pressure may be rexpressed as:
lX-I lR
n-I/n
12= [ (l-R)· P2
Po
(PJ
+R
=
_
0.5 if a>al;m;t 2·tanal 9={ rh*
R-l
2 7f
'-;'f(-,m
2
P2'*
N
DIDz
Poo
. ,J4,TJT,)lfa::;al;m;t
Eq.22
Obtaining the effective areas of the nozzles equivalent to the turbine stator and rotor Once R is established and considering how it is related to the pressure drops across the stator and the rotor, the calculation of the effective areas of the nozzles equivalent to the stator and the rotor can be completed by using the nozzle equation (Eq. 1). For the stator, one has:
A
eff _ st
=mr · ~ .~.(Poo) POO r PI
(r-rr
Yr
-1
r. ~[1-(J!L) r1 r 1 [
]
Eq.23
PO~
and for the rotor
Eq.24
The calculation of these effective areas has been carried out for three different turbines, a VGT, an AFT turbine and a fixed geometry turbine (FGT) whose characteristics are shown in Table 2. The FGT performance maps were supplied by the manufacturer. An obvious dependency on the effective area of the nozzle equivalent to the stator with the position of the turbine can be observed for the VGT and the AFT in Figure 7. AFT
VGT
'E600 5500 ~'" 400 Q)
~ 300 ~ Q)
~
00
~
o~~~~~~____~~~~~~~ 6 10 o AFT Displacement (mm)
~
200 100 0
""L,-.-~~~~~~~~~~~~ 40
50
60
70
60
90
Stator blades angle (")
Figure 7: Correlations obtained for the effective areas ofthe nozzles equivalent to the AFT and VGT stators
108
VGT
.-E
600
.s 600 m til 400
~
t&
"'0
400
<
300
200
~///
~
500
Q) (/)
.c
o
200
Q)
./
100
o //
o
I
1200
1000 m 750 til
1000 800 600
~
"n ~
!
5
o
..flO 60
12010U-
N
-JT01
0
Predicted 200
al ~
/'
o
300
400
~/
8
400 200 40 200
100
/
500
~
600
//
Predicted 200
400
600
800
1000 1200
(~) -JK
Figure 8: Correlations obtained for the effective areas of the nozzles equivalent to the AFT and VGT rotors It can also be noted that it is possible to correlate the effective area of the nozzle equivalent to the turbine rotor (Figure 8) as a function of the flow rate and the operation speed both corrected using the gas conditions at the rotor inlet. It can be shown that for a specific turbocharger speed, this effective area increases when the gas mass flow increases. Thereby, it seems obvious that the flow increase and the expansion ratio improve the use of the geometric section of flow passage through the rotor. In addition, for a specific flow rate, the effective area of the equivalent nozzle at the rotor decreases when the operation speed increases. This can be accounted for by considering that, as the operation speed increases, the centrifugal forces produced when the rotor turns also increase and the passage of the exhaust gas is obstructed. The model developed is also useful for FGT without guide stator blades. Assuming that R is 0.5 and using the data in the turbine maps provided by the the turbo manufacturer, the effective areas of the nozzles equivalent to the stator and the rotor of a fixed geometry turbine (Eq. 23 and Eq. 24) were calculated. The results obtained are shown in Figure 9, where it can be observed that the effective area of the nozzle equivalent to the stator increases, albeit slightly, when the flow rate is increased and the corrected operating speed is decreased. In Figure 7 it is also observed that, although the position of the variable geometry mechanism explains most of the variation seen in the effective section of the stator nozzle, it also depends on other turbine variables. Conversely, the effective area of the nozzle equivalent to the rotor continues to increase with the corrected flow rate and, in the case of a constant corrected flow rate (Table 2), to decrease with a corrected operating speed, as observed before for the variable geometry turbines.
109
~
E
.s til
~
:g
& LU 5
650 600 550 500 450 400 30 50
650
500
15
& LU
~ 0::
/'
/
./
././/
Predicted
400 400
1000
E -;; 1000 ~
/.r
450
/
~
.....,/
550
70 90 110 N* (rpsNK) 13015
!
//
//
600
900
900 800 700
450
500
550
600
650
~
1l" 0
800 700
21 22 23 m*(kg/s,lKlMPa)
27 3.1 23 . 1.5 1.9 r~p (TIs)
Predicted
600 600
700
800
900
1000
Figure 9: Effective areas of the nozzles equivalent to stator and rotor for an FGT The correlations obtained for the effective areas of the nozzles equivalent to the stator and the rotor in the three turbines studied are summarised in Table 2. Table 2: Summary of the correlations obtained for the effective Areas (mm2) of the equivalent nozzles that represent the stator and the rotor of the turbines A eff Stator VGT
k1 =0.82rO.09; k2 =1.44rO.11; k3=698r20
A eff Stator AFT
k1 =415r8; k2 = 0.257:r0.018; k3=22r2
r2=0.96
r=0.97
A eff Rotor VGT
A eff Rotor AFT
k1 =-1.04rO.03 k2 = 410r20 k3 = 0.92rO.05 k4 = O. 19r 0.02
r=0.97
k1 =-6.8r0.3 k2 = 1934r20 k3 = 0.043rO.006 k4 = 1.047rO.005
r=0.98
A eff Stator FGT
k1 =-4.95rO.19; k2 = -1.5rO.1; k3 = 0.194rO.009; k4 = 65H 13
A eff Rotor FGT
k1 =-4.52r O. 17; k2 =7.6rO.5 k3 =0. 154r 0.009; k4 =163.8r 1.5 k5 =390r 10
r=0.99
110
COMPARISON BETWEEN MEASUREMENT AND MODELLING
Once the turbines were characterised, the correlations obtained for the effective areas equivalent to the nozzles representing the stator and the rotor were implemented in a global gas dynamic model of an engine that supplies the inputs required by the turbine sub-model in order to calculate the effective areas. This is a one-dimensional, nonhomoentropic and unsteady wave action model whose details can be found from (12) to (20). The model has been validated using tests conducted in engine test bed, so that the turbines were coupled to the engines. The AFT is part of the turbo charging group of a 2.2 litre displacement compression ignition engine, the VGT is part of a 1.9 litre displacement compression ignition engine and the FGT is coupled to an 10.83 litre displacement engine. The features of these engines are shown in Table 3. The engines were installed in the test benches with all the equipment and instrumentation necessary to control their performance and measure their operational variables. In each test the most significant parameters related to the performance conditions of the turbine, such as the pressure and temperature at the inlet and outlet of the turbine, the flow rate passing through the engine, the operating speed of the turbo and the positions of the VGT and AFT stators, were measured. The results obtained from the global engine model were compared to the measurements carried out, so that it could be seen that the model was capable of reproducing the fluid-dynamic behaviour of different radial turbines conceived for the turbo charging of different engines. For the case of the AFT, in Figure 10 comparison is shown between the modelled and measured instantaneous pressures at the turbine inlet, at the turbine outlet and also at the exit of the intake manifold. The displacement of the stator rod, the engine speed and load and the rotating speed of the turbo for each point tested are also given. From the results obtained, the flow lamination at the turbine exit when it is closed (small rod movements) should be highlighted in comparison to the case when it is open at 60% (rod movements of approximately 6 mm). Indeed, this is consistent with the hypotheses previously discussed. Table 3: Characteristics of the engines
Type of injection Number of cylinders Displacement Compression ratio Rated power Rated torque Rated speed Bore Stroke Connecting rod length
Engine with VGT Direct
Egine with AFT Direct 4 4 2.21 1.91 18:1 18.3:1 88 kW / 4000rpm 98 kW / 4000rpm 300 Nm / 2000rpm 314Nm / 2000rpm 4500 rpm 4500 rpm 85mm 80mm 93mm 96mm 152mm 193mm
III
Engine with FGT Direct 6 10.83 I 16:1 340kW / 1800rpm 2200Nm / 1200rpm 1800 rpm 123mm 152mm 225mm
_.- Experimenlal
+
Turbine Inlel Pressure
+
Compressor Outlel Pressure
-41-
Turbine Outlel Pressure
3.5
3.5l 1.19rrm,1500rpm, fuilload,14OOOOrpm
2 mm, 1750 rpm, full load, 150000 rpm
~ 3.0
3.01
E
2.5~
2.5
~ c.. 2.0
c.. 2.0-1
1.51i
1.5
i
1.0·1~4
o
180
360 Crank angle
540
1.01_ o
720
2.0 ...~ 6 rrm. 2540 rpm, 51 % load, 94000 rpm
4.0
180
360
540
720
Crank angle
7.3 mm, 3500rpm, full load, 175000 rpm
1.8
E
3.0 1.6
c.. 1.4·
i
2.01 1.2
1.0·· -·--y--··-r---r--·-'.....,·--.-----1--'---1 o
180
360
540
1.0
I--.~~~~~~~~=,~~
+I-'--~-'-~-r---~-'-~__
o
720
Crank angle
180
360
540
720
Crank angle
Figure 10: Comparison of an AFT, for different measured and modelled points In the case of the AFT, the angle a] is kept constant. However, in the case ofa VGT, the variation of this angle causes an alteration in the passage section and it is precisely this angle which has to be known for each performance point. The measured-modelled comparison was carried out using steady tests of the 1.9 litre engine described above. The results obtained for different points are given as an example in Figure 11 where good agreement between the experimental values and those obtained from the modelling can be observed in all the cases. Additionally, in the case of the VGT, the average variables modelled and measured for the engine running at 2000 rpm and full load can be seen in Table 4 and it is worth noting that the relative errors in all variables are below 2%. Finally, in the case of the FGT, the comparison between measured and modelled instantaneous pressure for four points measured at full load conditions for different engine speeds (800, 1200, 1500 and 1800 rpm) is shown in Figure 12. These results cover a wide operating range for this engine size, and thus allow observing the good agreement achieved between the measured and modelled values.
112
Table 4: Comparison between the measured and modelled average values for an engine speed of 2000 rpm Engine with VGT - 2000 rpm Modelled 2.34 1.16 1056 900 0.99 2.22 302 421 0.065 2.21 324 165502
Turbine inlet pressure (bar) Turbine outlet pressure (bar) Turbine inlet temperature (K) Turbine outlet temperature (K) Compressor inlet pressure (bar) Compressor outlet pressure (bar) Compressor inlet temperature (K) Compressor outlet temperature (K) Air mass flow (kg/s) Intercooler outlet pressure (bar) Intercooler outlet temperature (K) Turbocharger speed (rpm)
Experimental 2.30 1.15 1033 898 0.98 2.23 307 422 0.066 2.22 322 164200
59",2000 rpm, full load, 166000rpm
51',1500 rpm, full load, 130000 rpm
3
Error (%) -1.7 -0.9 -2.2 -0.2 -1.0 0.4 1.3 0.2 1.5 0.5 -0.6 -0.8
2.5 ~
2
:J
III
~
a.
1.5
o
180
360
540
720
Crank Angle [0]
0
180
+--+--+ G----9---€)
360 Crank Angle
Experimental
["1
540
720
Turbine Outlet Compressor Outlet
63',2500 rpm, full load, 169000rpm
70',3000 rpm, full load, 173000 rpm
3 'i::'
'i::'
B
B
'"
'"
~
~
::J
"'"'~ a.
::J
"'"'~
2
2
a.
o
180
360
540
720
Crank Angle (0]
o
180
360
540
Crank Angle [0]
Figure 11: Comparison of measured and modelled pressures in a VGT for different blade angles and operating points
113
720
800 rpm
1200 rpm 5t-~~~~"~~~-A~~~
o 180
360
540
Crank Angle (OJ
-"'
- - - -
360
540
720
Crank Angle (OJ __ Turbine Inlo1
-Experimertal _ Turbine OuUel
1500 rpm
180
720
~ Compressor
Outlet
1800 rpm
-----: 6
e~
5
H~""'+---\---J--¥I,---,I---'~f--~--+--*I
~5~~~7-~*-~~~~~~#-~
I!!
~ 4f!~~~~~~~~~~~~~
:::I
'"'" 3 t-~~~~--~~~~~--~___1 I!!
n.
'"'" I!!
n.
3
+-~~~~---
----
+-~~~~~-,~~~~~
o
180
360
Crank angle (OJ
540
720
o
180
360
540
720
Crank Angle (OJ
Figure 12: Comparison of measured and modelled pressures in an FGT for different operating points CONCLUSIONS A new one-dimensional model for variable geometry turbines has been developed. This model is a natural evolution of a previous model for fixed geometry turbines in which the turbine is represented by two ideal nozzles, which reproduce the pressure drops across the stator and the rotor, and at an intennediate cavity, which reproduces mass accumulation in the system. The structure of the model presented is the following. Firstly, the reaction degree R of the variable geometry turbine under the desired perfonnance conditions is calculated. For this, a methodology has been proposed in which geometrical parameters and the corrected variables obtained using the maps supplied by the manufacturer or from the conditions of the gases at the turbine exit are used. Once R is established, and taking into account the type and the position of the stator blades of the variable geometry turbine together with the working conditions, it is possible to calculate the pressure drops produced at the stator and the rotor. Once these pressure drops are known, it is possible to calculate the effective areas of the equivalent nozzles representing the turbine stator and rotor. With regards to the VGT and the AFT, it was concluded that R has a high level of correlation with its opening and with the flow rate passing through the turbine. In addition, the effective area of the stator correlates with the position of the turbine, and that of the rotor with the corrected mass flow rate and the corrected turbine rotating speed. Thus, it was shown that the effective area of the nozzle equivalent to the rotor, for a specific rotating speed, increases as the gas flow increases. However, for a given flow rate, the effective area of the nozzle equivalent to the rotor diminishes when
114
rotating speed increases. This is caused by the centrifugal forces produced when the turbine rotor turns and by the obstruction imposed to the passage of exhaust gases. To validate the model, it has been introduced into a global engine model based on a wave action gas dynamic code. Results were compared to tests performed on an engine bench, and it was concluded that for both VGT and AFT, this model is able to reproduce the fluid dynamic and thermodynamic behaviour of the turbine with good accuracy. Finally, the model was applied to an FGT, using the measurements on the turbine map supplied by the manufacturer as input data. By calculating the effective areas of the nozzles equivalent to the stator and the rotor as a function of corrected variables supplied by the manufacturer, it was possible to achieve a highly precise reproduction of the fluid dynamic behaviour of this turbine when operating coupled to an internal combustion engine over a wide operating range.
ACKNOWLEDGEMENTS This work was sponsored by the Powertrain Engineering Division of Renault SAS. The authors would like to thank Dr. Antonio Torregrosa for his helpful suggestions.
BIBLIOGRAPHY (I) Benson, R.S. (1982) The Thermodynamics and Gas Dynamics of InternalCombustion Engines. Vol I Oxford University Press, Oxford. (2) Moraal, P. and Kolmanovsky, I. (1999) Turbocharger Modeling for Automotive Control Applications. SAE Technical Paper 1999-01-0908. (3) Chen, H., Hakeem, I. and Martinez-Botas, R. (1996). Modelling of a turbocharger turbine under pulsating inlet conditions. Proceedings of the Institution of Mechanical Engineers, 1996, VoI2JO, pp 397-407. (4) Iwasaki, M., Nobuyuki, I., Marutani, Y. and Kitazawa, T. (1994) Comparison of Turbocharger Performance Between Steady Flow and Pulsating Flow on Engines. SAE Technical Paper 940839. (5) Winterbone, D.E., Nikpour, B. and Frost, H. (1991) A contribution to understanding turbine performance in pulsating flow. Proceedings of the Institution of Mechanical Engineers, C4331011, pp 19-30. (6) Watson, N. and Janota, S. (1982) Turbocharging the internal combustion engine. McMillan Publishers Ltd. ISBN 0333242904. London. (7) Payri, F., Benajes, J., Jullien, J. and Duan, Q. (1991) Non-steady flow behaviour of supercharger turbine. Proceedings of the Third EAEC International Conference, pp. 347-351, Strasbourg. (8) Payri, F., Benajes, J. and Reyes, M. (1996) Modelling of Supercharger Turbines In Internal-Combustion Engines. International Journal of Mechanical Science, Numbers 89, pp. 853-869.
115
(9) Kessel, J.A., Schaffnit, J. and Schmidt, M. (1998) Modelling and real-time simulation of a turbocharger with variable turbine geometry (VTG). SAE Technical Paper 980770. (10) Nasser, S.H. and Playfoot, B.B. (1999) A turbocharger selection computer model. SAE Technical Paper 1999-01-0559. (11) Lujan J.M., Serrano, 1.R, BermUdez, V. and Cervello, C. Test bench for turbocharger groups characterization. SAE Technical Paper, 2002-01-0163. (12) Payri, F., Benajes, J. and Galindo, J. 1995. One-dimensional fluid-dynamic model for catalytic converters in automotive engines. SAE Technical Paper 950785. (13) Benajes, J., Reyes, E. and Lujan, J.M. 1996. Modelling study of the scavenging process in a turbocharged diesel engine with modified valve operation. Proc. IMechE, Part C: Journal ofMechanical Engineering Science, vol. 210. (14) Benajes, 1., Reyes, E., Galindo, 1. and Peidr6, J. 1997. Predesign model for intake manifolds in internal combustion engines. SAE Technical Paper 970055. (15) Benajes, 1., Reyes, E., Bermudez, V. and Serrano, 1.R. 1998. Predesign criteria for exhaust manifolds in IC automotive engines. SAE Technical Paper 980783. (16) Payri, F., Reyes, E. and Serrano, J.R., 2000. A model for load transients of turbocharged diesel engines. 1999 SAE Transactions-Journal of Engines, vol. 108-3, pp. 363-375. (17) Benajes, J., Lujan, 1.M. and Serrano, J.R. 2000. Predictive modelling study of the transient load response in a heavy-duty turbocharged diesel engine. SAE Technical Paper 2000-01-0583. (18) Payri, F., Reyes E. and Galindo J. 2001. Analysis and modelling of the fluiddynamic effects in branched exhaust junctions of ICE. International Journal for Gas Turbine and Power: Transactions of the ASME, vo!' 123, no. 1, pp. 197-203. (19) Payri, F., Benajes, J., Galindo, J. and Serrano, J.R., 2002. Modelling of turbocharged diesel engines in transient operation. Part 2: Wave action models for calculating the transient operation in a high speed direct injection engine. Proc. IMechE, Part D, D06501, vol. 216, pp. 479-493. (20) Galindo, 1., Lujan, J.M., Serrano, J.R., Dolz, V., Guilain, S. 2004. Design of an exhaust manifold to improve transient performance of a high-speed turbocharged diesel engine. Experimental Thermal and Fluid Science, vol. 28, pp. 863-875.
116
APPENDIX: CALCULATION OF VARIABLE GEOMETRY TURBINE
THE
REACTION
DEGREE
OF
A
The definition of R is usually based on the energy transferred, that is, the ratio between the energy transferred due to the pressure change in the rotor and the total variation of energy: R
= h. -~
A.I
hoo - h02
hoo - h02
Since no work is developed at the stator 1100=1101' If the fluid can be regarded as an ideal gas, it can be assumed that:
hoI -ho2 =c p • (Tol - 'fo2)
A.2
In addition, the energy transfer in the rotor can be represented as the product of the torque by the angular velocity
TV = llJ·TQ= m,(ulcol -U2C02 ) =m·c p • (Tol -To2 )
A. 3
From the previous equation, it is readily obtained that: -U C )
(UIC
OI 2 02 'fo1-'fo2 =-'---'---"-"--"--'~
cp
A.4
If, as a design hypothesis, swirl at the exit is neglected (C82=O ~ cl= cal) and it is assumed, also as a design hypthesis, that the radial velocity at the stator outlet (i.e., at the rotor inlet) it is equal to the axial velocity at the rotor exit (Crl=CaZ=CI'COS UI), from equations A.2 to AA one has:
2 2
2
2 2
R=I_CI-C2 =1_COl+Crl-Ca2 2ul cOl 2ul cOl
A.5
Taking into account the aforementioned conditions, and the velocity triangles, (Figure 2), this can be rearranged to give: sinal
ca2 ' - R=I-.5!L=I 2uI
cI'sinal
cosal =1_tanal .ca2 2 ul
2uI
A.6
2uI
Here, al is the gas entry angle to the rotor, which will be determined by the stator guide blades. In the case ofa rotor with radial blades, PI= 0, and without guide blades to direct the flow, COl=u], it is observed from equation A.6 that R is 0.5, equal to the case of the fixed geometry turbine without guide blades.
117
Analysis of Turbocharger Non-Adiabatic Performance S. Shaaban and J. R. Seume University of Hannover, Turbomachinery Laboratory ABSTRACT
Turbocharging is nowadays the most commonly used method of engine supercharging. One of the important factors affecting the turbocharger performance is the heat transfer inside the turbocharger and from the turbocharger to the ambient. Heat transfer takes place due to the high temperature gradient between the turbine and the other components of the turbocharger as well as between the turbine and the ambient. This heat transfer causes underestimation of the measured compressor efficiency and overestimation of the turbine efficiency. This results in an inaccurate estimation of the compressor power and the turbine power. The complex turbocharger geometry introduccs many geometrical and operating parameters that affect the turbocharger non-adiabatic performance. The purpose of the present paper is to analyse the turbocharger non-adiabatic performance and define the parameters affecting this performance. It also helps to understand and estimate the effect of each parameter on the turbocharger non-adiabatic performance. NOMENCLATURE
meridional velocity at the compressor p pressure [Pa] impeller outlet [m/s] specific heat [J/(kg.K)] q specific amount of heat transfer [J/kg] cp gas constant [J/(kg.K)] total enthalpy [J/(kg.K)] R h temperature [K] fraction of heat transfer T K mass flow rate [kg/s] U2 compressor peripheral speed [mls] compressor peripheral Mach number u 2 ~Kai,Rai,Tlt Greek Letters specific heat ratio !32.b blade angle at the compressor impeller K outlet [0] slip factor LI difference f..l ¢2 impeller flow coefficient C2n/U2 pressure ratio n: heat number 17 efficiency (" Subscripts non-adiabatic 1 compressor inlet dia exhaust gas 2 compressor impeller outlet exh heat for estimation of power 5 compressor outlet isentropic 6 turbine inlet is meas measured 8 turbine outlet static adi adiabatic s turbine after after the impeller T total air aIr turbo turbocharger ac actual heat transfer is tacked into consider before before the impeller * C compressor
1
119
1. INTRODUCTION
Heat transfer takes place inside the turbocharger from the hot turbine to the lubrication oil as well as from the hot turbine to the compressor through the bearing housing. It also takes place between the compressor and the lubrication oil. Heat transfer to the compressor causes deterioration of the measured isentropic compressor efficiency. This is because the total temperature difference across the compressor in this case results from the compressor aerodynamic work and the amount of heat transfer to the compressor. Adiabatic turbocharger performance would result in the absence of heat transfer between the components of the turbocharger and between the turbocharger and its surroundings. It is evident that adiabatic turbine performance is not likely to occur during normal operation of the turbine due to the high temperature of the exhaust gas. Even if the surface area of the turbine was insulated, heat transfer would take place from the turbine to the lubrication oil as well as to thc compressor. However, adiabatic turbine performance can be approximated in the laboratory by supplying the turbine with compressed air instead of hot exhaust gases. This type of experiments is designated as "cold measurements" below. Compressor performance likewise is actually non-adiabatic due to heat transfer from the turbine to the compressor. Adiabatic compressor performance can also be achieved by means of cold measurements. Chapman et al. [1], Malobabic and Rautenberg [4], Pucher et al. [5], and Pucher and Nickel [6] experimentally investigated the effect of heat transfer on the measured compressor and turbine efficiencies. It was clear in the experimental investigation of Pucher et al. [5] that small size turbochargers arc more affected by the amount of heat transfer to the compressor at low rotational speeds than larger ones. Malobabic et al. [3] reported that the heat transfer inside and from the turbocharger has a negative influence on the required compression power and the delivered turbine power.
2. TURBOCHARGER ISENTROPIC EFFICIENCIES The quality of the compression process through the compressor is normally expressed in terms of the isentropic compressor efficiency
.T(n(Kair -l)/Kair -1)
~h.
11
IS,e
c p•mr
=~=
'li"C
ac,C
It
t,C
cp,air
(1)
(T -T ) 5t
It
Direct estimation of the turbine efficiency using the measured temperature at the turbine outlet yields a highly overestimated turbine efficiency due to the high heat loss of the turbine. Therefore, turbine performance maps include the product of the turbine isentropic total-to-static efficiency and the turbocharger mechanical efficiency. This product can be estimated from measurements by means of defining a turbocharger overall efficiency in the form ill C c p,aJf. llturbo
=
T (n(Kair - 1)/Kair - 1) It
(( - Kexh-l
.
mT
t,C
Cp,cxh T6t
l-1I 618s ,T
J/)
(2)
Kexh
The product of the turbine efficiency and the mechanical efficiency is therefore given by llis,T
11m = lltufbo/llis,c
(3)
3. TURBOCHARGER NON-ADIABATIC PERFORMANCE Heat transfer to the compressor results in an increase in the total temperature at the compressor outlet. Therefore, the measured compressor efficiency under non-adiabatic operating conditions is lower than the actual compressor efficiency. This underestimated compressor efficiency causes an overestimation of the product of the turbine total-to-static efficiency and the mechanical efficiency according to
120
equation (3). Thc deviation of the measured compressor efficiency from the actual one can be analysed with the help of Figure I. This figure shows a schematic h-s diagram for the compression process in an adiabatic and a non-adiabatic compressor. In the absence of heat transfer to the compressor, the compression process would follow the path 1-5 adi . The true total temperature at the compressor outlet is therefore higher than the adiabatic total temperature due to the amount of heat transfer to the compressor. The compression process through the Tot PSt impeller is assumed to be adiabatic [2]. The amount of heat transfer to the compressor is divided into two fractions. The first fraction of heat transfer takes place at constant prcssurc before the impeller inlet. This amount of heat transfer causcs the total temperature to increase from T It to T ll*. The air is then compressed adiabatically in the impeller and this causes the total q C,velo", temperature at the impeller outlet to increase to T5t.adi*. The second fraction of heat transfer takes place at constant pressure after the impcllcr. This fraction Entropy s causes the total temperature to increase from T5t,adi* to Tst . Figure 1 Schematic h-s diagram of the compression The application of this model introduces process in an adiabatic and a non-adiabatic three different expressions for the compressor compressor efficiency. The first expression for the compressor efficiency is called the adiabatic compressor efficiency. This is the compressor efficiency in the absence of heat transfer to the compressor (4)
The second expression for the compressor efficiency is called the compressor non-adiabatic or diabatic efficiency lldia,C. This is the measured compressor cfficiency under non-adiabatic operating conditions lldia,C
=
~his,c / ~hdia,C
(5)
The third expression is called the compressor heat efficiency. This expression is used to define the actual comprcssor work under non-adiabatic operating conditions llheat,C
= ~his,c / ~hadi,C*
(6)
The denominator of equation (6) represents the actual compressor aerodynamic work under non-adiabatic operating conditions. While equation (5) represents the apparent compressor efficiency measured under non-adiabatic operating conditions, equation (6) must be used to estimate the actual compressor work whenever heat transfer takes place to/from the compressor. The three different expressions for the compressor efficiency coincide for an adiabatic compressor. For an adiabatic comprcssor, they also coincide with the isentropic compressor efficiency, equation (I). A presented compressor performance map can therefore be only used to estimate the actual compressor work if the compressor was adiabatic during the performance test. The same approach can be applied to the expansion process through the turbine as shown in Figure 2.
4. ACTUAL COMPRESSOR AND TURBINE WORK The ratio of thc compressor heat efficiency to the compressor adiabatic efficiency can be represented with the help of equations (4) and (6) in the form
121
Entropy s Figure 2 Schematic h-s diagram ofthe expansion proccss in a non-adiabatic turbine TJhcat,C
ll.his,c
Tlt
TJadi,C
ll.his,c'
Tit.
(7)
The total temperature at the impeller inlet under non-adiabatic opcrating conditions is givcn by (8)
Tit' = Tit + qc,bcforc/Cp,air
The specific heat at constant pressure is assumed to be constant throughout the present work The specific amount of heat transfer to the compressor before the impeller represents a fraction KC,before of the total specific heat transfer to the comprcssor. Equation (8) can therefore be rewritten as (9) Substituting equation (9) in equation (7) and rearranging TJhcat,c/TJadi,c =
(1 + KC,before Sh,e )-1
(10)
The compressor heat numbe~ Sh,e is defined as
Sh,c=~T cp,air
(11)
It
Equation (10) shows that the amount of heat transfer to the compressor causes an increase in the compressor aerodynamic work and hence in the compressor power demand, This increase in the compressor power demand under non-adiabatic conditions also depends on the distribution of the heat transfer. Figure 3 shows the ratio of the compressor heat efficiency to the compressor adiabatic efficiency as a function of the compressor heat number and the fraction of heat transfer before the impeller. The compressor heat efficiency equals the compressor adiabatic efficiency at zero heat number and/or heat transfer to the compressor only after the impeller (Ke,before = 0), Zero compressor heat number means adiabatic compression in the compressor and zero heat transfer before the impeller mcans adiabatic compression in the impeller. Increasing the fraction of heat transfer before the impeller results in an increase in the power demand of the compressor and hence a decrease in the compressor heat efficiency under non-adiabatic operating conditions. Thc decrease in the compressor heat efficiency becomes more significant with increasing compressor heat number. The ratio of turbine heat efficiency to turbine adiabatic efficiency can be likewise presented as TJheat,T /TJadi,T =
(12)
1- KT,before Sh,T
The turbine heat number presents the total amount of heat transfer from the turbine in non dimensional form
122
Sh T = qT , cp,exh T6t
(13)
Figure 4 shows the effects of the turbine heat number and thc fraction of heat transfer before the turbine rotor on the ratio of the turbine heat efficiency to the turbine adiabatic efficiency, The turbine heat efficiency coincides with the turbine adiabatic efficiency at zero turbine heat number and/or zero heat transfer before the turbine rotOL However, these conditions cannot be achieved for the turbine, The amount of heat transfer from the turbine cannot be zero during normal operation of the turbochargcr cven with insulating the turbine surface area because of the amount of heat transfer from the turbine to the other turbocharger components, The fraction of heat transfer before the turbine also cannot be zero because of the very large surface area of the turbine volute in comparison with the surface area of the rotor shroud, Therefore, the turbine heat efficiency is less than the adiabatic turbine efficiency, Increasing the amount of heat transfer from the turbinc may cause significant decrease in the turbine heat efficiency and hence in the power delivered by the turbine, 1.01-,--------------, 1.00 -k""""',----------------t
~ m"';~~:::~~:;_~:~-~~~~--- __ ] 0.95 ~
Compressor heat number ~hJ': ' - - 0 (
0.94 0.93 0.92
"
~}
'1
............
~
- - -0.025 ------ 0.050 ----0.075 -----0.100
0.91
0.1
0.2
0.3
0.4
0.5
0.6
0.7
O.X
0.9
1.0
0.0
KC.b,,,,,,[-J
Figure 3
Turbine heat number ~1~T:
- - 0 (adiabatic turbine) - - -0.025 ------ 0.050 ---- 0.075 -----0.100
0.88
0.90 +'-'-'-"""''''''''-'-r""T"T'"'-'-'-''''''''''''''-'-r""T-1
0.0
1.01 1.00 0.99 0.98 0.97 0.96 0.95 0.94 0.93 0.92 0.91 0.90 0.89
0.1
0.2
0.3
0.4
0.5
Kr.b,fore
Figure 4
Ratio of compressor heat efficiency to compressor adiabatic efficiency as a function of compressor heat number and fraction of heat transfer before the impeller
0.6
0.7
0.8
0.9
1.0
[-J
Ratio of turbine heat efficiency to turbine adiabatic efficiency as a function of turbine heat number and fraction of heat transfer before the turbine rotor
5. EFFECT OF HEAT TRANSFER ON THE MEASURED COMPRESSOR AND TURBINE EFFICIENCIES The compressor non-adiabatic efficiency can be expressed as Tldia,C =
h
L'1h isC
(14)
' t. adi,C* + qc
Rearranging equation (14) and substituting from equation (6) 1
1
qc
Tldia,C
Tlheat,C
L'1h is,c
--- = --- +---
(15)
Multiplying equation (15) by the adiabatic compressor efficiency and rearranging Tladi,C
lladi,c
lldia,c
llheat,C
qc
--=---+-...!.-'='-
TIt* cp,air Tit
(16)
t.hadi,C* Tit cp_air Tit
Substituting from equations (9), (11), and (12) in equation (16) and rearranging lladi,c lldia,c
=
(1 + K C,before . Sh,C )(1 + Sh,C cp,air Tit J Ah L:1
(17)
adi,C*
123
Equation (17) can be further developed by applying Euler's equation for turbomachinery to the compressor in the absence of compressor inlet guide vanes &adi,C* = u2 C2u
(18)
Thc swirl vclocity at the impeller outlet is given by
(19)
c2u =J.l(U2 -C2m/ tan 132,b) Substituting for C2u in equation (18) and rearranging L'lhadi,C* = J.l uHI- ~z/tan 132,b)
(20)
Since Cp,air =
Kair Rair Kair -I
(21)
Substituting from equations (20) and (21) in equation (17) and rearranging 11dia,C [ + -Sh,C 1 ' 1 - = (1 + KC,before Sh,C )-1 ,1 - - , --2 11adi,C Kair- 1 MU2 J.l(I-~z/tan132,b)
J-1
(22)
Equation (22) shows that the deviation of the measured compressor non-adiabatic efficiency from the compressor adiabatic efficiency depends on a, Compressor heat number Sh, C b, Impeller peripheral Mach number MU2 c, Impeller flow coefficient ~2 d, Fraction of heat transfer before the impeller KC,before e, Slip factor J.l f Blade angle at the impeller outlet 132,b These parameters are also affecting the measured product of the turbine total-to-static efficiency and the mechanical efficiency, This is because the product of the turbine efficiency and the mechanical efficiency is estimated according to equation (3) from the turbocharger overall efficiency and the measured compressor efficiency, The measured compressor efficiency according to the new definitions is the non-adiabatic compressor efficiency, Therefore
(23)
(l1hcat,T 11m )meas = 11 Turbo /l1dia,C
However, the actual product of the turbine efficiency and the mechanical efficiency should be estimated from the equation (l1heat,T 11m
t
= 11Turbo/11heat,C
(24)
Therefore 11dia,C: ( ) ( ) (25) 11heat,T 11m ac = 11heat,T 11m meas , - - 11heat,C This means that the relative systematic error (overestimation) in the measured product of the turbine efficiency and the mechanical efficiency due to the amount of heat transfer to the compressor is L'll1heat,T (l1hcat,T )ac
= 11heat,C: -1 11dia,C:
(26)
The ratio of the compressor heat efficiency to the compressor non-adiabatic efficiency can be obtained by dividing equation (10) by equation (22), The relative systematic error (overestimation) can therefore be represented as L'll1heat,T (27) (l1heat,T )ac 124
The above analysis defines the main parameters affecting the measured compressor and turbine efficiencies. The effect of each parameter on the measured compressor non-adiabatic efficiency and the systematic error (overestimate) in the estimated product of the turbine efficiency and the mechanical efficicncy is discussed in the flowing section with the help of Figure 5 and Figure 6. These figures are obtained by setting the geometrical and operating parameters in equation (22) and equation (27) at typical values ofthe GTl749V 55 Trim turbocharger of the company Honeywell Garrett. 6. FACTORS AFFECTING EFFICIENCIES
TURBINE
AND
COMPRESSOR
NON-ADIABATIC
6.1 Compressor Heat Number
The compressor heat number [equation (11) 1is a number that presents the amount of heat transfer to the compressor in non-dimensional form. Figure 5 shows that the compressor non-adiabatic efficiency coincides with the compressor adiabatic efficiency at zero compressor heat number. Increasing the amount of heat transfer to the compressor increascs the compressor heat numbcr and hencc decrcascs the compressor non-adiabatic efficiency. This effect is very significant at low compressor peripheral Mach number MU2 (low rotational speeds) because the compressor aerodynamic work decreases with decreasing thc compressor peripheral Mach number. Therefore, increasing the compressor heat number at low peripheral Mach number causes a significant increase of the ratio between the specific amount of heat transfer to the compressor and the compressor aerodynamic work. This makes the effect of heat transfer to the compressor very significant at low rotational speeds. The measured product of the turbinc efficiency and the mechanical efficiency is therefore significantly affected by the amount of heat transfer to the compressor at low rotational speeds. 6.2
Compressor Peripheral Mach Number
Compressor peripheral Mach number is one of the most important parameters affecting the non-adiabatic turbochargcr performance. This can be observed in equation (22) where the ratio of the compressor non-adiabatic efficiency to the compressor adiabatic efficiency depends on the square of thc compressor peripheral Mach number. This also applies to the systematic error (overestimation) in the product of the turbine efficiency and the mechanical efficiency as shown in equation (27). The ratio of the compressor non-adiabatic efficiency to the compressor adiabatic efficiency tends towards unity with increasing the compressor peripheral Mach number, Figure 5, because the compressor aerodynamic work increases with increasing the rotational speed. Therefore, the ratio between the specific amount of heat transfer to the compressor and the compressor aerodynamic work decreases with increasing the rotational speed. This causes the deviation of the compressor non-adiabatic efficiency from the compressor adiabatic efficiency to decrease with increasing the rotational speed. Figure 5 also shows that the compressor can be assumed to work adiabatically at high compressor peripheral Mach number and low compressor heat number. The deviation of the compressor non-adiabatic efficiency from the compressor adiabatic efficiency increases significantly with decreasing the rotational speed even at small heat numbers. This causes a high systematic error (overestimate) in the estimated product of the turbine efficiency and the mechanical efficiency at low rotational spceds, Figure 5. 6.3
Impeller Flow Coefficient
Increasing the impeller flow coefficient is always associatcd with decreasing the compressor aerodynamic work, equation (20). This causes an increase in the specific amount of heat transfer to the compressor in comparison with the compressor aerodynamic work. Therefore, the effect of heat transfer on the compressor non-adiabatic efficiency increases with increasing the impeller flow coefficient as shown in Figure 5. The systcmatic error (overestimation) in the product of the turbine efficiency and the mechanical efficiency likewise increases with increasing the impeller flow coefficient, Figure 5. 125
6.4
Fraction of Heat Transfer before the Compressor Impeller
The fraction of heat transfer before the impeller has a vcry small effect on the compressor non-adiabatic efficiency. This small effect is mainly due to the increase in the compressor aerodynamic work with increasing the fraction of heat transfer to the compressor before the impeller. The fraction of heat transfer before the impeller has no effect on the systematic error (overestimation) in the measured product of the turbine efficiency and the mechanical efficiency as shown III equation (27). q" :: 0.3, Mu, :: 0.8, K"..",,, :: 0.5 P,.• = 42', ;t =: 0.8, K"b,"" '" 0.5
D
D
q,,=0.1
,,,=0,7 1.0
1.0
0.8
J 0.6
....... ,
100
l:
'7f "-- , RO i :::- 60
SO
~ 60
"" ]
40
'\
r!
40
20 0
Figure 5 Effect of compressor heat number, compressor peripheral Mach number, and compressor impeller flow coefficicnt on the measured non-adiabatic turbocharger performance 6.5
Figure 6 Effect of blade angle at the compressor impeller outlet, and compressor slip factor on the measured turbocharger non-adiabatic pcrformancc
Compressor Slip Factor
The compressor slip factor directly affects the compressor aerodynamic work. High values of the slip factor at the same peripheral Mach number and impeller flow coefficient mean higher compressor aerodynamic work. Therefore, the ratio between the specific amount of heat transfer to the compressor and the compressor aerodynamic work decreases with increasing the compressor slip factor. This causes less deviation of the compressor non-adiabatic efficiency from the compressor adiabatic efficiency with increasing the value of the slip factor, Figure 6. The increase in the slip factor results in a decrease of the systematic error (overestimation) in the product of the turbine efficiency and the mechanical efficiency. This decrease is due to the decreasing effect of the amount of heat transfer to the compressor on the compressor non-adiabatic efficiency with increasing the value of the slip factor. 6.6
Blade Angle at the Impeller Outlet
Impellers with large blade angle at outlet deliver higher aerodynamic work to the fluid. This can also be seen in equation (20). This makes compressors with large blade angle at impeller outlet less 126
affected by the amount of heat transfer to the compressor, Figure 6. The effect of blade angle at the impeller outlet is very small at low heat numbers and increases with increasing the compressor heat number. This effect is generally less significant than the effect of compressor heat numbcr, comprcssor pcripheral Mach number, impeller flow coefficient, and slip factor. 7. EXPERIMENTAL INVESTIGATION
7.1
Procedure
Compressor non-adiabatic performance is investigated by comparing thc results of thc so called "cold measurements" with that of the "hot measurements". Cold measurements are conducted by supplying the turbinc with compressed air instead of the hot exhaust gases. The oil inlet temperaturc is kcpt lower than the total temperature at the compressor outlet to avoid any possible heat transfer from the oil to the compressor at the low rotational speeds encountered during cold measurements. Furthermore, previous measurements with and without isolating the compressor casing as well as simple lD estimation of convection and radiation heat transfer from the compressor casing show that the amount of heat transfer by convection and radiation from the compressor is very small and can be neglected in comparison with both, the conduction heat transfer from the turbine to the compressor through the bcaring housing and the compressor aerodynamic work. Therefore, the comprcssor can bc assumed to work adiabatically during the cold measurements and the measured compressor performance can be considered to be the adiabatic compressor performance. The turbine is supplied with hot exhaust gases in the hot measurements and the turbocharger performance maps are measured at constant total temperature of the cxhaust gas supplied to the turbine. The measurements of the turbocharger performance maps are repeated at different exhaust gas total temperatures at the turbine inlet. The hot measurements are then compared with the cold measurements. Any deviation of the measured compressor performance in the hot measurements from the adiabatic compressor performance is considered to be duc to thc amount ofhcat transfcr to the comprcssor.
7.2
Error analysis
An error analysis is performed to check the uncertainty of the measurements, especially at low rotational speeds. This error analysis is bascd on Taylor's thcorem, retaining thc linear tcrms and using the uncertainties of sensors and AID-converters as in their specifications. Figure 7 shows the estimated error of the measured compressor efficiency of the GT1749V 55 Trim turbocharger at an exhaust gas total temperature of 873K at the turbine entry. The maximum probable errors of the measured compressor efficiency are satisfactory. However, the maximum probable error of the compressor efficiency at nrcd, C = 60000 rpm is quite high. This is mainly because of the low temperature difference and the low pressure ratio across the compressor at low rotational speeds.
7.3
Experimental results
Figure 8 shows the variation of the measured compressor efficiency with the compressor reduced mass flow rate at differcnt compressor reduced rotational speeds and total temperatures at the turbine entry for the GT1749V 55 Trim compressor. The measured compressor efficiency at T6t = 305 K represents the compressor adiabatic efficiency (cold measurements). The adiabatic efficiency is measured only in the operating range extending from 60000 rpm to 120000 rpm. This is because the fluid power supplied to the turbinc in thc cold measurements is not sufficient to drive the compressor at higher rotational speeds. Therefore, the adiabatic compressor performance is measured only for rotational speeds 60000-120000 rpm. The parameters extracted from the comparison of the cold and the hot measurements are therefore presented only in the operating range extending from 60000 rpm to 120000 rpm. The measured compressor efficiency in the hot measurements represents the non-adiabatic compressor efficiency. Figure 8 shows a cIear effect of the total tcmpcraturc at the turbine entry on the compressor non-adiabatic efficiency. The measured compressor non-adiabatic
127
...!...
•.
0.8,..-------,-----:---.,-----,.--,-------,
u
is "'" 0.7 cQ)
~ 0.6 Q)
" ~
'5..
0.5
Q)
.~
o
0.4
U) U)
~ 0.3 E o
o
0.2 +---,---t---..--t--..-j--..--t--.-'-t----r"-t----;;:-j--..-'--l 0.00 0.02 0.04 0.06 0.08 0.10 0.12 0.14 0.16 Compressor reduced mass flow rate
rileed.
c [kg/sf
Figure 7 Error analysis of the measured isentropic efficiency of the GTl749V 55 Trim compressor efficiency is lower than the compressor adiabatic efficiency due to the amount of heat transfer from the turbine to the compressor. The deterioration of the compressor non-adiabatic efficiency is very significant at low rotational speed (low compressor peripheral Mach number). The effect of heat transfer from the turbine to the compressor decreases with increasing the compressor rotational speed (increasing the compressor peripheral Mach number). Therefore, the deterioration of the compressor non-adiabatic efficiency decreases with increasing the turbocharger rotational speed. The measured compressor non-adiabatic efficiency at 140000 rpm and 160000 rpm is almost independent of the exhaust gas temperature at the turbine entry. This behavior coincides very well with the results of the parameter analysis of the turbocharger non-adiabatic performancc presented above. The deviation of the compressor non-adiabatic efficiency from the compressor adiabatic efficiency is higher near the compressor surge linc. This bchavior is due to thc increase in the specific amount of heat transfer to the compressor at low mass flow rates. This increase causes the higher deterioration of the compressor non-adiabatic efficiency at low flow ratc. Note that the behavior of the comprcssor non-adiabatic efficiency with the exhaust gas temperature at the turbine entry is non-intuitive. It was expected that the amount of heat transfer to the comprcssor increases with increasing the exhaust gas temperature at the turbine entry. However, Figure 8 shows that thc mcasurcd compressor non-adiabatic efficiency is not directly dependent on the exhaust gas temperature at the turbine entry. The highest deterioration of the compressor non-adiabatic efficiency for the 60000 rpm compressor performance line takes place at T6t = 873 K. For thc 100000 rpm compressor performance line, this occurs at T6t = 973 K. This behavior implies that the exhaust gas temperature at the turbine entry is not the only parameter affecting the amount of hcat transfcr to the compressor. The reason is that heat is first transferred by conduction from the turbine casing to the bearing housing. Thc lubrication oil removes part of this conduction heat transfer and the rest flows to the compressor. Thus, the lubrication oil also works as a cooling fluid for the bearing housing. It has therefore a significant effect on the amount of heat transfer to the compressor and hence on the compressor heat number. Figure 7 and Figure 8 show that thc mcasured deterioration of the compressor efficiency due to the amount of heat transfer to the compressor during the hot measurements is higher than the maximum probable error in the measured compressor efficiency. This is also tme for low rotational spccd (60000 rpm) despite the maximum probable error being quite high at this rotational speed, Figure 7.
128
T ,,:
--0-
305 K _ _ B06 K -----A- B73 K
O.B..------'i"'c'c-_-~--
C,)
~
.g
~~ c ~
g~ ~
iij~~I
0.6
0.5
C,)
a c
~.~ ~::E
0.4
~
0.3-t .. ·~~~··~~+'1\~~·~~~~+~·~~~~·~~~:
a
a 0.2 +-~-i----T~-ir--.....--j U 0.00 0.04 O.OB 0.12 0.16 Compressor reduced mass flow rate ri'\ed c [kg/s]
U
O.B,------'C"-"--_-...:..,...--
C,)
~
0.7
7
-+- 973 K
need, c = BOOOO rpm
= 60000 rpm
O.B...-_---'n'''Cedc.;.c,--=_1o~o_o_o_O_rp~m_-----;
0.2+-~-i----r~-i-,.........j
0.00 0.04 O.OB 0.12 0.16 Compressor reduced mass flow rate m cod, c [kg/s]
C,)
~ co ...... 0.7
.0
~~.
J 0.6
§
::&l g 0.5
~ ~ 0.4
E a
0.2 +-~-i-~--ir--~-i-~--l 0.00 0.04 O.OB 0.12 0.16 Compressor reduced mass flow rate meed. c [kg/s] C,)
O.B
~
.0 m"l"
~~ § ?
0.7 0.6
:: ~ 0.5 I a c (J)
E a
Q)
,,
need, c = 140000 rpm :
Figure 8
0.2+-~-i----T~-i-,.........j
O.B -r-_-,n,,,,,ed,,,,c,--=_16~0_0_0_0-,rp~m_-----;
C,)
.0
~ co
=s.
0.7
§ .} 0.6
::&l g 0.5
~ ~ 0.4
U
0.3 0.00 0,04 O.OB 0.12 0.16 Compressor reduced mass flow rate meed, c [kg/s]
~
~
/' l!
Q}
U
~ ~ 0.4
E
0.3
a
0.2 0.04 O.OB 0.12 0.16 0.00 Compressor reduced mass flow rate meed. c [kg/s]
U
Q)
0.3
.~.....
III
0.2 +-~-+~-1-~-i-~--I 0.04 O.OB 0.12 0.16 0.00 Compressor reduced mass flow rate meed, c [kg/s]
Variation ofthe compressor non-adiabatic efficiency with the compressor reduced mass flow rate at different compressor reduced rotational speeds and total temperatures at the turbine entry for the GT1749V 55 Trim compressor
8. CONCLUSION
The present theoretical and experimental investigation of the turbocharger non-adiabatic performance shows that the deviation of the compressor non-adiabatic efficiency from the adiabatic efficiency due to heat transfer to the compressor depends on several geometrical and operating parameters. Among these, the compressor peripheral Mach number and the compressor heat number [equation (11)] are the most important parameters affecting the compressor non-adiabatic performance. The deviation of the compressor non-adiabatic efficiency from the adiabatic efficiency increases with decreasing the compressor peripheral Mach number and increasing thc comprcssor heat number, either by increasing the heat flow to the compressor or reducing the inlet total temperature. Therefore, the amount of heat transfcr to the compressor introduces a significant systematic error (overestimation) in the measured product of the turbine efficiency and the mechanical efficiency at 129
low rotational speeds. This theory is confinned by an experimental investigation of non-adiabatic perfonnance of a GT1749V 55 Trim turbocharger. 9.
REFERENCES
[1] Chapman, K.S., Nguru, R. and Shultz, J., 2002, "Simplified Methodology to Correct Turbochargcr Field Measurements for Heat Transfer and Other Effects". Final Report for Gas Research Institute, GRI-02/o156. [2] Hagelstein, D., Beyer, B., Seume, J., Hasemann, H. and Rautenberg, M., 2002, "Heuristical View on the Non-adiabatic Coupling System of Combustion Engine and Turbocharger". IMechE International Conference on Turbocharging and Turbochargers, London, UK. [3] Malobabic, M., Mobarak, A. and Rautenberg, M., 1983, "Influence of heat transfer between turbine and compressor on the performance of small turbochargers". GTSJ International Gas Turbine Congress, pp. 566-574, Tokyo, Japan. [4] Malobabic, M. and Rautenberg, M., 1987, "Adiabatic and Non-adiabatic Efficiencies of Small Turbochargers". International Gas Turbinc Congress, Papcr No. 87-Tokyo-IGTC-105, pp. 57-64, October 26-31, Tokyo, Japan. [5] Puchcr, H., Berndt, R., Grigoriadis, P., Nickel, J., Hagelstein, D., Abdelhamid, S. and Seume, J., 2003, "Erweiterte Darstellung und Extrapolation von Turbolader-Kennfeldern als Randbedingung der Motorprozesssimulation". 5th Stuttgart International Symposium, February 18-20, Stuttgart, Germany, ISBN 3 816921809, pp. 140-156. [6] Pucher, H. and Nickel, J., 2002, "Vermes sung erweiterter Kennfeldbereiche von Fahrzeugmotorcn-Turboladern". 8. Aufladetechnische Konferenz, pp. 321-339, October 1-2, Dresden, Germany.
© Institute of Turbomachinery and Fluid Dynamics
130
Part-load Performance Prediction of Turbocharged Engines
Shaaban, S.I, Seume, J. I; Berndt, R.2, Pucher, H.3; Linnhoff, H.J. 4 1. University of Hannover, Turbomachinery Laboratory 2. Technical University Berlin, presently Ingenieurgesellschaft Auto und Verkehr IA V GmbH 3. Technical University Berlin 4. Linnhoff Engineering
ABSTRACT Engine simulation programs presently include the simulation of the engine together with its manifolds and the turbocharger. Most of the existing engine simulation models are based on measured turbocharger performance maps. Usually, these maps cover only a limited operating range. Therefore, they cannot be used to simulate the turbocharger performance at every engine rotational speed and are unable to adequately model partload performance which in passenger cars prevails throughout most of engine operating life. The present paper aims at providing a physically meaningful method of turbocharger performance prediction. This method takes into account the heat flow in turbochargers as well as the aerodynamic performance of the compressor and the turbine, respectively. It uses the measured performance maps as a basis for the turbocharger performance simulation, extending the compressor performance prediction down to zero rotational speed. Engine operating points are simulated using the method proposed above at low turbocharger rotational speeds to investigate the effect of heat transfer from the turbine to the compressor on the engine simulation in this operating range. The computational results of both, zero-dimensional and one-dimensional engine process models, show a better match with experimental data than previous engine calculations with nonphysically based turbocharger models.
NOMENCLATURE A area [m 2] BMEP break mean effective pressure [bar] absolute velocity [mls] C cp specific heat [J/(kg.K)] spouting velocity [mls] cs GVP guide vanes position thermal conductivity K [W/(m2 . K)] L length [m] ill mass flow rate [kg/s] compressor peripheral Mach MU2 number
rotational speed [min-I] n nred.T turbine speed parameter [min- I/K°.5] p pressure [Pa] Q amount of heat transfer [W] q specific amount of heat transfer [J/kg] R gas constant [J/(kg.K)] T temperature [K] peripheral speed [m/s] u W work [J/kg]
131
Greek Letters relative flow angle [0] 8 heat reduction parameter (/12 compressor impeller flow coefficient 7J efficiency K specific heat ratio Subscripts 1 compressor inlet 2 compressor impeller outlet 5 compressor outlet 6 turbine inlet 7 turbine rotor inlet 8 turbine outlet air air B bearing housing b blade C compressor eff effective
fJ
TC
f-l ~
q If
pressure ratio slip factor turbine degree of reaction loss coefficient velocity coefficient
exh exhaust gas heat for estimation of power is isentropic oil oil red reduced s static turbine T TC turbocharger total u peripheral direction
INTRODUCTION
Engine simulation programs have developed and extended to include the simulation of the engine together with its manifolds and turbocharger. The simulation of turbocharger performance aims at providing the required boundary conditions for engine simulation. Most of the existing engine simulation models are based on performance maps provided by the manufacturer or measured in a laboratory. These maps cover only a limited turbocharger operating range. Therefore, they cannot be used to simulate the turbocharger performance at every engine rotational speed. Another significant factor in the simulation of the turbocharger performance is the heat transfer between the hot turbine and the cold compressor. This heat transfer causes an overestimation of the calculated compressor and turbine power when using the measured isentropic efficiencies. The effect of this error is particularly strong at low turbocharger speeds. The present paper aims at providing a physically meaningful method of turbocharger performance simulation. This method takes into account both, the non-adiabatic turbocharger performance as well as the aerodynamic performance of the compressor and the, turbine respectively. It uses the measured performance maps as a basis for the interpolation and extrapolation of the turbine and the compressor maps. METHODS OF TURBOCHARGER PERFORMANCE SIMULATION
Turbocharger performance can be estimated by means of mathematical interpolation, operational similarity, mean value modeling, and CFD modeling. Mathematical interpolation makes use of the common mathematical interpolation methods to estimate the turbocharger operating characteristics at operating points between the measured points. There is no physical basis to this method. Moreover, mathematical interpolation cannot reliably be used for the extrapolation of performance maps. 132
Operational similarity makes use of the common dimensionless groups to estimate every operating point. In this method, the turbocharger performance maps are presented as functions of different dimensionless groups. Operational similarity can be used efficiently for interpolating the performance maps. However, extrapolating the compressor map with this method may result in extremely wrong estimated compressor characteristics. This occurs particularly at low compressor rotational speeds because the peripheral Mach number MU2 tends to zero while the impeller flow coefficient <1>2 tends to infinity. Mean value modeling predicts the operating characteristics by means of modeling the flow in the compressor and the turbine along a mean streamline. This method can be used efficiently in the interpolation and the extrapolation of the performance maps by choosing probable models for the slip factor and the aerodynamic losses. Many authors have used mean value modeling for the simulation of compressor performance. Aungier (2000, 1995) presented aerodynamic performance prediction models for centrifugal compressors. Mueller et al. (1998) presented mean value modeling for the turbocharger compressor assuming the compressor to be adiabatic. Mean value modeling of compressor performance was also used by Kolmanovsky and Stefanopoulou (2000). Mean value modeling of turbine performance using empirical loss correlations was applied by Chen and Winterbone (1990) and Baines (1998). CFD codes can be used efficiently for predicting certain turbocharger operating points. However, the very long calculation time currently makes it impossible to use CFD codes within engine simulation programs. TURBOCHARGER PERFORMANCE SIMULATION
It is desired in the present paper to simulate the turbocharger performance with a minimum amount of data regarding the turbocharger geometry because this data may not be available for the users of the engine simulation programs. Therefore, regression equations are used to simulate the turbocharger performance. However, these regression equations are based on physically meaningful equations representing the turbocharger performance. This is necessary for the extension of the turbocharger performance maps towards very low rotational speeds. The present turbocharger performance simulation is based on fluid properties at thermodynamic states corresponding to different physical locations of the turbocharger. These physical locations are shown in Figure 1.
133
Figure 1 Physical locations of the turbocharger Compressor Performance Simulation
The first step in the simulation of the compressor performance is the estimation of the amount of heat transfer to the compressor. The energy equation of the compressor under non-adiabatic operating conditions and in the absence of compressor inlet guide vanes takes the form cp ,air(T51 -TIt )= f.tu 2c2u,b +qe
(1)
Equation (1) can be rewritten in the form cp,air(T51 -Tit) = f.t(I--_2_J+k u~ tanP2,b u~
(2)
The amount of heat transfer to the compressor is represented by a simplified one-dimensional conduction equation . kBAB( ) Qe = Coil - - - T61 - T51 LB
(3)
The heat reduction parameter Coil accounts for the fraction of heat removed by the oil. Equation (1) can therefore now be rewritten as C p ,air(T51
2
U2
-TIt )= (1--_2-J . kBAB (T -T ) f.t r\ + cOlI 2 61 51 tanl-'2,b LB me U2
134
(4)
j=1
j=M
Impeller flow coefficient
<1>,
Figure 2 Schematic diagram for the modeling of the non-adiabatic work coefficient The left hand side of equation (4) is the compressor non-adiabatic work coefficient. The non-adiabatic work coefficient is greater than the actual compressor work coefficient due to the amount of heat transfer to the compressor. The non-adiabatic work coefficient can be estimated from the measured data. Regression equations for the non-adiabatic compressor work coefficient as a function of the impeller flow coefficient at each constant peripheral Mach number are then obtained within the measured operating range. The non-adiabatic work coefficient as shown schematically in Figure 2 is presented as a two-dimensional matrix. The i-components of this matrix represent the values of the non-adiabatic work coefficient at constant peripheral Mach number. The j-components of this matrix represent the values of the non-adiabatic work coefficient at constant impeller flow coefficient. The compressor slip factor is assumed to be constant for all operating points with the same impeller flow coefficient. Thus the application of equation (4) to the operating points with the same impeller flow coefficient results in Cp,air(Ts~ - TIt )1 U2
cp,air(Tst - TIt )1 .. J,I
2 u2
= kB AB [
L
j,i+l
. (T6t - TSt )1 eml
B
.
2 IDe U2
1
. (T6t - TSt )1 eoil.
j)
2 IDe U2
(5)
j,i+1
A regression equation is obtained for the heat reduction parameter Boil as a function of the compressor operating parameters. The coefficients of this regression equation are estimated by fitting the results of equation (5). Thus, the heat reduction parameter and hence the slip factor and the amount of heat transfer to the compressor can be estimated at each compressor operating point. A regression equation for the compressor slip factor can therefore be obtained. The compressor aerodynamic loss coefficient is first estimated from the equation
k
(6)
= Cp,air(Tst -Tst,is)-qc
u~ A regression equation is obtained for the compressor loss coefficient by fitting the estimated values of the compressor loss coefficient. This regression equation is based on the main loss mechanisms of the compressor. It includes terms that represent a. Impeller friction losses b. Impeller passage losses c. Impeller diffusion losses d. Diffuser losses e. Volute losses
135
Modified performance line
---~-----
----
Performance /Iine
Figure 3 Schematic diagram of a compressor performance line with a negative slope near the surge point The impeller friction losses are estimated from equivalent pipe friction loss models. The impeller passage losses are considered to be proportional to the mean kinetic energy of the air. The impeller diffusion losses are assumed to be proportional to the diffusion coefficient. The diffuser losses are presented as a function of the diffuser friction coefficient, diffuser-width to diameter ratio, diffuser diameter ratio, and the flow angle at the impeller outlet (Rodgers, 1984). The volute losses are considered to be due to friction loss, loss of meridional velocity, and exit cone losses. The compressor total pressure ratio is estimated with the help of the compressor loss coefficient. The total pressure ratio near the surge point is presented as a linear function of the compressor reduced mass flow rate. This is because the slope of the compressor performance line tends to be zero or negative near the surge point. Small performance line slope may cause convergence problems in engine simulation programs. Negative slope of the compressor performance line causes additional calculation errors and instabilities, because a certain compressor total pressure ratio in this case has two corresponding compressor flow rates as shown schematically in Figure 3. Therefore, the compressor performance line is presented as a linear function of the reduced mass flow rate near the compressor surge line. The limiting value of the compressor performance line slope is given as an input to the calculations. The linear relation is also used to extrapolate the compressor performance beyond the surge line. This extrapolation is necessary because some engine operating points may be located beyond the surge line that was measured using combustion chamber test rigs. The absolute Mach number and the relative Mach number at the compressor inlet, the Mach number based on meridional velocity at the impeller outlet, and the Mach number at the inlet of the volute exit cone are also estimated. Additional shock losses are considered if any of these Mach numbers exceed unity. These losses are estimated from the normal shock relations. Compressor choking takes place when the absolute Mach number at the compressor inlet, the Mach number based on meridional velocity at the impeller outlet, or the Mach number at the inlet of the volute exit cone reaches 1.
Turbine Performance Simulation The turbine mass flow rate is estimated from the equation .
mT =
C
dT
,
A
P6t
I ' " R exh T6t
elf T
(7)
T
The turbine flow function T is given by 136
T
(8)
=
(9)
The turbine discharge coefficient is first estimated from the measured turbine mass flow rate. A regression equation is obtained for the turbine discharge coefficient by fitting the estimated values of the turbine discharge coefficient. This regression equation is a function of turbine rotational speed parameter nred,T, turbine total-to-static pressure ratio to the power 0.204, and inlet guide-vane position. The latter is used only for the turbines with inlet guide vanes. The exponent of the total-to-static pressure ratio in the regression equation is estimated by Pucher (Pucher et at., 2003) based on the measurements of Zinner (1961). The turbine rotational speed parameter accounts for the effect of the centrifugal force on the turbine mass flow rate. An expression for the turbine aerodynamic work can be obtained by applying Euler's equation for turbomachinery to the turbine rotor (10) The actual turbine efficiency (called the turbine heat efficiency in the present paper) can therefore be represented as
(11)
Equations for the swirl velocities at the turbine rotor inlet and outlet are obtained using the turbine velocity coefficients (Balje, 1981). The turbine velocity coefficient is defined as the ratio of the actual flow velocity to the ideal flow velocity. This applies for both the turbine stator and the turbine rotor. Thus (12) (13) (14) Substituting for
(lhjatr)
2 u7 cs
C7u
and
CS u
in equation (11) and rearranging
+&i(u7)=\Jfr,statoPO~7~I-iRy +\Jfr,rotor'co~Or .[iRy +\Jfr,stato!l-iRy) Cs
(15)
-2\Jfr,statopo~7(u7/cs}Jl-iRy +&i(U7/Cs)2 f.5 where:
OT
=US/U7
(16)
Equation (15) can be rewritten as
137
llheat,T 2(U7/Cs)
+Oi(~7s )=r[(U/cJ,7t6t8s,nred,T, GVP]
(17)
A regression equation is obtained for the left hand side of equation (17) as a function of turbine blade speed parameter U7/Cs, turbine rotational speed parameter, turbine total-to-static pressure ratio, and guide-vane position (only for turbines with inlet guide vanes). This regression equation is used to estimate the turbine heat efficiency at each turbine operating point. TURBOCHARGER PERFORMANCE SIMULATION PROGRAM "TC2005"
A turbocharger performance simulation program (called TC2005) is developed on the basis of the method described above. The "TC2005" stand-alone program consists of two main subroutines. The first subroutine estimates the performance of the compressor while the second one estimates the performance of the turbine. Both subroutines can be implemented in the engine simulation programs to estimate the turbocharger performance at each operating point. The two subroutines are also designed to simulate the performance of several turbochargers at the same time. This is very important for the simulation of engines with multistage turbo charging. Weighted regression equations are used in order to improve the results of the turbocharger performance simulation. The coefficients of these regression equations are obtained by comparing the models described above with the measured performance maps. A weighting factor is implemented in obtaining these coefficients. This weighting factor depends on the distance between the point being simulated and the measured points on the performance maps. The more closely is the simulated point to a measured point, the higher the weighting factor of this measured point in estimating the regression equations. A GT1749V turbocharger by Honeywell Garrett is simulated using the TC2005 Program. Figure 4 shows the estimated performance of the GT1749V turbocharger using the program TC2005 and the measured one. The use of weighted and physically based regression equations results in a very good agreement between the estimated and the measured performance maps. The difference between the estimated pressure ratio and the measured one near the surge line is due to the linear relation implemented in this region to stabilize engine simulation programs, as illustrated in Figure 3. An extrapolation test is also executed to prove the ability of the program to extrapolate the compressor performance towards very small rotational speeds. These small rotational speeds cannot be measured accurately due to the probable very high and unacceptable systematic measuring error. Therefore, the smallest measured turbocharger rotational speed (80000 rpm in this case) is eliminated from the measured performance map and a new input data file (without the measured performance at 80000 rpm) is used with the TC2005 Program. The calculated performance is then compared with the measured one as shown in Figure 5. This figure shows a very good agreement between the extrapolated and the measured performance.
138
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139
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Figure 6 Comparison between measured and calculated performance map of the GT1749V turbine A very good agreement between the measured and the calculated performance of the GT1749V turbine is shown in Figure 6. It can also be noted in Figure 6 that the physically based regression equations take also the effect of centrifugal force on the turbine performance map at different turbine guide vanes position.
ENGINE PERFORMANCE SIMULATION Engine Simulation Programs One of the main goals of the development of turbocharged engines for passenger cars is to optimize turbocharger response during acceleration from part load. The investigations typically start at very low turbocharger speeds, corresponding to low engine speed and torque. It is very important to have precise knowledge of the efficiencies of the turbine and the compressor in the lower speed range for a precise definition of these starting conditions in the calculation model of the engine. Furthermore, separating the effect of the heat transfer from the turbine to the compressor from the aerodynamic work gives more realistic results when calculating turbocharger dynamics which is vital for the prediction of car acceleration. The TC2005 Program has been implemented in two different engine simulation program models. The first program model is the zero-dimensional model THEMOS®. This model is based on the quasi-stationary filling and emptying method developed at the Technical University of Berlin, Institute of Internal Combustion Engines. The second engine simulation program model is the one-dimensional model PROMO which is based on unsteady, compressible pipe flow (Linnhoff, 1985). To be able to represent the behaviour of turbo machines in unsteady flow, efficiency and mass flow rate must be calculated for each iteration step of PROMO. Therefore accuracy and calculation speed are important demands on the software. These demands are met by the evaluation of the weighted regression equations. The user now can take advantage of a safer reproduction of extrapolated map data, a much more accurate distinction between similar compressors or turbines, and a much improved handling of the input data. 140
Both engine simulation programs were used to simulate engine operating points located inside as well as outside the measured turbocharger performance map to check the effect of the physically based turbocharger simulation model described above in comparison with the classical mathematically based turbocharger performance simulation. Results of Engine Simulation Identical measured turbocharger maps were used as input data for the TC2005 Program as well as for both engine process simulation programs. Running the TC2005 program under the assumption of adiabatic turbocharger behaviour generates a reference calculation. Both engine simulation programs were used with and without the TC2005 Program. Without the TC2005 Program means that the engine simulation program uses mathematical methods to simulate the turbocharger performance. A comparison between reference calculations with and without TC2005 Program shows good agreement for all tested full load and part load operation points (n = 4000 min- 1 / full load and BMEP = 2 bar, n = 2000 min- 1 / full load). This good agreement is due to the fact that the turbocharger operation points in these cases are either located inside the measured turbocharger maps or only a small extrapolation is required for the turbocharger performance simulation. These results were expected because of the small differences due to reproducibility of turbocharger performance maps for both calculation methods and the good agreements for the TC2005 Program already shown above. The operation point n= 2000 min- 1 / BMEP = 2 bar (Figure 6) corresponds to a very low turbocharger rotational speed and shows differences up to 10% because of the relatively high and therefore unacceptable systematic measuring error. Furthermore the influence of the temperature at the turbine inlet on the compressor efficiency has an increasing effect for this engine operation point. The turbocharger map data are based on a turbine inlet temperature of T6t = 873 K, whereas the corresponding measured temperature is 558 K at the engine under consideration, so that the compressor efficiencies are taken from the map are unrealistically small. Based on the experiences from experimental investigations on the GT1749V turbocharger and on the engine operation at steady-state and dynamic conditions it seems to be sufficient to verify the TC2005 Program with its non-adiabatic model at the engine operation point n = 2000 min- 1 / BMEP = 2 bar. Figure 6 includes a comparison between reference and non-adiabatic calculation model for both engine process simulation programs. This comparison shows that the estimation of the temperatures before and after the turbine as well as the temperature at the compressor outlet is clearly improved with the consideration of the turbocharger non-adiabatic performance. With the assumption that the turbocharger was adiabatic during measurements, the maximum percentage difference of the estimated temperatures can be as high as 9%. This high percentage difference is due to the consideration of the heat transfer inside the turbocharger as useful turbine and compressor work. The maximum difference in the estimated temperatures clearly decreases with the non-adiabatic calculation and reaches a maximum value of only 2% (right ordinate). The improvement in the estimated temperatures is achieved by THEMOS® as well as PROMO. The consideration of the turbocharger non-adiabatic performance also leads to a clear improvement of the estimated pressure at the compressor outlet and the turbine 141
inlet. The maximum difference in the calculated pressure at the turbine inlet with the adiabatic calculations is about 4.5%. This maximum difference is reduced to less than 1% (right ordinate) by the consideration of the turbocharger as non-adiabatic. In summary, the correlations which were described theoretically before starting this investigation are reproduced correctly. The compressor power from the non-adiabatic calculation is lower than that from the reference calculation. The turbine power also exhibits a correct behaviour. As a consequence of the different turbocharger power balance, the turbocharger speed and the turbine inlet guide vane position are different for both calculation methods. The operating point presented in Figure 6 was chosen to examine the ability of the presented physically based modeling of turbocharger performance to extrapolate the turbocharger performance maps down to very low rotational speeds where the effect of heat transfer inside the turbocharger is significant. The improved estimated performance with the application of the TC2005 Program proves the ability of the models described in the present paper to simulate the turbocharger performance correctly and on physical basis. However, other operating points will be simulated in future work to further explore the limits of the models described in the present paper 10
-1
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Figure 6 Comparison between measured performance and the performance predicted with THEMOS® and PROMO at n = 2000 min-I, BMEP = 2bar (Berndt and Pucher, 2005) left ordinate: deviations of the adiabatic reference calculation (= "basis") Note the larger scale. right ordinate: deviations of the non-adiabatic calculation
142
CONCLUSIONS
A method for the simulation of turbocharger performance maps is presented. This method is based on the measured turbocharger performance maps. It aims at extending the simulation of the turbocharger performance to very low rotational speeds with a minimum amount of data regarding the turbocharger geometry. The presented method uses physically based regression equations to simulate the turbocharger performance. These regression equations weight the operating points close to the computed operating condition more highly to achieve higher accuracy in the simulation of the turbocharger performance. A turbocharger performance simulation program is developed on the basis of the method presented. This program can operate either as a stand-alone program or as subroutines in engine simulation programs. Comparison with measured turbocharger performance maps shows a good agreement between calculated and measured turbocharger performance. This also applies for the extrapolation towards low rotational speeds. The subroutines of the TC200S program were implemented in two different engine simulation programs. The verification of the turbocharger performance simulation program included in engine process simulation programs confirms the expected behavior. REFERENCES
[1] Aungier, R.H. (199S): Mean Streamline Aerodynamic Performance Analysis of Centrifugal Compressors. ASME Journal of Turbo machinery, Vol. 117, pp. 360-366. [2] Aungier, R.H. (2000): Centrifugal Compressors; a Strategy for Aerodynamic Design and Analysis. ASME PRESS, New York, ISBN 0 791800938. [3] Baines, N.C. (1998): A Meanline Prediction Method for Radial Turbine Efficiency. IMechE 6th International Conference on Turbocharging and Air Management Systems, Paper No. CSS4/006/98, London, UK. [4] Berndt, R.; Pucher, H. (200S): Einfluss eines diabaten Turboladermodells auf die Gesamtprozess-Simulation abgasturboaufgeladener PKW-Dieselmotoren. 1st Conference Engine Process Simulation and Supercharging, June 30 - July 1, Berlin [S] Chen, H.; Winterbone, D.E. (1990): A Method to Predict Performance of Vaneless Radial Turbines under Steady and Unsteady Flow Conditions. IMechE 4th International Conference on Turbocharging and Turbochargers, Paper No. C40S/008, May 22-24, London, UK. [6] Kolmanovsky, 1.; Stefanopoulou, A.G. (2000): Evaluation of Turbocharger Power Assist System Using Optimal Control Techniques. Society of Automotive Engineers SAE, PaperNo. 2000-01-0S19. [7] Linnhoff, H.-J. (198S): Die Berechnung des Ladungswechsels und Ansprechverhaltens von Verbrennungsmotoren mit Abgasturboaufladung. Diss. RuhrUniversitiit Bochum, ISBN 3-89194-0S3-X [8] Mueller, M.; Hendricks, E.; Sorenson, S. C. (1998): Mean Value Modeling of Turbocharged Spark Ignition Engines. International Congress Exposition, Paper No. 980784,pp.12S-14S. [9] Pucher, H.; Berndt, R.; Grigoriadis, P.; Nickel, J.; Hagelstein, D.; Abdelhamid, S.; Seume, J. (2003): Erweiterte Darstellung und Extrapolation von TurboladerKennfeldern als Randbedingung der Motorprozesssimulation. Final Report for 143
Forschungsvereinigung Verbrennungskraftmaschinen FVV, FVV Project number 754, Heft 774. [10] Rodgers, C. (1984): Static Pressure Recovery Characteristics of some Radial Vaneless Diffusers. Canadian Aeronautics and Space Journal, Vol. 30, No.1, pp. 42-54. [11]Zinner, K. (1961) Diagramm zur Bestimmung des Betriebspunktes einstufiger Abgasturbolader. M.A.N. Forschungsheft 10. © Institute of Turbomachinery and Fluid Dynamics
144
Development of Electrically Assisted Turbocharger for Diesel Engine y YAMASHITA, S IBARAKI, and H OGITA Mitsubishi Heavy Industries, Ltd, Tokyo, Japan
We are developing a Hybrid-turbocharger. At first, we intended to confirm a elimination of turbo-lag and an increase of the engine torque. All components are our original design. This paper describes the design and experimental results. The ultra high-speed motor and inverter, which are key components, are realized by strong permanent magnet and unique circuit. The results indicated that at engine speed IOOOr/min and 1200r/min the engine torque could be increased approximately 17%. INTRODUCTION Background
An electrically assisted turbocharger (which we call "Hybrid-turbocharger") has been attracting attention in the automotive industry. The new-type boosting system has a possibility to respond demands for automotives from society such as corresponding a regulation of exhaust gas amount and fuel consumption, and improving of drivability in particular a elimination of turbo-lag. System composition As shown in Figure 1, a Hybrid-turbocharger is composed by components of conventional turbocharger and a ultra high-speed motor which drives a compressor electrically. The motor is located in a turbocharger. The inverter which drives the motor is located outside and connected to three windings of the motor. Also, the inverter is connected to a battery by two wires. The motor can control the torque independently of the turbine torque. Additionally the motor torque direction can be controlled not only positive to rotation direction but also negative direction. Therefore, it is possible to assist the turbine torque at low speed and to brake at high speed. It indicates that at high speed the surplus energy of exhaust gas which is ordinary discarded through a waste gate can be retrieved as electric energy. inverter
battery
intercooler
Figure 1 A schematic diagram of Hybrid turbocharger
147
Estimation of effects The improvement targets we have intended by the Hybrid-turbocharger are below. •
Improving a fuel consumption by 10% at low speed.
•
Increasing an engine torque by 50% at low speed.
•
Improving a turbo lag by 70% at an initial acceleration.
To determine how large output of the motor was needed, we conducted an engine simulation by an in-house code. The engine specification in the simulation is a turbocharged 2.Oliter diesel engine with an intercooler. The diameter of the turbine impeller and the compressor impeller are respectively 43mm and 44mm. The turbine efficiency and motor one were set to respectively 70% and 90%. The compressor efficiency was based on experimental results. The simulation results are shown in Figure 2. The steady simulation results indicate that a lkW motor input will allow the increase of the engine torque by 50%, and a 2kW will increase to double. As to the engine fuel consumption, a lkW motor input will improve by 8%; a 2kW will improve by 12%. The transient simulation results indicate that using a 1.3kW motor will shorten the turbo-lag by approximately 50%, and using a 2.0kW motor will shorten by 70%. Based on above results, the proto type motor power has determined 2kW. In following chapter, detail specifications of each component are illustrated .
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Figure 2 The engine simulation results with the Hybrid turbocharger 148
DESIGN Turbocharger and Compressor
The diameters of the turbine impeller and the compressor impeller are respectively 43mm and 44mm as same as the simulation. The turbine impeller is diagonal flow type, and the compressor impeller has a backward curve. The magnet rotor of the motor is located between the turbine impeller and the compressor impeller. There are two bearings between the turbine impeller and the magnet rotor. As a result, the rotor is supported as an overhang configuration. It means the rotor dynamics must be considered carefully to prevent a fatal vibration. The vibration analysis results have indicated that it is sure that under operating speed, 220 000 r/min, there are two critical speeds, but the effective damping of full floating bearing can prevent from a growth of vibration. Power electronics Ultra high speed motor
As an ultra high-speed motor, PMSM (permanent magnet synchronous motor) was adopted. The magnets used for a source of magnetic fields have very high performance. The material is Nd-Fe-B (Neodymium-Iron-Boron) magnet, and the remanent flux density Br is approximately 1.26 tesla. The strong magnet can make the motor size smaller than the other type of motors, such as a DC motor, an induction motor, a reluctance motor. Another approach for downsizing is a concentration winding method for coils that induce rotation magnetic fields. In addition the motor has a large gap between the magnet surface and the inner surface of the stator teeth. Since an inductance of motor windings is almost inverse proportional to gap length and react as a resistance for a rapid current change, a small gap leads to the quick response of a motor torque. In an ultra high-speed motor, it is very important to retain the rotor against its own centrifugal force. In our motor, the rotor is wound by a carbon fiber. As a negative side effect of downsizing, it is difficult to cool down a motor against its self-heating due to power loss. The excessive temperature rise causes a deterioration of insulation performance and irreversible demagnetization of magnets. In an ultra highspeed motor, the heat generation of magnet can not be negligible because the high frequent fluctuation of the magnet flux induces a eddy current loss in the magnet. To reduce the fluctuation of magnet flux on magnet surface, the stator has six teeth. Loss analysis of the motor was conducted by using a finite element method (FEM) and a boundary element method (BEM). For enough cooling, the stator and rotor of the motor are cooled by forced oil cooling and by forced air cooling respectively. By this cooling, the maximum temperature rises of rotor and stator are held under 140 degC and 155degC respectively.
149
Figure 3 Overview of the Hybrid turbocharger and motor stator
Inverter
In general, a voltage-source PWM inverter is widely used to operate a PMSM, where a vector control scheme is usually employed. However, the sinusoidal motor current regulation is hardly possible in the case of ultra high-speed drives because the fundamental frequency of the motor current is several kilo-hertz even at the rated speed, which does not allow the inverter to perform the pulse width modulation appropriately. Therefore, we adopted a pseudo-current source inverter for ultra high-speed drive, which is easier to control the current value and to synchronize the current phase to rotor phase. This inverter consists of two parts, i.e., a current-controlled buck-boost chopper and a 120 (electric) deg conduction 6-step inverter. The former circuit employs to control the motor current according to the current demand. In other words the circuit regulates the current by means of a pulse amplitude modulation (PAM) technique which adjusts the chopper duty ratio proper. The switching frequency of chopper is 30kHz. When the motor is in a motoring mode, the chopper acts as a buck chopper. On the contrary, when the motor is in a generating mode, the chopper acts as a boost chopper. The latter circuit employs to synchronize the current phase to the rotor phase which is detected by mechanical sensorless method. In permanent magnet motor, the rotor position can be detected by induced voltage of coil and additional analogue detect circuits without position sensor. Strictly speaking, since it is difficult to detect the induced voltage at low speed, the phase control is conducted by a open loop control at starting.
150
C
Vdc:DC power source, C: electrolytic capasitor, L: series reactor, D:Diode, PM:Permanent magnet Motor, Sc\, Sc2, S\, S2, S3, S4, S5, S6: FETs
Figure 4 The inverter circuit diagram and overview
EXPERIMENTAL RESULTS AND DISCUSSION Stand-alone experiment
Prior to the experiment with an engine, the stand-alone experiment has been conducted to verifY the basic performance. The turbine was given a torque and controlled its rotation speed by the source of air in laboratory. The voltage source of inverter which drove motor was not a battery but rectified DC voltage by 3-phase rectifier and capacitors. The reason was that the proto type Hybrid-turbocharger was designed as nv specification, and it was expected to operate for a certain amount of time without worrying about a limit of battery storage capacity. Table 2 shows a difference of rotation speed and pressure ratio between the normal turbocharger (without motor assist) and the Hybrid-turbocharger (with motor assist). Under this stand-alone experiment, the Hybrid-turbocharger's torque was controlled constant at about 0.l6 Nm which was equal to a power of2kW at 120000 r/min. Figure 5 shows comparisons of boost pressure and rotation speed between the normal turbocharger and the Hybrid-turbocharger. Focusing an attention on starting point where the effect of motor assist is significant, the time to the rotation speed 74 000 rlmin was cut to 33% and the time to boost pressure 26 kPa was cut to 41%. A few seconds after the start, the turbine torque was more dominant than motor torque. Therefore large differences of increasing rate between normal turbocharger and Hybrid-turbocharger were not seen.
151
Table 2 Comparisons of speed and pressure ratio between with and without motor assist Before assist 40000 81000 100000 122000 Speed[r/min] After assist 85000 109000 121 000 138000 Pressure Before assist 1.05 1.21 1.50 l.33 ratio After assist 1.23 1.41 1.51 1.66
90 80 70 ~ 60 ~ ~ 50 a(J)(J) 40 30 Q) ~ 20 8 10 0
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Figure 5 Comparisons of speed and pressure response between with and without motor assist
152
12
Experiment on engine The effect of the Hybrid-turbocharger for the engine performance has been verified on the engine test bench in our laboratory. The engine displacement is 1.7liter turbocharged with intercooler. The rated power is 75kW at 4600r/min, and rated torque is 240Nm at 2400r/min. In a default configuration, the turbocharger is VG (Variable Geometry) turbocharger whose diameters of the turbine and compressor are 40mm and 44mm respectively. When the motor speed was under 120 OOOr/min of the rated speed, the torque was controlled at O.l6Nm in constant, and above the rated speed the power was controlled at 2kW in constant. Figure 6 shows a difference of steady state engine torque between using the VG turbocharger and using the Hybrid-turbocharger. At engine speeds 1000r/min and 1200r/min, the engine torque was increased approximately 17%. This improvement is assumed to be caused mainly by an increase of air-fuel ratio. Figure 7 shows the difference of transient turbocharger speed and boost pressure between using the VG turbocharger and using the Hybrid-turbocharger. The start condition ofthe engine in this experiment was high idling, and suddenly engine was full loaded. In the Hybridturbocharger, the motor torque was generated at the same time as the engine load. The increasing rate of the Hybrid-turbocharger was larger than the VG turbocharger. Therefore, at least the response of the engine torque by Hybrid-turbocharger would be equal to the one by VG.
Assist motor output: 2kW
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154
CONCLUSION In this paper, it is mentioned that the effects of Hybrid-turbocharger which have a assist motor attached to conventional turbocharger are verified through experiments as follows: • At the starting point of the boost, by a 2kW-120 OOOr/min motor the time lag of speed could be cut to 33%, and the time lag of boost pressure could be cut to 41 %. • At engine speeds 1000r/min and 1200r/min, the engine torque could be increased approximately 17%. We should continue to discuss the following points in the near future: • Building up the 12V or 24V system without ill effect for other electrical component. (Note: The present system requires nv input. ) • Establishing the control strategy well matched to other component control, such as engine control and waste gate control. • Finding the most beneficial combination of turbine, compressor and motor by using a simulation model.
REFERENCES (1)
(2)
(3)
K Fieweger, H Paffrath, and N Schorn, Diesel Powertrain Systems, Ford Forschungszentrum Aachen, Germany, Drivability assessment of an HSDI Diesel engine with electrically assisted boosting systems, ImechE, London, 2002 S.George, G.Morris, J.Dixon, D.Pearce and G.Heslop, VISTEON U.K. Ltd., Optimal Boost Control for an Electrical Supercharging Application, 2004 SAE World Congress Craig Balis, Chris Middlemass, S.M.Shahed, Garrett Engine Boosting Systems, Design & development of e-turbo for SUV and light truck applications, DEER 2003
© Mitsubishi Heavy Industries Ltd.
155
The Design And Testing Of An Electrically Assisted Turbocharger For Heavy Duty Diesel Engines. Owen Ryder - Holset Engineering Co. Ltd. Herbert Sutter - ATE GmbH Laurentius Jaeger - Iveco Motorenforschung AG ABSTRACT
Better fuel consumption, quicker transient response and reduced emissions are claimed to be possible by using a turbocharger that has an integral electric motor/generator. This paper covers the mechanical design of several turbocharger and electric motor options and the selection of the design used in this EU supported project. The paper also covers the testing of the turbocharger on a gas stand where the performance capabilities of the motor and turbocharger were established, both in powered mode and as a generator. The effects of the heavier and longer rotor assembly on the dynamic shaft motion are also investigated. Finally, the paper summarises the testing done on a heavy duty truck diesel engine, demonstrating the benefits and limitations of such a system. INTRODUCTION
Electrically assisted turbochargers have been used for the purposes of increasing boost at low engine speeds and reducing turbo lag during transient events, thus giving performance benefits. Many of the benefits of electrically assisted turbochargers can be gained using variable geometry turbines, but further benefits can be obtained if the system is considered to be turbo assisted electrics. If the auxiliary equipment is electrically driven, then the electrical power to drive these can be obtained from the exhaust gas by the turbine, instead of the auxiliaries being mechanically driven parasitic loads. Simulation work (1,2) has shown that this type of system could reduce fuel consumption. The "ELEGT" project was set up to investigate the performance and benefits of such a system and is 50% funded by the European Commission together with Switzerland and incorporates six partners. These are:Holset (UK) Turbochargers A.T.E. (Germany) Electric motors Thien (Austria) Electronics Motor and system simulation work University of Durham (UK) Iveco Motorenforschung (Switzerland) Engine Iveco (Italy) Engine and Vehicle The electric motor specification is designed to give enough power to accelerate the turbo rapidly during transient events and enough generation capacity to power the vehicle's auxiliary systems. The motor specification is 1.0Nm up to 60 OOOrpm. The power is then capped at a maximum of 6kW above 60 OOOrpm in powered mode and 7.6kW in generation mode above 90 OOOrpm. The maximum design speed of the turbocharger is 130 OOOrpm but it should be capable of surviving 143 OOOrpm overspeed condition.
157
This paper describes the design processes for the hardware and the performance testing of the completed turbocharger, both on turbocharger gas stands and diesel engine test beds. The electric motor aspects of this project have been reported in IEE papers (3,4). ELECTRIC MOTOR TYPE SELECTION
Besides the performance specification of 6kW in powered mode and 7.6kW in generating mode, the motor must meet the following operating requirements: • able to survive at 300°C shaft temperature in worst case, • be accommodated in a small space with highest power density, • able to survive the highest speed of 143000 rpm (max. overspeed). Selection of motor type
A list of the motor types considered is shown in Table 1. The temperature limit of a permanent magnet machine would be of concern in a turbocharger application. Other drawbacks include the need for magnet reinforcement, magnet cost and complicated assembly requirements. An induction machine was used for this application as a robust motor able to cope with the high shaft temperatures and large air gap between rotor and stator. In order to achieve the maximum power density, the stator must be water cooled. In addition, an advanced potting material (220°C max. allowed temperature) was used to encapsulate the winding resulting in good heat transfer to the coolant. To reduce the stator copper losses, the filling of the copper windings inside the slots was maximized. In the rotor, pure copper bar material was used, the ends of which were soldered to the high stability copper end rings. The motor was equipped with special Cobalt Iron lamination material, being able to saturate the motor more than 2 Tesla without having excessive iron losses. In order to keep the mechanical stresses of the rotor to a minimum, it was necessary to keep the rotor diameter as small as possible. This was the main reason for choosing a radial flux machine over an axial flux machine. Considering the implications on the electronic driver properties, a two pole motor design was chosen in order to minimise the base frequency (2400 Hz). The motor drive electronics were designed to control field current independently of torque current to avoid unnecessary heating of the stator. An induction motor is relatively easy to control for motor as well as generator operation. A speed sensor was used to provide speed feedback to the drive electronics, although in the future, sensorless control should be possible. Simulation results
A.T.E. used its own in house single body thermal model software to simulate the winding temperature during generator operation of 7.6 kW @ 120 000 rpm. The simulations showed that the winding temperature would increase by up to 90 K. Mechanical stress distribution was calculated by Finite Element Analysis techniques at the rated speed of 130 000 rpm. This gave a maximum stress of around 320 MPa at the inside diameter of the rotor, as shown in Figure 1.
158
Table 1 - Different Types Of Motors Considered No. Motor Type
1
Induction asynchronous
2
Induction asynchronous
3
Permanent Magnet slotted
4
Permanent Magnet slotted
5
6
Permanent Magnetslotless
Permanent magnet slotless
Details • 2 poles • 3 phases • Cast Aluminium rotor bars & end rings
Remarks • Reinforcing of end rings required. • Potential of poor quality cast bars influencing the motor characteristics
• 2 poles • 3 phases • Cast Copper rotor bars and end rings
• Reinforcing of end rings required. • Good quality casting • Mass production / tool stress in the die-cast process • Possibility of integrated reinforcing of end rings • Reinforced ring magnet • Restriction of max. allowed magnet temperature
• 2 poles • 3 phases • Magnets places on rotor surface. • 2 poles • 3 phases • Magnets embedded below rotor surface • • • •
2 poles 3 phases Surface magnet Straight winding system on a plastic coil carrier system • Full diameter winding system
• • • •
•
Permanent
7 magnet -
• • • •
slotless •
2 poles 3 phases Surface magnet Straight winding system on a plastic coil carrier system Reduced winding pitch 2 poles 3 phases Surface magnet Straight winding system on a plastic coil carrier system Skewed winding system
159
• Mechanical strength of rotor shape. • Motor able to operate in field weakening range • Restriction in max. allowed magnet temperature • Reinforced ring magnet • Winding - system is in the magnetic field (eddy currents in the winding system) • Large end windings • Restriction of max. allowed magnet temperature • Reinforced ring magnet • Winding - system is in the magnetic field (eddy currents in the winding system) • Shorter end windings • Restriction in max. allowed magnet temperature • Reinforced ring magnet • Winding - system is in the magnetic field (eddy currents in the winding system) • No winding heads • Restriction in max. allowed magnet temperature
I-~-=--I====>,..----
Up to 200MPa at rotor surface
______----.
ANSYS 5.5.2
~~~4!;4~005
NODAL SOLUTION
~~~P:i TIME=l SEQV (AVG) PowerGraphics EFACET=l AVRES=Mat DMX =.025327 SMN =54.701 SMX =317.961 54.701 83.953 113.204 D 142.455 D 171. 706 D 200.957 D 230.208 D 259.459 CJ 288.71 317.961
Maximum stress 318MPa at rotor bore
-
Fig. 1: Mechanical stresses in motor rotor at 130 000 rpm. Improvements to design
During the course of the project, three different designs of induction motor were tested. The Mkl.O has a 056 x 38 mm stator, Iron Silicon laminations and an industry standard potting compound. In order to improve the thermal situation within the same dimensions, a Mk1.l motor was created with Iron Cobalt laminations and potting compound with a temperature capability up to 200°C. The Mk2 motor also has the Iron Cobalt laminations but the dimensions are larger and the potting compound has a limit of 220°C. TURBOCHARGER DESIGN Basic design
The turbocharger design had to incorporate the following features:1. Fitting a motor rotor onto a turbocharger shaft that gave acceptable shaft motion. 2. Location and assembly of bearings. 3. Accommodate the motor in a bearing housing that provided suitable cooling. The turbocharger design features a motor rotor heat shrunk onto a conventional shaft and wheel, with the turbine end bearing preassembled into a bearing housing end cap. The stator is pressed into the bearing housing which has an internal cooling passage. Conventional floating hydrodynamic bearings were specified. This allows operation at high speed and the rotor benefits from the cooling effect ofthe oil flow. Mkl turbocharger design
The first design had a 056 x 38 mm motor that would fit within the size of a conventional HY 40V turbocharger. This had spatial restrictions that meant several compromises were made, most notably that the cooling passage was within the bearing housing and not in direct contact with the stator. This was smaller than ideally required for a motor of 6kW output. After the first test work was done it became clear that a larger motor and better cooling would be required.
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Mk2 ELEGT Turbocharger • Motor 090 x 65 mm • Water cooling (wet stator) • 78mm bearing span
Fig. 2: Cross section through the Mia ELEGT turbocharger Mia turbocharger design
A second turbocharger was designed that placed motor requirements ahead of the turbocharger design. The stator was larger at 090 x 65 mm, requiring a longer shaft which may have caused an increase in shaft motion. The larger diameter also meant that the turbine volute was pushed away from the turbine entry and inlet vanes had to be used to maintain turbine efficiency. Shaft motion
The Mkl design of motor rotor was modelled along with a standard HY 40V shaft and wheel in Holset's rotor dynamics software. The results showed that natural frequencies were lowered and problems may be encountered at the maximum speed of 130 000 rpm but that improvements may be seen at lower speeds. A prototype turbocharger was tested with a dummy mass of the correct size and density and the test results were more encouraging, showing that the shrunk on motor rotor stiffened the shaft and reduced shaft flexing at high speeds. The end result was that the shaft motion was reduced and the oil film thickness increased relative to the standard HY 40V rotor system. For the Mk2 turbocharger, the longer shaft was expected to lower the frequency and weaken the shaft further, but the simulations show that the shrunk fit motor rotor is beneficial. The stiffening benefit allows the shaft to be increased in length by 20mm before shaft motion at the turbine wheel would become too large. -,--'
Fig. 3: Shaft motion predictions for the 20mm longer shaft. 161
Thermal considerations
The motor stator has to be kept below 220°C in order to avoid damage to the potting material and the winding insulation. Turbine temperatures are typically around 700°C and so water cooling was considered to be necessary. A full conjugate heat transfer calculation was performed on three design iterations using the CFX-5 code. The designs were: 1) Water jacket within the bearing housing. 2) Water jacket integrated into the stator. 3) Water jacket integrated into the stator with spiral guide vanes. Design 1 with water jacket within the bearing housing resulted in a hotspot where the oil drain cavity was located. Design 2 with the simple water jacket around the stator gave more even cooling than Design 3 with the spiral vanes. Whilst more time could have been spent optimising the vane design, the simpler design was expected to give sufficient cooling and so the simple design was chosen as it also gave more even cooling across the stator. An even temperature across the motor windings was considered important to ensure optimum motor performance and durability. Consideration was given to hot shut downs and it was found that the coolant must be circulated after the turbocharger is stopped in order to avoid the high metal temperatures in the turbine housing soaking through to the stator and causing damage. Assembly considerations
The assembly of the ELEGT turbocharger differs from a conventional design because the turbine end bearing is located in an end cap to allow the motor rotor to be heat shrunk onto the shaft. The end cap must be a press fit into the rest of the bearing housing to ensure accurate concentricity with the compressor end bearing. Oil must be supplied to the end cap for the turbine end bearing and sealed to avoid leakage to the turbine or atmosphere. In order to reduce the amount of oil splashing onto the motor parts, baffles are installed around the bearings to guide oil towards the oil drain and away from the rotor to stator gap. TURBOCHARGER TESTING
The ELEGT turbochargers were tested on a gas stand test rig at Holset in order to establish the performance of the motor, in particular, the length of time the motor could be operated at before reaching its maximum allowed temperature. Motor power only
From stationary, the motor was used to power the turbo up to a target speed of 30 000 rpm. The stator temperature was then monitored until a temperature rise of less than 1°C per minute was observed. Fig. 4 shows the comparison between the different motor designs. The Mkl motor with the first power controller gave a temperature of 150°C within 5 minutes. The Mkl.l motor with the second power controller was better at 135°C but the Mk2 motor gave a significant improvement running at 82°C in 26°C ambient conditions.
162
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.$ (f)
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Mk2 Motor - second controller with choke to reduce high frequency noise.
0
o
50
150
100
200
250 300 Time (sees)
400
350
450
500
Fig. 4: Motor stator temperatures whilst running at a target speed of 30 000 rpm powered by the electric motor only. This is a shaft power of around lkW. Altering turbo speed
Whilst speed increases were shown on the test rig, it was not possible to replicate the feedback loop that would occur on an engine, thus the true benefit of electric assist must be evaluated on the engine. The motor was tested on the gas stand up to the speed of 100 000 rpm to verify that the motor could operate at such conditions. A change in shaft power of2kW was achieved for more than 2 minutes. A reduction in shaft power of 2 kW was achieved at 100 000 rpm also for 2 minutes. A graph of the temperature rise is shown in Fig.5. Although unintentional at the time, a rapid braking event reducing the speed to 86,000rpm was achieved over a 3 second period. This represents a change in shaft power of 12kW and is shown towards the end of the data captured in Fig.5. Despite being significantly more than the rated power for the motor, this event did not damage the motor in any way. 200
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Fig. 5: Reduction in turbo speed on Mk1.1 turbo by generating electricity.
163
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- Stator Temp
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E
ENGINE TESTING The Mkl.1 ELEGT turbocharger was mounted on a standard production CURSOR 8 engine (7.8Iitre, 228kW), installed in a test cell at Iveco Motorenforschung AG in Arbon. The initial days of operation were to verify that the ELEGT is capable of: 1. Improving the transient response of a fixed geometry turbocharger, and to 2. Limit efficiently turbo speed, by generating electric energy. Fig. 6 is a schematic that shows the arrangement of hardware and information exchange. The electric power supply is from mains and any generated energy is dumped as heat via resistors. ,....------1 Test rig management
Combustion Engine
Power analyser Rapid prototyping tool
Chopper & resistance Power supply
Torque
Fig. 6: Scheme of ELEGT installation on test engine. Test equipment set-up A major part of the effort was dedicated to the electric part of the set up. The challenging control of a high power density, high speed electric motor is embedded right from the start in the electronic control system of the diesel engine. Lubrication and water cooling (not shown in the diagram) of the turbocharger are provided from the test facility, independent from the engine. Oil and coolant are supplied at the correct temperatures and pressures even without the IC engine running. Many functional tests of the bulky prototype electronics could thus be accomplished without exposing the first turbocharger prototype to high load conditions in a hot environment. It is noted that the Mkl.l electric motor can not accelerate the rotor to more than roughly 30 000 rpm, even with the compressor outlet disconnected from the intercooler, breathing free into ambient. Typical power flows through turbines are many times larger than what can be conveyed as current through an electric motor which is small enough to fit into a turbocharger's bearing housing. It is clear that the electric machine's assistance and recovery can account only for a "delta" around most turbocharger operation points. The ELEGT is built into a bearing housing that is derived from the Variable Geometry Turbine (VGT) turbocharger that is in production on CURSOR 8. The swallowing capacity of the ELEGT prototype can be matched for test purposes to the engine by means of a batch of fixed nozzle rings similar to the one that normally slides in ourVGT.
164
Load response, operation as a motor The chosen nozzle ring gives a good fuel economy at higher loads and speeds, at the price of poor response to demanded sudden load increases from part loads. In fact the engine operates in smoke limitation status for 7 seconds, during a typical response at lower speed, with this nozzle ring. This value is typical for wastegated turbo charging. Giving a 1 second long burst of electric power, corresponding to 0.3 Nm at the turbine shaft, the time in smoke limitation is reduced to 2 seconds. This is a value typical for VGT turbocharging, accomplished without changing the turbines geometry. Power for this test was provided by the mains.
Turbocharger speed limitation, operation as a generator The IC engine is operated at four speeds with 100 000 rpm turbo speed. Without changing the IC engine's fuelling, turbo speed is reduced to 96 000 rpm by generating about 1.5 kW electric, before converter. This corresponds to 0.09 Nm torque at the turbine shaft, applied during several minutes. Then the fuelling is increased until turbo speed is again 100 000 rpm. The previous turbo speed is run now with 5 to 9 % more power. The same scheme is shown here also for 90 000 rpm, where the 0.09 Nm torque at the turbine shaft reduce turbo speed to 85 000 rpm. This is shown in Fig.7. 20 18 16 14
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""
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1500
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Fig. 7: Engine operating conditions with and without ELEGT generating electrical power. Mkl.l ELEGT on Iveco Cursor 8 engine.
165
CONCLUSIONS The induction motor has been developed such that it can operate in a turbocharger, with the air gap that is necessary for the hydrodynamic bearings. The additional mass added to the rotor system is not detrimental to the shaft motion of the turbocharger and in some cases it improves it. The temperature of the stator is a limiting factor, with power levels of 2kW sustained continuously. Higher powers can be developed for short periods of time, providing the stator can be effectively cooled prior to the next event. The transient response of a turbocharger on the engine can be improved significantly with a built in electric motor, which reduces the engine's turbo lag and transient emissions. Mk2 will be able to generate electric power over a considerable area of the engine's operation range, whenever an air surplus allows this. Integration of an electric machine into the turbo-machinery for a diesel engine system has been achieved with a reliable first prototype, limited however in its temperature, respective power capabilities and electronics. A major effort is required on the integration of electric machinery into the architecture of future automotive powertrain systems.
REFERENCES 1) Millo, F., Mallamo, F., Pautasso, E., Ganio Mego, G.; The potential of Electric Exhaust Gas Turbocharging for HD diesel engines, Society of Automotive Engineers World Congress, Detroit, April 2006 2) Pautasso, E.: "Numerical simulation of the ELEGT (Electric Exhaust Gas Turbocharger) solution for Heavy-Duty diesel engines", Mechanical Engineering Graduation Thesis, Politecnico di Torino, Turin, Italy, 2004. 3) Bumby, J.R., Spooner, E., Carter, J., Tennant, H., Ganio-Mego, G., Dellora, G., Gstrein, W., Sutter, H., and Wagner, J.; "Electrical machines for use in Electrically assisted turbo-chargers", IEE Int. Conf. on Power Electronics, Machines and Drives, Edinburgh, MarchiApril2004, pp344-349. 4) Bumby, J.R., Crossland, S., Spooner, E., Carter ,J.: "The development of turbocharger accelerator motors and drives and their integration into vehicle electrical systems", Proc. of the IEE Conference on Automotive Electronics, London, March 15-16, 2005 © Holset Engineering Co Ltd
166
A Numerical Study of the Performance Characteristics of a Radial Turbine with Varying Inlet Blade Angle Liam Barr, MEng, AMIMechE Stephen W. T. Spence, CEng, PhD, MIMechE, MIEI Tony McNally, PhD, BSc Queen's University ofBelfast, UK SYNOPSIS The inlet blade angle of radial turbines is almost invariably fixed in the radial direction to avoid imposing additional bending stresses at the blade tips. However, a small number of published studies have shown that changing the inlet blade angle can shift the efficiency characteristic to bring improvements at off-design conditions. Such a move in the efficiency characteristic would benefit turbocharger performance where the turbine typically experiences lower than optimum velocity ratios while accelerating during engine transients. This paper details a numerical study of varying inlet blade angle. A ID performance prediction routine has been used to forecast changes in the overall performance characteristics of a turbine with non-radial inlet blading, revealing a shift in the peak efficiency point and an increase in efficiency at lower values of velocity ratio (UlC) when using a back swept blade angle of 30°. The numerical modelling strategy, which was fully validated against existing experimental measurements, was used to verify the findings of the 1D analysis and show how the back swept blade angle significantly reduced regions of flow recirculation at inlet during off-design operation. FEA revealed stress levels up to 51 % of the material yield strength at the inlet region of the pressure side on the 30° back swept blade. A brief review of new nanocomposite material technologies is presented to demonstrate the potential they provide for increased freedom in aerodynamic blade design. NOMENCLATURE 10
3D CFD FEA
One dimensional Three dimensional Computational Fluid Dynamics Finite Element Analysis
CNTs PR UlC
Carbon nanotubes Turbine pressure ratio Blade speed velocity ratio
INTRODUCTION The radial turbine is known to deliver optimum efficiency at a blade speed velocity ratio (UlC) of approximately 0.7. However, in vehicle turbocharging applications the turbine spends a significant part of its operating time accelerating from relatively low speeds in response to engine transients. Consequently, the turbine often operates at a UIC value well below the optimum design value, and rarely above the optimum value. Transient engine performance, and particularly transient exhaust emissions, are closely related to the level of boost pressure achievable. Rodgers [1] stated that a 1% increase in turbine performance would result in around 5% more torque available for turbocharger acceleration and significantly reduce the time taken for the turbocharger to reach full boost pressure during a transient. Therefore, improvements in turbine efficiency at
169
lower values of VIC would yield significant benefit during the engine transients that account for much of engine operating time. Turbine performance at lower than optimum VIC values may be enhanced by employing a back swept inlet blade angle. However, the need to avoid incurring additional bending stresses in highly stressed turbine blades means that there are practically no examples of commercial turbines with non-radial inlet blading and there have been few published studies of non-radial inlet blade angles. Most notably, Mulloy and Weber [2] published their findings from a series of experimental tests conducted on radial turbine rotors with unconventional inlet blade shapes intended to reduce the offdesign losses. Three rotors were produced and tested: a forward curved thick-bladed rotor, a round nosed thick-bladed rotor and a backward curved thick-bladed rotor. The actual value of inlet blade curvature from the radial direction was not specified. The tests, which were conducted by varying rotor speeds at constant turbine pressure ratio, showed that at low rotor speeds the backward curved blades appeared to align better with the flow yielding an efficiency increase. Conversely, the forward curved blades demonstrated better flow alignment and efficiency at higher VIC values. Mass flow rates followed similar trends, with the backward curved blades giving increased flow at lower UIC values. Figure 1 illustrates the change in measured performance. (Throughout the rest of this paper turbine blade angle is described by the terms 'forward swept' and 'back swept', as conventionally used to refer to centrifugal compressors, where forward swept blades are those with an inlet portion angled in the direction of rotation. Likewise, back swept blades are those with an inlet portion angled against the direction of rotation.) Efficiency Vs U/C at 1.S PR
so 75
65
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______ Round Nose Thick·Bladed Rotor
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_
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______ Std Thin·Bladed Rotor
0.45
0.50
0.55
0.60
0.65
0.70
0.75
o.so
U/C
Figure 1 Impact of rotor inlet blade angle on turbine efficiency [2] Mulloy and Weber commented on potential problems that might arise with these unconventional blade shapes. The thick blades produced a substantially tighter radius of curvature in the blade-to-blade plane, which it was thought could result in severe velocity gradients producing exit flow distortion, thus increasing the difficulty of
170
recovering the exit velocity head. Another concern was the increased moment of inertia and magnified stress levels that would arise from the thick blading. Meitner and Glassman [3] detailed the modifications made to an existing I D procedure for predicting the off-design performance of radial turbines to incorporate a rotor slip factor correlation that included rotor blade sweep at inlet. The resulting computer programme was used to predict the performance of a radial turbine with back swept inlet blade angles of 0°, 15° and 30° from the radial direction. The authors showed that there was an efficiency advantage at high values of work factor (corresponding to lower than optimum U/C values) as a consequence of employing backward swept blades at rotor inlet. Hakeem [4] presented an equation derived from a ID treatment of inlet and exit velocity triangles to show how the velocity ratio U/C corresponding to the peak efficiency point varies as a function of the inlet blade and flow angle.
(U) C
= optimum
0.707
tan fJinlet
(1)
tana;nlet
The derivation of equation 1 assumes zero incidence between the inlet flow angle and the inlet blade angle {3;nlet. For a purely radial inlet blade angle, tan{3;nlet = 0 and the optimum U/C is 0.7, but for a back swept inlet blade angle the optimum U/C is less than 0.7. Fredmonski et al [5] conducted a research program with the objective of designing, fabricating and testing compact radial inflow turbines with equal or better efficiencies relative to conventional designs while having an axial rotor length 40% less than that of conventionally proportioned radial turbines. Two rotors were designed and tested, one of which incorporated a slightly back swept leading edge with the intention of minimising losses associated with positive incidence. While it was not the focal point of their investigation, Fredmonski et al noted that the back swept rotor attained an efficiency of 88.1 % at its design velocity ratio of U/C = 0.65, which represented an increase of nearly half a percentage point over the purely radial rotor. The authors did not specify the back sweep angle used, but concluded that the improved efficiency may have resulted from the non-radial inlet blading. ainlel
ONE-DIMENSIONAL PERFORMANCE PREDICTION
In order to facilitate 3D numerical investigation of the impact of rotor inlet blade angle, 1D performance prediction modelling was first conducted focused around a duty corresponding to a typical turbocharging application. The ID modelling procedure was based on the method described by Connor and Flaxington [6], and the target design point performance along with the overall turbine dimensions are presented in Table 1. The initial rotor design employed radial inlet blading, which was subsequently varied between values of +30° and -30° while maintaining the other principal turbine dimensions. In this treatment, as in the publication of Meitner and Glassman [3], a back swept inlet blade angle is designated as a positive angle measured from the radial direction, since it aligns with what is conventionally considered to be positive incidence at inlet to a radial turbine.
171
I . Ta ble I DeSI2n . pomt S]!eci'fiIca f Ions an dd'Imenslons Iior 1D analysIs PR (t-s) 2.5 Inlet tip diameter 90mm
Speed
80,000 rev/min
Inlet blade height
20mm
Outlet pressure
101325 Pa
Exducer tip diameter
81 mm
Inlet temperature
873 K
Exducer hub diameter
25mm
Mass flow rate
0.5 kg/s
Exd. RMS blade angle
45°
Rotor blade number
10
Exducer throat area
2740 mm2
Figures 2 and 3 present the predicted efficiencies and flow rates for the five different inlet blade angles plotted against velocity ratio, VIC. The plots are for a constant pressure ratio and, therefore, the UIC value is proportional to rotor speed. Less than optimum UIC values can be considered to represent a transient condition where the turbine is accelerating from a lower speed towards its design speed. In agreement with the results of Weber and Mulloy [2] and Meitner and Glassman [3], it is clear from Figs. 2 and 3 that back swept inlet blades deliver improved efficiency at the lower VIC levels, and as expected, a corresponding increase in flow rate since losses are reduced. From a ID perspective, this is because of better blade alignment with the more tangential relative inlet flow angles that arise at the lower rotor speeds. At around two thirds of the turbine design speed (VIC = 0.46), the 30° back swept rotor delivers an efficiency advantage of2.4% over the radial bladed rotor. While this advantage exists over much of the 'acceleration side' of the D/C characteristic, the 30° back swept rotor suffers a penalty of 1.3% in peak efficiency. The forward swept rotors offer advantages at UIC values above 0.7; however, this is of little relevance to either turbocharger applications or small gas turbine engines. 0.82,-------------------------------, 0.80 0.78
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Figure 2 ID efficiency predictions of varying inlet blade angle at PR = 2.5
172
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0.85
0.90
Figure 3 ID mass flow rate predictions of varying inlet blade angle at PR = 2.5 Although not presented here, a range of pressure ratios between 1.5 and 3.9 were investigated and it was noted that while the overall trends were the same as presented in Figs. 2 and 3, the efficiency benefits of the back swept rotor at low VIC values were greater at lower pressure ratios. DEVELOPMENT AND VALIDATION OF NUMERICAL MODEL Since the initial I D analysis could not account for the real flow structures existing in the rotor inlet region, a 3D numerical investigation was undertaken using the CFX-IO Computational Fluid Dynamics (CFD) software. In order to carry out calibration and benchmarking of the numerical modelling strategy, validation was undertaken using results from an extensive experimental performance investigation of a 99 mm radial turbine with a vaned stator [7]. The construction of the turbine test rig in question and the measurements taken provided confidence that the turbine flow was axisymmetric in nature, and consequently only a single blade passage was modelled. CFX-TurboGrid 1.6, a turbomachinery specific grid generation package, was used to create a single passage mesh of the stator and rotor. The stator and rotor meshes each comprised 400 000 cells. Although this was more than the minimum requirement indicated by the grid independence study, it was considered desirable in order to effectively capture the complex 3D flow patterns within the blade passage. The cell density of the grid was similar to that used by Dunham and Meauze [8]. Both stator and rotor grids were constructed of H-grid cells, with particular attention being directed towards avoiding extreme skew angle values; the minimum skew angle was 25°, which occurred at the rotor trailing edge. There was no endwall clearance in the vaned stator, but the rotor included a blade tip clearance of 0.4 mm, which was modelled using 10 equispaced cells in the spanwise direction. The 'Stage' interface option was employed between the stationary and rotating domains, which takes the circumferentially averaged fluid properties at exit from the stator domain and applies them to the inlet of the rotor domain. Figure 4 shows the stator and rotor grids used, although the cell density has been reduced for the sake of clarity in the illustration.
173
Stator Domain
Rotor Domain
Figure 4 Stator and rotor grids used for the CFD validation study
The k-& turbulence model was used with a blend factor of 1.0 chosen as the advection scheme setting. Setting the blend factor to 1.0 is the equivalent to using second order differencing to calculate the advection terms within the discrete finite volume equations and is deemed to be the most accurate of the available advection scheme settings. The walls were assigned smooth, adiabatic, no slip conditions. The solution was considered to have reached convergence when the maximum values of all of the residuals and global imbalances fell below 5.0x 10-4• Experimental values for total pressure and total temperature at inlet and static pressure at outlet were used as boundary conditions. The flow upstream of the stator was assumed to approach in a radial direction. Figure 5 compares the numerically predicted and the measured turbine performance at a range of pressure ratios across a constant speed line corresponding to 55 000 rev/min corrected to 288K inlet temperature. The predicted efficiencies lie within a band of 0% to 4% below the measured values, while the predicted mass flow rates are within 1.7% and 3.5% above the measured values. With these results, validation had been achieved and a high degree of confidence had been established in the numerical modelling strategy described.
174
0.80
-r------------------------------.-
0.75
~
.. ····x·······x
0.70
1l
:e:LI
0.65
.~ ••. ~ ••.. ~ •...x •.... ~ ..... ·X··· ···.x
0.60
:::
~ E!
Measured efficiency Predicted efficiency - 0 - Measured mass flow rate x Predicted mass flow rate
-{)r-
~
x
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~
+----r--~--_.r---._--,_--_r--_.----+O.22 ~ 2.0
2.2
2.4
2.6
2.8 3.0 Pressure ratio (t-s)
3.2
3.4
3.6
Figure 5 Comparison of numerically predicted efficiency and mass flow rates with experimental data NUMERICAL STUDY OF INLET BLADE ANGLE A 3D radial element blade geometry typical of a turbocharger rotor was created using the CFX-BladeGen software based on the overall dimensions from the ID analysis given in Table 1. This baseline radial rotor geometry was then modified to incorporate blade curvature at inlet to yield two further rotor geometries for 15° and 30° back swept angles. Blade inlet angle was the only parameter adjusted; all other principal dimensions remained unchanged, including the rotor exit blade angle and the exducer throat area. The forward swept blades were not considered for further investigation since the potential aerodynamic advantages were clearly well outside the operating range relevant to turbochargers. Figure 6 shows the rotor blade geometry for the radial and 30° back swept rotors. The blade curvature was limited to a short distance of approximately 5% of the meridional length.
(a) Radial blade (b) 30° back swept blade Figure 6 Comparison of inlet curvature for the radial and 30° back swept blades Numerical models were constructed for each of the three rotor blade configurations in line with the strategy detailed in the preceding section. Two separate sets of operating conditions were modelled: a rotor speed of 80 000 rev/min and PR =
175
2.5 (Ule = 0.60), and a reduced rotor speed of 50 000 rev/min at the same PR = 2.5 (Ule = 0.38). The influence of the inlet blade angle on the efficiency and mass flow rate, as predicted by the numerical model, is shown in Fig. 7. At Ule = 0.6, the efficiencies of the different inlet angles are practically equal. At the off design condition where U/e = 0.38, the 30° back swept blade shows a clear efficiency advantage of 2% over the radial blade. Less significant are the changes in mass flow rates, which follow the trends of changing efficiency, as expected. 0.80.,---------------------.,t;.
0.75
Radial
o
o 15 deg. back swept
-;,
o 30 deg. back swept
60.70
!6.~ 0.65 WJ
0.60
o o
0.64 ~
t;.
0.55
0.62 o
0.60
o
8
t;.
0.40
0.45 0.50 0.55 Velocity ratio (U/C)
0.60
~
to!
~
0.58.g
+------,-----,-----.-----,------.----+ 0.35
C
0.56
~
0.65
Figure 7 Influence of inlet blade angle on efficiency and mass flow rate
The streamline distribution plots presented in Figs. 8 and 9 clearly show the reason for the variation in loss levels with the different blade angles. The streamline plots represent the mid-span location in the blade-to-blade plane; only the region around the blade leading edge has been shown, and not the full blade passage. At the design speed of 80 000 rev/min (Fig. 8) the relative flow is seen to approach the rotor in an almost radial direction, with the radial blade experiencing some flow reversal next to the pressure surface just after the leading edge. At the off-design condition of 50 000 rev/min (Fig. 9), strong positive incidence is evident at rotor inlet and a substantial zone of recirculation results as the flow separates from the suction side of the blade at the leading edge. However, in spite of the inlet curvature extending over only a short length of the blade, the back swept blade is much more tolerant of the positive incidence and, although still present, the extent of the recirculation on the suction side is much less than for the radial blade. The presence and extent of the recirculating flow was responsible for increased losses and the reduced mass flow rate experienced by the radial rotor at off-design conditions. Although further plots at other spanwise locations have not been shown here, it was apparent that the recirculation zone was concentrated in the half of the passage closest to the hub, and was less evident next to the shroud side. Hence, the angle of inlet blade curvature could be varied from hub to shroud for optimum aerodynamic advantage. It is clear that back swept blading at rotor inlet could offer aerodynamic performance gains for turbocharging applications, where the turbine spends a significant
176
part of its duty cycle at lower than optimum speeds, experiencing strong positive incidence.
(a) Radial rotor (b) 30 0 back swept rotor Figure 8 Streamline distribution at turbine inlet for the radial and 300 back swept blades at 80 000 rev/min (UIC = 0.60)
(a) Radial rotor (b) 30 0 back swept rotor Figure 9 Streamline distribution at turbine inlet for the radial and 300 back swept blades at 50 000 rev/min (UIC = 0.38)
FINITE ELEMENT ANALYSIS OF INLET BLADE ANGLE A stress analysis was performed on both the radial blade and 30° back swept blade using the Finite Element Analysis (FEA) package ABAQUS 6.5-1. Due to the rotor's cyclic symmetric geometry only a single periodic section of each rotor was modelled. Both models had a fillet radius of Imm along the blade root and contained a scallop. Material properties ofInconel 718 at 649°C were used. The rotors were analysed at an angular speed of 80 000 revs/min from which the associated blade root stresses where extracted and compared. A mesh seed independence study was performed in order to identity the number of cells required to adequately predict the levels of stress within the
177
model. It was found that using upward of 330,000 Tri cells for a single periodic model did not have a significant impact on the value of stress at various mesh nodes. Figure 10 shows the periodic section of the 30° back swept turbine (a) and periodic mesh (b) with reduced cell density for the purpose of clarity. The maximum blade root stress throughout the inlet region was found to be 26MPa and 522 MPa for the radial and 30° back swept rotors respectively, located on the pressure side of each rotor. This equates to an increase from 2.5% to 51 % of the material yield strength when changing the inlet blade angle from radial to 30° back swept. Although this was seen as quite a large increase in stress it was confined to a relatively small area within the inlet blade region, which could be further reduced with a larger fillet radius along the blade root. Localised yielding was observed in the centre of the hub on the back-face of both the radial and 30° back swept rotors. Figure 11 shows a contour plot of the stress distribution along the blade root at the inlet region of the radial and 30° back swept rotors.
(a) Periodic Section (b) Periodic mesh Figure 10 Periodic section and mesh of the 30° back swept rotor used for the FEA analysis
522MPa
/A=
385 MPa 309 MPa
Y
Y
Y
r
/
/
61 MPa 121 MPa
\"
/ .
26MPa
~.
182MPa
(! ,/ \. ! ;
/
)1
(a) 30° back swept rotor (b) Radial rotor Figure 11 Stress distribution at rotor inlet for the radial and 30° back swept rotors at 80 000 rev/min
178
METAL I CERAMIC NANOCOMPOSITES
Conventionally radial element blades have been used in turbine rotors to avoid additional bending stresses. However, advances in material technologies may provide the aerodynamic designer with freedom to depart from the radial constraints. Ceramic materials have a much lower density than conventional nickel alloys and stainless steels, and would therefore incur lower levels of bending stress in non-radial rotating blades. While much development testing has been conducted with ceramic turbine rotors in the past, they have not been widely employed in commercial products. A primary reason is the low toughness of ceramics, which can give rise to sudden failure and wide variations in rotor life. Recent developments in the use of nano-materials is enhancing the properties of conventional metals and ceramics and yielding new classes of materials called nanocomposites. These composites provide the potential for ceramics with significantly increased toughness levels, or metal alloys with improved high temperature properties. Nanocomposites constitute a novel class of material where a particle having at least one dimension on the nano-scale (10- 9 m) is dispersed in a matrix material, such as a polymer, metal or ceramic. Nanotubes, specifically carbon nanotubes (CNTs) are attracting intense research interest since they were identified by Iijima [9] working at the NEC corporation in Japan in 1991. CNTs have many unique properties, including having a tensile strength and modulus up to 20 times greater than stainless steel, conducting electricity better than copper and having similar thermal conductivity to diamond [10]. These properties combined with their size, shape, large aspect ratio and low density, typically one sixth that of steels, make them ideal candidates for use in novel composite materials. Guo-Dong Zhan and colleagues [11] at the University of California, Davis have reinforced alumina with up to 10 wt.% CNTs and reported a five fold increase in fracture toughness compared to the neat ceramic. Scientists at the NASA Glen Research Centre have developed CNT based coatings for two-phase oxide ceramic eutectics, AI203/Zr02 (Y203) which had enhanced friction and wear properties, both in air and in ultrahigh vacuum. Further research at the centre has also shown that the sintering resistance of certain ceramics can be increased by up to an order of magnitude on addition of CNTs. They reported that the grain size of an alumina ceramic after exposure at 1450 °C was decreased by an order of magnitude on addition of 5wt.% CNTs, indicating the potential for significant increase in the operating temperature of such materials. Bahizsi et al. [12] prepared carbon nanotube reinforced silicon nitride composites with improved bending strength and elastic modulus compared to matrices with added carbon fibre or graphite. Nanotubes made from other elements, such as Boron Nitride have also been synthesized recently. These nanotubes oxidize at higher temperatures than CNTs and their incorporation into ceramics offer higher temperature performance and have better anti-corrosion properties compared to CNT filled ceramics, but they are denser than CNTs. Carbon nanotubes have also been used to reinforce metal matrix composites. Dong et al. [13] reported an increase in fracture toughness, wear resistance and hardness of a CNT reinforced copper composite. Similar improvements were reported by Kuzumaki et al. [14] for a CNT titanium nanocomposite, and Laha et al. [15] having synthesized and characterized a CNT aluminium nanocomposite. Incorporation of CNTs into both ceramics and metals offer exciting developments in materials for turbomachinery applications. It may be possible to produce composites and coatings with a combination of enhanced toughness and hardness, increased thermal
179
conductivity and lower density. The machine-ability and formability of these materials may also be improved. It therefore seems that while past turbine designs have been constrained to use a radial element blade design, nanocomposite materials may provide the aerodynamicist with new freedom when defining blade shapes and angles. CONCLUSIONS
An increase in turbine efficiency, particularly at lower than optimum values of VIC, would result in more torque available for turbocharger acceleration, increasing boost air pressure during engine transients, benefiting engine response and emissions. A review of existing literature demonstrated the potential for enhancing turbine efficiency at low VIC through the use of back swept turbine inlet blades, although published performance data relating to varying inlet blade angle were very limited. A ID analysis for a typical turbocharger turbine indicated an efficiency increase of 2.4% at VIC = 0.46 when the inlet blade angle is back swept 30° from the radial direction. The back swept blade result in a peak efficiency decrement of 1.3% compared with the conventional radial blade. The turbine mass flow rate was found to increase with increasing efficiency. A thorough validation of the numerical modelling strategy was conducted using experimental performance measurements from a similar sized turbine rotor. Two versions of radial turbine geometry were designed with radial and 30° back swept inlet blading. Subsequent numerical modelling showed a 2% increase in efficiency with the back swept blade at VIC = 0.38, but identical efficiency levels at VIC = 0.6. The numerical models revealed a strong recirculation near the suction side of the radial blade at inlet under low VIC, (positive incidence) conditions, which was almost completely eradicated with reduced losses when using the back swept blade angle. The maximum stress throughout the inlet region occurred on the pressure side of the 30° back swept blade and equated to 51 % of the material yield strength. Stress considerations have conventionally constrained turbocharger turbines to adopt a radial element blade design. However, new advances in nanocomposite ceramics and metal materials present the possibility of non-radial blade designs to enhance turbine efficiency at more heavily loaded conditions. ACKNOWLEDGMENTS
The authors would like to acknowledge the technical support of staff at ANSYS Europe in the use of the CFX-IO.O software. Thanks are also due to John Doran for providing turbine measurements for validation purposes. REFERENCES
1. Rodgers, C., & Rochford, K., 2002, Small Turbocharger Turbomachinery, Instn. Mech. Engrs. Turbocharger Conference. C602/003/2002. 2. Mulloy, 1.M. & Weber, H.G., 1982, A Radial Inflow Turbine Impeller for Improved Off-Design Performance, 27th International Gas Turbine Conference and Exhibit. ASME International Gas Turbine Conference, Paper no. 82-GT-101. 3. Meitner, P.L. & Glassman, AJ., 1983, Computer Code for Off-Design Performance Analysis of Radial-Inflow Turbines with Rotor Blade Sweep, NASA TP-2199.
180
4.
5.
6.
7.
8.
9. 10. 11. 12. 13. 14. 15.
Hakeem, 1., 1995, Steady and Unsteady Performance of Mixed Flow Turbines for Automotive Turbochargers, Ph.D. Thesis, Imperial College of Science, Technology and Medicine, London. Fredmonski, A. J., Huber, F. W., Roelke, R.J., & Simonyi, S., 1991, Design and Experimental Evaluation of Compact Radial Inflow Turbines, NASA Lewis Research Centre, ReportAIAA-91-2127. Connor, W. A. and Flaxington, D., 1994, A One-Dimensional Performance Prediction Method for Radial Inflow Turbines, Instn. Mech. Engrs. Turbochargers and Turbocharging Conf., Paper No. C484/041194, pp. 271-282. Doran, W. J., 1999, An Experimental Assessment of the Effects of Shroud Profile on the Performance of a Radial Inflow Turbine, PhD Thesis, School of Mechanical and Manufacturing Engineering, Queen's University of Belfast. Dunham. J., and Meauze, G., 1998, An AGARD Working Group Study of 3D Navier-Stokes Codes Applied to Single Turbomachinery Blade Rows, ASME International Gas Turbine Conference, Paper no. 98-GT-50. S. Iijima, Nature, 354, 7, 56, 1991. J. Li, Y. Lu, Q. Ye, M. Cinke, J. Han, M. Meyyappan Nano Lett., 3, 929, 2003. G-D Zhan, J. D. Kuntz, J. E. Garay and A. K. Mukherjee App!. Phys. Lett. 83, 6, 1228,2003. C. Balazsi, Z. Konya, F. Weber, L. P. Biro, P. Arato Mater. Sci. Eng. C 23, 1133, 2003. S. Dong, J. Tu, X. Zhang Mater Sci. Eng. A 313,83,2001. T. Kuzumaki, K. Miyazawa, H. Ichinose, K. Ito J. Mater. Res. 13,2445, 1999. T. Laha, A. Agarwal, T. McKechnie, S. Seal Mater. Sci. Eng. A 381,249,2004.
181
EXPERIMENTAL STUDY ON THE PERFORMANCE OF A VARIABLE GEOMETRY MIXED FLOW TURBINE FOR AUTOMOTIVE TURBOCHARGER Srithar Rajoo and Ricardo Martinez-Botas
Department of Mechanical Engineering Imperial College London SW7 2AZ Exhibition Road, London SYNOPSIS
This paper investigates a variable geometry (VG) mixed-flow turbine with a novel, purposely designed pivoting nozzle vane ring. The nozzle vane ring was matched to the 3-dimensional aspect of the mixed-flow rotor leading edge. The VG mixed-flow turbine has been evaluated experimentally, in steady and unsteady conditions. The VG turbine shows higher efficiency and swallowing capacity at various vane angle settings compared to an equivalent nozzleless turbine. But the VG turbine unsteady performance was found to deviate substantially from the quasi-steady assumption, compared to a nozzleless turbine. The VG stator with the new unique vane design is expected to further enhance the mixed-flow turbine benefits in term of engine-turbocharger matching and transient performance. NOMENCLATURE
C c Cis
E M R T U W
Absolute flow velocity, mls Vane true chord, m Isentropic velocity, mls Turbine Power, W Mach number Universal gas constant, J/kg.K Temperature, K Rotor velocity, mls Relative flow velocity, mls
b 0
r s m
a
fJ r f//
Vane axial chord, m Nozzle throat width, m Radius, m Nozzle pitch, m Mass flow rate, kg/s Absolute flow angle, deg Relative flow angle, deg Specific heat ratio Tangential lift coefficient
subscript b Rotor blade v Nozzle vane o Total condition 4 Rotor inlet condition 5 Turbine exit condition(atmospheric) INTRODUCTION
Automotive turbochargers almost in their entirety are equipped with radial turbines, due to the efficiency superiority of radial design when compared to other turbine types. But radial turbines are limited in their potential, due to their radial leading edge. A mixedflow turbine differs from a radial turbine in that the leading edge is swept radially downward, as opposed to the zero blade angle of the radial turbine. Research has shown a substantial amount of exhaust gas energy to be available at velocity ratios of less than 183
0.7 - the point of highest energy recovery for a radial turbine. For the past three decades research has been carried out to explore the benefits of the mixed-flow turbine in terms of lower velocity ratio operation and higher swallowing capacity [1,2,3,4], For a nonzero rotor blade angle, the peak turbine efficiency point moves to a higher expansion ratio. Shifting the peak efficiency point to a higher expansion ratio is advantageous in a turbocharger application, which is subjected to pulsating flow from the reciprocating engine, where the greater energy of the flow is contained at high pressures. VGTs have been widely based on radial turbines, particularly in automotive application. Mixed-flow turbines with lower inertial characteristic, coupled with variable geometry operation have the capability of enhancing existing versions of VGTs. Unfortunately, there are not many VGTs currently utilizing mixed-flow turbines, and for the very few which exist [5], the nozzle vane used is adapted from radial turbines, these are invariably straight vanes. Due to the non-radial inlet of the mixedflow turbine, the use of straight vanes is not an optimized option as it creates nonuniform inter-space between the nozzle vane ring and the rotor, hence this work to study an alternative nozzle vane design and its implication on a mixed flow turbine performance. Most available published work on VGTs is still based on steady-state data. The present experimental facility offers the ability for both steady and pulsating flow testing of the VGT, equipped with a mixed flow turbine. It has been shown in the past research that the turbine performance during pulsating flow departs substantially from the quasisteady assumption [3,4]. Thus unsteady-state VGT testing is required in order to better understand the mixed flow turbine characteristics at different nozzle positions, hence the need for an optimized nozzle design for a mixed flow turbine. NOZZLE VANE DESIGN This section describes the design process of the nozzle vane for a mixed flow turbine. The nozzle vane ring was designed based on common axial turbine blade design method and through conformal transformation the designed axial blade was converted to circumferential nozzle ring. As the nozzle ring is preceded by a volute which will provide sufficient swirl to the flow, the nozzle vane is left un-cambered, and effort was concentrated on 3-dimensional variation of the nozzle vane to fit the leading edge of the mixed flow turbine. The nozzle vane profile was designed based on NACA airfoil 0015. The volute used in this study was designed based on a commercial turbocharger (HOLSET H3B), but enlarged for a nozzle vane ring fitting with an inlet area over radius ratio (Air) of 33mm and an exit flow angle of 70°. The nozzle vane ring is aimed for variable geometry operation and the operating range of the vane angles were determined based on the possibility of exploring wider range of turbine performance. The incidence at the rotor inlet is, i=fJ4-f3b (1) The optimum incidence is considered between _20 0 to -30 0 [6]. The rotor velocity, U4 is U4 = C4 sina4 -W4 sinfJ4 (2) Since Cm4 and Wm4 are the same,
W =C4cosa4/ 4 /cosfJ4 Substituting Eq. (3) into (2) will eliminate W4 , U 4 = C4(sina4 -cos a 4 . tan fJ4)
184
(3)
(4)
For the mixed flow rotor used in this study, the blade angle, fit, is 20 0 • By determining the incidence angle, relative blade angle, fJ4 is calculated using Eq. (1). The velocity of the rotor can also be expressed as, U 4 = m·r4, whereby m = 2nN;(;'0
(5)
and according to Euler Equation, the power of the turbine can be approximately deduced as,
E= TilUl (6) By determining the power output or the turbine revolution, U4 is calculated. Consequently, the maximum permitted U4 is determined as the maximum power and turbine rpm of the test-rig are known. With the values of fJ4 and U4 known, using Eq. (4) value of C4 is calculated for different absolute velocity angles, U4. The absolute flow velocity is then normalized and expressed as Mach number, M 4 , C4 C4 _ M 4(7) speed of sound ~ yRTo4 The flow Mach number at different absolute flow angle can be calculated for a range of turbine speeds using Eq (7). By limiting the Mach number to one, the suitable range of inlet flow angle for the power spectrum in the current application was found to be 40 0 to 80 0 • The determined inlet flow angle can be expressed as a function of the nozzle geometry [7], (8) The nozzle throat, 0 and pitch, s are dependant on the number of nozzle vanes. Number of nozzle vanes or indirectly the blade spacing is an important factor in loss contribution and it is determined through Zweifel's criterion, IJI
= 2(s I b)cos 2 a31tana2 -tana31, whereby
b = ccosav
(9)
The optimum pitch/chord (sib) ratio is a compromise between friction losses and good flow guidance. The optimum tangential lift coefficient, IJI is between 0.75-0.85, which will result in best pitch/chord compromise [6]. Since different nozzle exit flow angle will result in different pitch/chord ratios, it was decided to match the optimum pitch/chord ratio at mid-range of the flow angles. Another compromise considered in deciding the number of nozzle vanes and the chord length for a variable geometry system is the suitable nozzle space for capability of pivoting within the range needed. After an iterative analysis as mentioned above, the number of nozzle vanes was chosen to be 15. In order to match the nozzle vane to the mixed flow rotor leading edge, it was decided to apply lean stacking for the whole vane chord. The lean angle was chosen by trial and error and finally determined at 50 0 relative to volute hub surface, which is the same as the mixed flow rotor cone angle. The resulting meridional projection of the assembly (see Figure 1) shows good match between the nozzle vane and the rotor. Nevertheless, the lean stacking and conformal transformation leads to non-uniform alignment between the hub-hub and the shroud-shroud surfaces of the adjacent nozzle vanes. These result in non-uniform suction-pressure surface interaction. To resolve this, the shroud end vane profile was elongated to match the surface interaction at the hub end, and finally producing even interaction. As illustrated in the Figure 1, this results in a small degree of sweep in the nozzle vane trailing edge, but without jeopardizing the mixed flow rotor inlet match.
185
\) l/1j I~ l
I
I I __.JI
.
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Figure 1 Meridional Projection and the Lean Nozzle Vane Assembly
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Figure 2 Turbocharger Test-Rig Schematic Diagram EXPERIMENTAL SETUP
The experimental facility available in Imperial College London is a simulated reciprocating engine test bed for turbocharger testing. The facility has the capability of conducting steady state testing of single and twin entry turbines as well as unsteady testing with simulated engine pulsations, as previously reported [3,4]. Furthermore, the latest instalment of eddy current dynamometer enables turbine testing within larger velocity ratio range [8]. A schematic diagram of the turbine test rig is shown in Figure 2. The test-rig is supplied by Howden screw-type compressors, capable to deliver the necessary compressed air of up to 1.2 kg/s mass flow rate at a maximum pressure of 5 bar (absolute). Air flow is heated to avoid the occurrence of water vapour condensation during the expansion process within the turbine rotor. The air pulse generator was designed to experimentally simulate the exhaust gas pulsation based on the shape of the cut-outs of the chopper plates, for the unsteady flow turbine testing. A variable speed D.C. motor controls the rotating frequency of the chopper plates, hence the frequency of the pulsation. 186
0.9 80% Speed
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ute Is Ci'7
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Po/Ps Figure 3 Turbine Performance Plots at Different Vane Angles: (a) Efficiency versus Velocity Ratio, (b) Swallowing Capacity STEADY-STATE TURBINE PERFORMANCE The newly designed lean vane with the volute and the pivoting mechanism were coupled to a first generation mixed flow turbine developed in Imperial College London [2]. The turbine was tested for different nozzle vane angles and due to the cold flow testing; the flow non-dimensional parameters were matched to the turbine in an actual engine operation. All the necessary flow parameters were measured at the measuring plane (see Figure 2) before the volute inlet, and the turbine speed and torque were measured directly at the dynamometer. Figure 3 shows the turbine efficiency against velocity ratio and the turbine swallowing capacity. The figure shows results for an 80% of the design speed testing which is approximately 48000rpm. The results presented are for different nozzle vane angles and comparison shown for the same mixed-flow turbine tested with a nozzleless volute [4). One significant observation from Figure 3 is the range of velocity ratio in the current results in comparison to the nozzleless result This is due to the new eddy current dynamometer eliminates the surge and choke limits as opposed to conventional turbine testing with compressor coupling. 187
The peak efficiency of the nozzleless turbine is 75% and for the vane angle of 60°, 65° and 70°, the turbine peak efficiency shows improvement, which is 79%, 80% 77% respectively. Meanwhile, for the 40° and 50° vane angles, the peak efficiency drops to 61 % and 68% respectively. This is due to the increase in separation losses in the nozzle as the vanes opens and deviate from optimum incidence condition. Nevertheless, the 50° and 40° vane angles show higher swallowing capacity from the nozzleless turbine with approximately 2% and 8% mass flow parameter improvement respectively between the pressure ratio of 1.5 and 2.0. The swallowing capacity improvement is due to a bigger volute than the nozzleless version and the opening capability of the nozzle vane passage without choking. As for the higher vane angles, the turbine shows capability to achieve higher pressure ratio at lower mass flow parameter, for instance at the vane angle setting of 70°, pressure ratio of 1.7 is achieved with 34% lower mass flow parameter compared to the nozzle less turbine. The change in the nozzle vane setting from 40° to 70°, results in turbine efficiency to increase significantly in the first 25° and then drops slightly, meanwhile the turbine swallowing capacity to reduce slightly in the first 10° and then drops significantly. At 80% design speed, the vane angle setting of 65° shows best performance at peak as well as at higher velocity ratio region. This implies that this setting is suitable for variable geometry operation so as to adapt to the lower incoming mass flow feed. 0.9.,------------------____, •
0.8
~~
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~ 0.6
c .~ 0.5
ffi
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•
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0.7
0.6
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ute ls ~ 6.,-~~---------------____, x10-5
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Po/Ps Figure 4 Turbine Performance Plots with Lean and Straight Vane: (a) Efficiency versus Velocity Ratio, (b) Swallowing Capacity
188
The turbine performance with the new lean nozzle vane was also compared against an equivalent straight vane, which was initially used as baseline design. Figure 4 shows the turbine performance comparison at 80% design speed (48000rpm), with vane angle setting of 65°. As seen in the figure, the turbine efficiency with both the lean and the straight vane is fairly similar at peak condition, even though differences observed at offpeak condition. Due to the leaning, the new vane has higher wetted area than the straight vane. Nevertheless, with similar efficiency the lean nozzle vanes are capable of higher swallowing capacity, due to the larger nozzle area as an effect of leaning the vane. The pressure averaged mass flow parameter of the lean nozzle vane is approximately 10% higher than the straight vane. This indicates the suitability of the lean nozzle vane in turbocharger application, where higher swallowing capacity is often required without sacrificing the turbine efficiency. UNSTEADY TURBINE PERFORMANCE The performance of the turbine with lean nozzle vane was tested under pulsating flow condition for different vane angles. The pulsating frequency of the flow was 40Hz which approximately simulates a 3 cylinder engine running at 1600rpm. Testing was carried out at an average 80% design speed (48000rpm) and flow parameters were measured instantaneously for 1800 cycles. The detailed measuring and analysis procedures were described in Szymko et al. [9]. Figure 5 and Figure 6 illustrate the turbine instantaneous efficiency and swallowing capacity in a complete pulse cycle for 70° and 40° vane angle settings. The instantaneous efficiency is calculated based on isentropic condition measured at the measuring plane (see Figure 2), which is upstream of the volute inlet. Thus, the instantaneous efficiency refers to the stage from the measuring plane to the turbine exit. Also presented in the Figures 5 and 6 are the equivalent quasi-steady curve derived from steady flow testing of the turbine. It is observed that the turbine performance curves exhibit hysteresis loop, but with significant deviation from the quasi-steady condition, in comparison to the nozzleless turbine, as in Figure 7. The deviation is very substantial in the 70° vane angle setting, but reduces as the nozzle vanes open to 40°. During pulsating flow the turbine volute experiences continuous filling and emptying and in the 70° vane angle setting, due to high blockage of the nozzles, the emptying of the volute is delayed before the filling occurs in the consequent pulse,. 1.0
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Figure 5 Turbine Instantaneous Efficiency at 40Hz Pulsating Flow with 70° and 40° Nozzle Vane Angles 189
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Figure 7 Nozzleless Turbine Instantaneous Efficiency (a) and Swallowing Capacity (b) at 40Hz Pulsating Flow This creates a back pressure and eventually mass accumulation in the volute. It is suspected that the slow emptying process of the volute may have resulted in higher mass flow rate measurement, which in reality does not reflect the true instantaneous mass flow rate impacting momentum on the rotor. This eventually leads to the lower efficiency calculation for the 70° vane angle setting. The high mass flow rate reading is evident in the Figure 6 and it is also noticed that the turbine experiences choking at high pressure ratio region in 70° vane angle setting, which further indicate the possibility of mass accumulation upstream of the nozzle vane. Similar choking behaviour was also observed for the 65° and 60° vane angle settings. On the other hand, at 40° vane angle setting, the turbine did not experience choking and loops more closely to the quasisteady curve. Table 1 shows the cycle averaged efficiencies and velocity ratios of the turbine at different nozzle vane angle settings as well as the nozzleless condition. These parameters are averaged using isentropic power average method, which was described by Szymko et al. [9]. The equivalent quasi-steady efficiencies were derived from steady curve at the corresponding velocity ratios. It is noticed at all nozzle vane angle settings that the cycle averaged turbine efficiency are significantly lower than the 190
Table 1 Unsteady Cycle-Isentropic Power Averaged and the Equivalent QuasiSteady Efficiencies U/Cis
Nozzle Vane Settings 40° 50° 60° 65° 70°
0.61 0.61 0.62 0.62 0.60 0.62
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equivalent quasi-steady point and the nozzleless condition. At 70° vane angle, the unsteady cycle efficiency is about 33% lower than the equivalent quasi-steady efficiency and it gradually improves to about 17% at 40° vane angle. Figure 8 shows the turbine efficiency variation plotted for one complete pulse cycle for all the nozzle vane angle settings as well as the nozzleless condition. It can be noticed that the nozzleless turbine exhibit negative efficiency at the beginning and end of the cycle, which is the low pressure ratio region. The negative efficiency was explained by Szymko et a1. [9], as the consequence of turbine rotor impacting momentum on the flow at lower pressure ratio condition. As for the nozzled settings, the turbine did not exhibit negative efficiency at 70°,65° and 60° vane angle settings. This is due to the nozzle constantly providing sufficient flow momentum to the rotor during the pulse cycle. But at 40° and 50° vane angle settings, the turbine gradually exhibit negative efficiency, as the opening of the nozzle reduces the momentum of the flow in low pressure ratio region. It is also noticed that during the first 120° crank angles, where most ofthe isentropic power concentrated in the pulse, nozzleless turbine exhibits better efficiency than the nozzled turbine. For the rest of the cycle, the efficiency of the nozzleless turbine drops in comparison to the nozzled turbine. Another interesting observation is that at 65° vane angle setting, the turbine shows good efficiency towards the end of the cycle, which indicates better energy extraction from the lower pressure region of the pulse. 191
CONCLUSIONS A new nozzle vane ring has been designed to match the 3-dimensional feature of the mixed-flow rotor leading edge. The VG mixed-flow turbine performance was tested at different vane angle settings, under steady and unsteady conditions. The VG mixedflow turbine shows efficiencies at nozzle vane angle of 60°, 65° and 70° higher than the nozzleless turbine for much of the velocity ratio range, with highest peak efficiency of 80% obtained at 65° vane angle. The swallowing capacity of the VG turbine at fully opened condition (40° vane angle) is up to 8% higher than the nozzleless turbine. The VG turbine is capable of achieving high pressure ratio at lower inlet mass flow, which demonstrates its benefit in adapting to off-peak conditions during the turbine operation. Performance comparison with an equivalent straight vane design shows the higher swallowing capacity of the lean vane with similar efficiency at almost all pressure ratio condition. The VG turbine's unsteady performance shows substantial deviation from the equivalent quasi-steady assumption at all the vane angle settings, especially at close nozzle positions (60°, 65°, 70°), as much as 33% at 70° vane angle setting. The observed characteristics of the nozzled turbine under pulsating flow indicate the need to actively control the nozzle position to better adapt to the incoming flow, hence improving energy extraction. REFERENCES
2 3
4
5
6 7
8
9
Baines, N.C., Wallace, F.J. and Whitfield, A. 'Computer Aided Design of MixedFlow Turbines for Turbochargers', Transc. ASME, July 1979 VollOl 440-449. Abidat, M., Chen, H., Baines, N.C., 'Design of a Highly Loaded Mixed Flow Turbine', Proc. Instn. Mech. Engrs., 1992 Vol 206. Arcoumanis, C., Hakeem, 1., Khezzar, L. and Martinez-Botas, R.F. 'Performance of a Mixed Flow Turbocharger Turbine Under Pulsating Flow Conditions', Transc ASME 95-GT-210, 1995. Karamanis, N., Martinez-Botas, R.F. 'Mixed-Flow Turbines for Automotive Turbochargers: Steady and unsteady Performance' IMechE Int. J Engine Research, 2002 Vol 3 No.3. Baets, 1., Bernard, 0., Gamp, T., and Zehnder, M., 'Design and Performance of ABB Turbocharger TPS57 with Variable Turbine Geometry', 6th Int. Conf on Turbochargers and Turbocharging, Proc. of the IMechE, paper C554/017/98, 315325,1998. Japikse, D. and Baines, N.C., Introduction to Turbomachinery, Concept ETI Inc., USA and Oxford University Press, Oxford, 1994. Hiett, G.F. and Johnston, LH., 'Experiments Concerning the Aerodynamic Performance in Inward Radial Flow Turbines', Proc. of the IMechE, 1963 Vol. 178 Part 3I(II) 28-42. Szymko, S., Martinez-Botas, R.F., Pullen, K. R., McGlashan, N.R. and Chen, H., 'A High-Speed, Permanent Magnet Eddy-Current Dynamometer for Turbocharger Research' i h Int. Conf on Turbochargers and Turbocharging, Proc. of the IMechE, paper C602-026, 2002. Szymko, S., Martinez-Botas, R.F. and Pullen, K.R. 'Experimental Evaluation of Turbocharger Turbine Performance under Pulsating Flow Conditions', Proc. of ASME Turbo Expo, GT 2005-68878, 2005.
192
Turbocharger turbine performance under steady and unsteady flow: test bed analysis and correlation criteria Massimo Capobianco and Silvia Marelli Internal Combustion Engines Group (ICEG) Department of Thermal Machines Energy Systems and Transportation (DIMSET) University of Genoa - Italy
ABSTRACT Turbocharging is becoming a key technology for both gasoline and diesel automotive engines. A thorough knowledge of turbine behaviour under steady and unsteady flow conditions is a fundamental requirement for the improvement of engine performance, particularly in transient operation. To this end, a great deal of information can be obtained from investigations developed on dedicated test facilities. This paper presents a new arrangement of the turbocharger test rig operating at the University of Genoa. The results of an extensive experimental programme, focusing on the behaviour of the turbocharger regulating system, on turbine pulsating flow performance and its correlation criteria with steady flow results, are then analysed.
NOMENCLATURE Notations
f n p A K I
M p T
pulse frequency rotational speed pressure waste-gate opening degree ratio between mean pulsating and steady flow performance influence factor mass flow rate power temperature expansion ratio increase
Subscripts 3 4 nd t
M NSF p QSF S T WG
turbine entry turbine exit "non dimensional" (independent of inlet conditions) turbine referred to mass flow rate mean unsteady flow value referred to power average quasi-steady flow value static condition stagnation condition referred to waste-gate valve 193
Definitions and acronyms nnd = nJ-VT3 Mnd = (Mt . -VT3) / P3 et=P3/P4 GDI GP-IB NSF PC QSF VVA
turbine rotational speed factor turbine mass flow factor turbine expansion ratio Gasoline Direct Injection General Purpose Interface Bus Non Steady Flow Personal Computer Quasi Steady Flow Variable Valve Actuation
Note - Turbine performance parameters were defined according to total-to-static conditions in steady flow and static-to-static conditions in pulsating flow operation. INTRODUCTION
Due to the need to reduce CO2 emissions, coupled with increasing fuel prices, a great deal of research effort is currently being expended on the development of high efficiency powertrain concepts. In this regard, turbocharging can both improve automotive diesel engines and substantially enhance gasoline applications. For spark ignition engines, charge boosting is seen as a way of reducing engine displacement at constant rated power (downsizing concept) and its application in conjunction with other technologies, such as direct injection (GDI) (1) and fully flexible valve control systems (VVA) (2), can offer significant reductions in fuel consumption, especially at part load operation. In the meantime, the gasoline engine should be able to maintain its advantage over the diesel engine as regards exhaust pollutants. The application of turbocharging to automotive spark ignition engines has to face various problems related both to the specific operating environment (exhaust gas temperatures) and, principally, to functional aspects (3, 4, 5). There is clearly great interest, therefore, in carrying out dedicated investigations on small turbochargers for gasoline engine applications, focused on different targets such as: i) definition of compressor and turbine characteristics in an extended range; ii) evaluation of unsteady flow turbine performance and development of suitable correlation criteria between steady and pulsating flow operation; iii) analysis of the behaviour of applicable turbine regulating devices; iv) interactions between the turbocharger and the engine intake and exhaust system; v) development of unsteady flow performance prediction procedures. In order to further investigate the above aspects, measurements performed on a components test rig are particularly useful as they allow turbocharger performance to be investigated independently of the engine. In this case, the turbine is usually fed with compressed air which needs to be heated at least to a temperature level that prevents water condensation and outlet freezing problems. To this end, electrical heating is often preferred to a combustion chamber, due to its simplicity and ease of control (6). In any case, compressor and turbine performance characteristics are usually plotted in terms of "dimensionless" or "corrected" parameters regardless of inlet thermodynamic conditions (7). As regards steady flow operation, there are a number of difficulties connected with measuring turbocharger characteristics in a broad range. This is a basic requirement, especially for the turbine that usually operates under unsteady flow conditions and 194
instantaneously experiences expansion ratio levels varying in a substantially wider range than that considered in the maps provided by the turbocharger manufacturer. Extrapolation of turbine curves can generate significant errors, particularly in the field of the lowest expansion ratios. Consequently, if a quasi-steady flow approach is used to predict pulsating performance, a detailed definition of turbine steady flow curves is essential for improving calculation accuracy. Besides, experimental information on the effect of the regulating system setting (waste-gate valve or variable geometry device) on turbine characteristics is a fundamental input for theoretical simulation models and a pre-requisite for developing effective turbocharger control strategies. Test rig investigations are usually performed using the turbocharger compressor as a dynamometer. Since the compressor operating range is limited by chocked flow and surge, specific experimental techniques are required to widen the definition of turbine characteristics, especially as regards tests performed at low inlet air temperatures. To this end, the possibility of controlling the compressor supply pressure in a broad range, thus modulating its power absorption, allows the experimental definition of turbine curves be to considerably extended. If further expansion of measured characteristics for very low levels of the expansion ratio is required, specific devices can be used. For example, an original impulsive system acting on the compressor side of the turbocharger is presented in (8). This allowed the measurement of turbine characteristics near zero mass flow conditions to be significantly extended without modifying the turbocharger rotary assembly. Alternatively, high speed dynamometers can be used to absorb turbine power (9, 10). These devices can considerably extend the measurement of turbine characteristics and in some cases (10) also allow the actual torque produced by the turbocharger bearings to be evaluated. The main drawbacks of turbine dynamometers are complexity of design, restricted rotational speed range (usually below 80,000 rpm) and difficulties in coupling them to different turbines. Additionally, torque values are extremely low, thus requiring very sensitive measuring equipment. The design of a totally flexible turbocharger test rig also has to consider that the turbine usually operates under unsteady flow conditions and sometimes with partial admission. This adds further complexity to the test facility, requiring the introduction of a pulse generator system upstream of the turbine and the use of high speed measuring systems. The possibility of performing tests in unsteady flow conditions is an essential pre-requisite for investigating the effect of the main pulsating flow parameters on turbine performance, developing suitable comparison criteria with steady flow data and analysing the results obtained from theoretical models and quasi-steady flow prediction procedures (11). This paper presents the results of an extensive experimental programme developed on different turbochargers for automotive gasoline engines. The study was performed on the test rig operating at ICEG of the University of Genoa, which allows investigations to be performed both in steady and unsteady flow conditions. The experimental facility was recently upgraded (12) as regards its unsteady measuring capabilities: in the new configuration, tests can be focussed both on analysing the effect of the main pulsating flow parameters on turbine performance and on investigating the engine exhaust subsystem behaviour, including the effect of circuit geometry and of different engine valve actuation strategies. The steady flow characteristics of a small turbocharger turbine for downsized gasoline engines were measured over an extended operating range. The relevant results are presented and discussed in the paper, including the effect of the waste-gate valve 195
setting on turbine mass flow and efficiency. A detailed analysis on the sensitivity of the by-pass regulating device was also performed. Turbine unsteady flow operation was investigated with reference to a different turbocharger. Average pulsating flow performance was analysed, highlighting the influence of the main flow parameters and considering different correlation criteria between mean unsteady results and steady flow data. The turbine feeding circuit pressure diagrams and the results of quasi-steady flow calculations are also presented. EXPERIMENTAL FACILITY
The test rig operating at ICEG is a continuous flow apparatus which allows tests to be performed on individual components and subassemblies of automotive engine intake and exhaust circuits. The facility is particularly suitable for performing investigations on exhaust turbochargers due to the availability of two independent supply lines. In this case, the turbocharger turbine is fed with compressed air at low temperatures (up to 400 K) while the compressor, which acts as a dynamometer, can operate at a controlled pressure level by means of an appropriate regulating system. Turbine performance can also be investigated under unsteady flow conditions by using different pulse generator systems. The test facility layout has been fully described in previous papers (13, 14). Two different air compression stations are available, one of which consists of three electrically driven screw compressors providing a total mass flow rate of approximately 0.6 kg/s at a maximum pressure of 8 bar. Alternatively, a single stage centrifugal compressor with a delivery of up to 2.2 kg/s and a maximum compression ratio of 2.1 can be used. A proper regulating system makes it possible to feed both the turbine and the compressor supply circuit with air at controlled pressure levels. Mass flow rate in each line is evaluated by means of different measuring devices to allow accurate measurement of this parameter also in unsteady flow conditions. The turbine feeding line is fitted with an electrical heater to moderately raise the air temperature in order to avoid condensation and freezing problems during expansion. Two arrangements of the turbine supply circuit are currently available (Figure 1), allowing detailed investigations to be made both in steady and unsteady flow conditions. Pulsating flow at the turbine inlet can be provided by different pulse generator systems connected to the upstream portion of the feeding circuit through a plenum which acts as damping element and flow distributor. The two line configurations are addressed to the investigation of different aspects and can be easily interchanged by modifYing a few connections; in both cases, the upstream and downstream turbine measurement stations remain the same. The first circuit layout (referred to as arrangement A, Figure la) was extensively used during previous investigations developed at ICEG (8, 13, 14, 15, 16, 17) and was designed to perform parametric studies on the effect of the main unsteady flow parameters on turbine performance. Tests on single and two-entry components can be developed, independently controlling thermodynamic parameters at each entry. In both feed branches, pulsating flow is generated by a diametral slot rotating valve (15). The main pressure pulse parameters (amplitude and mean value) at each turbine entry can be controlled by correctly mixing two flow components (steady and pulsating one) in a Yjunction and adjusting upstream plenum pressure. The flow area diagram of each pulse generator can be changed by replacing a few stator and rotor parts, thus affecting the pulse shape. Dedicated flow control valves allow unequal admission conditions to be reproduced in the event of two-entry devices. A variable rotational speed electrical 196
a)
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Figure 1 Arrangements of the turbine supply circuit on the ICEG test rig
motor allows pulse frequency to be adjusted within the typical range of automotive engine intake and exhaust systems (10 - 200 Hz). For two-entry components, the pulse phase angle can be changed in steps of 1110 of the period. A second configuration of the turbine feeding circuit (arrangement B, Figure Ib) was recently set up in order to more precisely reproduce turbocharger unsteady flow operation when matched to an automotive engine. Furthermore, experimental investigations can be extended to a subsystem level, including the effect of the exhaust circuit geometry and different valve actuation strategies. In this alternative layout of the turbine supply circuit, heated compressed air at the damping plenum outlet enters a flow distributor, designed to reproduce the reference engine cylinder block, on top of which a motor-driven cylinder head is connected. The opening of the engine valves determines unsteady flow in the exhaust circuit, the geometry of which can be easily changed to reproduce various manifold configurations. The effect of different valve actuation strategies on turbine performance can also be investigated since the cylinder head can be fitted with a fully flexible valve actuation system (VVA) which allows any valve opening profile to be reproduced, including cylinder deactivation. In a next stage of the investigation, a further upgrade of the turbine supply circuit will be developed, including butterfly regulating valves in each duct of the flow distributor, with a view to reproducing engine transient flow operation. Turbocharger and cylinder head lubrication is provided by a dedicated circuit, based on two independent lines where oil temperature and pressure can be controlled and kept constant at set levels by means of suitable regulating systems. The oil mass flow rate and lubricant inlet and outlet temperatures are measured in order to estimate turbocharger mechanical losses. The test rig is fitted with a PC-controlled automatic data acquisition system. Average and instantaneous wall static pressures are evaluated by high frequency response straingauge transducers, while temperatures are measured by platinum resistance thermometers and turbocharger rotational speed and pulse frequency by inductive probes. Mean turbine and compressor mass flow rate are estimated by a laminar flow meter and a sharp edged orifice respectively. Transducer signals are managed and processed by various instruments connected to a GP-IB. In the event of unsteady flow tests, pressure transducers are mounted near the duct walls to avoid inaccuracies deriving from the damping effect of the signal connecting lines. High-speed sampling devices are used to simultaneously acquire different 197
pressure signals in transient conditions. Interactive procedures in LabVIEW® environment were set up to control the acquisition process and calculate turbocharger performance parameters both in steady and pulsating flow operation. Additionally, specific analytical tools were developed to allow post-processing of instantaneous pressure signals. DEFINITION OF TURBINE STEADY FLOW CURVES As mentioned above, the ICEG test facility allows an extended definition of components steady flow characteristics to be made. This aspect is particularly important in the case of turbocharger compressors and turbines which are generally simulated within engine global models using steady flow curves. Referring to a small turbocharger matched to a downsized gasoline engine, Figure 2 shows the turbine steady flow characteristics measured on the ICEG test rig. The
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100 I turbocharger was fitted with a ..--~ ._-,_.single entry nozzleless radial ,/ --flow turbine (rotor diameter / ----38.4 mm) and a waste-gate / valve as a regulating system. f 60 Turbine constant speed j '>------characteristics are reported in rI I terms of mass flow factor and ~=4000rpm I-i Fr: .JK I-overall efficiency (defined as I El,IS = 1.32 turbine total-to-static isentropic 1-- -. I efficiency multiplied by o 40 60 80 100 20 o turbocharger mechanical Wast"1Jate opening degree [%] efficiency) vs. total-to-static Figure 3 Waste-gate mass flow sensitivity expansion ratio. The investigation was extended to several turbine speed factor levels, ranging from 2000 to 7000 rpm/viK. Different wastegate openings (A) were considered for each turbine speed, being this parameter related to the linear displacement of the by-pass valve push rod and defined as a percentage of total rod displacement, ranging from 0 to 100 percent when varying the waste-gate flow area from zero to maximum. Four different waste-gate settings were selected (corresponding to A equal to 0, 5, 20 and 100 percent) on the basis of a preliminary investigation on the effect of the by-pass valve opening on turbine mass flow rate. The extension of the measured characteristics achieved by controlling the compressor supply pressure while working at almost constant turbine inlet air temperatures (about 400 K) can be clearly seen. The operating range explored proved to be considerably wider than that considered in the maps provided by the turbocharger manufacturer and allowed a satisfactory definition of steady flow curves to be achieved for application within simulation models. A substantial increase in turbine mass flow rate (approximately 50-60 percent) was observed when the waste-gate valve was fully opened. However, a detailed analysis developed at constant turbine rotational speed (n/viTT3 == 4000 rpm/vlK) and expansion ratio (StTS = 1.32) confirmed (14) a substantially higher sensitivity (defined as the ratio of change induced in the output, i.e., mass flow rate, to the related input change, i.e., push rod position) of this turbocharger regulating system in the field of lower by-pass settings. This aspect is highlighted in Figure 3, where the mass flow rate increase determined by the waste-gate opening (~WG) is plotted as a function of its opening degree (A). ~
.
-----
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It is clear that over 95 percent of the turbine swallowing rise was achieved at a waste-gate setting of 50 percent; it is also interesting to note that a by-pass aperture of 5 percent (corresponding to a rod displacement of 1.5 mm) determined a mass flow increase of approximately 40 percent of the total range. This behaviour is not unexpected and is mainly related to the geometric characteristics of the regulating device and the relationship between the connecting rod displacement and the related change of the by-pass flow area. All the same, this aspect should be carefully considered when designing the relative actuator system and when defining correct control
199
strategies, also taking into account the inaccuracies induced by mechanical clearances and hysteresis phenomena. When opening the waste-gate valve, overall turbine efficiency (defined with reference to the ideal work related to an isentropic expansion of the whole mass flow from stagnation conditions at the housing entry to outlet static pressure) decreased definitely (Figure 2). This result is mainly related to thermodynamic losses connected with the wasted enthalpy of the by-passed working fluid, even though a contribution to the efficiency drop may have been made by the flow pattern change in the turbine rotor induced by the by-pass opening, affecting the relative energy losses (15). UNSTEADY FLOW PERFORMANCE AND CORRELATION CRITERIA
In automotive applications, the turbocharger turbine usually operates under unsteady flow conditions and it is extremely difficult to evaluate pulsating performance since several parameters can affect measured data. As mentioned above, the ICEG test facility allows tests to be performed in unsteady flow conditions using different arrangements in order to highlight both the influence of the main pulsating flow characteristics on turbine performance and the behaviour of the entire exhaust circuit. This section presents the results of an extensive test programme developed on a current production turbocharger matched for application to a 4-cylinder 2-litre gasoline engine (Garrett GT2052 ELS). Unsteady flow measurements, performed by using both the available test rig pulsating flow configurations (arrangement A and B) were extended to different turbine speed factor levels and to a broad pulse frequency range (from 33.3 to 200 Hz). Since no experimental information was available on instantaneous mass flow rate at the turbine inlet under pulsating flow conditions, the analysis referred to static thermodynamic quantities. A detailed examination of pressure signals evaluated with the two available test rig configurations, including the effect of the waste-gate opening, is reported in (12). Referring to circuit arrangement B, Figure 4 shows an example of measured pressure diagrams in different sections ofthe turbine feeding line in a typical operating condition corresponding to intermediate engine speed. Pressure signals refer to one of the manifold pipes, to the exhaust manifold mixing volume and to the turbine inlet and outlet sections respectively. Exhaust manifold geometry clearly plays a fundamental role in wave propagation phenomena. In each pipe, the pressure pulse associated with the cylinder valve opening is evident, but pressure oscillations at higher frequencies can also be detected, with amplitudes that are often comparable with the main pulse. As widely reported in open literature (6, 18), the junction of four manifold branches in one mixing volume causes a substantial reduction in downstream flow unsteadiness, resulting in a small pulse amplitude at the turbine inlet and an even lower pressure oscillation at the turbine outlet. It is confirmed that the design of the feeding circuit significantly affects flow unsteadiness at the turbine inlet and consequently the available specific energy and turbine conversion efficiency (19). In this regard, different manifold geometries will be considered at a later stage of the investigation, the behaviour of which will be also evaluated on the basis of calculations performed through wave action models. The analysis of average turbine performance under unsteady flow conditions requires the definition of suitable methods of comparison that can highlight the effect of the main pulsating flow parameters and the shifts from steady flow results. To this end, several approaches have been proposed and discussed by various authors (10, 11, 19, 200
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Figure 4 Measured pressure diagrams in the turbine circuit (arrangement B)
20, 21), often conditioned by available experimental information. In any case, performance comparison developed at the same turbine average expansion ratio and rotational speed is an easy and immediate solution since its application requires just inlet and outlet pressure diagrams to be measured in order to calculate the turbine average expansion ratio. This procedure was used at ICEG during previous investigations (8, 14) referring to the K ratios between mean pulsating turbine performance (mass flow, torque or power) and the corresponding steady flow values. Using this approach, KM and Kp factors were calculated for different operating conditions in order to highlight the effect of the main pulsating flow parameters on average turbine performance. These results make reference to circuit arrangement A. Measured unsteady mass flow and power were generally different from steady flow values at the same average expansion ratio and rotational speed, with absolute shifts of up to 20 percent. These differences were lower when the waste-gate valve was opened, ---'!..-=4000 rpm
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Figure 5 Turbine inlet pressure waves at different pulse frequencies (arrangement A) 201
thus suggesting that flow unsteadiness was reduced on the by-passed fluid portion. As in the case of previous investigations performed on fixed and variable geometry turbines (8, 14), both unsteady mass flow and power were found to depend on pulse frequency though the actual magnitude of the effect was different at each frequency level. This result is probably related to the significant modifications in inlet pulse shape and amplitude which were observed when changing the pUlsating flow frequency (Figure 5), due to the wave action in the turbine feeding circuit (6,8, 14). At constant pulse frequency, a general trend to lower KM values and higher Kp levels was found when the average expansion ratio was increased. Figure 6 shows the results ; : 1.00 r-:~[~~[~~L~~L~;;~~~_·_··-"-r_-"------L---,j 0.98 +-< o referred to a pulse frequency of 1;~ 0.96 t:::::'~;t::L2;:~~T~~~~d 0.94 66.67 Hz, for which both ~ 0.92 average unsteady mass flow rate ~ 0.90 +--+-~--.- + and power were lower than the :::;: 0.88 +---'------1---'--+--1--. 1.4 1.6 1.7 1.8 1.5 1.3 corresponding steady flow Average expansion ratio values in the explored range. Even if the absolute values of KM and Kp proved to be affected by the selected pulse frequency level, the trend shown in Figure 6 was generally confirmed and 1.3 1.4 1.5 1.6 1.7 1.8 seems related to the Average expansion ratio corresponding change in inlet pulse amplitude when varying the average expansion ratio Figure 6 Comparison between average pulsating (Figure 7). The resulting link and steady turbine performance between inlet pulse amplitude and average turbine performance 1; 1.3 confirmed, at least as regards ..----i .!!. I mass flow rate, the conclusions -~ 1.1 of previous investigations (8, 14) ~--~ to.9 and the hypothesis of turbine ~ " ~ ~ quasi-steady flow behaviour in ] . 0.7 A=O I 1=66.67 Hz unsteady conditions, related to 4(0) t 1...----- /1 I ~ 0.5 the variable slope of constant 1.3 1.4 1.5 1.7 1.6 1.8 speed steady flow curves (8). Average expansion ratio Figure 6 shows that, Figure 7 Correlation between inlet pulse compared with a slight reduction amplitude and average expansion ratio in the turbine mass flow factor, power parameter substantially :;: 1.4 i ... o 1.3 I· grows when the expansion ratio r::== f 1.2 is increased (and consequently 1----' -'< i 5- 1.1 the inlet pulse amplitude). This ---i u 1.0 <;:: ~=4000 rpm /=66.67 liz result suggests a remarkable A=O ~ Jf, JK "1 0.9 influence of flow unsteadiness, Vl 0.8 0.5 0.6 0.7 1.0 1.1 1.2 0.9 0.' usually related to the pulse Inlet pulse amplitude [bar1 amplitude or to the ratio between Figure 8 Effect or inlet pulse amplitude on turbine this parameter and the average specific work ratio pressure level (19), on turbine ~
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specific work. In order to highlight this effect, turbine power levels under steady and pulsating flow conditions were compared for the same average mass flow rate. Figure 8 shows the relative result with reference to the operating conditions of Figure 6 and 7: in this case, the ratio K'p between mean pUlsating and steady turbine power was plotted against inlet pulse amplitude. Assuming that the coefficient K'p may be representative of the ratio between turbine specific work in the two flow situations, its behaviour is the outcome of the combined effect of the inlet energy increase for higher flow unsteadiness (i.e., pulse amplitude) (19) and the deterioration of turbine energy conversion efficiency in unsteady flow conditions (7). In the specific operating situation shown in Figure 8, the reduction of turbine efficiency seems to prevail over the inlet energy rise for the lower considered pulse amplitudes; on the contrary, for higher flow unsteadiness levels, K'p was always above unit, suggesting a prevalent effect of the available energy at the turbine inlet. The assumption of instantaneous steady flow behaviour of the turbine operating under pulsating flow conditions is often used to predict unsteady flow performance, particularly in the case of small automotive turbocharging units. Using this approach, instantaneous quasi-steady flow (QSF) turbine performance was calculated starting from inlet and outlet pressure diagrams and referring to steady flow characteristics, according to the procedure described in (11). Average calculated mass flow (MQSF) and power (P QSF ) were compared with the mean measured values (MNSF e PNSF) using the relevant influence factors 1M and Ip: - M NSF I MM QSF I p-- PNSF PQSF
Three different weighting techniques of instantaneous QSF levels (8) were applied: in addition to the mean arithmetic values (M'QSF and P'QSF) (11), average turbine performance was calculated as weighted average levels over the pulse period, assuming the instantaneous expansion ratio as weight (M"QSF and P"QSF) or on the basis of a suitable "mean equivalent expansion ratio", related to the pulse characteristics 1.3 1.4 1.5 1.6 1.7 1.8 Average expansion ratio through the instantaneous mass flow level (8) (M"'QSF and P"'QSF). In Figure 9, the estimated influence factors 1M and Ip are plotted against average turbine expansion ratio for the same operating condition considered in 1.3 1.4 1.5 1.6 1.7 1.8 previous comparisons. The Average expansion ratio results provided by the different calculation procedures proved to be almost unaffected by the Figure 9 Comparison of measured and QSF expansion ratio level (and then by calculated turbine performance 203
the inlet pulse amplitude) as regards turbine mass flow rate, while a noticeable influence of this parameter was found on Ip factors, with absolute values increasing with the expansion ratio. The best results were achieved using the traditional averaging procedure (I') and the weighted-average method (I"), while the process based on the "mean equivalent expansion ratio" produced the greatest deviations from mean experimental performance, overestimating measured values. In the case of mass flow rate, the best procedures gave satisfactory results (with shifts generally lower than 10 percent) confirming a quasi-steady flow behaviour of small turbocharger turbines operating under pulsating flow conditions referring to this parameter (8, 9). The approximation provided by QSF calculations as regards turbine power was less satisfactory: deviations from the average experimental levels resulted higher (even more than 20 percent), depending on the mean expansion ratio. The calculated influence factors also proved to be affected by the pulse frequency but no functional trend with respect to this parameter was found, probably due to the observed modifications of wave shapes. CONCLUSIONS AND FURTHER STEPS
The components test facility operating at ICEG is particularly suitable for developing studies on automotive turbochargers under steady and unsteady flow conditions. The paper presents a recent upgrading of the experimental apparatus: two different arrangements of the supply circuit allow the effect of the main pulsating flow parameters on turbine performance to be analysed and real engine exhaust subsystem behaviour, including the effect of the circuit geometry and of different engine valve actuation strategies, to be investigated. By means of specific experimental techniques, it is possible to defme turbine steady flow characteristics in an extended range while operating at relatively low air temperatures without removing the turbocharger compressor (used as a dynamometer). An example of the definition of turbine steady flow curves has been presented, referring to a small turbo charging unit for gasoline applications, for which mass flow and efficiency characteristics were measured in a broad range, also taking into account the effect of the waste-gate valve aperture and its sensitivity to the driving system position. Turbine pUlsating flow performance was investigated on the basis of various aspects: in an initial step, the effect of the exhaust manifold geometry on pressure diagrams measured in different sections of the turbine feeding circuit was discussed. A comparison between average turbine unsteady mass flow rate and power and the corresponding steady flow values was then performed, taking different correlation criteria into account. Measured unsteady parameters, compared at the same mean expansion ratio and rotational speed, generally proved to be different from steady flow ones. However, no functional link was found between measured shifts and the pulse frequency level, probably due to the associated modifications of wave shapes. On the contrary, a general trend towards a lower unsteady mass flow rate and higher power was found at constant pulse frequency when increasing the average expansion ratio, which proved to be directly related to the inlet pulse amplitude. Using a different correlation criteria, an increase in actual turbine specific work was found for higher inlet pulse amplitudes. Further work on this subject will be performed at ICEG in order to develop comparison procedures based on the flow energy content at the turbine inlet which can take into account the modifications of pulse shape occurring when unsteady flow frequency is varied, due to the wave action in the circuit pipes.
204
Quasi-steady flow calculations of turbine perfonnance were also perfonned, according to different instantaneous level weighting procedures: the relative results are presented and discussed in the paper. ACKNOWLEDGEMENTS
This work was partly developed with the financial support of the European Commission within the EU Integrated Project NICE - New Integrated Combustion System for Future Passenger Car Engines (6th Framework Programme, Priority 6.2, Contract TIP3-CT2004-506201). The authors would like to thank Mr Fabio Polidori for his assistance in the definition of turbine steady flow perfonnance. REFERENCES
2
3 4 5
6
7 8
9 10
11
12
13
B Lecointe and G Monnier, 'Downsizing a gasoline engine using turbo charging with direct injection', SAE paper 2003-01-0542. H Richter, D Schwarzenthal and L Spiegel, 'Variable valve actuation - key technology for high specific power output and low exhaust emissions', International Conference Spark Ignition Engine: the C02 Challenge, paper 02A501O, Venice, ATA, 2002. M Wirth, U Mayerhofer, W F Piock and G K Fraidl, 'Turbocharging the DI gasoline engine', SAE paper 2000-01-0251. T Lake, J Strokes, R Muphy and R Osborne, 'Turbocharging concepts for downsized DI gasoline engines', SAE paper 2004-01-0036. D Petitjean, L Bernardini, C Middlemass and S M Shahed, 'Advanced gasoline engine turbocharging technology for fuel economy improvements', SAE paper 2004-01-0988. D E Winterbone and R J Pearson, Design Techniques for Engine Manifolds: Wave Action Methods for IC Engines, London, Professional Engineering Publishing Ltd, 1999. N Watson and M S Janota, Turbocharging the Internal Combustion Engine, London, McMillan Press, 1982. M Capobianco and A Gambarotta, 'Unsteady flow perfonnance of turbocharger radial turbines', 4th International Conference Turbocharging and Turbochargers, Paper C405117, London, Instn Mech Engrs, 1990. A Dale and N Watson, 'Vaneless radial flow turbine perfonnance', 3rd International Conference Turbocharging and Turbochargers, Paper CI1O/86, London, Instn Mech Engrs, 1986. DE Winterbone, B Nikpour and G I Alexander, 'Measurement of the perfonnance of a radial inflow turbine in steady and unsteady flow', 4th International Conference Turbocharging and Turbochargers, Paper C405115, London, Instn Mech Engrs, 1990. R S Benson, 'Non steady flow in a turbocharger nozzleless radial gas turbine', SAE paper 740739. M Capobianco and S Marelli, 'Transient perfonnance of automotive turbochargers: test facility and preliminary experimental analysis', 7th International Conference Internal Combustion Engines (ICE 2005), Naples, SAE paper 2005-24-66. M Capobianco, A Gambarotta and G Cipolla, 'Influence of the pulsating flow operation on the turbine characteristics of a small internal combustion engine 205
14
15
16
17
18
19 20
21
turbocharger', International Conference The Small Internal Combustion Engine, paper C3 72/0 19, London, Inst Mech Engrs, 1989. M Capobianco and A Gambarotta, 'Variable geometry and waste-gated automotive turbochargers: measurements and comparison of turbine performance', ASME Transactions, Journal of Engineering for Gas Turbine and Power, 1992114,553560. M Capobianco, A Gambarotta and G Cipolla, 'Effect of inlet pulsating pressure characteristics on turbine performance of an automotive wastegated turbocharger', SAE Paper 900359. M Capobianco and A Gambarotta, 'Performance of a twin-entry automotive turbocharger turbine', Energy-sources Technology Conference and Exhibition, paper 93-ICE-2, Houston, ASME, 1993. M Capobianco, A Gambarotta and P Silvestri, 'Effects of volute configuration and flow control system on the performance of automotive turbocharger turbines', Internal Combustion Engine Division Spring Technical Conference, paper 95-ICE11, Marietta, ASME, 1995. R S Benson, The Thermodynamics and Gas Dynamics of Internal Combustion Engines (Edited by J H Horlock and D E Winterbone), Oxford, Clarendon Press, 1982. K Zinner, Supercharging of Internal Combustion Engines, Berlin, Springer Verlag, 1978. H Kosuge, N Yamanaka, I Ariga and I Watanabe, 'Performance of radial flow turbines under pulsating flow conditions', ASME Transactions, Journal of Engineeringfor Gas Turbine and Power, 19769853-59. D E Winterbone, B Nikpour and H Frost, 'A contribution to the understanding of turbocharger turbine performance in pulsating flow', International Conference Internal Combustion Engine Research, paper C433/011, London, Inst Mech Engrs, 1991.
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Flexible turbocharger turbine test rig MONA VI Dietmar Filsinger, Gerhard Fitzky, Bent Phillipsen ABB Turbo Systems Ltd, Baden, Switzerland ABSTRACT
In recent years the significant increase in the power density of internal combustion engines has been supported by the simultaneous progress in turbocharger development The driving forces for these developments were end-users' economical and operational considerations as well as future turbo charging requirements of new engine applications, e,g. emission limits. Rig testing is not only mandatory in the design and development phase of new turbocharger turbines, but also plays an important role in helping turbomachinery designers continue to improve their design tools and methods with accurate validation data. The data provided by rig testing- both overall performance data and internal datais required to develop and enhance the design tools that are used, including computational fluid dynamics (CFD). The turbocharger turbine test rig called MONA VI is described in the present paper. The article gives an overview of the rig installation and its integration into the infrastructure at ABB Turbo Systems Ltd. The configuration of the test rig to meet the requirements during the development process of a turbocharger turbine, considering not only time and cost but also flexibility, is highlighted. INTRODUCTION
The significant increase in the power density of internal combustion engines in recent years has been supported by the simultaneous progress in turbocharger development The driving forces for these developments were end-users' economical and operational considerations as well as future turbocharging requirements of new engine applications. Strict emission limits for NOx, particles, etc. require an overall optimization of the combustion processes and the turbocharging system. Overall turbocharger efficiency needs to be increased for improved part load behaviour and a simultaneous limitation of the thermal load of the engine. Higher compressor pressure ratios are in demand if the effective cylinder power is to be increased. A compact design is inevitable to increase the overall power density of the engine, while maintaining turbocharger weight limitations. This leads to the requirement for increased flow capacity of the components. These demands require the introduction of new components, innovative technology, and advanced design features (8). Time to market is the critical factor for design cycles of new components. In this context computational fluid dynamics (CFD) offers great possibilities. CFD allows the improvement of the design optimization cycle and its competent use can lead to higher performance designs, but CFD still has its limits. The accurate and reliable prediction of loss is difficult and influenced by effects that typically cannot be accurately modelled. Some examples of these effects are geometrical effects, such as surface roughness and flow path steps, uncertainty of the detailed boundary conditions and limitations of the CFD codes for the prediction of complicated flow phenomena, such as secondary flow effects, tip clearance flows and flow separation in diffusers. These limitations originate from limitations of computer resources to resolve the mentioned effects, averaging of
207
unsteady effects as well as limitations due to the modelling of the turbulent flow. Because of this, CFD modelling alone may not be adequate to ensure the performance and mechanical objectives oftoday's advanced turbomachinery designs. This leads to the conclusion that turbomachinery rig testing is and still will be a major activity. Rig testing is often the only way to evaluate the impact of most "realworld" effects and to accurately determine the component's performance. This statement is further emphasized when working outside the designer's normal design space, which is the case when the push for more highly loaded turbine stages is given. Then design optimization through purely analytical methods can be risky. To lessen or eliminate risk within these new design envelopes, it is essential to both calibrate design tools which are being used outside of their calibration range and simultaneously test the turbomachinery operating in these new envelopes. This is why both sophisticated testing facilities and extensive testing experience are needed to develop many of today's high performance turbomachinery components and systems. It is clearly unacceptable to learn that design and performance objectives have not been met after its first run in the actual application. Moreover, rig testing plays an important role in helping today's advanced turbomachinery designer continue to improve their design tools and methods with accurate validation data. The data provided by rig testing- both overall performance data and internal data- is required to develop and enhance the design tools being used, including CFD. Similar argumentation for rig testing can be found elsewhere, e.g. Hinch (7) gives a nice pleading for experimental rig testing. Consequently, it is essential for a turbocharger manufacturer to have access to a reliable, flexible, and- to keep development cost at a minimum- cost effective test bed adjusted to the specific requirements of turbochargers. ABB Turbo Systems Ltd. takes advantage of the turbocharger test centre CHARLES. It consists of several turbocharger test stands that are used to perform full-scale tests up to the largest turbocharger size. For "isolated" and, therefore, flexible investigations on components, a compressor rig and the turbine test rig MONA VI are integrated into this infrastructure. The specific requirements for the turbine test bed result from the on-engine operating conditions. OPERATING CONDITIONS OF TURBOCHARGER TURBINES Large turbochargers are installed on a variety of applications with different demands. If full load efficiency is critical for a diesel engine application a constant pressure charging system would be advantageous, since in that case the turbocharger reaches its highest efficiency. The turbine is fed with exhaust gas in a way that the pressure pulses are minimized and therefore the pressure in front of the turbine is nearly constant. This can be achieved by an exhaust system, which consists of large volumes. Thus, the exhaust gas flow coming from the single cylinders can be equalized. Figure 1 (left) illustrates the principle assembly of such an exhaust system, which is mainly in use on two-stroke diesel engines. If superior part load and transient behaviour is required, pulse charging systems will be favoured for four-stroke engines. Other charging systems can achieve similar beneticial part load behaviour only with substantial additional expenses for support equipment (2). Typical examples for such applications of four-stroke engines are main drives on ships with fixed pitch propellers. Pulse systems convert the exhaust energy more efficiently into turbine shaft work. The pressure pulses downstream of the outlet valve of the cylinders are preserved on their way through the exhaust pipes to the
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turbine inlet. On the other hand, the pressure pulses cause strongly variable inlet conditions in terms of turbine inlet pressure and temperature. The admission of the turbine is far from being steady and the turbocharger turbine has to operate with unsteady flow because of the intermittent cyclic operation of the diesel engine. This leads to considerable unsteady aerodynamic loads on the turbine, which have to be accounted for. The unsteady aerodynamic load can be expressed by the velocity ratio u/Cis, which gives the relation between the provided energy of the exhaust gas and the rotational speed of the turbine. Because of the inertia of the rotor, the speed of the turbine can be assumed to be constant, especially for large turbochargers, but the exhaust gas energy provided by the engine varies with time. Therefore, the velocity ratio u/Cis is varying over an engine cycle. Nevertheless, even for these operating conditions the highest possible efficiency is required. Moreover, in the case of a pulse charging system the turbine is provided with multiple entry gas inlets to separate the pulses. This leads to aerodynamic conditions that vary over the circumference at the same instant of time. As an example of such a charging system, Figure 1 (right) shows a pulse charging exhaust system, whereas the turbine is equipped with a two-entry gas inlet. cylinder
cylinder
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~ 1CH~
Figure 1: Principle assembly of an exhaust pipe system for constant pressure charging (left) and for pulse charging (right).
All these considerations lead to the requirements of a test rig for turbocharger turbines. It should be possible to experimentally characterize turbines for a wide range of inlet pressure and turbine speed. Measurement of efficiency should be possible for full load as well as for part load conditions. A variation of the velocity ratio should be possible over a wide range. Additionally, a simple, flexible, and therefore, costeffective installation of different types of turbines with their appropriate housings- inlet and exit- should be possible. TURBOCHARGER TURBINE TEST RIG MONA VI
Turbine test rigs are technically extensive and expensive to operate. They require skilled applied knowledge to be economically feasible and useful tools in the fast moving environment ofturbine design. A flexible modular concept of the test facility will offer a possible solution. Through a continuous development, evolution and improvement the rig layout in its 6th generation (MONA VI) is described. Figure 2 shows the relevant independent modules of the rig consisting of the test turbine with its peripheral components such as air source, air conditioner, pipe system and brake. The advantage of the modular concept is the flexible adaptation of each sub system to the varying requirements of the different test turbines. In the case of MONA, the test turbine will be fed with air from several parallelworking blowers. With a system of a cooler and an electric pre-heater the pressurized air's temperature is adjusted to 150 degree Celsius. This unusual low turbine inlet temperature requires the special consideration of the Mach similarity in test runs. But the low temperature has advantages in applying measurement techniques and in the choice of turbine part material.
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Controlling Sy_ Auxiliary Systems (oil, warer and electric systems)
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Figure 2 Sketch of the modular concept of the turbine test rig MONA The pipe system includes a duct-system, which provides a well-defined flow field at the turbine inlet and an emergency shut down valve. The test turbine is equipped with a water brake for power consumption and speed control. The turbine and water brake are coupled via a planetary gear. The brake and gear are mounted on a frame, which is arranged like a pendulum. This feature allows measuring the brake momentum with a swing arm. Additionally, the described rig components need auxiliary systems such as a conditioning system for lubrication oil, cooling water, brake water and an electric distributor and transformer. All rig components need a controlling system which helps to adjust the desired operating point of the turbine and which shuts down the machine in case of emergency. The primary purpose of the turbine test rig is to completely determine the performance of a given turbine configuration. Dimensional analysis shows, that the nine relevant parameters characterizing the thermodynamic performance may be reduced to three relations between five non-dimensional quantities. These nine parameters are: turbine efficiency, mass flow rate, inlet and outlet total pressures and temperatures, rotational speed, linear dimension as well as the gas properties: ratio of specific heats, gas constant and absolute viscosity. For operation with a given gas and neglect of the influence of size and the ratio of density to viscosity this is further reduced to three relations depending on two non-dimensional quantities. As efficiency is known from thermodynamic relations, when pressure ratio and relative temperature drop is known, the turbine characteristics can be completely defined by two relations. It is convenient to formulate these as relations between turbine efficiency and flow parameter each depending on pressure ratio and turbine velocity ratio. As well as enabling the measurement of these characteristics in a standardized way, the turbine test rig also allows detailed local investigations depending on the specific needs. The instrumentation is applied using standardized and permanently installed equipment as well as special equipment and is adapted considering the aim of each test series. The standard equipment consists of pressure and temperature measurement at the inlet and at the outlet of the turbine and a technique to determine the mass flow, rotational speed and torque of the water brake. The special equipment includes pressure tabs at other locations and pneumatic or optical measurement devices. The layout of the rig was chosen to make the installation of different designs of axial turbines possible. Different single stage turbines, the standard in today's large turbochargers, can be combined with the appropriate inlet and outlet casings. The 210
flexibility in the combination of the components opens a wide field of possible experimental investigations and optimization of given configurations. Figure 3 shows a photograph of the main components - turbine, gear box and water brake - of the test rig.
Figure 3 Turbocharger turbine test rig MONA VI with turbine, gear and water hrake (from right to left) Measurement uncertainties
The measurement of turbine performance requires high accuracy of the measured pressures, temperatures, mass flow and the power at the water brake. The definition of the main turbine characteristic, the turbine efficiency TlsT, is as follows:
PB: brake power; PL : bearing loss; mT: mass flow; TIT: turbine inlet temperature; POT: pressure at outlet; P*IT: total pressure at inlet; CIT: flow velocity at inlet; cp: heat capacity; R: gas constant.
It is important during turbine development that the accuracy of the measured efficiency is quite high. For the test rig MONA the uncertainty of the measurement of the efficiency is shown in Figure 4 as a function of the turbine expansion ratio. To generate this graph the measurement uncertainty was calculated by using the error propagation law from Gauss. This is applicable if the relevant measured variables are independent and it can be assumed that these measured variables are available with the same statistical certainty. The pressures and temperatures that according to the above formula are required for the determination of the efficiency are measured at several positions in each section of interest to obtain representative values. The data acquisition takes place with state-ofthe-art instruments of high precision. Thus the uncertainty of the efficiency value depends less on the accuracy of the temperature and pressure measurement than that of the mass flow rate and power measurement. The graphs in Figure 5 show the influence
211
of the uncertainty of mass flow and power on the total uncertainty of the measured efficiency. The diagrams show the dominant influence of the power at lower turbine expansion ratios. At full load conditions, the influence of mass flow is the determining factor in the uncertainty of the efficiency. Experience shows that substantial effort is needed to minimize the uncertainty of the mass flow and of the power to an acceptable level. The mass flow measurement is carried out with a calibrated V-Cone meter, which was developed by McCrometer Inc., USA. The outstanding attribute of the V-Cone meter is the marginal influence of swirling flow (13) and flow distortion due to pipe elbows (11) and valves (12). The accuracy ofthis technique is shown in Figure 6, right as a function of the pressure ratio for a test series on MONA.
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For measuring the turbine power, the momentum on the swing arm can be used. A load cell at the swing arm delivers the brake momentum. With the knowledge of the turbine speed the turbine power can be determined. The relative uncertainty of the power measurement is shown in Figure 6, left. It is important to note that if the determination of the turbine power would be carried out with the thermodynamic data at the inlet and at the outlet of the turbine, the influence of the uncertainty of the mass flow would be reduced. On the other hand, the required measurement value of the turbine outlet temperature would be quite inaccurate due to non-adiabatic effects. Therefore the accuracy of the efficiency determined with the measured power is approximately four times better. In the development process of new turbine components, the comparative test of different turbine versions is important. Therefore not only the total uncertainty but also the repeatability is of high importance. Fortunately, the repeatability of the turbine efficiency is significantly lower than the total uncertainty, particularly at medium and full load. Figure 4 shows the difference between these two uncertainties. The difference at high load is caused by the mass flow measurement. A significant fraction of the uncertainty of the mass flow is determined by systematic effects, which do not occur in repetitive measurements.
212
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Figure 6 Uncertainty of brake power and mass flow of the turbine test rig MONA Components of the test turbines The test turbine consists of the inlet housing, the single stage turbine and the outlet housing. Each component will have to fulfil certain efficiency requirements in an experimental verification. If necessary, an optimization guided through experiments can be performed. Flexible design of the test turbine is the key to a meaningful and economically feasible test series. The low turbine inlet temperature of 150 degree Celsius offers quite a few advantages in modelling the housings, the flow guiding components and the vanes and blades in an early phase of the development process. Housings
Prototype housings are preferably produced with the costly rapid prototype procedure. Another more economical possibility is a welded construction if the required shape allows this_ Due to the low inlet temperature, it is even possible to manufacture prototype gas outlets from wood. This is also cost effective and allows, if beneficial, smaller modifications of the flow-guiding contour during the test phase. Figure 7 illustrates examples of such wooden internals.
213
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Figure 7 Housing internals made from wood: turbine outlet diffuser (left) and turbine outlet housing (right) Turbine stage
The low turbine inlet temperature has further advantages. The guide vanes and the rotor blades can be manufactured from aluminium by milling. Five axis-milling machines make the manufacturing of vanes and blades flexible and inexpensive. Figure 8 shows the milling process of a guide vane and a ready-made blade.
Figure 8 Manufacturing of vanes (left) and blade made from aluminium (right) Measuring Positions
The third advantage of the low turbine inlet temperature is the reduced effort for the application of pressure taps. To obtain complete information about the thermodynamic behaviour of the stage, wall pressure taps are mounted in the hub and tip casings of the investigated measurement planes (entrance of gas inlet, exit of gas inlet, entrance of turbine stage, between stator and rotor, exit of turbine stage, entrance of gas outlet, exit of gas outlet). To trace the pressure recovery of the diffuser and the outlet housing, pressure taps are distributed along the streamlines. Also the total pressure loss of the inlet housing can be measured with pressure taps at the appropriate planes. The pictures in Figure 9 give an impression of the extensive application of the instrumentation wiring in the machine.
214
Figure 9 Pressure tap application in the turbine diffuser (left) and in the bearing casing (right) EXPERIMENTAL RESULTS The design of the turbocharger turbine is highly influenced by the requirements of the turbocharger compressor, the design of which in tum is dictated by the pressure ratio and mass flow requirements of the engine. Of major importance is the efficiency, whereas it is helpful to have the possibility to characterize the components separately. Gas Inlet At high specific flows, the loss in the turbine inlet casing might account for up to about 2 % points of turbine efficiency at high turbine pressure ratios and at low turbine pressure ratios even up to about 4.5 % points of turbine efficiency. The exhaust pipes of 4-stroke engines are usually narrow with a cross section area smaller than the turbine stator inflow area, so the gas inlet casing of the turbine has to be designed for flow deceleration. Flow deceleration usually leads to high losses. To improve this situation, the stator inlet area is reduced but a compromise has to be made, as this reduction also is a source of loss. Considering the requirement for a compact design of the turbocharger, this makes the design of the gas inlet casing (GIC) a challenging task. It is state of the art to use CFD for an optimization of the gas inlet contour. The design as a result of simulations can be confirmed by measurements on the test rig MONA. The determination of the pressure loss coefficient ;GIC can be achieved by characterizing the fluid flow at the inlet of the gas inlet (plane TI) and at the exit of the gas inlet (plane 0):
As can be seen from the equation, the total pressure loss is related to the dynamic pressure at the exit plane of the gas inlet. At first sight, this is an uncommon way to compare pressure losses of components. But in the case of the gas inlet of a turbocharger this is advantageous. Turbochargers are applied to many different engines with their specific charging systems. Because of this, different designs of gas inlet casings with varying inlet diameters and flange position are in use in combination with the same turbocharger turbine. Therefore, the relation to the velocity at the exit plane of the gas inlet with its constant cross section area is useful. With knowledge of the pressure loss coefficient, the loss in turbine efficiency caused by the losses in the gas inlet can be estimated. 215
Turbine Efficiency
A basic design decision for the turbocharger turbine is whether it is used on pulse charged engines or if it is used on engines with quasi-constant charging systems only. A detailed description of the specifics of the two different charging systems is given in (9). Nevertheless, the high excitation forces of the turbine blades on pulse charged engines require special measures to keep the blade vibration amplitudes at an allowable level throughout the operating range (5). The use of a lacing wire has been found to be a very effective way to achieve this. Influence of lacing wire
The lacing wire will however always be accompanied with aerodynamic loss (4). To quantify this loss experimentally, measurements were performed on the turbine test rig MONA. Tests were run with a freestanding blade design and were compared to test runs with blades equipped with a lacing wire. The results can be seen in Figure 10, which shows the turbine efficiency versus the turbine expansion ratio for two flow capacities of the turbine. It can be seen that over the whole range the loss caused by the lacing wire is about 2 to 3 %- points of turbine efficiency. This is substantial, so avoiding the lacing wire is advantageous for engines where pulse charging is not required, or in other words, when the mechanical requirements for the turbine allow doing so.
_ f-+--
:r: 0.85
~
'!'
0.80 -
~
r;c 0.75 ·u" ~
r[---
0.70
___ Free standing blade, lower mass flow
!!
:is
~
[--1-
..... Free standing blade, higher mass flow
0.65
- 0 'Blade with lacing wire, lower mass now
- I:l 'Blade with lacing wire, higher mass flow
++++-
0.60 1.0
1.5
2.0
2.5
3.0
3.5
4.0
Turbine expansion ratio [-]
Figure 10 Measured turbine efficiency versus expansion ratio for freestanding blade and blade with lacing wire
In this test series it was advantageous to use blades made from aluminium. Therefore, the realization of the blade variants- freestanding and with lacing wirecomprising the same aerodynamic layout was possible. A second example of using MONA to get a complete collection of data is the comparison of different turbine designs.
216
Comparison of different designs
Experience indicates that the turbocharger axial turbine without lacing wire has to be specifically designed to perform reliably. One aspect considering blade vibration issues is described in (6). For the present paper, a wide chord design of a free standing blade and a different blade layout with lacing wire was compared (see Figure 11). Details about the wide chord design (turbine 1) can be found in (3), whereas the turbine, which is applicable for pulse charging systems (turbine 2), is described in (10).
Figure 11 Wide chord design: turbine 1 (left); design with lacing wire: turbine 2 (right)
Figure 12 shows the results of the measurement of turbine efficiency of the two different designs. Due to the different applications of the two turbines, adequate gas inlet casings were used for the turbine measurements. Due to this, a comparison of the turbine stages requires a correction of the influence of the gas inlet on turbine efficiency. This can be achieved by calculating loss or gain in efficiency compared to a reference gas inlet casing, with a certain pressure loss (compare above). Figure 12 collects the results. For turbine 2 and lower pressure ratio a loss of nearly 4% points is caused by the gas inlet 2 compared to a chosen reference gas inlet. Gas inlet 1 exhibits only a slightly increased pressure loss as compared to the reference gas inlet. The maximum gain in efficiency is 0.5% points. Now a fair performance comparison of the turbine stages is possible. It shows the advantages of turbine 1 (wide chord design, free standing blades) for lower pressure ratios. For example, the efficiency for turbine 1 at a turbine expansion ratio of 1.6 is nearly 1.5% points higher than for the design of turbine 2 (blade with lacing wire). On the other hand, despite the use of a lacing wire for turbine 2, high efficiencies could be obtained throughout the measured range. Through a careful aerodynamic design, the performance penalty of the lacing wire could almost be compensated and the efficiency level of the present turbine without lacing wire has almost been reached. Comparisons with previous designs, which are presented in (3) and (10) respectively, show the superior efficiency of theses turbines and, therefore, the advancements in modem turbine design. As mentioned above, it is a specialty of turbochargers that even at constant engine load the operating conditions for the turbine are not constant. Therefore, the influence of the velocity ratio on turbine performance needs to be evaluated.
217
:::;: 0.85
i
• 0.80
~
~ 0.75
c:
'u'"
~ 11
0.70
- - 0 - Turbine
:e
1 gas inlet 1
I l - Turbine 2 gas inlet 2
~ 0.65
~Turbine
1 reference gas inlet
""'-Turbine 2 reference gas inlet
0.60 1.0
1.5
2.0
2.5
3.0
3.5
4.0
Turbine expansion ratio [-I
Figure 12 Measured turbine efficiency versus expansion ratio for two different turbine designs (compare Figure 11) Influence of velocity ratio uic;s
Figure 13 shows the turbine expansion ratio versus the velocity ratio for a turbine operating on an engine with a typical pulse charging system. This data is the result of an engine simulation (1) at 87% engine load and gives the operating conditions over one engine cycle. It can be seen that the turbine expansion ratio varies quite strongly since in a pulse charging exhaust system the pressure pulses from every single cylinder of the engine are isolated as much as possible from the engine to the turbine. This provides an advantageous part load behaviour since the high energy content of the unsteady pressure pulses can be utilized. On the other hand, the turbine will not constantly operate in its design point. The turbine expansion ratio as well as the velocity ratio is unsteady which leads to inevitable losses in turbine efficiency. Although the range of the unsteadily experienced velocity ratio even broadens at part load, pulse charging systems still provide advantageous part load behaviour without having to rely on supplementary components (2). For an evaluation of such turbines the influence of the velocity ratio on the turbine efficiency has to be known. This data can be collected on the turbocharger turbine test rig MONA. Figure 14 shows the result of a measurement. The turbine efficiency is shown for constant turbine expansion ratio while varying velocity ratio. This is possible since the turbine shaft power on the test rig can be adjusted by using the water brake. The variation of u!Cis is limited by the acceptable rotational speed of the turbine at the high end and by the maximum power, which is limited by the power consumption of the water brake at the low end. Figure 14 documents the typical behaviour of axial reaction turbines, which exhibit the optimum velocity ratio in the region between 0.55 and 0.70. The decrease of efficiency for higher and lower velocity ratios is due to additional losses caused by incidence effects for the incoming flow at rotor inlet and by swirl in the flow at rotor exit. Comparing Figure 13 and 14 it has to be mentioned, that only certain turbine expansion ratio- velocity ratio combinations are relevant for the operation under engine conditions! In Figure 13 and 14 these combinations are indicated for the given data. It can be seen that for a complete experimental evaluation of turbines for pulse 218
charging systems, measurements for different turbine expansion ratios and varying velocity ratios have to be performed!
,r-
-+-:-:.-!'-
If
I
t -f.-I I 1.0 +LLL..LLJ-'--"+--.LW-L..Ll--L.L+-,--'--'--L.L!lL..Ll-"+-'--.LL.LL..Ll--L.L+-,-L.LLLJ-'-'-----L.f :1 +1- t-++i-++++++t-t+-H+-+H
0.45
0.55
0.65
0.75
0.85
0.95
Velocity ratio u/c;. [-I
Figure 13 Turbine expansion ratio versus velocity ratio for a turbine integrated into a typical pulse charging system
The extraordinary value of the test rig MONA, where a wide variation of velocity ratios is possible, can also be demonstrated by Figure 14. The thin dashed line in the figure was extrapolated by using the measured efficiency data for a pressure ratio of 1.6. in a limited range of velocity ratio. The variation between 0.6 and 0.7 displays what would be possible on a conventional turbocharger test rig (hollow squares only). The extrapolation of these limited data, which would be necessary to evaluate actual onengine behaviour, would inevitably lead to an imprecise estimation of turbine efficiency for high and low velocity ratios.
7;
0.90 _
• expansion ratio = 2.5 _
0.85
• expansion ratio = 1 _6
'j 1ii
.a 0.80
s
'u O 75
f!
~
. . 0.70
:e?!.
0.65
.
.
.
0.60 -=±:jI:::rul1'!:il::l:TI:m~rnI:JTIl:TI:mTI:::D:1JJLillW, 0.45
0.55
0.65 Velocity ratio
0.75
0.85
0.95
u/c" [-I
Figure 14 Turbine efficiency versus velocity ratio for turbine 2 (compare Figure 12) at constant turbine expansion ratios
219
For high specific flow rate turbine designs, which are common for turbochargers, the aerodynamic performance of the inlet as well as the outlet casings have an important influence on the overall turbine performance. At high flow rate the specific kinetic energy just behind the rotor blade can account for more than 15% of the turbine enthalpy drop. If the exhaust casing cannot convert this kinetic energy to a static pressure rise, 15% of turbine efficiency is lost. Gas Outlet In practice, about 30% to 40% of the kinetic energy can be converted to static pressure rise in a well-designed exhaust casing. The specific design of an exhaust casing is a challenging task due to the complex flow structure in the exhaust casing, diffusing flow with separation effects, and coupling of turbine exit flow with the exhaust casing flow. In addition, the axial length of the exhaust casing has to be limited. For the engine builder a short design is advantageous since this supports a compact layout of the exhaust system. A new exhaust casing with reduced axial length was optimized in an approach combining experiments on MONA and CFD calculations. Figure 15 shows a sketch of the longer outlet casing (gas outlet 1) and the newly designed shorter gas outlet casing (gas outlet 2). The axial length was reduced by more than 30% and as Figure 16 illustrates, the same or even slightly higher turbine efficiencies could be achieved nonetheless. The diagram shows turbine efficiency based on total conditions at inlet and static conditions at outlet versus expansion ratio. The values refer to the turbocharger flange locations. The constant loss of the inlet and the design dependent loss of the outlet casing are included. For this investigation the previous described attributes of the turbine test rig MONA could be taken advantage of. Detailed insights could be realized due to the convenient instrumentation with sensors along the gas path. This also helped to accelerate the design phase with the approach combining experiments and calculation. Required modifications could easily be implemented. r- -
,....------..,..--...,
I
/
I
I \ \
"- .......
~---""':::.-"'tI Gas oudet 1 I
_
~--lilliiiI
Figure 15 Sketch of two different gas outlet designs
220
0.90
,
::!: 0.85
~"
..
, 0.80
~
,
~ 0.75
c .!!
ii.." 0.70 c
'-
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0.65
1.0
1.5
2.0
~Gas
outlet 1; lower mass flow
~Gas
outlet 1: higher mass flow
~
"'Gas outlet 2; lower mass flow
....e
-Gas outlet 2: higher mass flow
2.5
3.0
3.5
4.0
Turbine expansion ratio [-J
Figure 16 Measured turbine efficiency versus expansion ratio for two different designs of the gas outlet SUMMARY Rig testing is not only mandatory in the design and development phase of new turbocharger turbines, but also plays an important role in helping turbomachinery designers continue to improve their design tools and methods with accurate validation data. The turbocharger turbine test rig MONA was described in the present paper. The article gives an overview of the rig installation and its integration into the infrastructure at ABB Turbo Systems Ltd. The configuration of the test rig to meet the requirements during the development process of a turbocharger turbine, considering not only time and cost but also flexibility, is highlighted. One characteristic is the operation at low turbine inlet temperature to allow the use of materials that facilitate a rapid prototype manufacturing process of the components. Wooden parts can be used to modify the casings; aluminium can be used to produce stator vanes and rotor blades by milling. The rig is equipped with a water brake to realize a wide range of operating conditions for the tested turbine,. which correspond to the operating range of a given engine. To demonstrate the experimental possibilities of the turbocharger turbine test rig MONA, investigations concerning the pressure loss in gas inlet casings, the influence of a lacing wire on the aerodynamic turbine performance, the efficiency of different layouts of turbine stages, and the pressure recovery of diffusers in combination with the gas outlet casing are presented. ACKNOWLEDGEMENTS The authors would like to thank ABB Turbo Systems Ltd. for the permission to publish this work. The presented data could not have been collected without the help of quite a few members of the ABB Laboratory of Thermal Turbomachinery. The authors would like to explicitly mention their highly skilled colleagues from the application shop as well as Dr.-Ing. Ge Xu for manufacturing the prototype turbines. Dr. Douglas Robinson helped in "language polishing".
221
REFERENCES
1 T Bulaty, E Codan, M Skopil, "A flexible system for the simulation of turbocharged diesel engines and turbocharging systems", Spring Technical Conference of the ASME, Youngstown, Ohio, USA, ICE Vol. 26-3, pp. 57-63, 1996. 2 E Codan, "Die Aufladung zuktinftiger Grossdieselmotoren", Aufladetechnische Konferenz, Dresden, 2000. 3 R Dotl, D A Pesten, 0 Bernard, B Phillips en, "Turbocharger power for future railway applications", ASME- Paper ICEF2003-737, 2003. 4 J Dunham, B Phillipsen, "The application of streamline curvature code to the design of turbochargers", Proceedings of the 7th International Conference on Turbochargers and Turbocharging, pp. 169-179, ISBN 1 860583830, IMechE, Savoy Place London, 14.-15.05.2002. 5 D Filsinger, 0 Schafer, "Numerical calculation oflow order blade excitation in pulse charged axial turbocharger turbines and its experimental assessment", ASME Turbo Expo, GT2003-38182, 2003. 6 D Filsinger, Ch Frank, 0 Schafer, "Practical use of unsteady CFD and FEM forced response calculation in the design of axial turbocharger turbines", ASME Turbo Expo, GT2005-68439,2005. 7 D Hinch, "Rig testing most effectively validates and optimizes the performance of turbomachinery beyond the limits of analytical tools", SpinOffs published by Concepts NREC, Volume 1, Issue 4,2003/2004. 8 K-H Rohne, "Technologien fUr grosse Hochleistungsturbolader", 13. Aachener Kolloquium Fahrzeug- und Motorentechnik, Proceedings, Vol. 2, pp. 1019-1038,2004. 9 N Watson, M S Janota, "Turbocharging the internal combustion engine", The MacMillan Press Ltd., ISBN 0-333-24290-4, 1982. 10 D Wunderwald, K Heinrich, "Meeting the requirements of modern diesel and gas engines: The new TPL..-C turbocharger generation", 24th CIMAC World Congress, Kobe, PaperNo. 133,2004. 11 S Ifft, ED Mikkelsen, "Pipe elbow effects on the V-cone flowmeter", North Sea Flow Measurement Workshop, Peebles, Scotland, 1992. 12 S Ifft, "Partially closed valve effects on the V-cone flow meter", McCrometer Inc., Hemet, CA 92545, USA. 13 J S Shen, J Bosio, S Larsen, "A performance study of a V -cone meter in swirling flow", North Sea Flow Measurement Workshop, Lillehammer, Norway, 1995.
© ABB Turbo Systems Ltd.
222
ACTIVE CONTROL TURBOCHARGER FOR AUTOMOTIVE APPLICATION: AN EXPERIMENTAL EVALUATION Apostolos Pesiridis and Ricardo Martinez-Botas Department of Mechanical Engineering Imperial College London SW7 2AZ Exhibition Road, London
ABSTRACT The current paper presents the results from a comprehensive set of experimental tests on a prototype active control turbocharger. This is a continuing series of test work as part of the development of this new type of turbocharger. Driven by the need to comply to increasingly strict emissions regulations as well as a continuing strive for better overall performance the active control turbocharger is intended to provide an improvement over the current state-of-the-art in turbocharging. In this system, the nozzle is able to alter the throat inlet area of the turbine according to the pressure variation of each engine exhaust gas pulse thus imposing a substantially more 'active' form of control of the conditions at the turbine rotor inlet.
NOMENCLATURE U
m
FGT VGT ACT ER P T
Cp
MFP
ulC
i,
MFT LVDT CFRP FWG
blade tip speed mass flow rate Fixed Geometry Turbocharger Variable Geometry Turbocharger Active Control Turbocharger turbine expansion ratio or pressure ratio Pressure temperature specific heat at constant pressure mass flow parameter,
.rr. MFP=~
Greek symbols r ratio of specific heats 11 efficiency 8,AfJ Nozzle position and amplitude, respectively Subscripts is isentropic act min max
actual minimum area maximum area
0
total/stagnation condition
2 Po velocity ratio 3 Mixed Flow Turbine 4 Linear Variable Displacement Transducer Carbon Fibre Reinforced t-s Plastic Function Waveform Generator null
pulse generator turbine inlet turbine exit turbine total to static null position of actuator or nozzle
INTRODUCTION The continuous increase in requirements from charging systems as a part of modern internal combustion engines is driven by the rapid and very demanding introduction of
223
new emissions legislation, as well as ever higher power density requirements, whilst maintaining or improving the fuel efficiency. A relatively recently matured and increasingly popular charging technology is the Variable Geometry Turbocharger (VGT). The benefits accrued by the use of variable geometry devices for exhaust gas flow control to the turbine are well known and include improved transient response, fuel economy and more importantly reduced emissions in the face of ever stringent emissions regulations. The common problems encountered with VGT were reliability (for long periods of time while exposed to high temperature and corrosive exhaust gases), complexity because of the VGT actuation mechanism and control system, and subsequent, high cost [1]. However, recent research has tended to provide acceptable solutions to most of these problems and today VGTs have already had a significant impact in the design of small diesel engines. A fundamental issue that has so far not been addressed satisfactorily, however, is the less than ideal combination of a reciprocating engine providing energy to drive a rotodynamic machine such as a turbocharger turbine. Yet even with the advent of VGTs this mismatch is not eliminated, since a VGT responds to operating point changes only, i.e., for steady-state operation the nozzle setting of a VGT assumes one non-changing, optimal condition. However, regardless of the engine operating at steady-state or transient mode, the inlet conditions to the turbocharger still include a highly pulsating flow field with widely varying pressure and mass flow rate levels at all times [2]. It may still be possible to harness that energy by continually altering the effective throat area of the turbine by means of a fast-response nozzle. The Active Control Turbocharger (ACT) is a special type of VGT, where the nozzle (in this case a sliding wall-type restrictor) is able to alter the inlet area at the throat of the turbine inlet casing (volute) in phase and at the same frequency as that of the incoming exhaust pulses (Figure 1). For this purpose it is actuated by a suitable electrodynamic shaker - which supplied by a powerful amplifier - is capable of meeting the frequency and displacement requirements of this intensive and continuous operation.
Start of exhaust pr-ocess
! ... ~
!
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\lone gap B
« u
IJ' I])
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o
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'" N
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End of exhaust process
Fig 1 The ACT concept of operation during one Exhaust process cycle (240° CA) depicting adjustment of area as pressure changes during the course of one pulse. 224
In early production VGTs, open-loop control systems were used to provide the required boost pressure without taking into account engine volumetric efficiency. Current production VGT systems provide closed-loop control with engine volumetric efficiency being taken into account. An undue reduction in volumetric efficiency causes specific fuel consumption and emissions penalties. Excessive volumetric efficiency reduction is the result of excessive negative pressure drop across the engine i.e., when the exhaust manifold pressure instantaneously exceeds the inlet manifold pressure. In variable geometry mechanisms this is the result of excessive inlet area restriction However, there are a number of mechanisms available to prevent this from happening. Minimum turbine inlet areas need to be imposed in order to avoid this excessive pressure build-up and in the case of ACT operation the minimum inlet area per cycle can be phased in such a way as to avoid the valve overlap period of the engine. Minimal valve overlap is, therefore, a desirable engine characteristic. The minimal inlet areas that were tested were phases of 30°, 60°, 90° and 240° after the exhaust gas pulse start. Since, the valve overlap period occurs at any point after 180° of crankshaft rotation from the time of the exhaust valve opening the 240° and 30° cases are susceptible to poorer results. However, the actual phasing will be a function of exhaust manifold length as well as geometry and the wave dynamics of the particular exhaust process under consideration. The results of these tests are presented in later sections providing a measure of the capability of the ACT concept. In addition, the experimental setup of the laboratory is described and the design and operating details pertaining to ACT are explained. EXPERIMENTAL TEST FACILITY The turbocharger aerodynamic test facility is a simulated reciprocating engine test bed for turbocharger testing. In Figure 2(a) a schematic of the laboratory set-up for ACT testing is illustrated. Three screw compressors supply air at room temperature up to a maximum mass flow rate of 1.2 kg/so A 76kW heater stack is used to bring the air to the required test condition. The pulsations are generated by a pair of counter rotating plates situated 755 mm upstream of the measurement inlet plane. The unsteady flow generated replicates an engine exhaust manifold, albeit operating in cold flow; the nondimensional mass flow rate and speed parameters are correctly matched. The air flow is split into two pipes soon after leaving the heaters, this feature allows testing of twin entry turbos. The current turbine is part of a single entry system, thus the air flow is merged prior to entering the volute.
Instrumentation Figure 2(b) shows the hardware setup for high-speed data-logging and ACT control. Various sensors along the pipework upstream of the turbine and on the turbine itself are utilised. Pressures from the measurement plane, and turbine inlet casing stations, CT A (constant temperature hot wire anemometer), LVDT nozzle position and FWG (Function waveform generator) signals are directed to the high-speed analogue input data acquisition and control card, while the signal from the magnetic pickup sensor measuring pulse generator frequency and the signal from the turbine speed optical sensor are directed towards the high-speed, counter timer card.
225
Dynamometer Turbine Inlet casing Pulse Generator CTA___ Pressure Transducer (xi
High speed
counterftlmer card-U-H---r.-
High speed
Nozzle
FWG~~~;=====Jt~~~
analogue Input card
Ass_1y
(a)
(b)
Fig 2 (a) Turbocharger test rig schematic and (b) laboratory schematic of major high-speed data acqnisition hardware as well as the set-up for ACT control. Active control turbocharger Design The main ACT components can be seen in Figure 3. A nozzleless Holset H3B turbocharger turbine was modified to accept a nozzle for VGT testing. Due to the very small space available and dynamic requirements, a very thin (1.5mm thickness) section, tubular in shape, sliding nozzle was designed (Figure 4). Two embedded attachment points at the end away from the turbine inlet casing are used to mount the nozzle on to its actuating arm (yoke) via two small bearing pads, which fit between the nozzle and the yoke and assist in translating the pivoting movement of the yoke into a linear motion by the nozzle. The nozzle itself is guided towards the throat by an inner guide, which serves both as a bearing surface to the nozzle as well as forming the exducer section of the turbine. An outer guide houses the entire assembly and at its lower end is shaped into a bracket on which the yoke is attached and can freely pivot about. The entire assembly is held together and attached to the face of the turbine by six 50mm M5 bolts. The throat width to the turbine is 26mm wide and the nozzle can block 21.5mm off that passage width before reaching the stops. The materials used for the assembly were 6082-T6 aluminium alloy, for the entire assembly except for the nozzle, providing light weight, with adequate strength to overcome (in the case of the yoke) fatigue, which is the major issue in this type of operation. At the most severe operation the yoke achieved a safety factor of 10 with respect to the maximum bending stress applied to it at its corners by the bearing pads. The nozzle was of carbon fibre reinforced plastic (CFRP) construction, providing adequate strength for ultra low weight against the high pressure air flow used during testing - low nozzle weight being critical for the ACT in achieving good force performance.
226
BeatiugPad !onerGuide ~t'Sx50 StH Bolt 6 Yollile 1 pj"otPin Pivot Bearing
(a)
(b)
Fig 3 (a) Active control turbocharger exploded view and (b) assembly. An electrodynamic shaker was used to drive the nozzle in ACT mode (Figure 5). The shaker was chosen so as to be able to provide the required force for the limiting frequencies and amplitudes expected during testing (60Hz and up to 21.5mm, respectively). Direct coupling of the shaker and nozzle was not possible due to the nozzle being driven from the exhaust side of turbine, thus forcing the shaker to be offset to the side of the exducer. The simplest way to drive the nozzle from this position was with a single-piece, pivoting yoke-type actuator arm. This makes the design, simpler, more compact and more importantly allows for far better fatigue endurance with the added advantage of significant vibration damping due to the rocker-arm type operation of the yoke used to drive the nozzle, which largely damps out what would otherwise have been severe vibrations through the turbocharger. The total mass of all moving parts (Part Nos. 2, 3, 4,8 and 9 in Figure 3(a)) is only O.241kg, which was important in order to be able to achieve the force performance required for testing, even though it means a penalty in terms of vibration, since at present it is not possible to balance both sides of the yoke, as the top part (nozzle) is significantly lighter to the bottom part that also contains approximately O.2kg of the moving part of the shaker. Outer guide, \
I; Dynamometer
~9~~~E~~~/ Inner guide
2: !'ctive C:ontrnl "l'urbul'hHrgl'r 3: Actulitor (shHker) 4: Inlet pressure ITlwstJuccr
~ Rotor
Actuator yoke
Fig 4 ACT sectional view with MFT rotor 'D'. The thin section nozzle can be seen protruding axially at the inlet to the turbine.
Fig 5 Active control turbocharger with electrodynamic shaker and dynamometer
227
Operation The electrodynamic shaker is connected to the lower part of the yoke through an adjustable lever (Figure 2) used to set the nozzle restriction (iJVGT ) in VGT mode or the null point (iJnull ) in ACT mode. For VGT testing once iJVGT is set, its value does not change through either steady or unsteady testing, while for ACT testing iJnuli is set at the midpoint of the nozzle oscillation amplitude, between the maximum intended open nozzle area point (OVGT = Omax ) and the minimum area available (see Table I below). The amplitude, frequency and form of the controlling signal is controlled by a function waveform generator, which commands a control input in the form of a selected waveform to a 100W amplifier as well as to the controlling computer. From the amplifier the signal is routed to the shaker and transformed into motion of the nozzle. High speed pressure transducers located at the measurement plane as well as just upstream of the nozzle are used to measure the inlet pressure level. Timing is achieved through the use of the magnetic pickup sensor on the pulse generator to define accurately the pulse period and its start and end. The pulse generator opening signal is used to trigger the FWG, which in turn causes the nozzle to move in response to its signal. The known length of the pipework from the pulse generator to the nozzle is used to derive the pressure wave travel time, which travels at the acoustic velocity [3]. In addition the total system (control input to feedback) time response is known through testing and a time constant is thus derived that is input into the controlling software that controls operation of the ACT. Thus, the shaker is always in phase with the inlet pressure signal. Additional out-of-phase, resulting due to the fact that the total energy travel time is dependent on both the acoustic velocity of the pressure wave in addition to the local flow velocity [3], is detected through suitable software and adjusted manually as required, thus achieving a reasonably accurate phasing of inlet pressure and nozzle signals. All acquired data are phase shifted (lagged) eventually, from their measurement location (Figure 2(b)) to the turbine inlet entry according to the procedure described in [3]. For VGT testing iJVGT can be set by adjusting the lever only. For ACT testing, iJnull is adjusted manually, while b.iJACT and the shaker frequency are preset on the waveform generator according to the test parameters required. Typical iJVGT test points used during testing are included in Table I below (in normalized throat open area form); 'Ia hi e 1 Nozz Ie VGT setfmgs tor the 26 mmwi"de th roat area Nozzle restriction into throat (mm) iJV(JT (%)
0 4 8 12 16 20 2l.5
100 (Fully open area) 84.6 69.2 53.8 38.5 23.1 17.3 (Fully restricted area)
EXPERIMENTAL RESULTS The parameters tested included effects of varying frequency of exhaust pulses (engine rpm), different turbine rotational speeds and loads (expressed in millimetres distance 228
between the rotating eddy current dynamometer magnet and the two equi-spaced stators on either side with lOmm being the lowest and 0.5mm the highest load, respectively) set against a specific amount of isentropic energy provided upstream of the turbine. In addition, varying nozzle oscillating amplitudes (A 19) in ACT operation as well as the effect of different nozzle input signal phasings in relation to a baseline VGT inlet pressure trace were tested. The unsteady tests were carried out at three different VGT settings, 38.5%, 53.8% and 69.2% open throat area. These formed the reference point against which all the ACT tests are compared. Unlike the original idea to extract energy only at the lower (near ambient) energy levels of the pulse (Figure 1) where the minimum area provided by the nozzle occurs when the maximum throat area and the peak of the pulse are in-phase (i.e., the minimum area occurs 240° after pulse generator opening. Three other minimum nozzle area phase settings were tested: one, 30° after pulse generator opening, the second at 60° (at the nominal peak of the pulse) and the third at 90°. Figure 6, illustrates a reference inlet pressure signal of the VGT operating at a nominal constant 53.8% open throat area. The ACT nozzle amplitude during testing reached a maximum area equal to the VGT setting illustrated and a minimum area equal to the minimum physically possible (17.3%). The effect of the different phase settings is illustrated. The reason for reducing to a minimum the available throat area, at these four different points in the pulse was in order to evaluate the effect that these would have on energy extraction, which is the aim of this project. Indeed as already illustrated in previous work on this rig [3] the amount of energy available for a proportion of the pulse period is indeed negligible and, therefore, a modest overall energy extraction is to be expected from this region. Figure 7, illustrates the point of using phase settings other than 240°. This particular phase is barely an improvement over the basic VGT setting, but the 30° and 90° settings are notable improvements, while the 60° represents a very substantial improvement over the reference VGT trace. 1.7
1.8 1.7
- - Acr -
1.6
i
i 0
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1.6 1.5 06 1.4 0.4
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~
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0 60
120
100
240
300
0
360
120
100
240
300
Oank Angle (Degrees)
Oank Angle (Degrees)
Fig 7 Four ACT inlet pressures at different phases, shown here along with the baseline VGT inlet pressure signal.
Fig 6 Four different phasing schemes of the ACT nozzle oscillation tested, shown here along with the baseline VGT inlet pressure signal.
Figure 8, shows the equivalent energy extraction in terms of actual power for two different frequencies (40Hz and 60Hz) and at different loads. It is noteworthy that at the pulse peak the 90° setting achieves lower energy extraction than either the 240° setting or the VGT and FGT. This is due to the comparatively lower efficiency attained due to a 229
certain area restnctlOn already present at the peak due to the sinusoidally actuated nozzle. When the area restriction increases it becomes more optimum and the efficiency improves by comparison to the baseline and there clear work extraction is achieved. This phenomenon can be, also, observed in Figure 9. 38000
- - A C T - 2'10:::~1
34000
~
-
ACf· 240CBCj
-
VGT5-1B';.r~Jell
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- - A C T 3Cdeq --~V(?iT
a
120
60
180
240
180
120
60
Oank Angle (Degrees)
240
crank Angle (Degrees)
Fig 8 (a) ACT actual power absorbed against the baseline VGT actual power recovered at 60Hz, 4mm load and (b) at 40Hz and 3mm load. With reference to Figure 6, while at the pressure peak the other settings have a maximum throat area, the 90° setting due to its phasing has less than 40% available throat area. The consequence of this is higher pressure drop but a reduction in efficiency which may locally reduce the amount of power recovered. This is a well recorded tradeoff of variable inlet area turbines [4, 5, 6, 7 and 8]. However, in cycle-average terms there is a clear benefit from this type of operation as indicated by Table 2 below for the case of 60Hz at 4mm load:
Table 2 ACT energy extraction benefit over VGT at 60Hz, 4mm load. ACTI240deg
VGT Wact (W)
5495 1.55
ER
ACT290deg
240 deg% Increase
5902 1.69
5667 1.74
90deg % Increase
3.13 12.57
7.41 9.18
A 7.4% increase in actual power is observed for the 90° setting and a 3.13% for the 240° setting when compared to the reference VGT setting. This proves the superiority of the former setting although a useful increase is achieved by the 240° case as well. In the 40Hz case in Table 3 below, all four different settings are presented. The most beneficial phase setting among the four tested was, predictably, the 60° setting. This is due to the fact that the minimum throat area is in-phase with the peak energy levels of the pulse. However, this means a larger trade-off with efficiency and therefore the end improvement is not significantly higher than the 90° case.
Table 3 ACT energy extraction benefit over VGT at 40Hz, 5mm load VGT
ACT-
ACT-
ACT-
ACT-
30deg
60deg
90deg
240dcg
30deg-% Increase
60deg-% Increase
90deg-% Increase
240deg-%
Increase
Wac! (W)
1904
2001
2043
2040
1963
5.12
7.J 1
7.18
3.09
ER
1.29
1.39
1.44
1.38
1.31
6.97
10.94
6.71
1.42
230
In Figure 9, the equivalent efficiency fluctuations over the pulse cycle are given. The low efficiency levels at the ends of each pulse are evident. This is where the energy levels of the pulse drop to near ambient conditions and where the efficiency suffers most. As mentioned previously, the improvement in efficiency at the beginning and end of the pulse is evident with a loss approximately 90° after the start of the pulse due to the less than optimum small area restriction present when compared to the baseline VGT area restriction. 1.2
1.4 1.2
'U'
1.9
i
~ ·0 in
'U'
1.9
0.8 0.6
i
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0.8 0.6 0.4 0.2 0 -0.2 -0.4
60
120
180
240
300
360
--,--~-
60
0
Crank Angle (Degrees)
120
180
240
Crank Angle (Degrees)
Fig 9 (a) ACT efficiency variation at 60Hz, 4mm load and (b) at 40Hz and 3mm load. Another way of presenting the effect of active control at the turbine inlet is given in the form of unsteady turbine maps in Figure 10. In this case at 40Hz the behaviour of the turbine is effectively quasi-steady. Energy recovery is evident by shifting of the loops towards the right on the x-axis i.e., towards a higher expansion ratio starting at its lowest from the reference VGT loop and at its highest with the 60° setting.
9
8
-
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o 1.8
1.4
2.2
2.6
Expansion Ratio (tolal to static)
Fig 10 ACT mass flow parameter variation with expansion ratio at 40Hz, 3mm load.
231
300
360
CONCLUSIONS In addition to the development background on the design of the Active Control Turbocharger prototype and the experimental setup, a wide range of unsteady testing on an ACT was presented. A notable potential was demonstrated at different loads and turbine speeds. In this case 50% and 70% speed results are presented with powers ranging from lkW to lOkW cycle averaged with an amplitude of over 30kW. Actual power recovered ranges on average between 3% and 7% depending on nozzle area phasing throughout the range of testing carried out while pressure recovery is even higher but with an efficiency drop as a trade-off at certain periods when the area restriction is not significant. The potential of this system, however, noteworthy, is still hampered by the less than ideal match of the sleeve nozzle to a MFT inlet due to the creation of a relatively large interspace area formed between the nozzle and the turbine, which is responsible for part of the loss experienced by this turbine, as the flow is allowed to expand again for some distance before reaching the turbine rotor. A more specific design to match a nozzle mechanism to a MFT is already underway offering substantially greater improvement.
REFERENCES Watson, N., and Janota, M.S., "Turbocharging the Internal Combustion Engine," Wiley Interscience, New York, 1982 2 Palfreyman, D., and Martinez-Botas, R. F., "The pulsating flow field in a mixed flow turbine: An experimental and computational study", Proceedings of ASME TURBO EXPO 2004, ASME GT 2004-53143, 2004. 3 Szymko, S., and Martinez-Botas, R., "Experimental Evaluation of Turbocharger Turbine Performance Under Pulsating Conditions", Proceedings of GT2005: ASME TURBO EXPO 2005, ASME GT 2005-68878, June 6-9, 2005 4 Pesiridis, A, Szymko, S., Rajoo, S. and Martinez-Botas, R., "Development of Active Flow Control in a Turbocharger Turbine for Emission Reduction", Conference proceedings of the Internal Combustion Engine Performance and Emissions Conference, IMechE., 7-8 December 2004 5 Pesiridis, A, and Martinez-Botas, R., "Experimental Evaluation of Turbocharger Turbine Performance Under Pulsating Conditions", Proceedings of GT2005: ASME TURBO EXPO 2005, ASME GT 2005-68830, June 6-9, 2005 6 Flaxington, D., and Szuzupak, D. T., "Variable area radial-inflow turbine", 2nd Int. Conf. on Turbocharging and Turbochargers, Proc. of the IMechE, Paper C36/82, 1982 7 Franklin, P. c., and Walsham, B. E., "Variable geometry turbochargers in the field", 3rd Int. Conf. on Turbocharging and Turbochargers, Proc. of the IMechE, paper CI21/86, pp. 241-250,1986 8 Capobianco, M. and Gambarotta, A, "Variable Geometry and Waste-Gated Automotive Turbochargers: Measurements and Comparison of Turbine Performance," J. of Eng. For Gas Turbines and Power, Vol 114, p.p. 553-560, July 1992
232
Thermomechanical analysis of a turbo charger turbine wheel based on CHT -calculations and measurements Tom Heuer, Bertold Engels, Horst Heger, and Achim Klein BorgWarner Turbo Systems Engineering GmbH Marnheimer Str. 85/87 67292 Kirchheimbolanden, Germany
ABSTRACT CFD, FEA and experimental testing have been combined in order to find lifetime limiting design deficiencies of a turbo charger turbine wheel. Based on two steady state CHT-calculations the heating process of the wheel has been simulated in a transient calculation. The resulting temperature distribution serves as boundary condition for a structural analysis. Taking into account centrifugal forces, the influence of the thermal stress is evaluated. Moreover, a method has been developed to extrapolate the transient temperature results according to a modified turbine mass flow and rotational speed. Results from experiments are used for the extrapolation method as well as for the numerical boundary conditions. A modified design of the back contour has been calculated and compared to the original design.
NOMENCLATURE h 1'J ill
n p PS q (J
SS t t T
l
heat transfer coefficient angle mass flow rate rotational speed pressure pressure side specific heat flux rate stress Suction side temperature time temperature dimensionless wall distance
Wm~2
Kl ] ] ] ] ] ] ]
N mm~2
s K
1 ] ] ] ] ]
Subscripts bc in max M P
ref
reI rot st th tot #1-5
boundary condition inlet peak value von Mises Prinicipal reference
235
relative rotational static thermal total location of monitor points
INTRODUCTION Growing demands for turbocharger performance yield a need for various technical innovations like the method of exhaust gas recirculation (EGR) producing higher exhaust backpressure, variable turbine geometry (VTG), regulated 2 stage (R2S) turbocharging system or a combination of all of them. Information about these methods is given in Schmitt et al. [1, 2]. Regardless which system is preferred, there is a tendency towards higher turbine inlet temperatures in order to increase the available level of power from the hot gas. Moreover, the performance of a turbocharger is strongly related to the design of the turbine wheel. Hence, the first demand is to guarantee good aerodynamics to achieve high power output and efficiency. However, for daily use, the durability is even more important and three major design criteria have to be fulfilled: The temperature in the turbine wheel must not exceed the maximum allowable material temperature. The thermo-mechanical load of the blades must be considered. This includes the limitation of the centrifugal forces and high frequency blade vibrations realized by defining a rotational speed limit. A complex stress situation in the hub arises from the combination of centrifugal forces and temperature differences. These local temperature gradients exist at steady state and they even increase when the operating point is changed, resulting in an unsteady heating or cooling of the material with uneven wall thickness. In order to evaluate and to improve the design of a turbine wheel, the afore mentioned criteria have to be investigated. Certainly, the temperature distribution inside the wheel is crucial for a thermal analysis. Nevertheless, even for a structural or a frequency analysis the temperature distribution is of great interest due to its influence on the mechanical and physical properties of the material. By gaining knowledge about the thermal load of the wheel, it is possible to determine the influence of thermal stress on the overall stress that includes the impact of centrifugal force. In order to receive realistic results, an accurate and practicable method is desired which makes use of experimental and numerical results. The advantage is that test stand data can provide boundary conditions for the simulations and, on the other hand, they are useful for validation purpose. The highest stresses are expected to appear during the heating or cooling process of the turbine wheel. Thus, a transient investigation is indispensable.
METHODOLOGY A method has been developed which makes use of experimental and numerical results. Five Thermocouples are appliqued below the blade surface and the measurement signal is transmitted via radio telemetry. One measurement point is used to define a boundary condition for the calculations, while others help to evaluate computed results. The quality of the thermal stress analysis strongly depends on the accuracy of the computed temperature distribution in the turbine wheel. There are basically two methods to determine this temperature. For the conventional method, heat transfer coefficients have to be prescribed at the fluid/solid boundaries. Heat transfer coefficients are well known for simple geometries like flat plates, pipes, etc., but the investigated geometry is complex and flow phenomena like separations, shocks, transition, etc. have significant impact on the local heat transfer coefficient [3]. Hence, approved values or correlations are often a kept secret in the turbomachinery industry.
236
In contrast, the Conjugate Heat Transfer (CHT) method takes into account the interdependency between fluid and solid. The high accuracy of CHT-calculations applied to the field of turbomachinery was proven by Bohn et al. [4]. The maximum relative difference between measured and calculated surface temperatures of a convectively cooled turbine airfoil was less than 5% [5]. Average values were about I %. The ability to provide accurate temperature information at any position of the wheel makes the CHT-method the preferred method over the conventional method. Since a transient CHT -calculation is still very time consuming, a new approach has been developed [6], which simulates heating and cooling of the turbine housing, respectively. Two steady state CHT -calculations are performed to determine the flow and temperature field at two different operating conditions. Beginning with the temperature distribution at partial load, an immediate change of gas temperature and mass flow is assumed. Therefore, the flow condition at high load serves as boundary condition for a subsequent transient heat transfer calculation simulating the acceleration process. Monitor points in the simulation, defined at the position of measuring points in the turbine wheel, give information about local temperature gradients. At discrete time points corresponding to the fastest temperature rise, a steady state stress analysis detects the regions of the highest loads. This subsequent step is essential because the uneven wall thickness and the connection to the shaft results in stress maxima which are not necessarily located at the same locations as the temperature peaks. In a second stress analysis centrifugal forces are considered additionally. Pressure forces due to the flow are neglected because they are assumed to be of minor influence [7]. If the highest stresses exceed the allowable values, a further design loop is required (Fig. 1).
c
OJ
'(ii Ql
",
ztW
~
l'
::E 0: 1
~
tl
WI
~J
~
E
0
Ifinal design I Figure 1 Workflow of the simulation process EXPERIMENTS The purpose of the tests is to measure the turbine wheel temperature at different locations and at different operating points. Therefore, the measurement signal of the thermocouples must be transmitted wirelessly, which demands a special measurement technique. A similar method has been applied by Ludewig [8, 9] to measure the strains in a compressor wheel at significantly lower temperatures. The basic idea is to drill a hole through the entire shaft and additional holes from the thermocouples to the shaft cavity. From each thermocouple, a wire is passed through the shaft to the transmitter which is located at the colder compressor side. Little tubes protect the wire but they also limit the number of thermocouples per rotor. Hence, a turbine wheel of a commercial diesel turbo charger has been preferred to a smaller passenger car application allowing three measuring points. Two further points given in figure 2 have been recorded with a second rotor.
237
Experimental setup The transmitter is mounted in a sleeve at the end of the shaft on the compressor side, according to figures 2 and 3. It includes the multiplexer, the amplifier, and the modulator, which are powered by a Li-battery. The signal is transmitted from the rotating part to the stationary receiver containing amplifier, HF/IF converter, and demodulator. The measurand is visualized with an oscilloscope and it is printed and stored electronically when the temperature is in steady state. The functional efficiency of the thermocouples regarding sensitivity and linearity is checked in the laboratory. The thermocouples are calibrated and the measurement range is adjusted by a precision millivoltage-source. A voltage jump represents a temperature difference between the measuring point and the contacts of the transceiver where the reference cold junction temperature is measured by a further thermocouple. Hence, the reference temperature must be added to each ~T measured at the probes. Furthermore, a dynamic balancing has to be done to guarantee a failure free operation.
,----------------,------------l
to test bed measurement equipment
---------------------1 Turbine wheel ,---------...
measurement points
1- _
•
#1/#5
ps/ss
~ 1 1 1
amplifier modulato oscillator multi lexe
B
HF
demodulator
amplifier
Transmitter
Receiver
rotating part
Figure 2 Experimental setup Rotor preparation The fixation of the thermocouples under the blade surface is a special problem to solve. To avoid any changes of the blade profile, grooves are milled into two different blades at both pressure and suction side. The surface in the area of the holes and the grooves is roughened by sandblasting and, subsequently, it is coated with an aluminium layer by flame spraying (Figure 4). After the connecting wires are laid, the thermocouples are fixed with ceramic glue. In the groove, two steel straps are spot welded to the blade for a temporary fixing of the thermocouples (Figure 5). The glue needs to cure before the holes and grooves are filled with a second layer that recovers the original surface. After the second curing period the shaft hole on the turbine side is filled with glue. The preparation of the turbine rotor is finished by a third curing period.
238
Figure 3 Transmitter
Figure 4 Milled groove
Figure 5 Fixed thermocouple
COMPUTATIONAL MODEL
Geometry The investigated turbine wheel is assembled in a radial turbine which has been designed for a commercial diesel turbocharger application (Figure 6). It is characterized by a controlled bypass and twin-scroll housing. According to the tested hardware, the intake is modelled as a Y-branch pipe and the outlet as a straight pipe. The walls are included because they are crucial for a correct heat balance. The turbine wheel with a scallop shaped back contains 11 blades. The shaft is cut at height of the piston ring including the borehole and the glue plug at the opposite end.
q=O ...-------....,
bore hole: ambience: h, Tret glue plug
outlet:
heat shield:
q=O
shaft: T shaft
Figure 6 Geometry and boundary conditions Computational grid In the computational grid, hexahedrons are preferred because they allow larger aspect ratios and help to reduce the total grid size. Thus, only the volute is meshed with tetrahedrons and with prisms in the boundary layer. The 5 fluid and 8 solid domains contain 4.67 M nodes, 1.27 M tetrahedrons, 0.72 M prisms, and 3.85 M hexahedrons resulting in 5.84 M elements in total. A detailed view (Fig. 7) of the inlet region shows the O-grid structure of the fluid mesh and the high resolution of the boundary layer. In order to guarantee a sufficient resolution of the temperature gradient [10] the boundary layer has been refined to a dimensionless wall distance of y+ < 1 at all CHT -walls. In figure 7, the / -distribution is plotted for the full-load operating point where the highest velocities occur. The viscous sublayer is sufficiently discretized at all interior walls (fluid/solid interfaces) including hub, shroud, tip clearance, back wall, etc.
239
Y+1.0 O.B
0.6 0.4
Figure 7 Computational grid and y+.distribution Inconsistencies in the y+-distribution are attributed to an uneven thickness of the first grid cell around the twisted turbine blades.
Numerical model and boundary conditions For both designs, the same test data are used as boundary conditions for the operating point at full load and partial load, respectively.
Steady state conjugate calculation The fluid boundary conditions for the CHT-calculations are derived from the gas stand test data. With respect to figure 6 mass flow and total temperature are prescribed at the inlet and static pressure at the outlet. The bypass is closed during the gas stand tests, so that the flap position remains unchanged in all calculations. The number of rotor revolutions corresponds to the operating point. The CHT-method does not require any boundary conditions at interior walls, but at exterior boundaries heat transfer coefficients are given in combination with a reference temperature. The walls in the borehole of the shaft and the heat shield are assumed as adiabatic. Since the temperature at the cutting plane of the shaft is unknown, a special thermal boundary condition is applied at this location: In an iterative process the temperature at this cutting plane is elevated until the measured temperature of thermocouple #4 (Fig. 2) matches the corresponding monitor point of the calculation. The CFD code ANSYS CFX has been used in combination with the low Reynolds kro-SST turbulence model [11, 12]. This model is able to resolve the thermal boundary layer, which is necessary for CHT-calculations. Since different heat transfer rates occur at laminar and turbulent flow, further improvement can be achieved by the use of experimental correlations allowing a better treatment of the transition between laminar and turbulent flow. For CFX a transition model is recommended which solves two extra transport equations, one for the intermittency and one for the transition onset criteria in terms of momentum thickness Reynolds number. The expenses are 18 percent additional CPU time and app. 10 percent more RAM. Furthermore, large aspect ratios in the boundary layer lead to small matrix elements, which require a computation in double precision mode. The available RAM of the Linux Cluster demands to postpone the implementation of a transition model to future investigations or simpler geometries. To avoid a time-consuming transient CHT-calculation, which is often not costeffective in industrial application, a frozen rotor model has been chosen. Hence, the effects of centrifugal and Coriolis forces on the flow are included. Interfaces have been placed at rotor inlet and rotor exit. Viscosity and thermal conductivity of the exhaust gas, as well as material properties of the solid domains are a function of temperature.
240
Transient temperature calculation Since the same grid and the same solver are used for the steady state CHT- and for the transient calculation, the set of boundary conditions is almost identical. Only at the CHT -boundaries the heat transfer conditions have been changed. The conservation equations for mass, momentum, and turbulence are turned off and only the heat transfer equation is solved. This leads to a frozen fluid boundary condition, but the development of the wall temperature is still free. Thus, heat transfer coefficients are still not required. Two boundary conditions are changed during the transient calculation according to functions reflecting experimental experience. Firstly, the rotor is accelerated within 5 s. Secondly, the temperature at the cutting plane of the shaft is increased during 100 s. At the beginning of the cooling/heating process, small time-steps are used because in this period the highest stresses can be expected. From this point on, the time steps are progressively extended until a steady state is reached. The number of inner loops per time step is reduced during calculation, dependent on a residuum criterion. The temperature field is recorded at discrete time points to serve as input data for the succeeding thermal stress analysis.
Thermal stress analysis The FEM-code PERMAS has been used for the stress investigation. The mesh of the turbine wheel is imported with the corresponding temperature field at discrete time points. In the first static stress analysis the temperature field leads to thermal stress. In the second static stress analysis, thermal strain is combined with centrifugal forces which result from the rotation of the wheel. The FEA for the Hookean material works on basis of a linear elastic solid. Thus, it is assumed that the stress is a linear function of the infinitesimal strain. In the future the consideration of non-linear material behaviour could include effects like plastification resulting in a higher degree of accuracy. In this stage of development, the focus is on maximum stress regions, so-called "hot spots". Besides, there are further uncertainties like anisotropic material behaviour and casting defects which are unlikely to be ascertainable in the future. Since the cutting plane at the piston ring is too close to the wheel, the front end of the nose is fixed in tangential and axial direction.
Extrapolation of CHT results As to be discussed in the next paragraph, increasing the mass flow / rotor speed at constant turbine inlet temperature leads to an almost linear decrease of the wheel temperature. Hence, the temperature distribution is extrapolated according to the following procedure: Extrapolation of full load operating point: The full load temperature field is scaled to the elevated rotor speed according to experimental results considering the average gradient of the five measuring points between partial and full load. Calculation of each temperature field at discrete time points during transient heating: According to the temperature drop determined in step one, the heating characteristic of each node is adjusted. New run-up characteristic for acceleration during transient run: Since there is no information about the acceleration time, it is assumed that the new speed is reached in the same timeframe.
241
RESULTS AND DISCUSSION Experiment Exemplary for one turbine inlet temperature, the temperatures in the 5 thermocouples are plotted over 5 measured speeds and mass flows respectively (Fig. 8). A rising mass flow leads to an increased turbine speed, so that more enthalpy is converted into power transferred to the shaft. Since the turbine efficiency rises at the same time, the temperature in the 5 thermocouples decreases. Figure 9 shows the same behaviour for lower turbine inlet temperatures exemplified with thermocouple #1. 680,··························
,·······································r············· .............
r·························
,
660 500+---·-
640-
~ 400
E620 +--c::::-;;-;-r-+~~--"'-6-:C"""A+---....-O'>----i
:;:' 300
600
200
580
---
100
560~~~-+~~~~~~~~~~
o
20000
40000
60000
-----~-
-----~--+~~-----i
o~~~~~~~~~~-+~~~
80000
n [min- 1]
o
20000
40000
60000
80000
n [min- ' ]
Figure 8 Experimental temperature: ttot,in=720°C
Figure 9 Experimental temperature: thermocouple #1
Numerical simulations
Steady state conjugate calculation Some observations concerning the heat transfer in the turbine wheel are discussed for the high load operating point. The occurrence of high temperature gradients shows the effects more clearly than for partial load. In figure 10, the temperature distribution is shown at 50% span. Although a frozen rotor boundary conditions is used, the temperature is equally distributed. Only the tongue of the volute causes some distortions which are mixed out downstream. The blade-to-blade view of the fluid temperature clarifies the high thermal load of the pressure side compared to the suction side. Nevertheless, an equalized temperature distribution is found inside the blade. Along lines of constant radius given in figure 11, static temperature profiles at 6 locations give information about the direction of heat transfer. In figure 12, the tip area along line "A" is heated at both sides of the blade. Downstream, along line "B", the material is heated only at the pressure side and along lines "C" and "D" the blade is cooled on both sides. Further noticeable differences are visible in figure 13. In the rear part of the blade at 90% span the solid is cooled. The strong temperature gradient at the suction side arises from a jet of high kinetic energy in between the suction side and a separation bubble. The much thicker hub area at 10% span (line "F") is bi-directionally heated. However, the temperature gradient in the solid between pressure and suction side is negligible. Even in the thick hub region the differences are less than 3K. In the thin tip region, the high thermal conductivity leads to a gradient of less than 1K. Under consideration of varying wall thicknesses and the scallop shaped back with an undercut at the shaft, the importance of a transient temperature calculation is evident.
242
high thermal load on PS
,
.. /
./
895~
-850
equalized ,./'" "T-distribution in the vanes
~
'.'~~ ~.
,"
~
BCD
A
Figure 10 Temperature distribution at midspan
A
A B C
D E F
%span %stream 50 10 50 40 50 55 50 70 90 55 10 55
Figure 11 Locations of temperature profiles in fig. 12, 13
93°Tr~~===i----------~~----------:
920 910 900 890 ..: 880 870 860 850
-C: 55% stream
-~70%stream
g
f--------------=-='---\-----:------i
PS
A
_
!
_
PS~:.'
Jl[J~"=L Va- -
L7C- -
o
--
----~
840~~~~~~~~~~-r~~~~~~
30
40
50
70
60
80
[0]
~
Figure 12 Static temperature at 50% spanwise (profiles A, B, C, D)
g
930 920 910 900 890
E: 90% span - 0 : 50% span
I
F: 10% span
E
PSr;SS I
--1, ..::;<
..: 880 870 860 850 840
PS
J I
_.__ F
SS
.
5S
'-'
.... 0 L./'
30
40
50
60 ~
70
80
[0]
Figure 13 Static temperature at 50% streamwise (profiles D, E, F) 243
Transient temperature calculation As explained before, the transient solid body calculation serves as an input for the subsequent thermal stress analysis. Nevertheless, some conclusions can be drawn from the temperature calculation alone. The five monitor points given in figure 14 are identical with the measuring positions in figure 2. Points "# 1" and "#5" are located near the exit of the blade at 50% span. Here, the temperature rises considerably faster due to the thin wall. In contrast to that, the temperature increase at the back is delayed because of the long distance to the hot gas mainstream and because of the short distance to the oil-cooled shaft. The duration of the heating process is a result of the transient calculation. The experimental results symbolized by squares at t = 600s show, that the simulation is in good accordance with the experiments. The considerably different temperature gradients during the heating cycle generate further thermal stresses. The magnitude of these stresses in comparison with the stresses by centrifugal forces is subject of the subsequent FE-Analysis. 900
/5
850 800 750
g 700 I650 600 550 500 60
0
120
180
240
540
600
t[s]
Figure 14 Heating cycle: temperature at monitor points #1 - #5 Stress analysis The stress analysis is divided into the analysis of the steady states, and the transition between these states. The evaluation distinguishes between temperature load and thermo-mechanical load, i.e. thermal load combined with centrifugal load. In order to evaluate the influence of each load, the scale is set from zero to maximum stress value observed in the undercut between shaft and back wall. Original design
Since the thermal stress resulting from both steady state CHT calculations is very low, the investigation focuses on the transient part. As aforementioned, a series of linear elastic calculations is performed for discrete time points simulating the heating cycle. During this quasi-transient calculation, the maximum von Mises stress is recorded and illustrated in Figure 15. This artificial state shows local maxima at the fillet between blade, hub, and back. Since the extent of the maxima decreases in consideration of rotation, tensile and compressive stresses interact. The maximum principle stress shown in figure 16 proves the occurrence of compressive stress due to thermal load at the fillet radius between blade and back. The sources of the local stress maxima at the downstream blade-root fillet radius and at the undercut are both tensile stresses. 244
thermal + centrifugal load
thermal load
fillet radius (blade to back)
fillet radius (blade root)
undercut (back)
Figure 15 Maximum v. Mises stress during transient analysis (original design)
thermal load
thermal + centrifugal load
fillet radius (blade to back) ____ compressive stress
fillet radius (blade root) ____ tensile stress
undercut (back) ____ tensile stress
Figure 16 Maximum principal stress during transient analysis (original design) 1.0 0.9 0.8 0.7 ::;: 0.6 1! 0.5 "
o_ _ (Jlii rel~ ,th.:50t I (jrel,th
0.4
0.3 0.2 0.1 0.0
O'M(t): fillet radius
! {
'"
'f
,-/1"-,,, o
I 11
.........
.',
I
th + th
rot _i
60
180
120
o_ _ (J.,".I,thjiHiiot I
--,-----'t[CS];:::;:==--r---"""i"l 0.9 ~ ~~~____h"L.~~+_~~---W=ti:;-:;:r;;;i 0.831.0,t....1
0.7 ... 0.6 ~ 0.5
t-------/9-----t----H
" 0.4 0.3 0.2 0.1 iV-'---'-'--t--~ 0.0
-
O'M(t): undercut
'\-T~~~+-'-~~~j-.-""-:;;-""...,.....,~~~"""
o 60 120 180 240 Figure 17 v. Mises stress from partial load to full load at fillet radius and undercut (original design)
245
<Jrel,th
I 11
The von Mises stress in the fillet radius at the blade root and in the undercut is plotted versus time in figure 17. The upward and downward facing arrows in the diagram belong to the corresponding positions marked in the stress distribution plots. After 12 seconds, the thermal load leads to a tensile stress maximum even in consideration of centrifugal forces. The same analysis method applied to the undercut between shaft and back wall provides different results. The local thermal compressive stress maximum occurring after 12 seconds is shifted to 3 seconds in combination with tensile stress generated by centrifugal forces. Even at the end of the heating cycle, the tensile stress is comparatively high. It should be noted that the thermal boundary condition at the cutting plane of the shaft has an influence on the results and the conclusions drawn in this paragraph. At this location, the temperature is increased moderately according to a heating function given in figure 18 (T shaft ). Moreover, this qualitative plot shows the dependency between rotational loads, thermal loads, and stresses.
n
rotational loads / '
-----~<-
,.-
o
60
stresses
120
180
240
t[s]
Figure 18 Time dependent loads and stresses at undercut (original design) Original design: extrapolated operating point The preceding results have shown a high impact of the thermal loads on the combined thermal and rotational stress. However, the maximum rotor speed was only 316 m/s because of the sensitive measurement technique. To evaluate the importance of thermal loads on the overall stress at higher rotor speed, an operating point at 450 m/s has been derived according to the preceding paragraph "Extrapolation of CHT -results". Even the intermediate results of the transient run are adapted to the new conditions. Again, the investigation focuses on two areas, the fillet radius at the blade root and the undercut from shaft to back. In the style of figure 15, figure 19 shows that the thermal stress maxima recorded during the heating cycle are reduced. In contrast, the centrifugal stress increases from 70.000 to 100.000 rpm by a factor of about 2. The local combined stress maximum at t = 3s almost disappears (figure 20) resulting in a stress curve which is also qualitatively different. Nevertheless, even at the elevated rotor speed, the stress at the surface does not exceed the yield limit. On the other hand, it is exceeded in the bore, which is a special design feature allowing the instrumentation of the measurement technique. However, there is no bore in series parts. In addition to figure 16, the corresponding curves at increased rotational speed are diagramed in figure 21. They show qualitatively the influence of the speed on the thermal and thermo-mechanical loads. 246
thermal load
thermal + centrifugal load fillet radius (blade root) ~ tensile stress
O,el,P ~
compressive
1-
o undercut (back) tensile stress
-1
~
Figure 19 Maximum v. Mises stress: transient run to extrapolated operating point (original design) 1.5
1.0
:;:
J
----
0.5
0.0 <J'rel,th+rot
0
a
I I 11 O'rel,th
60
tl
1.5
120
o_ _ <Ji,el~,thi1+r,ot l
180
t[s]
O'rel,th
O"M(t): undercut 1.0
:;:
J 0.5
0.0
a
60
120
180
240
Figure 20 v. Mises stress from partial load to extrapolated full load at fillet radius and undercut (original design) !
'----
100.000 rpm 70.000 rpm
n
7------thermal + rotational stress
o
60
120 t[s]
180
O"th+rot
240
Figure 21 Time dependent loads and stresses from partial load to extrapolated full load at undercut (original design) 247
I 11
Superback design Does a geometric modification of the back lead to a significant reduction of the stresses? To answer this question, the same simulation procedure as for the original design is performed. Figure 22 shows the modified geometry with a so-called superback. The inclined back arises from the fact that the stress in a rotating disc with a conical-shaped cross section is the same in the thick root as in the thin tip.
original design -
super back design -
Figure 22 Comparison of geometries thermal load
thermal + centrifugal load
fillet radius (blade to
Figure 23 Maximum v. Mises stress during transient analysis (superback design) 1.0 .--~~~'--~~~'--~~r--,-up-e''''''ba''''Ck--:,':-h+"'-ro", 0.9 ~~+-~~~+-~---j- -original, thHot 0.8
-- --- -superback, th
0.7
th
:;: 0.6 .. 0.5 t) 0.4 0.3 0.2 0.1 0.0
~---
t-==:::t:s;;:l;;;;~<=Fr=:rJd~
O'rel,thHot
60
0 _11111 1 (Jrel,th
1.0
0.9
t1
t)
180
o
1------+-,-/_:::==:~~:::::::::::'-_l:~~="CC,u=-:pe='b=-:ac;:-k'''''h+'''ro::'"
_I: - original,
,
th+rot
~superback, th
0.7
0.6 0.5 0.4
----?-L:::t==:::::t;..;,.~-::;··o.~O':i i~,a~"~'hq; 1_''''''''''''---=---+~~~-~ __~~'-il
--+-
0.3
240 t2
'---'--~~'--~~~_-T_~-=_C::_:-==-l'~---"-,
0.8
~f
120 I [s]
~~ g/" ___ r-'
~.~ t-+~)~ 0.0
'1"
' / _ - 1 - O"M(t): undercut
+-'-f,--~~+-.~~~+-.-;:.o;..:,~"""~"",,,......-1
o
60
120
180
240
Figure 24 v. Mises stress from partial load to full load at fillet radius and undercut (superback design) 248
O'rel,th+rot
1 1 1 1 11
O'rel,th
In comparison to the original design (figure 15), the superback leads to a stress reduction of 55% in the undercut region for the analysed transient process (figure 23). In this evaluation, the undercut region of the superback includes the radius from the back to the shaft and it includes the undercut which has moved towards the bearing. The thermal stress in the undercut region of the superback design is slightly higher but Figure 24 shows lower combined stresses. Since the geometry of the blades has not changed, the fillet radius at the blade root is exposed almost to the same stresses as before.
DISCUSSION In further investigations, the modeled shaft length will be extended to the radial bearing. The problem of prescribing suitable thermal boundary conditions could be solved by defining heat transfer coefficients and the oil temperature as a reference temperature at the shaft surface. A further advantage is a more realistic clamped support boundary condition in the FE-Analysis. Since the rotational stress shows no change with the temperature distribution, a nonlinear analysis is necessary to evaluate the quality of the stress results.
CONCLUSIONS The main focus of this paper is to present a feasible optimization method for stress reduction. A combination of experimental data, steady state CHT-calculations, transient temperature computations, and thermal stress analysis has been applied successfully to a turbine wheel. The results are in good agreement with the gas stand test data, as the hollow shaft geometry for instrumentation is included in the model. However, in a series-production turbine wheel there is no bore. Instead of that, a blind hole in the shaft and in the turbine rotor respectively, forms a welding pocket due to friction welding. Furthermore, scratches from balancing, which weaken the wheel, cannot be included by simulation. Nevertheless, the numerical investigations provide some helpful information for the design process. The CHT analysis has shown only small temperature gradients in the blade between pressure and suction side despite large temperature gradients in the surrounding fluid. The transient thermal analysis points out the different heating rates dependent on the location in the wheel. Furthermore, it provides the input data for a subsequent FEA. The thermal stresses playa minor role compared to stresses generated by centrifugal forces but they are not to be neglected. Especially in the area of the undercut between shaft and back and in the area of the fillet radius between blade and hub they are important. At medium rotor speed, the thermal stress can amount 50% of the combined stress. Increasing the speed to a higher level, which is typical for engine application, the influence is reduced to circa 33%. Inclining the back of the turbine wheel back decreases the stress peaks. In the transition area from back to shaft, the combined stresses could be reduced by 45%. Changing the frozen rotor calculation to a complete transient computation would decrease the level of abstraction. However, for such a sophisticated approach a distinct increase in computational power is essential. Otherwise, the calculation will not fit into the workflow of the design process and there will be no benefit from the results.
249
REFERENCES I Schmitt, F., Engels, B., Schreiber, G., 2004, "Regulated 2-Stage (R2S'M) Charging System for High Specific Power Engines", 13 th Aachen Colloquium "Automobile and Engine Technology"
2 Schmitt, F., Weber, M., Gabriel, H., Lingenauber, R., and Schmalzl, H.-P., 2002, "New Investigations of Variable Turbine and Compressor Geometry for Passenger Car Turbocharger Applications", 23. Vienna Motor Symp. 3 Lakshminarayana, B., 1996, "Fluid Turbomachinery", John Wiley & Sons Inc.
Dynamics
and
Heat
Transfer
of
4 Bohn, D. and Bonhoff, B., 1994, "Berechnung der Klihl- und SWrwirkung eines filmgeklihlten transsonisch durchstromten Turbinengitters mit diabaten Wanden", VD1Berichte 1109, pp. 261-275 5 Bohn, D., Bonhoff, B., and Schonenborn, H., 1996, "Combined Aerodynamic and Thermal Analysis of a High-Pressure Turbine Nozzle Guide Vane", Proceedings of the 1995 Yokohama International Gas Turbine Congress 6 Heuer, T., Engels, B., and Wollscheid, P., 2005, "Thermo-Mechanical Analysis of a Turbocharger based on Conjugate Heat Transfer", ASME GT-2005-68059 7 Mc Donnel, G.T. and Roberts, Q.D.H., 2002, "Design of a turbine rotor for a highvane count variable geometry turbocharger", Proceedings of the 1MechE, paper C602101212002, pp.61-74 8 Ludewig, H., 1968, "Schwingungsuntersuchungen an Turbinenschaufeln von Abgasturboladern", offprint MTZ, Vol. 29, pp. 408-414
9 Ludewig, H., 1969, "Schwingungsmessungen an Turbinen-schaufeln von Hochtemperatur-Dehnungsmessstreifen", offprint Abgasturboladern mit Messtechnische Briere 111969 10 Schlichting, H., Gersten, K., 1997, "Grenzschicht-Theorie", Springer-Verlag, BerlinHeidelberg
1 I Menter, F.R., 1993, "Zonal Two Equation k-ro Turbulence Models for Aerodynamic Flows", A1AA paper 93-2906 12 Bardina, J.E., Huang, P.G., and Coakley, T., 1997, "Turbulence Modeling Validation", A1AA paper 97-2121 © BorgWarner Turbo Systems Engineering GmbH
250
DYNAMICS OF MISTUNED RADIAL TURBINE WHEELS X Sheng, DC Clay and J Allport Holset Engineering Co Ltd, St Andrew's Road, Huddersfield, HDI 6RA, England This paper presents investigations carried out at Holset into the dynamics of mistuned radial turbine wheels, including a literature review, a lumped parameter model, identification of the most responsive blade, distribution of the peak maximum order response and a method of mistuning identification. INTRODUCTION AND A LITERATURE REVIEW
Mistuning in bladed wheels is the phenomenon of random and unavoidable blade-toblade variations in geometry and material due to the casting process of the wheels. A mistuned wheel may exhibit vibration localisation and amplification, in which few blades have responses much greater than those of other blades and the tuned response. These mistuning effects not only significantly reduce the high cycle fatigue (HCF) life of the wheel, but also make it difficult to predict and measure the representative response. Mistuning has been studied for over 30 years. Mistuning may be investigated from a statistical point of view due to its random nature. In stochastic structural dynamics, efforts have been made to use the perturbation method to establish analytical relationships between probability density functions of random parameters in a structural dynamic system and those of the required outputs (e.g. natural frequencies and forced responses) [1, 2]. Perturbation method requires no natural frequency of the unperturbed (tuned) system is repeated. This is not the case for a tuned bladed wheel. An alternative to this approach is to derive a much reduced (compared to a conventional finite element) model of high computational efficiency with part of the model parameters being random variables. The distributions of the random parameters are estimated from measurement and statistical results are produced from the reduced model by performing a large number of calculations. Different techniques have been developed to produce a reduced model. Refs [3,4 and 5] represent a mistuned system using a lumped parameter model consisting of masses and springs. Such a model has a high computational efficiency but is difficult to capture vibration of high modes. Refs [6, 7 and 8] model a mistuned system using the component mode synthesis technique in which the first few modes, which are produced using FEM, of each component (substructure) of the system are employed to approximate vibration of the whole system. In addition to the substructural modes of a mistuned system, modes of the corresponding tuned system have also been used to synthesise the vibration of the mistuned system [9]. Other efficient modelling approaches are also attempted [10, 11]. Though statistics may be performed using a reduced model, mistuning identification for individual mistuned wheels is still desirable. Mistuning is identified normally from some, often inadequate, measured data, leaving the problem indeterminate. To obtain a set of unique parameters, extra conditions must be assumed. Two mistuning identification methods are suggested in Refs [12] and [13] in which blades are represented by multi-mass-spring systems coupled with each other and the hub (disk) is rigid and fixed. The first, termed the random modal stiffness approach (RMS), assumes mistuning exists in the stiffuess matrices of the blades only and the mistuned modal shapes are not significantly different from the tuned ones. Under such conditions, the mistuned stiffuess matrix of a blade can be determined straightforwardly by measuring 251
all the mistuned blade frequencies (not the bladed disk system frequencies). The second is tenned the maximum likelihood approach (ML). In this approach, both the mass and stiffuess matrices of a blade are allowed to be mistuned. By assuming a joint probability density function for the stiffuess and mass matrices, the mistuned mass and stiffuess matrices are estimated under the condition that they give the measured blade frequencies while at the same time make the probability density function maximum. It is illustrated that the ML approach works better than the RMS approach in tenns of the mean and standard deviation of the maximum forced vibration of the blades subject to order excitations. The difficulty in using the ML approach is how to choose correctly the type of the joint PDF and its parameters. It is also the fact that this method is based on the blade alone frequencies which may not be measurable for turbocharger turbine wheels since the blades cannot be removed from the hub. Another mistuning identification method, presented in Refs [14] and [15], is based on the fundamental mistuning model developed in Ref [11]. There are four assumptions in the fundamental mistuning model: a) only a single, isolated family of modes will be excited; b) the strain energy of that family's modes is primarily in the blades; c) the family's natural frequencies are closely spaced; and finally d) mistuning is small. There is another assumption in the model which has not been stated explicitly: modes in this family can be approximated by a weighted sum of a number of tuned modes. To identify mistuning, the mistuned frequencies and modal shapes of that family must be measured. Once the mistuning is detennined, the model can be used to predict responses to order excitations which mainly excite that family of modes. In other words, the excitation frequency must be close to the average of the frequencies of that family of modes. The challenge of using this method lies in accurate measurement of the mistuned modal shapes. This is true because the natural frequencies are closely spaced. Mistuning effects such as vibration localisation and amplification are demonstrated in some of the aforementioned Refs [e.g. 8] using the developed model. It is found that mistuning effects are particularly prominent near the so-called veering of the corresponding tuned system. Though the veering phenomenon is to be explained, the vibration localisation phenomenon is interpreted in a review paper [16] using the stability theory. It is numerically demonstrated in Ref [17] that intentional mistuning may be used to depress some of the negative effects of the random mistuning. Though effort is made in [17] to explore the mechanism of intentional mistuning, a satisfactory explanation has not yet been achieved. A possible route to this may be the use of singular value decomposition, as shown in Ref [18]. It should be realised that previous research is mainly concerned with aero engine turbofans and only a little is on small radial flow turbines such as those in Rolset turbochargers. Small radial flow turbines exhibit different dynamic behaviour from large aero turbofans and the effect of mistuning on the fonner is expected to be also different. It is also the fact that previous research has not considered the effect of centrifugal and Goliolis forces generated from the wheel rotation. A SIMPLE MODEL FOR RADIAL TURBINE WHEELS
FE modal analysis
For a cyclically symmetric structure consisting of N identical sectors, vibration modes can be grouped by the nodal-diameter (ND) number n, where n = 0,1, 2,A N / 2 if N is even or n = 0,1, 2,A (N -1) / 2 if N is odd. In tenns of a cylindrical coordinate system, the mode shapes, u rand u r+l, of the rth and (r + 1) th sectors are related by 252
u r +1 =ein2Jf(Nur for the n-ND modes [19], wherei =..r:l. This shows a travelling wave pattern which extends to all the sectors. A natural frequency associating with a nodaldiameter number different from zero and N / 2 (where N is even) repeats itself. Mistuning in general invalidates these modal properties. To assist the development of a simple, lumped parameter model so that some investigations can be performed, natural frequencies and modal shapes are calculated using Ansys Workbench for a 12-blade turbine wheel (Fig 1). In the calculation the shaft is cut off from the weld boss and the weld boss is constrained in the axial direction. The calculated natural frequencies at 21°C are listed in Table 1 excluding the three zero-frequencies of rigid modes. The first two modes are I-ND modes in which the hub rocks about a weld boss diameter. The frequencies of these two modes are denoted by IT! = wT! /21<. Frequencies of the third to eleventh modes correspond to nodal-diameter numbers from 2 to 6. They are close to each other (within 10 Hz, as the FE mesh gets finer, the discrepancies between these frequencies becomes smaller). These frequencies associate with modes in which the blades vibrate almost independently of each other and the hub stays motionless. This suggests that in these modes couplings between blades are weak and negligible. The average of these frequencies may be taken to be the first (cantilevered) blade frequency, denoted by IB = wB/21< . The twelfth mode, its frequency denoted by IR = wR/21< , is the first 0ND mode in which the hub rotates like a rigid body and the blades vibrate in-phase and at the same amplitude. The next two modes are still I-ND modes and the frequencies are denoted by In = wn /21< . Further higher modes (e.g. the 15 th and 16th modes in the table) have much higher frequencies and are not interested in this study.
Modes
1 (l-ND) 2 (l-ND)
3 4 5 6 7 8
Table 1 Natural frequencies of a 12-blade turbine wheel Frequency (liz) Frequency (liz) Modes
5725 5725 6423 6423 6425 6425 6427 6427
9
10 11
12 (O-ND) 13 (l-ND) 14 (l-ND) 15 16
Figure 1 A turbine wheel and its lumped parameter model
253
6430 6430 6433 6748 6940 6940 11510 11510
A lumped parameter model The FE calculation suggests that, for excitation at frequencies around or less than the blade frequency the wheel may be modelled as a rigid disk connecting a number of point masses through springs, as shown in Figure I. The jth blade is represented by mass, m j , connected to the disk by a spring of stiffness k i , where j = 1, 2, ... N. The position connecting the spring at the disk is described by the radius Rand anglea j measured from the x-axis. The disk is described by mass M and inertial moment 1. The disk is subject to stiffness, k x and k y , which model the rocking stiffness of the hub, and torsional stiffness, kg, from the shaft. Blade masses are restrained to vibrate in directions tangential to the disk periphery, so that the model has N + 3 degrees of freedom. The displacement of mj relative to the disk is denoted by Xj . Displacement relative to the hub is used since it is proportional to the strain/stress in the blade. The displacements of the mass centre of the disk in the x- and y-directions are denoted by x and y and the vibrational rotation angle of the disk by B. Without considering the spinning of the wheel, the differential equation of the model can be written as Mq+Kq =Q (1) where, (2) q = (x,y,B,xl'x 2 ,···,X N /
Q denotes the generalised force vector, M = (mu )i.j=I,2 .. N+3 and K mass and stiffness matrices with none-zero elements given by, N
mil =mn =M
are
N
+ Lmi
, m"
=]
i=1
(3a)
+R2Lmj j=1 N
N
m13 =-RLmjsina j
m 21.
,
= R" L..., m 1. cos a 1.
(3b)
i=1
i=1
m ji
=(k ij )i,j=I,2 ... N+3
= m j _3 , m lj = -m j _3 sin a j _3 , m 2j = m j _3 cos a j _3 , = k i - 3 , kll = k x, k22 = k" k33 = kg
m 3j
= Rm j _3
(3c)
(3d) k ji where j = 4,5, .. ·,N +3. For the nth order excitation at frequency OJ and of unit amplitude, the jth blade mass
is subject to a force defined by e infXj e iM . Thus in equation (I), Q = Qe iilX , and N
. . . -_ ( _" Q L..., e inrx
N j
j=1
•
.
".
SIn ai' L..., e i=1
T
N
ina]
cos ai' R"" L..., e ina '.
j
,e
ina}
,e
inaz . ..
,
e
inaN
)
(4)
j=1
Determination of model parameters for a tuned wheel For the model shown in Fig 1 to be equivalent to a real bladed disk, the model parameters must be determined correctly. The blades may to assumed to be uniformly spaced, therefore a i is given by
aj
= (.i-l)}J
(j=1,2, .. ·N)
(5)
The torsional stiffness of the shaft, kg, may be estimated from the shaft dimensions and material properties. For a tuned wheel in which the blades are identical to each other,
254
there are only six parameters to be determined, which are M, J, R, kx = ky, andm 1 = mj (j = 2,3A N). To determine them, natural frequencies of the O-ND and l-ND modes of the lumped parameter model and the blade frequency are derived. They must be equal to the counterparts from the FE model and this gives four conditions for parameter determination. Now at each blade mass, a unit harmonic force at frequency OJ is applied in the tangential direction. The displacement amplitude of the blade mass, observed from the ground, can be derived. A forced vibration FE analysis is performed under the same loading condition, i. e. a unit harmonic force at frequency OJ is applied at each blade tip in the direction normal to the blade surface. From the analysis the normal displacement amplitude, A1 , of the blade tip can be worked out. The displacement amplitudes from the two models are equalised to provide the fifth condition. Finally a torque of unit amplitude and frequency OJ is applied at the disk and the displacement amplitude of the blade mass observed from the ground is derived. This amplitude must be equal to the blade tip normal displacement amplitude, A2 of the FE model under the same excitation, so that the sixth condition is produced. From these six conditions it can be shown
k1
=k j
(6) (7) 2 2 -1 OJT~;T2
J- -11 1
(8)
(9)
i.e. the O-ND modal frequency is higher than the blade frequency. DETERMINATION OF THE MOST RESPONSIVE BLADE The displacement of a blade relative to the hub due to an order excitation is termed the order response which is proportional to the strain/stress in the blade. For a tuned wheel, the blades produce the same order response. This is not true for a mistuned wheel. There must be a blade experiencing the maximum order response of all the blades. As excitation frequency changes the maximum order response changes as well and may occur at a different blade. The maximum of the order responses of all the blades over all the excitation frequencies is termed the peak maximum order response (PMOR). It is important to determine which blade has the PMOR. If parameters of a mistuned wheel are completely known, a simple calculation can give the answer. Thus mistuning identification is practically important and will be dealt with below. An alternative is to use statistics. If a blade with a particular feature has a high probability (e.g. more than 90%) of having the PMOR and this feature can be easily identified, then the most responsive blade can be practically labelled. To do so, blades in a wheel must be numbered according to the blade feature. Blades from different wheels have the same number if the blades possess the same feature. 255
To number the blades of a mistuned wheel, measurement is performed for the tip-totip frequency response function matrix, denoted by H(co) , where, co is the radian frequency. This is a N x N matrix and is constructed from the normal displacements of the blade tips observed from the ground due to a unit normal point force at the same or another blade tip. Responses of the blades, denoted by q e (co) and calculated from q~(co)f = H(co)(e ina1 ,e ina2 ,A e inaN f (10) are termed the experimental order responses of the blades subject to the nth order excitation. Due to the vibration of the hub, an experimental order response is in general different from the corresponding order response which is observed from the hub. The maximum of the experimental order response of a blade over a frequency range is then determined and compared to those of other blades. Blades are then numbered based on their maximum experimental order responses: the higher the maximum experimental order response is, the lower is the blade number. Based on this blade numbering method and the lumped parameter model, calculations are carried out for a large number (3000) ofmistuned wheels to produce the occurrences for a blade to have the PMOR. The corresponding tuned wheel is the one shown in Fig I and Table 1. A loss factor of 0.005 is estimated for material damping. Mistuning is described by random blade frequencies which follow a normal distribution with the mean being 6427 Hz and the standard deviation 170 Hz. The probability of a blade to have the PMOR is shown in Fig 2 for five (the 4th , 5th , 6th , 7th and 8th) excitation orders. It can be seen that the No. 1 blade has more than 95% probability to have the PMOR. In other words, the blade having the peak maximum experimental order response almost certainly has the PMOR. It must be addressed that, the blade having the PMOR usually changes as the excitation order changes. qe(co)
= (qje (co),q; (co),A
'~It 1=. 1
2
: : . : ' ,~:J e
3 4 5 7 B 9 10 11 Maximum experimental order response (descending order)
12
Figure 2 Probability of a blade to have the peak maximum order response
DISTRIBUTION OF THE PEAK MAXIMUM ORDER RESPONSE Fig. 3 shows the PMOR of the 3000 wheel samples. To determine which distribution best fits the PM OR realisations, probability plots are produced using Minitab for a number of distributions. It is assumed that for any mistuned wheel, the frequency at 256
which the PMOR occurs will be excited at a 100% chance. This indicates that statistics should be performed for the PMOR against mistuned wheel samples. It is found two distributions fit the data well: the 3-parameter lognormal distribution (Fig. 4 for the 4th order excitation and a factor of IxlO-6 has been dropped) and the 3-parameter Gamma distribution (not shown here). The 3-parameter Weibull distribution is not a good fit, though it has been suggested for aero engine turbofan in some publications [e.g. 8].
1000
e
~ 500~--~~--~~--~~--==--~~--~. B 0 ~
i
1000'-~~~~~~~~==~~~~~~~
!
i
500
.; 1000
'---------O,-o-------C,Oc-----~c___--~----~--___,,J_
°L---;===;~~==.~~~=.';~~;;;;O=~~~~
::I
1000 1000
-
1500
OrderB, Max = 910.401; Mean
500
1000
2000
2500
=701.7645; SDV _ 59.2497 (x 18-6 m)
1500
2500
2000
3000
Reallutlons
Figure 3 Peak maximum order responses of the wheel samples Probability Plot of Order 4 3-Parameter Lognormai- 95% CI
99.99
l.oc Sca~
Thr...
~
1l
i
99
N
95
AD P-Value:
5.913 0.1585 329.9 3000 0.+41
.
80 50 20
0.01-'--'fL-=-----'----+--~--f__~_+_-'-_+_---'-_t_'
200
300
400
500
600
700
Order 4 - Threshold
Figure 4 Probability plots for the peak maximum 4th order responses of the wheel samples: 3-parameter lognormal
MISTUNING IDENTIFICATION The FE equation of motion of a mistuned bladed wheel may be written as (M + .MI)q + (K + M<:)q =Qe i..
257
(11)
where, matrices M and K represent the tuned mass and stiffness matrices while ~M and ~K denote mistuning in mass and stiffness. The order of the equation is reduced by expressing mistuned vibration in terms of a number of N tuned modes, i. e. q=['P1'P2A'PN]~eiwr=tI>~eiwr
(12)
where,tI> = ['P1'P2 A 'PN]is a matrix formed by the tuned modes and ~is the weight vector. An assumption has been made here that the mistuned vibration is still within the sub-space spanned by these tuned modes. Substituting Eq. (12) into (11) and premultiplyingtl>H , yields - w 2 (M* + tl>H ~MtI»~ + (K* + tl>H ~KtI»~ = tl>H Q (13) where M* and K *are diagonal matrices formed by the modal mass and stiffness of the tuned system. If mass and stiffness matrices mistuning (i.e. tl>H ~MtI>andtl>H ~KtI» is known, then forced vibration can be determined by solving this equation. To identify mistuning, use is made of the tip-to-tip frequency response matrix, H(w) , defined above. Define the force vector of an order excitation: F(t)=(F;,F2 ,A FNfe iwr = Fe iwr (14) in which the forces are applied at the blade tips and normal to the blade surfaces. The global force vector, Q, in Eq. (13) can be generated accordingly. The normal displacement amplitudes of the tips due to this order excitation can be calculated using qe = HF (15)
qe = ({g, q; ,K q~)T is the vector of the normal tip displacement amplitudes. According to Eq. (12), the tip displacement vector, (us, v s' Ws f of the sth blade, is
where,
given by (16) where, !PI(~' !P~~ and !pj~ are the x-, y- and z-components of the tip displacement of the sth blade in the mth tuned mode. If the directional cosines vector of the tip normal is denoted byas ' then from Eq. (16) (17)
In Eq. (17), the term on the left hand side is
q:. Denoting
a sm =a:(!PI<:!,!p~~,!p~~f then Eq. (17) becomes N
L,asmfJm
= q:, or
A~
(18)
= qe
(19)
m=1
from which ~=
A-Iqe
=
~
can be virtually expressed as
A-IHF
(20) Using Eq. (20), J3 can be calculated for different excitation frequencies and orders.
Two excitation frequencies, regular matrices. Let
WI
and w2 ' are chosen so that H(wl ) and H(w 2 ) are (21)
258
where, j
= 1, 2,A Nand
rj
is a number which mayor may not be an integer. The
orders, rj , should be chosen in such a way that the matrix corresponding global force vector is denoted by frequency and
Q (w2 ) j
[Fp F2 ,A FN]
Qj (WI)
is regular. The
for the first excitation
for the second. Two N x N matrices thus can be formed
B2 =A- IH(w2)[FI,F2,A FN ] (22) They are regular matrices. Inserting Eq. (22) into (13) yields equations from which
F
N ],
2
QN(WI)]B~I_[QI(W2)A QN(w2)]B;I}
;H 2 {wi[QI(WI)A W2 -
WI
(23a)
QN(WI)]B~I_WI2[QI(WJA QN(W2)]B;I}
(23b) Now Eq. (13) can be used to predict order responses of the blades in a mistuned wheel. Example calculations (not presented here due to page limit) show that the mistuning identification method is able to produce accurate order response predictions for frequencies around and below the blade frequencies. CONCLUSION This paper presents research performed at Holset into the dynamics of mistuned radial turbine wheels. Based on FE analysis, a simple lumped parameter model is developed for mistuned turbine wheels. Using this model, it is shown that the blade having the peak maximum experimental order response will at a probability of more than 95% have the peak maximum order response, thus the most responsive blade in a given wheel can be practically identified for stain gauging. Peak maximum order responses of mistuned turbine wheels are found to follow the 3-parameter lognormal distribution or 3-parameter Gamma distribution, rather than the Weibull distribution. A mistuning identification method is also proposed based on the tip-to-tip frequency response function matrix and this method is shown to produce accurate order response predictions. It should be pointed out that the effect of wheel rotation and temperature has not yet been considered. This must be one of the topics of further work since turbocharger rotors run at very high speeds and the turbine wheels are driven by high temperature gases. Further work will also be directed to other aspects: for example, the mechanism making a tuned wheel so sensitive to small mistuning, designs of intentional mistuning and explanation of the veer phenomenon, etc. REFERENCES 1 C S Manohar and R A Ibrahim, 'Progress in structural dynamics with stochastic parameter variations: 1987 to 1998', Applied Mechanics Review, 199952177-197. 2 S Adhiakri and M I Friswell, 'Random eigenvalue problems in structural dynamics', 45th AIAAlASME/ASCE/AHS/ASC Conf Structures, Structural Dynamics & Materials, Palm Springs, California, USA, April 2004.
3 D J Ewins, 'The effect of detuning upon the forced vibration of bladed disks', Journal of Sound and Vibration, 1969965-79. 259
4 J H Griffin and T M Hoosac, 'Model development and statistical investigation of turbine blade mistuning', ASME Journal of Vibration, Acoustics, Stress and Reliability Design, 1984106204-210. 5 A V Srinivasan, 'Flutter and resonant vibration characteristics of engine blades', ASME Journal of Engineering for Gas Turbine and Power, 1997 119742-775. 6 R Bladh, M P Castanier and C Pierre, 'Component-made-based reduced order modelling techniques for mistuned bladed disks-Part I: Theoretical models', ASME Journal of Engineering for Gas Turbine and Power, 2001123 89-99. 7 R Bladh, M P Castanier and C Pierre, 'Component-made-based reduced order modelling techniques for mistuned bladed disks-Part II: Application', ASME Journal of Engineering for Gas Turbine and Power, 2001 123 100-108. 8 R Bladh, C Pierre, M P Castanier and M J Kruse, 'Dynamic response predictions for a mistuned industrial turbomachinery rotor using reduced-order modelling', ASME Journal of Engineering for Gas Turbine and Power, 2002 124311-324. 9 F Moyroud, T Fransson and G Jacquet-Richardet, 'A comparison of two finite element reduction techniques for mistuned bladed disks', ASME Journal of Engineering for Gas Turbine and Power, 2002 124 942-952. 10 E P Petrov, K Y Sanliturk and D J Ewins, 'A new method for dynamic analysis of mistuned bladed disks based on the exact relationship between tuned and mistuned systems', ASME Journal of Engineering for Gas Turbine and Power, 2002124586-597. 11 D M Feiner and J H Griffin, 'A fundamental model of mistuning for a single family of modes', ASME Journal of Turbomachinery, 2002 124 597-605. 12 A J Rivas-Guerra, M P Mignolet and J P De1or, 'Identification of mistuning characteristics of bladed disks from free response data-Part 1', ASME Journal of Engineeringfor Gas Turbine and Power, 2001123 395-403. 13 A J Rivas-Guerra, M P Mignolet and J P De1or, 'Identification of mistuning characteristics of bladed disks from free response data-Part II', ASME Journal of Engineering for Gas Turbine and Power, 2001 123404-411. 14 D M Feiner and J H Griffin, 'Mistuning identification of bladed disks using a fundamental mistuning model-Part I: Theory', ASME Journal of Turbomachinery, 2004126150-158. 15 D M Feiner and J H Griffin, 'Mistuning identification of bladed disks using a fundamental mistuning model-Part II: Application', ASME Journal of Turbomachinery, 2004 126 159-165.
1600 Bendiksen, 'Localisation phenomena in structural dynamics', Chaos, Solitons and Fractals, 2000 111621-1660. 17 M P Castanier and C Pierre, 'Using Intentional Mistuning in the Design of Turbomachinery Rotors', AIAA Journal, 2002 40 2077-2086. 18 M S Shahruz, 'Elimination of vibration localisation in mistuned periodic structures', Journal of Sound and Vibration, 2005 281 452-462. 19 D L Thomas, 'Dynamics of rotationally periodic structures', International Journal of Numerical Methods in Engineering, 19791481-102. © Holset Engineering Co Ltd 260
IMPROVING ANALYSIS CAPABILITY IN ORDER TO REDUCE TURBINEHCF S T Kitson, D C Clay, D H Brown, R 0 Evans, D M Eastwood and P K Tootill Holset Engineering Co Ltd, Huddersjield, u.K. SYNOPSIS
This paper will describe the analysis and experimental methods used to understand and reduce turbine end High Cycle Fatigue (HCF) of a fixed geometry turbocharger. The turbine housing flow field is investigated using a computational fluid dynamics (CFD) code and subsequently compared to detailed pressure measurements around the volute exit. This study resulted in a new housing design which showed an improvement to the measured strain value. The paper then describes how an analysis technique was developed to integrate the aerodynamic and structural analysis of the wheel in order to predict the blade strain distribution at the mode I frequency. These predictions are successfully compared against the test data at two speeds relating to the different orders ofthis mode for both the old and new turbine housing stages. INTRODUCTION Background
Radial flow turbochargers are susceptible to blade vibration which can then lead to High Cycle Fatigue. High Cycle Fatigue is particularly prevalent on turbine wheels where the aerodynamic forces acting on the wheel induces the blades to resonate at their natural frequencies. When a wheel blade resonates it deflects from its natural shape. This deflection creates an alternating strain component which is highest at the point of maximum rate of change of deflection. If the strain is high enough the three stages of fatigue fracture will be induced, Stage 1 being initiation, Stage 2 being propagation and Stage 3 being final rapid fracture. This rapid fracture generally results in a piece of blade detaching, making the whole rotor system go out of balance and causing turbocharger system failure. This whole process is called High Cycle Fatigue because the blades resonate at a very high frequency (around 10,000 Hz). This is to distinguish it from Low Cycle Fatigue (LCF) where the alternating stresses are caused by the wheel speed varying due to the operation of the engine, e.g. in a bus application where the vehicle is accelerating and slowing down constantly. Technique Development
The blades deflect in a number of ways, from a basic bending of the blade to very complicated bending and twisting modes, the frequency increasing as the modes get more complex. The techniques developed can be applied to any mode of vibration, however, in this study the results are concentrated on mode I. The aerodynamics of a particular turbine stage were investigated to understand the measured high strain values and failures on test. This investigation took the form of an experimental study where the pressure distribution around the turbine volute outlet was measured at every degree. Traditionally this had been achieved by a series of pressure tappings in the turbine housing [1]. Unfortunately it is difficult to get sufficient 261
resolution, particularly around the tongue area where the pressure changes most rapidly. A new experimental technique was therefore developed which resulted in much better angular resolution and a more reliable pressure variation. The measurement technique gives a good description of the form of the excitation being imparted on the blade, however, a more analytical approach was needed in order fully understand the flow field and investigate alternative designs. The output from CFD calculations was compared against the detailed pressure measurements to give confidence in the tool before design modifications were investigated. The process of comparing designs using aerodynamics parameters e.g. pressure, around the periphery of the wheel is a very useful technique to enable successive iterations to home in on an optimised design. The aim being to minimise the disturbance in the flow that the wheel sees as it rotates around the housing. Reducing this disturbance gives the added bonus that the more uniform flow field leads to lower pressure loss in the system and hence higher turbine efficiencies. This investigation of the turbine housing characteristics takes no account of housing rotor interaction or the damping properties of the wheel. To make an assessment of rotor system life a calculation is needed which gives an overall vibration or strain level on the blade. This involves a much more complex aerodynamic solution of the turbine housing and wheel. The combined solution of housing and wheel requires a method of transferring the aerodynamic loading to the structural analysis. Different types of structural analysis were performed to determine the most efficient analysis in terms of accuracy and computing time. In order to validate the computational method a technique was used to directly measure the strain on a running turbocharger at different orders. Experience has been built up, correlating these measured strain values with field experience. The result of this is that design guidelines can now be set, such that target values can be aimed at during the design and development of the turbine wheel or housing. MEASUREMENT OF VOLUTE OUTLET PRESSURE DISTRIBUTION
An experimental technique was developed to enable the measurement of static pressure at the volute exit / wheel entry location. The technique uses a bladeless plug in place of the turbine wheel. The plug is essentially a bladeless wheel, having a hub section only (Figures 1). The plug is rotated automatically by a motor through a gearing arrangement and pressure tappings on the plug measure the static pressure as it is rotated around the volute. Four pressure tappings are located at 90° around the plug at different radii. The four resulting pressure traces can be compared for data accuracy. An additional check is made by over-rotating the plug by 20° so that each pressure trace has an overlap region. Pressure measurements have been carried out with and without a wheel [2]. This showed that the only the average level changed (due to the different swallowing capacity with and without wheel). The fundamental shape of the pressure distribution remains unchanged as this is determined by the volute design. During each test the plug is rotated through a full revolution whilst passing compressed air through the turbine volute. The turbine inlet conditions are varied and the pressure distribution measured for a variety of steady state running points. The resulting measured pressure distributions are Fourier transformed to extract the relevant orders of interest.
262
PLUG ROTATES IN DIRECTION OF FLOW
Figure 1 Rotating Plug in Housing Assembly This technique has been used successfully to measure the pressure distribution around various turbine housings and assess the forcing function responsible for turbine blade vibration. The process gives results very quickly and can be used as a screening for designs before performing more expensive engine testing.
CFDMETHOD The time dependent flow field and hence, the fluctuating pressures on the turbine blade surfaces is calculated using the computational fluid dynamics (CFD) code 'Newt' originally written by Prof. W.N. Dawes [3]. This is a time marching, explicit RANS (Reynolds Averaged Navier-Stokes) solver which is essentially an extension of the famous 'Dawes Code' - BtoB3D, using an unstructured tetrahedral mesh to allow the analysis of arbitrary geometries. It contains a low Reynolds number k-E turbulence model and semi-implicit residual smoothing. Newt has been used extensively at Holset for housing-only analysis to determine the static pressure distribution at what would be the turbine wheel leading edge position. This is a relatively straightforward problem to solve with no rotor-stator interaction. The flow is allowed to leave the computational domain via a 'bladeless hub' cavity mirroring the test work done with plug tappings. To simulate the performance of the entire turbine stage raises the simulation to a higher level of size and complexity. However, Newt is well conditioned and robust enough to handle this increase with relative ease. The connection between the stationary and rotating components is handled by a sliding mesh interface. This uses distance weighted nodal interpolation to preserve all information on crossing the interface. There is no circumferential averaging, or similar, which would lead to a loss of resolution across the boundary. The housing and full wheel simulation contains 3 million mesh cells. By arrangement with Prof. Dawes, Holset has access to the source code of the Newt solver. This has allowed us to customise the code in a number of areas and specifically to interface with the ANSYS structural solver. Two meshes are used in the overall simulation, the structural mesh and the fluids mesh. At the blade surface, the boundary between the two meshes, the nodal values of
263
pressure are superimposed from the fluids mesh to the structural mesh using a 'nearest node' approach. Because the resolution of the fluids mesh is much greater than the structural mesh there is no need for an interpolation step. The largest distance to the nearest node is generally of the order of O.OSmm, this compares to the blade height at the exducer of37mm. For this problem a 'one-way coupling' between the fluids and structural solver is used, i.e. we do not take account of the deformation of the structure by the fluid loads and hence, any change in the flow field that may arise form this deformation. It is generally agreed that the deformation of the blades is too small to have any meaningful effect of the flow field when compared to the large scale effect of the wheel-housing interaction. The boundary conditions at the stage inlet and exit are taken from test data at conditions giving maximum strain at a given wheel speed. The solid walls are modelled with a representative surface roughness and have the adiabatic heat transfer condition imposed. It has been found that a temporal resolution of 1 time-step = 1 degree of wheel revolution gives a satisfactory solution quality whilst maintaining solver stability. When the solver has settled into a regular periodic state (what one would called 'converged' if this were a steady state solution) which is after 2 to 3 full wheel revolutions, the solver output is turned on and pressures at every node on the wheel surface are written to file after every time-step. Thus we have 360 pressure values for each node describing one complete revolution of the wheel. For each node on the structural mesh the pressure v. time signal is then transformed into the frequency domain via a fast Fourier transform (FFT) to extract the mean load and the real and imaginary parts of the Fourier component of interest. COMPARISON OF PREDICTED AND MEASURED VOLUTE OUTLET PRESSURE AND FOURIER ANALYSIS
For housing A the CFD and measurements were run at the same conditions at inlet (total pressure, temperature and mass flow). The measured downstream pressure from the rotating plug was then compared to the prediction, Figure 2a Although the mean levels are not exact the CFD is predicting the shape of the distribution reasonably well and captures the acceleration under the tongue and pressure rise through the tongue. The repeatable cyclic nature of the pressure field can be approximated to a series of Fourier components of sine waves with an amplitude and phase angle. pee) = Ao + L: Ansin(ne+b n) Where; Ao is the mean pressure An is the amplitude of the nth Fourier component n is the order number e is the angle around the housing bn is the phase angle of the nth Fourier component When the wheel speed is such that its' multiple equals one of the wheels natural frequencies then forced resonance will occur, i.e. the force is pushing and pulling the blades at the exact frequency to cause the amplitude to grow. This multiplier determines the order of excitation. For example if the wheel is rotating at 120,000 rev/min then Sth order resonance will occur if the blade frequency equals (l20,000/60)xS = 1O,000Hz, i.e. the Sth order Fourier component of the pressure field is providing the alternating
264
force at the resonant frequency. The amplitude of the pressure distribution and the Fourier components is a key contributing factor to blade deflection and HCF. The comparison between the first 10 predicted and measured Fourier components of volute outlet pressure from Housing A is shown in Figure 2b. The trend of decreasing amplitude with increasing order is shown for both predicted and measured. The comparison looks quite favourable for most orders, although 1S\ 4th and 5th are over predicted by approximately 30%. It is important to recognise that the amplitudes of the Fourier components that affect strain within the running range is small compared with the average pressure. For example, the 5th order measured Fourier amplitude is only 1200 Pa compared to the average of 164000 Pa. This means that even to get within 30% accuracy on the alternating loads at the tip of the blade the CFD code has to be capable of predicting the pressure distribution to within 0.25% of the mean. 175000
/, A
170000
N <
~ 165000 ~
r~
I!!
-"" I~
:::I
= 160000
r\
""'\.
\ I.------'
I!! a.
'\
155000
·150
·120
-90
·60
/"'"
V
'\
~
·30
t-'"
V
V
tt
./
"'-
~ """
i"-E ~erl! l1.ent
1\ 1 \.Z
150000 ·180
~
i'
0
PI edictl d 30
60
90
120
150
180
Angle from Tongue (oJ
Figure 2a. Comparison of Measured and Predicted Pressure Distributions - Housing A 4500 4000 ~
~
~
3500
~ 3000 (I)
~ 2500
a.
~ 2000 ~
ill 1500 ~
r-- .
~
c.. 1000
r--
I--
500
r--
I--
.-
-
~.
~.
~.
~.
r--
o 2
3
4
oCFD
I
5
r;;! Measured
J
:00 rn JJa 6
7
8
9
10
Order
Figure 2b. Comparison of Measured and Predicted Fourier Components - Housing A 265
TURBINE HOUSING RE-DESIGN During performance testing at Holset, a particular turbine wheel experienced early life HCF failures when combined with housing A. Measurement of dynamic strain in the turbine blades confirmed high values. CFD results revealed a large variation in the pressure and mass flow entering the wheel. The volute was re-designed to smooth out the variations in the pressure and mass flow fields at the wheel entry location. Particular attention was given to the design of the tongue and the area schedule was developed by using the Centre Vortex method [4]. Analysis of the final re-design (Housing B) showed that the pressure and mass flow distributions were significantly smoothed out, particularly around the tongue area. Figure 3 shows the pressure and mass flux field variations for housings A and B. sth order Fourier components were shown to have reduced by 73% compared to Housing A. Strain measurements were made on the same wheel in both Housings A & B. At sth order, the strain was reduced by 70% (Figure 4). This compares well with the calculated reduction in Fourier amplitudes.
Housing B
Housing A
Figure 3a CFD Static Pressure Field
11) (I
Wl'.i{,,,,l?/$)
Housing A
Housing B
Figure 3b CFD Mass Flux Field
266
0.8
.. c
"i!
0.6
!I)
CD
> ~ Gi
IX
0.4
0.2
0 5th order
6th order
Figure 4 Strain Measurements on same wheel using Housing A and B HARMONIC ANALYSIS
The prediction of actual displacement and strain values is performed in the Ansys finite element package with a single blade model using harmonic analysis. Prior to the harmonic analysis, a modal analysis is required to calculate the mode one natural frequency of the blade of the finite element model, as this will be slightly different to a real blade. The exact frequency is required to catch the peaks in the harmonic response. A single blade is used, as opposed to a full wheel, to consider a tuned wheel with nominal blade geometry rather than a mis-tuned wheel. Prior to considering harmonic analysis, blade strain was determined by transient dynamic analysis. To perform transient dynamic analysis required the creation of a pressure loading file at everyone degree around the housing. These loads were then applied to the blade until a steady state response resulted, taking around 128 rotations or applications of each load step. To perform the transient dynamic analysis on a single blade took around six weeks run time. The transient dynamic analysis utilised a similar method to that described in [5] on a full wheel with coarser mesh. A similar result to that achieved in transient dynamic analysis can be achieved in harmonic analysis with a run time of only a few hours, using the same mesh, containing 35,000 elements. The finite element solid model of one blade includes the associated segment of the hub. It is meshed with second order tetrahedral elements resulting in around 150000 degrees of freedom. The model is constrained axially at the friction weld boss and by constraint equations linking the hub cut faces, which allow identical radial displacement at each cut face. A damping ratio of 0.0005 was used as measured in test work. The harmonic analysis calculates the response of the turbine blade to a given sinusoidal pressure applied across the blade surface. As previously explained the pressure loads are calculated by CFD for a given wheel and housing at a set running point. The relevant sinusoidal order of the pressure is extracted by FFT to load into the model. It should be noted that the amplitude and phase of the loading varies across the blade.
267
lIN
.7 r.,.,.. ~..••
~ .-n, __
Figure 5a Maximum Principal Strain
• pi
_~
.
_
, .~
_
I
".
..
~
')!,~
_ 1 _1/ S
Figure 5b Minimum Principal Strain
The pressures are applied as mean, real and imaginary components to the blade surface nodes by a macro in separate load steps. The blade is pre-loaded with the mean pressure in a static analysis and real and imaginary components are applied in a harmonic analysis. The harmonic analysis is performed over a range of frequencies centred on the mode one natural frequency as predicted by the modal analysis. During post processing the strain and displacement plots can be studied. Figures 5a, and 5b, show plots of mode one maximum and minimum principal strain as determined by harmonic analysis. For nodes of interest a macro can be used to plot strain against phase angle for each frequency and the peak values determined. MEASUREMENT OF BLADE STRAIN AND DAMPING
Measurement of dynamic blade strain is achieved through the use of specialist high temperature strain gauges and a telemetric data transfer system. Wheels are chosen for the test following detailed frequency measurement of a representative batch. Strain gauges are applied to the surface of selected turbine blades using high temperature ceramic cement, with gauge locations identified through FEA analysis of the blade. The selected wheels are modified to allow routing of the strain gauge lead wires through the hub and shaft material for connection to a short range radio telemetry transmitter mounted at the compressor end of the rotor. The instrumented turbocharger is applied to the matched engine, and operated to achieve speeds associated with critical excitation order conditions. The transmitted strain signal is processed on-line using a real time fast Fourier transform and monitored during the test to ensure peak resonant conditions have been achieved. Data is recorded via PC and reduced to provide peak dynamic strain level for each rotational order of interest together with the graphical output shown in Figure 6, displaying strain amplitude with respect to frequency and shaft speed. Due to variations in volute geometry as a result of the casting process, it is necessary to test a number of housings of a particular design in order to identify the potential variation in excitation level. For this study measurements were taken from a total of twenty turbine housings.
268
.1
Figure 6 Strain Amplitude Waterfall Plot Data from this study was also used to estimate damping for use in the harmonic prediction work. Tracked plots of dynamic blade strain amplitude versus rotor speed around the sth order operating conditions were produced and curve fitted assuming a single degree of freedom system behaviour. Damping was then estimated using the halfpower bandwidth approach described in Shock and Vibration Handbook [6], and referred to in blade vibration work by Griffin [7]. COMPARISON OF PREDICTED AND MEASURED STRAIN VALUES
Figure 7, shows the comparison of predicted and mean measured strain on the same turbine wheel combined with the housings A and B at two different orders. There is a slight under prediction at sth order for the Housing A and at 6th order for Housing B. However, in general the prediction is very accurate bearing in mind the difficulty of the calculation. This accuracy relies on: the boundary conditions of the CFD being correct, the accuracy of the complex unsteady stator rotor CFD calculation, the application of the blade loads and the blade harmonic analysis. Interestingly despite this, the comparison is better than when the Fourier components of the pressure distribution are compared in Figure 2b. This may be due to the grid quality in the particularly location or the inaccuracy of the measurement. The strain calculations rely on a distribution of load on the blade rather than individual Fourier amplitudes and therefore slight inaccuracies at particular locations will be averaged out. The analysis clearly is able to distinguish between a 'high' and 'low' strain turbine stages and therefore would be invaluable as a design tool for either the housing or wheel.
269
0.8 121 Measured Strain o Predicted Strain
c:
"i!
iii
0.6
~ ~ Gi 0.4
~ .. .. k%
W
0::
. . ~ ..
-~
0.2
~n ..
0 Housing A 5th order
Housing B 5th order
Housing A 6th order
Housing B 6th order
Figure 7 Comparison of predicted relative strain values at 5th and 6th orders CONCLUSIONS
The paper describes the analysis and experimental method used to understand the mechanism of turbine blade vibration. The CFD method has been validated by comparing the predicted volute outlet pressure distribution and Fourier components with measured data obtained by a technique which enable a detailed description of the flow field. A more complex CFD calculation was developed in order to obtain the unsteady load distribution along the blade surface. This load was decomposed into an amplitude and phase angle corresponding to the order of vibration so that it could be used as an input into a harmonic calculation. The resultant predicted strain was favourably compared with measured blade strain values obtained using a telemetry system. The accuracy of this method gives confidence that the technique can be used as a design tool for the turbine housing volute and wheel. The previous method concentrated on understanding the housing exit flow and a new housing was developed based on understanding and optimising the pressure distribution. The turbine stage with the optimised housing resulted in a 70% reduction in measured and predicted blade strain. This method can now be used to understand the housing/wheel interaction and also how different wheels behave under vibration. The method seems to work well for simple mode I vibration, however, further work is needed to validate the method for more complex modes. REFERENCES
I K H Scrimshaw and T J Williams, 'Size Effects in Small Radial Turbines', The American Society of Mechanical Engineers, 84-GT-21S 2 D E Winterbone, B Nikpour and G I Alexander, 'Measurement of the performance of a radial inflow turbine in steady and unsteady flow' I.Mech.E., C40S101S, 1990
270
3 W N Dawes, 'The development of a solution-adaptive 3D Navier-Stokes solver for turbomachinery', AIM Paper 91-2469, AIMISAEIASMEIASEE 27th Joint Propulsion Conference, Sacramento, June 1991 4 P M Chapple, P F Flynn and J M Mulloy, 'Aerodynamic Design of Fixed and Variable Geometry Nozzleless Turbine Casings' ASME, 1979, 79-GT-87 5 D Filsinger, M Sekavcnik, T Ihli, A Schulz and S Wittig, 'Vibration characteristics of a radial turbocharger impeller' Seventh International Conference on Turbochargers and Turbocharging (2002). 6 C M Harris and C E Crede, Shock and Vibration Handbook, fourth edition, McGrawHill, ISBN 0-07-026920-3 7 J H Griffin 'An Improved Method of Measuring Blade Vibration and Predicting Engine Durability' Aero Propulsion Laboratory, Air Force Wright Aeronautical Laboratories, AFWAL-TR-86-2118 © Holset Engineering Co. Ltd
271
Axial Load Capacity of V-Section Band Clamp Joints Kiumars Shoghi, PhD, MIMechE, CEng Borg Warner Turbo Systems Simon Barrans, PhD, MIMechE, CEng University of Huddersfield Prabhu Ramasamy, BSc University of Huddersfield SYNOPSIS
In this paper a method of predicting the axial load generated by V -section band clamps, taking into account both circumferential and transverse friction has been proposed. An experimental method of determining this axial load is also demonstrated and the theoretical predictions are shown to be accurate. The effect of this axial load on the turbocharger components making up the V-band joint has been investigated with the aid of finite element analysis. It has been demonstrated that this load can make a significant contribution to the level of stress in these components. NOMENCLATURE
Fa = Axial clamping load (N) F9 = Circumferential force at angle e (N) F~
= Force applied by the T-bolt (N)
R2 = Radius of the flange (m) f= Flange edge thickness (m) q = Load per unit length in the normal direction (N/m) s = Load per unit length in the axial direction (N/m) 8 = Angular position around the band (rad)
p = subtended angle of half the V-section band (rad). I..l = Coefficient of friction between the V-section band and flanges. ~
= Half angle of the v-section (rad)
INTRODUCTION
When a flange assembly is required to maintain a fluid tight seal, the axial capacity can be taken as the load necessary to overcome preload of the clamped joint and causing separation between two flanges. A typical turbocharger will have a housing in three parts; compressor and turbine housings mounted on either end of a central bearing housing. The interface between these parts is normally a pair of circular flanges clamped together in some way. In turbocharger applications it is important to have sufficient axial load capacity available in these clamped joints to prevent leakage and emission. Any such leakage would cause pressure drop between the wheel and the housing and result in loss of efficiency and performance of the engine system. In 273
addition, if the axial capacity available in the joint is too small, separation of the flanges may occur due to the loads applied from other engine system components. This could then allow contact between the compressor and turbine wheels and the profile of the housings to occur, quickly leading to turbocharger failure. During turbocharger operation temperature changes cause thermal expansion in the pipes for air intake, exhaust outlet, and compressed air supply to the engine. The loading due to these thermal effects combined with the loads due to vibration transmitted from the engine must not be large enough to overcome the internal clamping loads between the three turbocharger housing sections. Hence it is necessary to establish the value of the axial load capacity for these joints in order to specify maximum allowable external loads. In many modem turbochargers, these joints are formed by Vsection band clamps. The V -section of these bands captures the flanges and as the band is tightened in the circumferential direction both radial and axial loads are applied to the flanges. As part of the health and safety procedures for turbochargers, each new design has to be subjected to a burst and containment test. This test is designed to examine the ability of the turbocharger housing assembly to retain component fragments when a rotating wheel in the turbocharger is caused to fail by centrifugal action. An essential part of this containment is the ability of the V-section band clamp to hold the housing components under the action of the large axial loads that can be generated by wheel failure. An example of a V-band joint failure generated during a containment test is shown in figure l. Despite their wide use in a variety of industries, there is little or no guidance given regarding the relationship between the tensile load applied to the T-bolt and the resulting stress distribution in the V -band and load capacity of the V -band joint.
Figure 1. Failure of V-Band during burst and containment test 274
The majority of manufacturers base their designs on empirical equations of the type presented by Montford [1] and Hudson [2]. The limitation in such equations is associated with obtaining accurate experimental data over a wide range of design variables. Recently Takeuchi and Onoda [3] have proposed a theoretical model for V-band joints that recognises the variation of circumferential force around such joints. This model is based on the assumption that the clamp is of the form now being used in aerospace applications with a separate band strap and relatively short, discrete Vsegments. This model is therefore very similar to the flat band model recently demonstrated by Shoghi et al [4]. Within the Takeuchi-Onada model, the effect of the size of the V -section angle is absorbed into the coefficient of friction and is therefore difficult to employ across a range of V -band styles and applications. THEORY
Previously a method of determining the forces in a V-section band clamp has been presented by Shoghi et al [5]. In that work an equation was derived that predicted the axial load applied to the flange pair by the V-band. This theory accounted for friction acting between the V-band and the flange in the circumferential direction and hence, the variation in interface forces relative to circumferential position around the joint. An alternative method of determining axial load capacity is given by NASA [6] for Marman clamps. These clamps are similar to V-section band clamps but in Marman clamps the V -section and the band are separated. This is achieved by loosely attaching a number of short, V -section blocks to the inner diameter of a band strap. The attachment allows some circumferential movement of the blocks and also allows for differing coefficients of friction at the block-flange interface and the block-band interface. The installation procedure for these bands then specifies that strain gauges should be attached to the band to ensure that the circumferential load is uniform around the band. Hence, friction in the circumferential direction can be discounted. The remaining, transverse friction component acts in the plane of the section. For V-section band clamps of the type typically used in turbochargers, the circumferential friction effect cannot be neglected. However, transverse friction in the plane of the section will also be of importance. A method of predicting axial load capacity that takes account of both friction components is therefore required. The force components acting on the band section and a section of one flange are shown in figure 2 for the case of the band being tightened onto the flanges. The radial component of force acting on the band section is:
q sin ¢ + ,uq cos ¢
(1)
where: q is the normal load per unit length f1 is the coefficient of friction This force components is reacted by the internal circumferential force, Fe' in the band as derived in Shoghi et al. Hence, by radial equilibrium: q sin ¢ + ,uq cos ¢ = Fe 2R2
(2)
where R2 is the radius of the point of contact between flange and band 275
I
,/1.
,'1
.'--1I "
,/ ¢ /
! I
POSSIBLE RANGE OF CONTACT POSITION
FORCES ON BAND
FORCES ON FLANGE
f
CONTACT POINT
,, ,, ,, ,,
.,
~
... ~ ...._ • •~.. ...... .11....... , -
I
..",-,,~
Figure 2. Internal and external force components
Previously it was shown by Shoghi et al that the effect of friction III the circumferential direction causes a change in the internal circumferential load. With reference to figure 2, this is given by: (3) Combining equations 2 and 3 then gives: ------'-p-- de pcos¢ + sin ¢
276
(4)
Hence, the internal circumferential force can be related to the T -bolt force, Fp ' by: _Fp = exp[---,-Il....::.(fJ_-_(}--,-)_] Fe IlCOS¢ + sin¢
(5)
and the normal load per unit length is:
F exp[ q=
p
ll(fJ - (}) ] 11 cos¢ + sin ¢
(6)
2R2 (11 cos ¢ + sin ¢)
The axial force, Fa' generated by the band can be determined by integrating the axial load per unit length, s, around the band circumference:
(7)
The axial load per unit length canbe determined by considering axial equilibrium of a flange. With reference to figure 2:
s - qcos¢ + f1qsin¢
=
0
(8)
Combining equations 6, 7 and 8 yields, with some manipulation:
Fa
=
(l-lltan¢)Fp(IlCOs¢+sin¢)[ ( -llfJ 1 - exp ----'---'----CIl+ tan¢) (Ilcos¢+sin¢)
J]
(9)
For the case of axial load being applied to the joint, the direction of the in-plane friction component shown in figure 2 is reversed. The axial load is then given by:
Fa =
( - IlfJ (l + 11 tan ¢)Fp (sin ¢ - 11 cos ¢) [ 1 - exp ----'---'----(tan¢ -11) (sin¢ -Ilcos¢)
J]
(10)
A comparison of the previously presented axial load theory and the results obtained from equations 9 and lOis given in figure 3. The included half angle, ¢ , was set at 20 0 , the half angle of the band, fJ, was set at 167 0 and the T-bolt load was set at lOkN. These values are typical of the V-bands used on turbochargers although the T-bolt load recommended by the band manufacturers may often be less than this. It can be seen in this figure that the previously presented theory gave a result that would approximate both the tightening and axial load application cases derived here. Hence, it would be reasonable to use this equation as a first estimate. Figure 3 also shows that once assembled, the axial load applied to the V -band joint can be increased significantly without causing an increase in the T-bolt load. This is clearly beneficial in turbocharger applications where V -bands may be required to contain extraordinary events such as shaft or wheel failure. 277
80
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60
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ou.
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....... NEW THEORY: TIGHTENING
.... ....
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- - NEW THEORY: AXIAL LOAD
....
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,,
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~
20 10
o
0.2
0.1
0.3
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Figure 3. Comparison of old and new theories EXPERIMENTAL WORK Previous Work Previous attempts to determine the axial load generated in V-band joints (e.g Shoghi and Brown [7]) have typically used a pair of pressurized flange plates with an O-ring seaL Hydraulic pressure introduced between the plates is used to generate an axial load. The joint is normally judged to have failed when no increase in hydraulic pressure can be sustained. This peak pressure is then used to determine the axial load applied. The principal difficulty with the test described above is that the pressure at which leakage occurs is highly dependent on the O-ring and the O-ring groove. Although the flanges may separate, indicating that the axial load has been overcome, the O-ring may well maintain a pressure seaL Leakage and pressure loss will only occur as the O-ring starts to extrude between the flange plates. In much of the previous work, T-bolt tension had not been measured directly. Instead, the torque applied to the T-bolt was recorded (see for example Shoghi [8]). The geometry of the thread was then used to convert this value into a tensile load using an assumed value for the coefficient of friction within the thread. This introduces a number of undesirable sources of significant error. A New Methodology For the experimental work reported here, the objective was to measure the axial load as it was being generated, rather than measuring the pre-load effect due to the axial load. To do this, the flange plates were mounted in an Instron tensile test machine, as shown
278
in figure 4. The pull rods were screwed into the flange plates and a lock-nut tightened to eliminate movement due to thread clearance. The hydraulic ram was then used to bring the two flange plates together. The Instron control system was instructed to generate an axial compressive force of 0.5 kN between the plates. This ensured that the plates were touching. Switching to displacement control, the hydraulic ram was moved down a distance of 0.1 mm. This created a clearance between the flange plates and was designated the reference position. The V -band was then assembled loosely around the flange pair. A small 20 kN load washer was placed under the T-bolt nut to measure the T-bolt tension, FfJ' Previous experience with this type of load cell had shown that subjecting them to torsional load could generate significant errors. To guard against this, a roller thrust bearing was placed between the nut and the load cell.
t
+--
INSTRON LOAD CELL
... 1
I
I
INSTRON GRIPPE
FLANGES
THRUST BEARING
T-BOLT NUT
V-BAND
LOAD CELL
······1
.f---,
Figure 4. Test arrangement
279
INSTRON GRIPPER
With the components assembled, the T-bolt nut was tightened to generate the desired tension in the bolt. The axial load generated in the joint was then measured directly from the Instron control system. Readings were taken at the reference flange position and with greater flange separation. Between each flange position the V -band was fully relaxed. The flange and V -band characteristics used in this experimental work are shown in table 1.
PARAMETER BAND HALF ANGLE, ~ SECTION HALF ANGLE, ¢ FLANGE EDGE THICKNESS, f FLANGE LARGE DIAMETER, R2 SMAll
I
VALUE 1670 20 0 7.13 mm (min) 7.81 mm (max) 117.5 mm
I
55.88 mm Table 1. V-band properties
A second series of tests were carried out on the larger band. For this test the T-bolt was tightened to a nominal load with the flange plates at their reference position. The separation between the plates was then increased with both axial load and T -bolt tension being recorded at each stage. Results The results obtained for the larger flange pair are shown in figure 5. It can be seen here that there is little correlation between flange separation and axial load. This is as predicted by equation 9. Superimposed on this figure are the results predicted by equation 9 for the V -band data shown in table 1 and coefficients of friction of 0.1 and 0.2. These values span the range typically quoted for this type of application (see for example Takeuchi, S., Onoda [3] and Lancho et al [9]). It should be noted that higher levels of axial load can be generated with lower coefficients of friction. This fact is noted in the NASA design guides where it is recommended that a dry lubricant be applied to the flange surfaces prior to Marmon clamp installation. The results obtained from the tests carried out on the smaller band are shown in figure 6. Again, the results predicted by equation 9 for representative coefficients of friction are also shown. For the reference position and the first two deviations from this position, there is good correlation between the theoretical and experimental results. However, as flange separation is increased the experimental results deviate from the theory. There also appears to be a correlation between flange separation and axial load capacity. These two observations can be explained by considering the precise form of the V-section.
280
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ref +0.3
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+0.4
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+0.5
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oLL
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. THEORY (mu = 0.2)
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.
15
.
.
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•
.
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o
2
3
4
5
6
7
~
8
T-BOLT TENSION (kN)
Figure 5. Large flange results 50
ref
•
+0.2 mm
..
+0.4 mm
•
+0.6 mm
t:.
+0.8 mm
o
45 40
--THEORY (phi
=20; mu =0.1) o
35 .
.....
z
C 30
•
o
0
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25
... .'. o
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20
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D.
15
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_8'-'
10 5 0 0
2
4
6
8
T-BOL T TENSION (kN) --------
~----.--
Figure 6. Small flange results
281
10
12
14
It is assumed in the theory that contact between the flanges and the V -band takes place on a flat, conical surface between the bends in the section. For the bands studied here, this surface is specified as being inclined at an angle of 20° , as shown in figure 2. As flange separation increases, the flange edge thickness, f, will also effectively increase. This will force the V-band to expand and drive the point of contact towards the V-section legs. When the contact position reaches the bend in the section, the angle of inclination will increase dramatically. As shown in figure 6, this will significantly reduce the axial load generated by the clamp. The correlation between flange separation and axial load for moderate levels of separation is best understood by considering the V-band manufacturing process. The first stage of the manufacturing process is to generate the V-section by rolling material in flat strip form. At this stage the section will have the desired flat surfaces between the bends. After this rolling process the formed strip is bent to generate the circular band. During this second bending operation the section is warped, causing some outward displacement of the section legs. This displacement is generated by bending of the region between the top of the section and the legs. Hence, there will be some variation in the included angle of the section. As the effective flange edge thickness is increased, the point of contact will move down the section and hence to a region with a larger included angle. If the experimental results from low to moderate levels of flange separation for the small flanges were superimposed on those shown in figure 6 for the large band, it would be seen that they had almost identical linear trend lines. This confirms the perhaps surprising prediction from the theory that the level of axial load generated is independent of flange diameter. The results increasing the axial load on the larger band are shown in figure 7. Also shown in this figure are the results predicted by equation 9 for tightening and equation 10 for axial load application. Whilst results show the expected trend with a noticeable increase in axial load with no increase in T-bolt load, there is significant deviation from the theory. The theory actually predicts a significant decrease in T-bolt tension as the direction of the transverse friction component is reversed. This does not appear to be logical and this mode of operation deserves more detailed theoretical and experimental investigation.
Appraisal of the Methodology
The benefit of the methodology described above was that axial load was being measured directly. The use of a load cell to measure T-bolt tension was also an advance on previous work. However, there are a number oflimitations. These include: • It was found to be difficult to align the flange plates such that they were parallel. Sources of misalignment included the threads on the pull rods and in the flange plates, clamping of the pull rods in the Instron jaws and correct alignment of the Instronjaws. • The thrust bearing under the T-bolt nut was operating at levels approaching its load capacity. It was noted that some torsional load was being transmitted to the load cell. Whilst this does not appear to have affected the results, it could lead to damage in the long term.
282
60 50 2'40
==-
Cl
«
g 30 ...J
« ~ 20 10 0 0
2
4
6
8
10
T-80L T TENSION (kN)
Figure 7. Axial loading effect - large band OTHER COMPONENTS
Having determined the axial load capacity for a v-section band clamp on a turbocharger joint, it is important to examine the stress distribution in the clamped housings when subjected to an external load just below the value of the axial clamping load. At levels of axial load above this value flange separation would occur and components would clash leading to catastrophic failure of the turbo charger. The most appropriate technique for predicting the stress distribution in the complex castings making up the turbocharger housing is finite element analysis. Models were created of the compressor and bearing housings of a typical turbocharger employing the larger V -band studied above. The finite element package used was ANSYS, version 8.1. The geometry of the structures was initially defined within ProEngineer and imported into ANSYS. Due to the complexity of this geometry, tetrahedral elements with a linear interpolation polynomial were used and the mesh defined using the automatic, free meshing facility. The model of the compressor cover was constrained in the axial direction on edge of the flange that interfaced with bearing housing and V-band. This constraint simulated the presence of the V -band without having to include it in the model. Including the Vband would have required the specification of contact between the components and resulted in a prohibitively lengthy non-linear analysis. The axial load was applied as a distributed load on the air inlet connection. The results of this analysis showed that high stresses were generated at the root of the flange. However, these stresses were well within the limit of the material. The bearing housing model contained V-band flanges at both the compressor and turbine ends. It was assumed that the axial load could be transmitted from one flange to the other. The edge of the compressor flange was constrained in the axial direction and the axial load was distributed over the edge of the turbine flange. The stresses generated are shown in figure 8. Again, high stresses are shown at the root of the flange, 283
particularly where this runs into other casting features. Whilst the levels of stress observed were within the material limit, there was only a small safety margin. CONCLUSIONS
A method of predicting the axial load generated in a V-section band clamp joint taking account of both circumferential and transverse friction has been generated. The validity of the theoretical model has been demonstrated by practical testing of V-band clamps with extreme diameters. Agreement between the theoretical model and the practical results has been shown to fall within the error band due to uncertainty regarding the coefficient of friction within the joint. The effect of axial load generated by the V-band on other components within the turbocharger has been investigated by finite element analysis. It has been shown that in some cases, this effect can be significant. FURTHER WORK
The current theory can only predict the axial load generated by the clamp operating at levels of stress below yield. This theory could be usefully extended to allow ultimate failure of the V-band to be predicted. This would be a substantial development since ultimate failure of V-bands occurs in the internal comers of the section. To predict the approach to failure, the effect of residual stresses due to the forming process and bending of the section would have to be accounted for. The experimental method used here, whilst being a significant improvement on previous methods still has two sources of potentially significant error. These are undesirable torsional load being applied to the T-bolt load cell and misalignment of the flanges. It is proposed to overcome this second error source by running the two flanges on the same ground shaft within a purpose built test rig. Alternate methods of applying the tensile load to the T-bolt are currently being reviewed. The experimental work needs to be extended to demonstrate the joint behaviour as axial load is being applied. These tests should be taken to the ultimate failure point to validate the extensions to the theory suggested above. The finite element work has assumed that the axial load is uniformly distributed around the V-band flanges. Future work should model this load with the variation predicted by equations 8 and 6. This will allow the effects of V-band orientation relative to the housing components to be analysed.
284
NODAL 30LUTION SSP
9TEP=1 3UB =1
1 2005 11:26:43
TIME=l SEQV DMX
(AVG)
= . 075425
9MtI =.018301 3MX
=285.816
.018301
31.774
63.529
95.285
127.04 158.796
818
Figure 8. Von-Mises stress in bearing housing REFERENCES 1. Mountford, R. Design of clamp joints, Engineering Designer, pp. 37-40, 1980. 2. Hudson, C. vee-clamps solve flange coupling problems, Diesel Engineering, pp. 203-204, 1977. 3. Takeuchi, S., Onoda, 1 Estimation of separation shock of the Marmon clamp system by using a simple band-mass model, Trans, Japan Soc. Aero. Space Sci., Vol 45 (147), pp 53-60, 2002. 4. Shoghi, K., Rao, H.V. and Barrans, S.M., Stress in flat section band clamp, Int. 1 Mech. Sci. , volume 217, part c, pp 821-830,2003. 5. Shoghi, K., Barrans, S.M. and Rao, H.V., Stress in V-section band clamps, Int. 1 Mech. Sci. , volume 218, part c, pp 251-261, 2004. 6. Marrnan clamp system design guidelines, Guideline No. GD-ED-2214, NASA Goddard Space Flight Centre, 2000. 7. Shoghi, K., Brown, 1., Axial load capacity for V-bands, Teconnex - BorgWarner internal report, 2004. 8. Shoghi, K., Stress and strain analysis of flat and V-section band clamps, PhD thesis, University of Huddersfield, 2003. 9. Lancho, M., Larrauri, 1, Gomez-Molinero, V., CRSS: A separation system for launching very heavy payloads, Acta Astronautica, volume 47, part 2-9, pp 153-162, 2000.
285
Reliability Trends, Operating Issues and Acceptance Criteria related to Exhaust Gas Turbochargers used in the Marine Industry A Classification Society View Kian Banisoleiman* BSc, PhD, C.Eng, MIMarEST, MIMechE Norman Rattenburl C.Eng, FIMarEST
* - Lloyd's Register EMEA,
+ - Lloyd's Register 71 Fenchurch Street London EC3M 4BS
SYNOPSIS
Lloyd's Register is a leading international classification society with the objectives of enhancing its clients' quality, safety, environmental and business performance. In support of these objectives Lloyd's Register maintains technical Rules and regulations for the classification of ships and installed machinery, including engines and their exhaust gas turbochargers. This paper provides a current perspective of a classification society on exhaust gas turbochargers used in the marine industry from the view points of: • Rule requirements and issues relating to turbocharger response during engine load variations stemming from governor testing and the operation of electrical/propulsion systems • the most recurring in-service faults and their incidence statistics over the past decade for exhaust gas turbochargers • failure investigation cases related to turbocharger burst containment
INTRODUCTION
Lloyd's Register (LR) is a leading international classification society with the objectives of enhancing its clients' quality, safety, environmental and business performance. In support of these objectives LR maintains technical Rules and regulations for the classification of ships and installed machinery, including engines and their exhaust gas turbochargers. In 2005, Lloyd's Register amended its rules [1) relating to turbochargers by requiring the manufacturers to submit their turbocharger burst test assessment in addition to the usual turbocharger plans and particulars. This paper presents an overview of turbocharger rule requirements and issues related to turbocharger performance and response during engine load variation testing arising from engine capability and governor testing with operation of electrical systems. Technical data is presented in relation to the most common recurring in-service turbocharger defects relative to the other engine component defects for a wide range of auxiliary and ship main propulsion installations. Two failure investigation cases related to turbocharger burst containment are presented followed by analysis discussion and conclusions.
289
RULE REQUIREMENTS
Lloyd's Register's Rule requirements for turbochargers stem from the International Association of Classification Societies' (lACS) Unified Requirements M23 for design, construction and testing and M35 and M36 for alarms and safeguards. The Rules cover generic and additional requirements for mass-produced turbochargers. The generic requirements address the following: • plans and particulars to be submitted • type test requirements • dynamic balancing • overspeed test • mechanical running test • safety protection devices The requirements for mass-produced turbochargers are supplemental to the generic requirements and address the quality control of materials, manufacture, testing and certification. The plans and infonnation to be submitted include the following: • cross-sectional plans of the assembled turbocharger with main dimensions • fully dimensioned plans of the rotor • material particulars with details of welding and surface treatments • turbocharger operating and test data • manufacturer's burst assessment • details of alarms for exhaust gas temperature and lubricating temperature/pressure conditions
oil
Turbochargers intended for marine main propulsion engines are required to be constructed under survey. Lloyd's Register surveyors will be involved in the certification of materials, the applied standards and quality of workmanship together with testing at the manufacturer's and the engine builder's works. The installation and functioning of the required alanns and safeguards when installed on the engine onboard the ship are also verified. The turbocharger is a surveyable item subject to a survey regime as part of the engine system.
TURBOCHARGER PERFORMANCE The integration and matching of the turbochargers to particular engines is the responsibility of the engine builder. The classification society's role is to verify that the engine installation is able to meet the required load perfonnance capabilities, also stemming from lACS requirements. Load perfonnance capability verification for diesel engines is by means of type testing and an assessment of operational capability at the engine builders and during sea trials. During testing and trials, engines are required to run at different load/speed conditions up to 110% of the rated output for defined periods of time. The capability of the engine to quickly respond to different load conditions is essentially dependent on three support systems - the fuel system, charge air/exhaust gas arrangements and the governor characteristics. For engines driving electrical generators, the governing characteristics
290
and the ability of the engine and its support systems to respond to load variations is of key importance. Lloyd's Register has defined performance capability requirements for governors to provide a standard for assessment purposes. Governors are required to control the engine speed within 10% momentary variation and 5% permanent variation when full load is suddenly taken off or, when after having run on no-load for at least 15 minutes, load is suddenly applied as follows: • for engines with a brake mean effective pressure (BMEP) of less than 8 bar, full load application, or • for engines with BMEP greater than 8 bar, a percentage load calculated by the formula 800/BMEP %, (BMEP in bar), but not less than one-third of the full load, the full load being attained in not more than two additional equal stages as rapidly as possible. For larger engines, typically over 1 MW, the required capability of the turbocharger and associated charge air/exhaust system to respond to large load changes is sometimes difficult to achieve. On a case-by-case review acceptance of different load applications has been agreed on the basis of an analysis of the electrical load system. Such assessment is based on the maximum electrical step load that can be switched on or off, not causing the frequency variation of the electrical supply to exceed the following parameters: • permanent variations ±5% • transient variations due to step changes in load ± 10% with a maximum recovery time of 5 seconds.
IN-SERVICE TURBOCHARGER DEFECTS Background Lloyd's Register has maintained worldwide technical data on machinery defects through survey and failure investigation reports since 1958 [2]. In the early years, an edgepunched card system was used to record the technical details. In 1964 all data for our classed fleet, hull and machinery defects and basic data were transferred to 80-column punched cards which could be manipulated by an electronic sorter. This system quickly proved cumbersome as the demand for defects statistics and detailed information increased. In 1973, the technical data were transferred to an IBM mainframe computer. A database application was used for maintenance, searching and retrieval of information. In 1991 the old structure of the database was replaced by a new ship division database with enhanced capabilities [3]. For studies of machinery defects, the system could identify the population at risk, determine the period each component of the population had been at risk, find the appropriate defects attached to each component and calculate the defect rate in terms of defects per 10 ship years (this means number of defects which can occur on one ship in 10 years or 10 ships in one year). The technical records in the database were searched to produce information related to overall engine faults and specific turbocharger defects over a 10-year period to October 2005. The search was classified into three groups: for 4-stroke and 2-stroke main propulsion diesel engines, and for auxiliary diesel engines in all vessels classed with Lloyd's Register.
291
Published data from the marine insurer the Swedish Club [4] corroborate the overall trends seen the LR technical data specifically related to turbocharger defects on medium and slow speed main propulsion marine diesel engines. Auxiliary diesel engine defects overview The auxiliary engines comprised a population of 35,776 units for which 12,779 defects were reported. Analysis of these showed that the two most common defects related to crankshaft and connecting rods, comprising 23% and 16% respectively of total defects. Turbocharger defects were in seventh place, out of the 39 defect categories and contributed to 3% of the total defects. This was considered a small portion of engine defects and it was assumed that turbochargers on the auxiliary engines were robust compared to other engine components and were not analysed further. Main propulsion diesel engine defects overview 4-stroke engines featured a population of 11,093 units, average age 5.93 years, for which a total of 10,238 defects were recorded in 43 categories. Figure 1a presents a pie chart of the first 5 defects with largest proportion of the total. Figure la shows that the largest proportions were attributed to the connecting rod (21 %), crankshaft (16%), piston (11 %) and turbocharger (9%). It should be noted that this data comprises only the recorded defects and engine-component replacements due to the manufacturers recommended servicing intervals were not included in the data set.
Lloyd's Register [2,3)
Crankshaft
Fuel system
I Turbocharger I Conrod
Figure La - Lloyd's Register's data [2,3] 4-Stroke main propulsion marine diesel engine defects
292
Swedish Club [4)
Figure l.b - Swedish Club's data [4], medium speed engines 4-Stroke main propulsion marine diesel engine defects
Figure Ib shows a corresponding diagram based on the Swedish Club [4] data which has a total population of 82 defects spanning over 6 years to 2004. This data relates to major machinery damage claims (above 116,000 USD) for the marine propulsion medium speed diesel engines. The trend in the Swedish Club data ranked the first three most common faults being due to crankshaft, connecting rod and turbocharger with equal proportions of 25.6% of the total defects. Although used for different purposes, similarities may be drawn between the Lloyd's Register's and the Swedish Club's data, in that the most common defects related to engine connecting rods and the crankshafts. The turbocharger defects were ranked at fourth and first equal in Lloyd's Register's and Swedish Club's data respectively for the medium speed diesel engines. The data for 2-stroke diesel engines are shown on Figures 2a and 2b from Lloyd's Register's database and a corresponding diagram based on the Swedish Club data respectively. Two-stroke diesel engines comprised a population of 5,025 units, average age 6.1 years, for which a total of 6130 defects were reported in Lloyd's Register's database. A population of 135 defects comprised the Swedish Club's data. Figure 2a and 2b show that for the 2-stroke main propulsion diesel engines, turbocharger defects comprised 18% and 42.6% of the total defects for the Lloyd's Register's and the Swedish Club's data respectively and were the most recurring defects by a clear margin.
293
Lloyd's Register 12,3J
35,36.37.38
Crankshaft
30
29 28
27 26
Conrod
~_iiiiiii~~
23,24 25 •
10
22
11
Turbocharger 17
18
Liner
Figure 2,a - Lloyd's Register's data [2,3] 2-Stroke main propulsion marine diesel engine defects
Swedish Club [4J
I Piston rod I '-~~7s7w-n---1
~
___%-.'-___
Journal bearing
Liner
ITurbocharger I
Staybolts
Figure 2,b - Swedish Club's data[4], slow speed engines 2-Stroke main propulsion marine diesel engine defects
294
The general reliability trends in the data show that in moving from high to medium to slow speed diesel engines the proportion of turbocharger defects over all engine defects increased from 3% to 9% to 18 % respectively in the LR data and from 25.6% to 42.6% in the Swedish Club data. This means that the mass produced turbochargers for high speed engines are more reliable in comparison to those fitted to the medium and slow speed marine diesel engines. For the latter two engines, turbochargers fitted to slow speed engines are twice as likely to encounter defects as those on the medium speed engines in relation to other engine defects. This is attributed to the more arduous operating environment of the turbocharger associated with the slow speed diesel engines in terms of: • Firstly, amplitudes of low frequency vibrations which are excited by the motion of the engine, turbocharger support and connection pipes which excite the rigid body modes of the turbocharger. With slow speed engines the excitations at the supports and connections tend to be larger and occur at lower frequencies. The shock loading due to firing pressure is also included in this category. Moreover, the larger turbocharger size means that the natural frequencies ofthe turbocharger rigid body modes are lower. • Secondly, the high frequency vibrations originated from unbalance forces of the rotor related to turbocharger turbine fouling are more likely on the slow speed diesel engines due to the variable quality fuel oil operation of the 2stroke diesel engines. These two reasons explain the differences observed in the turbocharger defect trends presented for the auxiliary/high speed, 4-stroke/medium and 2-stroke/slow speed marine diesel engines. One incidence of turbocharger defects in 4-stroke main propulsion engines is in multiengine, diesel-electric applications at part load, with one engine held stationary to enable load-sharing at high rating amongst other engines, for improved efficiency. In such cases the vibrations transmitted to the roller bearings of the stationary turbocharger, via the foundations, have been known to damage the bearings and the rotor and some means of rotating the turbocharger rotor of the stationary engine would reduce the incidence of this type of defect. Further more, the fatigue life design strategy for the smaller turbochargers is different from the larger units in that the smaller turbochargers are designed for the life of the engine, whereas the larger turbochargers have limited fatigue life and turbocharger parts have to be replaced at specific number of service hours to ensure continuity in operation. This servicing requirement, which in most cases involves dismantling, replacing a component or components and reassembling the turbocharger can introduce defects. The incidence of turbocharger defects arising from of this type of activity are higher in the larger turbochargers.
Turbocharger defects Figure 3 shows the percentage of turbocharger defects for the Lloyd's Register's population 2- and 4-stroke main propulsion marine diesel engine. Further analysis of the Swedish Club data turbocharger defects was not possible as this information was not available. The percentage values were based on the total number of turbocharger defects only which were 918 and 1,074 for the 4 and the 2-stroke engines respectively.
295
Figure 3 shows that most recurring defects were due to the rotor (11.7-11 %), turbine nozzle (11.7-7.6%) and turbine blade (10-7%). The bearing lubrication oil pumps on both the compressor and turbine sides of turbochargers on 2-stroke engines were more prone to defects (4%) than those on 4-stroke engines (1.5%). Tables 1 and 2 show the details of recorded defects for the rotor and the turbine nozzle respectively which were the most recurring turbocharger defects on the 4- and 2-stroke engines. The recorded defect details are those recorded by Lloyd's Register's surveyors. These exclude replacement of the components due to manufacturer's recommended service instructions. The 10-year incidence statistics are also presented. Lloyd's Register Data [2,3)
ICompressor bearingl
Oi/pump
(%)
Oi/pump
1
3
5
7
9
11
13
15
17
19 21
23
25 27
29
31
33
35 37 39
41
43 45
47
49
51
53
55
Defect number (I)
1-'l'1JRBOCI!,i\RGER 2 - COMPRESSOR AIR FILTER 34BEARING ASSEMBLY JOURNAL BEARING 5HOUSING 67SHELL 8LINING ROLLER BEARING 910HOUSING 11CASING 12INLET 13OUTLET DIFFUSER 1415VANE 16IMPELLER 17BLADE 18INDUCER 19BLADE
2021222324252627282930313233343536373839-
LUBRICATING OIL SYSTEM LUBRICATING OIL PUMP LUBRICATING OIL PIPE OIL SEAL CENTRE SECTION BEARING ASSEMBLY JOURNAL BEARING SHELL ROLLER BEARING CASING LABYRINTH ROTOR ROTOR JOURNAL (CENTRE) ROTOR JOURNAL (COMPRESSOR) ROTOR JOURNAL ('l'lJRBINE ) SUPPORT BEAM SEATING HOLDING DOWN BOLTS THRUST BEARING ASSEMBLY
40- 'l'lJRBINE 41BEARING ASSEMBLY 42JOURNAL BEARING 43HOUSING 44SHELL 45LINING 46ROLLER BEARING 47CASING 48INLET 49OUTLET LUBRICATING OIL SYSTEM 5051LUBRICATING OIL PUMP 52LUBRICATING OIL PIPE 53OIL SEAL 54NOZZLE 55ROTOR BLADE
THRUST BEARING PAD
Figure 3. Main propulsion marine diesel engine turbocharger defects
296
DEFECTS BENT BROKEN/SHEARED COLLAPSED COMPONENT FAILURE CRACKED DISINTEGRATED ERODED EXCESSIVE WEAR FAILED FRETTED GROOVED HAMMERED LOOSE/SLACK MISALIGNED OVERHEATED SCORED SEIZED/JAMMED VIBRATION WIPEDIMELTED WORN (STATED FAIR WEAR AND T
4-Stroke engines defect 10 year numbers incideuce 7 0.0011 4 0.0006 1 0.0002
-
-
1
0.0002
2-Stroke engines defect 10 year numbers incidence 8 0.0026 10 0.0033 2 0.0007 1 0.0003 4 0.0013 0.001 3
-
-
1 6 75 1
0.0002 0.0009 0.0114 0.0002
-
-
1 74
0.0003 0.0241
-
-
-
1
0.0003
1 -
0.0002
0.0003 0.0003 0.0003 0.0016 0.001 0.0003 0.0007
-
-
-
6 3
0.0009 0.0005
-
-
1 1 1 5 3 1 2
1
0.0002
-
Table 1. Turbocharger rotor defect details
DEFECTS BENT BROKEN/SHEARED CORRODED CRACKED DEFORMED DISTORTED DISINTEGRATED ERODED EXCESSIVE WEAR FAILED GROOVED HAMMERED LOOSE/SLACK MECHANICAL ACTION SCORED TORN VIBRATION WORN CST A TED FAIR WEAR AND T WASTED
4-Stroke engines defect 10 year numbers incidence 0.0006 4 2 0.0003
-
-
18 1
0.0027 0.0002
2-Stroke engines defect 10 year numbers incidence
-
-
1 26 4 2
0.0003 0.0085 0.0013 0.0007
-
-
1
0.0002
-
-
-
-
1 66
0.0002 0.01
3 3 35 1
0.001 0.001 0.0114 0.0003
-
-
1 1 2 2 1
0.0002 0.0002 0.0003 0.0003 0.0002
-
-
3 4
0.0005 0.0006
-
-
-
-
1
0.0003
-
-
1 2 3
0.0003 0.0007 0.001
-
-
Table 2. Turbocharger nozzle defect details
Of actual turbochargers defects, the most catastrophic is due to overspeeding, which at worst can lead to bursting of the compressor impeller. Due to the high rotational speed of the impeller, large kinetic energies are involved and this means that this type of failure must be contained within the turbocharger casing so as not to compromise the safety of personnel and cause damage to engine room and other machinery.
297
TURBOCHARGER OVERSPEED INVESTIGATIONS Incidences of impeller wheel bursting are rare. Of the 7,769 main propulsion 2-stroke diesel engines which were analysed, there were 24 compressor impeller faults reported. Of these, six failures are known to be compressor wheel bursts due to overspeeding. This is a low incidence rate and indicates that this type of problem is not widespread. However, this is considered to be a dangerous failure. Lloyd's Register's Rules [I] specifies that for mass produced turbochargers all fully bladed rotor sections and impeller/inducer wheels are to be overspeed tested for three minutes at either 20% above the maximum permissible speed at room temperature, or 10% above the maximum permissible speed at the normal working temperature. This is a proof-test to ensure that the components do not contain defect sizes causing failure at overspeed, as well as prolonging the in service life of the components which survive the test. Moreover, the manufacturer's turbocharger burst test assessment, showing containment, are required to be submitted in addition to the usual turbocharger plans and particulars in order not to compromise safety from this type of failure. Case 1 An overspeed burst failure of the compressor impeller occurred in the unattended engine room of a double hulled 66,895 tonne dead weight oil tanker in November 2003 at ballast conditions and in a 'good' sea state. The vessel was powered by a 8.1 MW marine main propulsion 2-stroke diesel engine. The turbocharger had accumulated 62,121 running hours and had undergone renewal of bearings, lubricating oil pumps and shaft seals 5,500 hours before the incident occurred. Figure 4a shows the collected parts of a compressor impeller after the incident. It can be seen that the compressor impeller had fractured into four segments of approximate similar size. Figure 4b shows a fracture face 'A' which is typical of high strain rate overload failure. An arrow points to the location of the failure initiated at the bore of the impeller where striations radiating towards the outer section of the impeller hub show the rapid failure propagation. The radial/axial direction of fracture is typical of
b. Fracture face 'A' initiation
a. Re-collected parts
tT",'tll1rp
Figure 4. Burst compressor impeller (case 1)
298
the high speed burst and indicates that excessive hoop stresses were the probable principle cause of failure. There was no evidence of polishing or fretting in the bore of the impeller, or any rubbing contact on the labyrinth seal area on the back face of the impeller indicating that it had not been loose prior to fracture contact. Further, chemical analysis showed that the impeller had been manufactured from age hardening aluminium-copper-magnesium alloy of the 2618 type, which is a typical material for this type of compressor impeller with adequate tensile mechanical properties. Figure 5 shows the debris found in the turbocharger casing and in the exhaust pipe. In this case all debris was contained within the turbocharger casing, the engine and the exhaust manifold. There were no injuries and no damage to other machinery in the engine room. Measured engine data immediately prior to the incident were not available, but, turbocharger fouling and, compressor surging were considered as the probable case of overspeeding. Case 2 Figure 6a shows the re-collected compressor impeller from another case where the impeller burst into two sections after overspeeding. This incidence occurred on the main propulsion system of a 59,093 tonne dead weight containership powered by a 34.4 MW 2-stroke diesel engine in February 1996. Figure 6b shows the fracture surfaces. A fractured surface with characteristic striations similar to case 1 above was apparent. In this case the turbocharger casing fractured and debris were scattered around the engine room.
a. In turbocharger casing
b. In exhaust manifold
Figure 5. Burst compressor impeller debris (case 1)
299
b. Fracture face (fracture initiation arrow)
a. Re-collected parts
Figure 6. Burst compressor impeller (case 2) Chemical analysis showed several oxidation layers at the position of fracture initiation and indicates that the failure could have been initiated by material defects. Figure 7 presents the conditions on the compressor map in term of measured normal operating conditions, the matched load line and the operating conditions 1 hour prior to the failure. The data points were measured by an engine condition monitoring data acquisition system in terms of turbocharger speed and the compressor pressure ratio. This information was corrected for the inlet temperature and plotted on the compressor map in terms of the matched load line, measured operating conditions and operating conditions one hour before the incidence. 4.5
4.0 N300;:;;rpm*sqrt(300rrtn) (rpm)
3.5
3.0
2.5
2.0
9000 7800
1.5
1.0 4
10
12
14
16
1B
Q30D=Q'*sqrt(300fTIn) (m3ls)
Figure 7. Compressor map (case 2)
300
20
22
24
it can be seen that a sufficient surge margin exists for normal operation, but, just before the failure the compressor was operating very close to the surge fine. Finite element analysis The impeller stress distribution induced at the operating speed just before the failure was analysed using a finite element model. A three-dimensional solid element model of a 45° sector (l/Sth) of the impeller was used. Simulation was by means of linear static analysis. The simulation was used to evaluate the speed at the onset of impeller bursting. Table 3 shows the different load cases used in the finite element model and the predicted corresponding maximum and minimum radial and tangential hoop stresses on the impeller. Load case 2 was based on the operation of the compressor at the pressure ratio just before the failure. From the pressure ratio and the compressor impeller geometry details, an approximate blade passage pressure distribution was evaluated by computational fluid dynamics (CFD) calculations and was used as input in the finite element analysis. An impeller speed of 13,400 rpm was initially simulated in load case 3. This was then increased to 14,400 in load case 4. The onset of failure was detected by analysis at 15,500 rpm which is 3% beyond the maximum speed presented on the compressor map. In reality material properties act in a non-linear fashion, there would be a preload stress due to the grip of the impeller on the shaft, heat transfer effects from the turbine side and a non uniform pressure field due to surge, but these were not modelled in the linear analysis. With an impeller material ultimate tensile strength of 450 N/mm2, Table 3 shows that the centrifugal loading dominated the amplitude of the tangential hoop stress and it can be seen that onset of failure was predicted at an impeller speed of 15,500 rpm.
Radial stress (N/mm2)
Load cases
1- Gravity only 2- Blade pressure distribution 3- Centrifugal loading at 13400 rpm 4- Centrifugal loading at 14400 rpm 5- Centrifugal loading at 15500 rpm
Tangential stress CN/mm2)
Max.
Min.
1
-I
7 192 222 257
-5 -65 -75 -S9
Max. 1 3 353 40S 471
Min. -1 -2 -13 -15 -17
Table 3. Finite element analysis predicted stress results
Figure 8a and 8b show the predicted distribution of the von Mises stress at 13,400 and 15,500 rpm respectively through the across section of the impeller hub. It can be seen that the maximum stress occurred at the impeller bore in the hub at the similar location as in the damaged impellers, Figures 6b and 4b above, and corroborates the hypothesis that overspeeding was also a cause in the failure. Thus an over speed of 15% was evaluated based on the highest turbocharger operating condition of 13500 rpm on the matched load line with a clean turbocharger.
301
Iso-ContourValues:
Iso-ContourValues:
a-
b.
@
13400 rpm
@
15500 rpm
Figure 8. Compressor impeller FEA model predicted von Mises Stress (N/mm2) A case study was reported by Theotokatos [5] for a 2-stroke main propulsion marine diesel engine and a dirty turbocharger which had similarities to the cases above. This was in a containership propulsion system at a rating of 41.1 MW. The study was by means of a filling and emptying model of the engine and a dynamic model of the turbocharger capable of simulating the surge cycle were used to assess the surge cycle behaviour of the engine and the turbocharger. The simulated results show that during the surge cycle, when the compressor operating conditions change from the reverse to normal flow, the compressor speed could increase by 10% at excursions towards the choke line, compared to the normal steady operation with the non-surging clean turbocharger. Furthermore, Theotokatos [5] presents experimentally measured results from a surging compressor on a similar 2-stroke engine which shows that the turbocharger could overspeed by 14%. Thus, it was concluded that compressor surging could be a cause of overspeeding. Preventing surge, by automatic monitoring and advanced warning to reduce power would be one way to reduce the incidence of this type of failure. Yuya Nagayama [6] of ClassNK presents other causes of turbocharger overspeeding in marine diesel engines and identifies ways of reducing the incidence of such failure mode.
CONCLUSIONS Lloyd's Register's machinery defects technical data showed that in the 10-year period to October 2005, turbocharger defects on high speed auxiliary diesel engines, medium speed 4-stroke and slow speed 2-stroke main propulsion engines, with populations of 35,776, 11,093 and 5,025 units, comprised 3%, 9% and 18% of all the engine defects recorded and were ranked the ih, 4th and 1st most recurring defects respectively from 43 engine defects categories. 302
With objectives of enhancing its clients' quality, safety, environmental and business performance, Lloyd's Register's Rules relating to turbochargers were amended in 2005. The amended Rules [1] require turbocharger manufacturer's burst test assessment to be submitted for information, in addition to the usual turbocharger plans and particulars.
© Lloyd's Register
REFERENCES: 1- Lloyd's Register RULES AND REGULATIONS FOR THE CLASSIFICATION OF SHIPS PartS - Main and Auxiliary Machinery, Chapter 2, Section 10, July 2005 with Notice No.1.
2- T Sullivan Technical Records -1979 Lloyd's Register Technical Association Paper No.1, Session 1979-80 3- A Buckland, D W Woodcraft The New Ship Division Database and its Impact on Survey Reporting Lloyd's Register Technical Association Paper No.7, Session 1992-1993
4- The Swedish Club Highlights Main Engine Damage Update (1998-2004) Sep.200S Gullbergs Strandgata 6, P.O. Box 171, SE-401 22 Gotenborg, Sweden 5- G Theotokatos, N P Kyrtatos Analysis 0/ a Large 2-Stroke Marine Diesel Engine Transient Behaviour During Compressor Surging 7th Int. Conf. on Turbochargers and Turbocharging, IMechE, May 2002. 6- Yuya Nagayama A Study o/Turbocharger overrun in Marine Engines Using FTA Method Marine Engineers Review, pp 34-37, Dec/Jan 2006.
303
A novel method of high efficiency pressure charging Anthony o. Dye Epicam Ltd.
ABSTRACT
Turbo machinery pressure charging has served to achieve useful power density over many generations of piston engines. This has been mainly due to their virtually frictionless operation and ability to achieve efficient compression and expansion, albeit within a limited range of speeds, which have no direct relationship to the speed of the engine crankshaft. Integrating a piston diesel displacement engine, with steady state turbo devices, inevitably results in mis-matches and compromises, which limit operational effectiveness. The high cylinder volume ratio required for cold starting results in overcompression of the boosted charge and excessive NOx formation. Less than 40 percent of the pressure energy available is recovered at best, leaving most of the energy source to dissipate in the exhaust system. Internal friction has, until now, rendered alternative displacement devices ineffective to recover more of this wasted resource. The novel rotary displacement devices with low friction operation and variable geometry, which are introduced here, enable full advantage to be taken to largely eliminate these losses. Maximum boost can be applied at cranking speed to provide the cold engine with adequate compression for effective starting, and can be optimally reduced thereafter. The displacement action matches that of the piston engine at all speeds, by direct coupling to the crankshaft, which also returns the surplus of recovered pressure energy to the engine output shaft. Dynamometer test results of this new technology applied to a 1.9L and a 1.2L turbodiesel are described. Gains of 20 % in fuel economy and lower NOx emissions are indicated for the 1.9L engine, at the same boost level and power output as for the standard engine. Increased boost with the 1.2L engine can yield power output equal to that of the standard 1.9L engine, 40% improvement in fuel economy and peak cylinder pressures reduced below those of its standard build. THE ROLE OF TURBOCHARGERS IN AUTOMOTIVE DIESEL ENGINES
A pre-eminent role of the turbocharger, as the preferred technology for charge boosting, is seen wherever need arises to secure increased power density from diesel engines. The rapid acceleration in the preponderance of diesel cars in Europe, taking them from 20 - 50 % by market volume, in 15 years, is in large measure due to the improvements in engine performance that have been achieved. The role of turbochargers is clear in securing performance levels making diesels attractive to car buyers. The characteristics they possess which determine their pre-eminence are: a) efficient compression of charge air, b) use of exhaust gas expansion energy source to drive the compressor and c) low mechanical friction. Despite these clear advantages, turbochargers nevertheless, have several characteristics which do not fit well with the nature of the reciprocating piston displacement machine to which they are being harnessed. The limited speed range within which acceptably high efficiency is obtainable, places considerable constraints
305
upon control systems and limitations in practice, both on the engine charging and exhaust discharging systems. The extreme case here is that of cold starting. During cranking, gas flow from the engine exhaust manifold is well below the threshold at which the rotor speed generated by the turbine is capable of adding to the ambient charge density delivered by the compressor. In this case, the cold starting situation of the turbodiesel engine is essentially similar to that of the naturally aspirated diesel engine. A nominal incylinder charge volume ratio must therefore be configured, which is capable of yielding sufficient charge compression at TDC to provide the temperature required for combustion initiation. This nominal ratio (typically in the range 18 - 20: 1) becomes translated into an effective charge compression ratio which may exceed 40 : 1 during nonnal operation of the engine at conditions above idling, where the turbocharger is capable of progressive addition to the charge density. This means that for virtually the whole of its operating speed/load range, the engine is operating with charge compression beyond that required for optimum thennal efficiency. Moreover a degree of over-compression prevails, in which significant nitrogen oxide fonnation is taking place in the combustion chamber before the start of combustion. A second aspect of the limiting characteristics of turbochargers is becoming more noteworthy in the context of higher fuel prices and increasing awareness of the need to improve thennal efficiency. The most effective turbochargers recover less than 40% of the expansion energy present in the exhaust gas stream. This means that fuel economy improvements delivered via the use of turbochargers are derived from the increase in average BMEP as a result of engine downsizing, rather than from any improvement in BSFC. If improvement is to be realised by increasing the proportion of exhaust expansion energy recovered, then it seems that an alternative to the turbine is required as a more effective gas expander. RPC ENGINE ROTOR TECHNOLOGY
The limitations of turbocharger technology outlined above, became the starting point of the present exercise to investigate the use of novel rotary displacement technology as a practical alternative for automotive diesel supercharging. The rotary power couple illustrated in Fig. 1 was developed as a means to achieve volume change by displacement in compressible fluids using simple rotary motion. The displacement is effected by interaction between the two-lobe rotor and the three-pocket rotor as they are turned in a 3 : 2 speed ratio respectively on their axes between the close fitting end walls, in which their bearings are located. The original intention was to apply the technology in the fonn of a novel IC engine, for which a prototype was developed in the late 1990's and built in 2000 (1).
306
Figure 1 Model 2-lobe, 3-pocket rotary couple with perspex end waDs The rotary couple is equally effective as a compressor using input shaft power to deliver compressed gas or as an expander to convert available gas pressure energy into rotational shaft power. Its unique features include a) an extremely short cycle of approximately 90 degrees only, from ambient to maximum delivered pressure, as a compressor or vice versa as an expander and b) variable geometry which enables the swept volume to be extended through a range of approximately 4 : I during operation. Variation in swept volume is effected by means of the sliding containment wall, whose position can readily be adjusted in the axial direction. The helical rotor form provides a leakage flow area, which reduces exponentially as the volume of the transient chamber between the rotors reduces through the cycle. This enhances the control of gas leakage, such that when the pressure of the trapped charge is at its maximum, then the leakage flow area is at its minimum. It applies equally when the couple is used as a compressor and as an expander. It should also be clear that the volumetric delivery rate during the cycle reduces exponentially rather than linearly. Thus the technology is distinguished from Lysholm screw compressor technology, where the volumetric delivery is linear with respect to cycle angle and the cycle length extends over a much greater angular range with larger leakage flow areas. Early tests were conducted in 1996 with a prototype RPC compressor having rotors of 192 mm. diameter, with clearances of 0.05 mm. and swept volume of 420 cc. per cycle. The results showed that the compressor could deliver air at pressures above 1600 kPa. with volumetric compression ratio of 12 : 1, from ambient air intake supply. Its only real disadvantage was the small size of the port, which is visible in Fig.I. The difficulty in manufacturing the machine was considered to be more of a learning curve than an inherent drawback. Subsequent design modification using a windowed disk fitted to the delivery end of the pocket rotor allowed the port to be relocated more centrally between the shaft axes where very much more radial area is available (2). The increase in port flow area obtained this way, allowed an increase of approximately 450% in mass flow capacity through the port and extended the potential operating speed range from 3,000 to 15,000 RPM. The disk was fabricated in a phenolic resin matrix 307
with Kevlar fibre reinforcement. This material provided a very satisfactory interface with the port delivery walL However, it does have a temperature maximum limit of 200 degrees C, and is therefore unsuitable for testing at higher delivery pressures. This rig was re-tested with the enlarged port in 2004, at delivery pressures of up to 500 kPa. at speeds up to 5,000 RPM (lobe rotor shaft), when the delivered flow rate was 3,100 Litres/min. A thermodynamic model was used to simulate all aspects of delivery and leakage flows through the operating cycle and to convert incremental volume and pressure change directly into work increments. When aggregated over the cycle, this gives a measure of the ideal adiabatic work performed and a comparative base for indicating actual adiabatic efficiency. The tests indicated adiabatic efficiency of 53 % at 5,000 RPM and predicted that this would reach approximately 90% in the speed range 10 - 15,000 RPM. Heat loss to the rotors is very low, due to the extremely short time in which the rotor surfaces are in contact with the working fluid.
CONCEPT DEVICE FOR SUPERCHARGING AN AUTOMOTIVE DIESEL ENGINE Considering the variable displacement characteristics, and thermodynamic performance of the rotary couple devices outlined above, it became clear that the basis exists for a concept device, known as the 'dexpressor' (3), linking both a compressor and an expander via a common drive shaft. This shaft could be directly coupled to the crankshaft of an automotive diesel engine whose reciprocating displacement characteristics could be matched to those of the device via suitable direct drive with an appropriate speed ratio. In this way, the supercharging function of the turbocharger normally associated with the engine, could be replaced by a device with low mechanical friction, comparable with that of the turbocharger. In doing so, it would add new capabilities to provide variable effective compression ratio and recovery of virtually all of the pressure energy in the exhaust gas, at all operating conditions of the engine. The high delivery pressure capability also opens the way to substantial downsizing with low thermal and physical stresses and full recovery of the increased exhaust expansion energy. A VW 1.9L turbodiesel engine of 97 Kw. was selected as the host engine for an initial demonstration project, being a type having a well-established record within the European automotive industry. This type of engine also includes a 1.2L, 3 cylinder version which could be used to demonstrate the potential for downsizing. The thermodynamic model was adapted for use to model the compression and expansion processes to be performed by the concept device, when applied on this engine. The first objective of this prototype project was defined to provide the same level of boost as that delivered by the turbocharger, thus leaving the performance of the engine/dexpressor combination essentially unchanged from that of the original turbodiesel build. A secondary objective could then be undertaken, to use the same dexpressor device on the smaller engine to achieve the same power output as that of the standard 1.9L engine. The thermodynamic analysis for the first objective could thus determine the extent of any opportunity for fuel saving which would result when the power generated by the expander exceeded that absorbed by the compressor. The result ofthis analysis is shown in Fig. 2.
308
Engine speed
=4,500 RPM
Dexpressorspeed Fuel saving
=13,500 RPM
=21.0%
Power recovery
=27.1 Kw.
-
Transien! pressure Manifold pressures
Transient volume
BOO ""--Compressor cycle
---_~
4.5
I
_
-
- - Expandercyde - - - _ Exhaust manifold pressure
500
3.5
III
400
!
~
---'1;1------+ 300
~ &
!i
-'"
u 200 1.5
~ Expander
100
port open ------III-
0.5
o
200
400
600
800
1000
1200
1400
1600
1800
2000
2200
2400
Lobe rotor angle (degs. x 10)
Fig.2. Thermodynamic model of pressure charging and exhaust expansion energy recovery
The left half of the diagram in Fig. 2 represents one cycle of the compression process, which is provided by the dexpressor and the right half depicts one cycle of the expansion process. The model assumes that the compressor is receiving air at ambient conditions and delivering to the inlet manifold of the diesel engine, which acts as a plenum, maintained effectively at 200Kpa.. This assumption holds some validity because the compressor has a 4-lobe and 6 pocket rotor combination with the lobe rotor shaft being maintained at 3 times the crankshaft rotational speed. Thus, 12 cycles of the compressor are completed for every one revolution of the engine, 24 cycles for a complete 2-revolution cycle of the 4-stroke engine. It will also be seen that whilst the compressor sweeps its 200 cc transient chamber down to only the clearance volume from start to finish of its cycle, the pressure in the transient chamber rises from ambient to a peak of about 220 Kpa. The transient pressure then settles to the 200 kPa. boost level of the engine intake manifold which acts as its receiver, for the remaining duration of the cycle. The port timing is configured so as to start opening the port 20 degrees after the start of the compression cycle and close the port at the end of the cycle when the trapped volume is reduced to its minimum clearance volume. This configuration is designed to allow the transient pressure to reach receiver pressure before the port has opened far enough to permit some back-flow from the receiver. The condition illustrated here is at maximum engine speed, so that the brief peak pressure of about 20 kPa. above receiver pressure represents an acceptable compromise when considering that at low speed, it is desirable to avoid excessive back-flow. The expansion process shown in the right half of Fig. 2 is based on the assumption that the exhaust manifold pressure upstream of the expander device, is effectively maintained at 400 Kpa.. During the cycle shown, this represents the charging pressure Through this period, the transient entering the transient chamber of the expander. volume increases from the initial clearance volume to about 500 cc. when the charge is released at near ambient pressure. It will be seen that the duration of the whole cycle is 120 degrees rotation of the lobe rotor shaft. The port timing is configured so that 309
opening takes place in time for the start of inflow of the charge gas and the port closes 70 degrees later. This leaves 50 degrees of rotation for the expansion to near ambient pressure. Of course, this means that the next expansion cycle begins 30 degrees before Similarly, there is overlap on the the end of expansion of the current cycle. compression cycle, so that the current cycle starts compression before the end of delivery of the previous cycle. PRACTICAL APPLICATION OF THE CONCEPT IN A DEMONSTRATION PROTOTYPE It was apparent that design and build of a prototype device conforming with the ideal
solution indicated by the thermodynamic model would exceed the very limited development resources available. It was therefore decided to build a prototype, which could maximise the use of fully developed parts. Gear sets having 3 : 2 speed ratio with matched bearings capable of operation at 13,500 RPM were commercially available. It was considered that the swept volume of about 100 cc which would be possible, given the 80 mm. shaft centre distance implied by the use of these components, would require a speed ratio of 4 or 5 with respect to engine crankshaft. This arrangement, though not ideal, could permit demonstration of the technology at up to about 2,200 RPM of the dexpressor diesel engine.
Figure 3 Initial design layout for prototype dexpressor device
Thermodynamic analysis could provide an indication of the fuel efficiency implications of the application but could give no indication of the effects that the re-configured engine might have on emissions, especially those ofNOx. Delivery of charge boost to 200 Kpa. indicated that the nominal compression ratio should be adjusted by about half. Before doing so, it was decided to carry out a dynamometer test programme to map the SFC and raw NOx emissions over the operating envelope of the engine. The test engine selected for the programme had completed less than 4,000 miles of normal road use since new. Results of these tests are shown in Figs. 4 and 5.
310
Max Torque (Nm)
- 400 '","sf<:
Baseline Engine - Fuel Economy Map - Specific Fuel Consumption (GmsJKwJhr.) 250
90 -
350 lso-sfc
-
3001so·sfc
80 -
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Fig. 4 Baseline Engine - Fuel Economy Map
Baseline Engine - Raw MOx Emissions Map - Iso-ppm lines
-
250 r---------------------------------------------------------------, 9C
- NOx7~
80
200 ;1 -----------~~~~~~=====================-~
Max Torque (Nm ) NOx1500 NOx125O NOx1000 NOxSOO NOx250 Power Curve (approx)
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Figure 5. Nox Emissions Map - Baseline Engine
311
The cylinder pressure diagram for the baseline engine build, which is shown in Fig. 6 illustrates characteristics which were commonly seen over most of the operating speed range of the engine at near road-load conditions. The peak pressure is high, due to the nominal compression ratio of20 : 1 and supercharging pressure at 130 Kpa., resulting in temperatures very much higher than are needed for combustion initiation and which will necessarily result in NOx formation. Pressure is thus allowed to fall by almost 25% before fuel is injected (at about 5 degrees after TDC) and combustion takes place. This diagram is contrasted with that for the engine charged by the dexpressor compressor. Here, the peak cylinder pressure is more than 50% lower, due to the nominal compression ratio of only 10 : 1 and charging pressure of about 150 kPa.. Fuel is injected at nearer to TDC but the peak cylinder pressure remains substantially lower than in the Baseline engine. Combustion pressure will be further reduced when the dexpressor expander is added to the rig and lower fuel delivery is required at this load.
Cylinder Pressure - Turbocharged vs. Dexpressor charged engine Operating Condition - 2,000 RPM at 100 Nm. Torque
-
Turbocharged engine
-
Dexpressor charged engine
5.00E+01
4.50E+01
"":" 4.00E+01
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-
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1000
1200
1400
1600
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Figure 6. Comparison of cylinder pressure diagrams - turbo vs. dexpressor charging.
This situation clearly shows scope for reducing the configuration of the nominal cylinder volume ratio, so as to reduce peak cylinder pressure. Optimisation of the effective charge compression ratio by adjustment of dexpressor's variable volume facility could yield improvement in SFC in addition to reducing NOx emissions at this type of operating condition.. The modification made to the shape and size of the combustion chamber in the piston is shown in Fig. 7. No attempt was made to optimise the chamber shape for combustion or to modify the injector spray pattern for the changed geometry but smoke records were kept for all baseline engine test points for later comparison. Early tests with the dexpressor compressor coupled to the engine showed that the engine was able to achieve successful cold starting, using the modified Peak cylinder pressure during cranking was pistons and 10 : I compression ratio Subsequent tests showed that satisfactory operation of the measured at 185 Kpa. engine can be demonstrated over a useful range of speed and load. 312
Original combustion chamber - size = 24 cc.
New combustion
chamber - size
~
52 cc.
Fig. 7 Piston crown before and after machining
PROTOTYPE DEXPRESSOR COMPRESSOR The form of the prototype Dexpressor compressor is shown in Fig. 8 and the results of air flow tests up to 7,000 RPM are shown in Figs. 9 and 10.
POCI"ET ROTOR DISK
Fig. 8 Dexpressor prototype compressor design
313
The results of the airflow tests illustrated in Fig. 8 show that there is no sign of reaching peak flow capacity at the highest speed tested, nor indeed, over the -
Dxp Compressor 64 cc. Swept Volume • Delivered Airflow YS. Speed
Flow (UMin) Flow 112 Kpa. (UMin) Flow 124 Kpa. (UMin) Flow 136 Kpa (UMin) Flow 160 Kpa (UMin)
2~ r--------------------------------------------------------------L==JF~lo~w~2~0~ 0~ K~pa~(~ UM ~ inlU )
2000 ~------------------------------------------------------------~~~
C
1500 +-----------------------------------------------~~~~~~~--------_1
~ ~
'E ~ 1000 +-----------------------------~~~~~~~--------------------------_1
~ ~---~~~~~--~~~-----------------~
2000
1000
3000
4000
5000
6000
7000
Speed (Lobe Rotor RPM)
Fig. 8 Dexpressor Compressor Air Flow Delivery - - VEat112Kpa - - VE at 160 Kpa - - VE at 200 Kpa Adiabatic Effieiencyat 112 Kpa. (%) Adiabatic Effieiencyat 160 Kpa (%) , Adiabatic Efficiency at 200 Kpa (%)
Dexpressor Compressor 64 cc. Swept Volume Volumetric and Adiabatic Efficiency 120 110 100
60 90
§:
sol g
1;- 80
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m 40 10
30
m 1000
1500
2000
2500
3000
3500 4000 4500 _ (L_ RotorRPII)
5000
5500
6000
6500
0 7000
Figure 9. Dexpressor Compressor volumetric and adiabatic efficiency
314
performance range predicted by the thermodynamic model and shown in Fig. I O. It was not possible to bench test the prototype device at any higher speed because the drive was already at the maximum belt speed with the 2 : 1 drive ratio used. However, the belt speed for the engine drive is via a 4 : 1 drive ratio which will substantially reduce the belt speed and allow testing to over 2,000 RPM of the engine. -+-
110 cc_
sv at2 Bar
~ : : ~:~: :: ~.~:r~~odel)
Dexpressor Airflow Delivery Performance· Swept Volume 110 cc., 64cc., and 40 cc. Data from calibrated Thermodynamic Model
__ 64cc. s.v. at 2 Bar (Actual)
7~ '-------~~~~~~~~~~~----------~==~~~======~
6000 +-----
o ~--~--~--~--~~~~--~--~--~--~--~--~--~
2000
3000
4000
5000
6~
7000
8000
9000
1~
11000
12000
13000
14000
15000
Speed (Lobe Rotor RPM)
Figure 10. Actual and predicted dexpressor compressor airflow performance PROSPECTS FOR DEXPRESSOR APPLICATIONS It will be clear from the results shown, that the work is still at an experimental stage.
The expander is a very similar device to that of the compressor and its thermodynamic efficiency is unlikely to be found to differ significantly from that of the compressor. It therefore seems likely that up to half of the 21 % potential gains in specific fuel consumption which are indicated by the thermodynamic model for this engine application, will be realised with the present design of dexpressor components. The fact that the dexpressor compressor is able to offer sufficient charge boost at cranking speed to effect cold starting of the engine, opens up the possibilities for exploitation of new capabilities of commercial diesel engines. It becomes possible to operate the diesel engine with no greater compression than is ever required in order to achieve satisfactory combustion of the fuel. This implies m~or reductions in the generation of NOx in the combustion chamber when compared with today's turbodiesels. The substantial reduction in peak cylinder pressures also presages the development of a new generation of compact, lightweight diesel engines, capable of service operation at significantly higher BMEP and yet greatly reduced thermal and physical stresses compared with present day engines.
315
REFERENCE LIST
1
A.O. Dye, Rotary Fluid Engine, UK Patent No. 0500597, November 1990
2 A.O. Dye and M. Ardron, Rotary Device and an Engine, PCT/GB2005/001807 International Patent Application, November 2005 3 A.O. Dye, An engine and an apparatus for providing forced aspiration to an engine, WO 2005/080767, September 2005.
316
Turbine wheel design for Garrett advanced variable geometry turbines for commercial vehicle applications Dr Hua Chen Honeywell Turbotechnologies Ltd., Skelmersdale, UK NOTATION AVNT Advanced variable nozzled turbine BPF Blade Passing Frequency HCF High cycle fatigue Trim Square of inlet to exit diameter ratio of compressor wheel time 100 VNT Variable nozzled turbine SYNOPSIS
This paper describes the design and tests of a turbocharger turbine wheel for modern commercial diesel vehicle applications. The requirements of such applications are first discussed. The aerodynamic and mechanical designs of the wheel are then described and the test results of the wheel in a Garrett Advanced Variable Nozzle Turbine (A VNT) are presented. For such turbines, turbine wheel high cycle fatigue (HCF) is a serious threat to survive of the wheels. A right-at-the-first-time result was achieved through a combination of experiences, state-of-art simulations and experimental technologies.
INTRODUCTION Variable geometry turbines are used on commercial diesel vehicle applications to increase boost at low engine speeds and to reduce engine back pressure at high engine speeds. The boost is achieved by closing the flow guidance device (typically nozzles) upstream of the turbine wheel to increase the turbine inlet pressure. This increases the pressure at engine exhaust valve (engine back pressure). At the low cngine speeds this back pressure is relatively low so some increase of it does not affect engine performance and does not affect turbine life and reliability. When the engine speed is high, if the nozzles are closed, a high pressure drop can occur in the nozzles because of a high mass flow rate through a highly restricted flow passage. Engine back pressure is further increased. This phenomenon is utilised to slow down the engine for vehicle braking purpose (engine braking). Under the engine braking condition, shockwaves are often generated from the nozzle exits because the nozzles are usually choked at their throat and the flow further accelerates downstream along the pressure side of the nozzles. Fig. 1 shows CFD simulation results of Mach number distribution around the nozzles of a Garrett A VNT under engine braking condition. A strong shockwave is generated from the pressure side after the throat. These shockwaves impinge on the downstream turbine wheels, creating an excitation force with blade passing frequency (BPF), this force and its higher order harmonics can easily fail a turbine wheel. Fig. 2 shows the static pressure circumferential variation at the turbine wheel inlet (such variation has been confirmed by static pressure tapping measurement), and Fig. 3 shows the first three harmonic components of the shockwave. Fig. 4 shows a turbine wheel whose inducer blades were damaged by the BPF (first harmonic) component of such force and Fig. 5 shows another turbine wheel whose exducer blades were damaged by the second harmonic of the force. 317
The turbine wheel to be discussed was designed for working with the type of nozzles showing in Fig. I with engine braking capability. For structural integrity reason the nozzle shape could not be changed. In addition to survive the severe engine braking condition, the new wheel needs to be more efficient, lighter (for better transient response) and to have higher speed capability. AERO & MECHANICAL DESIGNS OF WHEEL Basic wheel geometry
The major geometrical parameters of the new wheel and baseline wheel are given in Table 1. Table 1 Summary of new and baseline wheel geometry
New wheel Tip diameter 72.Smm 13 Number of blades Inducer clipping ISdeg (mixed flow) Tip width (axial) l2.23mm Backdisk scalloping no Trim 84 Exducer hub diameter 24.6mm 27.lmm Axial length (hub) Relative design speed 106%
Baseline 72.Smm 13 non (radial flow) 12.23mm Yes 84 29.36mm 36.37mm 100%
A high blade count of 13 was mainly chosen for wheel BPF noise reason. High blade number increases the frequency of this noise making it less sensitive to human ears. The nozzle assembly of the turbine has 9 nozzles so this choice of rotor blade number also produces a nodal diameter of 4 that enables the vibration energy of the wheel caused by the upstream shockwave to be self contained in the wheel. To avoid HCF under engine braking condition played an important role in deciding the basic geometry of the wheel. The inducer was clipped by ISdeg for this reason. This enabled a non-scalloped (full) backdisk to be used to provide a better support to the inducer. This made it easy to avoid the inducer flap vibratory mode and to increase the natural frequencies of any inducer modes above the excitation frequency of nozzle shockwave (= wheel rotational speed x nozzle number). Although it is possible to achieve a similar result with scalloped radial turbine wheel, the design work would be much harder. Mechanical design of the wheel
With inducer natural frequencies clear of the first harmonic of the excitation force, the wheel must then survive the excitation of the second harmonic. This is a much harder job because resonance caused by the excitation is unavoidable. Here the inducer is saved because there are no isolated inducer modes due to the inducer clipping and the use of full backdisk. It is the exducer that is at risk. It is impossible to design an exducer whose natural frequencies are all higher than those of the second harmonic excitation 318
while still maintaining reasonable aero performances. In this particular case, the second harmonic has frequencies of wheel rotational speed time 18. Usually the excitation is strongest when the engine and turbine speeds are highest. To avoid exducer blade failure under resonance condition, the blade thickness of the exducer was carefully selected. An experience based method, using results from simple modal analysis, enabled different designs being quickly judged and a final design rapidly obtained. This design was then checked by a more elaborated simulation. In this simulation, the 3-D unsteady viscous flow of turbine nozzle and wheel assembly was first calculated so that the unsteady forces acting on the turbine wheel blades could be obtained. These forces were then fed into FEA simulation to obtain the wheel blade responses to the excitation in strain with an appropriate material damping factor. The wheel strain under combined centrifugal and thermal load was also calculated. The two types of the strains were then superimposed on the Goodman diagram of turbine wheel material and the HCF life of the wheel judged. Fig. 6 shows the Goodman safety factor of mode 2 (a blade vibration mode, see Fig. lO(a) for its mode shape) of the wheel under the first harmonic excitation and in resonance condition. The lowest safety factor is located at the exducer but is considered to be safe. The inducer safety factor is very high. Note that the frequency of this mode is never reached in real operating condition (the frequency is higher than the highest excitation force frequency in the operation), so the resonance would not occur in the real operation. However the simulation result gave confidence that the inducer would not fail in the operation before the wheel was actually made. Fig. 7 shows the same safety factor of mode 4 (a blade mode gave the lowest value of the safety factor, see Fig. lO(b) for its mode shape) of the wheel under the second harmonic excitation and in resonance condition. With this harmonic, the resonance condition does occur in the real operation. Again the wheel was judged to be safe. Aerodynamic design of the wheel Clipping of the inducer of a wheel generally increase turbine efficiency and flow capacity at high speeds but reduces turbine efficiency at low speeds. To compensate for the efficiency loss at low speeds, which comes as results of increased positive incidence and increased flow velocity, the inducer of the wheel was non-radially stacked to introduce some positive blade angles. The non-radial stacking needs to be done discreetly to avoid high blade mechanical stress. Tip3d CFD code by John Denton was used to quickly calculate turbine wheel performances and blade loading at two operating points of targeted engine. Geometry of the wheel was modified accordingly. A commercial CFD code was used to check the final few designs and to fine tune the design. All the CFD design work was done in steady state. No transient simulation was done to evaluate turbine transient performance. However, the clipping of the inducer alone would almost ensure a better transient response than the baseline radial turbine. In addition, the new wheel is lighter and more efficient. TEST RESULTS Performance comparison The new turbine wheel was tested in a gas-stand back-to-back with the baseline turbine wheel in a Garrett AVNT. Fig. 8 compares the performances of the two turbines. The 319
new wheel has higher efficiency over entire turbine operating range and it also has higher flow capacity. HCF test
The wheel natural frequencies were first measured. In addition to the more conventional pinging test, a laser scanning vibrometer was also used to obtain the blade natural frequency and mode shapes. This gave more accurate results particularly to the high frequency modes. Fig. 9 shows the setup of the vibrometer and Fig. lOa and Fig. lOb show the mode shapes of blade second and fourth modes. A light probe instrument capable of measuring the blade deflection in real gas-stand test condition was used to further identify modes with large amplitudes. Measured blade natural frequencies were plotted in Campbell Diagram Fig. 11 and the possible excitation frequencies identified. Garrett has several specially built, dedicated HCF test rigs whose sole function is to test HCF life of turbines. These rigs can ramp up and down turbine speed around potential resonance frequencies at specified turbine pressure ratios. Body born noise and gas noise are monitored during the test and resonances of the turbine wheel blades can be identified. At least 10 million cycles must be reached without failure before a pass is declared. Fig. 12 shows one of such rigs used for smaller passenger car turbocharger turbines. Several new turbine wheels were tested in these rigs under an inlet pressure greater than 4.5bar and passed. CONCLUSIONS Nozzle assembly of variable geometry turbines under engine braking condition can generate strong shockwaves impinging on downstream turbine wheels, causing resonances and damage to the blades. They also reduce turbine efficiency. Often, these shockwave can be weakened by redesigning the assembly. But in this case this could not be done due to other constrains. A new turbine wheel was designed to survive such severe working condition. A mixed flow type of inducer without scalloping was selected to provide support to the leading edge and blade thickness was carefully chosen to reduce blade stresses under resonance conditions. A simple, experience based method enabled different designs being evaluated quickly. A state-of-art, sophisticated simulation method was used to check the final design before hardware was made. These measures ensured the design pass the HCF test in the first place and no more design iteration was needed. With help of modern CFD, aerodynamic design of the wheel was done through a combination of performance calculation and blade loading and flow filed evaluation. This produced a lighter design with better performances. ACKNOWLEDGEMENTS The author wishes to thank his North American colleagues for their supports in this project, in particular Dr Fazal Rahman for providing the HCF calculation. He is also grateful to Dr William Connor of Honeywell Turbotechnologies, UK for his supervisory support.
320
REFERENCES 1.
2.
G T McDonnell & Q D H Roberts, Design of a turbine rotor for a high-vane count variable geometry turbocharger, i h International conference on turbochargers and turbocharging, IMechE, C602/01212002, 14-15 May 2002. D Filsinger, M Sekavcnik, T Ihli, A Schulz and S Wittig, Vibration characteristics of a radial turbocharger impeller, i h International conference on turbochargers and turbocharging, IMechE, C602100212002, 14-15 May 2002.
321
2.16e+00 2.07e+00 2.018+00 1.94e+00 1.888+00
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Contours of Mach Number
Jun 17, 2003 FLUENT 6.1 (2d, dp, ooupled imp, rke)
Fig. 1 Shockwave generated by turbine nozzles under engine braking condition by CFD (Turbine wheel inlet shown as a white circular arc)
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322
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Fig. 3 Harmonic components of wheel inlet pressure under engine braking condition by CFD
Fig. 4 Turbine wheel inducer failure under engine braking condition
323
Fig. 5 Turbine wheel exducer failure under engine braking condition .ANSYS 7 . 11e1 J\OIl 8 2004 PWT NJ . 13
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Gc:xx:lrrBn Safety Factor; Forero Resp:mse; W.xle 2 ,
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Fig. 6 Goodman safety factor of blade mode 2 in resonance condition
324
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Fig. 8 Comparison of measured turbine performances at const. Pressure ratios
325
1.05
15
Fig. 9 Set up of laser scanning vibrometer
(a) 2nd mode (b) 4th mode Fig. 10 Mode shapes of blade second and fourth vibratory modes (Red color has the smallest displacement and blue color the largest displacement)
326
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Fig. 11 Turbine wheel Campbell diagram
Fig. 12 HCF rig for testing HCF life of small variable geometry turbines
327
Compact Long-Route Exhaust Gas Recirculation Mixer Design and Optimization lunfei Yin, Nicolas Deschatrettes, Ocean Han, Philippe Renaud Honeywell Turbo Technologies Ltd., Skelmersdale, UK NOTATIONS d ---- Gas pipe diameter D ---- Main pipe diameter Pt ---- Total pressure Tt ---- Total temperature U ---- Velocity W ---- Mass flow rate 'Y ---- Uniformity Index A ---- Total pressure loss coefficient p ---- Density Subscripts 1 ---- Fresh air inlet 2 ---- Hot gas inlet SYNOPSIS
Reducing engine emissions becomes a very challenging target in automotive industries. Long-route exhaust gas recirculation (EGR) is considered as one potential solution to reduce emissions. The hot-gas is injected back into the turbocharger compressor wheel inlet. The hot gas will cause strong distortion in temperature and pressure fields. From the aerodynamics point of view, that will result in compressor performance loss and potential wheel overheating. In this paper, the design and CFD-optimization of long-route EGR mixers are discussed. Several design configurations were investigated. One selected design was optimized by changing different design parameters to achieve minimum loss and maximum uniformity of compressor inlet temperature. INTRODUCTION
Exhaust gas recirculation (EGR) has been used in recent years to reduce NOx emissions in light duty diesel engines. EGR involves diverting a fraction of the exhaust gas into the intake manifold where the recirculated exhaust gas mix with the incoming air before being inducted into the combustion chamber. EGR reduces NOx because it dilutes the intake charge and lowers the combustion temperature [1]. The US environmental protection agency has reviewed a variety of ways to implement EGR [2]. For passenger vehicles, many systems are under development. A long-route EGR mixer was designed in [3], but the configuration is not suitable for use with a turbocharger compressor. There are two EGR solutions based on such concept [4]:
329
(1) Long route system shown in figure 1: In a long route system, the pressure drops across the air intake and the stagnation pressure in the exhaust gas stream makes the EGR possible. Compressor Aftercooler
Figure 1 Long-route EGR layout (a filter not presented)
Compressor
Cooler
Turbine Figure 2 Short-route EGR layout
(2) Short route system shown in figure 2: These systems differ mainly in the method used to set up a positive pressure difference across the EGR circuit. Another way of controlling the EGR-rate is to use variable nozzle turbine (VNT). Most of the VNT systems have single entrance, which reduce the efficiency of the system by exhaust pulse separation. Cooled EGR should be supplied effectively. Lundquist[5] and others used a variable venturi, in which EGR-injector was allowed to move axially, thus varying the critical area. In the short-route EGR, exhaust gas is taken from after the exhaust valve and passed through a valve to the inlet of the engine. The long-route or low-pressure EGR, can also be used as a retrofit solution. The NOx reduction can be up to about 30-40%. In this case, particulate free exhaust gas is led from after a particulate filter (not shown here), then is cooled and introduced before entering the turbo. The hot-gas is injected back into the compressor wheel inlet in the turbocharger. The hot gas could cause strong temperature and pressure distortion at the compressor wheel inlet. From aerodynamic point of view, 330
that will result in compressor performance loss and potential wheel overheating. The following requirements are essential for an EGR mixer: Compact design due to limited space in a passenger car (the ratio of mixing length to pipe diameter is less than 1.0 ) Minimum flow restriction during the intake process, i.e. minimum (total) pressure loss from air inlet to compressor wheel inlet. Minimum total pressure loss from EGR hot gas inlet to compressor wheel inlet. In the EGR circuit, the loss is related to turbine power An acceptable mixing quality. The objective of this work is to design and optimize long-route EGR mixers in order to achieve the essential requirements. The CFD tool FLUENT was used to evaluate different designs. BASIC DESCRIPTION AND PERFORMANCE PARAMETERS Basic mixer model and conditions
The gas conditions for the fresh air inlet are ambient conditions: ambient pressure, ambient temperature; the mass-flow rate is 85% of the total mass-flow rate. The recirculation gas has a total Temperature of 500 Kelvin, and represents 15% of the total mass-flow rate. Two families of variables are defined to characterise the mixing process: Pressure loss coefficient, and Uniformity index. They are defined as follows. From the hot gas inlet to the outlet: /..(gas)= (Pt(gas inleW P1(outlet average»)/
{0.5 pU2 air inlet}
From the fresh air inlet to the outlet: /..(air)= (Pt(air in1et)~P1(<>ut1et average»)/ {0.5pU 2 i1ir inlet}
Uniformity Index y: y is the flow distribution index which is 1.0 for a uniform flow[6], and decreases with decreasing flow uniformity.
331
P~gas),
1'(gas), W(gas)
,~
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Pt(average) Tt(average) Total pressureloss coefficient A. Uniformity Index y
i
1,00
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Figure 3 Long-Route EGR mixer model
1~ r=l--L.,
2
i=1
Ailcv; - Vavg)1 A· Vavg
Where Ai is an elemental area and Vi can be Pt (Total Pressure) and Tt (Total Temperature) to calculate the index 'Y (Pt) or'Y (Tt). The higher the 'Y parameter, the more uniform the flows. 'Y gamma uniformity index is an integral measure of flow uniformity. These metrics can be used to define specific "pass/fail" criteria; the current long-route EGR design will fail if any of the indices falls outside a preset range. While these measures can be determined in FLUENT, it is preferable to have a post-processing utility which automates the process. A code has been written for this purpose. The flow uniformity characteristics of the long-route EGR mixer can therefore be quickly and easily determined. Models
Three cases have been examined, each with d/D=0.5. Case 01 is the simplest model in which there is a main flow pipe, the fresh air flows in and then mixes with the hot gas flow. The gas pipe is normal to the main pipe with a small diameter. In Case 02, there are a number of holes on the main pipe wall and the hole positions and areas are well distributed in order to achieve better mixing and smaller loss. The total area for the small holes are equal to the hot-gas pipe area. Case 03 has a Circumferential Slot on the main pipe wall; the highest and smaller slot widths are 32% and 10% of the hot-gas pipe diameter. CFD RESULTS AND ANALYSIS FOR THE STARTING MODELS
To take the best result from the starting models, the refined designs were made to optimise the number, area, position of the holes and other design parameters. The results on the three stations: 0.5D, ID and 2D were taken downstream from the centre of the hot gas pipe in figure 3, where D is the pipe diameter of the main pipe. The 332
main focus is on the 0.5D station, which represents the shortest EGR mixer and a compact design (the mixing length is.0.5D).
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Figure 5 Total-pressure loss coefficients at the O.SD station for the starting models Total pressure loss
Figure 5 presents the pressure loss coefficient at the O.5D station. The air total pressure losses are similar and around 0.5 in the three cases, but the hot gas total pressure losses are quite different. Case 01 presents the lowest total pressure loss at around 0.5; Case 02 gives the highest total pressure loss and is close to 4.0; Case 03 also presents very high total pressure loss. Therefore, the hot-gas total pressure loss is sensitive to the different designs. The air total pressure loss is not.
333
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Figure 6 The 'Y parameters for the starting models
Distortions in total pressure and temperature Figure 6 presents the y parameters for total pressure and total temperature at the 0.5D station. Case 01 presents the lowest y parameters; Case 02 gives the highest. Compared with Figure 5, it is clear that the higher the total pressure loss, the higher the y parameter. That consequently means the more uniform the flow is. The first priority of the EGR mixer is to reduce the temperature variations at the inlet of the compressor, in order to have the smallest impact possible on compressor wheel life. Case 02 was thus selected for further optimization; the purpose of the optimization is to improve further the temperature distribution by improving the mixing between the two gases, and to reduce as much as possible the losses across the EGR mixer. Case 02 can also keep the manufacturing costs down.
334
0.5
0.0
U/U max
Figure 7 Cross section view of the velocity contour at X = 0 1.0
1.0
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0.0
~.max
- P"min
Figure 8-a. Cross section view of the total pressure contour at the 0.5D station
0.0
1;._ - 1;.min
Figure 8-b. Cross section view of the total temperature contour at the 0.5D station
Analysing the flow field in Case 02, which is shown in figure 7; the velocities at the entry to the ring channel and before entering each hole are very high and those will cause higher losses in the hot-gas pipe. Figure 8 presents cross sectional views of the total temperature and total-pressure contour at the 0.5D station. There are big hot spots at the opposite side of the hot-gas inlet. This suggests that the distribution of the holes and their areas need to be optimized. FURTHER OPTIMIZATION OF CASE 02 In the next optimization steps, a number of design parameters were studied such as the angle, area and number of the holes, centre offset of the ring pipe and hot-gas pipe diameter as shown in figure 9. The final geometry, with 10 holes is also shown in figure 10. This was achieved after several designs iterations were performed.
335
Angle of 1IIe hole
Figure 10 Final Geometry
Figure 9 Optimisation plan
In the final design, the mass flow through each hole is roughly the same. The outer ring centre is off set about 5%D in order to reduce the losses.
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Figure 13 Total-pressure loss along the Figure 14 y parameters along the mixer mixer length length
336
A reasonable uniformity of y(Tt) >=0.96 and acceptable total-pressure losses have been achieved by varying the design parameters Figures 11 and 12 present the cross section view of the total temperature and total pressure contour at the 0.5D station. Compared with figure 7, the uniformity has been significantly improved and the hot spots have been weakened. The hot spots are well distributed and those profiles are likely to have a smaller impact on the compressor performance and wheel life due to the large reduction in the peak temperature. Influence of the mixing length Figures 13 and 14 present the influence of the mixing length on the mixer performance. The total pressure losses gradually rise along the mixing length. The uniformities have been improved, especially for y(Tt). That means that the longer the mixing length, the better the uniformity. Influence of number of the holes Figures 15 and 16 present the influence of the number of the small holes. The number of holes has very little influence on y(Pt), but y(Tt) increases with the number of the holes. The total pressure loss decreases significantly with the number of the holes in the same total hole area. 1.00
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Figure 16 Total-pressure loss coefficient
337
U/U max
1.0
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Figure 17 Cross section view of the velocity contour at X = 0 Influence of outer ring centre offset Figure 17 presents the velocity magnitude contour at X = O. At the entry to the outer ring, the flow area is bigger to cope with the larger mass flow because the hot gas mass flow decreases as the hot gas flows into the holes. The gas speed becomes evenly distributed along the ring channel, so the loss decreases. Due to the smaller-size hole, the velocities in each hole increase. Figures 18 and 19 present the influence of outer ring centre offset by about 5%D. This off-centre effect improves y(Tt) and reduces the total pressure loss significantly. Because of the offset, the mass flows are well distributed in each hole. The best design has 10 holes and an outer-ring centre offset of 0.05D. However, the current study is only based on CFD simulation. Further tests are needed to confirm the design and also to evaluate the reliability of the design. A LCF tests series has been arranged to check the mechanical and aerodynamic constraints in the mixer and compressor wheels. 0.962
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0.050
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Figure 19 Total-pressure loss coefficient 338
CONCLUSION
A compact design of long-route exhaust gas recirculation mixer has been achieved by optimization of design parameters such as the number of the holes, their position & angle as well as the area of the holes, outer-ring centre offset, and the hot gas inlet pipe configuration. The best design has 10 holes and an outer-ring centre offset of 0.05D. The design has achieved: - Compact design (the mixing length is within half of the main pipe diameter). - Minimum flow restriction - Acceptable total pressure loss from EGR hot gas inlet to compressor wheel inlet, about 1.5 times of the main fresh are pressure dynamic head. - Acceptable mixing quality. The total-temperature uniformity index is greater than 0.96. Further LCF tests are required to evaluate the reliability of the design. REFERENCES
1. A Gorel, J L Calabrese, Low pressure EGR system for diesel engines, US Patent Office, Pat No. 6 301887Bl, 2001 2. United States Environmental Protection Agency, 'Regulatory Impact Analysis: Control of Emissions of Air Pollution from Highway Heavy-Duty Engines', Air and Radiation, EPA420-R-00-OI0, 2000 3. H Gordon, D Holze, E L Keith, L M Samuel, Exhaust gas recirculation mixer apparatus and method, US Patent Office, Pat No.6 427671Bl, 2002 4. S Kohketsu, K Mori, K Sakai, T Hakozaki, 'EGR technologies for a turbocharged and inter-cooled -duty diesel engine', SAE 970347, 1997 5. U Lundquist, G Smedler, P Stalhammar, 'A comparison between different EGR systems for HD diesel engines and their effects on performance, fuel consumption and emissions', SAE 2000-01-0226, 2000 6. H Weltens, H Bressler, F Terres, H Neumaier, D Rammoser, 'Optimisation of catalytic converter gas flow distribution by CFD prediction", SAE paper No 930780, 1993. © Honeywell Turbo Technologies Ltd.
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Transient Performance Prediction of the Turbocharging system with the Variable Geometry Turbochargers Hiroshi Uchida, Akinobu Kashimoto, Yuzi Iwakiri TOYOTA Central R&D Labs. Inc.
1. INTRODUCTION Recently, for reducing C02 emission, the development of the high boost turbocharged engines are proceeding energetically to improve the automobile fuel consumption by means of the downsizing and lean boost. In order to improve the low-end torque and the transient response of the turbocharged engine, various turbo charging systems are developed. For example, variable nozzle turbocharger, electric assists turbocharger, electric boost compressor, variablc geometry compressor, and two stage turbocharging system, etc. The engine-turbocharger matching and control of those advanced turbocharging systems are more complicated than those of conventional system. Therefore it is important to predict both the steady state characteristics and transient characteristics of the advanced turbocharging systems in short calculating time for the development of the turbocharged engine with the control system. The wide flow range compressor both with the Variable Inlet Guide Vanes and the casing treatment has been developed in our laboratories. The synergy effect of the VIGV and the casing trcatment on the surge limits was found and its effect led to improve the surge limits significantly more than the individual effect of VIGV. Accordingly, it was possible to raise the boost pressure at low engine speeds and the significant improvement of the low-end torque was expected. It was also expected that the transient response of the engine torque would be improved by the optimum control of the VIGV. In order to predict the effect of the variable geometry turbocharger on the tubocharged engine performances, the one-dimensional simulation model of the turbocharging systems have been developed. It has the feature that the calculating accuracy of the transient characteristics of turbocharging systems is independent of the calculating time step within the range of b. t=O.1 to 0.8 sec. Accordingly, the transient characteristics after the arbitrary time can be predicted in a short calculating time. The simulation evaluated the effect of the VIGV on the characteristics of the turbocharging system. The result shows that the significant improvement both of the boost pressure at low engine speed and transient response of the turbocharger are achieved by means of the synergy effect of the variable inlet guide vanes and the casing treatment. 2. DEVELOPMENT OF THE VARIABLE GEOMETRY COMPRESSOR On the purpose of improving Ai the surge limits, the turbocharger compressor with the VIGV and the casing treatment has been studied experimentally and numerically. Fig.l shows the Fig.1 The structure of the compressor with VIGV structure of the compressor. and the casing treatment
341
The VIGV is located upstream of the casing treatment and be able to control the prewhirl of the impeller inlet flow. The casing treatment is composed of 6-cavities enclosed by the spoon-like curved wall and the two slits located at the shroud of the compressor housing. The shape of the cavities was designed to change smoothly the velocity of the re-circulating flow from circumferential direction to axial direction. 2.1 Experimental Result
100
r---I--:~~-:::::~F;::~-i"I
The performance of the compressor was measured by using the turbocharger performance test equipment on condition that the turbine inlet temperature was 873K. The 3.5 surge limits were detected by sensing the frequent pressure 3.0 oscillation at the compressor o .~ exit. The effect each of the 0::: 2.5 casing treatment, the vane ...::J setting angle of VIGV and the en ~ 2.0 impeller backward angle on the a. compressor performance was 1.5 evaluated experimentally. The location and the width of the 1.0 two slits of the casing treatment o 25 50 75 100 have been searched Grected Air Flow Rate (%) experimentally in order to improve the surge limits. The vane setting angle of the VIGV Fig.2 The experimental result of the compressor with VIGV and the casing treatment was changed from zero to 80 degrees from axial direction. Fig.2 shows the experimental results of the compressor. The figure shows the total to static performances and the relative efficiencies based on Casing Treatment at 100%. The surge flow rate of the compressor with the casing treatment was reduced 36% compared with the conventional compressor without the casing treatment. Furthermore, the compressor efficiency also was improved by the casing treatment. The effect of the VIGV vane setting angle on the surge limits of the compressor with the casing treatment also is indicated in Fig.2. The surge flow rate is decreased as the VIGV vane setting angle increases. So the surge flow rate is reduced 43% of the conventional compressor over the pressure ratio of 2.5 by means of the VIGV vane setting angle is changed from o to 80 degrees. On the other hand the individual effect ofVIGV on the surge limits has been investigated experimentally and the results shown that the reduction of the surge flow rate was less than 15%. Therefore it is thought that the drastic improvement of the surge limits is caused by the synergy effect of the VIGV and the casing treatment. Additionally, the influence of the impeller backward angle on the surge limits has been investigated experimentally about the compressor with the VIGV and the casing treatment. The result shown that the surge limits on the VIGV vane setting angle of 80deg was not changed by decreasing the backward angle from the radial direction. But both of the surge flow rate and the pressure ratio at the same rotational speed on the VIGV vane setting angle of 0 degree were increased by decreasing of the backward Q)
342
angle. As the results, the significant wide flow range compressor with the low centrifugal stress has been developed. It is expected that the developed compressor will improve both of the low-end torque and the transient response. The efficiency of the compressor with the casing treatment is a little bit higher than that of the conventional compressor in the low flow rate region. It is thought that the improvement of velocity distribution in the impeller caused by the casing treatment raises the compressor efficiency despite the increment of the re-circulating flow rate. The pressure loss through the VIGV increases as the vane angle increases in the high flow rate region. So the compressor peak efficiency decreases as the VIGV vane angle increases. But the efficiency in the low flow rate region does not decrease as shown in Fig.2. 2.2 Numerical Analysis The flow in the Impeller Trailing edge compressor have been analyzed by the 3-D computational fluid dynamics in order to investigate the reason why were the surge limits improved drastically Number of Mesh·1020QOQ by synergy effect of the VIGV and the Calculation mesh Conventional (45% flow rate) casing treatment. Fig.3 shows the calculation mesh and CT CT the result of the CFD analysis. In the case of the conventional compressor, the recirculating flow region is revealed widely near the tip clearance along the With Casing Treatment With VIGV and Casing Treatment impeller flow pass on (35% flow rate) (35% flow rate) condition of the surge flow rate. The Fig.3 The result of the CFD analysis of the compressor re-circulating flow with VIGV and the casing treatment region becomes small by use of the casing treatment but the re-circulation still occurs downstream the slit of the casing treatment. Furthermore the re-circulation flow almost disappears by the synergy effect of the VIGV and the casing treatment. Accordingly it is thought that the re-circulating flow near the tip clearance is one of the primary factors relating to the surge occurrences. 3. SIMULATION METHOD OF TURBOCHARGED ENGINE
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The I-dimensional performance prediction program of the turbocharged engine was developed to investigate both the engine-turbocharger matching and the transient response of the variable geometry turbocharging systems in short calculating time. It is composed of the simple engine model and the turbocharger model, which have the compressor model and turbine model. 3.1 Turbocharger Model The turbocharger models have been developed to be able to predict the performances of the compressor with the VIGV and the variable nozzle turbine by inputting the turbochargers' dimensions. The velocity triangle each of the compressor and the turbine is calculated by the one-dimensional flow calculation with the conservation of the angular momentum and with the pressure loss models. And, the pressure loss models have been developed by the one-dimensional flow analysis of test data obtained in our laboratories on the many kinds of turbochargers. The VIGV vane setting angle is related with the impeller inlet velocity triangle and the VN nozzle vane setting angle is related with the turbine rotor inlet velocity triangle. So the effect of the control both of the VIGV and variable nozzle on the turbocharging characteristics. The mechanical loss is in proportion to the square of the rotational speed. 3.1.1 Compressor Model The compressor input power Lc is shown as follows. Lc = Ga . Cpa(T3 - TO)
(I)
Ga:air flow rate(kg/s), Cpa:air specific heat(kcallkgK), TO,T3(K):temperature at inlet and outlet. The enthalpy rise in the compressor is shown as follows equation according to the angular momentum theory. ]·Cpa(T3-TO)=(U2·Cu2-Ul·Cul)/g (2) UI,U2 : circumferential velocity of impeller at inlet and exit(rnIs). Cul,Cu2: circumferential velocity of the flow at impeller inlet and exit(rnIs). J : the mechanical equivalent ofheat(kgrnlkcal). g: gravitational acceleration(rnIs 2) The adiabatic pressure at the compressor outlet on the condition of equal entropy is shown follows. P3ad=PO{(T3/TOtl (K-1L l} (3) PO : the compressor inlet pressure(kPa). K : specific heat ratio The compressor outlet total pressure is shown as follows by use of the total pressure loss f). P. P3=P3ad-~P
(4)
The total pressure loss is modeled as follows. ~P=~Pvigv+~Pinc+~Pfric+~Pch+~Pdif
(5) •Pressure loss in the VIGV ~Pvigv=func.(mach number, static pressure, vane setting angle) · Impeller incidence loss ~Pinc= func.(incidence angle, relative mach number, static pressure) · Impeller friction loss ~Pfric= func.(relative mach number, diffusion factor, static pressure)
344
· Impeller choke loss ~Pch= func.(mach number at the throat, incidence angle, static pressure) •Pressure loss in the diffuser ~Pdif= func.(mach number, inlet flow angle, static pressure) The compressor performance is calculated by inputting the air flow rate, the rotational speed, and the dimensions. This compressor model has the feature that the characteristics on the condition of negative efficiency can be predicted. The negative efficiency condition occurs when the turbocharged engine is operated at the idle. 3.1.2 Turbine Model The turbine input power Ltad is shown as follows. Ltad=Gg'Cpg' T4{l-(P4/Ps6r(Kg-I)/Kg} (6) Gg : gas flow rate, Cpg : gas specific heat, T4 : turbine inlet temperature. P4 : turbine inlet total pressure, Ps6 : turbine outlet static pressure. The enthalpy drop in the turbine is shown as follows equation according to the angular momentum theory. J. Cpg(T4 - T6) = (US' Cu5 - U6' Cu6)/g (7) U5,U6 : circumferential velocity of turbine rotor at inlet and exit(m1s). Cul,Cu2 : circumferential velocity of the flow at turbine rotor inlet and exit(m1s). The adiabatic pressure at the turbine outlet on the condition of equal entropy is shown follows. P6ad=P4{ 1 -(T4/T6r Kg/(Kg-I)} (8) Kg : gas specific heat ratio The turbine outlet pressure is shown as follows by use of the total pressure loss tJ. P. P6 = P6ad - ~P The total pressure loss is modeled as follows.
(9)
~P = ~Psc+~Pvn+~Pinc+~Pfric+~Pex
(10) • Scroll pressure loss ~Psc=func.(mach number at throat, static pressure) •Pressure loss in the VN ~Pvn= func.(mach number, vane setting angle, static pressure) · Rotor incidence loss ~Pinc=func.(incidence angle, relative mach number, static pressure) •Rotor friction loss L'l.Pfric= func.(relative mach number, diffusion factor, static pressure) •Discharging loss at rotor exit ~Pex= func.(relative mach number, velocity angle, static pressure) The turbine performance is calculated repeatedly as P6 become equal to the given value by inputting the pressure ratio, the rotational speed, and the dimensions. This turbine model also has the feature that the characteristics on the condition of negative efficiency can be predicted.
3.2 Transient Model of Turbocharged Engine The performance prediction program has been developed in order to predict primarily the characteristics of the turbocharging systems in short calculating time. So the steady state model like as the gas generator was made for the engine without considering the
345
cyclic phenomena. The turbocharger model is connected to the engine model by using the energy balance equation between the compressor and the turbine. The energy balance equation in the turbocharger is shown as follows. Lt+Lea=Lc+Lm+Ip·(w/2 7C)·U~wlD. t) (11) Lt: turbine output power, Lea: electric assist power, Lc: compressor input power, Lm: mechanical loss, w: angular velocity of the rotor, Ip: inertia momentum of the rotor, D. t: calculating time step. On the transient operating conditions, the factors of Lt, Lea, Lc, Lm, and Ware the function of time. Accordingly, the calculating tim e step D. t has to be very short for the good accuracy when the conventional cycle simulation is applied because the values of Lt, Lea, Lc, Lm, and w after the time step D. t are not able to be known. In our model, the steady state operating condition after the time step D. t is calculated firstly and then the transient operating 50 condition is calculated by use of the energy balance equation 0Q) with the mean values of Lt, Lea, ~ Lc, Lm, and w. The values of ,!J 40 Q) o .t:.t=0.1 Lt, Lea, Lc, Lm, and Ware '\il D .t:.t=0.2 calculated repeatedly by ~ 30 b. .t:. t=0.4 averaging the values at the time 0 u:: of t and t+ D. t. Therefore the ... X .t:.t=0.8 calculating accuracy of the < 20 transient characteristics does o 2 3 4 almost not depend on the Time (sec) calculating time step. The Fig.4 The effect of the calculating time step transient characteristics after the on the prediction accuracy time n D. t are calculated by integrating the equation (11) Turbine Inlet Pressure(experiment)' 300 - o -Turbine Inlet Pressure(calculation) I from t=O to n D. t. FigA shows 6. Boost Pressure(experiment) the result of the transient --Boost Pressure(calculation) performance prediction on the ~ 250 o & conditions of the various time "" OJ steps and indicates that the 5 200 calculating time step does not ""e change the calculation results. a. 150 Fig.S shows the predicted boost pressure and the turbine inlet pressure compared with 100 the experiment for the variable o 1000 2000 3000 4000 5000 nozzle turbocharged engine Engine Speed (rpm) operated on the full load Fig.5 The Predicted boost pressure and the back condition. The figure shows the pressure compared with the experiment good agreement of the prediction and the experiment. I
4. PERFORMANCE PREDICTION OF THE TURBOCHARGING SYSTEMS The performances of the variable geometry turbocharging systems have been investigated by the simulation on conditions of the steady state and transient operation.
346
The calculating time is about 20 minutes in the case of the transient simulation of the two stage turbocharging system on the recent personal computer. 4.1 Variable Geometry Turbocharging The performance prediction models of the compressors and the turbines were produced to evaluate the effects of the VIGV, the casing treatment, and the variable nozzle on the performances of the turbocharged engine. The gasoline engine with high power density was assumed. The maximum turbine inlet temperature is 113 OK at the maximum power operating point. Fig.6 shows the effects of the VIG V and the casing treatment on the boost pressure in the case of full load opcrating condition. The boost prcssure in the region of the low engine speeds is 300 significantly improved by use of the casing treatment and VIGV ~ 250 I----I---.f--+-.----..-+-because the compressor surge limits is drastically improved due to the synergy effect described ahead. Accordingly it is thought .... that the low-end torque and the ~ 150 t----~L-----.T-;"--+---I o transient response of the [lJ 100 ..._ - l_ _... turbocharged engine are improved by use of compressor 400 r--....,..--......-....,.~"-7., with the VIGV and the casing & c. treatment. On the other hand, the wide flow range turbine is necessary to achieve that high 250 I----I+-F~-t--...,.~-__I boost pressure over the wide ~ operating range. Fig.6 also shows :E 200 I---I-Ic---I-'\---+----I ., the turbine inlet pressures c 150 calculated to obtain the boost ::J f-- 100 ......- ........--""-----"--... of 'VIGV+Casing pressure 100 o 25 50 75 Treatment' shown. In the case of the conventional turbine with the Engine Speed (%) waste gate valve, the turbine inlet pressure increases as the engine Fig.6 The effect of VIGV on the boost pressure speed increases due to the increment of the wasted gas 300 flow. At high engine speed, ~ a. the turbine inlet pressure leads 25 250 to an unacceptable high value. ~ The high turbine inlet ::l en pressure causes the increase :il200 a. of the pumping loss and the ~ of knocking. Steady State condition occurrence g 150 -0- without VIGV Therefore the fuel co -6:- with VIGV Control consumption of the 100 turbocharged engine worsens. 0.0 1.0 2.0 3.0 4.0 5.0 On the other hand, in the case Time (sec) of the variable nozzle turbine, the turbine inlet pressure at Fig.7 The effect ofVIGV on the transient response
.,
~--.;._.:...._
:e
~
347
low engine speeds is higher than that of the conventional turbine. This is caused by the turbine efficiency reduction due to closing the variable nozzle. The pressure is lower than that of the conventional turbine at the middle and high engine speeds because the turbine efficiency is comparatively high. Additionally, the variable nozzle turbine is difficult to be used for the gasoline engine due to its complicated structure because the turbine inlet temperature is very high. As the results, it is thought that the variable geometry turbine having both the high efficiency and the wide flow range has to be developed for the turbocharged gasoline engine. r------------"I':::II""'I The effect of the VIGV control on the transient 2.5 response of the turbocharging Surge Line system was predicted by the 0 turbocharged engine simulation. :g Fig.7 shows the transient c:: characteristics of the boost ~::J 2 pressure. In the case of :il controlling the VIGV, the vane setting angle is maintained constant of 80 degree until - & - with VIGV control -=- without VIGV initial 0.6 sec, and is linearly 1.5 reduced to zero degree as the 30 40 50 20 time passes to 2 sec, and is maintained constant of 0 Corrected Air Flow Rate (%) degree after the time of 2sec. Fig.8 The operating line on the compressor map The transient response of the boost pressure is improved by 300 controlling the VIGV vane r0angle as mentioned above. The o. ""III 250 VIGV can change the rotational '::J speed of the turbocharger III III 200 maintaining the similar values III a.'both of the pressure ratio and +' III the efficiency of the 0 150 0 CD compressor. Accordingly, it is thought that the transient 100 response of the turbocharged E 200000 is improved by engine a. controlling the VIGV to cancel ~ "C 150000 the turbo-lag. The operating Q) Q) line on the compressor map is a. shown in Fig.8. The operating ~ 100000 c line shifts toward the lower o flow rate on the decelerating .~ 50000 c:: oLa_ _""condition of the turbocharged engine due to the turbo-lag. So, o the VIGV vane angle has to be 0.5 2.0 1.0 0.0 1.5 controlled to avoid the Time (sec) occurrence of the surge. Fig.9 The transient response of the two-stage turbocharging system
J:
~
348
4.2 Two Stage Turbocharging
The two-stage turbocharging system is also effective to improve the low-end torque and the transient response of the turbocharged engine. Then, the steady state and transient characteristics of the system were investigated by the engine simulation. The small turbocharger mainly charges air at the low engine speeds, on the contrary the large turbocharger mainly charges at the high engine speeds. So, the compressor surge does not restrict the charging pressure at low engine speeds. On the other hand, the two-stage turbocharging system is more complicated and larger size than the single turbocharging system. In many cases, it needs the bypass valve to avoid the choking of the small compressor at high engine speeds. Therefore, the engine-turbochargers matching technologies are especially important for the two-stage turbocharging system. Fig.9 shows the transient characteristics of the two-stage system compared with those of the single stage system. The inertia of the small turbocharger is 50% of the single stage turbocharger and the large turbocharger inertia is same as the single stage turbocharger inertia. The boost pressures of the two systems at each calculating time step are almost similar when each system is operated on the steady state condition. The small turbocharger of the two-stage system responds more quickly to the transient operation than the single stage turbocharger. Accordingly, it is thought that the two-stage turbocharging system improves the transient response ofthe turbocharged engine. 5. CONCLUSION
The wide flow range compressor with the Variable Inlet Guide Vane and the casing treatment has been developed for turbocharger. Then, the performance prediction program of the variable geometry turbocharging systems was produced to investigate the effect of the variable geometry turbocharger on the transient response. Then, the steady state and transient characteristics of the variable geometry turbo charging systems were evaluated by the simulation using the performance prediction program. The summary of the results is shown as follows. (1) The synergy effect of the VIGV and the casing treatment on the compressor surge limits was found, and the surge limits have been significantly improved due to the synergy effect. (2) The numerical analysis of the flow in the compressor with the VIGV and the casing treatment has shown that the re-circulating flow revealed near the shroud wall is restricted by the casing treatment and the VIGV. It is thought that the re-circulating flow is one of the primary factors that affect the surge limits. (3) The performance prediction models of the variable geometry turbocharging systems have been developed. The transient characteristics of the turbocharged engine can be calculated independently of the calculating time step. (4) The transient response of the turbocharged engine is improved by controlling the VIGV vane angle to cancel the turbo-lag.
349
6. REFERENCES 1) Robert J. McKee, Danny M. Deffenbaugh, 'Factors That Affect Surge Precursors in Centrifugal Compressors', GMRC Gas Machinery Conference, October 6-8, 2003. 2) S Yamaguti,et aI., 'The Development of Casing Treatment for Turbocharger Compressor', IMechE C602/0 1612002, 2002. 3) Tange, H., Ikeya, N., Takanashi, M., 'Variable Geometry of Turbocharger Compressor for Passenger Vehicles', SAE paper No.2003-01-005, 2003. 4) A.Whitfield, A.H Abdullah, 'The Performance of a Centrifugal Compressor with High Inlet Prewhirl', ASME Paper 97 -GT-182
Copyright© 2006 Hiroshi Uchida, TOYOTA Central R&D Labs.,Inc.
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PLAIN & FULL FLOATING BEARING SIMULATIONS WITH RIGID SHAFT DYNAMICS Dr Ian McLuckie, Mr Scott Barrett & Dr Boon Kai Teo Advanced Integrated Engineering Solutions Limited, P.D. Box 7784, Market Harborough, Leicestershire, UK T:+44 (0)1858 41485416 F:+44 (0)1858 414885 [email protected], & www.aiesl.co.uk ABSTRACT
Full floating bearings (FFB) are used in most automotive turbochargers, and due to emissions regulations there has been a push towards downsizing engines and applying turbo charging to generate optimized engine solutions for both gasoline and diesel applications. This paper shows a promising predictive bearing model that can be used to reduce the current turbocharger bearing system development times, which are mainly done by testing design parameters such as bearing geometry by a very time consuming experimentation process. The turbocharger rotor is regarded as being rigid, and the equations of motion are solved using the Bulirsch-Stoer integration scheme, and these are solved simultaneously with the bearing model. For this study the bearings are solved using a Rigid Hydro Dynamic (RHD) Finite Difference (SOR) scheme of Reynolds equation that includes both rotational and squeeze velocity terms. In order to keep the paper brief the results are shown only for the turbine bearing, as this bearing is the most highly loaded and operates at the highest temperatures. Two bearing types have been studied, the plain grooved (PGB) and the full floating bearing (FFB) for comparative purposes. The mathematical models used are generic and suitable for whole engine bearing studies, and the results in this paper show they seem very suitable in determining the onset of turbocharger bearing instability, and also how bearing instability may be suppressed. The current study has investigated forced response with the combined effects of gravity and unbalance. It is worth noting that the effects of both housing excitation and aerodynamic excitation [rom the compressor and turbine can be easily accommodated, and will be the subject of a future paper. INTRODUCTION
Due to the trend of engine down sizing and the general effects that global warming has on the development of engine systems, it is no small wonder that turbochargers have been seeing more demand in optimizing the power density of engines. At the same time they need to become more durable and operate over a wide range of vehicle and duty cycle applications. Therefore bearing design and development has to be painstakingly methodical in order to develop a turbocharger rotor system that is robust for a number of different vehicle applications.
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Figure 1 EngineDesigner typical V8 engine assembly Most if not all automotive turbocharger systems are developed on a gas stand, and sometimes in the transfer of the design from the test stand to a real application, this can result in problems that can be costly and not easily fixed. Our modelling approach will allow the future modelling of gas stand performance as well as applications in engines. We have extended our engine bearing models to turbocharger applications and the models used are also fully capable of incorporating aerodynamic forcing and forcing from the housing. The turbocharger is part of an Engine system package, and the Engine Design software used in the turbocharger bearing simulations is the culmination of many years of work carried out and recorded within references (l) to (8). McLuckie reference (1) is a comprehensive study of Gas bearing instability for high speed turbo rotors, and references (2) through to (8) form the basis of the engine bearing system design. These fundamental simulation techniques in conjunction with the work of reference (1) helped formulate the methods used for the turbocharger bearing instability simulations. ENGINE DESIGN ENVIRONMENT USED FOR TURBOCHARGER STUDIES The Engine Design environment used for the turbocharger bearings has been developed to enable rapid CAE design and analysis of complex systems. It is shown here in order to highlight an entirely novel approach to CAE analysis. The system environment is based on object-oriented principals and incorporates structural objects and fluid objects that interact via multi-physics. Figure 1 above gives the overview of a typical V8 engine configuration within the cngine design software. Engine bearing system models The engine bearing models offer a basic physical representation that is generic, so they can be easily applied to the Turbocharger bearing systems. Two rotor bearing systems were investigated in this paper. Firstly Plain Grooved Bearings (PGB) and secondly Full floating Bearings (FFB) For the turbocharger simulations we have 4 main bearing models that could be applied to the problem. SBA is a Short Bearing Approximation that is analytical and rapid but only accurate for solutions with low eccentricity ratios. Secondly SALBA is a Short And Long Bearing Approximation which is an analytical combination of SBA and LBA but is as accurate as RHD for plain bearings (no grooves or holes). Thirdly there is RHD which is a Rigid Hydro Dynamic numerical finite difference formulation. Lastly there is EHD, an Elasto HydroDynamic solution with coupled elasticity and hydrodynamics. For this paper the RHD solution was used as it can model grooving and
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pressure boundary conditions. Figure 2 below shows various bearings m our total engine friction system. Tribo logical objects
Figure 2 Typical Engine Tribological objects Engine system bearing mathematical models Two bearing types have been studied, the plain grooved (PBG) and the full floating bearing (FFB). The mathematical models used are generic and suitable for whole engine bearing studies. The bearing oil film forces are solved using a Rigid Hydro Dynamic (RHD) scheme and the values of pressure at the finite difference mesh points are solved by successive over relaxation. For incompressible flow and constant fluid properties, Reynolds Equation can be written.
where: p = pressure, h = film thickness, x = circumferential position, z = axial position, 11 = absolute viscosity, U = circumferential velocity and V = radial velocity. Note that U and V are not constant around the bearing surface The oil film forces are calculated in the RHD program. Currently, the housing is stationary, the journal and ring move. The equations of motion are in vector form and include the rotational acceleration terms for the floating ring. For the journal: For the ring:
= ~'HD(Xi,xJ + Fex,,,na/(t)- D.xj - K.Xj M,.x, =FRHD(Xr,X,) -FRHD(Xj,x,)-D.x, -K.x,
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The equations are solved by a Bulirsch-Stoer numerical integration method to obtain velocity and position at the next step from acceleration and velocity at the current step. TURBOCHARGER BEARING SIMULATION PARAMETERS This study has been compared with recent work by San Andres et al in references (9) and (10). However our typical rotor is assumed to be rigid, that is assumed to be operating below its 1st bending mode. A dynamic transient forced response of the turbocharger turbine end bearing has been studied for a typical rotor system and investigated by running up from standstill to 100,000 rpm. Speed steps below 20,000 rpm were 1000rpm and above 20,000 rpm they were 5,000rpm up to 50,000rpm and 10,000 rpm there onwards. The loads applied to the turbocharger rotor system were from gravity and unbalance loading. The finishing position of each time step is used as initial condition for the next speed calculation. Thc mass of the rotor was 0.485kg. The compressor bearing mass
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was O.291kg and the turbine bearing mass was 1.261 kg. Since the turbine bearing is the most severely loaded only this bearing was investigated in this paper. Unbalance loading was varied from O.023gmm, to 0.23gmm to a maximum of 2.3gmm. The lowest unbalance figure is bascd on BS (6861-1) grade G1.0. The other conditions of unbalance were G 10 and G 100. It is normal for rigid rotor turbochargers to be balanced to G2.5, so the study encompassed common turbocharger practices. The turbine end PGB bearing had a bore of 8 mm and a length of 3.5 mm with a circumferential groove of 1.0 mm. The FFB bearing had a bore of 8 mm with a length of 3.5 mm with a groove of 1.0 mm. The outside diameter is 14 mm and 6.5 mm long with a 1.0 mm wide circumferential groove. For the PGB bearings the diameteral clearance was varied from 16 microns to 8 microns to 4 microns. For the FFB bearings the inner bearing clearance was 2.5 microns and the outer diameteral clearance was 7 microns. Supply pressure to the bearings was varied from 4 bar, to 2 bar down to zero bar, and the oil for both bearings was assumed to be lOW 40 and the inlet temperature was 100 degrees centigrade. However this can be easily varied in the bearing models to accommodate any inlet conditions. The turbocharger configuration can be seen in figure 3a below. The bearing bush can be seen in figure 3b, and is 8 mm diameter by 3.5 mm long with 0.5 mm circumferential groove. PLAIN GROOVED BEARING (PGB) The loads acting on the plain grooved bearings consist of the superposition of gravity and unbalance loads. Figure 4 below shows typically the x and y force component diagrams for turbine bearing PGB2 at 20,000 rpm with an unbalance load of 2.3gmm.
Figure 3 a) A turbocharger rotor system with FFB bearings, b) Typical Full Floating bearing shell Loading at 20000 rpm
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Figure 5 (L-R) a) Turbine bearing PGB2 synchronous whirl at 3,000 rpm, b) PGB2 instability onset at 5,000 rpm, c) PGB2 sub-synchronous at 20,000 rpm, d) PGB2 sub-synchronous at 100,000 rpm (all with 0.023gmm unbalance, 4 micron Clearance & 4 bar supply) PGB orbits & sub-synchronous whirl onset The following sections describe the predicted results and comparisons to reference (9) and (10) from San Andres et al. Figure 5 shows a typical set of results for the turbocharger plain grooved bearing (PGB) mnning with a diameteral clearance of 4 microns and an oil inlet pressure of 4 bar and a low unbalance of 0.023gmm. Figure 5a shows synchronous whirl at a speed of 3000 rpm. However figure 5b shows the onset of instability (sub-synchronous whirl) at approximately 5,000 rpm. Thc orbit can be seen to be spiralling outwards and taking more than one revolution to complete the limit cycle loop, as the unbalance forces become greater than the gravity load, a common sign of instability onset. Figure 5c shows sub-synchronous whirl at 20,000 rpm, with the journal centre moving further towards the bearing centre as the unbalance forces increase with speed. Finally figure 5d shows sub-synchronous whirl (half frequency) orbit at 100,000 rpm where the eccentricity is taking up almost all the clearance space, and can be clearly seen centralised about the X and Y axes. This shows the condition where the unbalance forces are many orders of magnitude greater than the gravity load. PGB gravity & unbalance response The following figures show maximum bearing eccentricity ratio against speed, as the turbocharger mns up from zero rpm to its maximum speed of 100,000 rpm. It shows various curves which are the eccentricity ratio of the bearing plotted against speed. The gravity load remains constant as the unbalance load increases with speed.
Figure 6 Turbine bearing PGB2 (L - R), a) Eccentricity ratio versus speed with 4 16 micron clearance, 0.023 - 2.3gmm unbalance & 4 bar supply, b) Eccentricity ratio versus speed with 4 -16 micron clearance, 0.023gmm - 2.3gmm unbalance & o bar supply.
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Figure 6a shows the turbine bearing PGB2 response from gravity and unbalance loading. From the top the first 3 curves show results of the eccentricity ratio against speed with diameteral clearance at 16 microns and with oil supply at 4 bars. As speed increases the bearing eccentricity ratio increases until the load suddenly jumps, as the oil film cannot withstand the increasing unbalance load, and the journal touches the bearing shell. The next 3 curves show the results of the eccentricity ratio versus speed with diameteral clearance at 8 microns and supply at 4 bars. Here it can be seen that an unbalance of 2.3gmm (curve number 6) shows synchronous whirl right up to 100,000 rpm. The next 3 curves show rcsults of the eccentricity ratio versus speed with a diameteral clearance at 4 microns and oil supply at 4 bars. Again it can be seen that an unbalance of2.3gmm shows synchronous whirl right up to 100,000 rpm. Figure 6b shows the turbine bearing PGB2 response for gravity and unbalance loading. From the top the first 3 curves show results of eccentricity ratio versus speed with a diameteral clearance of 16 microns and zero bars supply pressure. For an unbalance of2.3gmm curve 3 shows the synchronous whirl up to 25,000 rpm where the approximate onset of sub-synchronous (half frequency) whirl takes place and continues up until 100,000 rpm. Thc next 3 curves show the results of eccentricity ratio versus speed with a diameteral clearance of 8 microns. For an unbalance of 2.3gmm curve 6 shows the synchronous whirl right up to 100,000 rpm. The last 3 curves show the results of the eccentricity ratio versus speed with a diameteral clearance at 4 microns. With an unbalance of2.3gmm the last curve shows synchronous whirl right up to 100,000 rpm. PGB bearing orbits and synchronous response Figure 7 (Left to Right) shows typical results for the turbine end bearing (PGB2) running with a diameteral clearance of 4 microns, oil inlet pressure of 4 bar and gross unbalance of 2.3 gmm. All the orbits are synchronous and stable, and shows that as the speed increases and hence the magnitude of the unbalance is many orders greater than the gravity load, that the journal orbit becomes virtually centralised in the bearing. Figure 7a is a very small synchronous whirl orbit at a speed of 3,000 rpm. Figure 7b is synchronous at 14,000 rpm with an increasingly larger orbit. Figure 7c shows a synchronous orbit at 18,000 rpm with an increasingly larger journal orbit. Figure 7d shows synchronous whirl at 100,000 rpm, where the eccentricity is taking up approximately 50 % of the clearance space.
Figure 7 Turbine bearing PGB2 eccentricity ratio curves (L-R) a) Synchronous whirl at 3,000 rpm, b) Synchronous whirl at 14,000 rpm, c) Synchronous whirl at 18,000 rpm, d) Synchronous whirl at 100,000 rpm (2.3gmm unbalance, 4 micron Clearance & 4 bar supply).
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Figure 8 Turbine end FFB2 (x and y) force components (1.26kg gravity & 2.3gmm unbalance @ 20,000 rpm) FULL FLOATING BEARING (FFB) OBJECTS Full floating bearings have two oil films, and the floating ring rotates normally in the region of 0.4 to 0.2 times the rotor speed. It has been observed by San Andres et al reference (9), that at low speed the ring ratio is about 0.4 and at high speed as the oil temperature increases the ring ratio reduces to about 0.2. We have taken the ring ratio to be fixed at 0.25 times the rotor speed, giving a 25,000 rpm ring speed at a maximum rotor speed of 100,000 rpm. A future paper concentrating on the FFB will consider variable ring ratio calculations. A typical load spectrum applied to the turbine FFB2 can be seen in figure 8 with the x and y forces at 20,000 rpm and with a gross unbalance of 2.3gmm. Inner oil mm sub-synchronous whirl onset Figure 9 shows a set of typical results for the turbocharger turbine end FFB2 inner film response running with an inner diameteral clearance of 2.5 microns and an outer diameteral clearance of 7 microns and an oil supply pressure of 4 bar and an unbalance of 0.23gmm. Figure 9a shows an extremely small synchronous whirl orbit at a speed of 1000 rpm. Figure 9b shows onset of sub-synchronous whirl at 3,000 rpm, the orbit can be seen to be spiralling outwards taking more than one revolution to complete the loop with a very small orbit, a common sign of instability onset. Figure 9c shows a subsynchronous orbit at 30,000 rpm with an increasingly larger orbit. Figure 9d shows subsynchronous whirl at 100,000 rpm, withjoumal orbit taking up approximately 30% plus of the clearance space, as the unbalance load increase in line with speed becoming many orders greater than the gravity load. 1.Z3g., .. _"nb.I+Z.III7 .. I"4b'f_81_3~G\IIl 00'5
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Figure 10 Turbine bearing FFB2 (L-R) a) Synchronous whirl at 3,000 rpm, b) Synchronous whirl at 10,000 rpm, c) Synchronous whirl at 12,000 rpm, d) Synchronous whirl at 100,000 rpm (2.3gmm, 2.5 micron innerl7 micron outer clearance, 4 bar) Inner oil film synchronous whirl Figure 10 shows a typical set of results for the turbine bearing FFB2 imler film response running with an inner diameteral clearance of 2.5 microns and a supply pressure of 4 bars, with a gross unbalance of 2.3gmm. The inner journal response can be seen to be synchronous right up to 100,000 rpm, which increases with speed as unbalance becomes many orders greater than the gravity loading. Figure lOa shows a very small synchronous orbit at a speed of 3000 rpm. Figure lOb shows synchronous whirl at 10,000 rpm. Figure IOc shows a synchronous orbit at 12,000 rpm with an increasingly larger eccentricity ratio. Figure IOd shows synchronous whirl at 100,000 rpm, where the eccentricity ratio takes up approximately 30 % of the clearance space.
Outer oil film sub-synchronous whirl onset Figure 11 shows a set of results for the turbine bearing FFB2 outer film response (ring) running with an outer diameteral clearance of 7 microns and a supply pressure of 4 bars, with an unbalance of 0.23gmm. It can be seen that as the ring rotates at about 0.25 times the shaft speed, that a 1/8 cyclc orbit of the shaft rotatcs at 112 frequcncy relative to the ring. This orbit gets increasing larger as the unbalance force becomes many orders greater than gravity loads with speed. Figure II a shows a 1/8 sub-synchronous whirl orbit (half frequency) at a speed of 6,000 rpm. Figure II b shows 1/8 sub-synchronous whirl at 14,000 rpm, with an orbit closely centred about the x axis and offset from the y. Figure Ilc shows a sub-synchronous orbit at 30,000 rpm with an orbit moving closer to the y axis. Figure lid shows sub-synchronous whirl at 100,000 rpm, where the eccentricity is taking up approximately 25 % of the clearance space. "".mm."" .. "~.'''m'...''_''_.."" I
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Figure 12 Turbine bearing FFB2 a) Ring sub-synchronous whirl at 6,000 rpm, b) Ring sub-synchronous whirl at 14,000 rpm, c) Ring synchronous whirl onset at 40,000 rpm, d) Ring synchronous whirl at 100,000 rpm (2.3gmm, 2.5 micron inner/7 micron outer clearance, 4 bars) Outer oil film synchronous whirl Figure 12 shows a set of results for the turbocharger full floating bearing FFB2 outer film response running with an outer diameteral clearance of 7 microns and a supply pressure of 4 bars, with a gross unbalance of 2.3gmm. Thc orbit responses show that as the ring rotates approximately at 0.25 times the shaft speed with an 118 cycle orbit of the shaft rotating at half frequency relative to the ring. This response also becomes synchronous at 40,000 rpm. Figure 12 a shows a sub-synchronous whirl orbit (half frequency) at a speed of 6,000 rpm. Figure 12b shows sub-synchronous whirl at 14,000 rpm, with an orbit closely centred about the x axis and offset from the y. Figure 12c shows a smaller sub-synchronous orbit at 40,000 rpm. Figure 12d shows synchronous whirl at 100,000 rpm, where the eccentricity is taking up approximately 15 % of the clearance space centred about the x and y axes.
FFB gravity & unbalance response The turbocharger FFB bearings were simulated in exactly the same way as the PGB bearings, the excitation input being by way of gravity and unbalance loading, which cause the bearing oil film to respond by generating a journal orbit as has been described earlier in the paper. The following figures show curves of the maximum bearing eccentricity ratio versus speed as the turbocharger runs up to a maximum speed of 100,000 rpm. Figure 13a shows the turbine bearing FFB2 inner oil film response for gravity loading plus unbalance, note that the speed along the x axis refers to the shaft rotation speed and hence the speed of unbalance excitation. The curves show the results of the eccentricity ratio versus speed with an inner diameteral clearance of 2.5 microns and an outer diameteral clearance of 7 microns and a supply pressure of 4 bars. With an unbalance of 2.3gmm curve 3 shows stabilised synchronous whirl up to100,000 rpm, where by it can be seen the benefits of gross unbalance in stabilising rotor behaviour. Figure 13b shows the turbine bearing FFB2 outer oil film response for gravity loading plus unbalance. From the top, the curves show the results of the eccentricity ratio versus speed with an inner diameteral clearance of 2.5 microns and outer diameteral clearance of 7 microns and a supply pressure of 4 bars. For an unbalance of 2.3gmm curve 3 shows sub-synchronous whirl up to 40,000 to 45,000 rpm where upon synchronous whirl takes over and this stabilised synchronous whirl continues right up to 100,000 rpm.
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BEARINGS, PGB2 AND FFB2, WATERFALL FREQUENCY SPECTRA Figure 14 above shows waterfall frequency spectra for the turbine bearing PGB2. Figure 14a has bearing details of 0.023 gmm unbalance, 4 microns bearing clearance and 4 bar supply pressure and shows almost entirely sub-synchronous response at approximately Yo shaft speed. It can be seen that the sub-synchronous ratio varies from O.S at low speed at 3,000 rpm to approximately 0.46 at 100,000 rpm. Figure 14b has bearing details of 2.3 gmm gross unbalance, 4 microns clearance and a 4 bar supply pressure and shows that the response is entirely synchronous over its entire speed range up to 100,000 rpm. Figures IS show the waterfall frequency spectra for the turbine bearing FFB2 with bearing details of 2.S/7 microns inner/outer clearance and 4 bar supply pressure. Figure ISa shows the results of an unbalance response of 0.23gmm. The trends of these results agree very well with those of reference San Andres et al (9) and (10). However unfortunately, the actual bearing and rotor details were not reported in these papers. Figure ISa shows an 118 order sub-synchronous response of the journal which is a 112 order of the ring at the extreme LHS. It also shows a S/8 order of the journal which is about 112 order ofthe shaft + ring speed. Figure ISa also shows a synchronous response of Ix shaft order. Figure ISb shows the same bcaring dctails of 2.S/7 microns inner/outer clearance and 4 bar supply pressure, but subjected to a gross unbalance of 2.3 gmm. One can see that there is a 118 up to about 40,000 to 4S,000 rpm where upon the response of the ring becomes synchronous right to 100,000 rpm. The response of the shaft has S/8 order also up to about 40,000 to 4S,000 rpm where upon it becomes also entirely synchronous up to 100,000 rpm. This is a fascinating phenomenon that needs more understanding, but appears to show the stabilising effect of gross unbalance forcing. The main issue now becomes how to load the bearing journals with gross unbalance with out causing large bending moments in the rotor shaft. This may well be reported in anothcr future paper.
TURBOCHARGER FRICTIONAL LOSSES The frictional losses of the turbocharger are made up of the compressor and the turbine end bearings, but here we are showing only the turbine bearing losses, and comparing both PGB and FFB losses for the case of 4 bar supply pressure, with unbalance varying from 0.023gmm to 2.3gmm, with the lowest clearance condition for both bearing types. We have carried out many more calculations but have no room to show them in this paper.
Figure 16 Turbine bearing PGB2 and FFB2 bearing friction torque versus speed (varying unbalance, supply pressure and clearance)
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Figure 16 shows that the highest friction is generated by the FFB bearing system, with the PGB bearing system with 4 micron clearance, 4 bar supply pressure and 2.3grnm unbalance producing the lowest frictional losses.
CONCLUSIONS This paper has shown some interesting trends with regard turbocharger bearing designs for the turbine bearing. This is the most heavily loaded and runs at relatively high temperatures. Both plain grooved bearings (PGB) and full floating bearings (FFB) have been investigated. The bearing models used, can determine any bearing design suitability with respect to forced response, whether the loads are derived from the combustion events, and inertial dynamic loading, or in this instance gravity and unbalance response. The turbine end PGB has been studied in order to prove the modelling approach prior to applying it to any FFB systems. The generic bearing model is an RHD Reynolds solution developed to simulate forced excitation and loading derived from combustion events and engine dynamics, and it can be seen that it is very effective in simulating the PGB. For the PGB to be effective for high speed turbo shafts, very small operating clearances are required. This is generally understood in thc industry. Also it can be seen that with a zero or low unbalance that half frequency whirl will be manifest even with a very small 4 micron bearing clearance. For bearing clearances of 8 microns or less, gross unbalance (2.3gmm) stabilises the rotor orbits extremely well. In fact low clearance and gross unbalance gives the best overall rotor bearing response as the whirl remains synchronous up to 100,000 rpm. However how to apply this only to the bearing journals is another story and this will prove to be a very interesting next paper. Additionally the turbine end FFB has been studied and shows interesting characteristics. The inner shaft bearing with a 0.23gmm unbalance and a 4 bar circumferential groove supply pressure shows it has a sub-synchronous orbit from as low as 3,000 rpm, and this is due to the centralizing effects of the 4 bar pressure from the circumferential oil groove. The sub-synchronous frequency is at 5/8 shaft order or 1/2 order of the shaft + ring speed. However with gross unbalance, 2.3gmm (GlOO balance levels), the inner shaft bearing has a 5/8 shaft order sub-synchronous response up to about 40,000 - 45,000 rpm where upon in then behaves in a synchronous manner right up to 100,000 rpm. The outer ring bearing on the other hand with 0.23gmm unbalance and a 4 bar pressure shows that it is unstable almost from ,'tart up and whirls at 1/8 order shaft frequency or 1/2 order ring frequency right up to 100,000 rpm. A very interesting phenomenon shown by the outer ring bearing with gross unbalance is that it starts with a with an 118 order shaft or 11 2 order ring and at about 40,000 to 45,000 rpm the response suddenly becomes synchronous and continues this behaviour right up to 100,000 rpm. These results tend to agree with the observations of San Andres et al (9) and (l0) and others. Their results of a turbocharger system show that bearings can be seen to operate sub-synchronously and suddenly at a particular speed this vanishes or switches to operate in a synchronous stable manner. The control of bearing instability and subs-synchronous response by gross unbalance has been long understood by the industry. But it has been shown that this cannot be applied to all bearing designs as a cure all. It is shown in this paper that gross unbalance (GIOO), for the PGB with clearances of 8 microns or less gives good results. For the FFB with clearances of 2.5/7 microns gross unbalance gives a stable orbit of the inner bearing and increases the range at which the outer bearing operates with synchronous whirl, we believe that with a different clearance ratio that range of stability of the outer bearing can be extended significantly. Future work will include the effects of a non-
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fixed speed ratio (currently under development) for the FFB and maybe the inclusion of housing and blade passing aerodynamic excitations.
ACKNOWLEDGMENTS We would like to thank all the team at Advanced Integrated Engineering Solutions Ltd for the efforts put into the EngineDesigner, and TurboDesigner bearing models.
REFERENCES I. I R W McLUCKIE - "Instability studies of an '0' - ring flexibly supported gas bearing mounted cool air unit". PhD Thesis 1990 Cranfield Institute of Technology. 2. D J S BARRETT, I R W McLUCKIE and M-T MA - "A valve train design analysis comparative study of three current diesel engines.", Der Virtuelle Motor conference. Munich. Germany. Haus Der Technik. October 2000. 3. D.J.S, BARRETT, A. El-ZAFRANY & I.R.W McLUCIUE - "Elastohydrodynamic analysis of bearings by boundary element and finite difference methods", Boundary Elements XXIV, Sintra, Portugal, 2002. 4. M-T. MA & IRW McLUCKIE - "An EHD study of a comlecting rod big end bearing including elasticity and inertia effects of the bearing structure", World Tribology Conference, Vienna, Austria, 2001. 5. IAN McLUCKIE - "From the engine to the vehicle and beyond", NAFEMS world congress, Orlando, Florida, USA, May 27 - 31 2003. 6. IAN McLUCKIE & SCOTT BARETT - "Objective Faster real world solutions". NAFEMS World Congress, Malta, 2005. 7. IAN McLUCKIE & SCOTT BARRETT - "V8 Engine Bearing Dynamics High performance with minimum friction" Proceedings of the ASME IDETC/CIE conference, Long Beach California. USA, September 24-28, 2005 8. IAN McLUCKIE - "Faster Real World CAE Solutions - with OOD Integrated Knowledge Based Systems". TCN CAE 2005, Leece, Italy, October 5-8, 2005. 9. L. SAN ANDRES &.J. KERTH "Thermal effects on the perfonnance of floating ring bearings for turbochargers" Proceedings of I MECHE Vol.218 Part 1. 2004 10.1. NARANJO, C. HOLT and L. SAN ANDRES. "Dynamic response of a rotor supported in a floating ring bearing" r' International Conference on Rotordynamics of MachinelY, ISCRMAl, paper 2005, Lake Tahoe, Nevada. USA, 2001.
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