CRYOCOOLERS 10
A publication of the International Cryocooler Conference
CRYOCOOLERS 10
Edited by
R. G. Ross, Jr. Jet Propulsion Laboratory California Institute of Technology Pasadena, California •
KLUWER ACADEMIC PUBLISHERS NEW YORK, BOSTON, DORDRECHT, LONDON, MOSCOW
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Preface The last two years have witnessed a continuation in the breakthrough shift toward pulse tube cryocoolers for long-life, high-reliability cryocooler applications with the development of mature products addressed to a wide variety of operating temperatures. On the commercial front, GiffordMcMahon cryocoolers with rare earth regenerators continue to make great progress in opening
up the 4 K market. Also in the commercial sector, continued interest is being shown in the development of long-life, low-cost cryocoolers for the emerging high temperature superconductor electronics market, particularly the cellular telephone base-station market. At higher temperature levels, closed-cycle J-T or throttle-cycle refrigerators are taking advantage of mixed refrigerant gases, spearheaded in the former USSR, to achieve low-cost cryocooler systems in the 65 - 80 K temperature range. Tactical Stirling cryocoolers, the mainstay of the defense industry, continue to find application in cost-constrained commercial applications and space missions, but continue to shrink in numbers as the defense industry continues its consolidation. To archive the latest developments and performance of this expanding stable of cryocoolers, this book draws upon the work of many of the international experts in the field of cryocoolers. In particular, Cryocoolers 10 is based on their contributions at the 10th International Cryocooler Conference, held in Monterey, California, in May 1998. The program of this conference consisted of 128 papers; of these, 101 are published here. Although this is the tenth meeting of the
conference, which has met every two years since 1980, the authors’ works have only been made available to the public in hardcover book form since 1994. This book is thus the third volume in this new series of hardcover texts for users and developers of cryocoolers. As a significant addition to this proceedings, Cryocoolers 10 contains ten articles highlighting cryocooler developments that have taken place in the former USSR over the past 20 years. Eight of these cover key accomplishments of the Special Research and Development Bureau (SR&DB) in Cryogenic Technology of the Institute for Low Temperature Physics and Engineering of the National Academy of Sciences in the Ukraine; they are listed in the subject index under: SR&DB of the Ukraine. Also, two articles authored by staff of the Kharkov State Polytechnic University in the Ukraine are included; they cover more recent research activities on pulse tube type coolers and provide insight into the teaching of cryocooler design in the Ukraine. The ten Ukrainian articles reflect a significant increase in collaboration between the cryocooler research centers in the former USSR and the broader worldwide cryocooler community. Because this book is designed to be an archival reference for users of cryocoolers as much as for developers of cryocoolers, extra effort has been made to provide a thorough Subject Index that covers the referenced cryocoolers by type and manufacturer’s name, as well as by the scientific or engineering subject matter. Extensive referencing of test and measurement data, and application and integration experience, is included under specific index entries. Contributing organizations are also listed in the Subject Index to assist in finding the work of a known institution, laboratory, or manufacturer. To aide those attempting to locate a particular contributor’s work, a separate Author Index is provided, listing all authors and coauthors. Prior to 1994, proceedings of the International Cryocooler Conference were published as informal reports by the particular government organization sponsoring the conference — typically a different organization for each conference. A listing of previous conference proceedings
v
vi
PREFACE
is presented in the Proceedings Index, at the rear of this book. Most of the previous proceedings were printed in limited quantity and are out of print at this time. The content of Cryocoolers 10 is organized into 15 chapters, starting first with an introductory chapter providing cooler overviews and summaries of major government cryocooler development programs. The next few chapters address cryocooler technologies organized by type of cooler, starting with Stirling cryocoolers, pulse tube cryocoolers, and associated research. Next, Brayton, Joule-Thomson, hybrid J-Ts, and sorption cryocoolers are covered in a progression of lowering temperatures. Gifford-McMahon cryocoolers and low-temperature regenerators in the 4 to 10 K range are covered next, followed by a glimpse into the future with miniature solid-state refrigerators and advanced refrigeration cycles. The last three chapters deal with cryocooler integration technologies and experience to date in a number of representative applications. The articles in these last three chapters contain a wealth of information for the potential user of cryocoolers, as well as for the developer. It is hoped that this book will serve as a valuable source of reference to all those faced with the challenges of taking advantage of the enabling physics of cryogenics temperatures. The expanding availability of low-cost, reliable cryocoolers is making major advances in a number of fields.
Ronald G. Ross, Jr. Jet Propulsion Laboratory California Institute of Technology
Acknowledgments The International Cryocooler Conference Board wishes to thank Lockheed Martin Advanced Technology Center, which hosted the 10th ICC, and to express its deepest appreciation to the Conference Organizing Committee, whose members dedicated many hours to organizing and managing the conduct of the Conference. Members of the Organizing Committee and Board for the 10th ICC include:
CONFERENCE CO-CHAIRS
Ted Nast, Lockheed Martin ATC
peter Kittel, NASA/ARC CONFERENCE ADMINISTRATOR
Aurie Pedronan, Lockheed Martin ATC
PROGRAM COMMITTEE
John Brisson, MIT William Burt, TRW David Glaister, Aerospace Corp. Geoffrey Green, NSWC
Tom Kawecki, NRL Lawrence Wade, JPL
PROGRAM CHAIRMAN
Peter Kerney, conductus CONFERENCE SECRETARY Jill Bruning, Nichols Research Corp.
PUBLICATIONS Ron Ross, Jet Propulsion Lab TREASURER
Ray Radebaugh, NIST
ADVISORY BOARD
Stephen castles, NASA/GSFC Chris Jewell, ESA Ralph Longsworth, APD Cryogenics Yoichi Matsubara, Nihon Univ., Japan Martin Nisenoff, NRL
Marko Stoyanof, AFRL Walter Swift, Creare Inc. Klaus Timmerhaus, U. of Colorado Jia Hua Xiao, NIST
In addition to the Committee and Board, key staff personnel made invaluable contributions to the preparations and conduct of the conference. Special recognition is due C. Stoyanof, C. Seeley, J. M. Lee, C. Nast, and C. Kerney.
vii
Contents Government Cryocooler Development and Test Programs
1
An Overview of the Performance and Maturity of Long Life Cryocoolers for Space Applications . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1 D. Glaister, The Aerospace Corp., Albuquerque, NM; M. Donabedian and D. Curran, The Aerospace Corp., El Segundo, CA; and T. Davis, AFRL, Kirtland AFB, NM
Air Force Research Laboratory Cryocooler Technology Development........ 21 T.M. Davis, J. Reilly, and Lt. B.J. Tomlinson, AFRL, Kirtland AFB, NM
Endurance Evaluation of Long-Life Space Cryocoolers at AFRL – an Update . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 33 Lt. B.J. Tomlinson, AFRL, Kirtland AFB, NM; and A. Gilbert and J. Bruning, NRC, Albuquerque, NM
DARPA Low Cost Cryocooler Performance Testing: Preliminary Results . . . . 43 T.G. Kawecki, NRL, Washington, DC; and S.C. James, AlliedSignal Tech. Services Corp., Camp Springs, MD
Development of Cryogenic Cooling Systems at the SR&DB in the Ukraine .. 55 S.I. Bondarenko and V.F. Getmanets, SR&DB, Kharkov, Ukraine
Stirling Cryocooler Developments
59
Qualification Test Results for a Dual-Temperature Stirling Cryocooler ...... 59 W.J. Gully, H. Carrington and W. Kiehl, Ball Aerospace & Tech. Corp., Boulder, CO; and T. Davis and B.J. Tomlinson, AFRL, Kirtland AFB, NM
Progress towards the Development of a 10K Closed Cycle Cooler for Space Use . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 67 A.H. Orlowska and T.W. Bradshaw, RAL, Didcot, UK; S. Scull, MMS, Bristol, UK; and Lt. B.J. Tomlinson, AFRL, Kirtland AFB, NM
Development of a Light Weight Linear Drive Cryocooler for Cryogenically Cooled Solid State Laser Systems . . . . . . . . . . . . . . . . . . . . . . . . . 77 L.B.Penswick, STC, Kennewick, WA; and B.P. Hoden, Decade Optical Systems, Inc., Albuquerque, NM
Low-Weight and Long-Life 65K Cooler . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 87 V.T. Arkhipov, V.N. Lubchenko, and L.V. Povstyany, SR&DB. Kharkov, Ukraine; and H. Stears, Orbita Ltd., Kensington, MD
ix
x
CONTENTS
Thermal Performance of the Texas Instruments 1-W Linear Drive
Cryocooler
. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 95
D.L. Johnson, JPL, Pasadena, CA
Qualification of the BEI B512 Cooler, Part 1 – Environmental Tests........ 105 D.T. Kuo, A.S. Loc, and S.W.K. Yuan, BEI Tech., Sylmar, CA
Use of Variable Reluctance Linear Motor for a Low Cost Stirling Cycle Cryocooler . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 111 M. Hanes, D. Chase, and A. O’Baid, STI, Santa Barbara, CA
Pulse Tube Cryocooler Developments
119
AIRS PFM Pulse Tube Cooler System-Level Performance . . . . . . . . . . . . . . . . . . . . . 119 R.G. Ross, Jr., D.L. Johnson, and S.A. Collins, JPL, Pasadena, CA; and K. Green and H. Wickman, LMIRIS, Lexington, MA
Multispectral Thermal Imager (MTI) Space Cryocooler Development, Integration, and Test . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 129 Lt. B.J. Tomlinson, AFRL, Kirtland AFB, NM; W. Burt, TRW, Redondo Beach, CA; D. Davidson and C. Lanes, Sandia Nat’l Lab, Albuquerque, NM; and A. Gilbert, NRC, Albuquerque, NM
IMAS Pulse Tube Cooler Development and Testing . . . . . . . . . . . . . . . . . . . . . . . 139 C.K. Chan, T. Nguyen, R. Colbert, and J. Raab, TRW, Redondo Beach, CA; and R.G. Ross, Jr. and D.L. Johnson, JPL, Pasadena, CA
Development of a 1 to 5 W at 80 K Stirling Pulse Tube Cryocooler . . . . . . . . . 149 Y. Hiratsuka and Y.M. Kang, Daikin Indus., Tsukuba, Japan; and Y. Matsubara,
Nihon Univ., Funabashi, Japan
Development of a 2 W at 60 K Pulse Tube Cryocooler for Spaceborne Operation . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 157 V. Kotsubo, J.R. Olson, and T.C. Nast, Lockheed Martin ATC, Palo Alto, CA
Performance of a Two-Stage Pulse Tube Cryocooler for Space Applications . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 163 J.R. Olson, V. Kotsubo, P.J. Champagne, and T.C. Nast, Lockheed Martin ATC, Palo Alto, CA
Development of Pulse Tube Cryocoolers for HTS Satellite Communications . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 171 V. Kotsubo, J.R. Olson, P. Champagne, B. Williams, B. Clappier, and T.C. Nast, Lockheed Martin ATC, Palo Alto, CA
A Pulse Tube Cryocooler for Telecommunications Applications ................ 181 J.L. Martin and C.M. Martin, Mesoscopic Devices, Golden, CO; and J. Corey, CFIC, Troy, NY
Design and Preliminary Testing of BEI’s CryoPulse 1000,
the Commercial One Watt Pulse Tube Cooler . . . . . . . . . . . . . . . . . . . . . . . . . . . . 191 S. W.K. Yuan, D. T. Kuo, and A.S. Loc, BEI Technologies, Slymar, CA
CONTENTS
xi
Pulse Tube Cryocooler Configuration Investigations
197
Optimal Design of a Compact Coaxial Miniature Pulse Tube Cooler . . . . . . . . 197 Y.L. Ju, Y. Zhou, J.T. Liang, and W.X. Zhu, Cryogenics Lab., Chinese Acad. of Sci., Beijing, China
Performances of Two Types of Miniature Multi-Bypass Coaxial Pulse Tube Refrigerators . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 205 J.T. Liang, J.H. Yang, W.X. Zhu, Y. Zhou, and Y.L. Ju , Cryogenics Lab, Chinese Acad. of Sci., Beijing, China
Development of a 5 to 20 W at 80 K GM Pulse Tube Cryocooler ............. 213 S. Fujimoto and Y.M. Kang, MEC Lab, Daikin Indus., Tsukuba, Japan; and Y. Matsubara,
Nihon Univ., Funabashi, Japan
Conceptual Design of Space Qualified 4 K Pulse Tube Cryocooler . . . . . . . . . . 221 G.R. Chandratilleke, Y. Ohtani, H. Nakagome, and K. Mimura, Toshiba Corp., Kawasaki,
Japan; N. Yoshimura and Y. Matsubara, Nihon Univ., Funabashi, Japan; H. Okuda, ISAS, Sagamihara, Japan; and T. Iida and S. Shinohara, NASDA, Tsukuba, Japan
Performance Dependence of 4K Pulse Tube Cryocooler on Working Pressure . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 227 N. Yoshimura, S.L. Zhou and Y. Matsubara, Nihon Univ., Funabashi, Japan; G.R. Chandratilleke, Y. Ohtani, and H. Nakagome, Toshiba R&D Center, Kawasaki, Japan; H. Okuda, ISAS, Sagamihara, Japan; and S. Shinohara, NASDA, Tsukuba, Japan
Research of Two-Stage Co-Axial Pulse Tube Coolers Driven by a Valveless Compressor . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 233 L.W. Yang, J.T. Liang, Y. Zhou, and J.J. Wang, Cryogenics Lab, Chinese Acad. of Sci., Beijing, China
Experimental Investigation of a Unique Pulse Tube Expander Design . . . . . . . 239 C.S. Kirkconnell, Raytheon Systems Co., El Segundo, CA
An Experimental Study on the Heat Transfer Characteristics of the Heat Exchangers in the Basic Pulse Tube Refrigerator . . . . . . . . . . . . . . . . . . 249 S. Jeong and K. Nam, Korea Adv. Institute of Sci. and Tech., Taejon, Korea
Double Vortex Tube as Heat Exchanger and Flow Impedance for a Pulse Tube Refrigerator . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 257 M.P. Mitchell, Mitchell/Stirling, Berkeley, CA; D. Fabris, Illinois Inst. of Tech., Chicago, IL; and B.J. Tomlinson, AFRL, Kirtland AFB, NM
Investigations on Regenerative Heat Exchangers . . . . . . . . . . . . . . . . . . . . . . . . . . 265 I. Rühlich and H. Quack, Univ. of Dresden, Dresden, Germany
Pressure Drop in Pulse Tube Cooler Components . . . . . . . . . . . . . . . . . . . . . . . . . 275 H.E. Chen, J.M. Bennett, S. Yoshida, A. Le, and T.H.K. Frederking, UCLA, Los Angeles, CA
Pulse Tube Flow and Operational Stability Investigations
281
Experimental Results on Inertance and Permanent Flow in
Pulse Tube Coolers
. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 281
L. Duband, I. Charles, A. Ravex, and L. Miquet, CEA/DRFMC, Grenoble, France; and C.I. Jewell, ESA-ESTEC, Noordwijk, The Netherlands
xii
CONTENTS
Experimental Results of Pulse Tube Cooler with Inertance Tube as Phase Shifter . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 291 K. V. Ravikumar, Atlas Scientific, NASA ARC, Moffett Field, CA; and Y. Matsubara, Nihon
Univ., Funabashi, Japan
Observation of DC Flows in a Double Inlet Pulse Tube . . . . . . . . . . . . . . . . . . . 299 V . Kotsubo, P. Huang, and T.C. Nast, Lockheed Martin ATC, Palo Alto, CA
Suppression of Acoustic Streaming in Tapered Pulse Tubes . . . . . . . . . . . . . . . . 307 J.R. Olson, Lockheed Martin ATC, Palo Alto, CA; and G. W. Swift, Los Alamos Nat’1 Lab., Los Alamos, NM
Performance of a Tapered Pulse Tube . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 315 G.W. Swift, M.S. Allen, and J.J. Wollan, Cryenco Inc., Denver, CO
Numerical Study of Pulse Tube Flow . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 321 Y. Hozumi, Chiyoda Corp., Yokohama, Japan; M. Murakami, Univ. of Tsukuba, Tsukuba, Japan; and T. Iida, NASDA, Tsukuba, Japan
Visualization Study of the Local Flow Field in an Orifice and Double-Inlet Pulse Tube Refrigerator
. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 329
M. Shiraishi and A. Nakano, Mech. Engin. Lab, Tsukuba, Japan; and N. Nakamura, K. Takamatsu, and M. Murakami, Univ. of Tsukuba, Tsukuba, Japan
Stability Study of Coaxial Pulse Tube Cooler Driven by Air Conditioning Compressor . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 337 L.W. Yang, J.T. Liang, Y. Zhou, P.S. Zhang, W.X. Zhu, and J.H. Cai, Cryogenics Lab, Chinese Acad. of Sci., Beijing, China
Gas Contamination Effects on Pulse Tube Performance . . . . . . . . . . . . . . . . . . . 343 J.L. Hall and R.G. Ross, Jr., JPL, Pasadena, CA
Pulse Tube Modeling and Diagnostic Measurements
351
Simple Two-Dimensional Corrections for One-Dimensional Pulse Tube Models . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 351 J.M. Lee and P. Kittel, NASA ARC, Moffett Field, CA; K.D. Timmerhaus, Univ. of Colorado, Boulder, CO; and R. Radebaugh, NIST, Boulder, CO
Pulse Tube Development Using Harmonic Simulations . . . . . . . . . . . . . . . . . . . 359 H.W.G. Hooijkaas, Eindhoven Univ. of Tech.; and A.A.J. Benschop, Signaal-USFA, Eindhoven, The Netherlands
Analysis of a Two Stage Pulse Tube Cooler by Modeling with Thermoacoustic Theory . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 369 A. Hofmann, Forschungszentrum Karlsruhe, Karlsruhe, Germany; and S. Wild, Univ. of Karlsruhe, Karlsruhe, Germany
Modeling Pulse Tube Coolers with the MS*2 Stirling Cycle Code . . . . . . . . . . 379 M.P. Mitchell, Mitchell/Stirling, Berkeley CA; and L. Bauwens, Univ. of Calgary, Calgary
Canada
Experimental Verification of a Thermodynamic Model for a Pulse Tube Cryocooler . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 387 J. Yuan and J.M. Pfotenhauer, Univ. of Wisconsin, Madison, WI
CONTENTS
xiii
Measurements of Gas Temperature in a Pulse Tube Using the Planar Laser Raleigh Scattering Method . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 395 K. Nara, Y. Hagiwara, and S. Ito, Adv. Mobile Telecommunication Tech. Inc., Aichi-ken, Japan
Mathematical Model of a Wave Cooler . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 405 V.N. Kukharenko, Kharkov State Polytechnic Univ., Kharkov, Ukraine
Pulse Tube Modeling as a Means of Teaching the Design of Cryogenic Refrigerators . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 413 V.N. Kukharenko, Kharkov State Polytechnic Univ., Kharkov, Ukraine
Brayton Cryocooler Developments
421
Design and Test of Low Capacity Reverse Brayton Cryocooler for
Refrigeration at 35 K and 60 K . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 421 J. McCormick, G. Nellis, W. Swift, and H. Sixsmith, Creare Inc., Hanover, NH; and J. Reilly, AFRL, Kirtland AFB, NM
Reverse Brayton Cryocooler for NICMOS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 431 G. Nellis, F. Dolan, J. McCormick, W. Swift, and H. Sixsmith, Creare Inc., Hanover, NH; and J. Gibbon and S. Castles, NASA GSFC, Greenbelt, MD
Design and Qualification of Flight Electronics for the HST NICMOS
Reverse Brayton Cryocooler
. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 439
C. Konkel and W. Bradley, Orbital Sciences Corp., Greenbelt, MD; and R. Smith, NASA GSFC, Greenbelt, MD
J-T and Throttle-Cycle Cryocooler Developments
449
Flight Demonstration of the Ball Joule-Thomson Cryocooler . . . . . . . . . . . . . . 449 R. Fernandez and R. Levenduski, Ball Aerospace & Tech., Boulder, CO
Design Optimization of the Throttle-Cycle Cooler with Mixed Refrigerant . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 457 M. Boiarski, A. Khatri, APD Cryogenics, Allentown, PA; and V. Kovalenko, Moscow Power Engin. Inst., Moscow, Russia
Long-Life Cryocooler for 84-90 K
. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 467
V.T. Arkhipov, A.V. Borisenko, V.F. Getmanets, R.S. Mikhalchenko and L.V. Povstiany, SR&DB, Kharkov, Ukraine; and H. Stears, Orbita, Ltd., Kensington, MD
Mixed Gas J-T Cryocooler with Precooling Stage . . . . . . . . . . . . . . . . . . . . . . . . 475 A. Alexeev, Ch. Haberstroh, and H. Quack, Univ. of Dresden, Germany
Experimental Comparison of Mixed-Refrigerant Joule-Thomson Cryocoolers
with Two Types of Counterflow Heat Exchangers . . . . . . . . . . . . . . . . . . . . . 481 E.C. Luo, M.Q. Gong, Y. Zhou, and J.T. Liang, Chinese Acad. of Sci., Beijing, China
Multicomponent Gas Mixtures for J-T Cryocoolers . . . . . . . . . . . . . . . . . . . . . . . . 487 V.T. Arkhipov, V.V. Yakuba, M.P. Lobko, and O.V. Yevdokimova, SR&DB, Kharkov, Ukraine; and H. Stears, Orbita Ltd., Kensington, MD
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CONTENTS
An Experimental Study and Numerical Simulation of Two-Phase Flow of Cryogenic Fluids through Micro-Channel Heat Exchanger . . . . . . . . . . . . 497 W.W. Yuen, UCSB, Santa Barbara, CA; and I.C. Hsu, Lockheed Martin ATC, Palo Alto, CA
Hybrid J-T Cryocooler Systems for Operation at 4-10 K
505
Hybrid 10 K Cryocooler for Space Applications . . . . . . . . . . . . . . . . . . . . . . . . . . 505 R. Levenduski, W. Gully, and J. Lester, Ball Aerospace & Tech., Boulder, CO
Design and Development of a 4 K Mechanical Cooler . . . . . . . . . . . . . . . . . . . . . 513 S.R. Scull and B.G. Jones, MMS, Bristol, UK; T.W. Bradshaw and A.H. Orlowska, RAL, Chilton, UK; and C.I. Jewell, ESA-ESTEC, Noordwijk, The Netherlands
Life Test and Performance Testing of a 4 K Cooler for Space
Applications . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 521 T.M. Bradshaw, A.H. Orlowska, RAL, Chilton, UK; and C.I. Jewell, ESA-ESTEC, Noordwijk, The Netherlands
Long-Life 5-10 K Space Cryocooler System with Cold Accumulator . . . . . . . . . 529 V.T. Arkhipov, V.F. Getmanets and A.Y. Levin, SR&DB, Kharkov, Ukraine; and H. Stears, Orbita Ltd, Kensington, MD
Sorption Cryocooler Developments
535
Periodic 10 K J-T Cryostat for Flight Demonstration . . . . . . . . . . . . . . . . . . . . . 535 R.C. Longsworth, A. Khatri, and D. Hill, APD Cryogenics, Allentown, PA
Characterization of Porous Metal Flow Restrictors for Use as the J-T Expander in Hydrogen Sorption Cryocoolers .............................. 545 A.R. Levy, UCSB, Santa Barbara, CA; and L.A. Wade, JPL, Pasadena, CA
Thermodynamic Considerations on a Microminiature Sorption Cooler ........ 553 J.F. Burger, H.J. Holland, H.J.M. ter Brake, H. Rogalla, Univ. of Twente, The Netherlands; and L.A. Wade, JPL, Pasadena, CA
Fast Gas-Gap Heat Switch for a Microcooler . . . . . . . . . . . . . . . . . . . . . . . . . . . . 565 J.F. Burger, H.J. Holland, H. van Egmond, M. Elwenspoek, H.J.M. ter Brake, and H. Rogalla, Univ. of Twente, The Netherlands
GM Refrigerators and Low-Temperature Regenerators
575
Development of a High Efficiency 0.5W Class 4 K GM Cryocooler . . . . . . . . . . 575 T. Satoh, R. Li, H. Asami, and Y. Kanazawa, Sumitomo Heavy Ind. R&D Center, Kanagawa, Japan; and A. Onishi, Sumitomo Heavy Ind. PPD, Tokyo, Japan
Development of a High Efficiency 4 K GM Refrigerator . . . . . . . . . . . . . . . . . 581 Y. Ohtani, H. Hatakeyama, and H. Nakagome, Toshiba R&D Center, Kawasaki, Japan; and T. Usami, T. Okamura, and S. Kabashima, Tokyo Inst. of Tech., Yokohama, Japan
CONTENTS
xv
Analysis of a High Efficiency 4 K GM Refrigerator Operating at a Lower Pressure Ratio . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 587 T. Usami, T. Okamura and S. Kabashima, Tokyo Inst. of Tech., Yokohama, Japan; and
Y. Ohtani, H. Hatakeyama, and H. Nakagome, Toshiba Corp., Kawasaki, Japan
Numerical Simulation of 4 K GM Refrigerator . . . . . . . . . . . . . . . . . . . . . . . . . . 593 T. Inaguchi, M. Nagao, K. Naka, and H. Yoshimura, Mitsubishi Electric Corp. Adv. Tech. R&D Center, Hyogo, Japan
Numerical Fluid Analysis of Pumping Loss
. . . . . . . . . . . . . . . . . . . . . . . . . . . . . 603
K. Naka, T. Inaguchi, M. Nagao, and H. Yoshimura, Mitsubishi Electric Corp. Adv. Tech. R&D Center, Hyogo, Japan
Multilayer Magnetic Regenerators with an Optimum Structure
around 4.2 K . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 611 H. Nakane, T. Hashimoto, and Y. Miyata, Kogakuin Univ., Tokyo, Japan; M. Okamura and H. Nakagome, Toshiba Corp., Kanagawa, Japan
Advances in Neodymium Ribbon Regenerator Materials . . . . . . . . . . . . . . . . . . 621 T. Felmley, Concurrent Tech. Corp., Johnstown, PA
Gd-Zn Alloys as Active Magnetic Regenerator Materials for Magnetic Refrigeration . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 629 V.K. Pecharsky and K.A. Gschneidner, Jr., Ames Lab, Iowa State Univ., Ames, IA
Magnetocaloric Properties of Gd3Al2 ................................................ 639 V.K. Pecharsky and K.A. Gschneidner Jr., Ames Laboratory, Iowa State Univ., Ames, IA;
and S. Y. Dan’kov and A.M. Tishin, Moscow State Univ., Moscow, Russia
Advanced Refrigeration Cycles and Developments
647
Development of a Dilution Refrigerator for Low-Temperature Microgravity Experiments . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 647 P.R. Roach, NASA ARC, Moffett Field, CA; and B. Helvensteijn, Sterling Software, Redwood Shores, CA
Preliminary Experimental Results Using a Two Stage Superfluid Stirling Refrigerator . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 655 A.B. Patel and J.G. Brisson, MIT, Cambridge, MA
Investigation of Microscale Cryocoolers
. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 663
J.M. Shire, A. Mujezinovic, and P.E. Phelan, ASU, Tempe, AZ
Cryocooler Integration and Test Technologies
671
Development of Advanced Cryogenic Integration Solutions . . . . . . . . . . . . . . . 671 D. Bugby and C. Stouffer, Swales Aerospace, Beltsville, MD; T. Davis, Lt. B.J. Tomlinson, and Lt. M. Rich, AFRL, Kirtland AFB, NM; J. Ku and T. Swanson, NASA GSFC, Greenbelt, MD; and D. Glaister, The Aerospace Corp., Albuquerque, NM
Cold Accumulators as a Way to Increase Lifetime and
Cryosystem Temperature Range . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 689 V.T. Arkhipov, V.F. Getmanets, A.Y. Levin, and R.S. Mikhalchenko, SR&BD, Kharkov, Ukraine; and H. Stears, Orbita Ltd, Kensington, MD
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CONTENTS
Test Results of a Nitrogen Triple-Point Thermal Storage Unit . . . . . . . . . . . . . 697 B.G. Williams and I.E. Spradley, Lockheed Martin ATC, Palo Alto, CA
Optimal Integration of Binary Current Lead and Cryocooler . . . . . . . . . . . . . . 707 H.M. Chang, Hong Ik Univ., Seoul, Korea; and S. W. Van Sciver, Nat’l High Magnetic Field
Lab, Tallahassee, FL
Cryogenic Systems Integration Model (CSIM) . . . . . . . . . . . . . . . . . . . . . . . . . . . 717 S.D. Miller and M. Donabedian, The Aerospace Corp., El Segundo, CA; and D.S. Glaister, The Aerospace Corp., Albuquerque, NM
Heat Rejection Effects on Cryocooler Performance Prediction . . . . . . . . . . . . . 723 Lt. B.J. Tomlinson, AFRL, Kirtland AFB, NM; and A. Gilbert and J. Bruning, NRC, Albuquerque, NM
Cryocooler Working Medium Influence on Outgassing Rate . . . . . . . . . . . . . . . 733 V.F. Getmanets, SR&DB, Kharkov, Ukraine; and G.G. Zhun', Kharkov State Polytechnic Univ., Kharkov, Ukraine
Accelerated Cryocooler Life Tests for Cryodeposit Failures . . . . . . . . . . . . . . . . 743 V.F. Getmanets and G. G. Zhun, SR&DB, Kharkov, Ukraine; and H. Stears, Orbita, Ltd,
Kensington, MD
Thermal Resistance across the Interstitial Material Kapton MT at Cryogenic Temperatures . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 753 L. Zhao and P.E. Phelan, ASU, Tempe, AZ
Space Cryocooler Applications
761
Cryocooler Subsystem Integration for the High Resolution Dynamics Limb Sounder (HIRDLS) Instrument . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 761 D.J. Berry, D. Gutow, J. Richards, and R. Stack, Ball Aerospace & Tech., Boulder, CO
EMI Performance of the AIRS Cooler and Electronics . . . . . . . . . . . . . . . . . . . . 771 D.L. Johnson, S.A. Collins, and R.G. Ross, Jr., JPL, Pasadena, CA
The Application and Integration of Mechanical Coolers . . . . . . . . . . . . . . . . . . . 781 R.M. Wilkinson, S.R. Scull, MMS, Bristol, England; A.H. Orlowska and T.W. Bradshaw, RAL, Chilton, England; and C.I. Jewell, ESA-ESTEC, Noordwijk, The Netherlands
Cooling System for Space Application . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 787 G.L. Ji and Y. Wu, Shanghai Inst. of Tech. Physics, Chinese Acad. of Sci., Shanghai, China
Drive and Control System for a Stirling Cryocooler . . . . . . . . . . . . . . . . . . . . . . 791 W. Biao, G. Ji, and Y. Wu, Shanghai Inst. of Tech. Physics, Chinese Acad. of Sci., Shanghai, China
Testing of Infrared Detectors Using a Zero Gravity Dilution Refrigerator . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 795 R.S. Bhatia, J.J. Bock, and P.V. Mason, CIT, Pasadena, CA; A. Benoît, CNRS, Grenoble, France; and M.J. Griffin, Queen Mary & Westfield College, London, UK
Design of a 90 K Cryogenic Passive Cooler for the IASI Instrument . . . . . . . . 805 D.J. Doornink, Fokker Space, Leiden, The Netherlands
CONTENTS
xvii
Cryocoolers for Human and Robotic Missions to Mars . . . . . . . . . . . . . . . . . . . . 815 P. Kittel and L.J. Salerno, NASA ARC, Moffett Field, CA; and D.W. Plachta, NASA LeRC, Cleveland, OH
Commercial Cryocooler Applications
823
Design Considerations for Industrial Cryocoolers . . . . . . . . . . . . . . . . . . . . . . . . 823 C.M. Martin and J.L. Martin, Mesoscopic Devices, LLC, Golden, CO
Survey of Cryocoolers for Electronic Applications (C-SEA) . . . . . . . . . . . . . . . . 829 J.L. Bruning and R. Torrison, NRC, Albuquerque, NM; R. Radebaugh, NIST, Boulder, CO; and M. Nisenoff, NRL, Washington, DC
Construction and Tests of a High-Tc SQUID-Based Heart Scanner Cooled by Small Stirling Cryocoolers . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 837 C.J.H.A. Blom, H.J.M. ter Brake, H.J. Holland, A.P. Rijpma, and H. Rogalla, Univ. of Twente, The Netherlands
Cryocooler Applications for High-Temperature Superconductor Magnetic Bearings . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 847 R.C Niemann and J.R. Hull, Argonne Nat’l Lab, Argonne, IL
Advanced Cryocooler Cooling for MRI Systems . . . . . . . . . . . . . . . . . . . . . . . . . . 857 R.A. Ackermann and K.G. Herd, GE Corp. R&D, Niskayuna, NY; and W.E. Chen, GE Medical Sys., Florence, SC
Indexes
869
Proceedings Index . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 869 Author Index . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 871 Subject Index . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 873
An Overview of the Performance and Maturity of Long Life Cryocoolers for Space Applications
D. S. Glaister The Aerospace Corporation Albuquerque, NM, USA 87119 M. Donabedian, D. G. T. Curran The Aerospace Corporation El Segundo, CA, USA 90245 T. Davis The Air Force Research Lab Kirtland AFB, NM, USA 87119
ABSTRACT A survey is made which identifies more than 30 long life coolers for space applications covering a wide variety of thermodynamic cycles and configuration types. These coolers range in capacities from a few milliwatts to over 10 W at temperatures from 10 K to over 120 K and include single and multi-stage designs. The primary objectives of this study were to provide a hardware summary and performance comparison for potential space cryocooler users and to serve as an aid to the Air Force in determining future cryocooler development. Funding for these coolers are being provided by the Department of Defense, NASA, and various other government agencies in the U.S. and abroad as well as by internal research and development moneys from a number of companies throughout the world. The study identifies several existing flight qualified coolers and at least 12 programs which are likely to provide flight qualified units for cooling in the range of 10 to 150 K before the turn of the century. The survey presents an overview and status of the maturity of the various cryocoolers and performance comparisons are made at 35 K, 60 K, and 100 K.
Cryocoolers 10, edited by R. G. Ross, Jr. Kluwer Academic/Plenum Publishers, 1999
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ACRONYMS
AATSR AFRL AIRS ATSR BAe BETSCE BMDO CCE COBE COOLLAR DOD EMD EOS ESA FDS FIRST GSFC HIRDLS HST HTSSE IMAS IRFPA ISAMS ISSC JPL LADS LMMS MIPAS MMRBC MMS MOPITT MSX MTI NICMOS PSC RAL
SBIRS-L SCRS SMC SMTS SPIRIT III SSC SSRB SSTI STS TES TMU UARS UCSB USU
Advanced Along Track Scanning Radiometer Air Force Research Laboratory Atmospheric Infrared Sounder Along Track Scanner Radiometer British Aerospace Systems, Limited Brilliant Eyes Ten Kelvin Sorption Cryocooler Experiment Ballistic Missile Defense Organization Cryocooler Control Electronics Cosmic Background Explorer Cryogenic On-Orbit Long Life Active Refrigeration Department of Defense Engineering and Manufacturing Development Earth Observation System European Space Agency Flight Demonstration System Far Infrared Space Telescope Goddard Space Flight Center High Resolution Dynamic Limb Sounder
Hubble Space Telescope High Temperature Superconductivity Space Experiment Integrated Multi-spectral Atmosphere Sounder Infrared Focal Plane Assembly Improved Stratospheric and Mesospheric Sounder
Improved Standard Spacecraft Cooler Jet Propulsion Laboratory Low Altitude Demonstration System
Lockheed Martin Missiles and Space Micholson Interferometer for Passive Atmosphere Sounding Multi-Stage Miniature Reverse Brayton Cryocooler
Matra Marconi Space Measurement of Pollution in the Tropopause Mid-Course Space Experiment Multi-spectral Thermal Imager Near Infrared Camera and Multi-Object Spectrometer Protoflight Space Cryocooler Rutherford Appleton Laboratories, U. K. Space Based Infrared Surveillance-Low Space Cryogenic Refrigeration Systems Space and Missiles System Center Space Missile Tracking System Space Infrared Imaging Telescope Standard Spacecraft Cooler Single-Stage Reverse Brayton Small Satellite Technology Initiative Space Transportation System Tropospheric Emission Spectrometer Thermo-Mechanical Unit Upper Atmosphere Research Satellite University of California at Santa Barbara Utah State University
OVERVIEW OF PERFORMANCE OF SPACE CRYOCOOLERS
3
INTRODUCTION
A variety of long life, mechanical cryogenic refrigerators (cryocoolers) for space are available or under development to provide cooling of infrared sensors and spectrometers, optical elements, low noise amplifiers, superconductivity devices and other scientific instruments for atmospheric monitoring and astronomy. The authors of this paper provide technical support to nearly every United States Department of Defense (DoD) spacecraft program either implementing or potentially implementing cryogenic cooling systems, as well as to the Air Force Research Laboratory (AFRL) at Kirtland AFB, NM, which is the DoD center for space cryocooler and cryogenic technology development. As part of that support, the authors are responsible for assessing the industry as well as routinely presenting overviews and summaries of space cryocoolers. The AFRL requested the preparation of an overview package of the status and maturity of space cryocoolers. The intent of this package is to help assess the state of the art and provide guidelines for future development of space cryocoolers. The purpose of this paper is to present a brief overview of the cryocooler data package. It is also the intent of the AFRL that this cryocooler data package be available to any user or vendor who requests it. Several caveats should be mentioned concerning the scope of this overview package. In recent years, the number of cryocoolers available with potential application to space has increased significantly. It is the authors’ intent to include every cryocooler whose primary purpose is for space application. However, it is likely that some vendors have been missed and the package is incomplete. In this regard, the authors’ encourage input from those vendors who were left off the summary in this paper.
SPACE CRYOCOOLER STATUS AND OVERVIEW As a result of the large number of coolers representing several different thermodynamic
cycles as well as several different hybrid combinations, and many different vendors, time or space does not permit including all of these coolers. Rather, a selected summary of some of the more prominent coolers and programs are provided in an attempt to provide a representative sample of the total population.
Tables 1a through 1e have been prepared to provide an overview. The coolers are listed by vendor in alphabetical order showing the model, cooler type, nominal performance, power input including electronics either estimated or measured, the current maturity level ( using the legend at the end of the table), applicable programs, sponsors, milestones or significant schedule dates, environmental and life testing completed or in progress and finally references cited for the source of the information. In some cases, multiple model designations are used when there is a difference between the vendor’s and sponsor’s designation. In the discussion to follow, the various coolers are grouped by the nominal operating temperature range for which the cooler is best suited. In some cases this will require identifying a specific cooler more than once if it is
being used over a broad range of temperatures. The four specific temperature ranges used to categorize the coolers are as follows: 1) 4-12 K, 2) 18-45 K, 3) 50-100 K and 4) over 100 K. It should be noted that new applications with varying temperature requirements and heat loads will need revised designs to optimize performance. This is necessary as the listed coolers have been optimized for single design points such as 2 W @ 60 K but are capable of providing cooling capabilities at lower and higher temperatures. Several of these applications demand weight reductions and higher efficiencies for ~100 K optics cooling due to considerations such as stringent gimbal mass limitations which also limit heat rejection capabilities. The very high specific powers at ~10K temperatures have also restricted the use of coolers in this range.
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OVERVIEW OF PERFORMANCE OF SPACE CRYOCOOLERS
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OVERVIEW OF PERFORMANCE OF SPACE CRYOCOOLERS
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OVERVIEW OF PERFORMANCE OF SPACE CRYOCOOLERS
Notes to Table 1. 1. TMU = Thermo-mechanical unit CCE = Cooler control electronics (assumed to be 6.0 kg if not
otherwise specified) 2. Maturity legend CD Conceptual design BB Brassboard EDM Engineering development model QM Qual model PF Protoflight model FM Flight model FP Flight proven
COTS Commercial off the shelf
9
3. Environmental tests completed a. Random vibration b. Sine vibration
c. d. e. f.
Thermal Cycle Thermal Vacuum
EMI/EMC Shock g. Flight qualifications 4. Total power input (including CCE) estimated to be equal to 10 watts + compressor input power/0.85 when not specifically measured. At 300 K unless otherwise noted (usually
measured or assumed to be at the cooler mounting flange). 5. Derivative of 50-80 K
6. Multistage requirement modified to singlestage requirement 7. Testing hours left blank when unknown or minimal
Also, now that clearance seal concepts with flexure and gas bearings are showing promise for long life, attention needs to be focused on improving performance by reducing cooler weight and reducing irreversible losses in the cooler. For these reasons, optimization techniques need to be introduced, such as those of Bejan46, to minimize both temperature and pressure drops in the cooler components such as critical cold end and hot end working fluid heat exchangers as well as
regenerators and recuperators. In addition both cooler and payload/spacecraft system designers
need to optimize the cooling capabilities and load requirements by utilizing temperature staging of cooling loads to improve overall system efficiency. 4-12 K Operating Range
Traditionally, cooling in this range has been accomplished with the use of superfluid liquid helium dewars such as IRAS44 and COBE45 at temperatures near 2 K or solid hydrogen cryostats such as the SPIRIT-III flown on the MSX Spacecraft for cooling near 10 K. There has been an incentive to develop cryocoolers for operation in this range because of the desire for longer life with reduced mass and volume relative to dewars. Coolers with capacity for
continuous cooling in the 50 to 100 milliwatt range are just now emerging. This region of operating temperatures is the least mature technology at this point in time.
MMS 4K Hybrid J-T/Stirling. This cooler is the culmination of a development activity started at RAL in the mid-1980’s. Under ESA funding, BAe (prior to being acquired by MMS)
developed an engineering model of a 4 K cooler which has since been qualified for spaceflight by MMS. The cooler consists of a two-stage Stirling pre-cooler which is integrated with a J-T cryostat to achieve final collection of liquid helium to provide cooling at near 4 K. The unit produces 5 to 10 milliwatts of cooling for somewhat over 200 W of input power. The total mass is about 50 kg and is a prime candidate for several astronomy missions including the Far Infrared Space Telescope (FIRST)29. MMS 10 K Multi-Stage Stirling. This program was initiated by AFRL/BMDO in early 1997 with the objective to quickly produce a protoflight quality cryocooler with at least 45 mW of
cooling at 10 K. Towards this goal of a fast turn around, the program leverages off of the MMS 20 K hardware and uses essentially a double set of the 20 K compressors (for a total of 4). Through regenerator improvements with rare earth materials, the program now has the potential to achieve up to 75 mW of cooling with a motor power of 2000 W/W or less. The program is on
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GOVERNMENT CRYOCOOLER DEVELOPMENT AND TEST PROGRAMS
schedule to complete a flight quality thermo-mechanical unit by April 1999. The intent of this program is not only to produce a 10 K flight cooler for customers who may need a cryocooler in the next 3 years, but also to serve as a pathfinder for future, more optimized Stirling and Pulse Tube 10 K programs. JPL BETSCE 10 K Periodic Sorption. The Brilliant Eyes Ten Kelvin Cryocooler Experiment
(BETSCE) was developed under funding from USAF/BMDO/AFRL to demonstrate the technology of a hydride sorption cryocooler for operation near 10 K and is described in detail by Bard15 and Wade16. The flight demonstration unit flown on STS-77 in May of 1996 was a periodic cooler supplemented with a Stirling cryocooler to provide 70 K pre-cooling. The flight unit provided 100 milliwatts of cooling at 10 K (on a 15 minute per 24 hour duty cycle ) starting from 70 K with a power input of about 200 W @ 290 K. This program was completed but the technology derived has been exploited further by application to continuous cooling cryocoolers in the 20-25 K region by Wade16 and Bowman17. These applications are described in the next section of this report for coolers in the 18-45 K range. A more detailed description of the design and performance of the 10 K JT cryostat of the BETSCE program is presented by Longsworth34.
Ball Hybrid J-T/Stirling (Redstone). This program was recently initiated by AFRL with the goal of pushing the state of the art and developing a very efficient (less than 1000 W/W motor power), light weight 10 K cryocooler for space applications (including doped Silicon infrared detectors). The design utilizes an enhanced 35/60 K Stirling unit (with some minor regenerator
improvements) to provide precooling at 15 to 18 K to a J-T cooler which provides at least 100 mW of cooling at 10 K. The J-T compressor uses a rotary vane design with significant heritage from terrestrial commercial applications, but which requires development for long life
application with a dry helium working fluid. This hybrid design takes advantage of the Stirling efficiency in cooling to cryogenic temperatures from ambient and the efficiency of the recuperative J-T cycle in cooling as the temperature approaches absolute zero. 18-45 K Operating Range
There are several coolers designed primarily for operation near the lower end of this temperature range and several that are best suited for the 35-45 K range but also operate very efficiently in the 50-80 K range. Engineering model coolers have been available for several
years in this operating temperature range and flight qualified coolers are becoming available. Qualified control electronics are lagging somewhat behind but are being developed. MMS 20 K Multistage Stirling. The MMS 20K cooler under development is based primarily on the 2-stage technology achieved at RAL under ESA funding. The program was initiated in early 1994 to provide 20 K cooling capability for the FIRST instrument. The back-to-back compressors are derived from the 50-80 K technology combined with a 2-stage displacer and produces approximately 120 milliwatts at 20 K plus 400 milliwatts at 30 K for a total input power of 122 W. An extensive flight qualification program was completed in early 199728. BALL 30 K Multi-Stage Stirling. This cooler was developed by NASA GSFC and was derived from earlier coolers using RAL/Oxford technology. The cooler uses diaphragm spiral flexure springs with integral back-to back compressors and a split expander with balancer. The expander
has incorporated a fixed regenerator design allowing for a lightweight displacer/piston and improved cooling efficiency. The cooler has undergone successful environmental and performance testing conducted at GSFC and is presently under life test with approximately 7000 hours of accumulated operation. BALL 35/60 K Dual Temperature Cooler. This cooler is being developed by AFRL as a candidate for cooling sensor IRFPA’s such as SBIRS LOW. It is a three-stage unit with the upper stage used as a shield for the lower stages that provide 0.4 W @ 35 K and 0.6 W at 60 K. The design was derived from the two stage NASA GSFC 30K program that also used the upper stage as a shield with 0.3 W @ 30 K. The cooler provides very efficient cooling with a TMU
OVERVIEW OF PERFORMANCE OF SPACE CRYOCOOLERS
11
specific power estimated to be 30 W/W at 60 K and 80 W/W at 35 K. The cooler will be delivered to AFRL for performance and endurance testing. JPL 20 K Continuous Sorption. JPL is currently developing a 20K continuous sorption cooler (LH2) for the Planck Surveyor program sponsored jointly by NASA and ESA (European Space Agency). The system utilizes hydride compressors connected to a J-T cryostat assembly to produce liquid hydrogen operating near 20 K. A passive radiator operating in deep space at
about 60 K allows the system to produce about 1.2 W of cooling at 20 K with around 400 W of input power. An EDM is scheduled to be delivered by the year 2000 with two flight units scheduled to be delivered to ESA in 2003. This cooler uses similar technology developed by BETSCE and the 25K UCSB Long Duration Balloon Cooler35. Creare 35K Single Stage Reverse Brayton or MMRBC. This cooler is being developed by AFRL as a candidate for cooling sensor IRFPA’s such as SBIRS LOW. NASA GSFC has also supported this technology and is currently using components of the SSRB and MSRB for the
NICMOS flight cooler and circulator. The technology involves component improvements over the SSRB to improve efficiency at lower cooling loads and temperatures. The original program requirements were changed from providing multi-stage cooling of 0.4 W @ 35 K and 0.6 W @ 60 K to a single stage providing 1.0 W at 35 K for approximately the same 100 watts of power. The new and smaller components involve using permanent magnets for both the compressor motor (replacing a less efficient induction motor) and cryogenic turboalternator and the use of a
more compact recuperator design for the heat exchanger. The turboalternator replaces the SSRB turboexpander allowing reduced parasitics and elimination of brake flow/cooling. This cooler
will be delivered to AFRL for performance and endurance testing. LMMS LADS 35 K Single Stage Stirling. This cooler is a slight modification of the L1710C
which is based on previous Lucas-built coolers. The cooler utilizes RAL/Oxford technology such as linear motors and spiral-flexure bearing supports to maintain clearance seals in the expander and back-to-back compressors.The modifications have included improved compression space porting and change in position sensor design. Current predictions based on L1710C test data will provide 0.5 W @ 35 K for a TMU specific power of 138 W/W. Two L1710C units have been built but have not been life tested. These units have accumulated a few thousand hours in various lab tests and one Lucas-built cooler has accumulated over 16,000 hours. Control electronics have been flight qualified to provide for cancellation of axial residual vibration and temperature stability of the cold-tip. An LMMS compressor has also demonstrated low lateral vibrations using tangential-arm flexures in limited testing.
TRW 35 K Model PTC-010A-035-I and PTC-020C-035-I Single Stage Pulse Tubes. These coolers were developed under separate AFRL programs and use previously developed technology under NASA/DoD contracts. The back-to-back compressors utilize flexure bearing supports for clearance seals and linear motor compressor drive as do their Stirling counterparts and are derived from Oxford technology. The pulse tube coolers have been sized to deliver 0.3 and 0.85W at 35K. These coolers are currently at AFRL for performance and endurance testing. These units were also performance characterized at JPL. The larger capacity unit has accumulated about 5,000 hours while the smaller unit has about 4,000 hours. Control electronics have been developed and flight qualified for the smaller capacity unit. Raytheon Single Stage Stirling 35 K PSC/SMTS. These TMU’s are similar to their SSC/ISSC counterparts. The SSC was designed under an AFRL development program to provide 2 W @ 60K. The ISSC was an improved version developed under IRAD and used for early SBIRS LOW life testing at 60 K which accumulated 46,000 hours on two units. Another unit successfully flew on a shuttle mission. The PSC program was initiated in late 1992 but the SMTS began only two years ago. Both units have benefited from numerous improvements to increase the TMU efficiency. The PSC unit has incorporated tangential-arm flexure springs for
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GOVERNMENT CRYOCOOLER DEVELOPMENT AND TEST PROGRAMS
clearance seal support in the back-to-back compressor which has recently been demonstrated at JPL to substantially reduce vibrations in both axial and lateral axes. This Aerospace Corp. tangential-arm design had previously been demonstrated only in flexure module tests and those mentioned for the LMMS compressor. The axial residual vibrations were canceled by control electronics feed-back techniques as with other controllers used on the TRW/Ball/LMMS units. The SMTS units incorporate a heat-intercept concept first demonstrated at JPL to substantially reduce cold-finger parasitics in Stirling coolers. The PSC will be delivered to AFRL for performance and endurance testing. An SMTS life test unit for the SBIRS LOW FDS program has accumulated ~9000 hours. Control electronics for the PSC are flight-type. The flight electronics for the SMTS coolers are rad-hard and are being flight qualified but are not standalone as they are integrated with other payload functions for FDS. 50-100 K Operating Range
The largest group of coolers are in this range with the majority of the units initially designed to operate in the 55 to 75 K range. This area represents the most mature technology. Several coolers have been flown since the early 1990’s primarily to support science experiments or technology demonstrations. Simple flight electronics have been tested and flown. However, fully qualified electronics with integrated vibration control capability have only just begun emerging during the last few years. These are being integrated into flight systems which will be
launched in the near future.
Ball 58 K HIRDLS Instrument. Ball began development of Stirling coolers in 1990 with purchase of a license to build clones of RAL “Oxford” type coolers. Beginning in 1992, Ball developed two lines of single-stage cryocoolers aimed at operation at 60 K. On IRAD, Ball built a series of two improved designs maintaining the RAL heritage. The initial design achieved 1.5 W of cooling at 55 K and is currently on life test at Ball with >23,000 hours. The second version of this unit (SA160) provides 2.5 W of cooling at 60 K for 116 W to the compressors. In parallel, Ball developed a single-stage version of its SB230 cryocooler. This cooler is a fixed regenerator design capable of lifting 1.6 W at 60 K for 53 W to the compressor. This cooler is scheduled to fly on the HIRDLS instrument on the NASA EOS Chem Platform in 2002. Ball COOLLAR 65/120K J-T. This program has been in existence for over 5 years and
culminated in a successful micro-gravity verification and characterization aboard the Space Shuttle in August of 1997. The design uses an oil lubricated compressor to provide the high pressure ratio for a multi-stage J-T (with some thermo-electric cooler assisted precooling) cold head which produces 5 W of cooling at 120 K and an average of 1.25 W at 65 K. Because this JT cooler produces and stores liquid nitrogen at the cold interfaces, it has the ability to load level a variable load at the 65 K stage. It was thus designed to meet a 65 K load profile with peak loads of 3.5 W. The cooler was also designed with long flexible lines leading to the J-T cold head to enable the remote mounting of the compressor. Creare 65 K Single Stage Reverse Brayton (SSRB). This cooler was designed for long life using high speed, small turbomachines with gas bearings allowing vibration and wear-free operation to provide 5 W of cooling at 65 K. The cooler consists of small-size precision devices.
The cycle operates with continuous flow of neon gas circulated by a compressor through a recuperative heat exchanger and turbine expander. Additional components include a high efficiency inverter/controller, aftercooler and load heat exchanger. This technology has been supported by both NASA and AFRL. The cooler has achieved its highest efficiency with a system (cooler and electronics) specific power of 37 W/W for 7.5 watts of cooling @ 65K and a heat rejection temperature of 280 K. The cooler is continuing endurance testing at AFRL and has accumulated ~25,000 hours.
OVERVIEW OF PERFORMANCE OF SPACE CRYOCOOLERS
13
Creare 65 K Stirling Diaphragm Cooler (SSC). In the late 1980’s both the USAF and BMDO
recognized the need for a long life cryocooler capable of handling small loads in the 50-80K temperature range. As part of the general standard spacecraft cooler (SSC) program, Creare was funded to develop an alternative technology to the flexure spring cryocoolers being developed by several sources. As part of this effort, Creare developed an engineering model of a cooler designed to provide 2 W of cooling at 65 The EDM was delivered to the AFRL in early 1994 and has accumulated over 10,000 hours of endurance testing after its performance was characterized and documented. The program has been completed. Creare 70 K NICMOS (SSRB). The Near-Infrared Camera and Multi-Object Spectrometer (NICMOS) is a second generation Hubble Space Telescope (HST) science instrument whose detectors are cooled to about 58 K by a solid Nitrogen cryostat. An anomaly discovered in late 1997 indicated that the operating life of the dewar would reduce the expected life from 48 months to about 1.7 years. As a result the NASA Goddard Spaceflight center (GSFC) decided to attempt to retrofit the solid cryostat with an improved version of the Creare 5 W SSRB. The new cryocooler is designed to provide approximately 7.7 W at 70 K for a total input power of around 315 W. The new system along with a newly developed flight electronics package is to be flight tested aboard the STS in late 1998 with final integration into the HST in early 2000. LMMS 1710-C/SCRS Stirling. Since 1987 Lockheed-Martin has been developing advanced
cryocooler systems based on the “Oxford” technology under a teaming arrangement with Lucas Aerospace (since terminated). Several different models emerged from this activity including the 1710-C and SCRS discussed in this paragraph and the LADS unit which was previously
discussed. The 1710-C consists of back-to-back compressors connected to a single displacer, providing about 2.0 W at 60 K with a total power input of 130 W including the control electronics package which has been flight qualified. Utilizing similar hardware with a slightly larger compressor piston diameter, an integrated system was developed as the SCRS, funded jointly by several USAF organizations as part of the SPAS-III flight experiment. Two compressors and two displacers which have their cold tip attached via flex couplings in a common vacuum housing provided about 1.2 W of cooling at 59 K. This assembly was tested and delivered to the Utah State University for overall system testing. Raytheon Stirling 60 K PSC/SMTS/ISSC/SSC. As stated for the Raytheon Stirling 35K units, these coolers have a common TMU heritage (e.g., expander/compressor sizing) in that they were
originally designed for cooling 2 W @ 60/65K. Both the PSC and SMTS coolers are the same units for 35 and 60 K application. The SMTS unit has a lower cooling requirement for 60K
operation of the same order as the 35K requirement. The SMTS life test cooler has been operated at 60 K as well as at 35 K mentioned above as part of the ~9000 hours of operation. Also as mentioned above one ISSC has accumulated nearly 24,000 hours while the other life test unit accumulated 22,000 hours. The respective TMU specific powers for the ISSC and SSC units @ 65 K are 45 W/W for 1.75 and 1.2 watts. The respective TMU specific powers for the PSC and SMTS (with heat intercept) are 27.5 and 22.4 W/W at 60 K. Without heat intercept the SMTS TMU specific power is about the same as the PSC TMU which does not use the heat intercept, i.e., 27.5 W/W. The ISSC #4 was modified to improve motor efficiency and thermal interfaces.
It was also tested under higher charge pressure and achieved improved performance. TRW 60K PTC Models for AIRS, TES, 6020, and IMAS Programs. These coolers in both integral (I) and split (S) configurations have been developed for both NASA and Air Force programs. They have common heritage to the above mentioned TRW 35K coolers. Two flight models of the PTC-010B-055-S (55 K) have been delivered to the NASA AIRS program and two flight units of the PTC-010C-057-I (57 K) will be delivered to JPL in 1999 for TES. PTC-010A060-I (60 K) will be delivered as a flight unit to MTI in 1998. An EDM of this unit was delivered to AFRL for endurance and performance testing after being performance characterized
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GOVERNMENT CRYOCOOLER DEVELOPMENT AND TEST PROGRAMS
at JPL. PTC-004A-055-I will be delivered to JPL in 6/98 as two EDM units for the IMAS program. TRW 60-150K Integral Miniature Stirling (SC-0-6A-65-I) and Integral Miniature Pulse
Tube (PTC-001A-115-I). These coolers were developed for DOD requirements and have been delivered for several applications including life testing and optics cooling for SBIRS LOW. Two mini-Stirling coolers were life tested for ~15,000 hours each to provide 0.16 W @ 60 K for an initial SBIRS LOW IRFPA requirement. The corresponding TMU specific power at 60 K is 88 W/W. At the same time TRW ran life tests on two IRAD coolers for the same program, one of which has accumulated ~27,000 hours to date. The Stirling cooler will be a part of the HTSSEII payload to be launched on the ARGOS spacecraft in December 1998. Two miniature P-T’s are currently operating on the CX payload and have accumulated 3 months of operation. MMS 50-80K Stirling. This cooler is a modification of the single stage 80K cooler formerly known as the BAe 80K. The 80K cooler was built and qualified for an ESA contract under a licensing agreement from Oxford University where the original developmental work was
performed. The heritage of this design is based on two single stage coolers developed by Oxford and RAL which were flown as part of the Improved Stratospheric and Mesospheric Sounder (ISAMS) instrument aboard the Upper Atmosphere Research Satellite (UARS) in September of
1991. These coolers accumulated well in excess of 25,000 hours in orbit. Several ground test units have accumulated over 50,000 hours each. Also, RAL has built two similar single stage coolers based on the same RAL/Oxford pressure modulator technology that was flown successfully for up to 7 years in the 1970’s before the instruments (Pioneer Venus, Nimbus 6 and 7) were turned off. The first unit flew as part of the Along Track Scanning Radiometer (ATSR) instrument aboard the Earth Resources Satellite (ERS-1) in July 1991 and was funded by ESA. The second unit flew on ATSR-2 which replaced ATSR-1. The total flight hours for both RAL coolers have exceeded 60,000 hours. The basic 50-80K unit which is also flight qualified has been made in batches of 15 and as of this writing over 45 units have been manufactured. A number of programs sponsored by both NASA and ESA are scheduled to fly this unit over the next several years including Micholson Interferometer for Passive Atmosphere Sounding (MIPAS )and the Measurement for Pollution in the Tropopause (MOPITT) instruments. The nominal performance of this unit is usually quoted as 1.7 W at 80 K but it is being used over a wide range of temperatures from about 58 K to 90 K. Details of the acceptance and qualification programs are defined by Jones43 and Davies27. Individual life test units have accumulated over 20,000 hours. Above 100 K
Although virtually all of the coolers in the previous two ranges can be operated at much higher temperatures, the performance of the TRW Miniature P-T (MPT) and miniature Stirling (MSC) units have been characterized at temperatures up to 150K specifically to be compatible with cooling optics, shields , etc. The MPT cooler TMU specific power for 1.5 W @ 115 K is ~13.4 W/W. for 2.5 W @ 150 K is 8.2 W/W, and for 1.25 W @ 175 K is 5.8 W/W. The MSC
TMU specific power for 0.5 W @ 100 K is 15.5 W/W, for 0.8 W @ 120 K is 9.7 W/W, and for 1.3 W @ 150 K is 6.4 W/W. However the Raytheon PSC/ISSC, TRW PTC’s, LMMS, Ball and Creare coolers have been proposed and in some cases tested for operation at the 100 to 120 K level. For gimbaled optics with limited heat rejection capabilities, projected requirements are
cooling loads of 6 to 10 watts at ~100K with TMU specific powers as low as 8 to 10 W/W and lightweight units of 3 to 5 kg. Current performance of the Raytheon PSC and TRW PT’s are in the 12 to 14 W/W range at this temperature with masses over 12 kg. All of the above coolers would require re-design to meet these 100 K cooling and mass requirements. Because of this deficiency, the AFRL has initiated the High Efficiency program to develop technology to meet the future gimbaled optics requirements.
OVERVIEW OF PERFORMANCE OF SPACE CRYOCOOLERS
15
SPACE CRYOCOOLER PERFORMANCE COMPARISON Methods and Criteria
The cryocooler capacity, power efficiency, and mass efficiency data are presented here at reference temperatures of 35, 60, and 100 K. The data was compiled from numerous vendor and government sources1-46. The data was interpolated or normalized to a reference cooling load temperature and a 300 K rejection temperature using the following Carnot cycle ratio:
where the Q is the cooling capacity and T is the temperature either at the cold tip (ct) or heat rejection (rej) interface. The use of this ratio assumes that the Carnot efficiency (or refrigerator coefficient of performance) is the same for both the data and reference conditions. The equation calculates the cooling capacity at the reference conditions for the same input power as the data. The Carnot ratio was also used to interpolate the upper temperature cooling loads of a few multistage cryocoolers to allow a (very) rough comparison to single stage units. There are many caveats to the database. The error associated with the Carnot extrapolation increases as the data conditions deviate from the reference. For cooling load temperature differences greater than about 5 K or rejection differences greater than about 20 K, the Carnot interpolation is suspect. Also, because the database for flight quality cryocooler electronics controllers is limited, the input power from the electrical bus for most of the data has been
estimated from the measured motor power. A “generic” electronics estimate of 6 kg with 85% efficiency and 10 W of quiescent power was used for cryocoolers which do not currently have flight like or flight quality electronics. The estimate of electronic power can easily result in inaccuracies of 5%. Overall, the database is to be used only for approximate (± 10-20% at best) comparisons between cryocoolers. Also, the reference cooling load temperature is at the cryocooler cold block interface, which can be significantly (typically 5 K for 1 to 2 W loads) colder than the cooled instrument. For cryocoolers or applications where the cooler cold head can be directly (without a thermal strap) mounted to the instrument interface, this temperature gradient can be significantly reduced. Because of their long flexible lines from their compressors to the cold head, the Joule-Thomson and Brayton cycle cryocoolers can more easily be mounted directly to the instrument interface. Thus, the Joule-Thomson and Brayton cryocoolers have the potential to be run at higher (about 3-5 K for 1-2 W loads) cold tip temperatures (and, thus, decreased input power and increased efficiency) while still maintaining the same instrument temperature. Since there are only a few existing flight electronic hardware units, a relative evaluation of the cryocooler motor powers is often more accurate than comparing the total input power. The motor power may be misleading for units or cycles which have significantly different electronic power requirements. The relative efficiency of the Brayton cryocooler improves using total power compared to using only motor power because of less power consumption in the electronics.
Cryocooler Mass and Power Performance at 35, 60, and 100 K Figures 1 and 2 are plots of cryocooler motor specific (input divided by cooling capacity) and total (motor and electronics) power, respectively, for cooling at 60 K as a function of cooling
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GOVERNMENT CRYOCOOLER DEVELOPMENT AND TEST PROGRAMS
Figure 1. Cryocooler motor specific power interpolated to 60 K cooling and 300 K reject.
Figure 3. Cryocooler total specific mass interpolated to 60 K cooling and 300 K reject.
Figure 2. Cryocooler total specific power interpolated to 60 K cooling and 300 K reject.
Figure 4. Cryocooler motor specific power interpolated to 35 K cooling and 300 K reject.
Figure 5. Cryocooler total specific power
Figure 6. Cryocooler total specific mass
interpolated to 35 K cooling and 300 K reject.
interpolated to 35 K cooling and 300 K reject.
OVERVIEW OF PERFORMANCE OF SPACE CRYOCOOLERS
Figure 7. Cryocooler motor specific power interpolated to 100 K cooling and 300 K reject.
17
Figure 8. Cryocooler total specific power interpolated to 100 K cooling and 300 K reject.
Figure 9. Cryocooler total specific mass interpolated to 100 K cooling and 300 K reject.
capacity. The Figures show the expected effect of efficiency increasing as the capacity increases and indicate the general trend at 60 K of the higher efficiency at low to medium loads of Stirling (especially the Raytheon PSC) and Pulse Tube cryocoolers compared with the Reverse Brayton. Figure 3 is a plot of cryocooler specific mass (SM or mass divided by cooling capacity) for the total unit (including electronics) for cooling at 60 K as a function of cooling capacity. The trend of increased mass efficiency with increased capacity is indicated. The light weight nature of the Brayton cryocoolers and the TRW IMAS design is also apparent. Figures 4 to 6 and 7 to 9 are similar plots of the specific power and mass as a function of cooling capacity for cooling at 35 K and 100 K, respectively. Of significance at 35 K is the potential (based only on component tests) improved relative performance of the recuperative Brayton cycle Creare MMRBC at lower temperatures. At 100 K, if successful, the new start AFRL High Efficiency program with goals of less than 10 W/W motor power and less than 1 kg/W total mass for 10 W of cooling should make significant advances over the state of the art. SUMMARY
An overview is presented of the status and performance for a wide range of long life cryocoolers being developed for space applications ranging in temperature from 10 K to at least
100 K. This survey identifies more than 30 coolers covering a variety of thermodynamic cycles and cooler types with capacities from a few milliwatts to over 10 W and includes single and multi-stage designs. The survey indicates that more than a dozen coolers are at or near flight model status and are undergoing flight qualification to be available for space applications before the turn of the century. Performance comparisons were made using plots of specific power and
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GOVERNMENT CRYOCOOLER DEVELOPMENT AND TEST PROGRAMS
specific mass as a function of capacity at temperatures of 35, 60, and 100 K. The comparisons show the relatively higher efficiencies of Stirling and Pulse Tube cycles near 60 K, the increased efficiency of all units with increasing capacity, and the potential increase in both power and mass efficiency for the new start AFRL High Efficiency program. ACKNOWLEDGMENT
The authors wish to acknowledge the support of personnel at the Air Force Research Lab and the Ballistic Missile Defense Organization. Special thanks are given to J. P. Reilly and Lt. B. J. Tomlinson of AFRL, S. Bard, B. Bowman, R. Ross, Jr., and L. Wade of the Jet Propulsion Lab, R. Fernandez, W. Gully, W. Kiehl, R. Levenduski, and R. Reinker of Ball, W. Swift of Creare, D. Gilman and K. Price of Raytheon, T. Nast and I. Spradley of Lockheed Martin, B. G. Jones of Matra Marconi Space, and C. K. Chan and M. Tward of TRW for providing a large portion of the cryocooler data presented here. REFERENCES 1. Gully, W., Personal Communication, Ball Aerospace and Technology, Boulder, Colorado (5 April 1998). 2. Horsley, W. J., “Test Results for the Ball Single-Stage Advanced-Flight Prototype Cryocooler,”
Cryocoolers 9, R. G. Ross, Jr., Ed., Plenum Press, New York (1997), pp. 55-58. 3. Horsley, W. J., “Test Results for Single-Stage Ball Flight Prototype Cooler,” Cryocoolers 8, R. G. Ross, Jr., Ed., Plenum Press, New York (1995), pp. 23-33. 4. Carrington, H., et al., “Multistage Coolers for Space Applications,” Cryocoolers 8, R. G. Ross, Jr., Ed., Plenum Press, New York (1995), pp. 93-102. 5. Berry D., “System Test Performance for the Ball Two-Stage Stirling Cycle Cryocooler,” Cryocoolers 9, R. G. Ross, Jr., Ed., Plenum Press, New York (1997), pp. 69-77. 6. Levenduski, R. and R. Scarlotti, “Joule-Thomson Cryocooler Development at Ball Aerospace,” Cryocoolers 9, R. G. Ross, Jr., Ed., Plenum Press, New York (1997), pp. 493-508. 7. Levenduski, R., et al., “Hybrid 10 K cooler for Space Applications,” to be presented at ICCC #10, Monterey, California (26-28 May 1998), paper #37. 8. Fernandez, R., “Flight Demonstration of the Ball Aerospace Joule-Thomson Cryocoolers,” to be presented at ICCC #10, Monterey, California (26-28 May 1998), paper #57. 9. Stacy, W. D., “Development and Demonstration of the Creare 65 K Standard Spacecraft Cooler,” Cryocoolers 9, R. G. Ross, Jr., Ed., Plenum Press New York, (1997), pp. 45-53. 10. Tomlinson, 1st. Lt. B. J., Personal Communication, Air Force Research Laboratory, Kirtland Air Force Base, Albuquerque, New Mexico (January 1998). 11. Levenduski, R., “Joule-Thomson Cryocooler Development at Ball Aerospace,” Cryocoolers 8, R. G. Ross, Jr., Ed., Plenum Press, New York (1995), pp. 543-558. 12. Swift, W. L., “Single-Stage Reverse Brayton Cryocooler: Performance of the Engineering Model,”
Cryocoolers 8, R. G. Ross, Jr., Ed., Plenum Press, New York (1995), pp. 499-506. 13. Nellis, G., et al., “Design and Test of a Low Capacity Reverse Brayton Cryocooler for Refrigeration at 35 K,” to be presented at ICCC #10, Monterey, California (26-28 May 1998), paper #36. 14. Dolan, F., et al., “Reverse Brayton Cooler for NICMOS,” to be presented at ICCC #10, Monterey, California (26-28 May 1998), paper #54. 15. Bard, S., “Flight Demonstration of a 10 K Sorption Cryocooler,” Cryocoolers 9, R. G. Ross, Jr., Ed., Plenum Press, New York (1997), pp. 567-576. 16. Wade, L., “Continuous and Periodic Sorption Cryocoolers for 10 K and Below,” Cryocoolers 9, R. G. Ross, Jr., Ed., Plenum Press, New York (1997), pp. 577-586. 17. Bowman, R., Personal Communication, Jet Propulsion Laboratory, Pasadena, California (February 1998). 18. Chan, C. K., et al., “Performance of the AIRS Pulse-Tube Engineering Model Cryocooler,” Cryocoolers 9, R. G. Ross, Jr., Ed., Plenum Press, New York (1997), pp. 195-212. 19. Ross, Jr., R. G., “AIRS Cryocooler Systems Design and Development,” Cryocoolers 9, R. G. Ross, Jr., Ed., Plenum Press, New York (1997), pp. 885-904.
OVERVIEW OF PERFORMANCE OF SPACE CRYOCOOLERS
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20. Chan, C. K., Personal Communication, TRW, Redondo Beach, California (5 April 1997). 21. Burt, W. W., “New Mid-Size High Efficiency Pulse-Tube Coolers,” Cryocoolers 9, R. G. Ross, Jr., Ed., Plenum Press, New York (1997), pp. 173-182. 22. Johnson, D. L., “Performance Characterization of the TRW 3503 and 6020 Pulse-Tube Coolers,” Cryocoolers 9, R. G. Ross, Jr., Ed., Plenum Press, New York (1997), pp. 183-193. 23. Tward, E., “Miniature Long-Life Space Qualified Pulse-Tube and Stirling Cryocoolers,” Cryocoolers 8, R. G. Ross, Jr., Ed., Plenum Press, New York (1995), pp. 329-336. 24. Burt, W. W., “Demonstration of a High Performance 35 K Pulse-Tube,” Cryocoolers 8, R. G. Ross,
Jr., Ed., Plenum Press, New York (1995), pp. 313-319. 25. Nast, T., Personal Communication, Lockheed Martin Missiles and Space, Organization H1-21, Palo Alto, California (15 April 1998). 26. Jones, B. G., Personal Communication, Matra Marconi Space, Filton, Bristol, England (20 April 1998).
27. Davies, S. W., “Product Specification, 50-80 K Mechanical Cooler,” Doc. Ref. no. PSP/MCC/A0426/MMB, (27 October 1997). 28. Scull, S. R., “Design and Development of a 20 K Stirling Cooler for FIRST,” Cryocoolers 9, R. G. Ross, Jr., Ed., Plenum Press, New York (1997), pp. 89-96. 29. Jones, B. G., “Qualification of a 4 K Mechanical Cooler for Space Applications,” Cryocoolers 8, R. G. Ross, Jr., Ed., Plenum Press, New York (1995), pp. 525-536. 30. Bradshaw, T. M., “Life Test and Performance Testing of a 4 K Cooler for Space Qualifications,” to
be presented at ICCC #10, Monterey, California (26-28 May 1998), paper #73. 31. Johnson, D. L., “EMI Performance of the AIRS Cooler and Electronics,” to be presented at ICCC
#10, Monterey, California (26-28 May 1998), paper #77. 32. Chan, C. K., “IMAS Pulse-Tube Cooler Development and Testing,” to be presented ICCC #10, Monterey, California (26-28 May 1998), paper #104. 33. Spradley, I. and T. Nast, Personal Communication, Lockheed Martin Missiles and Space, Palo Alto, California (10 April 1998).
34. Longsworth, R. C., “Periodic 10 K J-T Cryostat for Flight Demonstration,” to be presented at ICCC
#10, Monterey, California (26-28 May 1998), paper #8. 35. Levy, A. R., “Performance of a 25 Kelvin Sorption Cryocooler Designed for the UCSB Long
Duration Balloon Cosmic Microwave Background Radiation Experiment,” to be presented ICCC #10, Monterey, California (26-28 May 1998), paper #7.
36. Gilman, D., Personal Communication, Raytheon Systems Co., El Segundo, California (21 April 1998). 37. Price, K., Personal Communication, Raytheon Systems Co., El Segundo, California (21 April 1998).
38. Roberts, T. and J. Bruning, “Hughes Aircraft Co. Standard Spacecraft Cooler Acceptance Testing and Performance Mapping Results,” Phillips Laboratory Report #PL-TR-96-1163 (November 1996). 39. Tward, E., Personal Communication, TRW, Redondo Beach, California (21 April 1998).
40. Swift, W., Personal Communication, Creare, Inc., Hanover, New Hampshire (20 April 1998). 41. Nast, T., “Design, Performance, and Testing of the Lockheed Developed Mechanical Cryocooler,” Cryocoolers 8, R. G. Ross, Jr., Ed., Plenum Press, New York (1995), pp. 55-67. 42. Spradley, I. E. and W. G. Foster, “Space Cryogenic Refrigerator System (SCRS) Thermal Performance Test Results,” Cryocoolers 8, R. G. Ross, Jr., Ed., Plenum Press, New York (1995), pp. 13-22. 43. Jones, B. G., et al., “The Batch Manufacturing of Stirling Cycle Coolers for Space Applications Including Test Qualification and Integration Issues,” Cryocoolers 9, R. G. Ross, Jr., Ed., Plenum Press, New York (1997), pp. 59-68.
44. Petree, D. and P. V. Mason, “Infrared Astronomical Satellite (IRAS) Superfluid Helium Tank Temperature Control,” Adv. Cryo Engr, V. 29 (1983), pp. 661-667. 45. Volz, S. M., et al., “Cryogenic On-Orbit Performance of the NASA Cosmic Background Explorer
(COBE),” SPIE Conference, San Diego, California (July 1990). 46. Bejan, A., “Entropy Generation Minimization,” CRC Press, New York (1996).
Air Force Research Laboratory Cryocooler Technology Development Thomas M. Davis, John Reilly, and First Lt. B. J. Tomlinson, USAF
Air Force Research Laboratory Kirtland AFB, NM 87117-5776
ABSTRACT
This paper presents an overview of the cryogenic refrigerator and cryogenic integration programs in development and characterization under the Cryogenic Technology Group, Space Vehicles Directorate of the Air Force Research Laboratory (AFRL). The vision statement for the
group is to support the space community as the center of excellence for developing and transitioning space cryogenic thermal management technologies. The primary customers for the AFRL cryogenic technology development programs are Ballistic Missile Defense Organization (BMDO), the Air Force Space Based Infrared System (SBIRS) Low program office, and other
DoD space surveillance programs. This paper will describe the range of Stirling, pulse tube, reverse Brayton, Joule-Thomson cycle cryocoolers, and sorption cryocoolers currently under development to meet current and future Air Force and DoD requirements. The AFRL customer single stage cooling requirements
at 10 K, 35 K, 60 K, 150 K, and multi-stage cooling requirements at 35/60 K are addressed. In order to meet these various requirements, the AFRL Cryogenic Technology Group is pursuing
various strategic cryocooler and cryogenic integration options. The Air Force Research Laboratory is also developing several advanced cryogenic integra-
tion technologies that will result in the reduction in current cryogenic system integration penalties and design time. These technologies include the continued development of the Cryogenic Systems Integration Model (CSIM), 60 K and 100 K thermal storage units and heat pipes, cryogenic straps, thermal switches, and development of an Integrated Lightweight Cryogenic Bus (CRYOBUS). INTRODUCTION The use of long-life, active cryocoolers provides a significant improvement to DoD space surveillance and missile tracking missions. Cooling of infrared sensors in space at temperatures of 80 K and below has mainly been accomplished using stored cryogens. These expendable
cryogenic systems require the launch of heavy and complex dewars, which at best have a one or two year life. Cooled detectors allow vast improvements in identification and discrimination
Cryocoolers 10, edited by R. G. Ross, Jr. Kluwer Academic/Plenum Publishers, 1999
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GOVERNMENT CRYOCOOLER DEVELOPMENT AND TEST PROGRAMS
capability with a minimum of sensor apecture growth. Smaller apecture produces cheaper, lighter sensors; much easier to host in a space-based environment. Other space missions such as communications, remote sensing, and weather monitoring can benefit from subsystems using cryogenic technology including super conducting electronics, high data rate signal processors, and high speed/low power analog to digital converters. The objective of the Air Force Research Laboratory cryocooler effort is to develop and demonstrate space qualifiable cryogenic technologies required to meet future requirements for
Air Force and Department of Defense (DoD) missions. Pursuant of this objective, the Air Force Research Laboratory characterizes and evaluates the performance of development hardware, pursue advanced concepts for future spacecraft missions, and work to enhance cryocooler to spacecraft integration. Cryocooler development is considered a Military Critical Technology and is also tracked under the Defense Technology Objectives initiative. Performance improvement objectives have been established for life, power, mass, and vibration. Progress is reviewed annually at DoD level. Collaboration with other government development activities and private industry has been a major strength of the AFRL program. This has resulted in leveraging of scarce development funding and more rapid transition of cryocooler technology to the space community. Current AFRL cryocooler development programs include Stirling, pulse tube, reverse Brayton, Joule-Thomson, and sorption machines that produce cooling in the 10 K through 150 K temperature range. Compared with state-of-the-art dewars and cryogenic radiators, mechani-
cal cryocoolers offer space systems significant weight savings, performance improvements, and long life potential (greater than 5 years). After a specific cryocooler is developed, the unit is subjected to acceptance, characterization, and endurance tests based on customer and Air Force requirements. Acceptance tests are performed to determine if the unit meets contractual specifications. Characterization is then performed to determine the operating performance envelope of the cryocooler in nominal and off nominal conditions. Endurance tests are used to demonstrate operational hours, and identify and characterize long term, life limiting failure mechanisms and long term performance degradation. Components that significantly improve the efficiency, extend life, reduce mass, or limit induced vibration are developed and transitioned into next generation cryocooler designs. These technologies are typically developed under contractor sponsored In-house Research and Development (IRAD) efforts or from the Small Business Innovative Research (SBIR) program. Advanced, high pay-off cryogenic integration technologies are developed that reduce risk, complexity, mass, and volume of the cryogenic system. Utilization of improved integration technologies ensures an optimum cryogenic thermal management system is developed that limit or eliminates operational constraints imposed on the spacecraft platform. The warfighter payoffs for these innovations enable long-life, space based surveillance to support tactical and strategic missions. Additionally, cryocooler technology has been identified as enabling technology for national missile defense systems. CRYOGENIC TECHNOLOGY DEVELOPMENT REQUIREMENTS
Spacecraft cryocooler requirements differ quite dramatically from tactical cryocooler applications due to the imposition of a requirement for long life and continuous operation. Operational lifetimes for strategic spacecraft are usually in the 5-10 year range. With most Air Force spacecraft operating in orbits that preclude periodic maintenance, utilization of highly reliable components is critical. Current Air Force operational requirements result in the need for single and multi-stage cooling for short, medium and long wavelength infrared sensors. In some instances, DoD spacecraft are limited in terms of allowable payload power, mass, or volume. This results in an additional requirement for highly efficient thermodynamic machines requiring
minimal input power, lightweight structure, and packaged for as small a volume as possible. Additional constraints such as vibration suppression or minimization are required to mitigate or eliminate induced vibration from affecting sensor operation.
AFRL CRYOCOOLER TECHNOLOGY DEVELOPMENT
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In order to meet customer requirements, cryocooler development has explored different thermodynamic cycles and variations of thermodynamic cycles in order to develop a base of cryocoolers for specific and general applications. Each thermodynamic cycle or variation has its own inherent strengths and weaknesses. Additionally, the maturity of the technology for these cycles varies. The bottom line is that differing customer requirements dictate different thermodynamic cooling cycles to satisfy mission requirements. The following section, “Cryocooler Development Programs”, will highlight the different thermodynamic cycles and the development programs pursuing the different technology options. The requirements for cryogenic integration drive the development of components that improve the cryogenic system by reducing large system penalties and analysis errors. These technologies include thermal straps, thermal storage units for load leveling, thermal switches, cryogenic capillary pumped loops, diode heat pipes, thermal transport devices, thermal interface technologies, and thermal analysis tools. The sources of the technical requirements are USAF, BMDO, and DoD needs. Typical cryocooler programs include the development of flight quality electronics. Recent experience by AFRL and other agencies in the development of flight electronics have resulted in inflexible, ineffective, expensive, and outdated designs. In order to achieve a low cost and effective solution to this problem for users, AFRL has initiated an in-house effort to develop an approach for a common flight electronics design potentially applicable for use with Stirling and pulse tube cryocoolers. AFRL has assembled an in-house team to review the multiple technical issues and develop a common design for vibration control, temperature control, and cryocooler health monitoring. The initial phase of this effort is to demonstrate the feasibility of this concept by developing and evaluating breadboard hardware. Phase I of this effort was completed in
March 1997 and the final report is now being finalized. Phase II will focus on identifying industry requirements and the development of brassboard hardware for a proof of concept demonstration. The planned completion date planned for this effort is August 1998.
The technology issues involved in cooling of the gimbaled optics for the engineering and manufacturing design (EMD) phase of the SBIRS Low system is a high priority for both BMDO and the SBIRS Low program office. AFRL is investigating cryogenic solutions to the two axis gimbal problems to provide technology solutions to minimize the combination of system penalties for mass, volume, power, and flexibility to meet identified mission goals and requirements. CRYOCOOLER DEVELOPMENT PROGRAMS
AFRL cryocooler development includes various types of active cooling cryocoolers. In the following sections, Stirling (also pulse tube) cycle, reverse Brayton cycle, Joule-Thomson, and sorption cryocooler development programs at AFRL are detailed. Stirling Cycle Cryocoolers Background— The Stirling cycle cryocooler is the most mature design of the cryocoolers developed for the Air Force Research Laboratory. These devices operate under the Stirling thermodynamic cycle with a mechanical compressor and expander combined with a regenerator. The advantages inherent in this type of cooler are lowest combined volume and mass, less com-
pressor swept volume than pulse tubes, and the most power efficient cryocooling. The input power for this type of cooler is typically less than other coolers for temperatures greater than 20 K or loads less than 5 W. Some of the disadvantages for this cooler are the sub-mil ( ) clearance seals, concern with the cross axis vibration, and concerns with the reliability of the moving expander/regenerator. Raytheon (formerly Hughes Aircraft Company) PSC Cryocooler— The 60 K Protoflight Spacecraft Cryocooler (PSC) is under development by Hughes and funded by BMDO and SBIRS Low
(Figure 1). This design is the maturest in a series of cryocoolers developed by US government and US company sponsored IRAD resources. Specific objectives of this program are to develop
24
GOVERNMENT CRYOCOOLER DEVELOPMENT AND TEST PROGRAMS
a unit requiring less than 100 Watts of input power with 2 Watts of cooling, life in excess of 7 years, and total system mass less than 33 kg. The design of this cooler has the potential to meet the requirements for the MWIR and LWIR tracking sensor needs for the SBIRS Low EMD system. Design features include incorporation of a linear tangential flexure into the compressor. This innovation allows for improvement in radial stiffness and is expected to result in improved cryocooler reliability and performance. Titanium has been utilized within the housing and piston assembly to increase unit efficiency and result in reduced system mass. Hughes will also demonstrate the adequacy of piston alignment techniques necessary for ensuring adequate submil clearance of moving parts. The first protoflight unit was delivered in December 1997, and subsequently subjected to acceptance characterization tests and endurance evaluation at JPL and the Air Force Research Laboratory. Ball Aerospace Cryocooler— AFRL is developing a multi-stage Stirling cryocooler with Ball Aerospace. The objective of this effort is to simultaneously provide 0.4 Watt of cooling at 35 K and 0.6 Watt of cooling at 60 K. The Ball 35/60 K program is a candidate to meet Boeing’s Low Altitude Demonstration System ground demonstration tracking sensor imager requirements for simultaneous cooling at 35 K and 60 K (Figures 2 and 3). In addition, the cooler is a technology option for the SBIRS Low EMD tracking sensor. As a result, the cryocooler must be able to achieve 0.1 Nrms vibration requirement and a greater than 7 year lifetime. The Ball effort is a BMDO/Air Force funded program managed by NASA/GSFC. The Air Force program is building upon technology previously developed for a NASA 30K two-stage cryocooler3. Ball is modifying the NASA compressor design in order to help double the cooling capacity. To satisfy requirements, the displacer was redesigned from a two stage to a three-stage design. The cold finger features a fixed regenerator, which improves life, efficiency, and reduces induced vibration. Additional performance improvement is realized by the incorporation of precision piston alignment techniques that eliminate piston/cylinder contact in the cryocooler. A protoflight cryocooler with associated flight electronics is planned for a delivery to the Air Force Research Laboratory in March 1998 where it will undergo characterization and endurance evaluation.
Figure 1. Hughes 60K Protoflight Spacecraft Cryocooler.
Figure 2. Ball Three-Stage Cryocooler.
Figure 3. Ball Three-Stage Electronics.
AFRL CRYOCOOLER TECHNOLOGY DEVELOPMENT
Figure 4. MMS 20 K Cryocooler.
25
Figure 5. Defense Research Agency Cryocooler.
Matra Marconi Space (MMS) Cryocooler— Sponsored by BMDO, this program is intended to meet near-term requirements for 10 K Very Long Wave Infrared (VLWIR) sensor technology. The cooler under development is designed to meet a requirement of 45 mW @ 10 K. The design will utilize four of MMS’s standard compressors for their 20 K cryocooler with a multistage expander. The MMS 20 K cryocooler can be seen in Figure 4. The primary technology challenge is to identify regenerator materials capable of achieving the required cooling at 10 K. Simulation, modeling, and testing of potential regenerator designs were completed at Rutherford Appleton Laboratory in February 1998. The engineering design model cryocooler is scheduled for delivery to AFRL in May 1999. Defense Research Agency (DRA) Cryocooler— DRA, under BMDO sponsorship and AFRL technical management, has developed and delivered a cryocooler designed to meet the STRV mission requirements of 0.25 W @ 65 K. Figure 5 shows the DRA cryocooler without the compressor cover. Two versions of the cryocooler were manufactured. One version has a linear flexure spring design and the other version has a flexure leaf design. Both units have contacting, wearing seals and are predicted to have lifetimes on the order of 20,000 hours.
PULSE TUBE CRYOCOOLERS
Background— The pulse tube cryocooler is a Stirling thermodynamic cycle. However, this design approach has a compressor, regenerator, a passive pulse tube, and orifice to a gas reservoir. The Stirling expander has been replaced by a combination of the passive pulse tube and orifice to reservoir. The advantages associated with this approach are no cryogenic sub-mil clearance seals; higher reliability due to reduced number of moving parts, and reduced electromagnetic interference and vibration at the cold block. A disadvantage to this design is a slightly lower Carnot efficiency than the Stirling cooler (due to the irreversibilities in the pulse tube expansion) and integration difficulties due to the configuration of the cooler and the location of the cold block. TRW 35K Pulse Tube Cryocooler Program— AFRL, BMDO, and SBIRS Low program office sponsored the development of several single-stage pulse tube cryocoolers under the 35 K Pulse Tube contract with TRW. The goal of this effort was to improve pulse tube performance and reliability to match the maturity of Stirling cryocoolers. TRW delivered three engineering models under this effort (two units designed for 35 K and one unit designed for 60 K); all built as protoflight units4. The 35 K Pulse Tube program also focused on minimizing weight (< 18.1 kg) and reducing input power (<100 W) for the second (35 K) and third (60 K) pulse tubes. To accomplish this, TRW improved compressor performance and added a fixed regenerator. The compressor size was reduced from 20 cc to 10 cc for the 35 and 60 K cryocoolers. The compact design and smaller 10 cc compressor results in a mass of only 12.1 kg for both the 35 and 60 K
26
GOVERNMENT CRYOCOOLER DEVELOPMENT AND TEST PROGRAMS
pulse tubes. Once endurance testing is completed at AFRL, these three classes of cryocoolers
will be fully characterized in terms of both short and long term performance. The TRW 3585 delivers 0.85 W of cooling at 35K and requires 200 Watts of input power while rejecting heat at 290 K. This pulse tube has completed acceptance/characterization testing and is currently in endurance evaluation AFRL. A second pulse tube design (TRW 3503) delivers 0.3 W of cooling at 35 K while requiring only 83 Watts of input power and rejecting heat at 300 K. The 3503 machine completed acceptance/characterization testing at JPL and AFRL, and is also under endurance evaluation at AFRL. The final pulse tube cryocooler is the TRW 6020. This unit has a 10cc compressor and is capable of delivering 2 W @ 60 K. In addition to completing the rigorous acceptance and characterization testing at AFRL, this cooler was used to support the integration and test of the Multispectral Thermal Imager (MTI) mock focal plane sensor assembly. This test was valuable because of the design similarity of TRW 6020 to the
MTI cryocooler currently being fabricated (detailed below). This pulse tube is currently in endurance evaluation. TRW/Multispectral Thermal Imager (MTI) Pulse Tube— A variant of the TRW 6020 and being
developed for the DOE/Sandia National Laboratory Multispectral Thermal Imager (MTI) flight experiment, this cooler delivers 2.25 Watt @ 60 K (Figure 6). The flight electronics are adapted from the cryocooler electronics for the NASA/JPL AIRS instrument2. Delivery of the cryocooler and associated flight electronics is currently scheduled for April 1998. TRW 150K Mini Pulse Tube Cryocooler— Primary objectives for the 150 K Prototype Spacecraft Cryocooler program with TRW are to advance pulse tube technology available at 150 K to protoflight maturity. Under this contract, seven miniature pulse tube cryocoolers have been developed. These lightweight cryocoolers (2.0 Kg) provide 1.5 W @ 110 K, require only ~19 Watts input power, and reject heat at 300 K (Figure 7). Although the original objective of the program was to develop cryocoolers for the 150 K temperature range, the SBIRS Low program office has utilized this design for cooling of the gimbal optics on the Flight Demonstration System flight experiment. Two of these coolers are being used to provide 3 Watts of cooling @ 110 K. Designed and fabricated to protoflight levels, there are several space flight experiments, which plan
to incorporate the mini-pulse tubes to meet cooling requirements. One unit was provided for the NASA/SSTI (Lewis) experiment. Two new mini-pulse tubes have been developed and delivered
for an upcoming DOE/SNL CX mission. Lockheed Martin Coaxial Pulse Tube— The Co-axial Pulse Tube (CAPT) cryocooler was developed by Lockheed-Martin/Palo Alto under an AF PRDA. The major program objectives were to develop a unit providing 0.4 Watt at 35 K and 0.6 Watt at 60 K for an input power less than 100 Watts, life in excess of 5 years, and demonstrate performance that matches or exceeds similar Stirling cycle units. A unique aspect was the proposed use of several advanced
Figure 6. TRW 60 K Pulse Tube Cryocooler.
Figure 7. 150 K protoflight spacecraft cryocooler.
AFRL CRYOCOOLER TECHNOLOGY DEVELOPMENT
27
technology components developed under IRAD and SBIR programs. These innovations included application of co-axial pulse tube configuration that enhances integration, incorporation of tangential flexures for reduction in vibration and improved life, incorporation of a flexure cartridge that eases assembly time and increases lifetime, and incorporation of the etched foil regenerator that results in improved system efficiency. The basic phase of this program was
completed in September 1997. Due to funding considerations and technical issues, the option phase of this program was not exercised and the engineering design model cryocooler will be evaluated in the AFRL characterization facility. REVERSE BRAYTON CRYOCOOLERS
Background— Reverse Brayton cryocoolers employ a reversed power cycle that uses turbines for compression and expansion of the working fluid. Advancement in the technology uses micro-machined components for the turbines. The advantages to this approach are extremely low induced vibration, remote location of cold head from ambient temperature compressor, and capacity to lift high loads (> 5 W). Disadvantages range from relatively low hardware maturity to lower (relative) efficiency for small (< 5 W) heat loads and the complex micro-machining required. CREARE Single-Stage Reverse Brayton— Reverse Brayton cycle technology offers the potential for dramatic reduction in induced vibration over Stirling or pulse tube cryocoolers due to the operation of the turbine at high speeds that results in higher frequency vibration modes. NASA’s Near Infrared Camera and Multi-Object Spectrometer (NICMOS) program builds on the successes of an Air Force funded 65 K Engineering Demonstration Model (EDM) that is currently undergoing endurance tests at the Air Force Research Laboratory1. The EDM has demonstrated a robust design that has exceeded operational design specifications, even when using contaminated working fluid (Figure 8). Design of the flight-qualified unit utilizes the engineering unit heat exchanger and develops a miniaturized turboexpander and compressor. Miniaturization of these components results in a technical challenge to reduce component size without significantly decreasing efficiency because parasitic losses are a larger percentage of the overall system when miniaturized. The NICMOS cryocooler is a joint effort with the NASA Goddard Space Flight Center and will be flight demonstrated on a Shuttle mission in October 1998. Subsequently, the cooler is to be installed on the Hubble Telescope during a 1999 servicing mission. CREARE Miniature, Multi-Stage Reverse Brayton Cryocooler— Under BMDO sponsorship, the
Air Force Research Laboratory is developing the Miniature, Multi-stage Reverse Brayton Cryocooler (MMRBC) program with Creare, Inc. (Figure 9). The objective of this program is to develop dual load cooling of 0.4 Watt at 35 K and 0.6 Watt at 60 K. Total input power for either cooling requirement is not to exceed 125 Watts and
Figure 8. Creare 60 K reverse Brayton .
Figure 9. Miniature multi-stage reverse Brayton .
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GOVERNMENT CRYOCOOLER DEVELOPMENT AND TEST PROGRAMS
must demonstrate an operating life in excess of seven years. A primary potential application is for the SBIRS Low EMD tracking sensors. Exported vibration of this unit is negligible due to the operation of the turbine at such high speeds that results in high operating frequencies. In order to achieve these requirements, Creare plans to implement a cryogenic turboalternator that will reduce parasitic losses and result in increased efficiency of the unit. Current Reverse Brayton cycle machines rely on the working fluid and external turboexpander braking loads to control the speed of the turbine. Addition of the turboalternator will allow the turbine speed to be continuously optimized by the alternator. Advanced materials will be used in the design of the unit, which should result in additional motor efficiency improvements. The turbo alternator is projected to have 50% less parasitics and achieve 70% more efficiency compared to the turbo expander design. A radial flow heat exchanger is also being incorporated into the design which will reduce volume by 75% and be 70% lighter than the existing 65 K heat exchanger. Final design of the turbo alternator and heat exchanger is currently being completed. A protoflight unit is planned for delivery to the Air Force in early summer 1999. JOULE-THOMSON CRYOCOOLERS Background— These cryocoolers employ a Joule-Thomson (J-T), or isenthalpic expansion of the working gas to achieve cryogenic temperatures at the cold end heat exchanger. Actually, there are many cycles that employ the J-T expansion, such as the Linde-Hampson, Claude, and Joule-Brayton cycles. The advantages in this technology are inherent thermal capacity and load leveling, remote location of the cold end from the compressor, and negligible electromagnetic
interference or vibration in the cold end. Disadvantages include a very high-pressure ratio of compression or requiring another cryocooler for pre-cooling and the fact that J-T systems have inherent two-phase fluid management problems. Also, as a stand-alone cryocooler, this technology has a relatively low efficiency. Cryogenic On-Orbit Long Life Active Refrigerator (COOLLAR)—AFRL performed as technical advisors and were jointly responsible with NASA for the space demonstration of this technology for a DoD program office (Figure 10). Some of the objectives of the demonstration were the requirement for verification of the operation of the cooler in microgravity and to demonstrate temperature control a 65 K and 122 K. The experiment was successfully flown on STS-85 in August 1997.
SORPTION CRYOCOOLERS Background— Sorption cryocoolers utilize a heat driven hydride bed to provide the gas compression. The advantages and disadvantages to this technology are also similar to the J-T systems. However, additional advantages are that sorption compression has the potential for very long life, and produces negligible vibration. Principal disadvantages are the problems associated with the complicated set of valves for the multiple compressor beds and periodic versus continuous cooling and a relatively high input power requirement.
Figure 10. COOLLAR.
Figure 11. BETSCE .
AFRL CRYOCOOLER TECHNOLOGY DEVELOPMENT
29
Brilliant Eyes 10 K Sorption Cryocooler Experiment (BETSCE)— AFRL collaborated with NASA/JPL and USAF BE program office in the 10 K BETSCE program. This program was to demonstrate 10K cryocooler technology for space. The experiment was flown on the shuttle (STS-77). The basic configuration of the system was a Stirling cycle cryocooler (precooling to 60 K), a J-T expansion of the evolved hydrogen (cooling to 28 K), then sorption pump down to
10 K (generation of solid hydrogen ). This was to produce periodic cooling (30 minutes / 2 hours) of 0.1 W @ 10 K. However, the experiment was only able to achieve one successful cooldown. A faulty valve prevented the device from repeating the cooldown to 10 K (the cooler can be seen in Figure 11).
AFRL CRYOCOOLER CHARACTERIZATION FACILITY Objectives and Activities—A major concern in applying cryocooler technology for potential users is their unproven reliability. AFRL has several initiatives to investigate issues to meet the program requirements for reliability of > 7 years life. The AFRL characterization facility and Aerospace Corporation are also playing a critical role in improving reliability confidence of cryocoolers. Leveraging existing capabilities at JPL, GSFC, other government laboratories, and private industry, the AFRL facility supports performance verification and reliability evaluation of developed hardware in a simulated operational environment. The principal focus of the facility is to improve reliability confidence in cryocooler hardware and provide better understand of life-limiting factors. The end result of this process is to influence improvements in designs and reliability of future cryogenic systems. Cryocooler characterization and endurance testing is being accomplished on the Creare Turbo, TRW pulse tubes, and Hughes SSC units. Both the Hughes 60 K PSC and Ball Aerospace 35/60 K will begin a three-year endurance test simulating space operational conditions in early 1998. Several actions have been taken to make the facility more relevant and responsive to user needs. A major effort is being made to improve the feedback of evaluation data to the space community. A near-term solution recently implemented was to incorporate “load line” data into the Aerospace Corporation developed CSIM. Users of this model will now have near real time information from AFRL performance evaluations. Making cryocooler test data and other relevant information available on the AFRL Web page is also being evaluated. Test plans have expanded to include a broader range of temperature ranges during performance evaluation. AFRL recently established a working group to more formally address the range of cryocooler reliability issues. The group’s charter is to identify and implement actions to improve reliability confidence of spacecraft coolers. A collaborative effort with the Ukrainian Institute of Low Temperature Physics is assessing the possibility of applying accelerated lifetime testing methodology to selected AFRL cryocoolers. Aerospace Corporation has made significant progress in critical component reliability through the development of flexure bearings to satisfy low tolerance cryocooler design requirements. The Aerospace developed tangential spring has been baselined in the HAC 60 K PSC and LMMS CAPT cryocooler. Life testing has shown to date a reliability of 108 cycles. In addition to improving cryocooler reliability, The results of these activities provide will ultimately hasten the transition of cryocooler technology to DoD program offices and development contractors. CRYOGENIC INTEGRATION TECHNOLOGY Background—The objective of this activity is to develop components that improve the cryogenic system by reducing large system penalties and analysis errors. Technologies under development include a Cryogenic Systems Integration Model (CSIM), 60 and 100 K thermal storage units and heat transport devices, thermal switches, and flexible cryogenic joints and straps. The
technical challenges to be addressed during component technology development includes management of a two phase liquid in a zero-G environment, poor capillary wicking of a cryogenic heat transport device, poor phase change material thermal conductivity, and develop a thermal
30
GOVERNMENT CRYOCOOLER DEVELOPMENT AND TEST PROGRAMS
Figure 12. JPL Thermal switch
Figure 13. Swales / AFRL Thermal switch
switch that maintains low on resistance and high off resistance. Cryogenic Systems Integration Model— The Cryogenic Systems Integration Model (CSIM) is
an interactive, user friendly; PC Windows based software tool5. The objective of this effort is to develop an analysis tool that is capable of performing the integration of mechanical cryogenic systems and insure that an analysis tool is available to industry and the government. This package includes design algorithms for all cryogenic integration components and includes algorithms for heat load parasitics and heat rejection. The tool is also capable of accessing a database that includes materials, cryocoolers, and other components. Presently, CSIM is being distributed to industry and government personnel for evaluations against actual designs. Integrated Lightweight Cryogenic Thermal Bus (CRYOBUS)— Under an Air Force Research Laboratory contract, Swales Aerospace is developing the Integrated Lightweight Cryogenic Thermal Bus. The objective of this contract is to develop an integrated lightweight system and components for linking a mechanical cryocooler with a cooled component (principally an infrared sensor). The goal is to develop a class standardized interface between the cryocooler system and the sensor and to standardize the integration into the spacecraft. Phase I of this program has been completed and Swales has begun Phase II of this effort. At the conclusion of this effort, prototype systems, incorporating several advanced integration components, will be available to AFRL for in-house evaluation or use in flight demonstrations. Included in the CRYOBUS efforts are several cryogenic integration technologies. These technologies include commercialization efforts of other government-developed technologies, such as the JPL gas gap thermal switch in Figure 12 and the redesigned Swales / AFRL thermal switch in Figure 13. Other technologies evaluated under the program include a Neon cryogenic CPL (Figure 14) and a cryogenic thermal storage unit (CRYOTSU) (Figure 15). The CRYOTSU thermal storage technology will also be expanded to include potential use at ~100 K as a possible technical solution for optics cooling. CRYOBUS is also addressing the cooling of the gimbaled optics for the engineering and manufacturing design (EMD) phase of the SBIRS Low system. This has been identified as a critical area for cryogenic integration by AFRL and the SBIRS Low program office. AFRL will leverage cryogenic solutions to the two axis gimbal problem from the CRYOBUS program to provide solutions to minimize the system penalties for mass, volume, power, and flexibility.
Figure 14. Neon Cryogenic Capillary Pumped Loop.
Figure 15. CRYOTSU (disassembled).
AFRL CRYOCOOLER TECHNOLOGY DEVELOPMENT
31
Flexible Diode Heat Pipe— The Air Force Research Laboratory has also awarded contracts through the Small Business and Innovative Research (SBIR) program to Jackson and Tull and Swales & Associates for the development of flexible diode heat pipes and thermal storage units. Flight Demonstrations—The CRYOTSU thermal storage device and the Flexible Diode Heat
Pipes (CRYOFD) are focused on the design, development, and fabrication of devices that operate at 60 and 120K. Currently, both of these devices are scheduled for zero-G evaluations onboard Space Shuttle missions. The flight of the CRYOFD diode heat pipes occurred in July 1997 (the original flight of these units was canceled on orbit and scheduled for reflight on STS94). The delivery of the CRYOTSU thermal storage unit, the cryogenic thermal switch, and a NASA cryogenic capillary pumped loop was completed in February 1998 with a planned flight experiment for November 1998. Flight characterization of these units is required to determine that ability to manage multi-phase fluid in a zero-G environment and be able to correlate flight data with ground predictions. FUTURE ACTIVITIES
While continued development is planned in all temperature ranges, a major focus is being placed on 10 K cryocoolers to meet the requirements for VLWIR focal plane applications using Si:As detectors. In addition to the current MMS program, a new initiative is planned for FY98 with performance objectives of 100 milli-Watts at 10 K, greater than seven years lifetime, less than 100 Watts per Watt of input power, and reductions in mass and weight. Successful devel-
opment will support multiple spacecraft missions requiring focal planes with detector wavelength sensitivity greater than and improved uniformity. Current plans are for the development of both an engineering development and protoflight units. A second new program will
address the requirements for the SBIRS Low EMD optics cooling with an objective of 10 Watts cooling at 95 K. Long-life, reduced mass and input power are key program goals. Additional emphasis is being placed on cryocooler component research aimed at improving the efficiency of compressors and regenerators, reducing the cost and complexity of electronics, and increasing
reliability. Finally, long-term objectives will pursue advanced concepts with few or no moving parts and improved efficiency. SUMMARY
The Air Force Research Laboratory is currently developing a broad spectrum of cryocoolers and integration components necessary for DoD space missions. The implementation of cryocooler and cryogenic technology development has resulted in the successful teaming of government agencies, industry, and international sources through multiple sponsored programs. Sev-
eral spacecraft cryocoolers have reached the point of maturity to be potentially transitioned to operational systems. Applications of technology developed by the Air Force Research Laboratory and BMDO are now being baselined on the SBIRS Low demonstration spacecraft, the MTI spacecraft, and the NRL sponsored HTSSE-II superconductivity flight experiment6. In-house evaluation has also demonstrated cryocoolers are capable of operating continuously for long pe-
riods of time and characterized operation beyond the design point of the technology. Additionally, the in-house cryocooler characterization is focusing in on improvement and diagnosis of cryocooler technology reliability. This focus and the characterization data being generated in the laboratory is being fed back to the cryocooler manufacturers in order to improve follow-on cryocooler development. Continued technology development is required to improve cryocooler
efficiency and demonstrate operational lifetime in excess of seven years for DoD space missions. Additionally, cryogenic integration technology development and demonstration is providing reduction in system penalties and reduced design time for cryocooler to sensor integration.
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ACKNOWLEDGEMENTS Personnel from the Air Force Research Laboratory and the Aerospace Corporation carried out the work described in this paper. Mr. Erwin Myrick of the Ballistic Missile Defense Organization (BMDO/TOS) has played a crucial role in supporting both the requirement and necessary funding for spacecraft cryogenic cooler development. The Air Force SBIRS Low Program Office and DOE/Sandia National Laboratories also sponsored technical effort. The Jet Propulsion Laboratory/ California Institute of Technology through an agreement with NASA provided additional technical effort for the characterization of pulse tube cryocoolers. Administration and technical oversight of the Ball Aerospace 35/60 K cryocooler and Creare reverse Brayton cryocooler contracts are provided by the NASA/Goddard Space Flight Center. REFERENCES 1. Swift, W.L., “Single Stage Reverse Brayton Cryocooler: Performance of the Engineering Model”, Cryocoolers 8, Plenum Press, New York (1995), pp. 499-506.
2. Ross, R.G., “JPL Cryocooler Development and Test Program Overview”, Cryocoolers 8, Plenum Press, New York (1995), pp. 173-184. 3. Burt, W.W, and Chan, C.K., “Demonstration of a High Performance 35 K Pulse Tube Cryocooler”,
Cryocoolers 8, Plenum Press, New York (1995), pp. 313-319. 4. Sparr, L., et al. “NASA/GSFC Cryocooler Test Program Results for FY94”, Cryocoolers 8, Plenum
Press, New York (1995), pp. 221-232. 5. Donabedian, M., et al. “Cryogenic Systems Integration Model (CSIM)”, Cryocoolers 8, Plenum
Press, New York (1995), pp. 695-707. 6. Kawecki, T., “High Temperature Superconducting Space Experiment II (HTSSE II) overview and Preliminary Cryocooler Integration Experience”, Cryocoolers 8, Plenum Press, New York (1995), pp. 893-900.
Endurance Evaluation of Long-Life Space Cryocoolers at AFRL—An Update
1st Lt. B.J. Tomlinson Air Force Research Laboratory Kirtland AFB, NM 87117 A. Gilbert and J. Bruning Nichols Research Corporation Albuquerque, NM 87106
ABSTRACT The Air Force Research Laboratory (AFRL), under the sponsorship of the Ballistic Missile Defense Organization (BMDO), developed several long-life cryocooler concepts designed to meet military space system cryogenic refrigeration performance and mission life specifications. These coolers include Stirling cycle, pulse tube, and reverse Brayton cycle machines developed to varying stages of technological maturity. AFRL’s Cryocooler Characterization Laboratory (CCL) and the Jet Propulsion Laboratory (JPL) have extensively assessed the thermodynamic performance capabilities of these newly emerging coolers. However, a complete demonstration of the technological maturity of these coolers must include long-term endurance characterization to help determine the long-life potential of these machines. Endurance evaluation attempts to demonstrate that the baseline performance of these machines is maintainable for the 5-10 year design operational life in an environment simulating the conditions of their use in space. These trials also surface any wear out, drift, and fatigue related failure mechanisms associated with cooler designs, allowing for adequate design modifications to improve the long-life potential of these machines. Endurance evaluation also helps to quantify performance shifts over long time periods that must be accounted for in cooler control schemes and spacecraft thermal designs. This paper presents endurance evaluation results for long-life space coolers currently under evaluation at AFRL. Data includes updates on the reverse Brayton cooler and the diaphragm flexure Stirling cooler, both of which have been undergoing endurance evaluation since 1995 and 1996, respectively. It also presents data on new endurance trails associated with several pulse tube and Stirling cycle machines more recently developed by AFRL and delivered to the CCL for characterization and evaluation.
Cryocoolers 10, edited by R. G. Ross, Jr. Kluwer Academic/Plenum Publishers, 1999
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GOVERNMENT CRYOCOOLER DEVELOPMENT AND TEST PROGRAMS
INTRODUCTION
There are various parameters used to monitor and evaluate the suitability of coolers for both
military spacecraft use as well as ground-based applications. Of these parameters, lifetime and reliability distinguishes between the capabilities of short-lived tactical units and long-life space cryocoolers. Assessing cryocooler life and reliability is the current primary focus of AFRL’s cryocooler characterization and evaluation activities. However, contributing factors to cooler unreliability are as important to understand as the inherent life and reliability of the machines themselves. Some of the potential contributors to cryocooler unreliability include wear out, drift, fatigue, material creep, gaseous contamination, paniculate/compound contamination and clogging, material and workmanship defects, inadequate machining process development, assembly errors, material thermal expansion mismatch, and long-term alignment instability. However, all potential failure modes and life limiting mechanisms are not identified, and those that have been are not completely understood. Also, various cryocooler concepts can have vastly different inherent failure modes and effects, as well as “graceful degradation” characteristics. Because of the importance of validating cryocoolers as long-life spacecraft components, a very
high premium is placed on gathering cooler-specific life, reliability, and long-term performance trend data. The CCL’s test facilities, procedures, and equipment/instrumentation all serve to enable long-duration endurance evaluation data collection in environmental conditions simulating actual space use. Once underway, endurance evaluation is normally run until cooler
failure, but not less than its design life if the test article continues to exhibit adequate
performance. Information gathered during endurance evaluation is then made available to cooler developers and system integrators to help them refine both the cooler designs and the spacecraft intending to use them.1 EXPERIMENT CONDITIONS
Four cryocoolers are currently undergoing endurance evaluation at the CCL. Table 1 lists these coolers and their nominal operating conditions during the experiments. These operating conditions are maintained at constant levels during the entire endurance trial phase except for periodic heat load/cold end temperature comparisons to benchmark conditions shown in Figure 1. Load/Temperature checks against the baseline performance are usually conducted at one to two-month intervals and are used to quantify time-dependent performance drift. It should be
noted that most of the coolers’ heat rejection temperatures are cycled above and below the nominal condition listed in Table 1. Cycling rejection temperatures during the experiment more closely emulates the transient thermal effects experienced under normal space environmental
Figure 1. Nominal Benchmark Performance .
AFRL ENDURANCE EVALUATION OF SPACE CRYOCOOLERS
35
usage comparison purposes, 300 K is defined as the nominal rejection temperature for all coolers. The rejection temperature cyclic range is varied depending on intended orbital transient profiles, cooler design sensitivities to coefficient of thermal expansion effects, and thermodynamic
performance
limitations
usually
experienced
at
elevated
levels
above
nominal.
ONGOING ENDURANCE EVALUATION PROFILES
Single Stage Turbo Cryocooler (SSTC)
Design Overview. The SSTC is a 5 Watt/65 K engineering model of a turbomachinerybased reverse Brayton cycle design that has been under various stages of development for over ten years. It utilizes advances in miniaturization of components aimed at improving refrigeration efficiency while reducing cooler size, weight, and volume. It was developed as an alternative long-life cryocooler concept to the Stirling cycle coolers that allowed improved efficiency for relatively large heat load applications. The SSTC design uses an efficient inverter and induction motor to drive its high speed compressor operating on minimal wearing self-acting gas tilt-pad and thrust bearings. Gas is expanded through a miniature turboexpander also utilizing long-life gas bearings. Efficiency improvements were made possible by using a high effectiveness counter-flow recuperator comprised of stacked perforated metal plates achieving effectiveness levels greater than 0.98. The projected on-orbit life is greater than 50,000 hours of continuous operation.2,3 Experiment Set-Up and Procedures. The SSTC was assembled and integrated into a specially tailored experiment station (Figure 2) designed specifically for its performance and
Figure 2. SSTC Test Station Set-Up .
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endurance characterization. The set-up consists of a custom cabinet with an integral vacuum system capable of space environmental simulations with vacuum levels exceeding 10–5 Torr. It
has coolant loops routed throughout the station for heat rejection of its various components. Experiment monitoring and health status equipment and instrumentation are provided via the separate control rack. The compressor and inverter are attached to a vertical plate running longitudinally along the top of the cabinet. The expander, cold end heater, brake and heat exchanger are MLI wrapped and contained within the vacuum chamber mounted to the cabinet using reinforced steel framing. Environmental control panels and instruments are built into the front panel on the right side of the cabinet. Data acquisition (DAQ) is provided using a PC-based tabletop computer utilizing customized software. The SSTC is operated according to the experiment parameter conditions listed in Table 2. The cooler is run in a continuous duty-cycle mode except for intermittent stop/start cycling per design specifications, laboratory power supply interruptions, and periodic gas conditioning due to incomplete component cleaning during assembly. Heat rejection is maintained at 300 K during the bulk of the endurance evaluation period. Some heat rejection cycling is currently planned for the remainder of the experiment phase with periodic cooling load and cold end temperature trend analysis performed at the end of the transient cycle. The cooler operates in an unattended mode except for routine data recording, archival storage, instrument re-calibration, and vacuum system adjustments. Long-Term Performance Trends. Figure 3 shows the plot of the long-term cold end performance of the cooler over the entire period of endurance evaluation since April 1995. Figure 4 shows a similar plot over the same time period for input power versus time. Spikes in the performance plot are mostly due to lab power interruptions and periodic warm-
up periods when gas cleaning operations occurred. Over this period, the cold end temperature remained fairly constant at an average temperature of 66 K while cooling a 5 W heat load. Average daily cold end fluctuations were 0.05 K to 0.12 K. The cooler input power requirements have increased slightly over the three years of endurance evaluation from an initial 228 W to nearly 240 W. The expander speed upper boundary was reached in September 1997 and the cold end temperature has steadily increased from 65 K to 70 K over the ensuing 8-month period. The cooler has accumulated over 30,800 hours of total run time, with more than 24,000 hours of near continuous operation during the endurance evaluation period from April 1995 to May 1998. 65 K Diaphragm Flexure Standard Spacecraft Cryocooler Design Overview. The 2 W/65 Kelvin Standard Spacecraft Cryocooler (CSSC) is an
Figure 3. Time/Cold End Temperature History for Creare SSTC.
AFRL ENDURANCE EVALUATION OF SPACE CRYOCOOLERS
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Figure 4. SSTC Time/Input Power History.
engineering development model of a single-stage integral Stirling cryocooler utilizing six doubleacting elastic titanium metal diaphragms in the compressor and expander sections. The cooler is a low pressure, laminar flow machine operating at 32 hertz (nominally) with a mean fill pressure of 8.20 psig. There are no internal organic materials exposed to the helium working fluid. Nominal heat rejection is 300 K. The recuperative-type “flying tube” regenerator is comprised of three diaphragm supported concentric Titanium metal tubes allowing the helium working fluid to cycle in a counterflow fashion in the two annular spaces between them. There are two
compressor units aligned on a common axis operating 180 degrees out of phase for vibration cancellation. Each compressor unit is comprised of an inboard and outboard module. Compression and expansion of the helium is accomplished using the titanium diaphragms. The diaphragms in the compressor and expander sections are actuated by five linear oscillating electromagnetic motors operating in a “pull-only” fashion. The CSSC control incorporates Digital Signal Processing (DSP) methods using a wide bandwidth feedback control loop to manage motor forces. Diaphragm stroke sensing is provided through the use of Kaman position sensors providing feedback to a computer-based software-driven control/feedback scheme4. Experiment Set-Up and Procedures. The CSSC is integrated onto a tabletop experiment station as shown in Figure 5. The cooler is controlled via the electronics rack located at the end of the table. A PC computer is used to interface with the control rack to operate the CSSC. A separate Macintosh computer using LabVIEW™ data acquisition software is used to monitor the experiment and to archive data. The cooler’s cold end is shielded and insulated using MLI. The vacuum bonnet shown in the figure covers only the expander module. A common vacuum system
Figure 5. CSSC Experiment Station Set-Up.
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Figure 6. Time/Cold End Temperature History.
is coupled with the bonnet that also serves other coolers mounted on the same bench. Vacuum level is maintained near Torr during all experimentation. Since this engineering model cooler was not hermetically sealed, a helium working fluid gas makeup system is attached at three separate ports on the cooler to maintain constant internal charge pressure. Although this operating condition does not represent a true on-orbit configuration, the data acquired from the endurance evaluation of the engineering model is useful for possible follow-on diaphragm flexure cooler designs. Heat rejection is provided by a multiplexed single laboratory chiller plumbed to the bench where individual feeds are routed to each cooler under test. The CSSC uses a single inlet and outlet at the base of the compressors for all heat rejection. The CSSC key performance parameters are maintained at the levels shown in Table 3. The load and temperature are maintained constant while periodically cycling the rejection temperature +/-10 K about the nominal 300 K condition over a 60-day cycle. The cooler will be operated in this manner nearly continuously over a total cumulative run time of 80,000 hours or until failure. At the end of each 60-day transient heat rejection cycle, a load versus cold end temperature plot is measured to quantify any time dependent performance shifts that may have occurred. Long-Term Performance Trends. Figure 6 shows the plot of the long-term cold end performance of the cooler over the entire period of endurance evaluation since May 1997. A similar plot of input power versus time is shown in Figure 7. Again, spikes in the performance plots are due to lab power interruption or experiment equipment servicing and maintenance. Over this period, the cold end temperature remained very stable at an average temperature of 65.0 K while rejecting 0.78 +/-0.03 W heat load continuously. Average daily cold end temperature fluctuations were negligible. The cooler input power has remained nearly constant over the entire period of performance and endurance evaluation at the CCL. A total of more than 11,500 hours have accumulated on the machine, with over 4800 of those hours occurring during unattended endurance characterization. 35 K and 60 K Pulse Tube Cryocoolers
Design Overview. AFRL has been pursuing the development of pulse tube cryocoolers for space applications for nearly a decade. Current emphasis is on evolving this technology from
engineering model hardware to flight qualified next-generation coolers. Key thrusts include improving the efficiency of these machines and reducing weight, volume, and input power requirements while extending their useful operating ranges to 35 K and 60 K. Initial pulse tube developments (3585 Pulse Tube Cooler) resulted in the 20-cc compressor based design capable of operating at 35 K and 0.85 W heat lift for 200 W input power to the compressor. The
AFRL ENDURANCE EVALUATION OF SPACE CRYOCOOLERS
39
Figure 7. CSSC Time/Input Power History.
compressor technology was an extension of the spiral flexure spring designs associated with classic Oxford legacy Stirling coolers. After successfully developing the 20-cc compressor for the large capacity 35 K application, a 10-cc compressor-based scaled version of the cooler was built to provide 2 W cooling at 60 K (6020 Pulse Tube Cooler). The 60 K unit is a flight qualifiable design characterized at AFRL and JPL. From the high fidelity 60 K unit, a 35 K /0.3 W version was developed and similarly evaluated at JPL and AFRL. Of these three units, only the 20-cc 35 K and 10-cc 60 K units have entered endurance evaluation. The 10-cc 35 K unit will begin endurance evaluation in the spring of 1998 after completing the remainder of its flight qualification trials. Both pulse tubes in endurance evaluation share a common design heritage and are very
similar in makeup. Each have a compressor comprised of either back-to-back 5-cc or 10-cc modules with opposed pistons compressing the helium working fluid in a common internal compression space. Orifice pulse tube cold heads are attached to the center plate of the housing
assembly of the compressors. The cold end is near the junction of the compressor center plate and the regenerator location at the base of the cold head5. Experiment Set-Up and Procedures. The 3585 pulse tube cooler is integrated onto the same table-top experiment station as shown in Figure 5. The cooler is controlled using low distortion audio amplifiers and a sinusoidal voltage waveform generator. Control electronics are physically mounted in the rack located at the end of the experiment table as shown in Figure 8. A separate Macintosh computer using LabVIEW™ data acquisition software is used to
monitor the experiment and for data archiving. The cooler’s cold end is shielded and insulated using MLI. The vacuum bonnet shown in Figure 8 covers only the pulse tube cold head section
Figure 8. 3585 Pulse Tube Experiment Station.
Figure 9. 6020 Pulse Tube Experiment Station.
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of the cooler. The same common vacuum system described earlier for the CSSC is coupled with the 3585 bonnet. Vacuum level is maintained near Torr during all experimentation from this set-up. The common laboratory chiller is plumbed to the bench where individual feeds are routed to the 3585 cooler’s inlet and outlet ports at the heat rejection plates on each of the compressors. The 6020 pulse tube cooler is mounted in one of four thermally controlled 24-inch vacuum chambers utilizing a common heat rejection fluid loop and vacuum system for space environmental simulation. The vacuum and heat rejection systems in the multi-chamber set-up each have complete backups for fault tolerant operation of all environmental equipment. Together with the uninterruptable power supply, this set-up enables a stable endurance evaluation environment for the cryocooler. The 6020 cooler is mounted inside the 24-inch chamber in a
similar fashion to that shown in Figure 8, except that only MLI is used for shielding and insulating the cold head without the use of the internal vacuum bonnet. The control electronics,
DAQ computers, Lab VIEW™ software, and test instrumentation used for the 6020 unit are nearly identical to that described for the 3585 unit. The 6020 cooler is shown mounted in the experiment chamber in Figure 9.
The 3585 and 6020 coolers’ key performance parameters are maintained at the levels shown in Table 4. The load and temperature are maintained constant while periodically cycling the rejection temperature +/-10 K about the nominal 300 K condition over a 60-day cycle. As previously described for the other coolers undergoing endurance evaluation, these pulse tubes will also be operated in a near continuous mode over a total cumulative run time of 50,000 hours or until the units fail. At the end of each 60-day heat rejection variation cycle, a load versus cold end temperature plot is measured to quantify any time dependent performance shifts that may occur. Long-Term Performance Trends. Figure 10 shows the plot of the long-term cold end performance of both of the pulse tube coolers over the entire period of their endurance evaluation
Figure 10. Time/Cold End Temperature History.
AFRL ENDURANCE EVALUATION OF SPACE CRYOCOOLERS
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Figure 11. Pulse Tube Time/Input Power History.
phase. Figure 11 is a similar plot of cooler input power versus time for each of the machines. Spikes in the performance plots are due to lab power interruption. Gaps in the data since May 1997 indicate periods of time where the experiment was interrupted for system-level sensor demonstrations for the Multispectral Thermal Imager program, as well as fluid loop equipment upgrades for improved heat rejection control. Over this period, the cold end temperature remained fairly constant at an average temperature of 60.1 K for the 6020, and 36.5 K for the 3585 while rejecting 2.0 W +/-0.03 W and 0.85 W +/-0.01 W heat loads, respectively. Average daily cold end temperature fluctuations were +/-0.1 K per day for both coolers. At the start of the endurance evaluation phases for these coolers, input power and baseline performance maps had apparently drifted slightly since the beginning of performance evaluation at the CCL. Since then, no further drifts have been observed. However, further investigations into the causes of performance drift are needed and will be conducted now and culminate when out-of-bounds conditions are reached. A total of more than 6,500 hours have been accumulated on the 3585 unit, and approximately 3800 cumulative hours on the 6020 cooler. Roughly half of these total cumulative run-time figures have resulted from endurance evaluation.
Figure 12. Cycle Efficiency Histories of Coolers at AFRL.
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LONG-TERM PERFORMANCE COMPARISONS
Cycle efficiency is also a key figure of merit for long-term cooler performance comparisons. For the coolers presented in this paper, the cycle efficiency is considered to be a ratio of the loadto-input power normalized to the pulse tube’s Carnot efficiency by the following relationship:
Figure 12 shows the plots of the calculated cycle efficiencies of the coolers undergoing endurance evaluation at the CCL. As illustrated in the figure, all of the coolers undergoing endurance characterization exhibit fairly stable long-term performance (in terms of percent Carnot efficiency) over periods of near continuous experimentation.
SUMMARY AND CONCLUSIONS
The coolers undergoing endurance evaluation at AFRL are proving to be stable in their longterm performance. Most observation of performance to date have led to improvements in followon designs based on these early generation hardware experimentation. Performance drift
associated with the pulse tube coolers will require additional analysis if out-of-bounds conditions occur. However, predictions based on performance trends to date show refrigeration capabilities
of these two units are still highly maintainable and are within their excess design margins. Although long-term performance data gathered on AFRL coolers is far from conclusive at this stage, these experiments indicate that efforts to evolve cryocooler technology into hardware suitable for long-term space use are beginning to show promise. The wealth of information emanating from these trials will continue to serve as valuable tools to aid in performance, life, reliability, and spacecraft thermal control tradeoffs for mission-critical cryocooler space applications well into the future. REFERENCES 1.
Bruning, J., Pilson, T., “Phillips Laboratory Space Cryocooler Development and Test Program”, 1997 ICEC Conference Proceedings, Plenum Press, New York (1998).
2.
Pilson, T., Bruning, J., Gilbert, A., “65 K/5 W Single Stage Turbo Cryocooler Endurance Test”, Phillips Laboratory (AFRL) Interim Report, Kirtland AFB, NM (1996).
3.
Dolan, F., et al., “A Single Stage Reverse Brayton Cryocooler: Performance and Endurance Tests on the Engineering Model”, Cryocoolers 9, Plenum Press, New York (1997).
4.
Stacy, W. D., “Development of a Metal Diaphragm Type 65 K Standard Spacecraft Cryocooler”, Final Report, Creare, Inc., Air Force Phillips Laboratory TR-94-1156, Kirtland AFB, NM, 1994.
5.
Burt, W., “35 K Pulse Tube Cryocooler”, Final Report, TRW, Inc., Air Force Phillips Laboratory TR-94-1156, Kirtland AFB, NM, November 1995.
DARPA Low Cost Cryocooler Performance Testing: Preliminary Results
T.G. Kawecki Naval Research Laboratory Washington, DC, USA 20375 S. C. James AlliedSignal Technical Services Corporation Camp Springs, MD, USA 20746
ABSTRACT
Many semiconductor electronic technologies (such as Silicon CMOS and Gallium Arsenide HEMT) and High Temperature Superconducting (HTS) electronic devices offer significantly enhanced performance at cryogenic temperatures compared to ambient temperature operation. There have been few commercial or industrial products that exploit cryogenic temperature operation due to the cost and reliability of cryocooling system. To address this cryocooling need, the DARPA Low Cost Cryocooler Program has supported the development of a number of low cost, long life cryocoolers. The development goal in this program is to provide cooling capacity (that is required for a specific commercial application) with a minimum 3 year life at one thousand dollar cost target for a ten thousand unit per year production. Six types of cryocoolers were supported by this effort. One from each of the following vendors: Superconductor Technologies Inc., CTI, APD Cryogenics, CryoMech Inc., Raytheon (formerly Hughes Aircraft) and MMR Technologies, Inc. Demonstration systems from each of these vendors have been delivered to the Naval Research Laboratory and the testing of these systems has begun during the past year. In order to verify the performance of these new cryocoolers, the Naval Research Laboratory has established a long life cryocooler testing facility. Cryocooler performance parameters such as cryogenic thermal load, cold finger temperature, cryocooler input wall power and cryocooler rejection temperature are instrumented. Pertinent ambient conditions such as ambient room temperature and cold finger vacuum levels are also instrumented. An automated data acquisition system, utilizing a PC with Lab View software, is used to monitor, control, and archive cooler performance data. Approximately one year of testing has been completed on the first cryocoolers delivered under the Defense Advanced Research Projects Agency (DARPA) LCC program to date. A brief overview of the coolers under test is presented as well as selected test results. An overview of the existing test facility and future plans is presented.
Cryocoolers 10, edited by R. G. Ross, Jr.
Kluwer Academic/Plenum Publishers, 1999
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DARPA LOW COST CRYOCOOLER PROGRAM BACKGROUND
Shortly after the discovery of superconductivity with transition temperatures above the temperature of liquid nitrogen (at atmospheric pressures), DARPA started a program to exploit the properties of superconductivity at these elevated cryogenic temperatures in various system applications, both high power as well as electronic. From the beginning of this program, the importance of closed cycle cryogenic refrigeration technology was fully appreciated but it was not until 1993 that DARPA initiated a program to develop cryogenic refrigeration systems (“cryocooler”) which would be optimized for use in high temperature superconducting and cooled electronic applications. Specifically, the goals for the program were to develop cryocoolers with cooling capacities ranging from about 4 watts at 60 K (for a typical superconducting applications) as the lower temperature limit, up to a cooling capacity of 50 watts at 150 K (for a cooled microprocessor applications) as the upper temperature limit. The prospective contractors were asked to explore candidate system applications and, then formulate a concept for a cryogenic refrigerator, which could cool the system of interest with emphasis on low cost and high reliability. Once the critical thermodynamic requirements of the candidate application were defined, the other physical and electrical parameters were negotiable, consistent with the system requirements and with low cost and high reliability. The only constraints imposed on the cryocooler development program were the goals for low cost and high reliability. For viable commercial applications, a cost goal of US $ 1,000 per unit (when manufactured in quantities of 10,000 units per year) was, somewhat arbitrarily, selected. However, for expensive systems, a cost target of 10 percent of the total system cost might be acceptable. The other specified goal of the program was high reliability with “troublefree” operation for more than three years as a goal. Unfortunately, there is no accepted definition for reliability in military applications. The term Mean Time Before Failure (MTBF) is frequently used while for commercial applications. A reliability figure, for example, 95 percent reliable, over some period of time is commonly used. A precise definition of reliability needs to be accepted by the cryocooler community. As shown in Figure 1, the goals for both cost and reliability were significantly beyond what is commonly achieved for cryogenic refrigerators built for the commercial market. Thus, the LCC program was formulated to focus the interest of the cryocooler community on improving the performance of their products to satisfy the perceived needs of the HTS and cold electronics communities.
Figure 1. Cryocooler Cost Vs Life Time Trends.
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INTRODUCTION TO DARPA CRYOCOOLERS UNDER TEST
It is important to note that the DARPA cryocoolers are developmental cryocoolers, not production prototypes. They are an early step in the evolution toward long life lost cost production cryocoolers. The nominal performance of the DARPA coolers covers a wide performance range and is summarized in Table 1. A brief description of each cooler is contained in this section. The Superconductor Technologies Incorporated (STI) Cryocooler is a gas bearing/free piston, integral Stirling cycle cryocooler. The cryocooler features, a linear motor, hermetic seals and closed cycle control. The Cryogenic Technologies Inc. (CTI) cryocooler is a Gifford-McMahon (G-M) machine based on their compressor and cold head assemblies from their cryogenic vacuum pump product line. The CTI design has been modified to reduce cost of the cold head assembly by incorporating pneumatic valves and reducing the overall parts count. The CryoMech Inc. machine also uses an existing commercial G-M compressor (1 kW) but uses a pulse tube cold head designed by researchers at the University of Wisconsin-Madison. It is due for testing the summer of 1998. A single stage throttle cycle design utilizing a mixed working gas is used by the APD Cryogenics Inc. cryocooler. It is comprised of a modified commercial air conditioning compressor and a cold head expansion unit connected by a high-pressure gas hose. The Raytheon (formally Hughes Aircraft) pulse tube cryocooler utilizes a modified linear tactical cryocooler compressor connected via a short transfer line to a pulse tube expander. Due to an anomaly in the compressor design under the DARPA contract, the demonstrated
performance is 1.4 W @ 77K instead of the original 4 W goal. This unit will be delivered to NRL in the summer of 1998. The MMR Technologies Inc. LCC Cryocooler is based on the Klemenko Cycle that uses a multi-component refrigerant mixture in a single stream, cascade, throttleexpansion refrigeration cycle. This unit is configured as a stand-alone medical laboratory bench device to cryocool biological samples in a non-vacuum environment to 114 K. A gas mixture was evaluated for 90 K operation but after several hundred hours of operation, the cooler froze up. The available funding prevented exploring additional mixtures, which might work reliably at temperatures near 80 K. CRYOCOOLER TEST FACILITY DESCRIPTION AND PERFORMANCE
There are four main goals for the NRL LCC test facility, characterize fundamental cooler performance parameters, test at cryogenic loads and temperatures representative of commercial applications as much as possible, provide longevity performance feedback, and protect the cryocoolers under test. The primary goal of the NRL test effort is to provide independent early feedback on the performance potential of the DARPA LCC cryocoolers to manufacturers and to potential users.
The thermal load and cryogenic temperature test levels were selected by the cryocooler vendors. These levels were selected based on the cryocooler design and its targeted commercial applications. Testing is limited to continuous operation at constant thermal loads and ambient
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temperatures. NRL testing is expected to provide an indication of the longevity potential of the these cryocoolers, it is not intended to be as environmentally stressing as may be required for some commercial applications. Primary cryocooler performance parameters such as cryogenic temperature, cryogenic thermal load, compressor rejection temperature and wall input power are characterized. In addition, experiment parameters such as ambient temperature and vacuum levels are recorded. Instrumentation types and accuracy are shown in Table 2. At this time (May 1998), approximately one half to one full year of cumulative testing has been completed on a number of DARPA LCC units. Cryocooler cumulative test time to date depends on when it was delivered and the amount of failure down time it experienced, if any. The test goal is 2 to 3 years of operation. Cryocoolers currently under test include the two APD units, two CTI units, three MMR units, and two STI units. Raytheon and CryoMech coolers are due at NRL in summer 1998. The NRL LCC test facility consisted of two primary pieces of equipment, a cryopumped central vacuum system and instrumentation. All but one of the DARPA cryocoolers have cold fingers and require vacuum support to minimize parasitic thermal heat input to cold fingers. The exception is the MMR cooler that is designed as a standalone laboratory instrument for cooling biological samples within an integral ambient pressure container. A central vacuum hub provides vacuum to each test station through vacuum arms that are valved (with roughing ports) to allow coolers to be added and taken off line without interrupting other coolers under test. Vacuum has been maintained at each cryocooler test station to Torr or less to date except for planned and automatic cooler shutdowns. The MMR cryocoolers are not connected to the vacuum hardware and are run separately. Ambient temperatures in the cryocooler test lab have been a consistent The central element of the instrumentation is a PC based data acquisition and control system using Lab View software. Data is collected every two minutes for monitoring and recorded every 30 minutes on all cryocoolers as well as the test system itself. A green/yellow/red status system is used on all collected data. Yellow data conditions change nothing, it only serves as an alert to the operator. In the case of red condition data, only the red condition cooler is automatically shut down. If there is a test system failure, or loss of wall power, the test system and all coolers are automatically gracefully shut down. To date there have been only nine Lab View initiated cooler anomaly shutdowns in approximately forty thousand accumulated cooler test hours. The test facility runs continuously and is attended on a normal work week schedule at approximately a half man year level of effort. NRL test facility capabilities have continually evolved since the start of testing. To accumulate test time as early as possible, the first cryocoolers on test were initially run at cold no-load conditions. A closed-loop thermal load capability was added in the early phases of testing and then revised, after approximately 1500 hours of operation, in order to control thermal load power. Digital power analyzers, to characterize the wall-input power, were added to some
DARPA LOW-COST CRYOCOOLER PERFORMANCE TESTING
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coolers after testing began and adequate funding became available. There have been two planned system shutdowns for test facility upgrades. The first shutdown occurred in July 1997, lasting for 3 weeks, the second occurred in January 1998, lasting for 2 weeks. In addition there has been one planned building power outage, which caused the test facility to be shut down for twentyfour hours. There have been no unplanned test facility based cryocooler shut downs. PRELIMINARY DARPA CRYOCOOLER PERFORMANCE RESULTS
The two principal design challenges of these experimental cryocoolers are cost and longevity. Thus performance verses time is the focus of the results presented here. Full
cryocooler test results cannot be reported in this paper due to space limitations. Initially all coolers received at NRL are characterized for thermal load verses cryogenic temperature at ambient temperature to provide a benchmark. Cryocoolers are then put into longevity testing. Cooler performance testing is reported in cumulative operating hours. Note that these hours are not continuous operating hours due to three planned test facility shut downs and any Lab View initiated cooler anomaly shut downs that may have occurred. Cryocooler shut downs and subsequent actions are noted in the individual cooler performance time plots in Figures 2 through 7. Each data tick on the time axes illustrated in Figures 2 through 7 is an average of 30 minute archived data over a 24-hour period. A complete set of performance verses time data (cold finger temperature, thermal load, wall input power and rejection temperature) is presented for the APD Standard cryocooler as a representative sample of the primary cooler data collected by the NRL cryocooling test facility in Figures 2 & 3. Limited data is presented for the rest of the cryocoolers under test. The initial STI coolers under test failed and were replaced with updated designs. The original and updated STI cooler #1 designs are reported in Figures 4 and 5. The CTI #1 cooler is reported in Figure 6. Cold finger temperatures only are reported for the three MMR cryocoolers in Figure 7. A brief historical synopsis of the performance of each cryocooler under test follows. APD Cryogenics Incorporated
Two prototype APD LCC cryocooler designs are under test at NRL, they are designated the Standard and Modified units. The Standard unit is among the first coolers on test and is still running at nominal performance levels. It suffered a red status Lab View shut down at 4210 hours when the cold finger temperature climbed from 79°K to 85°K in a short time. It was restarted and is still operating at 7900 hours at a reduced thermal load from 5W to 4W. The Modified APD unit is experimental and has suffered gradual loss of performance. At 1034 hours a loss of charge pressure caused a red status shutdown. After a gas recharge and gas transfer line replacement the cooler was restarted with the thermal load reduced from 5 to 4 watts
to maintain 80K. operation. Another red status shut down occurred at 2680 hours due to climbing cold finger temperatures caused by low gas charge pressure. The Modified unit was inspected and recharged at APD and restarted in test at NRL, but again suffered a loss in performance from 4 W to 3 W to maintain the target 80K temperature. The final red status shut down occurred at 3590 hours when the cold finger temperature rose again. At that time the compressor was switched out for a new one and the modified cooler has since accumulated an additional 1000 hours giving a total of 4600 hours. Superconductor Technologies Incorporated
There are two STI cryocoolers under test. At 1690 hours the #1 unit electronic control box overheated and both #1 and #2 cooler electronics were sent back to STI for upgrades. At 2050 and 2040 hours on units #1 and #2 respectively, the original STI cryocooler designs suffered an apparent ceased piston and shut themselves down Both coolers were sent back to
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Figure 2. APD-Standard cooler cold finger temperature and cryogenic thermal load performance at 296K ambient temperature.
DARPA LOW-COST CRYOCOOLER PERFORMANCE TESTING
Figure 3. APD Cryogenics Standard cooler wall input power and rejection temperature performance at 296K ambient temperature.
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Figure 4. Superconductor Technologies original cooler #1, cold finger temperature and cryogenic thermal load performance at 296K ambient temperature.
DARPA LOW-COST CRYOCOOLER PERFORMANCE TESTING
Figure 5. Superconductor Technologies updated cooler design, cold finger temperature and cryogenic thermal load performance at 296K ambient temperature.
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Figure 6. CTI Cryogenics cooler #1 cold finger temperature and cryogenic thermal load performance at 296K ambient temperature.
DARPA LOW-COST CRYOCOOLER PERFORMANCE TESTING
Figure 7. MMR Technologies cold finger temperature performance for units #1-3 at 296K ambient temperature.
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STI for rebuild to a new engineering design. Both units are back under test, with approximately 1400 trouble free hours, operating at a thermal load of 4W at 78K. Cryogenic Technology Incorporated
There are two CTI cryocoolers under longevity testing. CTI cooler #1 operated for 5230 hours until a red status Lab View shut down occurred due to rising cold finger temperatures. Upon investigation a low compressor charge pressure was identified and the unit is currently back at CTI for inspection and recharging. CTI cooler #2 also experienced a loss in compressor charge pressure at 5030 hours of operation and is also at CTI for inspection and recharging. MMR Technologies Incorporated
All three MMR cryocoolers under test have logged approximately 5000 hours of testing each. There have been no anomalies or degradation in performance observed. SUMMARY AND FUTURE PLANS
Performance testing of the DARPA low cost program cryocoolers at NRL is still in the early stages. Two to eight thousand hours of testing has been completed on multiple experimental cryocoolers from APD, STI, CTI, and MMR. The test goal is to acquire two to three years of performance testing on these coolers. The NRL automated test facility is performing well. It has provided a consistent test environment and there has been no unplanned shut downs caused by the test facility. The most common problem in forty thousand hours of cryocooler testing in these units so far has been a loss of gas charge pressure. It is the principal cause of anomaly shut downs on the APD Modified and CTI units #1 and #2 cryocoolers. Significant gas loss in production long life refrigeration coolers and cryogenic vacuum pumps is well understood. It is likely that the loss of gas charge pressure would not be a difficult problem to solve in production units. The only catistrophic cooler failures have occurred on the original STI experimental crycoolers. New STI designs are now under test and performing consistently at nominal levels. In the fall of 1998 the DARPA sponsored CryoMech and Raytheon pulse tube cryocoolers should be under test for a total of 12 cryocoolers in performance testing. ACKNOWLEDGEMENTS
Sponsorship and funding for cryocooler testing is provided by the Defense Advanced Research Projects Agency (DARPA). The cooperation of the DARPA cryocooler vendors, APD Cryogenics, CTI, Superconductor Technologies, MMR Technologies, CryoMech and Raytheon, has been instrumental in the success of this testing.
Development of Cryogenic Cooling Systems at the SR&DB in the Ukraine S.I Bondarenko and V.F. Getmanets
Special Research and Development Bureau for Cryogenics Technologies Kharkov, Ukraine
ABSTRACT
This paper provides an overview of the cryogenic development activities carried out over the past 25 years by the Special Research & Development Bureau for Cryogenic Technologies (SR&DB) in the Ukraine. The SR&DB is a part of the Scientific and Technical Complex and Institute for Low Temperature Physics and Engineering, a unique center for cryogenic technology development within the ex-USSR.
INTRODUCTION Since 1971, all work in the ex-USSR in the field of cryogenic cooling systems has generally been carried out at the Institute for Low-Temperature Physics & Engineering – Scientific and Technical Complex (ILTPh&E–STC). That year, the Special R&D Bureau in Cryogenic Technologies (SR&DB) was established to carry out such tasks. The ILTPh&E–STC is comprised of: 1) the Low-Temperature Physical and Technical Institute, and 2) the SR&DB, which has two experimental facilities. The Institute has been conducting fundamental research in the field of low-temperature physics in the following technical areas:
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superconductivity, low-temperature magnetism physics of solidified gases physics of liquid helium and methods of obtaining hyper-low temperatures
The main directions for applied developments at the SR&DB have been:
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cryogenic cooling systems (space applications of first priority) facilities for simulation of space conditions and the study of various materials super-sensitive instruments for measurement of magnetic or thermal field parameters and of environmental gas components
There are about 1,500 personnel at the SR&DB, and the complex has its own liquefied gas production facility. The Scientific and Technical Complex is a member of the enterprises belonging to the National Academy of Sciences of the Ukraine. Cryocoolers 10, edited by R. G. Ross, Jr.
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Figure 1. Combined space cooling system.
Figure 2. Combined-environment space simulator.
GENERAL TRENDS, FEATURES AND RESULTS OF DEVELOPMENTS
A unique feature of the ILTPh&E–STC is their capability for solving tasks in the field of cryogenics, from fundamental physics research to the manufacture of new systems using complex methods, all in a single unified center. As a result, the STC has successfully developed, in Soviet Union times: • space-borne infrared-detector 84-90 K cooling system1 of 5-year orbit lifetime (Fig. 1) • the first search-oriented squid-magnetometers in the ex-USSR • SQUID-magnetometers for study of heart and brain magnetic fields • the first space-borne mass-spectrometer; it was applied for monitoring the gas-composition of the Venus-atmosphere • small-dimension airborne infrared radiometers with cooled detectors • unique simulator of space conditions (Fig. 2) capable of simultaneous reproduction of: - changing range of low temperatures - vacuum
- electron and proton flux - solar radiation flux a mock-up of a space-borne infrared telescope with optics cooled to 2K The SR&DB has gained significant experience in the development and fabrication of various types of cryogenic cooling systems including: • cryostats with liquid helium, hydrogen, and nitrogen • cold accumulators involving solidified gases such as argon, nitrogen, neon, and methane • throttle-type cryocoolers utilizing the Joule-Thomson effect • pulse tube cryocoolers • Stirling cryocoolers One of the most important developments of the SR&DB has been the engineering of a long life spaceborne 84-90 K cooling system1 (Fig. 1). Two key design philosophies have served as the system structure or design nucleus for this work and other SR&DB cryocooler systems2,3: 1) To increase system lifetime, a combination of a mechanical refrigerator and a stored cryogen system has been used4. The operating time on the mechanical refrigerator is thus greatly reduced and its life increased by only cycling it on periodically to refreeze the stored cryogen gas; the melting of the stored cryogen provides the continuous cooling load. 2) A low pressure mixed-gas working fluid5 is used to prolong the life of the cryocooler compressor in the Joule-Thomson system. This allows the use of non-wearing piston clearance seals and the use of ball bearing piston supports that do not require lubrication.
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Figure 3. Solid-body cold accumulator.
For some missions (like for the Soviet made Salyut-4 Space station) stored cryogen systems based on solid gases have been used as the sole cooling agent for specific applications such as cooling infrared detectors (Fig. 3). The logged lifetime of such systems is typically on the order of 0.5 – 1.0 year (Table 1). Simplicity, reliability and low cost are distinguishing features of these systems. Cryogenic accumulators4 have been developed by the SR&DB for different temperature ranges, from 10 to 190 K. The cooling power of the cooled object is typically on the order of 0.1 to 2.5 W.
A liquid cryogen cooling system using superfluid helium has also been developed by the SR&DB for a spaceborne infrared telescope (Fig. 4). The main problem, dimensional stability of the cooled structure when transitioning from 200 K to 2 K, was thoroughly solved based on a careful selection of the telescope structure design scheme, the choice of proper materials, and optimization of the cool-down and warm-up regimes. For cooling of superconducting elements in a spaceborne SQUID-magnetometer we have developed and manufactured planar Joule-Thomson type micro-refrigerators analogous to the well known developments6 of Prof. Little in the USA.
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Figure 4. Space-borne infrared telescope.
Other areas of technology development include: • Engineering of a lightweight, small-size tactical cryocooler for use at 65 K with approximately 1 W of cooling power • Development of physical approaches7 for the implementation of accelerated tests8 for cryocooler hardware destined for 5 to 10 years, or more, of operational life • Design and utilization of high-efficiency thermal insulation9 for cryogenic vessels of various applications ACKNOWLEDGMENT
For the past two years, the SR&DB has been cooperating with the U.S. Air Force Research Laboratories, the U.S. Naval Research Laboratories, and Orbita Ltd. in the area of cryogenic technologies. We wish to express our acknowledgment and gratitude to ESARD, OAID, AFRL, and Orbita Ltd. for their efforts in organizing our participation at the 10th ICC. REFERENCES 1. Arkhipov, V.T. et al., “Long Life Cryocooler for 84-90 K,” Cryocoolers 10, Plenum Press, New York, 1999.
2. Arkhipov, V.T. et al., “Low Weight and Long Life 65K Cooler,” Cryocoolers 10, Plenum Press, New York, 1999. 3. Arkhipov,V.T. et al., “Long-life 5-10 K Space Cryocooler System with Cold Accumulator,” Cryocoolers 10, Plenum Press, New York, 1999.
4. Arkhipov, V.T. et al., “Cold Accumulators as a Way to Increase Lifetime and Cryosystem Temperature Range,” Cryocoolers 10, Plenum Press, New York, 1999.
5. Arkhipov, V.T. et al., “Multicomponent Gas Mixtures for J-T Cryocoolers,” Cryocoolers 10, Plenum Press, New York, 1999. 6. Little, W.A., “Microminiature refrigerators,” Scientific Instruments, No. 5, 1984, pp. 3-26. 7. Getmanets, V.F. and Zhun', G.G., “Cryocooler Working Medium Influence on Outgassing Rate,”
Cryocoolers 10, Plenum Press, New York, 1999. 8. Getmanets, V.F. and Zhun, G.G., “Accelerated Cryocooler Life Tests for Cryodeposit Failures,” Cryocoolers 10, Plenum Press, New York, 1999. 9. Getmanets, V.F., et al., “Cryogenic Superinsulations with Increased Efficiency,” Advances in Cryogenic Engineering, Vol. 43B, Plenum Press, New York, 1998, pp. 1319-1325.
Qualification Test Results for a Dual-Temperature Stirling Cryocooler W.J. Gully, H. Carrington, and W. Kiehl Ball Aerospace & Technologies Corp. Boulder, CO, USA 80306 Thom Davis and B.J. Thomlinson U.S. Air Force Research Laboratory Albuquerque, NM, USA 87117
ABSTRACT We report on the latest of our Stirling-cycle mechanical cryocoolers. Our coolers are specifically designed to work in the space environment and to meet requirements such as power efficiency, compatibility with sensitive instruments, and high reliability. In this work we discuss a cooler tailored to efficiently provide refrigeration at two different temperatures simultaneously. This protoflight maturity-level cooler delivers approximately 0.4 W of cooling at 35 K and 0.6 W of cooling at 60 K for 80 W of input power from a dc supply. We will describe its general capabilities and some of the peculiarities associated with our approach to controlling two thermal loads with a single cooler. INTRODUCTION We have been funded by the Goddard Space Flight Center and the U.S. Air Force Research Laboratory to develop mechanical cryocoolers specifically for space applications. A space design emphasizes reliability, followed by efficiency, compactness, light weight, and the ability to work in the space environment. We began by building a new space cooler based on the outstanding “Oxford” linear technology. This cooler then evolved into a series of coolers with the same core technology but tailored to meet specific thermal needs. Our cryocooler consists of a compressor, displacer, and space-compatible control electronics. The original unit had a displacer with two expansion stages to achieve high power efficiency at It weighed 15 kg, operated from a 28 Vdc power source, and included features to suppress electrical emissions and vibration. From the beginning, the mechanisms were designed to support verification methods for the clearance in our seals. This cooler evolved into a series of coolers, each of which had its displacer cold head customized to meet a particular thermal requirement. In this work, we will describe a unit that we developed to provide cooling at two different temperatures simultaneously. The mechanical parts of this cooler are shown in Figure 1.
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Figure 1. Our multistage Stirling cryocooler with multiple interfaces for providing refrigeration at 35 K and 60 K simultaneously.
THREE KEY ASPECTS OF THE DUAL-TEMPERATURE COOLER
Complex infrared systems often need both detector cooling and separate optics cooling at a higher temperature. It would save weight and power if we could provide both of these needs with a single cooler. We developed this cooler to meet a specific requirement for 0.4 W of refrigeration at 35 K and 0.6 W of refrigeration of Since our existing 30 K two-stage cooler could not carry the combined watt of cooling ( ) at the lower temperature, we knew that we would either have to increase the efficiency of the machine or scale up the whole cooler. Motivated to minimize any design changes, we looked for a way to get the additional cooling by just modifying our displacer. After some detailed analysis, we found that we could achieve our goal by adding an additional expansion stage to our two-stage cold finger. The additional stage intercepts most of the displacer’s own internal parasitic heat load at a higher temperature, which frees up capacity at 60 K to meet the external needs. We also found it necessary to add an explicit heat exchanger at the upper stage, as we had at the cold tip, to minimize the thermal
resistance between the working fluid and the external load. This type of heat exchanger was possible only because of our fixed regenerator design, which has the regenerator stationary and
external to the displacer piston. This is the second of our research coolers equipped for monitoring its internal clearance seals. These coolers are equipped with a suite of internal sensors, whose feedthroughs are visible on the ends of the hardware in Figure 1. We began with the Oxford approach of supporting the moving armature with diaphragm springs. We extended this approach by adding sensors to confirm that the scheme worked as expected over the cooler’s life. The first of these coolers has been in life test at the Goddard Space Flight Center for about a year.
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This is the second cooler to complete a qualification test program representative of typical flight requirements. The tests and their key requirements are given in Table 1. QUALIFICATION TEST RESULTS
A capsule summary of the tests is shown in Table 1. We began these tests in March, and expect to complete the remaining test shortly. We will elaborate on each of these tests.
Thermal Performance
Typical thermal performance results are shown in Figure 2. At our baseline operating point of 36 Hz and 145 psig charge pressure, the cooler lifts about 0.6 W at 61 K and 0.45 W at 35 K for approximately 80 W of input power. In Figure 3 we provide a more complete set of data showing a range of heat loads on the upper stage as well as on the cold tip. The data in Figure 3 shows that there is little cross coupling of loads between the stages, which means that the temperature of one cooling stage is relatively insensitive to the heat load applied to the other stage. This occurs because the added heat at one stage produces offsetting heating and cooling trends at the other. Dual-Temperature Control. We also performed a series of tests in which we explored the ability to control two stages at the same time. The obvious difficulty is that the controls are intertwined because the cooling comes from a common source. As presently configured, we use
the cryocooler stroke to regulate the cold tip temperature to 35 K and a small makeup heater to maintain the upper stage at 60 K. This approach works only if there is excess refrigeration available at the upper stage. For example, when the cold tip carries a load of 0.4 W, the cooler is
Figure 2. Cold tip heat lift of the dual-temperature cooler with 0.6 W on the midstage.
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Figure 3. A heat lift matrix for both the cold tip and the midstage.
close to full stroke. Under these conditions the upper stage has a capacity of approximately 550 mW at 60 K. We can then rely on the makeup heater to regulate the upper-stage temperature to
60 K for an external heat load that lies between 0 and 550 mW. Because we are using stroke control, the heat load on the cold stage sets an upper bound on the cooling available at the midstage. Another consequence of the way we implement temperature control is a marked difference in the settling times of the two stages to a load change. The makeup heater on the upper stage is in a simple integrating loop with a time constant of about a second. A change in the external load at the midstage is compensated for rapidly, and there is no need for the cooler to react at all. But a change of load on the cold tip is a real change in the heat balance and takes much longer to settle because stroke control is rather slow. At present, each update requires at least 10 seconds, and the overall settling time for our “un-tuned” system can take several minutes as shown in Figure 4. In this particular experiment we made the cooler react by changing the heat load on the cold tip
by 100 mW. The uppermost curve in the figure shows how the stroke changes as it tries to keep the cold tip temperature (the lowest curve) at 35 K. But note how the 60 K stage stays reasonably constant as the makeup heater compensates rapidly for the change in refrigeration.
Figure 4. A transient test illustrating temperature control with a dual-temperature cooler.
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Noncontacting Operation The most common factors that limit cooler life are gaseous contamination and wear. We addressed the contamination problem with careful design and systematic processing of the materials in the cooler. Our approach to the wear problem was to ensure that the cooler operated elastically and had no rubbing parts. We started with the Oxford concept, where the armatures are supported on flexing diaphragm springs. We went on to add sensors to actually measure the
clearance gaps, first with the compressors, and later with the displacer. Design improvements and sophisticated alignment procedures have led to coolers that we believe do not contact. With our internal sensors we learned that there is a pneumatic side force on the piston in an
operating cooler that is not present statically. The pressure wave generated by the piston leads to a side force unless the piston is perfectly aligned in its cylinder. These side loads can be considerable, but can be minimized by using well-made parts and by precisely aligning the hardware. The goal is to limit the side loads to magnitudes that can be supported by the spring system. The critical area in the compressor is the piston clearance seal gap. This must be large enough to allow for the gravitational sag and for the dynamic excursions, yet still be small enough to function as a seal. The critical area in the displacer is the cold tip clearance gap. Even though
the armature is lighter and the pneumatic forces smaller, the tip is cantilevered a distance from the support and shaft bending becomes an additional concern. The cooler we are discussing now is the second of two research coolers equipped with these sensors. For a flight cooler, we use temporary sensors for alignment and for qualification test monitoring, but we will remove these before the unit is permanently sealed. These sensors are useful in a number of ways. One of the most immediate is to lay the cooler on its side and detect the armature's internal shift due to gravity. If the armature bumps into anything it immediately splays in an odd direction. Because of its simplicity, this is one of the first tests we perform after an event such as a launch vibration test. We next perform “stiction”
tests, which look for drag as the armature is slowly moved through its available stroke. Together these test for simple physical obstructions.
Contact during dynamics is a separate matter. Although originally intended for static measurements, the internal sensors were also able to monitor the lateral motions of the armature while the cooler was in full power operation. We illustrate the lateral motion of our compressor
piston in Figure 5 as a function of stroke. We contrast two cases, one at atmospheric and one at full pressure, to illustrate the importance of the dynamic pneumatic force. The nearly horizontal line represents the lateral deflection of the piston at one atmosphere. There is a small amount of apparent motion, which we have found in part to be spurious and due to imperfections in the
Figure 5. A plot of the lateral deflection of our compressor piston along its stroke as a function of the charge pressure in the cooler.
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sensor background signal. By contrast, when we pressurize the unit the sideways motion grows much larger. The deflection is a maximum as the pressure peaks while the compressor is moving towards top dead center. The plot resembles a pV diagram simply because the deflection is in phase with the pressure and not the stroke. The displacer deflection is plotted in a different way in Figure 6. When we use a lissajous pattern we lose the relationship between the sideways deflection and the stroke, but we see the deflection in both lateral axes simultaneously. This pattern is typical in that the deflections are in the main at the operating frequency of the cooler and polarized in one particular direction. Launch Vibration
We next subjected the mechanical hardware and the electronics to typical launch vibration levels. Our integrated random level was 14.1 Grms, with a two-minute dwell in each axis. We also performed a 15 G swept sine test at 4 octaves per minute, because each linear mechanism had a necessary soft mode in the axial direction. We damp this mode with a passive magnetic brake. We confirmed the function of this brake by monitoring the internal position of each armature during
the test. Afterwards we repeated a representative set of the performance tests after the launch vibration test to demonstrate that the cooler survived unchanged. EMI
We also took our 35/60 K set of electronics through a restricted set of EMI tests. Requirements were derived from the tables in the NASA General Environmental Verification Specification, but generally coincided with MIL-STD 461B. The 35/60 K electronics are considered “flightlike,” which means that they functioned properly and were an appropriate stand-in for all environmental tests, but are not built of costly flight parts. This set of electronics has been superceded by another generation being developed on the High Resolution Dynamics Limb Sounder (HIRDLS) program at Ball.
Figure 6. A lissajous pattern depicting the lateral motions of the displacer cold tip.
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As this was a second test of similar electronics, we performed only a few key emissions tests to evaluate modifications to our package. A more complete set of tests, including conduced and radiated susceptibilities, had been performed on our earlier cooler. We found that for our dual-temperature cooler, CE01/03 and RE02 emissions continue to be well below specification limits because of careful attention to suppressing spurious ground currents associated with our pulse-width modulation drive. Vibration Export Vibration export tests are sensitive to the method in which the cooler is mounted, and have to be performed in a special force dynamometer. Our dynamometer supports the compressor and displacer on independent stands, and is instrumented to measure both forces and moments from each separately. The vibration is a series of force harmonics because the cooler motion is periodic. We record the vibration on a spectrum analyzer while the cooler is running at full power and at low temperature. Our requirement is for 234 mN peak in each axis and at each harmonic. We find that the displacer meets this requirement in all directions without active axial vibration control. Without axial control the compressor’s axial vibrations exceed the requirement, but can be brought substantially below by our vibration control algorithm. Lateral vibrations are typically at or below the specification levels and are insensitive to attempts to control the axial vibration electronically. They depend primarily on the mode structure in the cooler, and are sensitive to the operating frequency and stroke amplitude. Thermal Vacuum Our final qualification test is a daily series of cycles in a vacuum test chamber for eight days. We control the mechanical unit heat sinks from –40 °C to +50 °C and back once a day. We separately adjust the electronics heat sink temperature between –20 °C and +50 °C every two days in order to test the temperature combinations. The test itself consists of a start, cooldown, and load line, which we repeat each cycle to check for the long-term repeatability of performance. SUMMARY
We have finished a number of qualification tests on the new dual-temperature cooler. We will be delivering the cooler to the Air Force Research Laboratory, where we will continue with the thermal vacuum testing. After we conclude these tests, we will perform an endurance test to evaluate the life and reliability of the unit. ACKNOWLEDGMENTS
We appreciate the long-standing support and funding of Steve Castles at the Goddard Space Flight Center, and the efforts of Kevin Byrne and Larry Crawford at the Air Force Research Laboratory during earlier phases of this work. REFERENCES 1. Carrington, H., W.J. Gully, W.K. Kiehl, S. Banks, E. James, and S. Castles, “Functional and Life Test Data for a Two-Stage Stirling Cycle Mechanical Cryocooler for Space Applications,” in Proc. of STAIF-98, Am. Inst. Phys., CP420 (1998), pp. 199-204. 2. Berry, D., H. Carrington, W.J. Gully, M. Luebbert. and M. Hubbard, “System Test Performance for the Ball Two-Stage Stirling Cycle Cryocooler,” in Proc. of the 9th International Cryocooler Conference, Plenum Press, New York (1966), pp. 69-77. 3.
Gully, W.J., H. Carrington, W.K. Kiehl, and K. Byrne, “A Mechanical Cooler for DualTemperature Applications,” in Proc. of STAIF-98, Am. Inst. Phys., CP420 (1998), pp. 205210.
Progress Towards the Development of a 10K
Closed Cycle Cooler for Space Use A .H. Orlowska, T. W. Bradshaw, S. Scull* and Lt. B. J. Tomlinson# Rutherford Appleton Laboratory Chilton, Didcot, OXON, OX11 OQX, UK *Matra Marconi Space Systems Filton, Bristol, BS12 7YB UK # USAFRL, Kirtland AFB, NM, USA
ABSTRACT
A 10K cooler is being developed for space applications. It is a further extension of a Rutherford Appleton Laboratory (RAL) 20K cooler being space qualified at Matra Marconi
Space Systems (MMS). New features include; optimised geometry, enhanced regenerators and a larger compressor system. This cooler differs from the 20K cooler in many ways. The larger pressure swing and lower
temperatures create problems when trying to model the thermodynamic processes. In this region the regenerator and the working gas heat capacities are strong functions of temperature and this creates difficulties in the modelling. In addition the pressure swing in the cooler is high, leading
to distortions in the pressure waveform in the cold end of the cooler. This paper describes some of the techniques used in developing an optimised geometry for this cooler. Various computer models have been used in conjunction with experimental data. In particular the pressure drop down the cooler has been measured in order to verify the computer model and to ensure that the pressure drop losses and regenerator geometry are correct. The hardware programme leading to a full proof of concept demonstration will be explained. The first hardware test involves the use of four MMS compressors into the standard RAL 20K cooler displacer. This doubles the compression space swept volume over the standard cooler. The outputs from the revised RAL model will be used to define the geometry of the 10K cooler. MMS will perform the detail design, manufacture and proof of concept programme. INTRODUCTION
Demonstration of compact, long term, reliable, and continuous heat lift at 10 K for Very Long Wave Infrared (VLWIR) detectors is a necessary technology for meeting United States Air Force (USAF), United States Department of Defense (DoD), and Ballistic Missile Defense Organizations (BMDO) near and far term imaging requirements. In an effort to support this requirement, the Air Force Research Laboratory (AFRL), with funding provided by BMDO,
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initiated the MMS 10K cryocooler program with Matra Marconi Space. This program will leverage proven MMS cryocooler technology to demonstrate the feasibility of multistage Stirling devices for cooling 10 K strategic space sensors. AFRL initiated an in-house study to determine the cooling requirements of Arsenic doped Silicon infrared detectors for VLWIR sensing. The IR detectors were based on the designs for Si:As BIB arrays. The Cryogenic Technologies Group at AFRL consolidated the technical input from industry, the Aerospace Corporation, and AFRL on detectors, focal planes, and thermal issues. The thermal sensor model, created by Aerospace Corporation personnel, is based on a parametric/generic analysis of an existing optics / cold cavity / focal plane design. Based on the D* focal plane sensitivity curve for these Si:As detectors, the adequate sensitivity, before the sensitivity begins to fall dramatically, is at 11 K. This equates to a focal plane temperature of 11 K +/-0.1 K. The AFRL thermal model assumes 10.9 K at the focal plane and calculated a 0.045 W heat load and a 0.7 K temperature drop through the detector, motherboard, base, and strap (with no margin). This provides the requirement for the Matra Marconi Space 10 K cryocooler of 0.045 W at 10.2 K at the cryocooler cold head. In 1996 the RAL two stage cooler achieved a base temperature of about 12 K, with approximately 100 mw of cooling power at This Stirling Cycle cooler consists of two compressors and a two stage displacer and consumes about 80 W input power. We now aim to improve this performance until a similar cooler can provide about 50 mW at around 10.3 K. The major problem that prevents a regenerative cycle from reaching very low temperatures is
one of regenerator heat capacity. As the temperature falls, the density of the gas passing through the regenerator increases, and hence the heat flow into the regenerator also rises. Unfortunately the heat capacity of most conventional regenerator materials falls as the temperature decreases. The second problem of low temperatures in any refrigeration system is one of efficiency. The maximum Carnot coefficient of performance at 20 K is 5.66%, and at 10 K it is 3.44 %. Typically a small Stirling cycle cooler would only operate at 2% of Carnot at 16 K and even less at 10 K. This implies that the input power required to obtain say 100 mw at 10 K would be well over 150 W, considerably greater than the maximum power presently used (around 100 W for two compressors). This would involve a redesign of the existing compressors, or the use of 4 instead of 2 compressors. It was agreed at the outset that the fundamental problem to be solved by this programme was the cold finger geometry optimisation and the regenerator composition and materials . In order to provide the increased pressure swing required for this cooler at minimum risk, four MMS compressors, based upon the qualified 50-80K cooler design but with 22mm pistons in place of the standard 20mm. This compressor system has not been optimised but provides a high reliability development tool for this work. Preliminary modelling of such a cooler indicated that 10 K could be reached if the pressure swing could be increased and the regenerator losses reduced. This paper provides more detailed modelling and presents the early results of running a cooler with four compressors. COMPUTER MODELLING
RAL has developed successful computer models of both the single stage and two stage Stirling cycle coolers2. These models assume adiabatic compression and expansion and include the effects of regenerator and shuttle heat transfer losses and phase shifts due to the finite speed of sound. Pressure drop is included in the single stage model but problems with convergence had precluded inclusion in the two stage model. This is modelled as a perturbation on the ideal cycle. In order to improve the model we have measured the pressure at the cold end of an operating cooler and used this data to correlate the model so that the effect of pressure drop through the regenerator could be included. This is likely to be particularly important in a cooler running with four compressors as the pressure swing under such conditions is very high.
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Figure 1 The data from the pressure transducers and position sensors.
The computer model of the two stage cooler has been described in detail elsewhere3. At high temperatures very good agreement with measured performance is obtained; however, at temperatures below 30 K difficulties in regenerator modelling together with ideal gas assumptions, leads to discrepancies between modelled and measured cooling powers although
relative trends can be accurately shown. It will, however, always be more difficult to produce an accurate model for a cooler operating near its base temperature where the cooling power
produced is almost matched by the internal losses that consume it. PRESSURE DROP MEASUREMENTS DOWN THE COOLER
The pressure swing in the cooler was measured in two places - at the cold end of the cooler in the expansion space and at the “T” piece close to the compressors. These results were then compared with the pressure drop calculated in the computer model of the cooler. The measured data from the experiment is shown in Figure 1. In this figure the position sensor outputs are referenced to the second Y axis. The pressure transducer outputs are on the primary axis. Note that the apparent dc offset in the displacer stroke is a function of the measuring equipment and that the displacer was in fact operating around the zero position. The extra dead volume from the pressure transducer located at the cold end was also added to the cold end dead volume. The effect of this extra dead volume in the cold end is to limit the pressure swing and increase the mass flow through the displacer. This reduces the gross cooling power and increases the regenerator loss. The base temperature was therefore limited to around 30K. The results of the calculations and the experimental data are shown in Figure 2. The calculated results were factored by 85% to more closely agree with the warm end pressure swing. The, reasons for this are discussed below :
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It was noticed after the results were taken that the average pressure on the transducers was significantly less than the fill pressure. The mean DC level was found to be 8.6 - 8.73 bar whilst the fill pressure was 9 bar abs. This is due to a pumping action on the Stirling cycle compressor pistons. This is a reduction in the mean pressure by approximately 5%. Unfortunately only the AC pressure swings were recorded on the oscilloscope traces.
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At these frequencies the clearance seals on the compressors are slightly leaky and the pressure swing is not at its maximum for a given stroke.
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Figure 2 Measured and calculated results.
A further source of error is that the laboratory control electronics is not capable of driving the moving parts of the cooler in a purely sinusoidal manner. This generates distortions in the pressure waveform. In particular, the position sensor output on the compressor is generally slightly “peaky” with a broader negative region. This may account for the discrepancy between the calculated and measured traces in the negative region. Despite this the model does show some of the finer features observed in the experimental results which suggests that the model is a reasonable approximation to reality. The absolute values underestimate the pressure drop by approximately a factor of two and the pressure swing by 18%. The two stage cooler has two regenerators, the warm regenerator acting between room temperature and the mid stage and the cold regenerator acting between the mid stage and the cold tip. The calculated pressure drop down each section of the warm regenerator is shown in Figure 3. The regenerator is graded with different porosity and/or materials. The section labelled “A” is at the warm end of the regenerator and accounts for much of the pressure drop. These sections are not equal length - sections A and B account for approximately 30% of the length each with C,D and E making up the remainder. The parameters for this calculation are the same as earlier i.e. with the pressure transducer on the cold end. These results suggest that there may be a case for coarsening the mesh in the warm part of the cold regenerator. The experimental results suggest that the computer programme is underestimating the pressure drops along the cooler so the problem may well be more serious than the calculations suggest. RESULTS OF FURTHER MODELLING The model was used to refine the geometry for a four compressors cooler designed to produce significant cooling at about 10 K. The current displacer was used as a baseline. The model gives good agreement at the lowest temperature but overestimates the cooling power at higher temperatures. This could be due to problems with heat transfer for high heat lifts. At the present time, since we are interested in low heat lifts at low temperatures where the model and data are in good agreement, the model has been used for geometry optimisation. The program has been found to correctly predict the optimum phase for the present geometry. The optimum working phase between the compressor and the displacer has been found to be about 30° from both model (Figure 4) and data. This phase has been used in the remainder of the modelling.
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Figure 3 The calculated pressure drop down each section of the cold regenerator.
The geometry of the displacer can be modified by changing the cold and warm displacer diameters and the displacer length. The cold displacer diameter was modified first, keeping the warm displacer diameter at 14 mm. The results from this analysis is shown in Figure 5. It can be seen that the cooling power appears to increase as the cold displacer diameter rises,
but that this is most pronounced at the higher temperatures. The slopes of the load lines indicate that at the lowest temperatures a cold displacer diameter of 10 or 11 mm would give best results. The warm displacer diameter must be increased also in order to give cooling power at the middle
stage to enable it to reach at least 170 K. Although the effect of increasing the warm displacer diameter appears to be to reduce the cooling power at the cold end, this is necessary in order to increase the cooling power at the midstage and lower its temperature. The influence of the warm displacer diameter on the cooling power at the cold tip is shown in Figure 6.
Figure 4 Effect of phase on modelled performance.
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Figure 5 The effect of cold displacer diameter on modelled performance.
The previous modelling was carried out at 7.5 mm compressor strokes. If the strokes are increase to 9 mm the performance is substantially improved (Figure 7). The remainder of the modelling, varying the cold regenerator length, has been carried out at 11mm cold regenerator diameter and 15mm warm regenerator diameter. The warm regenerator diameter was chosen as a compromise to retain the cooling power at the mid stage whilst ensuring the cold end performance. Figure 8 shows that the performance improves as the regenerator is lengthened. The main limitation on the lengthening is not given by the model; it is mechanical. The optimum regenerator length will be the longest practical, commensurate with the difficulties of manufacture and strength and robustness in use. The model is probably overestimating the performance at temperatures below 10 K due to regenerator losses; it is unlikely that the base temperature would be below 8K.
Figure 6 The effect of warm displacer diameter on modelled performance.
DEVELOPMENT OF 10K COOLER FOR SPACE USE
Figure 7 The effect of compressors stroke on modelled performance.
Figure 8 The effect of increasing the cold regenerator length on modelled performance.
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Figure 9 The forces acting on the standard displacer from 2 compressors.
DISPLACER MOTOR MODELING One of the problems anticipated in a cooler with such large pressure swings was the sizing of
the displacer drive motor. The forces acting on the displacer have been analysed and compared with the capabilities of the existing design of displacer drive motor. These are shown in Figure 9. The forces are from the pressure drop along the regenerators Pdrop, from the pneumatic driving force on the displacer shaft Pneu, from the motion and the diaphragm spring suspension system. Changing to four compressors (Figure 10) increases the gas flow and the subsequent increase in
the force from the pressure drop along the regenerators, Pdrop, is marked - it has increased by over a factor of two from the standard cooler leading to a swing away from a balanced system.
Figure 10 Results for
mm compressor strokes.
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Figure 11 The forces on the displacer for various compression space swept volumes.
Results as a Function of Compression Space Swept Volume The calculations for the total force on the displacer drive motor were repeated for different compressor swept volumes and these are presented in Figure 11. The peak to peak force required is in the region of 6N for the worst case.
The force on the displacer for various phase angles The force on the displacer depends on the phase angle between the compressor and the displacer. Some of the forces depend on the motion of the compressor piston and other on the motion of the displacer so the phase relationship between these values will also vary. The total force on the displacer for various phase angles is shown in Figure 12. This plot is interesting as the resultant force on the displacer has a lower harmonic content at high compression space swept volumes
than at lower.
Figure 12 The force for various phases.
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At the lower phases used for the performance modelling the total force is about 2.5 N. The existing displacer magnetic circuit and coil have now been modelled to calculated the force that can be accommodated by the motor. The present design will give a force of 3.7 N for a current density of 5.12 amps/sq.mm,
which is equivalent to 41.5 ampere turns in the existing coil cross section. Since the coil has about 100 turns (neglecting those which fall outside the main field) this is a current of 0.4 A. This current is small and implies that there will be no problem driving the displacer with the existing motor. REGENERATOR DEVELOPMENT RAL has been working on regenerator improvements specifically for work at low temperatures. Much has been published in the last few years on the improvements to the performance of GM machines when rare earth materials with magnetic transitions at low
temperatures are used in the regenerator. These materials can have large specific heat anomalies as the transitions are reached. Many of these materials are brittle, and processing is problematic. GM machines operate at high pressures and low frequencies compared to Stirling machines so pressure drop is not as significant and sphere regenerators are usually used. We are experimenting with different ways of packing materials to find optimum regenerators. The standard regenerator allowed the cooler to reach 10.85 K, with 100 mW at 15.8 K and 200 mW at 20 K. (no attempt was made to optimise the operating conditions to maximise the heat lift at the higher temperatures) FURTHER WORK Work is continuing at RAL in the optimisation of the regenerator. Meanwhile, MMS are taking the output of the RAL cold finger optimisation modelling and designing a new displacer/momentum balancer assembly based upon the 20K unit under qualification for the FIRST/PLANCK programme. The final deliverable prototype cooler will thus comprise a displacer with optimised cold
finger and regenerator assembled to the four off compressors used to date in the development together with laboratory standard cooler drive electronics. Throughout the development the hardware build has remained close to the existing space qualified design and heritage, thus giving a high degree of confidence in the ability to achieve a flight qualified cooler with minimum risk. The programme is scheduled for completion during 1999. REFERENCES 1. T W. Bradshaw, A. H. Orlowska, C. Jewell, B. G. Jones and S. Scull, “Improvements to the cooling power of a space qualified two-stage Stirling cycle cooler”, presented at the 9th International
Cryocooler Conference, New Hampshire, USA, (1996), “Cryocoolers 9”, Plenum, R G Ross (ed), p 79-88.
2. T W Bradshaw, A H Orlowska and J Hieatt, “Computer Modelling of Stirling Cycle Coolers”, Proceedings of the 7th International Cryocooler Conference, 17-19th November 1992, Santa Fe, New Mexico, US, PL-CP--93-1001, p621. 3. TW Bradshaw, SFJ Read and DA Cragg, “A Design Study on a Multi-Stage Stirling Cycle Cooler for Space Applications”, 1986, a report issued under ESA contact number 6342/85/NL/PB(SC).
Development of a Light Weight Linear Drive Cryocooler for Cryogenically Cooled Solid State Laser Systems L. Barry Penswick
Brian P. Hoden
Stirling Technology Company Kennewick, WA, USA 99336
Decade Optical Systems, Inc. Albuquerque, NM, USA 87109
ABSTRACT
This paper describes the development and testing of a free piston, linear drive cryocooler designed to meet specific requirements for a light weight, airborne, cryogenically cooled
semiconductor laser system. The function of the laser and how the cryocooler is used in the laser will be described. The basic cryocooler is an in-line configuration which was initially developed to support
high temperature super conductor (HTS) device cooling. Because of its relatively high capacity (8 Watts at 65 K) and high degree of modularity, this unit was selected as the baseline for the significant weight reduction process which was required to meet laser system weight constraints. The cryocooler employs a moving iron linear drive motor and flexurally supported moving components to eliminate wear as potential cause for performance degradation. The weight reduction process started with a 60 pound cooler and reduced the weight to 30 pounds without changing its performance significantly. A second phase of this program is currently ongoing to reduce the weight to 20 pounds. Finite element analysis was used to optimize the weight and strength.
A description of the cryocooler operating requirements along with mechanical details of the gas compressor, cold head, heat rejection system, and supporting electronics is provided. Issues concerning the hardware changes required in the weight reduction program and their impact on cooler operating characteristics are discussed. A comparison between the performance of the InLine cooler and the light weight version is presented. INTRODUCTION
Decade Optical Systems, Inc. (DOS, Inc.) is a R&D and production company for ruggedized lasers. DOS, Inc. packaged an Optically Pumped Semiconductor Laser (OPSL) in 1995 for field
test applications operating in the mid-IR wavelength band. This laser incorporated a Sunpower Stirling cryocooler. In 1997, DOS, Inc. was awarded a contract to further develop the OPSL to be lighter, more powerful, and consume less input power to be used on a rotary wing aircraft. In order to maintain the semiconductor laser material at 80 K during lasing, a 16 Watt heat lift cryocooler was needed. To meet a power budget of 310 Watts and a weight of 20 pounds, a stirling cycle cryocooler was selected as the baseline. Stirling Technology Company (STC)
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produced a cooler in this category except the weight was 60 pounds. Through a joint effort between STC and DOS, Inc., this cryocooler was reduced in weight without changing the performance. LASER DESIGN AND CRYOCOOLER INTERFACE
The semiconductor laser specifications (complete package, including power supplies and cooling) which impacts the cryocooler design include: 40 lbs., 400 W input power, operating temperature of 0°C to 50°C, 16 Watt heat lift at 80 K, and 3.5 minute cool down time. The laser without the cryocooler weighs 20 pounds and consumes 90 Watts during cool down. This leaves 20 pounds and 310 Watts of input power for the cryocooler. The laser design is extremely aggressive and requires dense packing of all of its components. The cooling system is a series of customized heat exchangers and custom computerized controls. In order to lightweight the laser, a custom carbon epoxy/honeycomb structure and housing has been designed. A compact vacuum chamber with a bellows is used to isolate the vibration when attached to the cryocooler. Figure 1 shows the configuration of the laser with the STC cryocooler. In order for the laser to produce the correct mid-IR wavelength, semiconductor laser material developed by MIT Lincoln Labs was used. This material (AlGaSb) is optically pumped in the near-IR with a diode laser array. The laser material operates most efficiently at 80 K or colder. The heat load generated at the laser material interface with the cryocooler cold finger is about 20 Watts at 80 K. The laser run time is short and has a cool down time period between runs; therefore, a thermal mass has been designed to connect to the cold finger that stores the heat during the run time. The 3.5 minute cool down from 30 K to 80 K required for the cryocooler is a function of the thermal mass which needs to be between 90 to 120 grams. This weight requirement lead to extensive finite element analysis and design of the cryocooler cold finger.
Figure 1. Semiconductor Laser Design.
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Cryocooler Description
The In-Line free piston, linear drive cryocooler employed in the current program is a member of a cryocooler family, which has been under development by STC for commercial applications. These units are characterized by relatively high cooling capacities (10 to 25 Watts
at 80 K), the use of moving iron linear drive motors, flexural bearings used throughout, and a highly modular mechanical configuration which allows the basic components to be arranged in a number of different ways to quickly meet end user requirements. The “standard” In-Line cryocooler is shown in Figure 2 and represented the starting point for the low weight, laser cooling system. Cooling capacity at various cold head temperatures are noted in Figure 3 for drive power levels up to the 310 Watt limit dictated by the laser system power budget. The following sections provide further details on the In-Line cooler and discuss the conversion process from the standard In-Line to the lightweight laser cooling system. Hardware Description and Testing As previously noted, the laser has a number of weight, size, and power requirements which play an important role in the transformation of the In-Line cooler hardware to meet the laser cooling goals.
Linear Drive Motor. The linear drive motor represents the heaviest component within the cryocooler and was the focus of the initial weight reduction effort. The moving iron linear motor employed in STC’s cryocoolers is unique in the sense that it is made up of a fundamental element that can be “stacked” to provide the necessary drive power to the cooler. This basic process is shown schematically in Figure 4. An individual stack contains the stationary copper
Figure 2. In-Line Cryocooler Configuration.
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Figure 3. Cooling Capacity vs. Cold Head Temperatures (Electrical Input Power < 310 Watts).
coils, high strength permanent magnets, and the stator laminations. The moving part of the motor is made up of laminations, which complete the magnet circuit. When combined into a multi stack configuration, only the copper coils need to be changed allowing considerable design flexibility. Based on the low electrical input power constraints, a 4-stack version of the linear motor was utilized rather than the five stack in the standard In-Line unit. A number of the moving components were also modified to minimize reciprocating weight. This in turn allowed the number of flexures to be reduced. Due to the cool down time requirements, it was felt that the system should be capable of operating over a range of drive frequencies so as to potentially maximize cooling capacity at various temperatures while retaining a reasonable linear motor efficiency. Modeling of the system during cool down conditions indicated that the potential frequency range was between 48 and 52 Hz; therefore, the motor was “tuned” for a nominal 50 Hz operating point.
Figure 4. Schematic of Moving Iron Linear Motor Configuration.
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Compressor Assembly. The compressor assembly is made up of the linear drive motor, the power piston and its respective cylinder housing, the piston drive rod, flexures, and the pressure vessel, which encloses all of these components. To minimize the compressor assembly weight, the primary focus was on the power piston housing and the rear pressure vessel. The In-Line
piston housing and pressure vessel was redesigned to minimize weight. Extensive FEA was performed by DOS, Inc. to optimize the weight to strength based on a 375 psi internal pressure.
The results are shown in Figure 5. The rear pressure vessel was replaced with a thin walled pressure vessel fabricated from high strength aluminum. A portion of the pressure vessel was integral with the piston housing, with the final close out carried out by welding the pressure vessel end cap to the cylinder. Figure 6 compares the standard In-Line system compressor assembly with the equivalent component for the lightweight version. Cold Head Assembly. Because there are interactions between the cold head and the cooled laser components, it was necessary that the cycle’s helium side heat exchanger be mechanically and thermally compatible with the thermal “coupling” system employed. Simultaneously, the mass of the heat exchanger had to be minimized so as to reduce system cool down time. This combination of requirements led to a flat cold head heat exchanger that incorporated a series of radial flow passages for the helium. The portion of the heat exchanger, which is coupled to the cooled laser components, is fabricated from OFHC copper and contains the radial fins, which form the helium flow passages. The copper section is vacuum brazed to a stainless steel backing plate which provides a significant portion of the cold head mechanical strength and alignment to the displacer cylinder liner.
Figure 5. Results of Structural Analysis.
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Figure 6. Standard In-Line and SLS Compressor Assemblies.
Controller. STC modified its standard cooler control system to meet the rapid cool down requirements. The controller scheme incorporates Hall effect sensors mounted externally to define the positions of the power piston and displacer. Assuming both values are within their allowable range and the cold head is warmer than desired, the controller commands the power supply to increase the voltage applied to the motor, increasing piston amplitude, and in turn displacer amplitude. This voltage increase continues until the displacer amplitude limit is reached or the cold head is at the desired temperature. In the latter case, a simple PID controller
is utilized to maintain a fixed cold head temperature. The STC power supply and controller weighs 45 pounds. DOS, Inc. redesigned and repackaged this power supply/controller and reduced the weight to two pounds. The controller was designed to vary the frequency as a function of temperature to utilize the maximum amount of power available. Cryocooler Unit. Figure 7 shows the final configuration of the cryocooler that evolved
from the process described above. Figure 8 shows a comparison of the In-Line cryocooler to the 30 pound cooler. Test Results. A series of tests were performed to fully characterize the cooler system.
These included both steady state and rapid cool down measurements and are briefly discussed below.
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Figure 7. Cryocooler Configuration.
Operation of the cooler at various operating frequencies is shown in Figure 9. As can be clearly seen, the input power levels for specific cold head temperatures with different operating frequencies vary dramatically. A comparison of the cool down times for the 30 pound and the In-Line (60 pound) cryocooler is shown in Figure 10. The longer cool down times are due to the inherent thermal masses; however, the cool down time for 116 gram thermal mass matches for both of the coolers which proves that the lightweighting did not effect the performance. Further testing is currently underway to develop a controller algorithm that provides optimal cooling conditions during the cool down process. The cooling capacity of the unit as a function of cold head temperature while operating in the displacer limited condition is depicted in Figure 11. As can be seen the cooling capacity starts out at a very high value and falls continuously as the temperature is reduced. It is
Figure 8. Comparison of the In-Line With the 30 Pound Cryocooler.
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Figure 9. Cooler Operating Characteristic’s at Various Operating Frequencies.
Figure 10. 30 Pound vs. In-Line Cryocooler Cool Down.
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Figure 11. Net Cooling Capacity vs. Cold Head Temperature.
important to note that the drop off in capacity shown in Figure 11 is highly dependent on the specific operating conditions of the cooler since the dynamic characteristics of the free piston / displacer combination will define the power available to the Stirling cycle and in turn the cooling capacity. Operation at other frequencies or charge pressures can result in significantly different cooling capacities at various cold head temperatures. The steady state cooling capacity values fall very close to those of the standard In-Line cooler system, clearly indicating that the modifications to the critical cold head assembly had little if any impact on cooling capacity.
CONCLUSIONS
A significant physical size and weight reduction effort was performed on the In-Line cryocooler without compromising cryocooler performance to meet the laser program requirements. The success of this program was based on exploiting the unique modular nature of the basic cryocooler configuration. The cryocooler portion of the system is currently under going further refinements to meet the final system requirements. Key to this effort is a further reduction in the weight of the unit to approximately 20 pounds while maintaining the same cooling capacity within the existing power budget. The design of this system has been completed and is in fabrication. Testing is planned for the third quarter of 1998. STC is actively investigating a number of alternative commercial applications for this unit or its variants.
Low-Weight and Long-Life 65K Cooler V.T.Arkhipov, V.N.Lubchenko, L.V.Povstyany
Special Research and Development Bureau for Cryogenics Technologies Kharkov, Ukraine
H. Stears Orbita Ltd. Kensington, MD, USA
ABSTRACT
The Research and Development Bureau at the B. Verkin Institute for Low Temperature Physics & Engineering, in a joint project with Orbita Ltd, USA, has created a flexible cryogenic refrigeration system for use in continuously operating equipment. The refrigeration system uses a thermoaccumulator in combination with an intermittently active Stirling cryocooler. The thermoaccumulator is a cryostat with an interior temperature of maintained by the solidification of nitrogen. The cryocooler, a linear drive split Stirling design, periodically refreezes the liquid nitrogen. The temperature is controlled by a closed-loop control system included in the package. The system operates from a variety of power sources, including 110V, 60 Hz AC, 110V, 50 Hz AC, and 27 volts DC. INTRODUCTION The most up-to-date, effective, and widely used cryocoolers for use in the 65 to 70 K tem-
perature range are Stirling-cycle machines. Such machines are manufactured in both single- and two-stage versions. In many applications the cooled object is mounted directly upon the coldtip of the cryocooler displacer. Such a scheme often minimizes the parasitic heat load. On the other hand, with such an intimate interface, the cooled object can suffer from cooler-generated vibration and electromagnetic interference. Also, Stirling cryocoolers are somewhat unreliable, and
failures occurring within the units will invariably cause severe and unallowable malfunctions of the entire system. Backup coolers are sometimes used to address such risks, although such a decision is not always successful. The main issue is the additional thermal load due to the thermal
conduction through the coldfinger of the inactive cooler. In addition, a backup cooler requires a prearranged “T” scheme, which adds complexity and installation time. Thus, the use of a backup cooler increases the system costs without a guarantee of improved system reliability. The most efficient Stirling cryocoolers are often integral configurations with a crankshaft
drive that ensures proper piston motion and optimal phase angle between compressor and displacer. Similar efficiency levels can be obtained with Stirling coolers using electronically-controlled linear drives in which the phase shift is maintained by closed-loop electronic control. This
is the technique used in Oxford-style Stirling coolers.2,3,6 Cryocoolers 10, edited by R. G. Ross, Jr. Kluwer Academic/Plenum Publishers, 1999
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Crankshaft driven coolers are unbalanced, but thanks to miniaturization, it is possible to reduce vibration to levels that can be tolerated by typical cooled devices. Since the compressor causes most of the vibration, it is advantageous to use a split-Stirling configuration with the compressor separated from the sensitive cooled load. The next step may be the incorporation of a pulse tube expander. Pulse tubes seem to be highly reliable, with minimum induced interference, yet operate with high efficiency. However, another solution for the stable maintenance of cryogenic temperatures is the use of a system consisting of a thermoaccumulator mounted between a cryocooler and the cooled object. A thermoaccumulator utilizing a phase transition such as the heat of melting can act as a buffer and maintain the system temperature during routine or emergency cryocooler shutdowns. For the case where redundant cryocoolers are utilized, the buffer also provides the time-reserve necessary for actuation of the standby cooler, or even a second standby device. Being of a significant weight, thermoaccumulators also neutralize cryocooler vibration at the cold-load interface. Furthermore, since the thermoaccumulator provides temporary heat storage, the cryocooler can be shut off entirely for periods of time to allow a series of high precision measurements to be made. Although it is especially attractive to combine thermoaccumulators with Joule-Thomson (J-T) type cryocoolers, J-T systems are often not as efficient below 80K as Stirling machines. This has led to the system design described in this paper. An earlier work5 discusses a thermoaccumulator in a system in which the cryocooler’s refrigeration capacity is less than the maximum load. This report describes a system where the thermoaccumulator is designed for relatively low load rates. The design considered here is mainly for ground applications, although it can be modified for space use. STRUCTURAL FEATURES
The cryogenic refrigeration system under consideration has three subsystems, a thermoaccumulator (TA) with an expanding tank-receiver, a split-Stirling type cryocooler, and a power conditioning module that operates the cooler. A schematic of the system is shown in Fig. 1. In this system the cooled object is mounted onto the bottom of the thermoaccumulator. The thermoaccumulator is an efficient nitrogen dewar, with multilayer insulation located within its vacuum walls. There is a provision for initial pumpout, and an absorption type cryopump is provided to maintain the vacuum while cold. Externally, the nitrogen dewar is attached to a
Figure 1.
A scheme of the cryocooler: 1-TA-nitrogen vessel; 2-pump-off valve; 3-payload heat exchanger; 4- cooled object; 5-tank-receiver; 6-control device; 7-window payload; 8-temperature sensors;
9-electric power supply device; 10-air fan; 11-filling valve; 12- compressor unit; 13- coolerheat exchanger; 14-displacer GCM; 15- heat exchanger for accelerated freezing TA; 16filling valve.
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storage unit, and a filling valve is provided for the initial charging. The nitrogen dewar is also connected to the coldtip of the Stirling cooler at the payload heat exchanger, described later. Two thermometers, one at this interface, and one at the load, are used for system temperature control. The split-Stirling machine consists of two parts, an expander unit and a dual opposed compressor; they are interconnected by a flexible transfer line. The cryocooler is filled with helium and sealed off with a valve. The system includes a fan for cooling the compressor, the ambient cryocooler heat exchanger interface, and the power conditioning module. The power conditioning module includes the control system that operates the cryocooler. When the refrigeration system is turned on, the Stirling compressors provide a pressure wave to the expander assembly, and the system begins to cool. As the thermoaccumulator volume cools, its nitrogen charge first cools, then condenses, and then solidifies. The pressure in the nitrogen system drops from 2-3 MPa to 94 torr (the triple point pressure of nitrogen at 63.15 K). At this point the cryocooler shuts off to prevent the complete solidification of the nitrogen and thereby preserves the temperature regulating properties of the thermoaccumulator. After a predetermined interval, during which the solid nitrogen has melted to some degree to satisfy the external heat load, the cryocooler starts up and begins to cool the system again, and the cycle repeats. When the refrigeration system is shut off permanently, the system warms up completely, and the nitrogen returns to the external storage container. The key features of the cyclic operation are illustrated in Fig. 2. The figure displays the duty cycle of the cryocooler, the fraction of solid N2 in the accumulator, the load temperature, and the nitrogen pressure as a function of time over a complete operational cycle of the system. Two variants are shown (designated by continuous and dotted lines, respectively), one for a routine cooldown, and one for an initial accelerated cooldown where the gaseous nitrogen is precooled by an external liquid nitrogen source which relieves the cryocooler of the task of initially condensing the liquid.
The duty cycle of the cryocooler affects the lifetime of the system, for the lifetime is proportional to the operating time of the cryocooler. A compressor cycle-operation lasts for: where and are, respectively, the cryocooler pause duration time and the time increment required for the cryocooler to recover to the setpoint temperature level; and are the heat inflows to the TA-working volume during the cryocooler switched-off, and during the cryocooler switched-on periods, respectfully.
Figure 2. A cryocooler-performance cyclogram.
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A similar ratio helps to quantify the potential lifetime of the system as a whole: where is the cryocooler operational lifetime. It’s apparent from Equations (1 & 2) that to increase the lifetime of the system, it is expedient to use a cryocooler with a large cooling capacity, and to extend both the cryocooler operating (switched-on) time period and the cryocooler non-operating (switched-off) time period. This will naturally result in a need to increase the mass of the thermoaccumulation substance and, hence, the weight, dimensions, and consumed power of the cryocooler as well. The load and timeduration adjustments are made by control-panel manual setups for the compressor switched-on and switched-off time periods. Typical computed parameters for the system described in this paper include: • Cryocooler heat capability: 4 W • Parasitic lead load of TA-cryostat: 0.4 W • Net useful continuous load capability: 0.1 to 0.5 W In order to provide for 3 years operational lifetime, the required longevity for the cryocooler is estimated as 5,500 to 9,500 hours for a cooler with a 3 W capacity, and 4,000 to 7,000 hours for a cooler with a 4 W capacity. THERMOACCUMULATOR
The thermoaccumulator subsystem is composed of the cryostat proper (represented by the thermal melting-based accumulator), the receiver, the connecting pipeline, the heat exchanger, and accessories. It is designed for accumulation and storage of the solid-body cooling agent (such as nitrogen), along with direct cooling of a cryogenic load. The thermoaccumulator subsystem described in this paper has the following parameters or features: • Working substance: nitrogen (or argon or gas-mixtures) • Effective volume : • Operating pressure: to 4.5 MPa (as a pilot-pressure) • Heat ejection power: 0.1 to 3 W • Parasitic thermal load: • Unit weight: Figure 3 displays the overall dimensions of the thermoaccumulator in both longitudinal and cross-sections.
Figure 3. Design of a thermal accumulator (TA).
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The nitrogen-filled cryostat volume is a cylindrical (toroidal) shape vessel made of stainless steel with a series of copper-foil ribs inside the cylinder to provide for temperature stabilization, and for effective nitrogen-freezing during performance of the refrigeration system. The bottom side of the nitrogen vessel is provided with a special contact area for mounting of the cooled object (Fig. 3, -3). The nitrogen vessel is fixed to the top lid (-5) of the unit body (made also of stainless steel) by means of a fiberglass tubular element (-4). The upper lid of the cryostat body is also furnished with: • Nitrogen input / output tube • Pump-out valve (-7) for pumping the guard vacuum space of the TA-unit • A pressure-sensor inside the guard vacuum space (for pilot versions of the TA-unit) • A mounting device for the Stirling cryocooler displacer (-8) • Two electrical connectors for connection of temperature sensors and control heaters
On the bottom side of the unit body, there is attached a removable hatch-window (-10) used to install the cooled object and optical filters. The removable hatch is furnished with an electric receptacle (-11) for connecting cooled-object signal wires. The unit body is provided with a security-membrane element (-12). Specific features taken into consideration within the design include: • Low parasitic heat influx levels achieved through the use of optimized materials for thermal insulation and support elements, and to the use of a long filling tube • A large contact area for mounting various cooling objects • Structural design and materials for the TA-core to provide for cooled-object temperature stability during the cyclic operation of the Stirling cryocooler • The thermoaccumulator design enables the possibility of exchanging the cryocooler displacer with a redundant unit in the case of a failure
The task of providing high-performance characteristics for the thermoaccumulator is complicated by the small size of the cryostat. Heat inflows from the cryocooler displacer (during shutoff) comprise over half of the thermal load from the structure (0.23 W). Size restrictions do not permit further reduction of the heat inflows by means of heat-bridge length increases. The tubular support structure has been accepted as the most appropriate to achieve rigidity of the cooled-object and to simplify the cryostat assembly process. If it is required to further increase the non-operating time of the cryocooler, it would be necessary to enlarge the mass of the TA-working substance and to enlarge the volume of the nitrogen vessel. Yet these procedures do not result in a considerable increase in parasitic heat inflows. For this reason it is better to utilize a larger volume thermoaccumulator (unless weight or size limits are specified), and mostly due to the time of the cryocooler initial cooldown. Similarly, with respect to the cryocooler cooling capacity, the cooling load of initial cooldown for the thermoaccumulating substance and structure from 300K to 63K is large compared to that of the active in-cycle operation with the steady state cryogenic load. High stability of the cryogenic temperature is achieved by means of intensive heat-mass transfer inside the cryostat. Here, the structure also ensures: • Temperature leveling within the solid-phase volume obtained by cryopumping • Relatively thin solid-phase thicknesses upon cooled surfaces resulting from the desublimation process • Good liquid-phase contact that maintains the triple-point temperature against contact surfaces of cold-duct and cooled object To reduce the initial cooling time, a heat exchanger is provided to preliminary and rapidly cool down the gas that is coming to the TA from the receiver. The cooldown is achieved using liquid nitrogen, which actually liquefies the nitrogen working gas flowing through the heat exchanger; recall that the working gas is initially under higher pressure.
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The major conditions for normal operation of the TA are:
• High level of purity of the working gas (no water or other contaminant gases) • Fully leak-tight TA-structure • High vacuum available within the cryostat guard-vacuum insulation volume SPLIT-STIRLING TYPE CRYOCOOLER
The low-cost split Stirling-cycle cryocooler was specifically designed to operate with the thermoaccumulator in our proposed system. The primary concerns were reliability, service life, and maintaining system operation up to 10,000 hours. Another important consideration was to configure the cryocooler in order to shield the thermoaccumulator and the cooled object from the cooler’s undesirable EMI and vibration attributes as much as possible. To minimize vibration, a dual opposed linear compressor was selected and used in a split configuration, remotely located from the load. The free-piston expander was also divided into two parts, a displacer unit, and a detached regenerator, as shown in Fig. 4. Details of the cryocooler design are summarized in Table 1. Materials for pistons, cylinders, and magneto-ducts fabrication have been selected with conforming linear-expansion thermal coefficients.
Figure 4. Cross-sectional view of the Stirling cooler displacer unit: 1- flange, 2- displacer piston, 3- heat
exchanger payload, 4- regenerator, 5- transducer of piston position, 6- delivery pipe.
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Another important design consideration of the Stirling cooler design was achieving minimal off-state parasitic loads. It is understood that this split configuration may be a little less thermally efficient than a more conventional configuration with the regenerator located within the displacer. However, it is felt that this is offset in the present system, where the cooler is run intermittently, by the reduction in the displacer off-state conduction achieved by this separation. In the present system, the accumulator is expected to have a heat leak of 0.1 to 0.2 watts, with the load itself of 0.1 to 0.5 watts. In contrast, an integral displacer regenerator would have an off-conduction of 1 to 1.5 watts. Clearly, the isolation is important. We predict that the cooler, as configured, will lift 4.0 watts at 62 K. By only mounting the regenerator to the thermoaccumulator at the cold heat exchanger, a high level of mechanical and thermal de-coupling was also achieved from the displacer because of the interconnecting tubing. The regenerator and displacer are, however, mounted to a common flange at the warm end to simplify independent testing and integration into the load. In continuous operation, the split configuration will be a little less power efficient, but power efficiency is not paramount for the ground applications envisioned. Finally, it is expected that the thermoaccumulator will provide for the necessary temperature regulation of the load. POWER SUPPLY AND CONTROL BLOCKS
The electronic controller circuit is designed to be operated from a 27 volt DC power source, but includes a built-in converter for rectifying AC so that it can be powered either from 110 volt 60 Hz or 127 V 50 Hz. The controller provides the variable frequency commutation drive for the linear compressor electromagnetic coils, which accounts for most of the power draw of the system. An optronic device provides for isolation between the low current and power sections of the controller in order to reduce electromagnetic interference. The unit also contains circuits to drive the various sensors in the system. These include position sensors used to locate the moving armatures, thermometers for monitoring working temperatures, and pressure gauges for monitoring the cryocooler and thermoaccumulator charge pressures. The electronic circuit provides for the long term temperature control of the thermoaccumulator. The controlled temperature is actually that at the interface between the displacer coldfinger and the thermoaccumulator, while the temperature at the load interface reflects the cooled object status. The controller switches the compressor off when the coldfinger temperature starts to drop below the triple point temperature. The compressor then turns back on after a preset delay time chosen by the operator. As long as the cooler turns on before all of the solid has melted, the load will be maintained at the triple point indefinitely. OVERALL DESIGN CONFIGURATION
The components are mounted into a general frame, as shown in Fig. 5; it has overall dimensions of In the version considered here the refrigerated object is inserted through a window on the underside of the cooler. In an actual system, the cryostat itself could be decoupled from the frame and attached to the exterior object to be cooled.
CRYOCOOLER SPECIFICATION
• • • • • •
Working temperature: Useful cold production efficiency: Consumed power: Unit weight: Operational life, expected: Electric power source:
0.1 to 0.5 W 150 W 14 kg over 5 years 110V 60 Hz
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Figure 5. Overall configuration of the complete refrigeration system: 1 - thermoaccumulator; 2- displacer; 3- receiver; 4- compressor unit; 5- power/control package; 6- monitor.
Depending on the Customer’s requirements and operational conditions, the cryocooler may be custom-manufactured with a detached exterior control and monitoring package. This would enable additional control, or, when necessary, adjustment of the compressor switch-on periodicity levels for particular cooling conditions such as seasonal, or day / night temperature contrasts. The store-
receiver could also be located separately. For long continuous operation the receiver could be removed and substituted with a security membrane-valve. Presently the cryocooler unit is undergoing optimization tests for its components. Available test and operational experience to date inspires us to hope for successful commitment through the year. An second variant of this model provides a serially manufactured compressor unit of similar technical characteristics. In conclusion, the authors would like to acknowledge the U.S. Naval Research Laboratory, and Dr. M. Nissenoff in person, for task formulation and useful discussion and inputs on the content of this paper. REFERENCES 1.
2.
Arkhipov, V.T., et al., “Long Life Cryocooler for 84-90K,” Cryocoolers 10, Plenum Press, New York (1999). Ross, R.G., Jr., “JPL Cryocooler Development and Test Program Overview,” Cryocoolers 8, Plenum Press, New York (1995), pp. 173-184.
3.
Jones, B.G, “Development for Space Use of BAe’s Improved Single-stage Stirling Cycle Cooler for Applications in the Range 50-85K,” Cryocoolers 8, Plenum Press, New York (1995), pp. 1-12.
4.
Burt, W.W., et al., “Demonstration of high Performance 35K Pulse Tube Cryocooler,” Cryocoolers 8, Plenum Press, New York (1995), pp. 747-764.
5.
Bugby, D.C., “Development of a 60K Thermal Storage Unit,” Cryocoolers 9, Plenum Press, New
6.
Werrett, S.T., et al., “Development of Small Stirling Cycle Cooler for Space Flight Applications,” Advances in Cryogenic Engineering, vol. 31 (1986), pp. 791-799.
York (1995), pp. 747-764.
Thermal Performance of the Texas Instruments 1-W Linear Drive Cryocooler D. L. Johnson Jet Propulsion Laboratory California Institute of Technology Pasadena, California USA 91109
ABSTRACT The thermal performance of the Texas Instruments (TI) 1 -watt linear drive cryocooler was measured for coldtip temperatures varying from 23 K to 130 K over a range of heat reject temperatures (+50°C to –54°C), input voltage levels (7 Vrms to 10 Vrms), and operating frequencies (50 Hz to 65 Hz). The cooler was driven with a linear amplifier during performance testing to remove the effects of the cooler drive electronics to better understand cooler operation and thermal performance for a variety of environmental conditions that may be encompassed by different users. The wide heat sink temperature span provided a large variation in the refrigeration capacity of the cooler in terms of both the lowest attainable coldtip temperature and the refrigeration capacity at a given coldtip temperature. Over this large span of heat sink temperatures, the lowest achievable coldtip temperature varied from 45 K to 23 K, while the refrigeration capacity at a given temperature changed by as much as 650 mW. As one example, the refrigeration capacity at 77 K varied from 0.82 watts at +50°C to 1.45 watts at -54°C, with corresponding cooler input power levels of 31 watts and 27 watts, respectively. Additional cooler tests were performed using the TI voltage-mode cooler drive electronics and DC-DC voltage converters to measure the electronics efficiencies in a simulated spacecraft operating mode. The cooler drive electronics was observed to operate with a nominal 70% efficiency, while this particular set of voltage converters operated with an nominal 83% efficiency. The efficiency of the electronics plays a large part in determining the overall spacecraft power requirements to operate the cooler as well as determining the thermal dissipation characteristics of the various electrical components. The results of all the performance measurements are presented in this paper. INTRODUCTION There is an increasing demand for low-cost cooling options for focal planes on short duration (1-2 year), low cost, space flight experiments, both on spacecraft that fly in a low earth orbit and on landers and rovers for planetary study. These applications place high demand on both the cooler operating conditions and on the thermal control of the cooler’s environment because of the extreme temperature variations of the ambient environment and because of the limited available power. With the emphasis on low cost, the success of recent flight experiments utilizing the Ricor 506B cryocooler1 and the Texas Instruments (TI) 0.15 W cryocooler2 have piqued the interest of the space science community in the use of these coolers. The Ricor cooler is representative of the rotary drive Cryocoolers 10, edited by R. G. Ross, Jr. Kluwer Academic/Plenum Publishers, 1999
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cooler technology and the TI 0.15 W cooler is representative of the linear drive cooler technology. Coolers of both technology types were originally developed for military applications but are used extensively in commercial applications as well. As a second generation technology developed to overcome the limitations of the rotary cooler design, the linear drive cryocoolers provide all the necessary features for a short duration mission: low cost, low input power, low vibration, light weight, rugged, with a mean-time-to-failure (MTTF) between 3000 and 8000 hours, and are operable over a wide range of heat rejection temperatures (+71°C to –54°C). (These coolers are distinguishable from the linear drive, flexure-bearing Stirling and Pulse Tube cooler technology designed for multi-year life in space-borne applications.) These cryocoolers are ideal for applications where wide temperature extremes are encountered or
where it is desirable to get the most performance out of a cooler by operating at low ambient temperatures. The linear drive cryocoolers are made by a host of manufacturers, but there is little in the way of a data base that describes in detail the performance of the cryocoolers over their entire operating space. The lack of performance data adds to the difficulty of the user community to select the cooler best suited for the prospective application. The focus of this paper is to present the thermal performance capabilities of the TI 1-watt linear drive Stirling-cycle cryocooler over a wide ambient temperature range and cooler operating
parameters to provide a large data base usable for a thermal designer. The performance measurements were made with the cooler driven with a low distortion linear amplifier to remove the effects of the cooler drive electronics. Significant levels of testing of the cooler were made to measure the performance over a wide range of cooler drive frequencies and input voltages, and with applied coldtip loads from 0 watts to 3 watts, covering coldtip temperatures from less than 30 K up to 130 K. Additional testing of the cooler was performed using the TI voltage-mode drive electronics and DCDC voltage converters to measure the efficiency of the electronics in a simulated spacecraft electrical configuration. The results are reported in this paper. COOLER DESCRIPTION
The Texas Instruments 1 -watt linear drive split-Stirling cooler is sized to provide 1 watt of refrigeration at 77 K for a nominal input power into the cooler drive electronics of 32 watts3. The cryocooler and drive electronics can be operated in ambient temperatures ranging from +71°C to –54°C. The MTTF rating as specified by TI is for 4000 hours, but has been reported4 to be at least 5500 hours as an interim MTTF value collected during the course of ongoing endurance tests for coolers operating over a heat sink temperature range between +52°C and –32°C. The dual-opposed linear drive pistons provide minimal piston motion, and the balanced piston pair helps reduce cooler-generated vibration. The transfer line geometry for the cooler is determined by the user; the cooler used in these performance measurements had a transfer line 5 cm long, providing a compact cooler profile with minimal performance loss because of the short transfer line.
Figure 1. Test setup for thermal performance measurements of the cryocooler.
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TEST APPARATUS AND TEST SETUP
The cooler was mounted on a copper heat sink plate supported on stainless steel thin-wall tube legs for thermal isolation from the vacuum chamber. The copper heat sink plate was cooled via a fluid loop connected to a recirculating chiller. The cooler body was supported with aluminum clam shells machined to the curvature of the compressor and expander bodies as shown in Fig. 1; these support clamps provide
the heat transport mechanism to get the heat from the cooler to the heat sink plate. A thin sheet of indium was placed between the cooler and the clam shells to fill any voids. Thermocouples were placed on the compressor, the heat sink plate, and at the top and bottom of the clam shells to measure the temperature gradient in the clam shells. The entire cooler and heat sink plate were wrapped with MLI to minimize radiative losses between the cooler and the chamber wall. The cooler was driven with a sinusoidal voltage waveform from a function generator amplified through a low distortion power amplifier. This permitted the drive frequency and compressor input voltage to be parametrically varied for performance mapping of the cooler. A true rms power meter was used to measure the voltage, current, and power into the compressor. A 20.8-gram heater block, outfitted with a resistive heater and two temperature sensing diodes, was attached to the coldtip. The coldfinger was completely wrapped with MLI to minimize radiation to the coldfinger. THERMAL PERFORMANCE
The cooler was comprehensively tested over a broad range of heat sink temperatures from +50°C to –54°C and for coldtip temperatures ranging from the no-applied load temperature to 130 K,
representing applied coldtip loads from 0 to 3 watts. At each heat sink temperature, the cooler was first operated at the nominal 55-Hz drive frequency with a 10-Vrms input level (roughly 85% of full drive voltage capability) to acquire a load line. Additional 10-Vrms load lines were then measured for
other drive frequencies, varied in 5-Hz increments, to determine the optimal drive frequency. Load line performance data was recorded for the cooler operating at several different input voltage levels. The thermal performance measurements are plotted on multivariable plots to describe the cooler thermal performance dependence on the heat sink temperature, the cooler input voltage, coldtip load, and drive frequency. From these parameters, one can see the net effect on the overall cooler input power and attainable coldtip temperature, and on the cooler specific power (the cooler input power required to provide 1 W of refrigeration) under these operating conditions. The data are presented in the following figures. Drive Voltage. Figures 2-7 show the performance of the cooler as a function of the input voltage for heat sink temperatures varying from +50°C to -54°C. The load line performance data were obtained by running the cooler at several constant input voltage levels, starting from a high of 10 Vrms. (The maximum input voltage to the compressor coils representing full stroke capability was approximately 11.6 Vrms as observed during initial cooldown of the cooler when operating with the cooler drive electronics.) Varying the input voltage between 10 Vrms and 9 Vrms results in an approximate 6 watt change in input power; this corresponds to a nominal change in refrigeration capacity of 150 mW at 50 K and 300 mW at 80 K. Heat sink temperature. The heat sink temperature was varied over a very broad temperature range from +50°C to -54°C. The heat sink temperature was measured and maintained at the copper heat sink plate, with the cooler body temperature allowed to rise above the heat sink temperature. A temperature gradient of 5°C was measured between the heat sink plate and the end cap of the compressor body (next to the helium transfer line flange) when the cooler was operating with 35 watts of input power. The temperature at the expander helium transfer line inlet increased as much as 25°C above the heat sink plate temperature. The cooler performance was found to be very sensitive to the heat sink temperature. The
minimum coldtip temperature attainable for the 10 Vrms input load lines varied from 45 K with the +50°C heat sink, to 23 K with the –54°C heat sink. This corresponds to a change in an applied load
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Figure 2. Thermal performance sensitivity to input voltage at +50-°C heat sink temperature.
of about 600 mW at 45 K. As another example, the cooling load at 77 K for the 10 Vrms input
increased from 0.82 watts to 1.45 watts as the heat sink temperature was changed from its warmest to its coldest temperature. In terms of the specific power, this 77-K performance improved from 38 W/ W at +50°C to 19 W/W at –54°C. In general it can be observed from the figures that a 20-°C temperature change in the heat sink resulted in an approximate 4-K change in coldtip temperature for temperatures below 60 K (or equivalently, to a 125-mW change in cooling capacity for a constant coldtip temperature). At temperatures above 100 K, there was as much as a 10-K change in coldtip temperature for a 20-°C heat sink temperature change (or similarly, this was equivalent to an approximate 125-mW change in cooling capacity at a given coldtip temperature).
Figure 3. Thermal performance sensitivity to input voltage at +20-°C heat sink temperature.
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Figure 4. Thermal performance sensitivity to input voltage at 0-°C heat sink temperature.
Drive frequency. The nominal drive frequency during these tests was 55 Hz. The frequency was varied in 5-Hz increments to compare the performance sensitivity at the different heat sink temperatures.
Operating at or near the optimal drive frequency requires the least amount of cooler input power and also yielded the lowest specific power (watts of input power per watt of refrigeration). Over the range of heat sink temperatures the cooler was operated, the optimal drive frequency varied by more than 5 hertz. At the +50-°C heat sink temperature the cooler operated well at both 55 Hz and 60 Hz (Fig. 8), with nearly identical specific powers at all but the warmest of coldtip temperatures. But the cooler required
slightly less (2 watts) input power to operate at 55 Hz. At the –54-°C heat sink temperature, the cooler performed better with a 50-Hz drive frequency, requiring lower input power and operating with lower specific power over the entire range of coldtip temperatures observed. (Fig. 9).
Figure 5. Thermal performance sensitivity to input voltage at –20-°C heat sink temperature.
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Figure 6. Thermal performance sensitivity to input voltage at –40-°C heat sink temperature.
COOLER EFFICIENCY
Coefficient of Performance. Heat sink temperature is a very important external variable in cryocooler operation in that it directly enters into the Carnot efficiency of the cryocooler. The thermodynamic coefficient of performance (COP), the figure of merit for cryocoolers, is defined as the ratio of the net cooling power to the net applied input power (input electrical power heating of the coil), and is expressed as a percentage of the ideal Carnot COP. Figure 10 shows the percent Carnot COP values for the 10-Vrms load line at each heat sink temperature. Motor efficiency. The dominant losses in the motor design are due to the losses within the coil, which are a result of the coil resistance and the capacitive or inductive circulating currents (eddy currents). The motor efficiency can be defined as the ratio of the (input power - ) to the input power. The motor efficiencies were found to be quite high also, ranging between 96–98% for the cooler for all heat sink tempera-
Figure 7. Thermal performance sensitivity to input voltage at –54-°C heat sink temperature.
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Figure 8. Thermal performance sensitivity to drive frequency at +50-°C heat sink temperature.
tures. The motor efficiency was found to be dependent on the drive frequency and the input voltage, increasing with increasing drive frequency and with increasing input voltage. Power factors. The power factor is defined as the ratio of the input power to the product of the measured true rms voltage and true rms current. Non-unity power factors are caused by the presence of compressor drive forces, such as inertial forces and the mechanical-spring and gas-spring forces, that are not in phase with the compressor velocity. Thus minimizing these forces helps to achieve high power factor. The highest power factors recorded (.98–.99) were at the highest drive frequencies tested at the various heat sink temperatures and were independent of coldtip temperature. Operating the cooler at lower drive frequencies resulted in slightly lower power factors overall, and that were also found to vary from .96 for low coldtip temperatures to .82 at high coldtip temperatures. The high power factors indicate these motors are highly tuned and operating near the mechanical resonance.
Figure 9. Thermal performance sensitivity to drive frequency at –54-°C heat sink temperature.
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Figure 10. Sensitivity of the % Carnot COP to heat sink temperature for the 10-Vrms load lines.
COOLDOWN TIMES
Cooldown times were measured for various heat sink temperatures, input voltages, applied heat loads, and copper thermal masses. All parameters were held constant during the cool down period. Cooldown times were measured for thermal masses of 0 gm (bare coldtip), 20.8 gm, and 55.3 gm and give a first order linear dependence of the cooldown time versus thermal mass. Table 1 lists the cooldown times to reach 50 K under several operating conditions and cold tip thermal masses. It should be noted that when operating the cooler with the cooler drive electronics, the cooler input voltage during cooldown is greater than 11 Vrms (see the next section). Operating with input voltages of less than 11 Vrms to the cooler during cooldowns would occur if the supply voltage to the cooler drive electronics were less than the minimum 17 VDC required by the electronics. In these instances the cooler drive electronics is effectively starving the cooler, preventing it from operating at full stroke.
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DRIVE ELECTRONICS AND POWER SUPPLY EFFICIENCY
As a final test to simulate the operation of the cooler on a spacecraft, the cooler was driven with the Texas Instrument’s voltage-mode drive electronics card and DC–DC converters to measure the efficiency of the electronics under various operating conditions. A pair of Interpoint DC–DC converters (model MFLHP2812S) provide 24 VDC power to the cooler drive electronics, and were used in parallel to reduce the current load on the converters. The converters are used to provide isolation between the cooler and the spacecraft power bus, and are capable of providing the constant output voltage for DC input voltages ranging between 19 VDC and 40 VDC. The cooler drive electronics are operable with conditioned input voltages ranging from 17 VDC to 32 VDC. Using the cooler drive electronics, the cooler operated at a drive frequency of 56 Hz. During cooldown, the cooler drive electronics drives the cooler at the maximum stroke by supplying full voltage to the cooler, starting with 11.6 Vrms at the start of the cooldown and decreasing to 11.2 Vrms as the cooler coldtip temperature drops to the 50-K range. The drive electronics continues to supply full voltage to the cooler until the coldtip temperature reaches the set point temperature, as monitored with a control diode. As this set point temperature is reached, the drive electronics reduces the input voltage (input power) to the cooler to a level where the cooler produces refrigeration equal to the thermal load on the coldfinger. If the thermal load on the coldfinger changes, the supplied input voltage to the cooler will likewise change to meet the new refrigeration requirements at the set point temperature. During the test, the DC–DC converters were kept in ambient air and cooled with a fan. The cooler drive electronics card was wedge-locked into an aluminum card cage that fastened onto a cooling plate within the vacuum chamber, and was cooled to the same heat sink temperature as the cooler. The DC–DC converters and drive card efficiencies were measured using a three-channel true rms power meter inserted into the circuit at the input and output of the devices for simultaneous
measurements. The measured efficiency of the converters varied between 81–84%, and the efficiency level was more dependent on the supply voltage than on the throughput to the cooler. The efficiency of the drive electronics ramped down from 75% at the start of the cooler cooldown to around 70% when the set point temperature was reached. The drive electronics efficiency remained around 70%, varying little as a function of either throughput power to the cooler or with changing heat sink temperature. These measured efficiency numbers suggest that the overall efficiency of the electronics circuitry is on the order of 58%, that the DC power required to operate the cooler is 1.72 times larger than the power required by the cooler. Knowing these efficiency numbers also helps understand the heat dissipation characteristics of the cooler and cooler electronics. SUMMARY The Texas Instruments 1 -watt cryocooler has been tested over a heat sink temperature range from +50°C to –54°C for a variety of input voltages and drive frequencies. The cooler was driven at several input voltages when driven with the external oscillator and amplifier; the highest input voltage was 10 Vrms, about 85% of maximum input voltage capability. The optimal drive frequency for the cooler performance depended on the heat sink temperature; a 55-Hz to 60-Hz drive frequency range was preferable at the warmest heat sink temperatures, and 45 Hz to 50 Hz was the preferred drive frequency range for the coldest heat sink temperatures. Under these conditions, the compressor input power (for the 10 Vrms input voltage) could be kept below 35 watts. Over the full heat sink temperature range the attainable noload coldtip temperature varied from 45 K to under 24 K, corresponding to a 600-mW increase in refrigeration capacity at 45 K. At a 77-K coldtip temperature, a 630-mW change in refrigeration performance was measured over this range in heat sink temperatures. During operation of the cooler with its drive electronics and simulated spacecraft circuit, the cooler drive frequency was 56 Hz and was driven with an input voltage of over 11 Vrms during the cooldown, and its set point operating temperature could be varied. The operating efficiency of the drive electronics was measured at a nominal 70% and the DC–DC converter efficiency was measured at a nominal 83%.
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ACKNOWLEDGMENTS
The work described in this paper was carried out by the Jet Propulsion Laboratory, California Institute of Technology. The work was sponsored by the Ballistic Missile Defense Organization under a contract with the National Aeronautics and Space Administration. Particular credit is due S. Leland for assembling the test setup and providing the Labview data acquisition program, and A. Lepe for the assistance in the data reduction and graphical presentations. The author wishes to thank Texas Instruments for the gracious loan of the cooler to perform these tests. REFERENCES 1.
Priest, R.E., et al., "Ricor K.506B Cryocooler Performance during the Clementine Mission and Ground Testing: A Status Report," Cryocoolers 8, Plenum Press, New York (1995), pp. 883-892.
2.
Glaser, R., Ross, R.G., Jr. and Johnson, D.L., "STRV Cryocooler Tip Motion Suppression," Cryocoolers 8, Plenum Press, New York (1995), pp. 455-463.
3.
Rawlings, R.M., Granger III, C.E., and Hindrichs, G.W., “Linear Drive Stirling Cryocooler: Qualification and Life Testing Results,” Cryocoolers 8, Plenum Press, New York (1995), pp. 121-127.
4.
Dunmire, H, and Shaffer, J., "The DOD Family of Linear Drive Coolers for Weapon Systems," Cryocoolers 9, Plenum Press, New York (1997), pp. 17-24.
Qualification of the BEI B512 Cooler,
Part 1 - Environmental Tests D.T. Kuo, A.S. Loc, and S.W.K. Yuan
Cryocooler Group, Edcliff Division BEI Technologies
Sylmar,CA91342 ABSTRACT
BEI’s involvement in cryocoolers began with a corporate sponsored project in 1991 to develop a closed-cycle Joule-Thomson cooler for a high-temperature-superconductivity application. After achieving limited success with a mixed gas refrigerant, the company leveraged its expertise in linear compressor technology to develop a miniature Stirling-cycle refrigerator for IR detector cooling. The first miniature cooler designed for 150mW capacity at 78K has been well received in the infrared user community1. BEI has recently enhanced the performance of this cooler by as much as In this paper, the qualification of this cooler to the Military Standard is discussed
in detail.
INTRODUCTION This is the first of two papers regarding the qualification of BEI’s B512 cooler (Figure 1). The environmental tests performed are summarized here, while the results of the on-going life test are reported elsewhere.
Table 1 summarizes the performance specification at three different case temperatures. The following example illustrates the performance at 22 C. With an external heat load of 120 mW,
the cooler cools down a 4.5 gram thermal mass from 22 C to 78K in less than 9 minutes. With an external heat load of 300 mW, the cooler maintains a cold tip temperature of 78K, with less than 14W of input power. With an input power of 20W, the cooler provides a minimum refrigeration of 400 mW.
Cryocoolers 10, edited by R. G. Ross, Jr. Kluwer Academic/Plenum Publishers, 1999
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Figure 1. The BEI B512B Cooler. ENVIRONMENTAL TESTS
The Environmental Tests are comprised of the audible noise test, vibration output test, basic shock test, non-operational vibration test, and gun-firing shock test, and will be discussed in the following sections.
Audible Noise Test The audible noise test was performed by the Night Vision and Electronic Sensors Directorate (NVESD) on a BEI B512C cooler (Serial number N0043). Sound pressure was measured along six axes of the cooler, one meter from the microphone. Table 2 summarizes the measured sound levels in dB as a function of center frequency, Hz. To convert the above sound levels to a distance of 10 m (instead of 1 m), the following
equation was employed.
where
is the 1 m distance, and
is the 10 m distance.
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107
Figure 2. Measured sound level vs. center frequency.
Figure 3. Vibration output forces. Figure 2 shows the measured sound levels (at a distance of 10 m) as a function of the center frequency. The cooler met the specification for all frequencies.
Vibration Output The vibration output test was also performed by the Night Vision and Electronic Sensors
Directorate (NVESD) on a BEI B512C cooler (Serial number N0043). Vibration output forces were measured in three directions. Figure 3 shows the vibration output forces along the three axes. The maximum vibration level was recorded on the axis parallel to the compressor. As noted, the forces are well below the specification of 0.51bf.
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Basic Shock
The basic shock test was performed at the facility of Environment Associates, Inc. Chatsworth, California. The cooler was subjected to shock impulses (half sine wave) of 30g for 18ms per MIL-STD-810C, Method 516.1, Procedure 1. Three shock impulses in each direction (+ and -) were imposed on all three axes for a total of 18 shocks. An ATP was performed immediately following the Basic Shock Test. When the cooler was first turned on, there was a large jerking motion and the unit then operated roughly. A clamped cable was discovered indicating that the cable was damaged by being clamped accidentally between the brackets of the shock/vibration fixture. It was verified that the cooler had been shorted. After careful inspection, it was noticed that the displacer spring bond was broken due to the shorted cable. The spring was reattached and bonded, and the cable repaired. The cooler then passed the performance ATP. Non-Operational Vibration
The cooler was then subjected to the Non-Operational Vibration Test according to the spectral density profiles outlined in Tables 3, 4 and 5, of MIL-STD-810E, Method 514.4, Procedure I, Category 8. The test duration was 60 minutes per axis. An ATP was performed immediately following the Non-Operational Vibration Test. The cooler was found to operate in a rough manner for a couple of seconds and the abnormality subsided. The cooler then passed the ATP test.
Figure 4. Basic shock profile.
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Figure 5. Gun-Firing shock profile.
Gun-Firing Shock
The Gun-Firing Shock test was also performed at the facility of Environment Associates, Inc. Chatsworth, California. The cooler was subjected to shock impulses (half sine wave) of 120g for 1ms per MIL-STD-810C, Method 516.1, Procedure 1. Three shock impulses in each direction (+ and -) were imposed on all three axes for a total of 18 shocks. The profile of the gun-firing shock is presented in Figure 5. An ATP was performed immediately following the Basic Shock Test. The cooler was found to run rough. The operator tapped on the fill-port side of the compressor a few times and the roughness subsided. The cooler then passed the performance ATP.
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POST-TEST ATP The post-qualification ATP was performed a week after the other tests. The cooler passed the ATP with cooldown times of 7:57 minutes and 10:01 at 23 C and 82 C, respectively. Maximum refrigeration capacity was measured to be 500 mW at 23 C and 370 mW at 82 C. The input power was 12 W for 400 mW load at 23C and 19.6 W for 300 mW load at 82C.
CONCLUSION The cooler passed all ATP tests following each environmental test described in this report. The life test of this cooler will be reported elsewhere. REFERENCES 1
Kuo,D.T., A.S. Loc, and S.W.K.Yuan, Experimental and Predicted Performance of the BEI MiniLinear Cooler, Cryocoolers 9, Plenum Press, New York (1997) p.l19.
2
Yuan, S.W.K., D.T.Kuo, and A.S.Loc, Enhanced Performance of the BEI 0.5 Watt Mini-Linear Stirling Cooler, to be published in Proc. of Advances in Cryogenic Engineering, Vol. 43, 1997.
Use of Variable Reluctance Linear Motor for a Low Cost Stirling Cycle Cryocooler
M. Hanes, D. Chase, and A. O’Baid
Superconductor Technologies Incorporated Santa Barbara, CA 93111
ABSTRACT
In the implementation of a compressor for a linear Stirling cryocooler, one major cost factor is the volume of magnet material required in the motor. Since the magnet material used is oftentimes the high energy-density rare earth types for size and high ambient operating temperature considerations, any reduction in the total magnet volume will result in a cost savings. The use of a variable reluctance linear motor configuration can be implemented into a compressor for a specified peak force, with a reduced amount of magnet material required. However, a variable reluctance motor has a non-linear force profile which also must be taken into account for optimal performance. This is especially critical when implementing cryocoolers which use gas bearings for long-life operation. Design data, and resulting performance for a linear free-piston Stirling-Cycle Cryocooler implemented using a variable reluctance motor are presented and compared to a conventional linear motor implementation.
INTRODUCTION Superconductor Technologies Inc. (STI) has as its primary product high temperature superconductors (HTS) which are utilized as filters in cellular applications to provide increased performance over conventional filters. One of the ramifications of utilizing HTS in the filters is the necessity of having to maintain the HTS at cryogenic temperatures. There are a multitude of cryogenic refrigerators which could be considered for this application; STI chose to use a Stirling cycle, free piston, linear motor design. This design provides for a compact, efficient, long life system without the need for helium lines connecting the cold end to the compressor, as with a Gifford McMahon type cooler. If the application involves mounting the HTS products on a tower, or other remote locations, the elimination of the helium lines becomes a more significant advantage. The Stirling cycle cooler designed for this application must have a long life to allow the cryogenically cooled filters to compete effectively with the conventional, ambient temperature filters. Cryocoolers 10, edited by R. G. Ross, Jr.
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The performance requirements for this application were determined to be 4 watts of lift at
room ambient with less than 150 watts input power, an operating range of -40°C to 60°C, a mass of less than 5 kg. The motor is configured to utilize an external coil and laminations, which
alleviates the need for electric feedthroughs, removes the potentially dirty coil winding from the working volume and reduces the internal heat load by eliminating the losses associated with the coil. This motor uses magnets which are magnetized axially, as opposed to radially. Since there is no magnetizing volume limitations, as with small, radially polarized magnets, the magnets can be made in one piece and then magnetized, as opposed to manufactured in multiple sections, then magnetized and then assembled. This decreases the cost of the magnet assembly. A gas bearing is utilized to prevent the compressor piston from coming into contact with the compressor cylinder. This eliminates both the wear and debris generation which would result if there is contact between these surfaces. The targets for the cooler life is 40,000 hours with no required maintenance. GAS BEARING DESIGN PRINCIPLES One of the key reasons for choosing this motor is the fact that it requires a relatively small amount of magnet volume, thus reducing the cost of the motor. This motor incorporates a small air gap between the internal iron and the external laminations, which reduces the reluctance of the magnetic circuit, thereforee allowing this motor to perform with low magnet volume. However, one of the drawbacks of this small air gap is a high sensitivity of the radial side force between the internal and external components of the motor and the eccentricity between these parts. This radial side force can easily exceed the weight of the piston - magnet assembly. In order to overcome this force this cooler utilizes a gas bearing scheme to ensure long cooler life. The gas bearing eliminates virtually all contact between the compressor piston and the compressor cylinder, hence eliminating friction and wear. The piston essentially floats on a thin layer of helium gas, which is the same gas used as the working fluid for the thermodynamic processes within the cooler. A cross section of the piston - gas bearing - cylinder assembly is shown below. During cooler operation, the high pressure reservoir is kept at a relatively constant and high pressure by the action of the check valve. During the portion of the cycle where the working pressure in the warm end of the cooler is higher than the pressure of the high pressure reservoir, helium flows from the warm end into the reservoir and “recharges” it. During the time when the warm end pressure is lower than the reservoir pressure, the check valve is closed, preventing helium from escaping from the reservoir. During the entire cycle, helium is flowing from the
Figure 1. Typical piston gas bearing layout.
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reservoir through the piston flow restrictors and into the bounce volume. The three pressures within the system are shown in Figure 2. As shown on the graph, all the pressures initially start at the same level. As the cooler begins to run, the pressure in the reservoir begins to pump up to an almost constant level. The magnitude of the fluctuation in the reservoir pressure is a function of the reservoir volume and the piston flow restrictor flow rates. Therefore, if these parameters are designed correctly, the gas bearing will operate over an almost constant pressure difference, in spite of the oscillatory nature of the pressure in the warm end, or working, volume of the cooler. Figure 3 shows an expanded piston - cylinder gap to illustrate the principles of a gas bearing supported piston.
Figure 2. Pressures in various cooler volumes.
Figure 3. Gas bearing with piston off axis.
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A piston which is supported by a gas bearing will have the flow resistance of the piston flow restrictor approximately equal to the flow resistance of the annular gap between the piston and the cylinder, when the piston is centered in the cylinder. This results in the pressure in the gas bearing pad being approximately halfway between the reservoir and the bounce volume pressures. Also, when the piston is centered, the pressures in the pads are equal on all sides of the piston, and there are no net bearing forces acting on the piston. However, when the piston is forced off center, as depicted in the above drawing, the resistance of “gap 2” becomes lower than that of “gap 1” and the pressure in the gas bearing pad associated with gap 1 increases (becomes more closely coupled to the higher pressure reservoir), while at the same time the pressure in the pad on the opposite side decreases (becomes more closely coupled to the lower pressure bounce volume). This results in a pressure difference between the two sides of the piston, which act upon the projected area of the piston to provide a centering force. Since the flow resistance of the gap is proportional to the inverse of the gap width cubed, large pressure differences will exist for very small piston offsets. This self centering gas bearing will have a spring constant in the range of 10,000 1b/in per set of gas bearings, which is adequate to prevent piston to cylinder contact, and ensure the longevity of the cooler.
MINIMIZING MOTOR SIDE FORCES THROUGH DESIGN AND ALIGNMENT
Design
The side force associated with a given amount of non concentricity between the internal iron - magnet assembly and the external laminations is proportional to the magnetic flux in the air gap between the internal magnet assembly and the external laminations, and, hence, can be controlled by varying the size of this air gap. As the air gap is increased, the flux in the gap decreases, resulting in lower radial forces between the assemblies. The graph in Figure 4 shows the effect of the air gap on the radial side force.
Figure 4. Effect of air gap on magnetic side force.
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The negative aspect of increasing the air gap is that the force constant and efficiency of the motor decrease. Through a combination of performance testing the motor, both by itself and in a cooler, and analysis of the system, a compromise which reduces the side forces to acceptable levels and maintains good overall cooler performance was attained. The optimum air gap for this application was determined to be .025 inch. Alignment
Although a gas bearing system is used to prevent contact between the piston and cylinder, it is still of paramount importance to minimize the magnitude of the motor side force in the basic cooler design. This is accomplished by insuring the internal iron - magnet assembly is as close to concentric with the external laminations as possible. This is accomplished by insuring proper alignment during two assembly procedures during the cooler build; the piston OD to the internal iron - magnet assembly OD and the compressor cylinder ID to the pressure vessel OD. The pressure vessel OD is the surface to which the external lamination are mounted. The maximum allowable total indicated runout (TIR) between each of these surfaces is .0015 inch. This allows for a maximum runout of .003 inch between the internal iron - magnet assembly and the ID of the external laminations, which is equivalent to an offset of .0015 inch (half the runout). In order to avoid excessively tight tolerances, and to keep the cost of the piece parts to a minimum, precision tooling is utilized in the assembly, braze and weld processes to insure these alignments are met. As can be seen in Figure 4, the magnitude of the motor side force at the worse case alignment is approximately 1 lb. Based on the gas bearings having a spring constant of about 10,000 1b/in and the radial gap between the piston and the cylinder being nominally .0006 inch, the gas bearings can support 6 lb of radial force before contact between the piston and cylinder
will occur. Even with the weight of the piston - magnet assembly, 0.65 lb, adding to the side
force, this resultant load is still well within the 6 lb limit of the gas bearing. In addition to measuring the pressure vessel to cylinder runout, the location of the runout extreme is recorded,
and, when the cooler is mounted horizontally, it is oriented such that the magnetic side force is opposing the gravitational force.
MOTOR CONTROL
The other concern with this motor is the non linearity of the motor force constant, motor inductance and radial side force as a function of axial displacement. This characteristic is shown in Figure 5. These non-linearities make controlling the cryocooler a more difficult undertaking. Driving Considerations
In a conventional linear motor, the motor terminal impedance is conventionally represented as a series combination inductance, resistance, and back-EMF generator as show in Figure 6. In this figure, Lm represents the armature inductance. Rac represents the total dissipative losses at the operating frequency, taking in account not only the DC coil resistance, but dynamic losses such as eddy current losses, hysteresis losses and AC winding resistance. The voltage source represents the back-emf generated in response to the linear velocity (dx/dt) of the piston. In actual operation, the motor current is servo-locked to a sinusoidal reference, resulting in a sinusoidal current, which in turn results in a sinusoidal force applied to the piston with magnitude determined by Kf (numerically identical to Kv through appropriate units). Piston stroke amplitude can be measured and limited by de-imbedding the Lm and Rac terms to
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Figure 5. Axial force vs. piston location.
Figure 6. Armature electronic equivalent circuit of a linear motor.
determine the back-EMF term. By determining the back EMF magnitude, the velocity, and hence the peak-to-peak piston amplitude can be controlled. In a variable reluctance linear motor, the model becomes more complex to account for the impact of the changing reluctance, as shown in Figure 7. In this case, the inductance Lm the back EMF ‘constant’ Kv and the force ‘constant’ Kf are all now functions of piston position (x), which considerably complicates the driver circuitry and the mechanism to limit piston travel. More specifically, due to the position-dependent axial force profile, a sinusoidal current no longer insures a sinusoidal force profile on the piston.
Figure 7. Armature electronic equivalent circuit of a variable reluctance linear motor.
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Sinusoidal piston motion could be obtained through the use of a position sensor on the piston in conjunction with a suitable control system. For cost and reliability reasons, however, a sensor was not included in the cooler itself. Consequently, operation without a position sensor feedback was required. In actual implementation, given the linear motor is operated at resonance, the resultant mechanical band-pass filter is reasonably effective at attenuating harmonics of the forcing function. Consequently, if the cooler is driven with a sinusoidal current waveform, the harmonic distortion generated by the Kf(x) term is attenuated and does not appear to have a large affect on the overall cooler performance. Nevertheless, ongoing work is being done using pre-distortion of the applied current waveform to yield a sinusoidal forcing function with lower harmonic content, for increased efficiency and lower vibration considerations. The Lm(x) and Kv(x) terms, produce a terminal voltage which is a strongly non-linear function of the current, because of their inherent dependence on axial position. The biggest impact of this is in the control of the peak-to-peak piston amplitude. While the velocity of the piston can be extracted in a linear motor by ‘de-imbedding’ the back EMF term, this same approach cannot be used with the variable reluctance motor without considerable increase in controller complexity. A simple and effective approach to piston stroke limitation was implemented by exploiting
the fact that the Kv(x) term becomes extremely non-linear at the ends of the piston stroke. This is detected by first reducing the loop bandwidth of the current drive control loop to 300Hz, then examining the amplitude of harmonics of the resulting current waveform in the 400Hz-700Hz range. As the piston moved to the end of its range, the strong non-linearity of Kv(x) generated a rich harmonic content in the current waveform, which was detected by the controller and used to limit the drive level. As the characteristics of the cooler change, (e.g. as its cold-tip cools, or over ambient temperature changes) this approach appears to adapt readily and allow maximum drive into the cooler without piston overstroke.
COOLER PERFORMANCE SUMMARY
Table 1 lists the design goals of the cooler and the actual achieved results and Figure 8 shows the performance of the cooler at various input power levels over a wide ambient temperature range.
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Figure 8. Cooler performance at various input powers and heat rejector temperatures.
CONCLUSION
Using a variable reluctance linear motor in a cryocooler does present some associated difficulties, such as relatively high magnetic side forces and non linearities associated with axial displacement. It is the relatively small air gaps which contribute to the side forces, however, it is
this small air gap which allows the use of low volume magnets. Through proper motor design, accurate assembly of the cryocooler and innovative electronic controller design, these potential drawbacks have been overcome and this motor is providing an inexpensive approach for a long life cryocooler suited to commercial applications.
AIRS PFM Pulse Tube Cooler System-Level Performance R.G. Ross, Jr., D.L. Johnson, and S.A. Collins Jet Propulsion Laboratory California Institute of Technology Pasadena, CA 91109 K. Green and H. Wickman Lockheed Martin IR Imaging Systems (LMIRIS) Lexington, MA 02173
ABSTRACT JPL’s Atmospheric Infrared Sounder (AIRS) instrument is being built to make precision measurements of air temperature over the surface of the Earth as a function of elevation; the flight instrument is in the final stages of assembly and checkout at this time, and uses a pair of TRW pulse tube cryocoolers operating at 55 K to cool its sensitive IR focal plane. The cryocooler development activity is a highly collaborative effort involving cooler design and fabrication at TRW, cooler characterization and qualification testing at TRW and JPL, and system-level performance characterization and instrument integration at LMIRIS. During the past few months the Engineering Model AIRS cooler has been integrated with the instrument focal plane assembly and measurements have been made on the overall thermal and operational performance of the cryosystem including vibration compatibility, ∆T from cooler to focal plane, and temperature control stability. At the same time the AIRS flight (PFM) coolers have undergone qualification and characterization testing at JPL prior to shipment to LMIRIS in January 1998, where they are now undergoing integration and system-level testing with the AIRS flight instrument. This paper presents the measured system-level performance of the AIRS flight coolers including detailed thermal, vibration, and temperature control performance with the EM and flight instrument boundary conditions. INTRODUCTION The objective of the Atmospheric Infrared Sounder (AIRS) instrument is to make precision measurements of atmospheric air temperature over the surface of the Earth as a function of elevation. The AIRS instrument is scheduled to be flown on NASA's Earth Observing System PM platform in the year 2000, and is being developed under JPL contract by Lockheed Martin IR Imaging Systems (LMIRIS) of Lexington, MA. In Spring 1994, TRW of Redondo Beach, CA was awarded the contract to develop and produce the flight coolers for the AIRS instrument. After delivering the Engineering Model (EM) cooler1 for testing and integration studies in July Cryocoolers 10, edited by R. G. Ross, Jr. Kluwer Academic/Plenum Publishers, 1999
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Figure 1. AIRS flight cryocoolers and drive electronics
1996, the flight (PFM) coolers, shown in Fig. 1, were delivered in November 1997.2 The foundation of the AIRS instrument is the 58-K infrared HgCdTe focal plane cooled by a pair of TRW 55-K pulse tube cryocoolers. The focal plane is hard mounted to the infrared spectrometer, which in-turn is cooled to 150 K by a 150K/190K two-stage cryogenic radiator. The spectrometer operates over a wavelength range from visible through and places demanding requirements on the thermal, vibration, and EMI performance of the cryocoolers. Figure 2 illustrates the overall instrument construction and highlights the key assemblies. Physically, the instrument is approximately in size, with a mass of 150 kg and an input power of 220 watts. Configurationally, the 58-K IR focal plane assembly is mounted integrally with the 150-K optical bench, which is in-turn shielded from the ambient portion of the instrument by the 190-K thermal radiation shield and MLI blankets. The ambient portion of the instrument contains the high power dissipation components including the instrument electronics and the cryocoolers and their electronics. These high-power-dissipation components have their heat rejection interface to a set of coldplates that conduct the heat to spacecraft-mounted radiators via a system of heatpipes. Extensive characterization of the cooler's performance has been carried out during the qualification testing and instrument integration phases at TRW, JPL and LMIRIS. These test results are described in the remainder of this paper.
Figure 2. Operational elements of the overall AIRS instrument
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AIRS CRYOCOOLER SYSTEM-LEVEL PERFORMANCE AIRS Cryosystem Design and Thermal Interface Attributes Early in the design of the AIRS instrument, key decisions of design philosophy were established that served as fundamental ground rules for the cryocooler system design. These included: • Totally redundant cryocoolers—to avoid one cooler being a single-point failure • No heat switches—to avoid increased complexity, cost and unreliability • Ambient heat rejection to spacecraft-supplied cold plates operating between 10 and 25°C
• Cooler drive fixed at 44.625 Hz, synchronized to the instrument electronics—to minimize asynchronous vibration or EMI noise pickup from the cryocooler
• Cold-end load (focal plane) mechanically mounted and aligned to the 150K optical bench with a maximum vibration jitter on the order of 0.2 mm
• Focal plane calibration (for temperature, motion, etc.) every 2.67 sec (every Earth scan) • Cooler input power goal of 100 watts (22 to 35 volts dc), and mass goal of 35 kg • Cooler drive electronics fully isolated (dc-dc) from input power bus; EMI consistent with MIL-STD-461C
Based on the above fundamental ground rules, the AIRS cryosystem design, shown in Fig. 3, was developed.3,4 This system incorporates two independent 55K cryocoolers, a primary and a non-operating backup, each connected to the 58K focal plane using a common high-conductance coldlink assembly. Ambient heat from the operating cooler is rejected to the coldplates located in the plane of the instrument/spacecraft interface. Table 1 provides a breakdown of the overall cryocooler beginning-of-life (BOL) refrigeration load measured on the AIRS Engineering Model (EM) instrument, and projections of representative end-of-life (EOL) properties. A key determiner of these BOL/EOL loads is the BOL/EOL temperature of the optical bench and pulse tube vacuum housing—assumed to be 145 K/160 K and 309 K/314 K, respectively. Cooler Thermal Integration Considerations
To minimize thermal conduction losses between the focal plane and the cryocooler, the pulse tube coldblock needs to be located close to the focal plane. Unfortunately, in addition to providing refrigeration, the expander of a high-efficiency pulse tube refrigerator also dissipates a large amount of ambient heat — for AIRS, nearly 50% of the total compressor input power.
Figure 3. Schematic of AIRS cryogenic system and cooler interfaces.
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Thus, the expander also needs to be mounted close to the instrument heat rejection system in order to minimize its operating temperature and maximize its efficiency. With the AIRS instrument, the distance between the focal plane and the instrument heat-dissipation cold plates is approximately 45 cm (18 inches), and is spanned by a combination of the cooler-focal plane coldlink assembly and the pulse tube expander heat-rejection mounts as shown in Fig. 3. Sapphire Coldrod/Flexlink Assembly Thermal Conductance. The sapphire coldlink assembly—designed and fabricated by LMIRIS—contains a copper-braid flexlink section at one end to accommodate the relative motion that occurs between the pulse tube and the spectrometermounted focal plane dewar during launch and during cooldown of the instrument to cryogenic temperatures. The copper flexlink assembly bolts directly onto the two pulse tube coldblocks at one end, and at the other end attaches to the gold-plated sapphire coldrod using a molybdenum/ aluminum shrink-fit interface. The total measured thermal resistance of the complete coldrod assembly from the pulse tube coldblock to the focal plane active elements is approximately 3.5 K/ W as detailed in Table 2. In addition to the copper-braid section that connects the pulse tube coldblocks to the sapphire rod, the cold link assembly also contains copper braids that connect the coldblocks to one another so that the appreciable (~0.5 watt) off-state conduction of the redundant cryocooler pulse tube does not have to be conducted to the sapphire rod and back to the operating cooler. Pulse Tube and Compressor Heat Rejection Performance. The pulse tube and compressor heatsink mounts, illustrated in Fig. 4, were designed and fabricated by TRW as part of the cooler structural support, and delivered as part of the cryocooler system. These mounts are required to conduct up to 40 watts from the operating expander to the cryocooler heat-rejection coldplate, and up to 70 watts from the operating compressor to the coldplate. The heatsink design strives to simultaneously minimize the rejection temperature of the pulse tube and compressor
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Figure 4. Temperature rise of the compressor and pulse tube regenerator and orifice block above the 25°C coldplate temperature as a function of compressor power level.
and the total required mass. Figure 4 describes the measured gradients between cold plate interface temperature, nominally 25°C, and the operating temperatures achieved at the pulse tube regenerator, the pulse tube orifice block, and the compressor body. Cryocooler Refrigeration Performance
One of the key attributes of the TRW AIRS pulse tube cryocoolers is their excellent thermal performance as highlighted in Figs. 5 through 7. These data are for a coldplate interface temperature of 25°C and include the effects of the significant thermal gradients detailed in Fig. 4. Note that the two flight coolers both achieve approximately 50 WAV at 55 K, but are slightly different in thermal performance, representing unit-to-unit differences. Also note that Fig. 5 is in terms of compressor input power, while Fig. 6 is in terms of total cooler system input power, including the inefficiency of the electronics. Figure 7 describes the sensitivity of the measured cooler performance to heatsink temperature. The 4-K shift in the isotherms for the 20°C change in heatsink temperature gives the 1-to-5 temperature-sensitivity ratio typical of previous TRW coolers.5
Figure 5. Measured thermal performance of the AIRS flight pulse tube coolers in terms of compressor input power with 25°C heat rejection coldplate temperature.
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Figure 6. Measured thermal performance of the AIRS flight pulse tube coolers in terms of total cooler
system input power including electronics with 25°C heat rejection coldplate temperature.
Figure 7. Measured sensitivity of the thermal performance of the AIRS flight pulse tube cryocoolers to changing heatsink temperature.
Cryocooler Electronics Performance
Included in the performance data of Fig. 6 is the efficiency performance of the AIRS cryocooler drive electronics. These electronics, shown earlier in Fig. 1, are a key part of the overall AIRS cryocooler system and play a critical role in the overall cooler performance. Figure 8 describes the details of the cooler electronics electrical efficiency as a function of load. Note that although the electronics draw on the order of 15 watts when the compressor is at zero power input, the extrapolated tare power is less than 5 watts when the compressor is running at its design load in the AIRS instrument.
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Figure 8. Relationship between total cooler system input power (including electronics) and compressor input power for compressor power levels from 0 to 100 watts.
In addition to being required to drive the compressors with high electrical efficiency, the cryocooler electronics are also required to perform a number of vital control, noise suppression, and data acquisition functions. These additional design attributes include: • Full (dc-dc) transformer isolation from the input 28 Vdc power bus • Built-in shorting relays to suppress cooler piston motion during launch • Cooler drive fixed at 44.625 Hz, synchronized to the instrument electronics—to minimize asynchronous vibration or EMI noise pickup from the cryocooler • Very high degrees of EMI shielding, consistent with MIL-STD-461C • Advanced feedforward vibration suppression system with accelerometer-based closed-loop nulling of the first 16 cooler vibration harmonics • Closed-loop cooler coldblock temperature control via piston stroke control • Built-in monitoring of cooler operational variables and performance data • Built-in low-frequency stiction test drive waveform Other aspects of the performance of the cryocooler electronics are described below under Electromagnetic Interference, Self Induced Vibration, and Coldblock Temperature Control.
Electromagnetic Interference An important attribute of both the AIRS mechanical cooler and its electronics is generated EMI, particularly AC magnetic fields (Figure 9), radiated electric fields, and AC ripple current fed onto the 28 Vdc power bus. As described in detail in a companion paper,6 the AIRS cooler
Figure 9. AC magnetic fields radiated from the AIRS mechanical cooler before and after the addition of the flight mu-metal shields — compared with the requirements of MIL-STD-461C RE01.
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design incorporates special external mu-metal magnetic shielding to suppress AC magnetic fields from the mechanical compressor drive motors, and special EMI-suppression packaging of the electronics to control radiated electric fields. With these provisions, the AIRS cooler meets all EMI requirements except in the area of AC ripple currents on the input power bus. Excessive ripple current is a particularly demanding issue for linear coolers of the Oxford type because the motor drive current varies sinusoidally at the relatively low operating frequency of the cooler — 44.625 Hz for the AIRS cooler. For AIRS, the solution involved the integration of a special
power supply within the spacecraft that is able to accommodate very high cooler ripple currents, and the addition of a supplemental EMI filter within the AIRS instrument. Table 3 summarizes the input ripple current attributes of the two AIRS coolers as a function of power level.
Self Induced Vibration Another important function of the AIRS cooler drive electronics is suppression of self induced vibration through the use of an advanced feedforward vibration suppression system with
accelerometer-based closed-loop nulling of the first 16 cooler vibration harmonics. The AIRS instrument has a strong sensitivity to vibration and jitter, allowing no more that 0.2µm movement
between the focal plane and the incident optical beam during any single 2.67 second scan. Figure 10 describes the measured vibration forces generated by each of the two AIRS coolers when
Figure 10. Vibration forces measured from the AIRS S/N 301 and S/N 302 cryocoolers during qualification acceptance testing with the active vibration suppression activated.
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Figure 11. Focal plane and cryocooler coldblock temperature history during a system-level test involving turning the focal plane on (a ~200 mW step increase in cryocooler load) from a previous off condition.
operated individually during qualification acceptance testing. Note that the vibration is reduced to very low levels except for the 6th harmonic, which is greatly amplified by the presence of a principal structural mode in the cooler support structure around 270 Hz. Because the resonant frequencies will be different (and lower) in the final instrument configuration, the ultimate test of vibration compliance will be during system-level operation in the flight instrument. So far, cooler operation with the AIRS EM instrument has shown no measurable effects from coolergenerated vibration. Coldblock Temperature Control
As noted above, coldblock temperature control is another demanding system-level function carried out by the AIRS cryocooler electronics. The requirement for this feature stems from a very strong sensitivity of focal plane background noise level to focal plane temperature, and a noise-rejection algorithm that requires a high level of noise stability during any individual 2.67 second scan cycle. The result is a requirement for short term focal plane temperature fluctuations no greater that and a corresponding cooler coldblock temperature fluctuation no greater than 10 mK. The AIRS cooler electronics performs this temperature control using digital control of compressor stroke amplitude based on temperature instrumentation on the pulse tube coldblock. Figure 11 illustrates the quality of control achieved during a severe system-level test involving turning the focal plane on (a ~200 mW step increase in cryocooler load) from a previous off condition. Note that the recovery time to reach stability is approximately one hour and the level of control is approximately Cryocooler System Mass
As a final characterization of the AIRS pulse tube cryocooler system, Table 4 highlights the mass breakdown by element. SUMMARY AND CONCLUSIONS
The AIRS cryocooler system development activity is a key part of the AIRS instrument development and focuses on developing and integrating the cryocoolers so as to maximize the performance of the overall instrument; it is a highly collaborative effort involving development
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contracts with Lockheed Martin and TRW, and cryocooler characterization testing at JPL. To date, the overall cryocooler integration approach has been developed and refined, and the stateof-the-art TRW pulse tube cooler has demonstrated excellent thermal performance and light weight. Results have been presented detailing the cryogenic loads on the cooler, the overall cryocooler thermal performance margins achieved, and thermal heatsinking considerations. Mass properties of the cryocooler system, and thermal properties of the developed coldlink assembly have also been presented. ACKNOWLEDGMENT
The work described in this paper was carried out by the Jet Propulsion Laboratory, California Institute of Technology, Lockheed Martin IR Imaging Systems, and TRW, Inc; it was sponsored by the NASA EOS AIRS Project through an agreement with the National Aeronautics and Space Administration. REFERENCES 1. Chan, C.K., et al., "Performance of the AIRS Pulse Tube Engineering Model Cryocooler," Cryocoolers 9, Plenum Publishing Corp., New York, 1997 pp. 195-202. 2. Chan, C.K., Raab, J., Colbert, R. , Carlson, C. and Orsini, R.,“Pulse Tube Coolers for NASA AIRS Flight Instrument,” Proceedings of ICEC 17, 14-17 July 1998, Bournemouth, UK. 3. Ross, R.G., Jr. and Green K., "AIRS Cryocooler System Design and Development," Cryocoolers 9, Plenum Publishing Corp., New York, 1997, pp. 885-894. 4. Chan, C.K., et al., "AIRS Pulse Tube Cryocooler System," Cryocoolers 9, Plenum Publishing Corp., New York, 1997, pp. 895-903. 5. Ross, R.G., Jr. and Johnson, D.L., “Effect of Heat Rejection Conditions on Cryocooler Operational Stability,” Advances in Cryogenic Engineering, Vol. 43, 1998. 6. Johnson, D.L., Collins, S.A. and Ross, R.G., Jr., "EMI Performance of the AIRS Cooler and Electronics," Cryocoolers 10, Plenum Publishing Corp., New York, 1999.
Multispectral Thermal Imager (MTI) Space Cryocooler Development, Integration, and Test 1st Lt. B. J. Tomlinson
D. Davidson and C. Lanes
Air Force Research Laboratory Kirtland AFB, NM 87117
Albuquerque, NM 87185
W. Burt
A. Gilbert
TRW Space & Electronics Group Redondo Beach, CA 90278
Nichols Research Corporation Albuquerque, NM 87106
Sandia National Laboratories
ABSTRACT
This paper presents the selection rationale, development, flight qualification, flight integration, and characterization testing for the Multispectral Thermal Imager (MTI) cryocooler. The MTI program is a technology demonstration effort led by Sandia National Laboratories (SNL) under the sponsorship of the Department of Energy. As part of this effort, Sandia will build, deploy, and operate a small satellite, the core of which is a telescope and focal plane that is cooled to 75 Kelvin using a TRW supplied pulse tube cryocooler. This application, a milestone in the application of active cryocooling, is intended to demonstrate the utility of large capacity, non-split, cryocoolers for space missions.
The cooler for MTI is a near identical, tailored version of the TRW 6020 pulse tube cryocooler, a product of recent long life cooler technology advances, which can provide greater than 2 Watts of cooling at 60 Kelvin. The focal plane is under development at Santa Barbara Research Center (SBRC). Both items are one-of-kind products whose integration occurs in the later phases of the MTI system build, thereby posing potential schedule risks to the overall program success. The existence of a TRW 6020 pulse tube at the Air Force Research Laboratory (AFRL), and it’s similarity to the MTI cryocooler, allowed for characterization testing to be performed with a focal plane thermal mock-up to mitigate this risk. This test and the development of a new thermal strap design will also be presented. SNL SELECTION RATIONALE
The MTI program considered various cooling options and cooler configurations. A closedcycle, mechanical refrigerator was selected to meet the requirement for a continuously cooled
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focal plane for the duration of the three-year mission life. The high reliability of the cooler accommodated a non-redundant cooling system, eliminating the associated difficulties and inefficiencies of such a design. The pulse tube cooler was chosen over a piston-displacer cooler primarily due to its inherent increased reliability, lower shaking forces, and lower mass. Considering the constraints of the MTI payload, an integrally configured cooler was deemed superior to a split expander/compressor configuration. The justification being; the improved efficiency of an integral cooler, the rigidity of the transfer line between the expander and compressor precluded Optical Assembly/Payload Structure sway, and the waste heat generated on the split expander was difficult to remove. TRW FLIGHT COOLER SYSTEM
The MTI cryocooler refrigeration system consists of the elements shown in Figure 1, including the Thermal Mechanical Unit (TMU), the Cryocooler Control Electronics (CCE), the cryogenic thermal strap for connection to the user’s instrument, and flight cables for connecting the TMU and CCE. An ancillary electronics unit, known as the Sensor Conditioning Assembly (SCA), provides signal buffering for the dual redundant accelerometers used for vibration cancellation. The SCA also provides conditioning for the TMU internal position sensors that monitor piston displacement and provide signals to maintain DC centering of the pistons and provide end stop trip in the anomalous event of a stroke excursion. The nominal operating point of the cooler is 2.25 W at 65 K at 300 K. The TMU has nearly double this capacity at extended stroke and this was a major factor in the selection of the cooler for the MTI application. The system is designed to exceed the 3-year, 10-krad MTI requirements. The TMU is a head-to-head configuration using Oxford flexure bearing technology and is a
clone of the 6020 cooler reported elsewhere 1,3 except for exterior mounting and heat pipe interface tailoring. The avionics is a TRW standard product model B130 design, designed to be highly compact, low weight, efficient, with low EMI signature. To accommodate the large cooling capacity of the cooler, the CCE is designed to source up to 130 W to the cooler with 85% nominal efficiency over the operating range. It is controlled internally by a RTX2010 processor and is designed for autonomous operation in the event of loss of communication with the payload. New operating code can be uploaded to this unit from the ground station in the event this is desired. The three-year calculated system reliability is 0.9411 and is limited by the control electronics.
Figure 1. MTI Cryocooler System Hardware.
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Table 1 provides the mass of the system components; total weight is 19.6 kg. Utah State University (USU) developed the cryogenic thermal strap specifically for TRW in collaboration with the TRW author. The novel aluminum foil design uses swaging to eliminate soldering interfaces and is light enough not to amplify launch loads on the cold head yet efficiently conducting 2-3 W in a flexibly compliant configuration. A prototype of this strap was provided to SNL and used for the 6020 performance tests at AFRL (discussed later).
TRW COOLER FLIGHT QUALIFICATION Load lines for the MTI cooler are presented in Figure 2. The MTI cooler is less efficient than the 6020 cooler. An intentional variation in the cold head design to slightly improve performance based on laboratory data in practice had a result of reducing it, but still left the cooler within performance requirements. The cause of the small but unexpected variation has not been established.
Figure 2. MTI Cryocooler Load Lines.
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Figure 3. Thermal Cycle Tests.
Previous vibration qualification allowed for testing to acceptance levels, whereas shock testing was performed to qualification levels. These levels are shown in Table 2. The shock loads are severe and similar to spacecraft-level shock loads. The components passed shock and vibration testing without difficulty except for a minor de-bond of a temperature sensor inside the
Figure 4. TMU Residual Vibration Output with Cancellation Control Loop Enabled.
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avionics unit determined to be improperly bonded to a solder-plate surface and a lead failure of a CCE caging transistor that was traced to a mounting support design error. The TMU and electronics were thermal cycled to the levels shown in Figure 3. All tests were performed at the nominal operating point with temperature and vibration control loops active and operating. In subsequent testing the TMU was operated at –20 °C. The measured vibration output of the TMU is shown in Figure 4. For test expediency, the measurements were made while rigidly mounted on a commercial Kistler dynamometer using the Kistler sensors for error signal feedback. In this way force measurements are derived directly and the user’s mounting structure was not a factor in reducing the data. The cooler system’s EMI output and EMC susceptibility were measured in accordance with MIL-STD-461D for CE101, CE102, RE101, RE102, CS101, CS115, RS101, RS103. It was found that the system passed with the following exceptions; the CE102, RE102, and the CS tests were marginal, while CE101 was significantly exceeded. The conducted emission results are attributed to the large periodic current draw, which is typical of oscillating drive mechanical cryocoolers. A SNL in-line power filter will reduce the effects on the bus to acceptable levels. SNL FLIGHT INTEGRATION
The cooler integration design effort was primarily concerned with; providing power and command interface, structural support for the CCE and TMU, waste heat removal, cryogenic heat conduction from the focal plane, and insulation of cryogenic surfaces. The payload configuration is illustrated in Figure 5. The cooler system power is supplied to the CCE from the 28V spacecraft bus, the CCE thereafter powers the TMU. The power to the cooler system is conditioned by a passive power filter which reduces what would otherwise be excessive current ripple, resulting from the cyclic
operation of the cooler. The cooler is commanded via a RS422 link from the payload processor. The use of the integral cooler design allows the implementation of a straightforward structural support and waste heat removal design. The Focal Plane Assembly (FPA) is mounted
Figure 5. MTI Payload Configuration.
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Figure 6. Cooler/FPA Integration.
Figure 7. Cooler Integration Hardware.
directly to the telescope (or, Optical Assembly, OA), that is supported by Payload Structure via six kinematic links. The kinematic links impose some flexibility between the OA and the Payload Structure, essential to OA performance, but a complexity to the cooler integration. The Payload Structure is used to support the TMU and its associated waste heat radiators, allowing the use of a conventional, rigid heat pipe design. Variable conductance heat pipes transport the TMU waste heat to the radiators. The system is designed to provide a –10°C heat sink for the cooler that is stable to +/- 5°C. The TMU cold block is conductively coupled to the FPA via the USU cryogenic Thermal Strap. Thermal gaskets, made from 0.002” thick gold foil, are used to reduce the thermal resistance in the bolted joints between the FPA, Thermal Strap, and cooler Cold Block. Gold plated Ultem “hard-radiation shields” are designed to reduce the thermal radiative load on the cryogenic thermal strap while reducing the amount of multi-layer insulation (MLI), and its associated contamination, in the vacuum space. A flexible welded bellows and the Vacuum Shroud provide a closed vacuum space around the cooler Cold Head and FPA. The closed vacuum space allows for cold focal plane ground testing with the payload in an ambient laboratory environment. The cooler integration is illustrated in Figure 6, the actual hardware in Figure 7. During flight operations the payload is primarily configured in either of two modes; stand-by
or image. The cooler system continuously will maintain the FPA at operational temperature (75 K), this is done to reduce the detrimental effects of thermal cycling the FPA. During standby mode a gold narcissus mirror is rotated into the Lyot stop of the system, thereby reducing the radiative loads to the FPA. Prior to imaging the narcissus mirror is rotated out of view, imparting the OA radiative and image scene loads onto the FPA. AFRL / SNL CHARACTERIZATION TESTING
The developmental nature of the cooler and FPA potentially posed schedule and performance risks to the overall MTI program. To mitigate this risk SNL planned for early characterization testing of the cooler system and FPA. To support this test SNL fabricated a mock-up OA and contracted with SBRC to provide a Thermal Mass Mock-up (TMM) of the flight FPA. The
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existence of a TRW 6020 pulse tube at the AFRL, and its similarity to the MTI flight cooler, allowed for the characterization testing to be performed with “flight-like” hardware at AFRL. The different cold block configuration of the MTI cooler required that an additional adapter block be
inserted between the 6020 cold block and the thermal strap. The total lift required from the TMU was approximately 2.25 W (depending on configuration mode), which for purposes of the thermal budget, is divided into radiative, conductive, and active FPA power loads. FPA power dissi-
pation was simulated by powering fixed resistance heaters placed on the TMM and different operating modes simulated by varying the voltage across the heaters. The hardware was integrated and the testing performed at AFRL to characterize the following and the results of these tests are detailed below: 1. Cooldown from ambient - measure transient performance of the system. 2. Operational temperature variation – measure system performance at different focal plane operating temperatures. 3. Reject temperature variation - measure system performance while operating at different cooler rejection temperatures. 4. Stand-by and imaging loads - measure the cooler lift required while maintaining the focal plane temperatures under each load condition. 5. Hardshield effectiveness - MLI was removed from the thermal strap to measure the effectiveness of the Ultem hard shields. 6. Pulse tube orientation – the test system was rotated from vertical to horizontal to determine whether gravitational effects would alter its performance. 1. Cooldown from Ambient
The system required approximately 5 hours to cool the focal plane from 296 K to 75 K,
Figure 8. Cooldown from Ambient, Environmental Temperatures.
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Figure 9. Compressor Input Power as a Function of Focal Plane Temperature.
cooler reject was maintained at 285 K during this period. Figure 8 shows the interface temperatures (focal plane, pedestal, thermal strap, adapter block, and cold block) over the cooldown period. The cooldown was performed in the image configuration with no active heating of the focal plane. 2. Operational Temperature Variation
Figure 9 is a plot of the cryocooler compressor input power as a function of the focal plane temperature, for each of the three focal plane temperature points. The data recorded for each of the focal plane temperatures is tabulated in Table 3. The temperature drops between the cryocooler cold block adapter and the focal plane pedestal is recorded in Table 4 for each focal plane temperature. Note that although the total load is not measured it is still imposed and only the active FPA power is shown.
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3. Rejection Temperature Variation
4. Standby and Imaging Loads
Table 6 is a tabulation of the steady state data for the standby and imaging configurations. 5. Hardshield Effectiveness
During the previous tests the thermal strap was completely wrapped with MLI. The system
was warmed to ambient, re-pressurized, and a section of thermal strap wrap was removed. The system was then returned to operational temperatures. The results are tabulated in Table 7. This test was run in the image configuration.
6. Pulse Tube Orientation Table 8 tabulates the data recorded with the pulse tube in the vertical and horizontal positions. Both data sets were taken with the cooler only (FPA de-mated) using a heater on the cold block to apply the load.
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SUMMARY The MTI program requires a focal plane cooled to 75 K for the duration of its three-year life. The program selected a variant of the TRW 6020 pulse tube to fulfill this requirement
and,through the non-redundant integration into the MTI payload, intends to demonstrate the utility of active cooling, and related technologies, for space missions. Although an essential element, the cooler ultimately plays a supporting role to the MTI program, therefore SNL teamed with TRW and AFRL to develop, integrate, and test the cooler system, to gain confidence in the ability of the cooler system to perform. A rigorous program was undertaken at TRW to develop and qualify the cooler. The result-
ing TMU and CCE can produce in excess of 3 W at 62 K while rejecting to –10 °C. A test series conducted at AFRL using the similar TRW 6020 pulse tube characterized the integrated flight system. The tests showed that the system could be cooled to operating temperatures in 5 hours while maintaining small temperature gradients. The average change in compressor power for a change in focal plane temperature, near the 75 K design point, is approximately 1.8 W/K. The average change in compressor power for a change in cooler reject temperature is 0.4 W/K. The non-imaging load for a 75 K focal plane is 1.98 W, requiring 57.8 W to the cryocooler compressor and approximately 62 % of full stroke. The imaging load increases to 2.23 W with a corresponding increase in compressor input power of 69.6 W and approximately 66 % of full stroke. The thermal resistance of the USU thermal strap, measured from the non-imaging test, was 1.6 K/W. No significant change in the thermal load was noted with the removal of a section of MLI around the thermal strap or by changing the pulse tube orientation from vertical to
horizontal. ACKNOWLEDGMENT
The authors of this paper would like to acknowledge the efforts of the many supporting individuals at SNL, TRW, and AFRL.
REFERENCES 1. Burt, W. W. and Chan, C. K., “New Mid-Size High Efficiency Pulse Tube Coolers”, Cryocoolers 9, Plenum Press, New York (1997), pp. 173-182. 2. Davidson, D., “MTI/FPA test report”, Sandia National Laboratories, Albuquerque, New Mexico (1997).
3. Johnson, D. L., et al., “Performance Characterization of the TRW 3503 and 6020 Pulse Tube Coolers”, Cryocoolers 9, Plenum Press, New York (1997), pp. 183-193.
4. Williams, B., Jensen, S., and Batty, J. C., “An Advanced Solderless Flexible Thermal Link”, Cryocoolers 9, Plenum Press, New York (1997), pp. 807-812.
IMAS Pulse Tube Cooler Development and Testing C.K. Chan, T. Nguyen, R. Colbert, and J. Raab TRW Space & Technology Division Redondo Beach, CA 90278
R.G. Ross, Jr. and D.L. Johnson Jet Propulsion Laboratory California Institute of Technology Pasadena, CA 91109
ABSTRACT
The Integrated Multispectral Atmospheric Sounder (IMAS) cryocooler has been developed to provide 0.5-watt cooling at 55K in a lightweight compact configuration. The design goal for the cooler was a factor-of-three in size and mass reduction over the AIRS cooler design, with a compressor input power goal of less than 50 W/W — 50% lower than the AIRS cooler at the 0.5-watt cooling capacity. The developed cooler incorporates a vibrationally balanced compressor with heat spreader in the center plate; this further increases the total system efficiency by maintaining a temperature difference of <5°C between the after-cooler and the heat-rejection interface. Two different coldheads have been designed for the IMAS application: an integrallinear option, and a split-coaxial option. The integral-linear option offers efficient performance, and a single warm mechanical/thermal interface. The split-coaxial option offers compactness and some cold interface system advantages. The development of the IMAS cryocooler is presented together with thermal, vibration, and EMI performance data gathered on the cooler both at TRW and at JPL. INTRODUCTION
The Integrated Multispectral Atmospheric Sounder (IMAS) instrument is an advanced concept instrument being examined by JPL as a second-generation atmospheric sounder for making precision air temperature measurements from space. Key to reducing the mass and power of the IMAS instrument is achieving a new long-life cryocooler with significant mass and size reductions over the AIRS cryocooler1-5 for the needed 0.5-watts at 55 K focal plane load. This cooling requirement falls midway between the robust cooling capability of existing AIRS-class coolers (1.75 W at 55K) and the capability of miniature long-life coolers, such as the TRW mini pulse tube,6,7 which has a capacity of approximately 0.5 watt at 75 K. The IMAS cryocooler development effort was carried out by TRW, with JPL as an active integrated product team partner. Figure 1 shows a size comparison between the new IMAS cooler and the larger AIRS redundant cooler system. Cryocoolers 10, edited by R. G. Ross, Jr. Kluwer Academic/Plenum Publishers, 1999
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Figure 1. IMAS cryocooler (left) in comparison to the AIRS cooler system (right).
Design Motivation and Requirements
The design load of 0.5 watt at 55 K was derived from detailed calculations of the operational IMAS cryogenic cooling loads from beginning-of-life to end-of-life for the IMAS cryosystem conceptual design shown in Fig. 2. In this system, the compact pulse-tube cryocooler is mounted directly to an instrument-mounted radiator to which ambient heat from the operating cooler is rejected at approximately 270 K. Connection to the 55K focal plane is made using a highconductance coldlink assembly containing a flexible link to accommodate relative motion created during cooldown and launch. The cold link is supported from the focal plane and provides minimal loads into the pulse tube. Other fundamental ground rules for the cryocooler system design include: • Use of a single high-reliability non-redundant cooler to avoid the significant mass and power penalty associated with redundant cryocoolers • Cooler efficiency goal of 50 WAV with a 0.5 W load at 55K — 50% better than the excellent AIRS cooler at the same power level (see Fig. 3) • Total input power goal of 50 watts, and total mass goal of 10 kg, for the mechanical cooler together with its drive electronics • Compressor and pulse tube reject temperature less than 5°C above thermal interface temperature to maximize operational efficiency. • Cooler drive frequency fixed at 54 Hz and synchronized to the instrument electronics to eliminate pickup of asynchronous vibration and EMI noise from the cryocooler • Cooler drive electronics isolated from input power bus; EMI consistent with MIL-STD-461C
Figure 2. Conceptual design of the IMAS
cryosystem.
Figure 3. Comparison of IMAS efficiency goal
with AIRS efficiency performance.
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Figure 4. Split coaxial (left) and integral linear (right) pulse tube concepts investigated as part of the IMAS cryocooler development effort.
IMAS COOLER DEVELOPMENT
Because the needed cryocooler performance was outside of the capabilities of existing cryocoolers, a collaborative TRW/JPL teaming approach was established to achieve the necessary cryocooler technology advances. This approach led to the investigation of two pulse tube configurations for the IMAS cooler (see Fig. 4): 1) an integral-linear approach built on the heritage of the highly successful AIRS cooler, and 2) a new split-coaxial design with the promise of reduced mass and improved instrument interfaces. Fundamental to both designs was a new TRW two-piston head-to-head low-vibration compressor. The high-capacity miniaturized compressor is derived from a joint effort between TRW and Oxford University to develop a next-generation generic flexure bearing compressor of lighter weight, better efficiency, lower EMI, and high-capacity (same class capacity as the AIRS compressor). The majority of the light-weighing potential for a pulse tube cooler lies in the compressor. The IMAS compressor size and mass (as shown in Fig. 5) have been greatly reduced over the current generation AIRS-type compressor by using an entirely new approach to the basic layout of the motor design. The innovative new patented motor and moving coil suspension concept allows great force and stroke in a small package along with lower radiated DC and AC magnetic fields. The compressor motor has a 150 W maximum power capability and has been qualified to a 14.14 Grms launch vibration environment.
Figure 5. Characteristics of the IMAS cooler, shown schematically with integral linear pulse tube.
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Figure 6. IMAS cryocooler showing linear pulse tube (left) and heat spreader (right). The new compressor concept was designed to operate efficiently over a range of swept
volumes and operating frequencies by variation of the piston diameter, fill pressure, and moving mass. Two versions have been built to date: the 4.4 cm3 swept-volume IMAS compressor, shown in Fig. 6, and a larger-piston generic 6.5 cm3 version that has the same motor design and external
dimensions as the IMAS compressor. The moving mass and the smaller piston diameter in the IMAS compressor were selected to provide the needed capacity at the required IMAS operating frequency of 54 Hz, which is established by the need for synchronization with the instrument
data-acquisition functions. In the compressor, the piston shafts are supported fore and aft by flexure springs, which are
designed and test-verified for infinite fatigue life. Non-contacting tight gas clearance seals between the piston and the cylinder provide the compression seal in the traditional Oxford-cooler manner. The absence of rubbing, maintained by the flexure bearings, allows multi-year-life capa-
bility. To assure helium retention well past the useful life of the cooler, aluminum-jacketed Crings located between the helium working fluid and ambient provide hermetic metal-to-metal
seals. Piston motion in the new compressor is actuated by moving-coil linear drive motors. The stator motor field is generated by high-strength NdBFe permanent magnets using a cobalt iron return flux path and iron pole pieces. These high-performance materials maximize the field in the
motor gap for the least weight, and as with the AIRS motors, provide the potential of >90% motor efficiency. Internal wiring is stranded, ETFE-insulated or Kapton flexible cable. All
wiring exits the bulkhead through ceramic-insulated pins in feedthroughs wired through a pigtail to common D-shell connectors for the power, thermometry, and accelerometers.
To maximize the operational efficiency of the cooler when integrated into the instrument, the cooler has been designed for direct mounting to an instrument radiator or heatpipe interface with less than a 5°C thermal rise above the heatsink temperature. Temperature rise from heatsink to
cooler has been a critical issues with previous coolers, and achieving minimal rise is an important design focus for the IMAS cooler. The IMAS cooler uses highly thermally conductive aluminum center plate and end-caps to remove the compressor heat while providing a good thermal expansion match, light weight, and ease of fabrication. Figure 6 shows the closely-coupled integral heatsink/structural mounting interface on the rear side of the IMAS prototype cooler. The IMAS baseline coldhead is the integral-linear design shown in Fig 6. This design is derived from the successful AIRS coldheads. The linear configuration offers design maturity,
higher efficiency, elimination of flow straightener, demonstrated producibility and ease of interface. The H-bar behind the coldhead provides structure rigidity and the heat conduction path from
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Figure 7. Prototype coaxial pulse tube coldhead.
the warm end heat exchanger to the center plate of the compressor. In addition to the linear configuration, a coaxial coldhead, derived from TRW IRAD, was also designed for the IMAS cooling load. The coaxial configuration, shown in Fig. 7, offers an alternative focal plane interface and a potential cooler mass reduction through elimination of the pulse tube structural support. Either the linear or the coaxial coldhead can be integrated with the IMAS compressor in an integral or in a split configuration. The coaxial split configuration provides an alternate for an advanced focal plane design, especially if the base of the regenerator can be thermally mounted onto a lower-temperature radiator and the focal plane can be mounted directly onto the coldtip. The IMAS coaxial configuration offers a 50% reduction in coldhead mass because it does not require the H-bar. From the system point of view, the cooling load of the coaxial split configuration is also reduced. The mass of the IMAS coolers in both linear and coaxial configurations is summarized in Table 1. IMAS Cooler Electronics Another key issue addressed by the IMAS cooler design is compatibility with the sensitive IR and millimeter-wave detectors and electronics. To reduce noise input to the detector circuits to very low levels, the IMAS cooler baselines the use of TRW’s flight qualified, radiation hardened, and high efficiency AIRS/SMTS/TES cooler electronics family. These electronics provide electrical isolation from the spacecraft power bus and use digitally generated piston waveforms to provide precise closed-loop suppression of generated vibration and to provide millikelvin temperature control of the cooler coldfinger so as to achieve the needed fractional millikelvin stability at the focal plane. The key driver on the temperature control is the fluctuating temperature of
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Figure 8. Packaging concept and design goals for the new IMAS cryocooler drive electronics.
the cryocooler heatsink due to orbital variations in the effective thermal radiation environment, including periodic solar input in some circumstances. From previous measurements of space coolers it is known that every 5°C change in heatsink temperature maps into approximately a 1 K change in coldtip temperature for the same stroke or input power.8,1 In addition to the vibration and temperature control functions, the software programmable drive electronics also provide for cooler operational control, and acquisition and transmission of cooler operational data to the IMAS instrument. Relays in the electronics short the compressor motor drive coils during launch to prevent excessive launch-induced piston motion. Design goals for the IMAS electronics are summarized in Fig. 8. As an augmentation to the AIRS/SMTS/TES cooler electronics, the IMAS cooler develop-
ment effort is exploring incorporating active ripple current suppression into the cooler electronics. Excessive ripple current fed onto the input power bus is a common problem for all lowfrequency linear drive coolers,9,10 and has been solved to date by the addition of a separate ripple
filter in the spacecraft power system. The new ripple-suppression cooler electronics being examined as part of the IMAS cryocooler development effort could result in a savings of several kilograms of total system mass, and greatly improve spacecraft accommodation. Preliminary test data of the new IMAS design indicates the feasibility of ripple current reduction to levels consis-
tent with typical spacecraft power systems. The weight of the IMAS cooler electronics with ripple current suppression is estimated to be around 6.5 kg, as is shown in Table 1. Without the integral ripple filter, AIRS measurements1 suggest that the projected electrical efficiency is well modeled as P(total input) = P(compressor input)/0.85+5 watts. Because of the addition of the ripple current filter, the IMAS electronics is projected to be P(total input)=P(compressor input)/0.78 + 2 watts. COOLER PERFORMANCE MEASUREMENTS
As part of the development process, extensive measurements of the performance of the IMAS cooler have been carried out, both at TRW and at JPL. These are summarized in the area of thermal refrigeration performance, vibration performance, and EMI performance.
Thermal Performance Figure 9 describes the measured refrigeration performance of the IMAS S/N 102 cooler as a
function of stroke, input power, coldblock load, and coldblock temperature. In the process of developing the IMAS cooler, several pulse tube coldhead designs have been both analytically and experimentally evaluated. The S/N 102 cooler represents the best performance achieved as of December 1997, prior to delivery of the first IMAS unit to JPL. As noted in Figure 10 the
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Figure 9. Thermal performance of the S/N 102 IMAS cryocooler as a function of input power, coldblock load, and coldblock temperature.
Figure 10. Load lines for the linear-integral and split-coaxial pulse tubes in comparison with the load line for the AIRS cooler at a similar power level and 300 K heatsink temperature.
performance of the IMAS S/N 102 cooler with linear coldhead is better than that of the coaxial coldhead at this point in its development. It also has surpassed the performance of the AIRS cooler1 at the same 0.5-watt at 55K power level. Self-Induced Vibration Performance As part of the exploratory testing effort, the self-induced vibration of the IMAS cooler was tested with both a Texas Instruments Standard Cooler CCA drive electronics (Part No. 2954026-2),
which generates a square waveform rather than a sinusoidal waveform, and with a low-harmonicdistortion sinusoidal-waveform laboratory drive electronics. Figure 11 compares the vibration results from the two tests. Although the self-induced vibration levels with the two electronics are quite similar, in the square-waveform case, about 20% of the cooling capacity at 75K was lost for the same input power.
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Figure 11. Self induced vibration of the S/N 102 IMAS cryocooler when powered by low-distortion lab electronics (top), and square-wave tactical cooler drive electronics (bottom).
EMI Performance Another feature of the smaller, lower-power motors in the IMAS cooler is lower levels of AC magnetic fields. Two sets of AC magnetic field measurements were made to quantify the IMAS cryocooler AC magnetic field emissions: 1) at a 7-cm distance, corresponding to the MILSTD-461C RE01 test specification9, and 2) at a 1-m distance, corresponding to a MIL-STD-462 RE04 test method. Figure 12 shows the measured RE01 magnetic field performance of the IMAS compressor at 75 watt input power, contrasted with that of the AIRS compressor shown for 105 watts of input power10; the data are plotted in decibels above 1 pT. The magnetic field emission levels of the IMAS cooler are quite low compared to other space coolers.9
Figure 12. Radiated magnetic fields of the S/N 102 IMAS cryocooler (left) in contrast to those from the larger AIRS cooler (right).
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SUMMARY AND CONCLUSIONS
The IMAS TechDemo cryocooler development has been carried out in close collaboration with the IMAS instrument development so as to maximize the performance of the overall instrument. The cooler development is a collaborative effort involving development activities at TRW, and cryocooler characterization testing at JPL. The state-of-the-art pulse tube cooler has demonstrated excellent thermal performance, light weight, low self-induced vibration and low magnetic field emission. Results have been presented detailing the overall cryocooler thermal performance
achieved, the cooler’s vibration and EMI attributes, and its mass properties. ACKNOWLEDGMENT
The work described in this paper was carried out by TRW, Inc. and the Jet Propulsion Laboratory, California Institute of Technology; it was sponsored by the NASA EOS IMAS
TechDemo Project through an agreement with the National Aeronautics and Space Administration. REFERENCES
1. Ross, R.G., Jr., Johnson, D.L., Collins, S.A., Green K. and Wickman, H. “AIRS PFM Pulse Tube Cooler System-level Performance,” Cryocoolers 10, Plenum Publishing Corp., New York, 1999. 2. Chan, C.K., Raab, J., Colbert, R. , Carlson, C, and Orsini, R.,“Pulse Tube Coolers for NASA AIRS Flight Instrument”, Proceedings of ICEC 17, 14-17 July 1998, Bournemouth, UK.
3. Ross, R.G., Jr. and Green K., “AIRS Cryocooler System Design and Development,” Cryocoolers 9, Plenum Publishing Corp., New York, 1997, pp. 885-894. 4. Chan, C.K., et al., “Performance of the AIRS Pulse Tube Engineering Model Cryocooler,” Cryocoolers 9, Plenum Publishing Corp., New York, 1997, pp. 195-202. 5. Chan, C.K., et al., “AIRS Pulse Tube Cryocooler System,” Cryocoolers 9, Plenum Publishing Corp., New York, 1997, pp. 895-903.
6. Tward, E., et al., “Miniature Long-Life Space-Qualified Pulse Tube and Stirling Cryocoolers,” Cryocoolers 8, Plenum Publishing Corp., New York, 1995, pp. 329-336.
7. Chan, C.K., Jaco, C. and Nguyen, T., “Advanced Pulse Tube Cold Head Development,” Cryocoolers 9, Plenum Publishing Corp., New York, 1997, pp. 203-212. 8. Ross, R.G., Jr. and Johnson, D.L., “Effect of Heat Rejection Conditions on Cryocooler Operational Stability,” Advances in Cryogenic Engineering, Vol. 43, 1998. 9. Johnson, D.L., Collins, S.A. and Ross, R.G., Jr., “EMI Performance of the AIRS Cooler and Electronics,” Cryocoolers 10, Plenum Publishing Corp., New York, 1999. 10. Johnson, D.L., et al., "Cryocooler Electromagnetic Compatibility," Cryocoolers 8, Plenum Publishing Corp., New York, 1995, pp. 209-220.
Development of a 1 to 5 W at 80 K Stirling Pulse Tube Cryocooler Y. Hiratsuka and Y. M. Kang
MEC Laboratory, DAIKIN INDUSTRIES, LTD. 3 Miyukigaoka, Tsukuba 305-0841, Japan Y. Matsubara
Atomic Energy Research Institute, Nihon University 7-24-1 Narashinodai, Funabashi 274-0063, Chiba, Japan
ABSTRACT
Pulse tube cryocoolers have recently received considerable attention due to their advantages and reliability and to their reduced cold head vibrations compared with other regenerative
cryocoolers ( e.g., GM and Stirling cryocoolers ). The pulse tube cryocoolers used in this study are similar to Stirling cryocoolers but they do not require any moving displacer in the cold head. This is beneficial for mechanical simplicity, reliability, reduction of induced vibrations. Performance comparable to Stirling cryocoolers are achievable in terms of cooling power, cooling temperature and efficiency. To cool high-Tc superconductive devices, we have design ed and tested a prototype of Stirling pulse tube cryocooler that has a cooling capacity of 1 to 5W at 80K and a dual-piston linear drive motor compressor. Presently for a compressor input power of 200W, this prototype has achieved a cooling capacity of a few watt at 80K. Measured test parameters include cooling loads, cooling temperature, input power, operating frequency and pressures. We evaluated two types phase shifter mechanisms, a double inlet type and an inertance tube type. To aid the design of cryocoolers, we developed a numerical analysis technique and validated it with the experimental results. This report presents experimental and numerical results for our pulse tube cryocooler design. To study the compatibility for SQUID applications, we also measured the magnetic noise of cryocooler operation with a fluxgate magnetometer outside the magnetically shield room.
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INTRODUCTION
Pulse tube cryocoolers, which have no moving parts at the cold section, are more attractive than other small cryocoolers because of their high reliability, simpler construction, and lower vibration levels. Pulse tube cryocoolers cool by adiabatic expansion of a gas piston. The gas piston phase angle is controlled and optimized by an orifice valve. These types of cryocoolers may be roughly divided into GM and Stirling pulse tube cryocoolers, depending on what type of compressor is used. In 1992, TRW1 developed a Stirling type pulse tube cryocooler with a cooling capacity of 1W@80K, similar to power required for a Stirling cryocooler. In 1994, Haruyama2 and Kanao3 designed prototype pulse tube cryocoolers similar to the TRW cryocooler. To develop the feasibility of fabricating a long-life and a low-noise cryocooler, we designed a prototype cryocooler. We also developed a numerical analysis model of pulse tube cryocooler performance for optimizing cycle efficiency. The reports for SQUID applications cooled by a small cryocooler (e.g., Stirling and Pulse tube cryocoolers) have been published4,5 In a Stirling type cryocooler the compressor pistons are driven by linear motors tuned to a drive frequency of 50 Hz. Each piston is moved by feeding an alternating current through a coil that is located in a permanent magnetic field and connected to the piston. This drive system produces magnetic noise. To study the influence of cryocooler magnetic noise on SQUID measurement systems, we measured the magnetic noise of the cryocooler located outside the magnetically shielded room to the magnetic field generated by the cryocooler.
This paper describes the development of a prototype cryocooler with a design target of 30W input power and a cooling capacity of 1W at 80K or 100W input power and a cooling capacity of 5W at 80K. We describe both measured and calculated results for the cryocooler performance, including measurements of the magnetic field around of the cryocooler. GENERAL DESIGN
A schematic drawing of the proposed pulse tube cryocooler is shown in Figure 1 and its specifications are shown in Table 1. The cryocooler is a split Stirling type double inlet pulse tube and opposed pistons driven by a linear motor. We developed the two sizes of compressor, its swept volume is 3 and 7cc at 60 and 200W input power, respectively, and the compressor is 68 and 90mm in outer diameter and 130 and 180mm in length, respectively. The regenerator consists of #400 SUS stacked mesh sheets. The compressor is connected to the expander by tubes of various lengths ( 100 to 300mm ) . A thermocouple and heater are mounted on the cold head, which is made of copper. The piston position is monitored using a laser vibrometer. A pressure transducer is mounted near the compressor discharge head and before the orifice and double inlet valve, and used to determine the mass flow rate through the orifice valve. The compressor input power is constant. These measurements are used to calculate both the pressurevolume (P-V) work of the compressor and the equivalent P-V work of the expander.
Figure 1. Schematic drawing of the prototype cryocooler.
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EQUIVALENT P-V METHOD
The equivalent P-V work is calculated both an experimental value and a computational value to use for sensitivity studies, to identify key parameters, and cooling performance (presented by
Y.Matsubara6). The gas in the pulse tube can be divided into the three parts : I) the hot end of the pulse tube, II) the gas displacer, and III) the cold head of the pulse tube. From an experimental value, mass flow rate through the orifices calculated by the pressure obtained from the experimental result. The mass flow rate of orifice, double inlet valve and regenerator can written as
where of, dv, rg is coefficient of discharge at orifice, double inlet valve and regenerator, respectively, we used values measured during a steady-flow test. The gas displacer volume in the pulse tube can therefore be determined by solving the differential equations. The coefficient of discharge can be determined by ADS ( Automated Design Synthesis ) so that the maximum cooling capacity can be determined. The equivalent P-V work and the cooling capacity at cold head can then be calculated as follow:
where Hrg is a regenerator loss, a conduction loss and a radiation loss, in the case of infinity large
matrix heat capacity a regenerative effectiveness of equation becomes Then for packed screens, the friction facter and Nusselt number is referred to a experimental counter flow equations. Moreover is a pulse tube enthalpy loss, which is calculated similar to the shuttle loss, and can be expressed as follows:
where k is the thermal conductivities of helium gas in the pulse tube,
stroke,
is the gas displacer
is the gap of between the pulse tube and the gas displacer.
NUMERICAL ANALYSIS MODEL8
To understand the process occurring in the pulse tube cryocooler in detail and to design it, a numerical analysis model has been developed. The cryocooler was divided into a number of subsections, and each subsection was divided into control volumes (Figure 2 and Table 2 ). Each subsystem can exchange work, heat, and mass with its surroundings through its section boundaries. For simplicity, expansion and compression in the reservoir were considered to be adiabatic.
Figure 2. Mathematical model.
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Assuming the working fluid is an ideal gas, the equation of state and the conservation of mass, momentum, and energy for the control volumes can be written as
the temperature of enthalpy transferred across the interface of a control volume is supplied to the adjacent upstream control volume. For the matrix in the regenerator, The mass flow rate at the orifice can be written as
The differential equations were integrated using the Runge-Kutta Gill method. The conditions for iteration convergence were that the relative discrepancies in temperature and pressure
between two cycles were less than tube were less than 0.1W.
and that the energy reserve in the regenerator and pulse
RESULT & DISCUSSION
Cryocooler Performance & Numerical Analysis Result Figure 3 and 4 shows the measured and simulated cooling performance, respectively, of the
prototype pulse tube cryocooler. For 60W input power, the no-load temprature was 57K, the cooling capacity was 0.65W at 80K, the compressor efficiency was 50%, and the P-V work was about 30W. For 200W input power, the no-load temperature was 50K, the cooling capacity was 2.0W at 80K, the compressor efficiency was about 55%, and the P-V work was about 110W. The simulated performance agrees roughly with the measured performance, indicating that
the numerical analysis model accurately represents the operating cycle of the cryocooler, and that such models can be effectively used to design pulse tube cryocoolers. To improve the cooling capacity, a phase shifter was inserted in place of the orifice valve by using an inertance tube (Figure 5). The cooling capacity with the inertance tube increased about 20% compared to that with the double inlet (Figure 6). The no-load minimum temperature decreased to 48K and the cooling capacity was 2.4W@80K. Moreover the cooling capacity can increase about 10% with replacement a conect tube of 300mm length by one of 150 mm length.
Figure 3. Cryocooler performance at 1W class.
Figure 4. Cryocooler performance at 5W class.
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Figure 5. Method of phase shifter.
Figure 7. Heat loss analysis at 1W class (80K).
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Figure 6. Influence of the phase shifter.
Figure 8. Heat loss analysis at 5W class (80K).
Comparison of Exparimental & Various Models Results The results of various simulation models are presented in Table 3. We observe that results of the equivalent P-V method are identical to those of the numerical analysis model and the experimental. After that, it needs futher study about other parameter to increase accuracy. Heat Loss Analysis
Figures 7 and 8 show the heat loss analysis of the expander for 1 and 5W class cryocoolers, respectively. For the 1W class cryocooler, the measured equivalent P-V work was 3.3W, composed of 0.65W of cooling, 0.7W of combined conduction and radiation loss, and an unknown loss of 2.0W. Similarly, for the 5W class cryocooler, the measured equivalent P-V work was 8.6W, composed of 2.0W of cooling, 1W of combined conduction and radiation loss, and an unknown loss of 6.6W. Our numerical analysis showed that this unknown loss may consist of regeneration loss and a pulse tube loss. To increase the cooling capacity, these losses must be identified and minimized.
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ESTIMATION OF MAGNETIC NOISE Measurement of Magnetic Field
For application to high-Tc SQUID, we developed a split Stirling type pulse tube cryocooler. To study the influence of magnetic noise on SQUID, we first made measurements outside the magnetically shielded room. We measured DC and AC magnetic fields with a Pb permalloy
shield around the compressor and without a shield on the flux gate type magnetometer (Figure 9). The results are shown in Figures 10 and 11, respectively. The magnetic noise of the shielded compressor was small compared to that without a shield and that the magnetic field decreased with increasing distance from the compressor. Calculation of The Magnetic Field
At a distance relatively large compared to the coil diameter, each coil can be considered as a magnetic dipole having a dipole moment. The strength of a dipole magnetic field was calculated using the equations shown in Table 4. Comparison of the measured and calculated AC magnetic field (Figure 12) indicates that at a distance of 1m the AC magnetic field was the same level as the background environmental magnetic fields.
Figure 9. Measurement apparatus of magnetic field.
Figure 10. DC magnetic field.
Figure 11. AC magnetic field.
Figure 12. Comparison of the experimental and the calculation data.
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CONCLUSION
We developed a double inlet split Stirling pulse tube cryocooler and analyzed its performance experimentally and numerically. The significant results of our research are:
(1) For 60W input power, the no-load temperature was 57K and the cooling capacity was 0.65 W@80K. For 200W input power, the no-load temperature was 48K and the cooling capacity was 2.4W@80K. (2) The simulated performance from our numerical analysis model roughly agreed with the
performance determined from our experimental measurements. (3) Heat loss was estimated by the equivalent P-V work in the expander and partitioned into the fraction
used for cooling, heat transfer to the surroundings, and a portion that was unaccounted for. (4) At a distance of 1m the AC magnetic field or with a Pb permalloy shield around the compressor was the same level as the background environmental magnetic fields.
REFERENCE 1. E. Tward : "Miniature Pulse tube Cooler":ICC 7th `92 2. T. Haruyama : "Cooling performance of a Prototype Miniature Pulse Tube Refrigerator with a Flexure Spring Compressor" : ICC 8th `94
3. K. Kanao : "A miniature pulse tube refrigerator for temperature below 100K": ICEC15th `94 4. M.David : "80K Miniture Pulse Tube Refrigerator Performance" : ICC 9th `95 5. H.J.M. ter Branke : "MAGNETIC NOISE OF SMALL STIRLING COOLERS: Advances in Cryogenics Engineering Vol. 39 `94 6. Y. Matsubara: "Work-loss Distribution on GM-type Pulse Tube Coolers" : Cryogenic Engineering
Vol. 33 No.4 `98 (in Japanese) 7. M. Tanaka : "Flow and Heat Transfer Characteristics of the Stirling Engine Regenerator in an Oscillating Flow" : JSME Vol. 33 `90 8. I. Urieli: "Computer simulation of Stirling cycle machines" : 12th IECEC `77
Development of a 2 W at 60 K Pulse Tube Cryocooler for Spaceborne Operation V. Kotsubo, J.R. Olson, and T. C. Nast Lockheed Martin Advanced Technology Center Palo Alto, CA, USA 94304-1191
ABSTRACT Lockheed Martin has been developing pulse tube systems principally for spacecraft operation for several years. Major emphasis in our pulse tube development has been on improving the thermodynamic efficiency, since other benefits such as low vibration, enhanced reliability, and reduced cost are well established. We have developed a pulse tube cooler which provides 2.4 W at 60 K and 4.7 W at 80 K, achieving PV specific powers of 22 WAV at 60 K and 12 W/W at 80 K, with a no-load temperature of 32 K. This system consists of an in-line pulse tube cryocooler, our flight-qualified, flexurebearing, clearance seal compressor, and our flight-qualified electronic controller. INTRODUCTION Lockheed Martin has been developing advanced mechanical cryocooler systems for spacecraft applications since 1987, and recently introduced a complete flight-qualified Stirling cooler system in 1995.1 This system consists of an Oxford heritage, flexure-bearing compressor with dual-opposed pistons for low-exported vibration, a driven displacer with an active balancer, and a state-of-the-art electronic controller that drives the compressor, displacer and balancer, and provides feedback control of exported vibration and of the coldtip temperature. As pulse tube technology continued its rapid evolution, it became evident that a highly efficient pulse tube coldhead would eventually replace the Stirling displacer as the technology of choice because of the obvious advantages of the no-moving-parts coldhead, including no vibration at the cold interface, increased reliability, elimination of the intricate assembly of a long-life displacer, and elimination of the displacer drive and vibration cancellation electronics. Lockheed Martin began in 1994 to develop pulse tube cryocoolers for flight applications, with the goal of achieving efficiencies competitive with those of Stirling coolers. Because of the maturity of our flight-qualified compressor and electronic controller, we targeted a coldhead specifically to replace the displacer for this system. As a result of this effort, we successfully developed the Mark III pulse tube cooler, demonstrating 2.4 W cooling at 60 K, 4.7 W of cooling at 80 K, and 7.0 W cooling at 100 K, with 53 W of PV power and 100 W of compressor power, with a heat rejection temperature of 295 K. Despite this relatively low motor efficiency, the overall cooler efficiency above 60 K is comparable to or exceeds that of almost all reported single-stage coolers. We have applied the same design methodology used to develop the Mark III pulse tube to develop several other pulse tubes, reported in other publications in this conference. We have
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achieved 0.30 W cooling at 65 K with 15 W total compressor power in a miniature pulse tube2
sponsored by a DARPA program for HTS satellite communications, and by a NASA-GFC program for the development of a miniature cryocooler. We achieved 9.66 W cooling at 77 K with 175 W total compressor power in a large pulse tube, sponsored by the same DARPA HTS program.2 Under Air Force Phillips Laboratory sponsorship, we developed a two-stage intermediate bypass pulse tube which simultaneously provided 0.32 W at 35 K and 0.60 W at 56 K with 95 W of PV power, and 175 W total compressor power.3 SYSTEM DESCRIPTION Coldhead Description The Lockheed Martin Mark III pulse tube is a single-stage coldhead in an in-line configuration, operated at 44 Hz. The regenerator tube and pulse tube were thin-walled titanium. The particular unit tested had a 5-cm-long transfer line between the compressor body and the coldhead, but a longer transfer line could be used with only a small reduction in power efficiency.2 The phenomenology in the open literature on regenerators,4 gas wall interactions,5 and compression space behavior6 is sufficiently mature to allow the optimized design of highly efficient pulse tubes. Compressor The compressor is our flight-qualified L-2010 compressor1 with dual head-to-head pistons, a
common compression space, and
swept volume. The compressor utilizes clearance seal
flexure-bearing technology, and is designed for low vibration, high reliability and long life. The
piston diameters are 20 mm, instead of the 17 mm diameters of the 1710C compressor in Ref. 1. The motor is a moving coil design, and uses Kaman eddy current sensors to monitor position. The compressor body is made of aluminum for good heat rejection. The compressor envelope lies within a right circular cylinder with a 112 mm diameter and a 274 mm length, and the compressor assembly has a mass of 7.2 kg. Using feedback vibration control, the residual axial forces have been reduced to less than 0.04 N, and the residual lateral forces to less than 0.4 N. Electronic Controller The flight-qualified electronic controller1 contains electronics to drive two compressors and a single Stirling displacer and its active balancer. Digital Error Correction System (DECS), a Lockheed-Martin developed and patented control algorithm, provides feedback control on the piston and displacer motions for exported vibration reduction and temperature control. The controller can accept input signals from various temperature sensors, accelerometers, and load cells for use by the control algorithms. The 28 Vdc controller uses pulse width modulation (PWM) amplifiers with 80% power conversion efficiency. Tare power for the controller is 16 W. The controller mass, without cabling, is 5.2 kg, including the unnecessary displacer electronics, and the envelope is 195.6 mm wide by 222.2 mm long by 198.1 mm high. The weight, size, and tare power will all be reduced for electronics required for the pulse tube system, since the functions for the displacer and balancer are eliminated. PULSE TUBE PERFORMANCE Pulse tube thermodynamic performance tests were performed with the coldtip only mounted in a vacuum can, and wrapped with several layers of MLI. Heat was rejected to a coldplate cooled by chilled circulating water at 295 K. Figure 1 shows the cooling power of the Mark III pulse tube as a function oft emperature, with 100 W of compressor power, 53 W of PV power, and a 295 K heat rejection temperature. Shown are data for a regenerator designed to maximize efficiency at 60 K, as well as data for a second regenerator, built after interest was expressed in achieving cooling at 35 K. Our methodology allowed us to redesign the regenerator to increase cooling capacity at 35 K at the expense of cooling capacity at higher temperatures. In this same manner, it would also be possible to improve
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Figure 1. Cooling power as a function of temperature for the Lockheed Martin Mark III pulse tube.
the higher-temperature performance with simple regenerator modifications, while sacrificing some of the low-temperature cooling capacity. The PV specific power is shown in Figure 2. Figures 1 and 2 are for maximum piston stroke, but our power efficiency is relatively insensitive to the stroke. This is shown in Figure 3, where we show the compressor-specific power as a function of the compressor power. At 100 K and 80 K, the specific power is nearly independent of compressor power, even down to just 20% of the maximum power. The 60 K data show an appreciable increase in specific power below about 60% of the maximum power. MOTOR IMPROVEMENTS
Since the original development of the 1710C compressor, there have been advances made in both magnetic materials and magnetic circuit designs. The existing L-2010 compressor uses outdated motor technology which used samarium cobalt magnets and soft iron. We will be replacing the magnets with higher strength neodymium iron boron, and the soft iron with high-permeability
Figure 2. PV specific power divided by cooling power as a function of temperature for the Lockheed Martin Mark III pulse tube.
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Figure 3. Total compressor power divided by cooling power, as a function of total compressor power, for the Lockheed Martin Mark III pulse tube. This cooler maintains its efficiency even at just 20% of full
power, except for the 60 K case where some performance is lost. The system on which these measurements were taken was not fully optimized, so it had about 20% less cooling power than the 60 K optimized system shown in Figures 1 and 2.
vanadium permendur. These simple drop-in replacements will reduce motor losses by about 40%, which will reduce the compressor power for the Mark III cooler from 100 W to 82 W and increase
motor efficiency from 53% to 65%. We are also developing a magnetic circuit upgrade which will reduce motor losses by an additional 40%, which will reduce the compressor power to 70 W and increase motor efficiency to 76%. This cooler will have a total specific power of 29 W/W at 60 K, 15 W/W at 80 K, and 10 W/W at 100 K. SUMMARY
Lockheed Martin has developed a high-efficiency single-stage pulse tube cryocooler for spacecraft applications, capable of 2.4 W cooling at 60 K and 4.7 W cooling at 80 K, with 53 W of PV power and 100 W of compressor power. This cooler reached a no-load temperature of 32 K. The compressor and electronic controller are flight-qualified, developed for our Stirling cooler. We are working on motor improvements which will reduce motor losses by 65%, which will reduce the compressor power to 70 W for the same cooling power. ACKNOWLEDGMENTS
This work was supported by internal research and development funding from Lockheed Martin. REFERENCES 1. Nast, T.C., Champagne, P.J., Isaac, D., Pryor, G.M., von Savoye, R.L. and Naes, L.G., “Design, Performance and Testing of the Lockheed-Developed Mechanical Cryocooler,” Proc. 8th International Cryocooler Conference, Vail, CO (1994). 2. Kotsubo, V., Olson, J.R., Champagne, P.J., Williams, B., Clappier, B., and Nast, T.C., “Development of Pulse Tube Cryocoolers for HTS Satellite Communications,” to be published elsewhere in these proceedings. 3. Olson, J.R., Kotsubo, V., Champagne, P.J., and Nast, T.C., “Performance of a Two-Stage Pulse Tube Cryocooler for Space Applications,” to be published elsewhere in these proceedings.
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4. Kays, W.M., and London, A.L., Compact Heat Exchangers, McGraw-Hill, New York (1984); Armour,
J.C., and Cannon, J.N., “Fluid Flow Through Woven Screens,” AIChE Journal (1968), p. 415. 5. Swift, G.W., “Thermoacoustic Engines,” J. Acoust. Soc. Am. 84 (4) (1998); Lee, J. P., Kittel, P., Timmerhaus, K.D., and Radebaugh, R., “Higher Order Pulse Tube Modeling,” Cryocoolers 9, Plenum Press, New York (1996); Olson, J.R. and Swift, G, “Acoustic Streaming in Pulse Tube Refrigerators: Tapered Pulse Tubes,” Cryogenics, Vol. 37 (1997), p. 769; Xiao, J.H., Yang, J.H., and Tao, Z.D., “Miniature Double Inlet Pulse Tube Cryocooler: Design by Thermoacoustic Theory Compared with Preliminary Experimental Results,” Adv. Cryo. Eng, 41B (1995), p. 1435.
6. Kornhauser, A.A., “A Model of In-Cylinder Heat Transfer with Inflow-Produced Turbulence,” 27th IECEC, Vol. 5, (1992), p. 5.523.
Performance of a Two-Stage Pulse Tube Cryocooler for Space Applications J.R. Olson, V. Kotsubo, P.J. Champagne, and T.C. Nast Lockheed Martin Advanced Technology Center Palo Alto, CA, USA 94304-1191
ABSTRACT
Lockheed Martin has developed a two-stage pulse tube cryocooler under contract to the Air Force Phillips Laboratory. The cooler is an intermediate bypass design in an in-line configuration. At a reject temperature of 295 K, the cooler simultaneously produced 0.6 W at 56 K and 0.32 W at 35 K. With no heat load on either stage, the cold stage reached 23 K and the intermediate stage reached 48 K. This cooler was driven by a swept volume flexure bearing compressor, developed by Lockheed Martin for low-cost cryocooler applications. We also observed evidence of dc mass flow through the intermediate capillary and speculate that this resulted in some loss in second-stage cooling power. We observed unstable behavior of the second-stage temperature which we also attribute to dc flow. We evaluated an etched foil regenerator. While analysis suggested significant improvement with this technology, tests in both pulse tubes and Stirling-cycle coolers showed poorer performance compared to the conventional screen regenerators. As part of this contract we also evaluated linear flexures for reduction of lateral exported compressor vibration. In a direct comparison with spiral flexures in the same compressor, the linear flexures reduced the lateral vibrations by a factor of three to four for the higher harmonics. INTRODUCTION A growing number of space flight cryogenic instruments require cooling at several tempera-
tures, either because of multiple detectors with differing operating temperatures, simultaneous cooling of detector and optics, or the reduction of parasitic heat loads from support structures and electrical leads. For very low temperatures, multi-staging may be necessary simply to be able to reach the required temperatures. Many configurations are possible for multitemperature cooling using various combinations of passive radiators for higher temperature stages, multiple coolers, and multistage coolers. To minimize system complexity, a multistage coldhead driven by a single compressor is advantageous over a multi-cooler system with several compressors. Using a pulse tube coldhead instead of a Stirling coldhead provides further simplification, since the pulse tube has no moving parts, minimal vibration, no displacer motor drive that dissipates power and requires electronic control, and no mechanical precision alignment associated with maintaining the non-contacting displacer clearance gap required for long-life operation. Supported by an Air Force Research Laboratory PRDA, we undertook a program to demonstrate the feasibility of a multistage pulse tube for cooling detectors at two temperatures. Targeted
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were specific cooling requirements of 0.4 W at 35 K and 0.6 W at 60 K. These presented a challenge because of the specific cooling capacity goals, and the fact that the 60 K first stage is substantially lower than the optimum temperature for maximizing the efficiency of the second stage. Using an intermediate bypass configuration,1 we succeeded in simultaneously producing 0.32 W at 35 K, 0.6 W at 56 K with 85 W of PV power at a reject temperature of 295 K. By lowering the heat rejection temperature to 280 K, we met the requirements, with 0.4 W at 35 K, 0.6 W at 56 K, and 90 W of PV power. During some of the tests we observed indications of dc flow2,3,4,5 circulating through the closed loop formed by the second-stage regenerator and pulse tube and the intermediate capillary. We believe that this resulted in some loss in second-stage cooling power but have not been able to quantify it. We also observed unstable temperature behavior in the pulse tube during one test, which we attributed to this flow. We evaluated etch foil regenerators6 for this cooler. Analysis using flow correlation obtained from CFD modeling7 indicated that etch foils would significantly improve the performance over conventional regenerators. However, we have tested four etched foil regenerators in both Stirlingcycle and pulse tube coolers, and in all cases the conventional screen regenerator performed better than the etched foil regenerator. Finally, as part of the PRDA, we also evaluated linear-arm flexures,8 comparing compressor exported lateral vibration forces between linear flexures and spiral flexures for the same compressor. The linear flexure force levels were typically a factor of two to four lower than the spiral arms for the harmonics. PULSE TUBE DESCRIPTION From among the large variety of configurations for multistaging of pulse tubes, we baselined the multibypass configuration first introduced by Zhou and Han1. The pulse tube, shown schematically in Figure 1, consisted of two regenerator sections and two pulse tube sections. At the cold end of the first-stage regenerator, a capillary bypassed the second-stage regenerator and pulse tube, and entered the junction between the first- and second-stage pulse tubes, forming the first stage of the pulse tube. Since the first stage cooled from 300 K to 60 K, a factor of 5 in temperature, while the second stage only spanned a factor of 1.7, from 60 K down to 35 K, most of the demand was on the first stage. Hence, the first stage regenerator and pulse tube volumes were significantly larger than the second stage, and the mass flow through the bypass capillary was significantly larger than the mass flow entering the second-stage regenerator. The ratio of cooling power between the first and second stage was determined by the ratio of
the mass flow between the second-stage regenerator and the first-stage pulse tube, controlled by means of the diameter of the bypass capillary. The diameter was determined experimentally because of the difficulty in predicting the flow impedance due to turbulent flow, with Reynolds number on the order of and capillary entrance and exit effects. This pulse tube required a large compressor swept volume, so we used a swept volume compressor originally developed under NASA Goddard support for low-cost commercial
Figure 1. Two-stage intermediate bypass pulse tube cryocooler.
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applications.9 This compressor is a flexure bearing, moving magnet design with self-aligning features for rapid assembly, and unlimited lifetime. PULSE TUBE TEST RESULTS The pulse tube was instrumented with pressure sensors near the compression space, at the warm end of the pulse tube, and in the reservoir volume. A Kaman eddy current sensor was used to detect the piston position. Compressor PV measurements were made by digitizing the piston position and pressure waveforms, and integrating the product of the pressure and the differential piston displacement over one cycle. Thermocouples were placed along the outer walls of the regenerators and pulse tubes for observation of temperature profiles. Standard platinum resistance thermometers were mounted at the first stage at both the compressor side and pulse tube side of the bypass capillary, and at the second-stage coldhead. The coldhead itself was wrapped with multilayer insulation and mounted in a vacuum can pumped to about torr. The compressor was mounted outside of the vacuum space, and the compressor and orifice heat rejectors were water jacketed for heat rejection. Figure 2 shows the load map for the final configuration. With a 295 K reject temperature, 85 W of PV power, and 33.5 Hz operation frequency, we achieved 0.32 W at 35 K and 0.6 W at 56 K. At the specified operating temperature of 60 K and 35 K, the pulse tube achieved 1.25 W and 0.18 W of cooling. With no load on either stage the second stage reached 23 K under these operating conditions. If the cold stage temperature is at 40 K and the warm stage at 60 K, 0.48 W and 1 W are attained. In all of these measurements, the temperatures on the regenerator and pulse tube sides of the bypass capillary were the same to within 1 K. Figure 3 shows the cooling capacities as a function of drive frequency. The ratio of the two capacities is dependent on frequency, so in addition to the capillary diameter, operating frequency could also be used to adjust the relative cooling capacities. These data show that the 60 K capacity was a factor of two higher than the requirement, so further reduction of the capillary tube diameter was expected to reduce mass flow bypassing the second stage, decreasing the first-stage cooling power and increasing the second-stage cooling power, and resulting in a coldhead closer to meeting the specifications with a 295 K reject temperature. This was generally the pattern observed during the previous iterations in adjusting the capillary diameter. Unfortunately, further capillary adjustments did not bring the cooling power ratios closer to the requirements. We speculated that dc flows were hindering the fine tuning of the cooling capacity ratios. Recently, several groups have recognized that the double or multiple inlet configurations have a
Figure 2. Load map for two-stage pulse tube cryocooler. Compressor power = 173 W, PV power = 85 W, heat rejection temperature = 295 K.
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Figure 3. Simultaneous cooling powers at 35 K and 55 K, as a function of drive frequency. Compressor power = 173 W, PV power = 85 W, heat rejection temperature = 295 K. Frequency can be used to vary the relative cooling powers of the two stages. These data were measured with a non–optimum configuration and are not identical to those in Figure 2.
closed loop fluid path that allows a dc circulation of mass through the system, and experiments have demonstrated loss of cooling power associated with this phenomenon.5 Work done both at Lockheed Martin3 and at Grenoble4 have also shown that the temperature profiles of the pulse tube and regenerator can indicate the presence of these flows. Figure 4 shows the temperature profiles in the second-stage regenerator and pulse tube, during one particular cooldown, at 23 minutes after turning on the cooler, and at 45 minutes. Because thermocouples were used, the absolute temperatures were not accurate, but changes in temperature at each spatial point are sig-
Figure 4. Evidence of dc flow.
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nificant in these plots. The temperature profile at the later time has a slight shift in the direction of the pulse tube, which indicates dc flow toward the pulse tube. (The lowest temperature points on the pulse tube also show a surprising inversion in the later time data.) Our speculation is that the magnitude of this dc flow changed as we changed capillary diameters, which in turn changed the parasitic load on the second stage. We also observed behavior suggestive of flow instabilities in the second stage. After per-
forming characterizations of an early non-optimum pulse tube, with the first stage at 55 K and the second-stage at 35 K, the heaters on both stages were turned off and the temperature was recorded as a function of time. For an hour, the cold-end temperature underwent dramatic oscillations with
an amplitude of nearly 10 K before finally settling down at a temperature of about 31 K. This is shown in Figure 5. We speculate that this also was due to unstable behavior of the dc flows. ETCHED FOIL REGENERATOR EVALUATION
The standard regenerator matrix such as screen stacks or sphere beds has compactness factors about a factor of five less than the geometries of parallel plates or tubes, suggesting that there is substantial room for improvement in regenerators. Previous attempts at parallel plate regenerators have been unsuccessful, with speculation that the poor performances were due to flow maldistribution. Recently, Yaron6 introduced a regenerator concept approaching the flow characteristics of parallel plates, while providing adequate lateral flows to prevent channelization. This involved state-of-the-art metal film etching techniques to fabricate a highly specialized local geometry, designed with the aid of CFD analysis7. Based on friction factor and Nusselt number correlations from the simulations, analysis indicated a potential for significant improvement in
both pulse tube and Stirling cooler performance. We selected foil geometries intended to be drop-in replacements for the standard screen stack regenerators in our Stirling and pulse tubes, fabricated samples, and performed tests in both Stirling cycle and pulse tube cryocoolers. We tested four etched foil regenerators, and in all cases the conventional regenerator outperformed the etched foil regenerator.
Figure 5. Temperature oscillation suggestive of dc flow or flow instability.
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During the fabrication and assembly process, it became evident that the final geometry was somewhat less than ideal. The etching process did not produce foils of uniform thickness, and the rolling of the foils into a cylinder for insertion into a regenerator tube resulted in non-uniform gaps between foils. The etched foil technology requires, at the very least, further development of fabrication techniques. TANGENTIAL FLEXURE TESTS
Many imaging instruments require extremely low cryocooler exported vibration levels. Counterbalanced pistons controlled with active feedback loops can achieve significant reduction in vibration in the axial direction, parallel to the piston motion. However, reduction in lateral vibration is problematic because there is no controllable mechanism that can be used for vibration cancellation unless secondary actuators are used. Thus, reducing lateral vibrations ideally would be accomplished by refinement of the piston–flexure suspension system, and increasing the lateral stiffness of the flexure springs has been suggested as one method. Wong, Pan, and Johnson8 recently introduced a novel flexure arrangement, replacing the conventional spiral arms with tangential linear arms, as shown in Figure 6. This provided higher lateral stiffness/axial stiffness ratios for similar spring stress levels. To determine whether this higher lateral stiffness would actually reduce lateral exported vibrations, we performed dynamometer tests on both types of springs in our 1710 space-qualified compressor. The spiral springs
used were the original springs, whereas the tangential springs were designed and fabricated by Peckham Engineering. The design of the tangential springs was constrained by the existing geometry of the compressor. Figure 7 shows the results of FEA analysis on the tangential springs vs. the spiral springs. Over the full range of extension, the tangential springs show a much higher ratio of lateral/axial stiffness than the spiral springs. Both of these springs were designed within the infinite fatigue stress limits, with the peak stress in the tangential springs at 290 MPa, and the spiral springs at 311 MPa. Tests were performed in a 3-axis dynamometer developed by Lockheed Martin. The piston motions were closed-loop controlled to a sinusoidal motion with about 1% harmonic distortion. Shown in Figure 8 are data for a typical operating condition of 60 Hz at a 5 mm peak-to-peak stroke. With the exception of the fundamental, all of the higher harmonics in all cases showed reduction in lateral vibrations by a factor of three to four with the tangential springs. The residual vibration at the fundamental frequency can be explained by a slight misalignment of the pistons of 0.1 degrees. During these tests we also observed that the lateral vibrations for the spiral springs
were sensitive to dc offsets in the piston position, while the tangential springs were insensitive to offsets.
Figure 6. Geometries of spiral and tangential flexures.
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Figure 7. Comparison of spiral and tangential flexure stiffnesses.
SUMMARY
We have succeeded in demonstrating the feasibility of a multi-stage pulse tube coldhead, achieving 0.32 W at 35 K, 0.6 W at 56 K with 85 W of PV power and 295 K reject temperature. With a reject temperature of 280 K we met the requirements with 0.4 W at 35 K and 0.6 W at 56 K, with 90 W of PV power. We have observed indications of dc flow and have observed instabilities in the second-stage temperature possibly attributable to these flows. In addition, we have identified several factors which suggest that the multi–inlet pulse tube may not be the most thermodynamically efficient configuration for multistage cooling. Tests on four etched foil regenerators, in both Stirling and pulse tube cryocoolers, showed worse performance in all cases than the conventional screen regenerators. Our conclusion is that the fabrication methods are not sufficiently mature to produce regenerators with better performance than traditional screen regenerators.
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Figure 8. These test data show the reduction in induced vibration when tangential springs are substituted for spiral springs for the lateral axis.
A dynamometer comparison on compressor lateral exported vibration was made between tangential flexures and spiral flexures. With the exception of the fundamental, the tangential flexures reduced the measured vibration levels by a factor of three to four. ACKNOWLEDGEMENTS
This work was supported under contract with Air Force Phillips Laboratory, Contract F29601-95-C-0213.
REFERENCES 1. Zhou, Y., and Han, Y.J., “Pulse Tube Refrigerator Research,” Proc. 7th Int. Cryocooler Conf., Santa Fe, NM(1992). 2. Gedeon, D., DC Gas Flows in Stirling and Pulse Tube Cryocoolers,” Proc. 9th Int. Cryocooler Conf., Waterville Valley, NH (1996). 3. Kotsubo, V, Huang, P., and Nast, T., “Observation of DC Flows in a Double Inlet Pulse Tube,” to be published elsewhere in this conference.
4. Ravex, A., Poncet, J.M., Charles, I., and Bleuze, P., “Development of Low Frequency Operation Pulse Tube Refrigerators,” to be published in Adv. Cryo. Eng. 5. Ju, Y.L., Wang, C., and Zhou, Y., “Dynamic experimental investigation of a multi-bypass pulse tube refrigerator,” Cryogenics 37, pp. 357-361 (1997). 6. Yaron, R., Shokralla, S., Yuan, J., Bradley, P., and Radebaugh, R., “Etched Foil Regenerator,” Adv. in Cryo.Eng.,Vol. 41 (1996). 7. Nigen, J.S., Yaron, R., Karki, K.C., Patankar, S.V. and Radebaugh, R., “3-D Flow Model for Cryo-
cooler Regenerators,” Proc. 9th Int. Cryocooler Conf., Waterville Valley, NH (1996).
8. Wong, T.E., Pan. R.B. and Johnson, A.L., “Novel Linear Flexure Bearing,” Proc. 7th Int. Cryocooler
Conf., Santa Fe, NM (1992). 9. Nast, T., Champagne, P., and Kotsubo, V., Development of a Low Cost Unlimited Life Pulse Tube Cryocooler for Commercial Applications, to be published in Proc. 1997 CEC/ICMC, Portland, Oregon (July 1997).
Development of Pulse Tube Cryocoolers for HTS Satellite Communications V. Kotsubo, J.R. Olson, P. Champagne, B. Williams, B. Clappier, and T. C. Nast Lockheed Martin Advanced Technology Center Palo Alto, CA, USA 94304-1191
ABSTRACT
The Advanced Technology Center of Lockheed Martin is developing high-capacity (10 W @ 77 K) and low-capacity (0.5 W @ 77 K) pulse tube cryocoolers as part of a program to develop reduced weight satellite communications payloads utilizing high-temperature superconducting (HTS) microwave circuits. A consortium of Com Dev, DuPont, and Lockheed Martin is working together with NASA Lewis Research Center to develop this technology under a DARPA program. The cooling performance and system characteristics of both the high-power and low power cryocooler systems are described in this paper. INTRODUCTION
The discovery of the high-temperature superconducting materials in 1986 expanded the potential for wide scale applications for superconductive devices, since cooling could now be provided by single-stage cryocoolers. Planar film technology has evolved to the point where highquality, low-surface-resistance films are readily available, allowing state-of-the-art, highly compact passive RF components. The most mature component is the filter, where a number of companies are exploring potential commercial and military applications where high out-of-band rejection, low noise, or compact size and low weight are system drivers. Since this technology is attractive for satellite communications systems, a consortium consisting of Lockheed Martin, Com Dev, DuPont, and NASA Lewis has formed under a Defense Advanced Research Projects Agency (DARPA) Technology Reinvestment Program (TRP) to develop HTS subsystems for these applications. The goal is to develop completely self-contained subsystems, including the HTS devices, cryopackaging, and the cryocooler and associated drive electronics. Systems trade studies have indicated that these systems will be competitive on size and weight reduction, rather than performance. While the HTS devices themselves are extremely compact in comparison to conventional components, the overall system will be competitive only with an extremely compact, lightweight cooler. In addition, limited power budgets on satellites place stringent requirements on cryocooler power draw. Finally, typical commercial communication satellites now have 15-year-lifetime requirements, placing a high demand on cooler reliability. As part of the consortium, the Lockheed Martin Advanced Technology Center is developing two pulse tube cryocoolers to cool these devices. A low-capacity cooler (0.5 W @ 77 K) will be Cryocoolers 10, edited by R. G. Ross, Jr. Kluwer Academic/Plenum Publishers, 1999
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used for receive-side subsystems, where RF power dissipation is negligible, and the dominant heat load is conducted heat from the coaxial signal cables. A high-capacity cooler (10 W @ 77 K) will be used for transmit side subsystems, where there is significant power dissipation, both in the HTS devices and coaxial signal cables. The compressors use the proven, long-life, flexure-bearing, linear drive architecture. These particular compressors are second generation versions, where a number of low-cost manufacturing features have been implemented. Pulse tubes were selected for the coldheads primarily for high reliability to meet the 15-year-lifetime requirement. In addition, pulse tubes now offer excellent thermodynamic performance surpassing that of Stirling coolers at 77 K as demonstrated in recent work by Lockheed Martin, both in this program and in a separately IRAD funded program1. An engineering model of the high-capacity pulse tube has been developed and tested. This unit uses a dual opposed piston compressor based on a compressor originally developed by Lockheed Martin for low-cost commercial applications under a NASA Aerospace Industry Technology Program (AITP) program2. This cooler produced 9.66 W at 77 K with a total compressor input power of 175 W for a specific power of 18.1 W/W. An engineering model of the low-capacity cooler is presently under development. We are
implementing low-cost manufacturing features similar to that of the compressors used for the 10 W cooler. It will be a dual head-to-head configuration, and the total mass of the cooler will be about 1.25 kg. The cryocooler is concurrently under development under a NASA-GSFC contract. Laboratory coldheads for this cooler have been built and tested, using Lockheed Martin’s spacequalified 1710C compressor3 for testing. One coldhead, designed for the projected operating frequency of 60 Hz, achieved 0.5 W of cooling at 77 K with 10 W of PV power. Since this coldhead was designed for 60 Hz, the compressor was operated off-resonance so the total input power was
relatively high at 28 W. We also designed a second coldhead specifically to match the 1710C compressor, and achieved 0.5 W of cooling at 77 K with 15 W total compressor power. All of these coldheads were designed using methods that were developed in an IRAD supported program for a 2 W at 60 K pulse tube for spaceborne applications. This work is presented in a publication presented elsewhere in this conference1. Requirements Definition
The requirements for the coolers were determined by heat load analysis of the configurations for the HTS subsystems. The subsystems were selected by a study evaluating the applicability
and the maturity of HTS subsystems for several existing and future communications satellites. For the receive side, a 60-channel C-band MUX for IntelSat 8 was determined to have highest potential for near-term deployment.
The package for this system has been designed specifically with the goals of minimizing total package mass and minimizing the cooling load on the cryocooler. The filters and circulators are arranged in a stacked cylindrical configuration. The cold components are mechanically supported by a central structure consisting of three re-entrant fiberglass tubes, designed to maximize the thermal path length from ambient to the cold stage, while having a reasonably compact vol-
ume. The pulse tube coldfinger inserts into the bore of these tubes and is connected to the cold platform with a flexible copper braid link and a shrink-fit interface, which allows decoupling of the cooler from the package. An aluminum shroud surrounds the cold components, and the overall package is wrapped with MLI. Coaxial signal cables pass through both a warm and cold connector plate, and are bundled to pass through the MLI blanketing. The overall package has been structurally designed for survivability during launch. The cooler will consist of a redundant pair of compressors connected to a single pulse tube coldhead. High-reliability, space-qualified pyro valves, one normally open and the other normally closed, will be used to switch from one compressor to the other in case of failure. A single plate at the base of the package is used both for the mechanical interface and as the heat rejection surface to the spacecraft.
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Table 1 summarizes preliminary estimates for the heat load on the cryocooler, with a warm boundary temperature of 300 K. The dominant load, 230 mW, comes from conduction through the coaxial cables, which were commercially available coaxes selected to minimize conduction load, while meeting RF insertion loss requirements. The cable heat load analysis included estimates of
the radiation load on the coax from the environment. The heat load from RF power dissipation was negligible. The fiberglass support structure contributed about 100 mW, while the MIL blanketing contributed about 75 mW. Based on these analysis, 0.5 W cooling power at 77 K was selected for a 20% margin on the package design. In the event that the final package has a slightly higher heat load, the pulse tube can produce 1 W at 77 K with minimal changes in the design. The high cooling capacity requirement for the transmit subsystem was based on heat load evaluations of several different strawman multiplexer configurations. These configurations included different numbers of channels, various cable types, possible use of RF thermal isolators to reduce conducted heat loads, and various assumptions on power dissipation within the filters themselves. These analyses indicated that the heat load on the cooler is dominated by the dissipation within the cold RF components and conductive heat and ohmic power dissipation in the coaxial cables. The heat leaks conducted through the support structure and radiative loads through the MLI blanketing were negligible. The resulting heat loads ranged from 8 to 33 watts. For these cooling loads, assuming a highly efficient 15 W/W cryocooler, the compressor power would range from 120 W to nearly 500 W. This results in a significant equivalent weight penalty for the spacecraft power system and radiators necessary to supply this power and dissipate the rejected heat. Using a power equivalent weight estimate of the burden on the satellite system alone for 500 W of compressor power is 91 kg, which essentially prevents the larger dissipation HTS system from being competitive with conventional technology. Currently, more studies are underway to further define the transmit side subsystem and to determine feasibility in light of the significant weight penalty. Despite the lack of a finalized subsystem, to proceed with cooler development, a cooling goal was set at 10 W at 77 K, based on the assumption that a feasible transmit subsystem will likely be one with lower cooling requirements.
High-Capacity Pulse Tube An engineering model of the nominal 10 W at 77 K pulse tube cryocooler has been developed and tested for thermodynamic performance. The dual opposed flexure-bearing compressor was based on a design originally developed under a NASA AITP from NASA GFC for low-cost commercial cryocoolers. This compressor incorporates a number of unique design features to reduce manufacturing costs, yet enhance the long-life characteristics of the flexure bearing approach. The details of the compressor have already been presented at the 1997 CEC in Portland, OR.2 This motor uses a moving magnet, which eliminates flexing leads required for a moving coil motor, and allows the coil windings to be placed outside of the pressure vessel, removing coil potting as a potential contamination source. The external coil also eliminates the penetrations
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through the pressure wall for the electrical feedthroughs. The external coil reduces cost and increases reliability. The piston/flexure suspension system self-aligns during assembly, providing on-axis motion of the piston. The critical piston-cylinder alignment uses a unique alignment process which eliminates the tedious alignment procedure of earlier designs. This alignment technique uses load cells, optical position sensors, and micropositioners which allow the entire alignment procedure to be computer automated. This innovative compressor design significantly reduced both the piecepart count and assembly time over the previous design. The coldhead was an in-line configuration, operating at a frequency of 33 Hz. For test purposes, chilled circulating water at 295 K was used to extract heat at both the warm end of the regenerator and at the warm end of the pulse tube. The coldtip only was inserted into the vacuum vessel, and was wrapped with several layers of MLI. Figure 1 shows load lines for input powers of 100, 140, and 175 W, where the lower power levels correspond to lower compressor strokes.
The 175 W load line was near the maximum stroke capability of the unit. This cooler produced 9.66 W at the design point of 77 K, and reached a minimum temperature of about 35 K. Figure 2 shows the total compressor specific power. At 77 K and 175 W input power, the specific power is 18.1 W/W, and is slightly higher at the lower power levels. The motor efficiency was about 67% and depended on operating conditions. Because many applications require vibration isolation between the coldtip and compressor, we also tested a 100-cm-long transfer line for this cooler. Test results showed about a 0.5 W loss in cooling power in comparison to the original 30-cm-long transfer line.
Figure 3 shows the dual compressor pair. Total mass for this system excluding electronics is 13.01 kg, with the compressor weight 12.4 Hz. The envelope for the compressor is a right circular cylinder 11.5 cm diameter and 45.8 cm long.
Low-Capacity Pulse Tube
The 0.5 W space-qualifiable pulse tube engineering model under development will be driven by a miniaturized version of the low-cost compressor, utilizing the same low-cost manufacturing concepts. A dual opposed piston configuration will be used for vibration cancellation. The unit will have an envelope of 6.6 cm diameter by 17.6 cm long.
Figure 1. Leadlines for the high-capacity cooler at compressor input power levels of 100, 140, and 175 W.
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Figure 2. Total compressor specific power for the high-capacity cooler.
Two in-line versions of the coldhead have been designed and tested, using a single piston version of our 1710C compressor3 for test purposes only. One coldhead was designed for 60 Hz operation corresponding to the resonant frequency of the compressor under development. The loadlines for four different PV powers for this coldhead are shown in Figure 4, taken with a 295 K heat rejection temperature. PV power is plotted instead of total power, because the 60 Hz frequency was off resonance for the 1710C compressor. At the design point of 77 K, 0.5 W of cooling power could be achieved with about 10 W of PV power, for a PV specific power of 20W/W. The PV specific power is shown in Figure 5.
A second coldhead specifically matched to the 1710C compressor to demonstrate overall power efficiency was also built and tested. This cooler ran at 45 Hz to operate the compressor on
resonance. The load lines for this pulse tube for three different total power levels are shown in Figure 6. At 77 K. over 0.5 W of cooling was produced with 15 W of total power, for a total specific power of 30 W/W. The total specific power for this cooler is shown in Figure 7.
Figure 3. The dual opposed piston compressor for the high-capacity cooler.
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Figure 4. Load lines for the 60 Hz coldhead for the small-capacity cooler.
We are developing a U-tube version of this coldhead for the final configuration, for ease of integration into the HTS system. The system layout is shown in Figure 8. Our analysis indicates
that the U-tube turning duct results in some loss in performance. On the other hand, we have modified the design of the 60 Hz coldhead based on the results of these tests, and anticipate that these
Figure 5. PV specific power for the 60 Hz coldhead for the small-capacity cooler.
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Figure 6. Load lines for the 45 Hz coldhead for the small-capacity cooler.
design improvements will compensate for the loss in performance due to the U-tube so that the overall performance will be comparable. The U-tube coldhead will be under test in July, 1998 with tests of the compressor and coldhead planned for late 1998. The compressor and pulse tube are being developed under joint funding by DARPA and NASA GSFC.
Figure 7. Total compressor specific power for the 45 Hz coldhead for the small-capacity cooler.
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Figure 8. Schematic of the U-tube configuration of the small-capacity cooler currently under development.
A new set of control electronics is being developed for this cooler. The new electronics will be smaller, lighter, more efficient, and use less overhead power than Lockheed Martin’s controller for Stirling cryocoolers, due to elimination of displacer drive and control functions, and because of the small compressor motor. The control algorithm will be simplified by controlling only the piston amplitude and offset, rather than the complete waveform control as with the existing controller, although optional modules for closed loop temperature control and active vibration cancellation can be included. The logic circuits will be reduced to fit in a gate array, eliminating the need for a microprocessor. PWM amplifiers with high conversion efficiency will provide current to the motors.
SUMMARY Under a DARPA TRP program, a consortium of Lockheed Martin, Com Dev, and DuPont is jointly developing HTS subsystems for satellite communications applications. The cryocoolers are required to have high efficiency, low mass, and a 15-year life. Trade studies on potential subsystems have set the cooling requirements at about 0.5 W at 77 K for a receive side system, and 10 W at 77 K for a transmit side system. Lockheed Martin has developed an engineering model of the high-capacity cooler, demonstrating 9.66 W of cooling at 77 K with 175 W of total compressor power. The coldhead was driven by a dual opposed piston compressor, originally developed under a NASA program for low-cost cryocoolers. A low-capacity, 0.5 W of cooling at 77 K engineering model is currently under development. The compressor will be a dual opposed unit using the same architecture as the low-cost cooler
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compressor, and the total cooler mass will be 1.25 kg. With 15 W compressor power, the unit is predicted to produce over 0.5 W of cooling power at 77 K. Laboratory tests have been performed on the pulse tube coldheads. A 60 Hz version, designed to match the compressor under development, has demonstrated 0.5 W at 77 K with about 10 W of PV power. A 45 Hz coldhead specifically designed to match the laboratory compressor has demonstrated 0.5 W of cooling with 15 W of total compressor power. ACKNOWLEDGMENTS
This work was supported by DARPA Cooperative Agreement No. NCC3-517 and Lockheed Martin Internal Research and Development funding. The development of the low-capacity cooler is also supported by NASA-GSFC under contract NAS5-97218. REFERENCES 1. Kotsubo, V., Olson, J, R., and Nast, T. C., “Development of a 2 W @ 60 K Pulse Tube Cryocooler for Spacebome Operation,” these proceedings. 2. Nast, T., Champagne, P, and Kotsubo, V, “Development of a Low-cost Unlimited-life Pulse Tube Cryocooler for Commercial Applications,” to be published in Adv. Cryo. Engin. 3. Nast, T. C., Champagne, P.J., Isaac, D., Pryor, G. M., von Savoye, R. L. and Naes, L. G., “Design, Performance and Testing of the Lockheed-Developed Mechanical Cryocooler,” Proc. 8th International Cryocooler Conference, Vail, CO (1994). 4. Lockheed Martin Internal Communications, 7/23/97.
A Pulse Tube Cryocooler for Telecommunications Applications
J.L. Martin†, J.A. Corey‡, C.M. Martin† †
Mesoscopic Devices, LLC, Golden, Colorado, USA 80401 CFIC, Troy, New York, USA 12180
‡
ABSTRACT
Superconducting Core Technologies (SCT) is developing a pulse tube cryocooler designed for cooling telecommunications equipment in outdoor installations. The cooler is based on a moving-magnet, flexure-bearing-supported linear compressor that is both inexpensive and scalable to a variety of power levels. The cryocooler is designed to support HTS superconducting filters and cryocooled low noise amplifiers in a commercial wireless telephone range extension product. The key components of the SCT cryocooler are the linear compressor, an efficient single-stage, in-line pulse tube cold head, a high-reliability electronic drive and an air-cooled heat rejection system.
This paper presents preliminary measurements of cooling performance, describes system weight and size envelope, and describes the trade-offs between system performance and production costs. Performance measurements for the 300 W, passively balanced compressor are given, along with predicted production costs in low and moderate volumes. INTRODUCTION
Superconducting Core Technologies (SCT) is a manufacturer of cryogenically cooled frontend receivers for cellular telephone base stations. These receivers utilize high-temperature superconductor thin-film circuits and cooled low noise amplifiers (LNA’s) to provide a combination of excellent filtering and low-noise pre-amplification which extends the range and
capacity of a base station. All of SCT’s receivers are designed for towertop mounting. SCT’s first-generation systems utilized Gifford-McMahon cryocoolers with ground-mounted compressors and helium lines leading to the coldhead inside the towertop receiver. For a secondgeneration system, SCT wanted to replace this split arrangement (compressor on ground, cold-
head on tower) with a integrated, tower-mounted system. In January of 1997, SCT began developing a pulse tube cryocooler specifically for this telecommunications application. This paper describes the development and early testing of this system.
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SYSTEM DESCRIPTION SCT’s front end receivers include between two and six radio-frequency (RF) channels, each with a superconducting filter and cryogenically cooled low-noise amplifier. The microwave components are contained in an evacuated cryostat, and RF signals pass into and out of this cryostat along small-diameter coaxial cables. Outside the cryostat, room-temperature electronics provide power, status monitoring, system control, lightning protection and RF bypass functions. The cryostat and room-temperature electronics, along with the cold head and compressor, are contained in an outer aluminum can which provides weather shielding.
Performance goals
The target application for the pulse tube cryocooler developed at SCT was a “fully-sectored” cellular-band receiver with six RF channels, utilizing thallium-based superconducting filters and low-gain (13 dB) LNA’s. In this configuration, approximately three-quarters of the heat load on the cryocooler is due to conduction along the RF cables, and one-quarter is due to the power dissipation in the LNA’s. Total heat load on the system when operating in a 20°C ambient is approximately 4.3 W. This rises to 4.8 W at the system’s maximum design temperature of 60°C (140°F). Design cold tip temperature was 70 to 80 K, depending on the exact configuration of filters and amplifiers. The design of the cryocooler was driven primarily by challenging cost and reliability targets. Due to the high cost of servicing tower-mounted equipment, a maintenance cycle of at least five years was desired to minimize life-cycle costs for the receiver. This >40,000 hr continuous duty life significantly exceeds that demonstrated by tactical cryocoolers, and limits the design options available. The production cost goal of <$3,000 in quantities of 200/month for the complete cryocooler meant that this cooler would need to achieve a lower cost per watt than any other long-life cryocooler yet demonstrated. Table 1 summarizes the requirements for the system. To meet these requirements, we selected an in-line pulse tube design using a moving magnet compressor with flexure bearings. SCT contracted with Clever Fellows Innovation Consortium (CFIC) to design the compressor for the system. The following sections describe the key components of the cryocooler in more detail.
Compressor The compressor is based on CFIC’s patented2,3,4 STAR™ resonant reciprocating motor. The assembly consists of the coil-wound stator; integrated piston and permanent magnet plunger; cylinder/motor support; and passive dynamic balancer. These are enclosed in a common pressure vessel with a single-plane interface to the pulse tube. A section drawing is given in Figure 1. A photograph of the internal assembly is shown in Figure 2. This unit has a motor diameter of 162 mm and a total prototype vessel dimension of 175 mm diameter by 220 mm long. The mass is 10 kg without vessel, 15 kg with the prototype cover. Production vessels are intended to be welded shut, not bolted, saving 2 kg.
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Figure 1. Compressor cross-section.
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Figure 2. Compressor internal assembly.
The motor is an eight-pole design with stacked, transverse laminations (much like a conventional rotary motor stator). The plunger carries eight rectangular, dual-polarity permanent magnets of iron-neodymium at 1.25 Tesla remnance that reciprocate between pairs of stator poles (Figure 3), producing extremely low sidepull forces and external flux spillage. To enable this economical and effective construction, the plunger is supported on unusual planar flexure suspension elements which completely eliminate torsional motion (and vibration) typical of conventional spiral planar springs. Only two flexures are required for the motor (one fore, one aft). Each flexure is thin enough (0.3 mm) to be economically stamped from valve steel sheet. These flexures have been fatigue tested to more than cycles at up to 30% over maximum stroke and re-proven in compressor testing. The motor is designed for 270 watt mechanical
Figure 3. STAR motor plane section.
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output from 300 watt input at 60 Hz and 14 mm stroke. It is wound for a maximum 14 volts rms input to take advantage of a 24 volt DC supply. The piston moves without rubbing contact in a cylinder to produce the pressure wave that
drives the pulse tube. The clearance seal is maintained in a centered radial position by the flexures. The radial clearance is 15-25 µm. This robust and inexpensive design prevents wear and limits seal loss to approximately 6% of delivered PV power at rated conditions (14 mm
stroke for
swept volume, 60 Hz, with 200 kPa amplitude 45° ahead of the volume
delivery wave.)
The piston/cylinder is much smaller in diameter than the STAR motor. This provides room for a compact passive balancer in the annular space around the cylinder. This balancer uses six coil springs in parallel and two more planar flexures that prevent all but the desired mode of
motion. The mass-spring system is tuned to match the operating frequency of the motor, but partial cancellation of 80 to 90% was chosen to meet the residual acceleration limit of 1g (amplitude) while broadening the band of useful balanced frequencies to approximately 3 Hz.
This is achieved by shaping and fitting the moving mass to enhance its windage drag.
Cold head The SCT pulse tube design utilizes an in-line pulse tube with an inertance tube to provide the phase shift. Figure 4 shows the cold head layout for the water-cooled laboratory version. The production version is similar, but uses an air-cooled aftercooler and a slightly different vacuum flange. The cold head uses stacked screens for the regenerator, heat exchangers and aftercooler. Key parameters for the pulse tube cold head are summarized in Table 2.
Control and power electronics The electronics for the cryocooler consist of two primary subsystems: a power inverter and a
closed-loop control system. The inverter converts DC power to a sinusoidal waveform to drive the compressor. Using an inverter, rather than driving the compressor at the power grid frequency has several advantages. By using DC power, the system can more easily be supported by the base station’s battery backup system. Since loss of AC power is the single most important cause of cryocooler non-availability, using the station’s battery backup systems significantly increases the overall system reliability. Second, using an inverter allows a single system to be used worldwide, independent of the frequency of the local power grid. Finally, a low-voltage DC system is more easily licensed and installed, as some countries limit the voltage that can be used on power cables running up the tower.
Figure 4. Pulse tube cold head.
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Figure 5. Prototype inverter.
Early specifications for the receiver required the system to operate on either US standard 24 V nominal power supplies, or on European standard 48 V systems. This requirement lead to the choice of the 14 V rms design voltage for the motor. This relatively low voltage allows operation of the cryocooler off of 24 V nominal supplies, however, it does require relatively large
capacitors to minimize the reflected current ripple. The requirement for 24 V operation was later dropped from the specifications, and second-generation systems will be able to take advantage of the narrower voltage range.
Figure 5 is a photograph of the prototype inverter board. The inverter was designed to SCT specifications by Electronic Power Conversion, of Vestal, NY. The board is designed to mount directly to the back of the compressor, and incorporates both an aluminum heat sink and standoffs for the cooling fan. The inverter uses a pulse-width-modulated (PWM) architecture with a 22 kHz switching frequency. In addition to normal operation under closed-loop control, the inverter also incorporates several features intended to protect the compressor during offnormal conditions. The inverter monitors compressor stroke and will reduce output power when an overstroke condition is detected. It also incorporates a soft-start feature and low-voltage shutdown circuitry. The inverter is an all-analog design. Control signals to the inverter are carried across an optically-isolated input port. The port takes a 100 Hz TTL square wave as an input and produces a voltage proportional to the input signal duty cycle. Optical isolation of the inverter simplifies the development of lightning protection for the system by allowing the cryocooler to float with respect to the RF circuitry ground. The inverter is designed to operate in outdoor conditions, including salt fog and rain. Production boards will be conformally coated and will use sealed connectors for connections to the control electronics and the compressor. The control electronics are built into SCT’s REMOTE™ control, diagnostic and health monitoring subsystem. The REMOTE board includes a microprocessor which controls the bypass switches, monitors the performance of the low-noise amplifiers, and controls the cryocooler to maintain the desired temperature. The control board includes a constant-current
source to drive a silicon diode temperature sensor attached to the filter packages. The board senses the voltage from the diode and uses a lookup table to determine the temperature. The accuracy of the system is approximately while the resolution is approximately 0.1 K.
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The sensed temperature is compared to the set-point temperature and used in a closed-loop control system to maintain the filters at their optimal operating temperature by adjusting the inverter output voltage and hence the compressor stroke. On a similar cryoccoler, this control system was shown to maintain the filter temperature to within of the setpoint temperature over a 100°C variation in ambient temperature. PRELIMINARY PERFORMANCE MEASUREMENTS
Compressor. The compressor was first tested with a dummy load consisting of an enclosed test volume connected through a variable orifice to the back vessel space. In this configuration, the motor was tested to over 400 watts input and 322 maximum output. Compressor efficiency was not measured separately, but with a pulse tube in place, the ratio of PV power to input AC power was found to peak at 76%. PV power was derived from a reflective optical position sensor on the piston and a dynamic pressure sensor in the working space. Maximum measured combined efficiency of the compressor (motor + piston/cylinder) is 85% at 98 watts PV. Total seal losses were estimated at <4.7% of PV power at design conditions, based on measurements of seal leakage under DC flow conditions.
Inverter efficiency. We measured the efficiency of the inverter while it was powering a dummy load with the same resistance and inductance as the actual compressor. The inverter
efficiency was measured at 94% when operating at 90% of rated load (270 W). The efficiency remains above 85% down to 30% of rated power. Load curve. Figure 6 shows the load curve for the cryocooler. Since no multi-layer insulation was used in these tests, the figure shows both the measured load curve and the predicted load curve with the estimated 0.75 W of radiation heat transfer subtracted. These tests were conducted on a water-cooled version of the cryocooler. Cooling water temperature for these tests was maintained at 20°C. The load curve shown in Figure 6 was taken with the compressor charged to 2.0 MPa mean pressure, operating at 57.3 Hz and 91% of rated stroke (6.4 mm amplitude). Compressor mechanical power at this condition was estimated by simultaneously
measuring the piston position and the pressure in the compression space and integrating over several cycles. The mechanical power during the tests of Figure 6 was approximately 180 W. We estimate the seal losses in the compressor to be between 3 and 6% of the mechanical power, so the fluid power delivered to the cold head is approximately 172 W. With the radiation load subtracted, the available cooling power is 5 W at 70 K, 7.75 W at 80 K, and 10.75 W at 90 K. No-load temperature is approximately 54 K. These tests were conducted on a preliminary version of the cryocooler; we expect somewhat better performance in production models due to a cleaner internal geometry and optimization of the pulse tube parameters.
Figure 6. Load curve.
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SYSTEM INTEGRATION Figure 7 shows a photograph of the towertop receiver into which the cryocooler was incorporated. The receiver has a larger-diameter, hermetic upper section that houses the RF bypass switches, the cryostat with the filters and amplifiers, the lightning protection hardware, and the control and health monitoring circuitry. The twelve RF connections (six inputs and six outputs for a three-sector base station with two antennas per sector) are located in an annular ring at the bottom of this upper section. The compressor, inverter, cooling fan and aftercooler are located in the smaller-diameter lower section of the unit. This lower section is not hermetic, but has a continuous air flow through it provided by a variable-speed DC fan. Airflow from the fan is directed past the inverter, then upward through an annular space around the compressor. It then turns radially inward, passes through the aftercooler, and exits radially just below the upper section. This “mushroom” configuration minimizes the RF path length from the antennas to the filters, provides room for the twelve large RF connectors, simplifies the cooling of the power
electronics and minimizes the visual cross-section of the system. Figure 8 shows a cross-section of the receiver, showing the location of the components inside the outer weather shield. MANUFACTURABILITY
The cost target for 200 per month (lots of 1,000) is $2,600 for a complete compressor, pulse tube and electronics. This was met and bettered with this design. Manufacturability is a key to the design and all parts were developed with specific production processes identified.
Figure 7. Towertop receiver.
Figure 8. Towertop receiver cross-section.
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Cost estimates for the design were developed from complete engineering drawings and vendor quotations with at least two corroborating responses for all but minor parts. Additionally,
cost estimates were developed for single units, lots of 100 assemblies and for 10,000 per year (for alternate consumer applications). Table 2 gives our expected sell prices for complete
cryocoolers at these production levels. Processes appropriate to each production rate were established and costed with supplier input. At low rates, all parts are machined, including vessel and flexures. All assembly is manual,
with minimal fixturing. At the design rate of 1,000 pieces, vessels are spun and flexures are stamped. Fixtures are limited to stator and plunger assembly. At 10,000 pieces, many aluminum parts are cast and coil and stator production are automated. Fixtures are limited to stator and
plunger assembly. At 10,000 pieces per year, many aluminum parts are die cast and stator production is automated. At low rates, the main costs are in machined parts and assembly, although magnet cost is
also high. At design rates, magnets and machined parts remain the major cost elements. In highrate production, magnet cost (at commodity rates) remains the most costly part of this compressor. By careful design for assembly and manufacture at the target production rate, this compressor can be produced for less cost than any other in this capacity range of which we are aware.
SUMMARY SCT is developing a pulse tube cryocooler specifically for the towertop telecommunications
application. The cryocooler has the potential for very high reliability and low cost. With an expected fabrication cost of <$600/W at the 5W/70K/333K condition and a design lifetime in excess of five years, this cryocooler should achieve a price-to-performance ratio better than any other system with comparable cooling power and lifetime. The system was designed from the beginning for an outdoor, air-cooled application, and can operate in a wide range of industrial
environments. By trading some size, weight and performance for manufacturing simplicity, the SCT/CFIC design allows high reliability and good performance without high prices. Preliminary
results indicate that the system will be able to meet aggressive price/performance goals, although further work is required to optimize the performance of the pulse tube and to characterize the system performance at high and low ambient temperatures.
ACKNOWLEDGEMENTS This work was funded by SCT, with support from NIST and Los Alamos National Lab.
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REFERENCES 1.
Gardner, D.L. and Swift, G.W., “Use of Inertance in Orifice Pulse Tube Refrigerators”, Cryogenics, Volume 37, (1997), pp. 117-121
2.
Yarr, G.A. and Corey, J.A., “Linear Electrodynamic Machine”, U.S. patent 5,389,844, Feb. 1995
3. Corey, J.A., “Linear Suspension Device”, U.S. patent 5,139,242, Aug. 1992 4.
Corey, J.A., “Linear Reciprocating Alternator”, U.S. patent 5,146,123, Sep. 1992
Design and Preliminary Testing of BEI’s CryoPulse 1000, the Commercial One Watt Pulse Tube Cooler S.W.K. Yuan, D.T. Kuo, and A.S. Loc Cryocooler Group, Edcliff Division
BEI Technologies Sylmar, California 91342
ABSTRACT
This paper describes BEI’s efforts in producing a one watt pulse tube cooler, the CryoPulse 1000, for commercial applications. The cooler was designed by a computer model which has been validated against various Stirling and pulse tube coolers in the literature.1-8 The cooler was designed, fabricated and tested. Preliminary testing showed that the CryoPulse 1000 produces good cooling. However, the heat exchanger at the hot end of the regenerator needs to be modified to more efficiently reject the heat energy generated. Detailed test results will be presented in a future paper. INTRODUCTION
BEI is a manufacturer of Stirling-cycle cryocoolers based on the concept of clearance seal, pneumatically driven displacer and linear drive motors. Several cooler models are available covering refrigeration requirements ranging from 150 mW to 5.0 Watts of cooling at 78 K. In
pursuit of a long-life cooler, BEI has started an in-house program to design and manufacture a cooler based on the concept of pulse tube and flexure bearings. CryoPulse 1000 is the first step toward this goal, by replacing the pneumatically driven displacer of our B1000 Stirling Cooler9 with a Pulse Tube. The B1000 Stirling Cooler satisfies the U.S. Army’s specification for the SADA II cooler; the specification requires one watt of cooling at 77K with a maximum input power of 40W (ambient temperature of 23 degree Celsius). At an elevated ambient temperature of 71 degree Celsius, the cooler should provide 0.55W of refrigeration capacity with a maximum input power of 60W. The actual performance of BEI’s B1000 Stirling cooler far exceeds that of the SADA II requirement. Thus, although pulse tubes are generally less efficient than their Stirling cousins, the BEI CryoPulse 1000 cooler is still expected to produce one watt of cooling at 78K (ambient case temperature). COMPUTER ANALYSIS
At BEI, a pulse tube computer model has been developed which is very similar to the Pulse Tube Performance Model (PTRM) that was validated against two different pulse tube coolers Cryocoolers 10, edited by R.G. Ross, Jr.
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(References 3 and 4), including a blind test. This model was modified from the Stirling Refrigerator Performance Model (SRPM), which is a third order model that has been validated extensively against various Stirling coolers in the literature. They include the Lucas-Lockheed 60K unit,7 the NASA/Philips Magnetic Bearing unit,6 the Oxford refrigerators,5 and the Atmospheric Infrared Sounder (AIRS) units A, B, and C. A detailed description of the model can be found in Reference 8. The BEI Pulse Tube Model was used to design the CryoPulse 1000 Cooler. The equations and assumptions used in the PTRM model were discussed elsewhere.8 The model breaks up the pulse tube cooler into a number of nodes. The number of nodes in each
section depends on the value of the state variables. For example, more nodes are required in the regenerator because of the large temperature difference and large pressure drop in the axial direction. Conservation of energy, momentum and mass are solved until the solutions converge. Equations of state and empirical equations for pressure drop and heat transfer are also used. No fudge factors are used in the program. The PTRM model was modified from the Stirling Refrigerator Performance Model. The expansion space of the SRPM was replaced by the pulse tube with an orifice (or inertance tube) and the surge volume. The volumetric variation and the flow passage (to the displacer motor) at the hot end of the regenerator (of the Stirling model) were also eliminated. The gas transport in the pulse tube is modeled as unidirectional laminar or turbulent flow, depending on the Reynolds number. Heat transfer in the axial direction is modeled as enthalpy flow whereas the radial heat transport is predicted by the forced flow heat transfer coefficient (for both laminar and turbulent regimes). The transport across the orifice is modeled by the discharge flow coefficient. Figure 1 compares the experimental and predicted performance of pulse tube coolers in the literature. It includes a commercial 30-watt unit built by NIST (Reference 3), which the PTRM model was validated against without prior knowledge of the performance. All the experimental data fall below the first order analysis, i.e., net cooling is proportional to the cold tip temperature (Tc), the swept volume (Vc), the mean pressure (Pm), compression ratio (Ph/Pl), and frequency (f). This indicates the optimism of the first order analysis, not being able to incorporate all losses (for example, losses associated with dead volume, streaming effect, and boundary layer effect). Also included in the plot are the predicted performance based on the third-order BEI Pulse Tube Model. They include the blind test of the NIST unit and the current predicted performance of the One Watt CryoPulse cooler discussed in this paper.
Figure 1. Comparison of model prediction with experimental data.
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Figure 2. Input power vs. Refrigeration Capacity at 80K.
Refrigeration capacity versus input power (at 80K) is presented in Figure 2. For an input power of 55 watts, CryoPulse provides 1 watt of refrigeration. Refrigeration capacity as a function of coldtip temperature is shown in Figure 3 for a constant input power of 60W. CryoPulse
1000 produces over 5W of cooling at 200K and can cool down to 30K with no load. The effect of orifice size on the cooling capacity is shown in Figure 4 for two operating frequencies. Based on past experience, the computer model tends to over-predict the orifice size and a smaller diameter was used in the experiment. The performance of this cooler can be further enhanced by using an etched-foil regenerator together with by-pass(es) between the regenerator and the pulse tube and/or a double-inlet design.
Figure 3. Refrigeration capacity versus coldtip temperature.
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Figure 4. Effect of orifice size on refrigeration capacity. THE DESIGN
A standard one-watt compressor for the Stirling refrigerator is used to drive the pulse tube. The pulse tube was designed to be a retrofit of the displacer of BEI’s one watt Stirling cooler (B1000). In order to fit inside the SADA II coldfmger dewar, a concentric design was employed with the regenerator wrapping around the central pulse tube as shown in Figure 5. Helium gas is transported via the transfer line, which passes through a gap regenerator and a conventional
regenerator before entering the pulse tube. At the hot end of the pulse tube, an inertance tube is used to tune the phasing for maximum refrigeration. The old gas-spring volume of the Stirling expander is used for the surge volume. RESULTS AND DISCUSSION
CryoPulse 1000 was fabricated and preliminary tests were performed. The pulse tube was found to have a lot of advantages over the pneumatically driven displacer, with less parts, less tight tolerances, and less contamination issues. Figure 6 shows a picture of CryoPulse in operation with a frost-ball building up at the coldtip of the cooler. Based on our experience, the size of
Figure 5. A schematic drawing of the pulse tube.
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Figure 6. The BEI CryoPulse 1000 Cooler.
the frost-ball tells us the pulse tube is producing good cooling. However, the gap regenerator (see Figure 5) gets extremely hot during testing and has to be cooled by convection with a fan. In order to accurately monitor the performance of CryoPulse 1000, a temperature sensor and heater have to be mounted at the coldtip and the entire coldfinger inserted into a vacuum jacket. Further tests will be performed and results will be reported elsewhere. CONCLUSIONS
BEI has been successful in the design and fabrication of our prototype commercial pulse tube cooler, the CryoPulse 1000. Detailed testing will be performed and results will be reported in a later paper. The gap regenerator of the pulse tube was found to get extremely hot during operation. A redesign of this heat exchanger is called for to enhance the performance of CryoPulse 1000. It is also proposed to make the coldfinger out of titanium to reduce the heat conduction from the gap regenerator down the coldfinger. REFERENCES 1.
D. Kuo, T. Loc, and S.W.K. Yuan, Experimental and Predicted Performance of the BEI Mini-linear Cooler, in Proc. the 9th International Cryocooler Conference, (1997) p. 119.
2.
D.T. Kuo, A.S. Loc, and S.W.K. Yuan, Enhanced Performance of the BEI 0.5 Watt Mini-Linear Stirling Cooler, to be published in the proc. Of Advances in Cryogenic Engineering, 1997.
3.
S.W.K. Yuan and R. Radebaugh, A Blind Test on the Pulse Tube Refrigerator Model, in “Advances
4
S.W.K. Yuan, Validation of the Pulse Tube Refrigerator Model Against a Lockheed Built Pulse Tube Cooler, in “Cryogenics”, Vol. 36, No.10, (1996) p.871, S.W.K. Yuan, I.E. Spradley, and W.G. Foster, Validation of the Stirling Refrigerator Performance Model Against the Oxford Refrigerator, in: “Advances in Cryogenic Engineering”, Vol. 39, (1994)
in Cryogenic Engineering”, Vol. 41, (1996) p.1383.
5
p. 1359.
6.
S.W.K. Yuan and I.E. Spradley, Validation of the Stirling Refrigerator Performance Model Against the Philips/NASA Magnetic Bearing Refrigerator, in: “Proc. 7th Int. Cryocooler Conf.”, Vol. 1, Phillips Lab, USA (1993) p.280.
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7.
S.W.K. Yuan, I.E. Spradley, P.M. Yang, and T.C. Nast, Computer Simulation Model for Lucas Stirling Refrigerators, in: “Cryogenics”, Vol. 32, (1992) p. 143.
8.
S.W.K. Yuan and I.E. Spradley, A Third Order Computer Model for Stirling Refrigerators, in: “Advances in Cryogenic Engineering”, Vol. 37 (1992) p. 1055.
9
S.W.K. Yuan, D.T.Kuo, A.S.Loc, and T.D. Lody, Experimental Performance of the BEI One Watt Linear (OWL) Stirling Cooler, to be published in the proc. Of the Advances in Cryogenic Engineering, 1997.
Optimal Design of a Compact Coaxial Miniature Pulse Tube Cooler Y. L. Ju, Y. Zhou, J. T. Liang, and W. X. Zhu Cryogenic Laboratory, Chinese Academy of Sciences P. O. Box 2711, Beijing 100080, China
ABSTRACT This paper presents an optimal design of a 1W / 80K co-axial miniature pulse tube cooler based on the aid of numerical model that incorporates both thermodynamics and hydrodynamics for the oscillating flow and the regenerator matrix. The simulation allowed clear visualization of the instantaneous parameters variations and the entire process occurring in the pulse tube cooler and gave quantitative information on the dominant influence in the cooling capacity of the coaxial miniature pulse tube cooler. The cooler should produce 1W of cooling power at 80K with a total input power of less than 30W and with a heat sink temperature of 300K. Based on the numerical analysis, a compact co-axial miniature pulse tube cooler with a rotary drive 1.66cc compressor was designed and studied. The flow channels of the pulse tube and regenerator were designed to fit the cold heat exchanger. A lowest temperature of 66.5K and cooling capacity of 0.6W at 80K with a total input power of 66W was achieved with the system pressure of 4.0MPa and the operating frequency of around 30Hz. This design was very compact and could be made into a commercially available pulse tube cooler. INTRODUCTION Owing to no moving component in the low temperature region, the performance and lifetime of the miniature pulse tube coolers for space applications have been drastically improved in recent years[1-3]. There are three types for the configuration of pulse tube cooler, which are U type, line type and co-axial (concentric) pulse tube cooler. Obviously, the co-axial pulse tube cooler is very compact and convenient for practical applications. Chan et al.[2] in TRW developed a linear type pulse tube miniature cooler that achieved a cooling power of 0.53W at 80K for input power of 17.8W to a 1cc swept volume linear drive compressor. Recently, they improved the cold head and demonstrated that the new cold head has 3.7 times the cooling power at 65K of the previous design[3]. Kanao et al.[4] presented a U-type miniature pulse tube cooler that achieved a lowest temperature of 98K. The lowest temperature of 68.5K was obtained by Wang et al[5] with a miniature co-axial pulse tube cooler. The co-axial pulse tube cooler was very compact for practical applications. However, its efficiency was still less than that of the Stirling cooler. Our approaches in this paper to improve the efficiency of the co-axial miniature pulse tube
Cryocoolers 10, edited by R. G. Ross, Jr. Kluwer Academic/Plenum Publishers, 1999
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cooler were to determine the operating conditions and the dimensions and geometry of each component to achieve optimal cooling capacity at 80K with minimal input power from the compressor by the aid of numerical analysis. The simulation allowed clear visualization of the instantaneous parameters variations and the entire process occurring in the pulse tube cooler and gave quantitative information on the dominant influence in the cooling capacity of the co-axial
miniature pulse tube cooler. Based on the numerical analysis, a practical co-axial miniature pulse tube cooler with a rotary drive compressor was designed and studied in this paper. A lowest temperature of 66.5K and cooling capacity of 0.6W at 80K was achieved with system pressure of 4.0MPa and operating frequency of around 30Hz. NUMERICAL MODEL
The internal working processes in pulse tube coolers are very complex owning to the unsteady, oscillating compressible gas flow, the porous media in the regenerator and the addition
of the orifice, double-inlet and multi-bypass valves. Therefore, we use one-dimensional model for the numerical simulation to reduce the CPU time and avoid any difficulty involved with the two- or three-dimensional models, and the porous medium approach in regenerator, in which the regenerator is treated as a uniformly distributed flow resistance unit. A detailed description of the numerical model has been given in references 6 and 7. The operating parameters for the miniature pulse tube cooler are: the wall temperatures of the hot (heat sink) and cold heat exchangers are 300K and 80K, respectively. Since we could not find the friction factor and heat transfer coefficient for the oscillation gas flow, we used the correlation equation based on the steady flow through a stack of screens given by Kays and London[8] with correction factor based on our experimental results for the high frequency oscillating flow regenerator[9] in which the value of the cycled-averaged pressure drop of the oscillating flow in the regenerator is two to three times higher than that of a steady flow at the same Reynolds numbers based on the cross-sectional mean velocity. After a grid refinement study, all calculations presented here were carried out using a 200 non-uniform grid size, the calculated results with finer grid do not reveal noticeable changes. The grid is finer for the pulse tube and regenerator volume near the heater exchanger in order to identify effectively the shape gradient of temperature profile. The solution was obtained through the entire pulse tube cooler, from compressor to reservoir. With the oscillation compressible gas flow condition and the conjugate thermal conduction and convection between the gas flow and the solid matrix in regenerator, the solution took a long time to obtain a convergent solution, requiring 200~300 CPU mm on HP 5/90C mainframe. Use of previous converged solution at different conditions reduce the run times somewhat. NUMERICAL ANALYSIS
As we know, the right conditions for cooling to occur in the pulse tube cooler are that the amplitude of the gas mass flow and pressure oscillations must be large enough and the phase shift between the gas mass flow and pressure must be appropriate to carry the heat away (by enthalpy flow) from the cold point (cold heat exchanger). The size and geometry of each component of the pulse tube cooler and the orifice, double-inlet and multi-bypass work together to control the amplitude and phase shift between the gas mass flow and pressure to maximize the cooling
capacity of the pulse tube cooler. The optimal size and geometry of the pulse tube cooler and the opening condition of the valves can be obtained with numerical analysis and proper design. Our design approach using the numerical simulation to the co-axial miniature pulse tube cooler has three stages. The first stage is to choose the main size and geometry of each
component, such as the regenerator, pulse tube, aftercooler, cold and hot end heat exchangers. The optimal size and geometry of each component for this co-axial miniature pulse tube
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cooler were determined by the principle of achieving maximum cooling capacity at 80K with minimal input power from the compressor by the aid of numerical analysis. The cooler utilized a valveless rotary drive compressor. The swept volume of the compressor is 1.66cc. The simple sinusoidal oscillations for the rotary drive compressor was assumed in our numerical simulation.
The main optimal size of each component for the co-axial miniature pulse tube cooler are as follows: the pulse tube is 5mm in inner diameter and 70mm in length; the regenerator is 8mm in inner diameter and 58mm in length, filled with 400 mesh stainless steel screen; the pulse tube and the regenerator are made of stainless steel with thickness of 0.15mm; the length and the outside diameter of the cold head were 8mm and 6mm, respectively; the reservoir volume is 22cc. When the main size and geometry of each component for the pulse tube cooler were obtained. The second stage of the numerical simulation is to predict the oscillating flow characteristics and determine the optimal opening values of the orifice, double-inlet and multibypass valves for the pulse tube cooler. A typical calculation was made for the pulse tube cooler to determine these opening values. The operating wall temperature of the cold end and hot end heat exchangers are 80K and 300K, respectively. The operating mean pressure is 2.0MPa, the frequency f = 25Hz. Figures la~f showed the effects of the orifice opening values on the dynamic parameters in the pulse tube cooler. Figures 1a and 1b showed the instantaneous temperature at the two ends of the pulse tube for different orifice setting. The cold end temperature of the pulse tube cryocooler achieved the lowest temperature at the orifice setting of Figures 1c and 1d illustrated the gas mass flow rate at the two ends of the pulse tube for different orifice opening values. The orifice opening has large effect on the mass flow rate. The larger the orifice opening, the higher the gas mass flow rate. While the double-inlet opening has little effect on the gas mass flow rate in our numerical simulation. The dynamic pressures in the compressor and in the pulse tube for different opening values of the orifice valve were given in figures 1e and 1f. The pressure wave in the compressor remains almost the same for different orifice opening values, while the pressure wave in pulse tube decreased with the increased of the orifice opening.
The instantaneous variations of gas temperature, mass flow rate and dynamic pressure in the pulse tube cooler at the optimum opening value of the orifice, double-inlet and multi-bypass valves were compared in figure 2a~f. From these figures, we can clearly see the influences of the orifice, double-inlet and multi-bypass versions on the amplitude and phase shift between the mass flow rate and dynamic pressure in the miniature pulse tube cooler. Figure 2(a) showed that the multi-bypass version can achieve the lowest temperature for the pulse tube cooler. Figure 2(c) showed that the multi-bypass version can decrease the gas mass flow into the cold end of the pulse tube and the dead gas volume in the regenerator because the multi-bypass version allows a part of the gas flow into the pulse tube directly from the regenerator, results in increasing the performance of the regenerator. Figures 2(e) and 2(f) showed that the multi-bypass version can increase the pressure ratio, therefore, it can improve the refrigeration performance for the miniature pulse tube cooler. The third stage is to choose the optimal system mean pressure and operating frequency to the maximum cooling capacity and minimum input PV work based on the optimal size of each component of the pulse tube cooler and the optimal opening conditions of the orifice, doubleinlet and multi-bypass valves. Figures 3 and 4 showed the predicted cooling capacity and PV work at 80K of the co-axial miniature pulse tube cooler with the system mean pressure and operating frequency, respectively. From these figures, we concluded that the optimal operating frequency is about 30Hz and the optimal system pressure is around 3.0MPa for the co-axial miniature the pulse tube cooler. The cooler should produce 1W of cooling power at 80K with a total input power of less than 30W and with a heat sink temperature of 300 K.
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Figure 1. Variation of the dynamic parameters with the orifice opening values. (a) temperature at cold end of pulse tube; (b) temperature at hot end of pulse tube;
(c) mass flow rate at cold end of pulse tube; (d) mass flow rate at hot end of pulse tube (e) dynamic pressure in the compressor; (f) dynamic pressure in the pulse tube.
DESIGN OF A COMPACT COAXIAL PULSE TUBE COOLER
Figure 2. Comparison of the instantaneous parameters for three versions of pulse tube coolers. (a) temperature at cold end of pulse tube; (b) temperature at hot end of pulse tube;
(c) mass flow rate at cold end of pulse tube; (d) mass flow rate at hot end of pulse tube; (e) dynamic pressure in the compressor; (f) dynamic pressure in the pulse tube
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Figure 3. Cooling capacity with the system mean pressure and operating frequency.
Figure 4. PV works with the system mean pressure and operating frequency.
EXPERIMENTAL DESIGN AND RESULTS
A co-axial miniature pulse tube cooler that was the similar to the split Stirling cooler was designed and constructed in our laboratory. This cooler utilized a valveless rotary drive 1.66cc compressor made by Cryotechnologies Corporation. The hot ends of the pulse tube and regenerator were mounted on a gas flow control unit. The gas flow from compressor to the pulse tube and the gas flow through the double-inlet and multibypass valves was also controlled by the unit. A double inlet valve was used between the hot end of the pulse tube and the regenerator, and a multi-bypass tube was added to connect the middle parts of the pulse tube and regenerator. A reservoir volume of was mounted on the backside of the unit through an orifice valve. The orifice and double-inlet valves were all selfmade adjustable needle valves.
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Figure 5. Cooling capacity as a function of the cold head temperature.
The pulse tube and regenerator were made of stainless steel with thickness of 0.15mm. The pulse tube was placed within the annular regenerator. The flow channels of the pulse tube and regenerator were connected by the cold heat exchanger. This design was very compact and could be made into a commercially available pulse tube cooler. The measured cooling capacity for different input power to the compressor as a function of the cold head temperature was shown in figure 5. A lowest temperature of 66.5K was obtained with operating frequency of 30Hz and system mean pressure of 4.0MPa. It took about 10 minutes
to reach the lowest no-load temperature. The maximum cooling power of 0.6W at 80K was achieved with a total input electronic power of 66W. CONCLUSIONS
A study and design of a 1W / 80K co-axial miniature pulse tube cooler was reported in this paper based on the analyses of a numerical simulation for understanding the dynamic parameters
and internal process occurring in a pulse tube cooler. The simulation allowed clear visualization of the instantaneous parameters variations and the entire process occurring in the pulse tube cooler and gave quantitative information on the dominant influence in the cooling capacity of the
co-axial miniature pulse tube cooler. The three stages of the design approach by the aid of numerical analysis on the pulse tube cooler were presented. The cooler should produce 1W of cooling capacity at 80K with a total input power of less than 30W from numerical prediction. Based on the numerical analyses, a practical co-axial miniature pulse tube cooler with a rotary drive compressor was designed and constructed. The swept volume of the compressor is 1.66cc. A lowest temperature of 66.5K and cooling capacity of 0.6W at 80K with a total input electronic power of 66W was achieved with system pressure of 4.0MPa and operating frequency of around 30Hz. The pulse tube cooler was very compact and commercially available for practical applications. ACKNOWLEDGMENT
The author gratefully acknowledges the support of the National Natural Science Foundation of China and K. C. Wang Education Foundation, Hong Kong.
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REFERENCES
1. Jr. Ross, R. G. JPL cooler development and test program overview, Cryocoolers 8, Plenum Press, New York (1995), pp. 173-181. 2. Chan, C. K. et al. Miniature pulse tube cooler, Cryocoolers 7, Plenum Press, New York (1993), pp.113-120, 1993.
3. Chan, C. K. Jaco, C and Nguyen, T. Advanced pulse tube head development, Cryocoolers 9, Plenum Press, New York (1997), pp.203-210.
4. Kanao, K. et al. A miniature pulse tube cooler for the temperature below 100K, Cryogenics, vol.34,
no.2(1994), pp.164-169. 5. Wang, C. Cai, J. H. Zhu, W. and Zhou, Y. Miniaturization of a co-axial pulse tube refrigerator with linear motor drive compressor, Adv. Cry. Eng., vol.41B (1996), pp. 1419-11426.
6. Ju, Y. L. Wang, C. and Zhou, Y. Numerical simulation and experimental verification of the oscillating flow in pulse tube refrigerator, Cryogenics, vol.38, No.2 (1998), pp.169-176.
7. Ju, Y. L. Wang, C. and Zhou, Y. Numerical simulation and experimental verification of oscillating flow in a multi-bypass pulse tube cooler with area change, Adv. Cry. Eng., vol.43 (1998) (in press). 8. Kays, W. M. and London, A. L. Compact Heat Exchangers, 2nd McGraw-Hill, New York (1964). 9. Ju, Y. L. Jiang, Y and Zhou, Y. Experimental study of the oscillating flow characteristics for the regenerator in pulse tube cryocooler, Cryogenics, vol.38 (1998) (in press).
Performances of Two Types of Miniature Multi-Bypass Coaxial Pulse Tube Refrigerators J. T. Liang, J. H. Yang, W. X. Zhu, Y. Zhou, and Y. L. Ju
Cryogenic Laboratory, Chinese Academy of Sciences, P. O. Box 2711 Beijing 100080, China
ABSTRACT For the purpose of cooling space-borne infrared devices, two types of miniature multi-bypass coaxial pulse tube refrigerators are currently under development in our laboratory. The Oxfordstyle compressors with flexure bearings and clearance seals made by our laboratory are used to generate the pressure waves. The first type of miniature pulse tube refrigerator is designed for a compressor with about 2 cc swept volume. It is expected to provide 1 W of cooling power at 80 K. The second type of miniature pulse tube refrigerator uses a compressor with 0.79 cc swept volume. It is designed to have a cooling capacity of 150 mW at 85 K. At present the first one, coupled with a 1.66 cc compressor, reaches 66 K with no cooling load and produces 0.6 W of cooling power at 80 K with 66 W compressor input power, while the second one reaches 82 K and produces 100 mW at 85 K with 29 W input power.
INTRODUCTION The radiant cooler is a well developed technology for cooling the photo-electrical detector on long term practical satellites. However the radiant cooler may be inadequate in some cases due to
insufficient cooling capacity, posture of the satellite, etc. Miniature mechanical cryocoolers with small cooling power at 80 - 85 K are thus required. Such mechanical cryocoolers should have high efficiency, long life time, and high reliability. With the continuous improvements in efficiency in recent years, the pulse tube refrigerator (PTR) is now a good choice for such low power space applications.1 Compared with Stirling refrigerator, the PTR has no moving parts at low temperature. So it is simple and reliable, and its cold head produces low vibration and low electromagnetic interference. In our laboratory, the efficiency of PTR is presently lower than that of Stirling cryocoolers, although a PTR with efficiencies higher than aerospace Stirling machines has been reported by TRW. 2 According to our calculation and experiments, the multi-bypass method is an effective way to improve the efficiency of PTR.3 The effect of the multi-bypass method depends on the temperature distribution along the regenerator and the pulse tube, the position and opening of each bypass, etc. On the other hand it increases the mechanical complexity of the cold head. The regenerator and the pulse tube may be arranged in reciprocal U-shape, in line, and in coaxial configuration. The coaxial design is the most compact and the most convenient for
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coupling with the devices to be cooled. Particularly, with a coaxial design, it is easy to use the multi-bypass technique to improve the refrigeration performance. In this case, no connecting capillaries are necessary for the multi-bypass between the regenerator and the pulse tube. A few holes on the wall of pulse tube suffice. However, the coaxial design is more complicated than the linear and the U-shape designs. It may involve more losses. Careful design for the flow channel is important in this case. By optimizing the coaxial configuration, refrigeration performances comparable to the U-shape design can be achieved. The Oxford-style compressors are also under development in our laboratory.4 They are used to drive the pulse tube cold head. The following are the performances of two types of miniature multi-bypass coaxial pulse tube refrigerators. THE 2 CC CLASS MINIATURE PTR
The cold head of this type of pulse tube cryocooler is designed for an Oxford-style compressor with 2 cc swept volume. The pulse tube is set inside the annular regenerator with preset holes on its
wall for the multi-bypass effects. The regenerator wall and the gas reservoir are made of stainless steel. The cold head is made of copper to enhance heat transfer. Stainless steel screens are used as regenerator matrix. The working fluid is pure helium gas. The orifice and the double inlet passages are all integrated in the gas reservoir which is directly installed on the hot end of regenerator and pulse tube. The cooler looks like a split-Stirling cryocooler. A capillary tube connects the cold
head to the compressor. The 2 cc compressor with flexure bearing and clearance seal is being developed in our laboratory and it is not yet available for the experiment. To test the cold heads designed and
fabricated, two other compressors have been used. The first one was a 2 cc linear dual piston compressor with coil springs. The second one was a rotary compressor with 1.6 cc swept volume. The test results are as follows. Figures 1 and 2 show the experimental results obtained with the 2 cc linear compressor with coil spring. The working frequency is 50 Hz and the reject temperature is 300 K. At the fill pressure of 1.3 MPa, the lowest temperature reached was 63 K and 420 mW of net cooling power at 82 K have
Figure 1. Cooling power vs. temperature for PTR with 2 cc coil spring linear compressor.
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Figure 2. Cooling power at 150 K vs. Power input for the PTR with 2 cc coil spring linear compressor.
been achieved. The electric power input decreases as the cold end temperature increases. At 63 K the electric power input is at its maximum, 71 W, while at 70 K the electric power input is reduced to 47 W. It can be seen from Figure 2 that at the fill pressure of 1.45 MPa, about half Watt of cooling power is available at 150 K with 13 W of electric power input. The experimental results obtained with the 1.6 cc rotary compressor are shown in Figures 3, 4 and 5. Compared with the 2 cc linear compressor with coil springs, the 1.6 cc rotary compressor operates at lower frequencies and at higher fill pressures. Its frequency is 30 Hz, and the fill pressure can be up to 4 MPa. At the fill pressure of 4 MPa, the cold end temperature of 66 K was
Figure 3. Cool down curve for the PTR with 1.6 cc rotary compressor.
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Figure 4. Cooling power vs. Temperature for the PTR with 1.6 cc rotary compressor.
reached with no cooling load and about 600 mW of cooling power was achieved at 80 K, the compressor power input being 66 W. Figure 3 shows the cool down curve of the cooler. Figure 4 shows the cooling power in function of temperature at different fill pressures. The refrigeration performance improves significantly with the increase of fill pressure. As can be seen from Figure 5 in which the vertical axis indicates the cold end temperature in term of the output of thermocouple, the operation of the cooler is quite stable. Since the cold heads are not specially designed for these two compressors, the above experimental results are not the optimized results. But these results can give us an idea of what
Figure 5. Temperature stability of the PTR with 1.6 cc rotary compressor.
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Figure 6. Schematic of 0.79 cc linear compressor with flexure bearing and clearance seal.
performance we can get when the multi-bypass coaxial cold head is coupled with the Oxford-style compressor for which it is designed. For a pulse tube refrigerator with linear compressor, the optimization of the system is much more difficult than with a rotary compressor. That is because the swept volume of a rotary compressor is always constant, whereas the swept volume of a linear compressor is not constant and the resonant characteristics must be taken into account. When the 2 cc Oxford-style compressor is ready, better performances can be achieved by adjusting both the compressor and the cold head so that they will be well adapted to each other. THE 0.79 CC CLASS MINIATURE PTR
This type of miniature multi-bypass coaxial pulse tube refrigerator uses a 0.79 cc swept volume Oxford-style compressor. The 0.79 cc linear compressor developed by our laboratory is schematically shown in Figure 6. It is a one piston compressor with flexure bearing and clearance
Figure 7. A cool down curve of the PTR with a 0.79 cc linear compressor.
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Figure 8. Cold end temperature vs. Frequency for the PTR with a 0.79 cc linear compressor.
seal. It is about 80 mm in outer diameter and 120 mm in length. Its weight is less than 1.8 Kg.
According to the design, 0.79 cc is the maximum swept volume of it. The configuration of the cold head is similar to that of the 2 cc class pulse tube cooler except that the dimensions are smaller. In the cold head the connections are made by welding. The weight of the cold head (including the gas reservoir) is less than 0.2 Kg.
Figure 7 shows a cool down curve of the cooler at the frequency of 67 Hz and the fill pressure of 2.7 MPa. The lowest temperature of 81.5 K was reached in about 25 minutes with no cooling load applied at the cold end. The cooler was capable of providing about 100 mW of cooling power at 85
Figure 9. Cold end temperature vs. Fill pressure for the PTR with a 0.79 cc linear compressor.
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K with 29 W compressor power input. The dependence of the cold end temperature on the operation frequency is shown in Figure 8. For this cooler 67 Hz is the optimum frequency. The
effect of the fill pressure on the lowest achievable cold end temperature has also been studied. As can be seen from Figure 9, the performance improves as the fill pressure increases. We are currently improving both the compressor and the coaxial multi-bypass cold head in
order to reduce the compressor power input and to increase the cooling power to 150 mW at 85 K. CONCLUSION
Two types of miniature coaxial multi-bypass pulse tube cryocooler destined for space infrared cooling have been developed and tested. The 2 cc class miniature PTR reached the lowest temperature of 63 K and produced 420 mW cooling power at 82 K when driven by a 2 cc swept volume linear dual-piston compressor with coil springs. It reached the lowest temperature of 66 K and produced 600 mW cooling power at 80 K when driven by a 1.6 cc swept volume rotary compressor. The 0.79 cc class miniature PTR with an Oxford-style linear compressor reached the lowest temperature of 81.5 K and produced nearly 100 mW cooling power at 85 K. The performance of the miniature PTR’s which are not yet optimized show that the compact multibypass design is suitable for the miniature high frequency pulse tube cryocoolers. ACKNOWLEDGEMENT
This work is supported by the National Natural Science Foundation of China. The author gratefully acknowledges the support of K. C. Wong Education Foundation, Hong Kong. REFERENCES 1.
Duband, L., Ravex, A., Bradshaw, T., Orlowska, A., Jewell, C., and Jones, B., “50 - 80 K Pulse Tube Cryocooler Development”, Cryocoolers 9, Plenum Press, New York (1997), pp. 213-221.
2.
Burt, W.W. and Chan, C.K., “New Mid-Size High Efficiency Pulse Tube Coolers”, Cryocoolers 9, Plenum Press, New York (1997), pp. 173-182.
3.
Cai, J.H., Wang, J.J., Zhu, W.X., and Zhou, Y., “Experimental Analysis of the Multi-bypass Principle in Pulse Tube Refrigerators”, Cryogenics, vol. 34, no. 9 (1994), pp. 713-715.
4. Yang, J.H., Zhou, Y., Tao, Z.D., Zhu, W.X., Cai, J.H., Liang, J.T., and Ju, Y.L., “Experimental Study
on Miniature Pulse Tube Cryocooler”, Proceedings of the 5th Japanese-Sino Joint Seminar on
Cryocooler and Its Applications (JSJS-5), Osaka, Japan, September 16-20, 1997, pp. 179-183.
Development of a 5 to 20 W at 80 K GM Pulse Tube Cryocooler S. Fujimoto and Y.M. Kang
MEC Laboratory, DAIKIN INDUSTRIES, LTD. 3 Miyukigaoka, Tsukuba 305-0841, Japan Y. Matsubara Nihon University, Atomic Energy Research Institute 7-24-1 Narashinodai, Funabashi, Chiba 274-0063, Japan
ABSTRACT
In order to select the most suitable phase shifting mechanism for the GM pulse tube coolers applied to cryoelectronic devices, we evaluated the performance of both double-inlet and fourvalve type phase shifting mechanisms by using a single-stage GM pulse tube cooler and an on-
off timing control system designed for valved cryocoolers. The lowest temperature achieved was 29.1K and the largest cooling capacity at 80K and 60K was 13 and 10W, respectively. For the four-valve type cooler, the cooling capacity at 80K was 72% of that for the double inlet type cooler. We also evaluated the dependency of the operating frequency on the inclination of the pulse tube. When the operating frequency was higher than 6Hz, the effect of the pulse tube inclination was negligible. We estimated the cooling performance by simplified numerical loss analysis based on the equivalent PV method. The calculated performance was larger than the measured performance both for the double-inlet and for the four-valve cryocoolers. The calculated results also indicate that the stroke of the gas displacer in the pulse tube is large relative to the pulsetube length, so that the enthalpy loss in the pulse tube may be significant. INTRODUCTION
Pulse tube cryocoolers applied to cryoelectronic devices must be reliable and efficient. The phase-shifting mechanism strongly affects the reliability and efficiency of the system including the cryoelectronic devices. We evaluated both double-inlet and four-valve type phase-shifting mechanisms by using a single-stage GM pulse tube cooler and an on-off timing control system for valved cryocoolers. For various applications, cryocoolers need to be inclined during operation. When the pulse tube cylinder is inclined to the vertical direction, natural convection occurs, and this decreases the cooling capacity. We therefore analyzed the influence of the cryocooler inclination on cooling capacity. For cooling cryoelectronic devices that are sensitive to electromagnetic noise, such as SQUID, the valve motor unit must be separated from the cooler. We therefore also evaluated the influence of the separation length on the cooling capacity To evaluate the experimental results and to improve the cryocooler performance, we Cryocoolers 10, edited by R. G. Ross, Jr. Kluwer Academic/Plenum Publishers, 1999
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estimated the the PV work in the cold end volume of the pulse tube and heat loss in the cooler by numerical analysis based on the equivalent PV method. EXPERIMENTAL APPARATUS
Figure 1 shows the schematic of the test apparatus. To test various types of valved pulse tube cryocoolers, we made the valve box unit controlled by the personal computer. Of the ten solenoid valves in the valve box, four were used to control the high pressure gas from the compressor to the regenerator, four were used to exhaust the gas from the regenerator to the compressor, and two were used to control the gas to and from the hot end of the pulse tube. The valve timings were controlled by a control system mounted in a personal computer. To imitate rotary valve conditions, the valves for the regenerator were opened sequentially. The valve timing data was input into the personal computer, so that even after the valve operation was started, the timing of every valve could be changed during cooler operation. For example, the valve timing chart for the four-valve type cooler is shown in Figure 2. The compressor unit used in these tests was the DAIKIN model U102CW, which is commonly used for GM cryocoolers and has a rated input power of 1.5kW at 50Hz. A sectional view of the pulse tube cooler is shown in Figure 3. Separate flanges were welded to the pulse tube cylinder and the regenerator cylinder. This allowed us to test various combinations of the two cylinders. For a given cylinder diameter, we can test various lengths of the pulse tube by using a sliding mechanism of the hot end strainer, which allowed us to change the pulse tube length during operation. We also used spacers of different length in the regenerator hot end. The original length of the pulse tube and regenerator cylinders was 200mm and 180mm, respectively. We were able to vary the pulse tube length between 170mm and 230 mm during the operation. NUMERICAL ANALYSIS
To evaluate the experimental results and to improve the cryocooler performance, we estimated the PV work in the cold end volume and heat loss in the cooler by numerical analysis
Figure 1. Schematic of test apparatus.
Figure 2. Sample valve timing chart.
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Figure 3. Sectional view of single-stage GM pulse tube cooler.
based on the equivalent PV method
1
. The pulse tube can be divided into three parts: I) the hot
end of the pulse tube, II) the gas displacer, and III) the cold end of the pulse tube. The mass flow
rate through the orifices and the regenerator can be calculated as follow :
where
are the coefficients of discharge at the orifice valve, the double inlet valve,
and the regenerator, respectively, and are the pressures in the pulse tube, the buffer tank, and the regenerator hot end, respectively. For the coefficients of discharge, we used values measured during a steady-flow test. We obtained the gas displacer volume in the pulse tube by
solving the differential equations. The equivalent PV work and the cooling capacity at the cold end can be represented as
where is a regenerator loss. We used the Schalkwijk's equation for the effectiveness of the regenerator 2 . For heat transfer characteristics of the stacked screen matrix, we used the experimental data by Kays and London 3 . Hoth represents other heat losses, such as conductive
and radiative heat losses. RESULTS AND DISCUSSION
Double-Inlet Type and Four-Valve Type
The best performance of the cooler was obtained with the double-inlet configuration and when the rotary valve unit was used. For the valve operated at 2.4Hz and a pulse tube length of
220mm, the cool-down time from room temperature to 80K was about 27 minutes.
The lowest
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temperature was 29.1K. The cooling capacity at 80K and 60K was 13 and 10W, respectively. Although the best performance of the cooler was achieved with the double-inlet configuration, the four-valve type pulse tube cooler has other advantages compared with the double-inlet type cooler. For example, simple analysis showed that the four-valve type cooler is more efficient than the double-inlet type 4 . We tested the efficiency of the four-valve type cooler using the test apparatus shown in Figure 1.
To compare the efficiency of the four-valve with the double-inlet type cooler, first we evaluated the performance of the valve box unit controlled by the personal computer. We did this by operating the double-inlet type cooler using the valve box and compared the cooling performance with that by using the rotary valve. The lowest temperature achieved was 30.5K,
similar to that achieved using the rotary valve. The cooling performance degraded about 10% compared to using the rotary valve. Next, we operated the four-valve type cooler using the valve box. The size of the pulse tube and the regenerator and the operating frequency were the same as
those for the double-inlet type. Figure 4 shows the cooling performance of the four-valve and double-inlet type coolers. The lowest temperature achieved was 31 .0K. The cooling capacity of the four-valve type at 80K was 72% of that for the double-inlet type. For the four-valve type, the reason for the cooling performance degradation may be related to the effect of the valve timing on the phase shift of the gas flow to the regenerator (main
valves) and the gas flow to the pulse tube hot end (sub valves). The valve timing may not be optimized and the flow pattern in the sub valves of the solenoid valves may be different from that of the rotary valve. Numerical results shown later in this section support this conclusion. Pulse Tube Size
To evaluate the influence of the pulse tube size, we operated the double-inlet type cooler at a frequency of 2.4Hz and varied the pulse tube length between 170mm and 230mm. The cooling capacity at 80K (Figure 5) increased with increasing pulse-tube length. Figure 5 shows that for a
frequency of 2.4Hz, the optimum length was larger than 230mm. This indicates that the compressor capacity was too large for the pulse tube used in these tests. Effect of Pulse Tube Inclination
When the pulse tube is inclined to the vertical direction, the cooling capacity is degraded by
convective heat loss in the tube5 . Because pulse tube coolers are used in applications that may require such inclination, operating conditions must be determined to minimize the effect of this inclination. In general, increasing operating frequency reduces the convective heat loss. We therefore measured the effect of the operating frequency on the cooling capacity. At first, the cooler was oriented as shown in Figure 3, with the cold end downward, and a pulse tube length of 220mm. We varied the operating frequency from 2 to 10Hz. At each condition, we measured the lowest temperature achieved and the cooling capacity (Figure 6a). The lowest temperature
Figure 4. Performance of four-valve type pulse tube cooler.
Figure 5. Pulse tube length and performance.
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Figure 6. Influence of pulse tube inclination.
was achieved at an operating frequency of 2.4Hz, and the maximum cooling capacity at 80K was achieved at an operating frequency of 3.0Hz (Figure 6c). This demonstrates that increasing
frequency decreases the cooling performance. To determine the effect of inclination of the pulse tube on the performance, we turned the cooler upside down and repeated the measurements (Figure 6b). At a frequency of 6Hz the lowest temperature achieved was 47.5K and the cooling capacity at 80K was 10.5W. The cooling capacity was almost independent of operating frequency between 4 and 8Hz. At a frequency of 2.4Hz, the lowest temperature achieved was about 60K and the cooling capacity at 80K was approximately 4W, which is less than a third of the cooling capacity achieved with the cold end oriented downward. The cooling capacity at 80K for both downward and upward orientations is shown in Figure 6(c). For an operating frequency higher than 6Hz, the difference between the two orientations is small, indicating that convective heat loss in the pulse tube becomes insignificant when the operating frequency is higher than 6Hz.
Effect of Directional Characteristic of Double-Inlet Valve
When a needle valve is used as the double-inlet valve, the reliability and the performance of the cooler is degraded by the directional characteristics of the valve. We used needle valves for the double-inlet and the orifice valves. We evaluated the directional characteristics of these valves by a steady-flow test. The difference in the coefficient of discharge between the positive
direction and the negative direction was about 10%. At first, we oriented the valve in the positive direction, as shown by the arrow in Figure 7. The cooling performance for this orientation is indicated by the circles and the dashed line in Figure 7.
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Figure 7. Effect of directional characteristic of double inlet valve.
Similarly, the performance corresponding to the valve oriented in the negative direction is indicated by the triangles and the dashed line. The difference between these two orientations is noted by the lower temperature and increased cooling capacity achieved with the valve oriented in the positive direction. To compensate for the directional characteristics, we connected a buffer tank through a bypass line to either the exhaust line or to the intake line of the compressor through another needle valve. When the double-inlet valve was oriented in the positive direction, the bypass line should be connected to the intake line (low pressure volume). The performance under these conditions is indicated by the circles and the solid line. The performance corresponding to the double inlet valve oriented in the negative direction and the bypass line connected to the exhaust line (high pressure volume) is indicated by the triangles and the solid line. These results show that the bypass line between the buffer tank and the high/low pressure volume is therefore useful for compensating the directional characteristics of the double inlet valve. This bypass line may function similar to double-orifice valves in two-stage pulse-tube cryocoolers 6. Valve Motor Separation
When the GM pulse tube cooler is used to cool sensors that are sensitive to electromagnetic noise, such as SQUID, the valve motor unit must be separated from the cooler. For example, in the GM/JT cryocooler used to cool SQUID sensors in biomagnetic measurement systems, the valve motor unit is separated from the expander and set outside a magnetically shielded room. In this system, the separation is about 3m7 . We tested the performance degradation of the valveseparated pulse tube cooler. When the valve motor was separated either 1 or 2m, the cooling capacity decreased about 10 or 15%, respectively (Figure 8). Numerical Estimate of Cooling Capacity Based on Equivalent PV Method Using the numerical analysis based on the equivalent PV method, we calculated the cooling capacity of the double-inlet and four-valve coolers. We summarize the results in Table 1, where
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Figure 8.. Effect of valve motor separation.
is the stroke of the gas displacer and Lpt is the length of the pulse tube. The calculated performance is higher than the experimental performance for both types of coolers. The reason for this difference may be because the stroke of the gas displacer in the pulse tube is large compared with the pulse-tube length, so that the enthalpy loss in the pulse tube may be significant. CONCLUSIONS
We developed a single-stage GM pulse-tube cooler and evaluated the efficiency of the double-inlet and the four-valve type coolers. We also evaluated the dependency of the operating frequency on the inclination of the pulse tube. The major results are as follows. 1. The optimum cooling capacity was achieved for the double-inlet configuration using the rotary valve. The lowest temperature achieved was 29.1K and the cooling capacity at 80K and 60 K was 13 and 10W, respectively. 2. For the four-valve type cooler, the cooling capacity at 80K was 72% of that for the doubleinlet type cooler. The cause of the cooling performance degradation may be improper valve timing and the difference in flow pattern in the solenoid valves for the pulse-tube hot end compared with that in the rotary valve. 3. When the operating frequency was higher than 6Hz, the effect of the pulse tube inclination on the cooling capacity was negligible. This indicates that convective heat loss in the pulse tube is negligible. 4. A bypass line between the buffer tank and either the high- or low-pressure volume was effective in compensating the directional characteristics of the double-inlet valve.
5. Our numerical analysis based on the equivalent PV method indicates that the stroke of the gas
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displacer in the pulse tube is large relative to the pulse-tube length, so that the enthalpy loss in the pulse tube may be significant. This indicates that the pulse tube size is too small for the capacity of the compressor. REFERENCES 1. Zhu, S., Wu, P. and Chen, Z., "Double inlet pulse tube refrigerators : an important improvement", Cryogenics, Vol. 30 (1990), pp.514
2. Schalkwijk, W. F., "A Simplified Regenerator Theory", Trans. of the ASME (1959), pp.142 3. Kays, W. M. and London, A., L., "Compact Heat Exchangers", McGraw-Hill Book Company (1976), pp.140 4. Matsubara, Y., "Work-loss Distribution on GM-type Pulse Tube Coolers", Jap. Journ. of Cryog. Eng., Vol.33, No.4(1998), pp.207 5. Thummes, G., Schreiber, M., Landgraf, R. and Heiden, C., "Convective Heat Losses in Pulse Tube Coolers : Effect of Pulse Tube Inclination", Cryocoolers 9, Plenum Press, New York (1997), pp.393 6. Chen, G., Qiu, L., Zheng, J., Yan, P., Gan, Z., Bai, X. and Huang, Z., "Experimental study on a double-orifice two-stage pulse tube refrigerator", Cryogenics, Vol.37 (1997), pp.271 7. Sata, K., Fujimoto, S., Fukui, N., Haraguchi, E., Kido, T., Nishiguchi, K. and Kang, Y., M., "A 61-channel SQUID System for MEG Measurement Cooled by a GM/JT Cryocooler", IEEE Trans. appl. Supercond., Vol.7, No. 2 (1997), pp.2526
Conceptual Design of Space Qualified 4K Pulse Tube Cryocooler G.R. Chandratilleke, Y. Ohtani, H. Nakagome, K. Mimura Toshiba Corporation, Kawasaki, Japan, 210-0862 N. Yoshimura and Y. Matsubara Nihon University, Funabashi, Japan, 274 H. Okuda Institute of Space and Astronautical Science (ISAS) Sagamihara, Japan, 229 T. Iida and S. Shinohara National Space Development Agency (NASDA) Tsukuba, Japan, 305
ABSTRACT Pulse tube cryocoolers for space applications in general operate at high frequencies and high mean pressures in order to make the cooler system compact. However, with high operating frequencies, it is know that there is a limit to the lowest temperature a cryocooler can reach because of degradation of regenerator performance. There is evidence to show that a cryocooler in multistage configuration has to operate at 1 to 2 Hz to obtain temperatures in the 4K region. Therefore, in order to obtain a compact cryocooler producing 4K, we devised a multistage configuration in which a single stage 4K pulse tube is thermally coupled to a conventional cryocooler. Thereby, it is possible to operate the 4K pulse tube at a low frequency and the high-temperature stage at a high frequency and achieve a reasonably compact system for space use. We shall qualitatively investigate in this paper why a low frequency operation is required for achieving low temperature levels. INTRODUCTION
For a pulse tube cryocooler to be qualified for space applications, it has to address the following factors: reliability, efficiency, weight and volume. Reliability implies life of the cryocooler, its temperature stability and vibration-free operation both mechanically and magnetically. In addressing the size of the pulse tube to used, a major factor will be the operating frequency of the pulse tube. On the other hand, in addressing the reliability and efficiency of the cryocooler, the type of compressor used plays an important role. From the viewpoint of the driving frequency, pulse tube cryocoolers can be classified into two categories: low frequency (a few Hertz) and high frequency (several tens of Hertz) pulse tube
cryocoolers. In low-frequency cryocoolers, a compressor producing a continuous high pressure mass flow uses a set of flow switching valves to generate pressure oscillations in the pulse tube. This method of producing a pressure oscillation is called valved-compressor method or GM type compressor method, where GM is the acronym for Gifford-McMahon. In high-frequency pulse tubes, pressure oscillations are created by the movement of a piston directly connected to the pulse tube and controlled by a linear motor. This method is called valueless-compressor method or Stirling-type compressor method. Selection of the gas compression method for a pulse tube depends on the temperature range that the cryocooler is intended for. For instance, in the case of a single-stage pulse tube Cryocoolers 10, edited by R. G. Ross, Jr.
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cooler that provides temperatures in 30K to 80K range, the Stirling type compressor is more appropriate because it has a high efficiency and results in a lower total weight. The linear compressor is also appropriate in this case because it works stably at high frequencies than at low frequencies, which has no ill-effect in the temperature range concerned.(1) On the other hand, to attain temperatures lower than 30K, it will be required to use a lower frequency and therefore the GM-type gas compression method is more suitable. In the Stirling type compressor, if the driving frequency is lower than 7Hz, piston's resonance condition will require the piston to be heavier and the spring to be less stiff; Under this condition, it is difficult to stabilize the piston against changes in pressure or gravity. Several experimental work have in fact (2,3) indicated that low-frequency operation is required for achieving low terminal temperatures. In addition, to achieve lower temperatures, pulse tube coolers have to be multi(2,3) staged . In a conventional 3-stage pulse tube cooler we developed using in its third stage regenerator, the stage temperatures were 88K, 33K and 10.6K at l.7MPa mean pressure and pressure ratio of 2.(4) In this paper, we shall investigate why a low-frequency operation is required for achieving 4 K temperature levels, and then propose a new hybrid cryocooler for space-applications requiring 4 K temperature level. We will first examine whether the regenerator performance will degrade owing to occurrence of poor heat transfer between gas and regenerator material. Next, we shall pay attention to enthalpy flow within the regenerator to see if it has a frequency dependency. COMPARISON OF DIFFUSION DEPTHS We first checked on diffusion depths in helium gas and regenerator material Er3Ni to find if they have an effect on regenerator degradation. The diffusion depth, d, in a semi-infinite solid medium is given by the following equation for the condition that its surface temperature varies at frequency, f, and the depth is defined as the distance from the surface to a location where the temperature drops to 1/e of the surface temperature amplitude. and c respectively are thermal conductivity and specific heat of the medium.
Figure 1 shows the diffusion depth in helium gas at 1.5MPa for the frequencies between 2Hz and 20Hz. The diffusion depth shows a minimum as a result of helium having a peak in specific heat near the critical point. There is a general tendency that the diffusion depth reduces as the frequency increases.
Figure 1 Diffusion depth in helium gas at 1.5MPa.
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Figure 2. Diffusion depth in regenerator material, Er 3 Ni.
Figure 2 shows the diffusion depth in in the temperature range 5 K to 20K as the frequency is varied between 1 Hz to 20 Hz. Here also the general tendency exists that when the frequency is increased, the diffusion depth in reduces. At larger frequencies like 20Hz, has a diffusion depth less than 0.18 mm in the temperature range 5 K to 20 K, and near 5 K it has dropped to about 0.13 mm. On the other hand, helium in the same temperature range, has a diffusion depth greater than 0.13 mm. Considering the fact that helium space is interstitial and is generally much less in size than 0.13 mm, it is therefore understood that heat diffusion from helium gas to the regenerator material is not a critical issue at high frequency.
ENTHALPY FLOW IN REGENERATOR Another factor that can affect cold-end cooling capacity is the enthalpy flow within the regenerator. We did the following analysis to evaluate available cold-end cooling capacity as the frequency is changed. Regenerator details and other conditions are given in Table. 1. To analyze enthalpy flow within the regenerator, we numerically solved by finite difference formulation the three equations: continuity equation, energy equation, and heat transfer equation, one-dimensionally. In addition, helium was treated as an ideal gas and pressure drop neglected. As for the time and space divisions, we divided one pressure cycle into 90 divisions while the regenerator length was divided into 60 equal divisions. In the calculations, we changed the pressure ratio by changing the pressure amplitude, with the mean pressure kept constant.
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Figure 3. Enthalpy flows at the 80K cold-end of regenerator for two different pressure ratios when mass flow rates are adjusted to produce the identical work flow at regenerator cold-end.
Figure 3 shows the enthalpy flow as a function of frequency for two pressure ratios
1.5 and 2 when the cold-end expansion work (PV work x frequency) is kept constant. To do this, the mass flow rate has to be changed from 0.2 g/sec at a pressure ratio of 2 to 0.34 g/sec at the lower pressure ratio, 1.5. It is seen that frequency dependence of enthalpy flow is different depending on the mass flow rate through the regenerator: even if mass flow rate is low, enthalpy flow is large if the pressure ratio is high. Irrespective of the value of pressure ratio, the enthalpy flow increases with frequency but the work flow remains independent of
the frequency. This implies that the cooling capacity available at the cold-end decreases with increased frequency, the maximum cooling capacity being the difference between the work flow and enthalpy flow.
To examine in more details why the cooling capacity decreases with increased frequency, temperature distribution in the regenerator was examined. Figures 4 and 5 show mean
temperature distributions in the regenerator for the same pressure ratios and mass flow rates as in Fig. 3 at the frequencies, 20 Hz and 4 Hz respectively. These two frequencies were selected because, for these frequencies, dependency of enthalpy on frequency is vastly different with the other parameters being constant.
Figure 4. Temperature distributions in the regenerator at the pressure ratios of 1.5 and 2 at 20Hz frequency when mass flow rates are adjusted to produce the identical work flow at the cold-end.
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Figure 5. Temperature distribution in the regenerator for two different pressure ratios at 4 Hz frequency when mass flow rates are adjusted to produce the identical work flow at the cold-end.
Comparing Figs. 4 and 5, it is seen that at high frequency, the temperature distribution has changed from a much linear one at 4 Hz as shown in Fig. 5 to a one that is steeper at low-temperature end as shown in Fig. 4. As a result, at high frequency, the regenerator has its heat exchange taking place mostly in the cold-end region and therefore, its effective length appears short. This causes the regenerator to become less efficient and the enthalpy flow through the regenerator to increase with an increase in frequency. Above calculations were done in the temperature range, 80 to 300 K. To examine the effect of frequency at a much lower cold-end temperature, the cold-end temperature was set to 30 K and the calculations redone. The mass flow rate was increased from 0.34 g/sec to 1.2 g/sec to compensate for the thermodynamic reduction of work at the cold-end. Figure 6 shows these results for the conditions: l.5MPa mean pressure and 0.3MPa pressure amplitude or a 1.5 pressure ratio. When the cold-end temperature is 30 K, the work and enthalpy lines intersect at about 8 Hz in contrast to 28 Hz when the cold-end temperature is 80 K. This shows that the lower the cold-end temperature of a regenerator, the lower the maximum frequency at which the regenerator can produce a cooling power.
Figure 6 Comparison of enthalpy flows at two cold-end temperature when the cold-end work flows are at comparable levels.
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Figure 7. Proposed thermally-coupled frequency-wise separated pulse tube configurations.
CONFIGURATIONS FOR 4 K SPACE QUALIFIED PULSE TUBE CRYOCOOLER
In order to realize a compact cryocooler for space applications, it is necessary that pulse tube cryocoolers be operated at high frequencies and high mean pressures. However, at high operation frequencies, we showed that the enthalpy flow in a regenerator is increased, the cooling load is reduced and the maximum frequency that produces a cooling power is also reduced. In other words, this implies that to obtain a high cooling power or a lower cold temperature, the frequency should be reduced. Therefore, it is mandatory that the low-temperature stage of the pulse tube be operated
at a low frequency. On the other hand, if the high temperature stage of the cryocooler is operated at high frequency, it is possible to obtain a higher efficiency and a lower overall
weight. (1) These two ideas can be put together in a single cryocooler model, where the
low-temperature stage operating at a low frequency is thermally coupled with a single or several high-temperature single stages operating at higher frequencies but thermally coupled with each other as shown in Fig. 7. By doing this, we can realize a compact cryocooler to produce 4K for space use. Such a configuration has the added advantage that cold-end phase shift control can be done in each stage independent of the other stages, which had been hitherto difficult.(4)
CONCLUSIONS It was shown that high frequency operation of a pulse tube is not suitable for achieving low temperatures owing to increased enthalpy loss in the regenerator. Though low frequency operation is essential for producing 4 K level temperatures, it is considered to be unsuitable
for space use in terms of cryocooler’s compactness. This issue can be resolved to realize a
compact 4 K pulse tube cryocooler for space applications by a new cryocooler concept. In this concept, a low frequency single stage cryocooler producing 4 K is thermally coupled with a high-frequency cryocooler that is efficient and compact. REFERENCES 1. Ohtani, Y., et al., “High Efficient Pulse Tube Refrigerator with Linear-Motor Drive Compressor”, Advances in Cryogenics, Plenum Press, New York (1994), pp. 1441-1448. 2. Matsubara, Y., and Gao, J. L., “Novel Configuration of Three-stage Pulse Tube Refrigerator for Temperatures below 4 K”, Cryogenics, vol. 37, no. 4 (1994), pp. 259-262. 3. Wang, C., et al., “A Two-Stage Pulse Tube Cooler Operating below 4 K ”, Cryogenics, vol. 37, no. 3 (1997), pp; 159-164.
4.
Ohtani, Y., et al., “Development of a Three Stage Pulse Tube Refrigerator”, Proceedings
of Fifth Japanese-Sino Joint Seminar on Cryocooler and its Applications, Osaka (1997), pp. 118-121.
Performance Dependence of 4K Pulse Tube Cryocooler on Working Pressure N. Yoshimura, S. L. Zhou and Y. Matsubara Nihon University Funabashi, 274 Japan
G. R. Chandratilleke, Y. Ohtani, and H. Nakagome
R&D Center, Toshiba Corporation Kawasaki, 210 Japan
H. Okuda Institute of Space and Astronautical Science (ISAS) Sagamihara, 229 Japan S. Shinohara National Space Development Agency of Japan (NASDA) Tsukuba, 305 Japan
ABSTRACT
Performance of pulse tube cryocoolers working around 4 K region is affected by the fact that helium gas deviates from the ideal gaseous state. Previous experimental studies are mostly done in the working pressure region of 1 to 2 MPa, similar to the commercially available 4 K G-
M coolers. However, the non-ideal state of the helium gas near 4 K region suggests that use of lower working gas pressure should give a better performance. In order to verify this hypothesis, an experimental apparatus consisting of a single stage pulse tube cryocooler was fabricated and tested at different base pressures down to 1 atmosphere. A two-staged G-M cooler was used as the pre-cooler to keep the pre-cooling temperature constant Experimental results show that a significant increase of cooling performance can be obtained near the plateau temperature region on T-S diagram by decreasing the mean working gas pressure as anticipated. INTRODUCTION
The minimum cooling temperature that a pulse tube cooler can provide, has reached liquid helium temperatures by using magnetic regenerator materials and multi-staging pulse tube. Operating conditions of such coolers are almost the same as that of G-M coolers, where a mean
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Figure 1. Schematics of experimental set up.
pressure of 1 to 2 MPa, a pressure ratio of 2 to 3 and a frequency of 1 to 2 Hz are used. From the thermodynamic viewpoint of helium gas as the working fluid, however, it is indicated that the cooling performance (cooling capacity / net input work) of such an operating condition may not be optimized for the temperature range below about 8 K where the non-ideality of the working gas properties becomes significant. This paper describes the performance dependence of a 4 K pulse tube cryocooler on working pressure. The mean working pressures of 1.5, 1.2 and 0.25 MPa (referred to as high, medium and low pressure cycle respectively) have been studied experimentally. EXPERIMENTAL RESULTS
Figure 1 shows the schematics of the experimental set up. It is known that a pulse tube can reach the temperature below 4 K in its cold end even if the hot end is kept at room temperature 1-5. Only the required condition is to pre-cool the regenerator by means of upper stage pulse tubes or by external cooling sources such as cryogens or other cryocoolers. In this experiment, a twostaged G-M cooler instead of a multi-staged pulse tube was used as the pre-cooler to avoid the interaction of upper stage pulse tube.
Cold ends each of first and second stage regenerators consisting of copper mesh and lead shot are pre-cooled by the first and second stage cold heads of the G-M cooler respectively Here the performances of these regenerators are ignored. Third stage regenerator consists of shot (0.3 mm in diameter) installed in a SUS304 tube of 17 mm in diameter, 64 mm in length and 0.5 mm in thickness. The inner diameter of pulse tube is 10 mm, thickness is 0.5 mm and length is 350 mm. Hot end temperature of third stage regenerator is kept at 12~13 K by the G-M cooler throughout the experiment. The pulse tube is controlled by double inlet method with additional second orifice proposed by Chen Guobang4. Oscillating pressure is given by a set of rotary valve and continuous flow compressor. A mass flow meter is installed on the low pressure line between the rotary valve and the compressor. Operating frequency of the pulse tube cooler is 0.5 Hz. Figure 2 shows the pressure wave measured at the hot end of the pulse tube when the cold
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Figure 2. Temperature and pressure wave .
Figure 3. Temperature and pressure trace on TS diagram.
end temperature of the pulse tube was at its minimum temperature. A calibrated Germanium thermometer attached to the connecting tube between the pulse tube and the regenerator cold end
has been used to measure this temperature with oscillations as shown in figure 2. Comparing the pressure wave obtained from high pressure cycle and low pressure cycle, the work loss within
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Figure 4. Cooling capacities of different pressure range.
Figure 5. Normalized Cooling capacities.
the regenerator for low pressure cycle may not be negligibly small in spite of the low driving frequency of 0.5 Hz The cooling capacity was measured by an electric heater provided on the cold end of the third stage regenerator. Results are given in figure 3 and figure 4. Figure 3 shows the measured pressure and temperature trace on a TS diagram of helium gas. Traces between the isobaric lines of 2MPa and 1MPa, 0.5MPa and 0.12MPa are the results of high pressure cycle and low pressure cycle respectively. The trace marked with “0(mW)” corresponds to the zero-cooling capacity, it changes along the direction of the arrow when heating power is increased. The uppermost one shown in this figure corresponds to the heating power of 213mW. It should be noted that in the case of low pressure cycle, some of the traces
cover the liquid phase area. That means helium gas within the pulse tube is partially liquefied. Figure 4 shows the experimental results of cooling capacity for high, medium and low pressure cycles. Ph and Pl refer to maximum and minimum pressure respectively. The cooling capacity of 80 mW has been obtained at almost the same temperature (~5 K) for these three cycles with different mean pressures. The required compressor work, Win, for each pressure cycle, however, is different because of different mass flow rate and pressure ratio. Assuming isothermal compression work for the
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measured mass flow rates, 472 W, 383 W and 170 W are calculated as the required work for high,
medium and low pressure cycles respectively. Therefore, each cooling capacity should be normalized by the respective isothermal compression work to obtain the performance dependence on working pressure.
Results of normalized cooling capacities are shown in figure 5. Horizontal bar of each data point indicates the amplitude of the temperature oscillation. It clearly shows that the low pressure
cycle is advantageous over the high and medium pressure cycles for the cooling temperatures above 5 K. The performance of the medium pressure cycle is also increased slightly better than that of the high pressure cycle above this temperature, whereas the lowest cooling temperature is achieved by the high pressure cycle and is 2.5 K. DISCUSSIONS
To confirm these experimental results, following calculations have been carried out. Assuming an ideal regenerator below the constant pre-cooling temperature, like an idealized counter flow heat exchanger, relation between the work flow and enthalpy flow for helium gas was calculated. Figure 6 and figure 7 show the results of the calculation for the fixed pre-cooling temperature of 13K. of the regenerator. Numbers indicated for each curve is referred to as maximum and minimum pressures in MPa, and corresponds to that of the experiment. The cooling capacity, Q, is the difference between the work flow, Wf, at the cold end and the enthalpy flow, Hf, through the regenerator. The input work, Win, is the minimum required work at room temperature. The performance tendency of each pressure cycle well represents the experimental results as shown in figure 5. The threshold temperature of 4.6 K obtained from the calculation of low pressure cycle is corresponded to the minimum temperature of the oscillating amplitude in figure
5. The discrepancy in the zero-cooling capacity temperature will be caused by the actual regenerator inefficiency resulting from the limited heat transfer surface and heat capacity of the regenerator material. This induces additional enthalpy flow, which is not included in the calculation.
To realize a cooling temperature below 4 K efficiently in the low pressure cycle, the minimum pressure of the cycle must be decreased to a sub-atmosphere level. In this case, the upper stage regenerator should be carefully designed to minimize the flow friction of the regenerator. In figure 7, a negative value of Hf/Wf means that the enthalpy flow is from the cold end of the regenerator to the hot end. It is to be noted that the increase in the negative enthalpy flow through the third regenerator implies an increase in the required cooling capacity of the second
Figure 6. Calculated cooling performance.
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Figure 7. Calculated enthalpy flow.
stage pre-cooler. This indicates that low pressure cycle is good for the following conditions: the
second stage regenerator has a high performance, or either the second stage pulse tube or precooler has a large cooling capacity. CONCLUSION
The possibility of increasing the cooling performance of 4 K pulse tube cryocooler was studied by reducing the working pressure. The experimental study indicates the superiority of the low pressure cycle in terms of cooling power compared to the medium or high pressure cycle for the cooling temperature above 5 K. To reduce this threshold temperature below 4 K, the minimum pressure of the cycle must be decreased below one atmosphere, which is not very suitable. The cooling temperature of 5 to 6 K, however, could be used in superconductor applications, where the cooler performance can be competitively used over superconductor performance. Further experimental studies to include the pre-cooler performance are required to demonstrate the superiority of the low pressure cycle. REFERENCES 1. Gao, J.L. and Matsubara, Y., “Experimental Investigation of 4 K Pulse Tube Refrigerator”, Cryogenics, vol. 34. no. 1 (1994), pp 25-30. 2. Matsubara, Y. and Gao, J.L., “Multi-staged Pulse Tube Refrigerator for Temperature below 4 K”, Cryocoolers 8, Plenum Press, New York (1995), pp 345-352. 3. Wang, C., Thummes, G. and Heiden, H., ”A Two-stage Pulse Tube Cooler operating below 4 K”, Cryogenics, vol. 37, no. 3 (1997), pp 159-164.
4. Chen, G., Qiu, L, Zheng, J., Yan, P., Gan, Z., Bai, X. and Huang, Z., ”Experimental Study on a Doubleorifice Two-stage Pulse Tube Refrigerator”, Cryogenics, vol. 37, no. 5 (1997), pp 271-273. 5. Wang, C., Thummes, G. and Heiden, H., ”Experimental Study of Staging Method for Two-stage Pulse Tube refrigerators for Liquid 4He Temperatures”, Cryogenics, vol. 37, no. 12 (1997), pp 857-863.
Research of Two-Stage Co-Axial Pulse Tube Coolers Driven by a Valveless Compressor L.W. Yang, J.T. Liang, Y. Zhou, and J.J. Wang Cryogenics Laboratory Chinese Academy of Sciences Beijing, 100080 China
ABSTRACT
Multi-stage pulse tube refrigerators driven by Stirling-type compressors have potential application in space and military. Here a 16 Hz, 55 cc rotary drive compressor was used to do twostage pulse tube cooler research. The largest input power was less than 700 W. Experiments and analyses included two kinds of pulse tube cooler structures: 1) totally co-axial, and 2) first stage co-axial and second stage U-type. For the totally co-axial structure, the lowest temperature reached was only 75.7 K. However, for the other structure (first-stage co-axial and second-stage U-type), a lowest temperature of 25 K was reached after optimization and with a radiation shield.
When no radiation shield was used for the second stage, the lowest temperature was 28K and the cooling capacity at 33K was 0.5 W. INTRODUCTION
Pulse tube refrigerators have the advantage of high reliability and low vibration, so they are
an excellent substitute for Stirling refrigerators and G-M refrigerators. Our multi-stage pulse tube refrigerator was mainly driven by a G-M type compressor, which
needs a rotary valve to control the frequency of the refrigerator. The G-M type pulse tube refrigerator, which is operated at a frequency of 1 to 2 Hz, has reached a refrigeration temperature below However, as compared to another multi-stage pulse tube cooler driven by a
Stirling-type compressor, the lowest temperature is still rather high; this structure has the advantage of compactness and high efficiency, and its frequency is generally higher than 10Hz. In 1991, Zhou et al. achieved a lowest temperature of 31 K with a two-stage pulse tube driven by a valveless compressor.3 In 1996, the U.S. Air Force Phillips Laboratory space cryocooler development4 shows their plan for two-stage and three-stage pulse tube coolers driven by Stirling-type compressors: 35K/60K for two-stage, and 10K/35K/80K for three stage. In 1997, in the SinoJapan Joint conference, Japan’s plan for multi-stage pulse tube coolers for space was shown together with some results. Their three-stage pulse tube cooler driven by a 94 cc linear compressor reached a lowest temperature of 20 K at a frequency of about Based on our 55 cc, 16 Hz valveless compressor, we have conducted research on multistage pulse tube coolers. The main test results are presented in this paper.
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Figure 1. Two kinds of two-stage pulse tube refrigerator structures.
CO-AXIAL TWO-STAGE PULSE TUBE REFRIGERATOR
There are several types of structures available for two-stage total co-axial pulse tube coolers. Figure 1 (A) is one kind of structure we used. The main feature is that the length of the second stage pulse tube is almost the same as the second stage regenerator and a small diameter long tube
is connected from the first stage cold tip to the ambient temperature orifice valve and reservoir. Because the first stage cold end is connected to ambient temperature through a small tube, heat transfer loss due to the gas stream in the small tube may be large. However, based on two considerations we adopted such a structure, studied the heat transfer loss, and concluded that an inertance tube may remedy the heat transfer loss. In fact, the refrigeration effect was not good. Figure 2 shows a typical set of cooldown curves. While the temperature of the second stage pulse tube drops to 74.5K, the coldtip temperature of first stage is 146 K; and on the other limit, when second stage two valves are closed, the final temperatures are 115.5 K for first stage coldtip and 122.8 K for second stage coldtip.
Figure 2. Typical change of two cold tip temperatures.
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Figure 3. Cooling down curves of first-stage pulse tube refrigerators.
This shows that such a co-axial structure makes the losses too large to drop the temperature further. In addtion, because of the co-axial structure, a change of dimensions is not easy. One thing worth mentioning is the effect of the orifice valve and double inlet valve on the refrigeration temperature. When the second stage orifice valve opens, the temperature of the second stage coldtip drops, but the first stage temperature rises; when the second-stage doubleinlet valve opens, the temperature of the first stage coldtip drops, but the second stage temperature changes little. Finally, the first stage double-inlet valve has little effect on coldtip temperature. PARTIALLY CO-AXIAL STRUCTURE
From the above experiment, we found two problems: 1) the performance of the first stage needs to be improved, and 2) connection between the first stage coldtip and ambient with an inertance tube has a negative effect. After analysis, we changed the dimensions of the pulse tube and regenerator, and the structure was changed to Figure 1(B). That is, a first stage co-axial structure, and a second stage U-type structure. The first pulse tube was optimized and its performance was improved greatly. Figure 3 shows the cooldown curve of the two kinds of first stage co-axial cooler: the newer one has a lowest temperature of 44 K, while the former has a lowest temperature of 95 K only when driven by 55cc compressor alone. The dimension of the new first stage pulse tube is 24 mm outside diameter for regenerator, 11 mm inside diameter for the pulse tube, and the length is 150 mm; the regenerator was filled with 250-mesh stainless steel screens. SECOND STAGE REGENERATOR WITH PB ONLY
Regenerator filling materials are very important to pulse tube performance. After analysis we adopted Pb only as the regenerator material for the second stage regenerator, because when the temperature is below 80K, Pb has a larger thermal capacity than steel or copper. In order to systematically study the effect of the regenerator, the length of the regenerator was studied. During the experiments, the temperature of the hot end of the regenerator was about 330 K, and the temperature of the first stage coldtip was about 90 K. Figure 4 shows how the second stage temperature changes with regenerator length. One key result is: the longer the second stage regenerator, the higher the temperature. This is a little different than general phenomena. The main reason for this is that the pressure wave in the second stage pulse tube is too low, its pressure ratio is less than 1.2; thus, most of the gas flow in the second stage pulse tube could not
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Figure 4. Effect of pressure and regenerator length on temperature.
reach the first stage coldtip. Another feature is the higher the average pressure, the lower the pressure. The lowest temperature with Pb only was 32 K. Another different phenomena between the structure of Figure 1(A) and Figure 1(B) is the function of the second-stage double-inlet valve. Different to (A), the second stage double inlet in Figure 1(B) could lower the temperature of first stage coldtip and second stage coldtip at the same time. The first stage coldtip temperature is 5 K lower generally, and second stage coldtip temperature is 20 K lower. SECOND STAGE REGENERATOR WITH STAINLESS STEEL AND PB
The above experiments show that with Pb as the regenerator material, the temperature was still rather high. After analysis, we found that stainless steel may help at the second stage. When part of the Pb in the 75mm long second stage regenerator was replaced by a 25mm length of 350mesh stainless steel screen, the temperature was further lowered. Figure 5 gives combined results showing that in 20 minutes, the first and second coldtips reached 85K and 28K, respectively.
Figure 5. Cooling down curve of 28K test.
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Figure 6. Temperature change with cooling capacity.
Cooling capacity was also measured, as shown in Figure 6, yielding 0.5W cooling capacity at 33K. A typical feature in the cooling capacity curve is that when second stage temperature increased, the first stage cold tip temperature dropped, though not very much. This is similar to other two-stage pulse tube results.2 The reason is that with cooling capacity added, the pressure ratio increased, the first stage refrigeration effect increased and at the same time the second stage regenerator loss decreased.
Generally, for the second stage, radiation losses may be large, but using the first stage cooling capacity to decrease the second loss is usually an effective method to lower second stage temperature further.2 Based on such a consideration, a radiation shield attached to the first stage coldtip was used; test results are shown as in Figure 7. The cooling rate was evidently slowed and the final temperatures were 25 K and 108 K, respectively, for second and first stage. CONCLUSION
Through tests and improvement, a lowest temperature of 25 K was achieved with a 16 Hz, 55cc valveless compressor. Tests show that the two-stage pulse tube cooler driven by a Stirling-
Figure 7. Cooling down curve of 25K test.
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type compressor is different from GM-type pulse tube coolers. The one key problem is the first stage pulse tube design; its performance may decrease considerably when coupled with the second stage. Another problem is the choice of regenerator materials for the second stage. Through research, we believe after further improvement, a multi-stage pulse tube driven by a Stirling-type compressor has the possibility to reach 10 K or even lower. ACKNOWLEDGMENT
The author gratefully acknowledges the support of the Superconductivity Center of China and K.C. Wang Education Foundation of Hong Kong. REFERENCES 1.
Matsubara, Y., Gao, J.L., “Novel configuration of three-stage pulse tube refrigerator for temperature below 4K”, Cryogenics, vol. 34, No. 4 (1994), pp. 259-262.
2.
Wang, C., Thummes, G. and Heiden, C., “A two stage pulse tube cooler operating below 4K”, Cryogenics, vol. 37, No. 2 (1997), p. 159.
3.
Zhou, Y., Zhu, W.X., Liang J.T., “Two-stage Pulse Tube Refrigerator”, Proceedings of the fifth International Cryocooler Conference, 1988, pp. 137-141. Bruning, J., Pilson, T., “Phillips Laboratory Space Cryocooler Development and Test Program,” Adv. in Cryogenics Engineering, vol. 43B (1998), Plenum Press, New York, pp. 1651-1660.
4. 5.
Ohtani, Y., Chandratilleke, G.R., et al., “Development of a three stage pulse tube refrigerator”, Proceedings of the fifth Japanese-Sino joint seminar on cryocooler and its applications (JSJS-5), 1997, pp. 118-122.
Experimental Investigation of a Unique Pulse Tube Expander Design C. S. Kirkconnell Raytheon Systems Company El Segundo, California, U.S.A., 90245
ABSTRACT The performance of a pulse tube expander in which the regenerator volume is distributed among three parallel tubes arranged symmetrically around the pulse tube has been experimentally investigated. This “4-tube” expander configuration, which was recently patented by Raytheon Systems Company (formerly Hughes Aircraft), is of interest because it has structural advantages over the more common U-tube design. The improved strength of the 4tube design permits reduced wall thickness, higher allowable side loads, and related design and system integration advantages. Furthermore, the 4-tube expander does not have the problem of conductive coupling between the regenerator and pulse tube, which is a source of lost refrigeration capacity in the traditional concentric configuration, a competing rigid expander design. Experiments were performed with and without the regenerator tubes linked by conductive straps, and the expander was shown to work more efficiently with the straps. The data reveal that the thermodynamic efficiency of the 4-tube expander is comparable to that of the more common pulse tube expander configurations. INTRODUCTION As pulse tube cryocoolers have become competitive with the more well-established Stirling cryocoolers in the realm of thermodynamic efficiency and have thus become a practical alternative in many applications, there has been an increased interest in pulse tube cryocooler implementation and system integration issues [1,2,3]. Allowable side load, residual vibration amplitude, and cold mount location are examples of common mechanical integration concerns. With respect to the latter, locating the cold region at one end of the expander, as in a Stirling or folded pulse tube, eases the design integration of the cooler into the rest of the system. Residual vibration control is typically accomplished through an active balancer in the compressor module, a balanced dual-piston compressor design, a split expander-compressor configuration connected via a transfer line, or a combination of these techniques. The maximum allowable side load concern has been traditionally addressed through the use of light-weight, flexible cold straps to minimize the mass which must be supported and the structural loads which are transmitted to the cold tip. The 4-tube pulse tube expander configuration specifically addresses two of these system integration issues, cold mount location and allowable side load. Like the concentric and U-tube designs which have been discussed previously [4,5,6], the 4-tube is a variation on the folded Cryocoolers 10, edited by R. G. Ross, Jr. Kluwer Academic/Plenum Publishers, 1999
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pulse tube theme with a distinct cold tip at one end of the expander. The rigidity of the 4-tube
expander is theoretically higher than that of a similarly-sized U-tube because of the symmetry of the design, and analysis is presently underway to substantiate and quantify that assertion. Though the concentric expander yields these same system integration advantages, the thermodynamic performance of the 4-tube is not hampered by radial conduction between the pulse tube and regenerator, as in the standard concentric. (A variation on the concentric design
in which the pulse tube and regenerator are separated by an annular vacuum has been discussed [7], and such a design is obviously not affected by pulse tube-to-regenerator radial conduction). Though the system-integration advantages of the 4-tube are potentially significant, certain aspects of the 4-tube expander design suggest its maximum cooling efficiency may be less than that of the U-tube. The concern stems from the 73% larger perimeter-to-cross sectional area ratio caused by arranging identical regenerator volumes as three parallel tubes versus a single tube of the same length. First, the total regenerator wall cross sectional area is increased proportionately, assuming identical wall thicknesses, so the relative size of the parasitic loss due to conduction through the regenerator walls is higher in the 4-tube. Also due to the larger perimeter-to-area ratio, the percentage of regenerator volume affected by possible boundary layer effects in the vicinity of the walls is greater in the 4-tube. Though the former can be accurately quantified analytically, the latter is not as yielding to simple analytic tools. For these reasons, experiments were performed to measure the performance of a 4-tube pulse tube expander. Experimental data were obtained for three slightly varying expander designs: a baseline design, an intermediate configuration with the three regenerator tubes thermally linked with conductive straps, and a final configuration which includes the thermal straps and an increase in the cross sectional area of some internal flow passages. EXPERIMENTAL SETUP
4-Tube Pulse Tube Expander Design
A side view of the patented Raytheon 4-tube expander is provided in Figure 1. The total regenerator volume is distributed among the three identical outside tubes, and the smaller center tube is the pulse tube component. The four identical-length stainless steel tubes are brazed into copper heat exchanger manifolds at either end. The large rectangular block is the warm end heat exchanger manifold, and it houses the inlet and rejection heat exchangers. The three inlet heat exchangers, which are located between the compressor and the regenerator in the flow circuit, are located at the warm ends of the regenerator ports. The single rejection heat exchanger is located at the warm end of the pulse tube. This particular expander is of the double-inlet configuration [8]; two external metering valves on the warm-end manifold (not shown in the figure) serve as the bypass and surge valves. The small manifold with the three fingers at the opposite end of the expander houses the cold, or load, heat exchanger. Three “fingers” link the regenerator tubes to the cold heat exchanger at the manifold centerline to allow the four tubes to shrink asymmetrically, a phenomenon which occurs due to uneven cooling, without inducing inordinately large stresses in the braze joints. Such structural compliance is important for any
folded tube expander because pulse tubes and regenerators have different temperature profiles [9, 10]. For this particular expander, this design feature is particularly important because the three regenerator tubes themselves also tend to develop differential temperature gradients, as is discussed herein. Other aspects of the design are provided in the patent [11]. Test Apparatus and Instrumentation
The test apparatus consists of the 4-tube expander and a convenient wall-powered laboratory compressor. The compressor, which has been used in other similar studies, is a fixed swept volume (20 cc), fixed frequency (21 Hz) rotary device [12]. The compressor and expander were connected with a seven centimeter transfer line. The expander was specifically sized for use
INVESTIGATION OF A UNIQUE PT EXPANDER DESIGN
241
Figure 1. Side view of a 4-tube pulse tube expander.
with this compressor, so the performance data provided herein are a fair assessment of the
expander’s capabilities. Figure 2 illustrates the cryocooler assembly and associated instrumentation. The rejection temperature at the warm end heat exchanger manifold was maintained at 291 K using a recirculating glycol-water bath for all of these experiments. The cold end heat exchanger is temperature-controlled with a LakeShore controller, DT470 temperature diodes, and an resistive heater. The pressure waves in the transfer line, the surge tank, and on the pulse tube side of the surge valve are measured using high-rate Paine strain gauges and amplifiers. A
piezoelectric differential pressure transducer in the surge tank provides the amplitude of the small surge volume pressure fluctuations, a direct measurement which facilitates an indirect measurement of the gross refrigeration power produced within the expander. The phase of the compressor piston stroke is provided by a proximity sensor. A cam is mounted to the drive shaft such that it is identically in phase with the drive piston, and the proximity sensor measures the position of the cam, hence providing the piston phase angle. The vacuum environment is maintained by a Leybold vacuum system. Temperatures are measured at the locations shown using welded type-T thermocouples. Note that the temperature of the three regenerator tubes is measured at an identical axial location (1.0” from cold end) to provide a means of evaluating the thermal balance of the regenerator during operation. EXPERIMENTAL RESULTS
Baseline Expander Design (No Copper Straps)
The initial experiments were performed with the expander as shown in Figure 1, i.e., without copper straps thermally linking the tubes together. A variety of mean operating pressures and valve settings were investigated in an effort to identify the optimum operating condition, defined
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PULSE TUBE CRYOCOOLER CONFIGURATION INVESTIGATIONS
Figure 2. Experimental test setup and instrumentation. P = absolute pressure, dP = differential pressure, T = temperature, LS = LakeShore diode, TC = type-T thermocouple, Q = electrical heat load. See fig. 1
for actual expander geometry.
as that combination of settings which yielded the maximum net refrigeration at 70 K. The piston swept volume and operating frequency are fixed by the design of the compressor, so these parameters were obviously not part of the optimization process. A mean operating pressure of 25.4 atm and a 5.5:1 ratio between the bypass valve and surge valve flow coefficients were
identified as the optimum operating parameters. This same operating point was used for the other two configurations to maintain a relevant basis for comparison between the data sets. The net refrigeration and input PV power versus temperature for the optimum operating
condition are shown in Figure 3. (The input PV power is estimated from the measured values of the mean pressure, pressure ratio, and piston phase angle. See reference [12] for details.) The cryocooler yielded 1.5 W net refrigeration at 70 K for 125.4 W input PV power, yielding a specific power (SP; input PV power/net refrigeration) of 83.6. This performance fell well short of the 3.2 W and SP of 41.2 which had been projected for this cold tip temperature from the numerical model (model described elsewhere [13]). A likely explanation for the shortfall was revealed by the regenerator thermocouples, located at an identical axial location on each tube (1.0” from the edge of the cold-end copper manifold). Figure 4 shows the temperatures on the three regenerator tubes for each of the load points from Figure 3. Note the large temperature variations between the three tubes for each operating point, the difference between the maximum and minimum temperatures ranging from 73.1 K at no load to 62.5 K for a 90 K cold tip. This operating condition was neither predicted by nor accounted for in the numerical model. Given the identical size, environmental loading, and heat exchanger interfaces for the three tubes, the most plausible explanation is an uneven flow distribution between the three tubes due to subtle differences in tube diameter and/or packing density. The next section describes an effort to mitigate this impact of this effect. Intermediate Test Results (With Copper Straps)
In an effort to balance out the regenerator temperature distribution, copper straps were added at four locations to thermally link the three tubes. An improvement in the balance of the flow distribution and a reduction in the regenerator parasitic losses were anticipated with an improved
INVESTIGATION OF A UNIQUE PT EXPANDER DESIGN
243
Figure 3. Comparison of experimental and numerically-predicted load and power curves for baseline expander design (no thermal straps). ‘exp’ = experimental; ‘num’ = numerical.
Figure 4. Regenerator temperatures versus cold tip temperature for baseline expander configuration. Cold tip temperatures correspond to experimental load points from Fig.
at no load (61.9 K).
thermal balance. The exact locations of the straps are shown in Figure 5. Each strap was constructed of 3 layers of 9/32 inch wide, 10 mil thick OFHC copper strips with the layers bonded to each other and to the regenerator tubes with “Scotch” 966 “Hi-Temperature” acrylic laminating adhesive. A higher conductivity adhesive could possibly have been used to obtain better performing thermal straps, but the results obtained were impressive, nevertheless. The reduction in temperature differentials between the regenerator tubes is demonstrated in Figure 6. The maximum temperature delta between tubes was reduced from 73.1 K to 17.2 K, and this translated into improved cryocooler performance. The no-load cold tip temperature was reduced by 3.3 K to 58.6 K, and the load capacity at 70 K was increased from 1.5 W to 2.3 W. Furthermore, the improved flow balance reduced the overall system pressure drop, so the input
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PULSE TUBE CRYOCOOLER CONFIGURATION INVESTIGATIONS
Figure 5. Location of regenerator thermal straps. Each strap constructed of (3) layers of 10 mil copper, bonded to regenerator tubes and bonded layer-to-layer. Straps located at identical axial location on all three regenerator tubes. Dimensions given in inches.
PV power dropped to 117 W, yielding a specific power of 50.9 at 70 K. Though still short of the design goal, this performance marked a significant improvement beyond that which had been
obtained without the straps. Final Test Results (With Copper Straps and Reduced Flow Restrictions)
As described above, the addition of the thermal straps reduced the thermal imbalance between the regenerator tubes and improved the cooling capacity and efficiency of the expander. With the link between a more balanced regenerator and improved performance having been established, an additional step was taken to further reduce this imbalance. Consider the initial cool down of the refrigerator. At time zero, the entire expander is in
thermal equilibrium, including each of the three regenerator tubes. Immediately upon starting the compressor, however, the temperatures between the tubes begin to diverge. Evidently,
Figure 6. Comparison of regenerator temperatures with and without thermal straps. (1) No-load with thermal straps, (2) No-load without thermal straps. Note the significant decrease in temperature differences between the regenerator tubes due to the addition of the straps.
INVESTIGATION OF A UNIQUE PT EXPANDER DESIGN
245
slightly more refrigeration is produced in the cold tip region of the more efficient tube(s). This
results in a faster localized cooling of the gas, which in turn causes the density of the gas in that particular tube to decrease more rapidly, which results in a lower pressure drop and an even larger percentage of the flow being shuttled through the more efficient tube. This leads to an even more rapid cooling of this more efficient tube in relation to the other tubes, hence the temperature differentials develop and persist. Given this relationship between the fluid dynamic and thermal imbalances, it was hypothesized that a reduction in the regenerator pressure drops would reduce the magnitude and impact of any flow imbalances, hence retarding the development of the regenerator thermal mismatch. Towards this end, the regenerator was disassembled and re-packed with a slightly less restrictive screen pack design. The effects of this change were modest. Figure 7 shows that the reduced pressure drops in the regenerator did indeed result in a further decrease of the regenerator thermal imbalance, with the maximum temperature differential at no load falling to 7.2 K. However, as revealed by the load curve provided in Figure 8, the cooling capacity was unchanged (2.3 W @ 70 K) from the previous configuration, and the input power was only slightly less (110 W vs. 117 W @ 70 K). The substantial improvements in expander performance versus the initial data set (compare Figures 3 and 8) must therefore be attributed to the addition of the thermal straps, not the modification of the internal regenerator design. DISCUSSION AND CONCLUSION
A pulse tube expander in which the regenerator volume is divided between three parallel tubes has been constructed and tested. After the addition of thermal straps, which yielded a substantial improvement, and the implementation of a slightly less flow restrictive regenerator design, which provided only a modest increase in thermodynamic efficiency, the performance of
this 4-Tube expander was shown to be comparable to that of the more conventional pulse tube designs. For example, a U-Tube expander of similar size described in a previous paper [12]
provided 4.5 W @ 80 K for 117 W input PV power while the 4-Tube yielded 4.3 W @ 80 K for 107 W With respect to the modeling projection of at 70 K, the SP of 47.6 achieved for the final configuration at 70 K is in reasonably sound agreement.
Figure 7. Regenerator temperatures versus cold tip temperature for final expander configuration. Note a further reduction in regenerator tube temperature deltas due to reduced pressure drops. See Fig. 5 for
comparison.
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PULSE TUBE CRYOCOOLER CONFIGURATION INVESTIGATIONS
Figure 8. Load curves for improved expander design (thermal straps and reduced pressure drops). The significant improvement versus the baseline performance attributable primarily to thermal straps.
By virtue of simple thermodynamics, the observed increase in cooling capacity in going from
the baseline to final configurations must be due to an increase in gross refrigeration produced, a reduction in parasitic losses (conduction, regenerator inefficiency, etc.), or a combination of the two. The relative performance of the initial and final expander configurations was evaluated to characterize the individual contribution of these two effects. The 70 K refrigeration data are
provided in Table 1 as a typical example of the comparative measurements taken over the temperature range investigated (58 K to 90 K). Note that the gross refrigeration is essentially the
same for both cases, with the strapped expander actually yielding slightly less gross capacity. Since the gross refrigeration value is a derived measurement, it is desirable to support the validity of the values reported with direct measurements. The pressure ratio measurement (Pmax/Pmin) before the surge valve provides that support. Since the pressure drops in the pulse tube and rejection heat exchanger are very small, the pressure ratio measured at the valve is representative of the pressure ratio in the cold expansion space. The data reveal that the amplitude of the pressure wave in the expansion space is the same for both configurations, which supports the assertion that the gross refrigeration is about the same. Therefore, the improved cooling capacity is evidently due almost entirely to a reduction in the parasitic losses. As a final note, the slight reduction in the amplitude of the compressor-side pressure ratio for the final configuration is consistent with the reported decrease in required input power discussed previously.
INVESTIGATION OF A UNIQUE PT EXPANDER DESIGN
247
The data obtained indicate that the 4-tube configuration is a viable part of the trade space available to Raytheon in considering pulse tube designs for particular systems, especially those applications requiring a robust expander design with a high allowable side load. On a more general note, the results reveal the difficulty in achieving an even, predictable flow distribution in parallel hydrodynamic systems, particularly those systems in which significant heating or cooling is taking place in localized regions, and how the flow distribution can be somewhat controlled by adding external thermal connections between the parallel circuits. ACKNOWLEDGMENTS
The author would like to acknowledge the support and assistance of the 4-tube expander coinventors: Alan Rattray (lead inventor), Ken Price, Steve Soloski, and Sam Russo. REFERENCES 1.
Duband, L., A. Ravex, T. Bradshaw, A. Orlowska, C. Jewell, and B. Jones, “50-80 K Pulse Tube Cryocooler Development”, Proc. of the 9th Intl. Cryocooler Conf., Cryocoolers 9, Plenum Press, New York (1997), p. 213.
2.
Russo, S. C. and G. R. Pruitt, “Development of a Low-Cost Cryocooler for HTS Applications”, Proc. of the 9th Intl. Cryocooler Conf., Cryocoolers 9, Plenum Press, New York (1997), p. 229.
3.
Burt, W. W. and C. K. Chan, “New Mid-Size High Efficiency Pulse Tube Coolers”, Proc. of the 9th Intl. Cryocooler Conference, Cryocoolers 9, Plenum Press, New York (1997), p. 173.
4.
Curlier, P., “Cryocooler Technologies: 2. Pulse Tube Miniature Cryocoolers”, Proc. of SPIE, vol. 2552, SPIE, Berlin, Germany (1995), p. 791.
5.
Liang, J., A. Ravex, and P. Rolland, “Study on Pulse Tube Refrigeration, Part 3: Experimental Verification”, Cryogenics, vol. 36, Elsevier Science, Ltd., Oxford, UK (1996), p. 101.
6.
Richardson, “Development of a Practical Pulse Tube Refrigerator: Co-Axial Designs and the Influence of Viscosity”, Cryogenics, vol. 28, Elsevier Science, Ltd., Oxford, UK (1988), p. 516.
7.
Rattray, A., S. C. Soloski, and F. N. Mastrup, “Concentric Pulse Tube Expander with Vacuum Insulator”, U.S. Patent Number 5680768 (1997).
8.
Zhu, S. W., P. Y. Wu, and Z. Q. Chen, “Double Inlet Pulse Tube Refrigeration: An Important Improvement”, Cryogenics, vol. 30, Elsevier Science, Ltd., Oxford, UK (1990), p. 514.
9.
Kirkconnell, C. S. and G. T. Colwell, “Parametric Studies on a Numerical, Nonlinear Pulse Tube Flow”, Journal of Fluids Engineering, vol. 119, no. 4, American Society of Mechanical Engineers,
10.
New York, (1997), p. 831. Bauwens, Luc, “Near-Isothermal Regenerator: A Perturbation Analysis”, Journal of Thermophysics and Heat Transfer, vol. 9, no. 4, ASME, New York, (1995), p. 749.
11.
Rattray, A., C. S. Kirkconnell, K. D. Price, S. C. Soloski, and S. C. Russo, “Cooling System Using a Pulse-Tube Expander”, U.S. Patent Number 5647219 (1997).
12.
Kirkconnell, C. S., “Experiments on the Thermodynamic Performance of a ‘U-Tube’ Pulse Tube Expander”, accepted for publication in the Proc. of the Cryogenic Engineering Conference (1997).
13.
Kirkconnell, C. S., “Numerical Analysis of the Mass Flow and Thermal Behavior in HighFrequency Pulse Tubes,” Ph. D. thesis, Georgia Institute of Technology, Atlanta, GA (1995).
An Experimental Study on the Heat Transfer Characteristics of the Heat Exchangers in the Basic Pulse Tube Refrigerator
S. Jeong and K. Nam
Cryogenic Engineering Laboratory Korea Advanced Institute of Science and Technology Taejon, Korea 305-701
ABSTRACT
A basic pulse tube refrigerator has been constructed with extensive instrumentation to study the heat transfer characteristics of the heat exchanger experimentally under the oscillating
pressure and the oscillating flow. The extensive instrumentation includes the pressure transducers across the regenerator, several wall temperature sensors, the gas temperature sensors and the heat flux gages to measure the instantaneous heat transfer rate in the heat exchangers. This paper describes some sequential experiments with the basic pulse tube refrigerator. The experiments were performed for various cycle frequencies under the square pressure waveforms. First, the heat flux was measured through the cycle at the both cold and warm end heat
exchangers without the regenerator. The experimental results were compared to the previous theoretical analysis of the oscillating pressure and the oscillating flow. Second, the regenerator
was added to the pulse tube to make a basic pulse tube refrigerator configuration. The experiment showed the great impact of the regenerator on the temperature and the heat flux profiles. The performance of the pulse tube refrigerator is sensitive to the oscillating heat transfer phenomena. In order to enhance the thermal communication capability of the heat exchanger with the gas at low operating frequencies, a unique design of the triangular shape radial fin concept was applied to the heat exchangers. The thermal contact resistance was virtually eliminated. This paper presents the performance comparison of the new fin type heat exchanger and the conventional screen-type one. INTRODUCTION
The pulse tube refrigerator (PTR) has an advantage over Stirling and Gifford-McMahon (GM) cryocoolers that it has no moving part at the cold end, which results in mechanical simplicity, small vibration and long term reliability. For this reason, pulse tube refrigerator has been developed extensively for adapting it to the infra-red sensor of the space application or for
Cryocoolers 10, edited by R. G. Ross, Jr. Kluwer Academic/Plenum Publishers, 1999
249
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PULSE TUBE CRYOCOOLER CONFIGURATION INVESTIGATIONS
Figure 1. Schematic diagram of the experimental apparatus. ( : wall temperature sensor, : gas temperature sensor, P : pressure sensor, Q : heat flux sensor).
cooling electronic devices with high temperature superconductor (HTS).1 Researches for the pulse tube refrigerator have been mainly focused on the performance enhancement or theoretical
analysis of thermodynamic behavior of the working gas. The refrigeration mechanism in the basic pulse tube refrigerator could be explained by the ‘surface heat pumping effect’2 that has been since Gifford and Longsworth who first conceived the basic pulse tube refrigerator. The heat transfer analysis under oscillating flow and oscillating pressure has been well developed for the Stirling machine, the experimental approaches for the heat transfer characteristics of the pulse tube refrigerator, however, have not been performed so well.3,4 This paper describes the experiments of the real time measurement of the heat flux and the gas temperature at the heat exchangers of the pulse tube refrigerator under the input pressure of the square waveforms. Two different types of the heat exchangers were devised to investigate their heat transfer characteristics under oscillating pressure environment in the pulse tube refrigerator. The instantaneously measured experimental data of the heat flux in the heat exchanger were also compared with the theoretical predictions that had been developed for the heat transfer condition under oscillating pressure. EXPERIMENTAL CONFIGURATIONS
An experimental apparatus was constructed to study the heat transfer characteristics in the basic pulse tube refrigerator. A schematic diagram of the experimental setup is shown in Fig. 1.
A helium compressor (CTI-cryogenics model 8200) supplies helium gas as working fluid. The rotary valve system provides the basic pulse tube refrigerator with oscillating gas pressure and flow. The frequency of the oscillating pressure was adjusted by the rotating speed of the rotary valve, which was controlled by the stepper motor and the function generator. The basic pulse
STUDY OF HEAT EXCHANGERS IN BASIC PT REFRIGERATOR
Figure 2. Schematic diagram of the heat exchangers and the sensor installation. (unit: mm) (a) Mesh heat exchanger (b) Fin heat exchanger ( : gas temperature sensor, Q : heat flux sensor).
tube refrigerator was installed in the vacuum chamber (not shown in the figure) to minimize the heat load due to convection heat transfer. The regenerator and the pulse tube were made of stainless steel tubes with outer diameters of 37.5 mm and 19 mm and lengths of 100 mm and 200 mm, respectively. The regenerative matrix in the regenerator consisted of 1300 stainless steel screens of 400 mesh. A stainless steel tube of 0.25 mm wall thickness was used for the pulse tube to reduce the axial conduction heat transfer along the wall. The cold and warm end heat exchangers were brazed to the pulse tube as shown in Fig. 1. Fig. 2 illustrates the configurations of two different heat exchangers. Fig. 2(a) is a conventional mesh heat exchanger. For this configuration, 5 copper screens of 100 mesh were inserted in the cold end and 40 screens in the warm end. Fig. 2(b) is a radial fin heat exchanger. It was fabricated by EDM (Electric Discharge Machining) to arrange 24 internal fins radially.5 The center pole was inserted afterward to eliminate unuseful void space of the heat exchanger. The heat flux at the heat exchanger wall was measured by the heat flux gauge (RdF model 20455-2). It was a micro-foil type gauge and its time constant for the step input was 60 ms according to the manufacturer. Due to the difficulty of the sensor mounting and the leakage problem, the heat flux sensor was installed at the outside wall of the heat exchanger as shown in Fig. 2. The heat flux signal was calibrated to have a positive value in the case of the net heat transfer toward the gas and a negative value for the heat transfer from the gas. A fast response thermocouple (Type E) was also inserted in the heat
exchangers. Its diameter was about 0.05 mm and the detail of the thermocouple instrumentation
was presented in the previous paper.5 The strain gauge type pressure transducer was installed at the inlet of the pulse tube to monitor the oscillating pressure. All the experimental data were stored in the computer for the analysis when the cyclic steady state was reached. EXPERIMENTAL RESULTS AND DISCUSSIONS
Heat transfer analysis of the pulse tube without regenerator
Under the oscillating flow and the oscillating pressure, it has been known that there is a phase difference between the gas temperature and the heat flux.5 Apart from Newton’s law of cooling for the steady state condition, Kornhauser proposed a new heat transfer correlation as follows6:
251
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PULSE TUBE CRYOCOOLER CONFIGURATION INVESTIGATIONS
where
k f
: thermal conductivity : hydraulic diameter : oscillating frequency : real part of the complex Nusselt number : imaginary part of the complex Nusselt number
The complex Nusselt number in Eq. (1) was determined by using the least square method for the experimental data.5 On the other hand, Jeong has derived the following correlation from the twodimensional energy equation for the case of the fluid flow between two flat plates.7 The heat flux at the wall was expressed as:
where
2H : distance between two flat plates R : gas constant P : pressure of the working gas k : thermal conductivity of the working gas The fundamental difference of Eq. (2) from Eq. (1) is that the heat flux term associated with the
oscillation is derived from the direct instantaneous pressure change. Although the temperature gradient term in Eq. (1) is replaced by the pressure gradient term in Eq. (2), two equations describe virtually the same physical phenomenon because the temperature is varied by the pressure change. To compare the experimental data and the theoretical analysis, the heat flux at the inside wall of the heat exchanger was estimated quantitatively from the measured heat flux data at the outside wall. The heat exchanger was assumed as the one-dimensional plate, then the heat flux at the inside wall could be estimated by the following equation8:
where
L
: effective thickness of the heat exchanger : thermal diffusivity of the heat exchanger material
: specific heat of the heat exchanger material : temperature at the outside wall of the heat exchanger
: heat flux at the inside wall : heat flux at the outside wall
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253
The heat flux was calculated by Eq. (1) and Eq. (2) using the measured gas temperature, the wall temperature and the oscillating pressure. The predicted heat flux calculation was compared to the experimental data for the fin heat exchanger at the cold end as shown in Fig. 3. Both correlations give more accurate predictions at low oscillation frequency. During a cycle of the trapezoidal pressure wave, the calculated heat fluxes are negligible at low-pressure period. As the frequency
grows, however, both correlations start to fail in predicting the heat flux at low-pressure period. The gas expands as the pressure drops and the heat flux from the outside toward the expanded cold gas continuously existed even during the low-pressure period. This tendency is more evident at high oscillation frequency case. The experiment of Fig. 3 was performed without regenerator. Therefore, the outward (negative) heat flux exists during the high-pressure period at the cold end heat exchanger. The complex Nusselt number used in Eq. (1) is a cyclic averaged value. Since it is constant during the cycle, an excessively undamped oscillation of the heat flux
is calculated near the sudden pressure rise or fall due to the effect of the slope change in the gas temperature. The applicability of Eq. (1) is actually uncertain in the case of such a sudden pressure change. Eq. (2), on the other hand, has the coefficients, and that are functions of the oscillating pressure. They are similar to the complex Nusselt number in Eq. (1). and are instantaneously changing over one cycle. Fig. 3 shows that the heat flux calculated by Eq. (2) follows the trends of the experimental data more accurately in the period of high-pressure gradient than the heat flux by Eq. (1). Fig. 4 shows the experimental data for the mesh heat exchanger. Different from the fin heat exchanger, the heat flux gauge at the outside wall of the cold end did not respond quickly to the gas temperature change. The measured span of the heat flux value was considerably smaller than
Figure 3. Comparison of the measured heat flux and the theoretical predictions of the fin heat
exchanger at the cold end. (a) 0.2 Hz (b) 0.5 Hz (c) 0.7 Hz (d) 1.0 Hz.
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PULSE TUBE CRYOCOOLER CONFIGURATION INVESTIGATIONS
Fig. 4 Temperature, heat flux and pressure data of the mesh heat exchanger at the cold end. (a) 0.5 Hz (b) 0.7 Hz (c) 1.0 Hz (d) 2.0 Hz.
the previous case. This was due to the radial thermal contact resistance of the mesh heat exchanger. As the frequency became high, the net heat flux changed its sign due to the increase of the compression work and the gas temperature. It was evident that the mesh heat exchanger was underdesigned for this pulse tube. The heat flux data for the different types of heat exchangers were very distinctive. The performance of the heat exchanger, therefore, could be evaluated from this heat flux data.
Heat transfer analysis of the pulse tube with regenerator The pulse tube described in this section was accompanied by the regenerator so that they became a basic pulse tube refrigerator. The cooling performance, however, was very poor due to the significant loss by excessive instrumentation and the dead volume between the regenerator and the pulse tube. Table 1 shows the cyclic mean heat transfer rate at the various frequencies for the case of the mesh heat exchanger. was the heat transfer rate at the aftercooler of the regenerator. was calculated from the flow rate, the inlet and outlet temperature of the cooling water. and were the heat transfer rate at the warm end and the cold end heat exchangers,
respectively. These values were obtained from the integration of the heat flux data. Table 1 shows that the heat transfer rate at the aftercooler increased as the cyclic frequency became high. This result was due to the increase of the flow friction at the regenerator, which was evident in the oscillating pressure condition. The heat load always existed at the cold end in spite of the no
STUDY OF HEAT EXCHANGERS IN BASIC PT REFRIGERATOR
255
load condition. This result actually represented the direct measurement of the parasitic heat losses at the cold end. The parasitic heat losses consisted of the radiation loss, the convection loss, the axial conduction loss along the pulse tube wall and the conduction loss through the instrumentation wires. The axial conduction loss along the pulse tube wall was considered as the major heat loss. Table 1 also shows that the heat transfer rate at the warm end heat exchanger had its maximum value near the frequency of 2 Hz. This phenomenon can be explained by the experimental data at the warm end as shown in Fig. 5. The performance of the basic pulse tube refrigerator is determined by the compression ratio and the heat transfer capability at the heat
Figure 5. Temperature, heat flux and pressure data of the mesh heat exchanger at the warm end. (a) 0.5 Hz (b) 2 Hz (c) 3 Hz (d) 5 Hz.
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PULSE TUBE CRYOCOOLER CONFIGURATION INVESTIGATIONS
exchanger. For frequencies lower than 2 Hz, the compression ratio was high enough to compress the gas in the pulse tube effectively, thus increasing the gas temperature at the warm end. On the other hand, the heat flux was not large due to the slow motion of the working gas during the cycle. In the case of frequencies higher than 2 Hz, the heat transfer coefficient could increase due
to the fast oscillating flow, but the compression ratio was essentially reduced due to the increased flow friction and the heat transfer time was also reduced. As a result, the difference between the gas temperature and the wall temperature was relatively small. The heat rejection to the outside was not efficient so that the whole refrigerator performance became poor. The minimum cold end temperature around 235 K was obtained at 2 Hz. CONCLUSIONS
(1) The heat flux of the pulse tube was measured instantaneously. The phase difference between the gas temperature and the heat flux was clearly observed. There was a significant difference
in the heat flux data between the two types of heat exchangers. This phenomenon was mainly due to the thermal contact resistance of the heat exchanger.
(2) The heat flux data at the cold end were compared with the theoretical heat transfer relation of the oscillating flow and the oscillating pressure in the fin heat exchanger. Even though the heat flux tendency during the cycle could be predicted reasonably well by the previous theoretical analysis, further heat transfer analysis would be necessary to produce more accurate results. (3) The amount of the heat rejection at the warm end had a maximum at the specific operating frequency for the basic pulse tube refrigerator. This experimental result could be explained by the measured data at the warm end heat exchanger. ACKNOWLEDGEMENT
This research was supported in part by the Korea Science and Engineering Foundation under the contract no. 981-1006-038-2. REFERENCES
1.
2.
Thummes, G., Landgraf, R., Giebeler, F., Mück, M. and Heiden, C., "Pulse tube refrigerator for high-Tc SQUID operation", Advances in Cryogenic Engineering, Vol. 41B, Plenum Press, New York (1995), pp. 1463~1470. Gifford, W.E. and Longsworth, R.C., "Surface heat pumping", Advances in Cryogenic Engineering,
Vol. 11, Plenum Press, New York (1966), pp. 171~179.
3.
Shiraishi, M., Seo, K. and Murakami, M., "Pressure and temperature oscillations of working gas in a pulse tube refrigerator", Cryocoolers 8, Plenum Press, New York (1995), pp. 403~410.
4.
Seo, K., Shiraishi, M., Nakamura, N. and Murakami, M., "Investigation of radial temperature and velocity profiles in oscillating flows inside a pulse tube refrigerator", in: Cryocoolers 9, R.G. Ross, Jr. Ed., Plenum Press, New York (1997), pp. 365~374.
5.
Jeong, S. and Nam, K., "Investigation on the heat transfer characteristics in pulse tube refrigerator", presented at the Cryogenic Engineering Conference at Portland, Oregon, USA (1997) paper APG5. Kornhauser, A. A., "Gas-wall heat transfer during compression and expansion", Sc.D. Thesis, Dept. of Mech. Eng., Massachusetts Institute of Technology, Cambridge, MA (1989).
6. 7.
Jeong, E.S. and Smith, Jr., J.L., "An analytic model of heat transfer with oscillating pressure",
General papers in Heat Transfer, ASME, Vol. 204 (1992), pp. 97~104.
8.
Burggraf, O. R., "An exact solution of the inverse problem in heat conduction theory and applications," Journal of heat transfer, Trans. ASME, Vol. 86(1964), pp. 373-382.
Double Vortex Tube as Heat Exchanger and Flow Impedance for a Pulse Tube Refrigerator Matthew P. Mitchell1, Drazen Fabris2 and Lt. B. J. Tomlinson3
1. Mitchell/Stirling, Berkeley, California 2. Illinois Institute of Technology, Chicago, Illinois 3. Air Force Research Laboratory, Kirtland AFB, New Mexico
ABSTRACT
Orifice pulse tube refrigerators reject heat and alter phase through fluid flows at the warm end of the pulse tube. Work otherwise wasted in the orifice can be put to use in a novel device based upon the principle of the Ranque/Hilsch vortex tube. By separating flows in both directions into hot and cold streams, and rejecting heat from the hot streams, enhanced heat rejection can be accomplished. A shell of hot fluid rotating rapidly in the bore of the vortex tube ensures effective heat exchange from each hot stream. Impedance of a double vortex tube controls flows as would an orifice. Reduction in fluid temperature at the warm end of the pulse tube reduces conduction losses and improves overall efficiency of the system. Preliminary experiments indicate that the double vortex tube works. Some issues in optimization of the double vortex tube as a component of a pulse tube refrigerator system are discussed. INTRODUCTION
Orifice pulse tube coolers, shown schematically in Figure 1, operate on familiar principles. The flow restriction of the orifice establishes a phase difference between the pressure wave and the flow direction. Unlike a Stirling cooler, a pulse tube cooler does not use the work of expansion effectively; the work at the warm end of the pulse tube simply compresses the fluid in the reservoir. Because the regenerator thermally isolates the pulse tube, orifice, and reservoir from the compressor and aftercooler, the pulse tube, orifice, and reservoir act as a mechanically forced but thermally isolated cycle. In order to achieve an energy balance in the portion of the system outboard of the regenerator from the compressor, heat recovered at the cold heat exchanger must be rejected at the warm heat exchanger. (1) A dynamic heat exchanger, driven by the work that would otherwise be dissipated in the orifice, can enhance the heat transfer at the warm end of the pulse tube, lower the temperature at which heat is rejected there, and thereby enhance cooling performance.
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PULSE TUBE CRYOCOOLER CONFIGURATION INVESTIGATIONS
Figure 1. Orifice Pulse Tube Cooler.
PRINCIPLE OF THE RANQUE/HILSCH VORTEX TUBE
The Ranque/Hilsch vortex tube is used in a variety of commercial applications where spot cooling is required and a cheap source of compressed gas is available. Several vendors offer similar products with similar applications (2, 3). All are open systems. In essence, the vortex tube uses the pressure drop between source pressure and ambient pressure to split a fluid stream into two separate streams, one hot and one cold. A commercial Ranque/Hilsch vortex tube is shown
schematically in Figure 2. (2) The separation of a flow into hot and cold streams is accomplished by generating a vortex in the fluid in the tube. Inflow is through one or more jets that enter one end of the vortex tube tangent to its bore. Typically, "vortex generators" containing multiple tangential inlet jets are
employed. The warm flow exits from the periphery of the vortex tube at the far end. The cold flow exits through a small opening at the axial center of the tube immediately adjacent to the inlet jets. That opening can be be insulated to reduce conduction between the body of the vortex tube and the cold fluid. The usual working gas is clean, dry compressed air; helium also works well. The relative flows at the hot and cold ends of a vortex tube can be adjusted by controlling the size of the apertures at one end or both. Commercial manufacturers of open-system, airdriven vortex tubes publish tables of temperature rise and fall for various "cold fractions" at various input pressures. Those tables assume discharge to atmospheric pressure. With air entering the vortex tube at 0.8 Mpa, flow can be adjusted to create a cold flow 70 C below or 105 C above inlet temperature. With equal hot and cold flow fractions, a temperature difference of 108 C between hot and cold streams is attainable. With an 80% cold fraction, the temperature spread is 137 C. (3) If the hot and cold streams were to be insulated and recombined as shown in Figure 3, the First Law demands that the result be no different from the result in simply venting the pressurized gas to the ambient air. However, if heat were to be removed from one of the streams before they
Figure 2. Ranque/Hilsch Vortex Tube.
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Figure 3. Recombined Flows From Vortex Tube.
were recombined, the effect would be to lower the total internal energy and the temperature of the resulting combined stream. The inherent nature of the vortex tube aids heat rejection to its wall. Since the hot stream rotates at the periphery, it is in contact with a large surface area, which promotes heat transfer. That attribute of the vortex tube is advantageous for the pulse tube application. PRINCIPLE OF THE DOUBLE VORTEX TUBE
An important issue in applying a vortex tube to a pulse tube cooler is the reversal of the flow direction over the course of a cycle. A novel double-acting vortex tube is shown in Figure 4. It provides the same flow features regardless of the direction of the fluid flow. One inlet/outlet
tube is connected to the warm end of a pulse tube and the other inlet/outlet is connected to the reservoir. In each direction of flow, the flow splits. The cold stream is insulated, and the hot stream is in contact with the heat sink through the body of the vortex tube and the output path. Thus, heat is removed from the hot stream before it is recombined with the cold stream. The effect is that the temperature of the combined stream is lower than the temperature that would have been achieved had all of the flow rejected heat to the heat sink from the incoming fluid temperature, and potentially lower than the temperature of the heat sink itself. Flow through the double vortex tube involves some subtleties. As shown in Fig. 5, flow entering the bore of the vortex tube has two paths available (B and C), but moves preferentially along B. Fluid enters the bore of the vortex tube tangentially, setting up the vortex. The exiting flow leaves through both the tangential passage connected to path D and the axial port connected to path E. Return flows follow the same pattern in reverse direction.
Figure 4. Schematic Double Vortex Tube.
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Figure 5. Flow Paths in Double Vortex Tube.
The hot fluid is always in contact with the vortex tube wall and moves along a separate path from the cold fluid whether the flow comes from or goes into the pulse tube. The result is that heat can be rejected when fluid is moving from the reservoir back to the pulse tube, whereas, in a conventional orifice pulse tube cooler, the drop in pressure during that phase of the cycle cools the fluid temperature to below the temperature of the heat exchanger, thereby limiting the total heat transfer over the cycle. An appropriately-designed double vortex tube of appropriate size also offers substantial resistance to the flow. Since the flow features are identical regardless of flow direction, the double vortex tube can fulfill the function of an orifice under reversing flow. The effectiveness of a vortex tube relates to the total momentum and mass flux of the working fluid. For high pressure helium, the density is comparable to air at atmospheric pressure. Therefore, designs similar to conventional vortex tubes may be applicable. In the absence of
Figure 6. Hypothesized Temperature Changes Over a Cycle.
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experimental data for helium in a closed system with varying pressure drop, the information in commercial "cold fraction" temperature tables for air can be used to approximate predicted performance of a double vortex tube relative to the performance of a conventional heat exchanger by determining the change between the temperature of the fluid leaving the pulse tube and fluid returning to the pulse tube in a later part of the cycle. Since the mass of fluid leaving and returning to the pulse tube is the same for all cases, the difference between the temperature at which the fluid exits the pulse tube and the temperature at which it returns is directly proportional to the amount of heat that the fluid has rejected during the cycle. For purposes of illustration, a hypothetical analysis is illustrated in Figure 6. It assumes that the heat sink is always at 300K. Fluid is assumed to exit the warm end of the pulse tube at a temperature 20 K greater than the 300 K sink temperature. The fluid flow is then split 50/50
between hot and cold streams. The assumed change in temperature in each stream is 40 K, up or down. With the cold stream insulated and partially-effective heat transfer between the warm stream and the 300 K heat sink, a 280 K cold stream re-unites with a 310 K cold stream at 295 K.
That outflow temperature, in turn, is assumed as the input temperature for return flow from the reservoir back to the pulse tube. On the return flow, the stream is again split 50/50 between streams at 335 K and 255 K. Again, the cold flow is insulated. The temperature of the warm flow again approaches but does not reach the heat sink temperature. The re-combined stream re-enters the pulse tube at 282.5 K, or 17.5 K below heat sink temperature. Implications for design
The temperature spread between hot and cold streams in the double vortex tube depends upon the pressure drop through the system; high pressure drop produces high temperature differences. Because it is relative pressure drop that controls, the desired pressure drop can thus be obtained by adjusting mean pressure (which makes relative pressure changes larger or smaller in absolute terms) or by adjusting pressure ratio, or both. Changing the amplitude of the cyclical pressure ratio also changes the temperature ratio between maximum and minimum temperatures in the presumably-adiabatic pulse tube. Some optimization thus is required to determine the ideal combination of absolute pressure and pressure ratio. Commercial vortex tubes are offered in a variety of capacities ranging from 2 to 150 standard cubic feet per minute of compressed air. One vendor (Exair) offers vortex tubes ranging in length from 105 to 280 mm (3). Some reported vortex tube research has been conducted with far larger apparatus (4). Finding the correct combination of dimensions for a vortex tube to be incorporated into a specific pulse tube cooler system is a challenge. As a starting point, commercial vortex tubes can be measured to determine appropriate proportions for a typical
vortex tube intended to create a stream of cold air in an open system. However, the open commercial vortex tube and the closed-system double vortex tube use different fluids, different
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PULSE TUBE CRYOCOOLER CONFIGURATION INVESTIGATIONS
Figure 7. Double Vortex Tube experiment
pressures, different flow patterns and different approaches to heat rejection. Moreover, in the closed double vortex tube system, dead volume and flow resistance must be optimized in relation to the optimum flow in and out of the warm end of the pulse tube. For the experimental work described here, the dimensions of the double vortex tube were dictated primarily by the sizes of commercially available vortex generators. That work was intended as a preliminary starting point for future research and design. Tests Initial work on heat transfer in vortex tubes was done with a unidirectional Ranque/Hilsch vortex tube of conventional internal design and unconventional external design. The vortex tube itself is 3.86 mm in internal diameter and 65.6 mm long. It is bored into a finned aluminum bar 37.8 mm in diameter, giving the assembly a thermal mass (with fittings) of 329 J/K. A simple test with helium was to exhaust in an open system to ambient pressure. Although imprecise (i.e. relative flow fractions are unknown) these heat transfer rates were observed: Test model of the double vortex tube
First tests of a double vortex tube as a warm end heat exchanger and flow impedance for a pulse tube were run at the NASA-Ames Research Center's cyrogenics laboratory. The experimental double vortex tube shown in Figure 7 was designed for a degree of flexibility in the choice of vortex generators and cold flow outlets. The vortex tube itself is bored through the center of a round aluminum bar 63.3 mm in diameter. The vortex generators are replaceable brass rings into which tangential passages have been machined. Radial passages near
the ends of the bar connect to annular plenums that surround the vortex generators. The radial passages connect to tube fittings that are screwed into the sides of the bar. The vortex tube itself is 71 mm long and 4.2 mm in diameter. Three sets of Vortec vortex generators, slightly modified in some cases, were used. They are rated for 2, 4, and 8 cfm with air at 0.68 MPa (100 psi) gauge pressure. The size of the tangential passages was determined by the size of the openings in the set of vortex generators selected, resulting in a wide but coarse grid of possible openings. The vortex generators are held in place by plastic plugs that contain the axial cold outlet openings. The plastic plugs, in turn, are held in place by end caps. The insulated cold passages and the point of re-combination of warm and cold flows are inside the aluminum bar. Eventually,
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the plastic end-plugs were drilled out from 1 to 1.5 mm internal diameter, increasing cross section of that flow passage by more than 100%. Again, the adjustment covered a substantial range, but in a large step. Test rig
Tests were run using a linear-drive compressor operating at 60 Hz, with variable displacement. Displacement of about 16 ml was used in early runs and 13 ml was used later in the experiments. Charge pressure was 2 MPa. Pulse tube and regenerators furnished by NASA were of conventional, in-line design. The regenerator housing was 67 mm long and 15.875 mm in mm in internal diameter, with 200 mesh stainless steel screens. The pulse tube was 77 mm long and 8.94 mm in internal diameter. The reservoir was a 1.0 liter flask. The pulse tube terminated in a flange. A variety of warm heat exchangers could be
clamped to the flange, including a conventional warm heat exchanger consisting of copper screens forced into a large-diameter tube. With the screen-type heat exchanger, flow between the pulse tube and reservoir was adjusted with an integral needle valve. For experiments with the double vortex tube, the NASA pulse tube was connected to the double vortex tube through a diffuser cone that transitions the pulse tube's inner diameter of 9 mm to a 3/16" fitting with internal diameter of 3.05 mm. All experiments were done without vacuum jacketing on the pulse tube and regenerator, using plastic insulation. For some tests, a loose wrap of aluminized plastic foil was used to reduce radiation and convection. In later experiments, a close-fitting box of rigid plastic foam was used. None of the insulation was air tight, but the plastic foam reduced frost buildup, and thus the load imposed by condensing and freezing water vapor.
Results Using the NASA screen-type warm heat exchanger and aluminized plastic insulation, a noload temperature of 133 K was obtained by careful adjustment of the needle valve. That confirmed the best no-load result that NASA had previously obtained with that configuration. Using the same pulse tube configuration but substituting the double vortex tube for the warm heat exchanger and needle valve, and foam insulation for aluminized plastic foil, a no-load temperature of 122 K was attained. The improvement does not appear to have been wholly attributable to the change in insulation. For that test, the double vortex tube was equipped with 2 scfm vortex generators and 1mm cold outlets. Other tests under similar operating conditions but with larger hot and cold stream openings produced less satisfactory results. However, when the double vortex tube was equipped with its largest vortex generators (8 scfm), small cold outlets (1 mm) and coupled to an intertance tube to improve phasing, a no-load temperatures of 104 K was attained. A variety of inertance tubes were used, and those combinations routinely produced no-load temperatures
below 110 K. It should be noted, however, that a slightly better result (97 K) was obtained on one occasion with the inertance tube alone. When the cold outlets were bored out from 1 mm to 1.5 mm, no combination of double vortex tube and inertance tube matched the earlier performance. Like the unidirectional Ranque/Hilsch vortex tube used in initial heat-transfer experiments, the double vortex tube has substantial thermal mass of about 985 J/K. Based upon rates of change of temperature of that thermal mass, 8 W of heat was being rejected to the double vortex tube with the cold joint between regenerator and pulse tube of the NASA test rig at 122K. CONCLUSION
The double vortex tube is a heat exchanger with unique properties. By splitting the incoming fluid flow into hot and cold streams and rejecting the heat from the hot streams, the double vortex tube can emit a stream of fluid that is potentially colder than the heat sink itself and
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thus colder than can be obtained with a conventional heat exchanger. The high pressure drop typical of this device can be calibrated to satisfy the requirement of the flow restriction for the pulse tube. Although test results are preliminary, they are encouraging. More research needs to be done to confirm the validity of this concept and optimize the hardware. ACKNOWLEDGEMENTS
Much of the work described above was financed by the Ballistic Missile Defense Organization under an SBIR Phase I research contract managed by Air Force Phillips Laboratory at Kirtland Air Force Base. The U.S. Government has certain rights in pending patent applications for the double vortex tube. Work at NASA-Ames Research Center was conducted under a nonreimbursable Space Act Agreement. The helpful assistance of Peter Kittel, Pat Roach, Jeff Lee, Ali Kashani, Ben Helvensteijn, and Al Spivak is gratefully acknowledged. REFERENCES
1. David, M., Marechal, J-C., Simon, Y., "A Method to Estimate the Pulse Tube Refrigerator Performances," Proceedings, 7th International Cryocooler Conference (1993), pp. 1078-1085. 2. ARTX, Reference Manual, Compressed Air Productivity Tools, ARTX, Fairfield, OH, (1996) p. 6. 3. EXAIR, Compressed Air Products for Industry, EXAIR, Catalog No. 96, Cincinatti, p. 5. 4. Sibulkin, M., "Unsteady, viscous, circular flow Part 3. Application to the Ranque-Hilsch Vortex Tube", Journal of Fluid Mechanics, vol. 12, part 2, (1962), pp. 269-293.
Investigations on Regenerative Heat Exchangers I. Rühlich and H. Quack Lehrstuhl fuer Kaelte- und Kryotechnik Technische Universitaet Dresden 01062 Dresden, Germany
ABSTRACT A comparison of the thermal attributes of different heat transfer surfaces shows a large potential to improve regenerators with respect to the ratio between pressure drop and heat transfer. The lower limit of this ratio is given by a single plate. With the use of computational fluid dynamics (CFD), shapes and arrangements of matrix elements were investigated systematically. As the most influential parameter, we determined the velocity changes within the matrix due to changes in the free flow areas and the aspect ratio of the single elements. The optimum geometry consists of slim elements in the flow direction, arranged in a staggered overlapping pattern. Such regenerators could provide thermal performance equal to stacked screens, but with up to five times lower pressure drop.
INTRODUCTION The efficiency of regenerative cryocoolers like Gifford-McMahon, Stirling, and Pulse Tube refrigerators depends to a large degree on the performance of the regenerators. This performance is, on one hand, determined by the heat capacity of the matrix. Lots of interesting research is presently being performed especially for the temperature range below 10 K. But good performance also requires very good heat transfer between the fluid and the matrix, and at the same time a low pressure drop of the flow. These are conflicting requirements. To
improve the heat transfer one normally has to pay with higher pressure drop. A further parameter that influences the thermal efficiency is the longitudinal thermal conduction in the matrix and in the gas; this should be as small as possible [1],[2]. Cryocoolers want to
produce large temperature spreads between the warm and cold end of the regenerator. This leads to an overall long and slim shape and a matrix with rather short micro-elements in the longitudinal direction. Present day regenerators consist of beds of spheres or stacked screens with typical element diameters of 0.05 to 0.3 mm. These are quite handy solutions, because regenerators can be manufactured quite easily from “semi-finished” products. But are these really good solutions, e.g. concerning the pressure drop that one has to accommodate to obtain the required heat transfer? Radebaugh et al. [3] have compared a number of different geometries concerning such a ratio between heat transfer and pressure drop. Their results are shown in Figure 1. As the abscissa, they use the Reynolds number, and as the ordinate, they use the ratio between the Colburn modulus j and the friction factor f. The higher up a geometrical shape is in this figure, the better the shape is for use in a regenerator with respect to heat transfer and pressure drop. Cryocoolers 10, edited by R. G. Ross, Jr. Kluwer Academic/Plenum Publishers, 1999
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PULSE TUBE CRYOCOOLER CONFIGURATION INVESTIGATIONS
Figure 1. Ratio j/f versus Re [3].
This evaluation shows that beds of spheres and stacked screens, the present day choices for the geometry of regenerator elements, are more than 5 times worse than e.g. parallel plates or pipes; tube banks lie somewhere in the middle. Although parallel plates and pipes are not well suited for regenerators because of their large longitudinal conduction, the fact that there seem to exist geometries that are much better than, e.g. screens, prompted us to ask the question: Which are the physical parameters that determine the ratio between pressure drop and heat transfer in regenerator beds? The aim was, of course, to identify the best possible internal geometry of regenerators.
DEFINITIONS
We started by redoing the literature search of Radebaugh et al. [3] [2] and noticed that different authors use diverse definitions for some of the parameters. To be able to compare different geometries in dimensionless quantities, one should use consistent definitions, e.g. for the reference velocity and the characteristic length. We have chosen the following definitions (Figure 2): A volume V with height H, width W, and length L is filled with elements of the matrix that occupy a volume and have a total surface S. From this, one can define porosity and hydraulic diameter:
A mass flow with density velocity is given by:
is entering through the front area
The free flow
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Figure 2. Reference geometry.
Inside the regenerator one can define a theoretical mean free flow area which leads to a mean velocity
:
in the main flow direction:
For each geometrical shape and arrangement of the matrix elements one can identify a minimum free flow area where the fluid has the highest velocity (see also Fig. 4):
For the definition of the friction factor f in the pressure drop equation, we use the pressure head connected to the mean velocity :
The transferred heat can be expressed by the change of the fluid temperature or by the heat transferred from the matrix surface: The temperature difference between matrix and flow is changing periodically (see Fig. 5). For the definition of the “mean” temperature difference we use, quite arbitrarily, the temperature difference at the midpoint of a matrix element: is the flow averaged temperature
For the dimensionless parameters following definitions apply:
To describe the pressure drop in a certain length of the regenerator, we introduce the term “number of pressure heads” NPH:
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To describe the quality of the heat transfer, we apply the principle of “number of transfer units”
NTU:
To describe the suitability of a geometrical configuration for use as a regenerator we calculate the ratio NPH/NTU:
When taking experimental or calculated data for f and Nu from other authors one has to be
very careful, because often different reference velocities and lengths are being used. In the American literature often Colburn modulus j is defined as:
Whereas the characteristic length, which is used in Nu and Re, cancels out in the definition
of j, one has to make a choice for the velocity applied in the Reynolds number. Some authors use some whereas we find to be the most appropriate reference velocity for an investigation dealing with regenerators, because this way one can compare geometries with different porosities on an equal basis. It is a well known fact, that the values for Nu depend on the thermal boundary condition (constant wall temperature, constant heat flux etc.). In regenerators a linear change of matrix temperature with the regenerator length is the most realistic boundary condition. Concerning the influence of the Prandtl number on the Nu number, this varies slightly from geometry to geometry and also with the Reynolds number. Since we are primarily interested in the behavior of gases like helium or nitrogen, we limit the investigation to the special case of The data from the literature were arranged in a diagram with the Reynolds number as abscissa and the ratio NPH/NTU as ordinate (Fig. 3). The data for parallel plates is taken from analytical solutions [4],[5]. From the same sources one can get the following table for flow inside some characteristic channel geometries:
Values for the sphere bed, stacked screens, tube banks, crossed rods and the square channel are taken from Kays/London [5]. Results of measurements for screens vary a lot. New data tend to lie close to the lower limit of the given range. Data for serrated plate-fins are taken from [6]. The Data for “Etched Foil Regenerator” are taken from [7]. Analytical solutions normally do not include the effect of longitudinal conduction, whereas the experimental do. The single plate is a special case, since it is not possible to construct a regenerator out of a single plate. Furthermore the calculation of a pressure drop is not directly adaptable to a single
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plate, but in spite of this, the ratio NPH/NTU can be derived by use of the boundary layer theory
[8] as follows: We assume a thin single plate with length L in the laminar flow region and a uniform heat flux over the plate length. The averaged Nu number for the whole length can be taken from the boundary layer theory [8],[9]:
As the thin plate has no form drag, the friction factor can be calculated directly from the wall shear stress:
Figure 3. NPH/NTU.
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PULSE TUBE CRYOCOOLER CONFIGURATION INVESTIGATIONS
Figure 3 contains basically the same information as Figure 1 with an inverse scale on the
ordinate. But now the scale of the ordinate has a clear physical meaning. And the base lines “single plate” and “parallel plates” determine the best ‘possible limits’ , against which all other
geometries can be evaluated. The curves show a very small dependence on the Reynoldsnumber. A rise of the curves on smaller Reynolds numbers is due to longitudinal conduction effects. The rise of the curves to higher Reynolds numbers indicates flow separation or the onset of turbulence. When inspecting Figure 3 one again sees the large difference between the ‘smooth’ surfaces and the cross flow across rods, screens or spheres. NUMERICAL SIMULATION
To get a deeper understanding of the principles standing behind the fluid flow and heat transfer in a regenerator matrix, we decided to apply a numerical calculation of the full flow and
temperature field. We chose a commercial CFD-package as tool for the calculations (TASCflow). This solver is in principle suitable to calculate the compressible and non-stationary Navier-Stokes equations in combination with the energy equation. So far we have restricted the investigations to noncompressible steady state calculations. Geometries were varied in a systematic way, concerning: • Shapes of the single matrix elements, e.g. rods with round, elliptic or lenticular cross sections of varying aspect ratios,
•
Arrangement of the elements in the matrix in staggered or in-line way with varying distances
•
Porosity. Figure 4 shows a unit cell of the calculation with the location for the smallest flow area For the evaluation of the steady state heat transfer, an overall linear temperature profile was imposed on the matrix, but the temperature of each single element was assumed to be constant. This resulted in a stepwise – sometimes overlapping temperature profile of the matrix (see Figure 5).
Figure 4. Unit cell.
Figure 5. Thermal entrance behavior of fluid flow.
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The calculation of the flow and temperature field was extended along a certain length until entrance effects had died out and a periodic velocity and temperature profile had established. In Fig. 5 such a development is qualitatively shown for two different values of the free flow velocity In the middle of the element in the longitudinal direction the “mean” temperature difference was evaluated. is generally increasing with increasing For a lenticular shaped element in a staggered overlapping arrangement the calculated flow pattern is shown in Fig. 6 and the temperature field is shown in Fig. 7.
Figure 6. Flow pattern at Re=500.
Figure 7. Temperature field at Re=500.
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Figure 9. Variation of ratio A m /A m i n
Figure 8. NPH/NTU for Re=500.
Figure 10. Geometries with porosity 0.6: shape 1, ellipse (slimness 0.3 and 0.5), cylinder.
From the velocity and temperature fields, the values of NTU/ L and NPH / L for different
velocities can be derived, and the ratio NPH / NTU can be calculated for varying Re numbers. It should be pointed out that our calculations include the effect of longitudinal conduction. MAIN INFLUENCE PARAMETERS
When inspecting the results of the calculations for many different geometries we found that there are two main parameters: • Velocity changes in the matrix • The aspect ratio of the single elements. This can be explained in Fig. 8, where for a fixed Re number the values for NPH/NTU are shown as a function of the ratio of the geometry. The ratio is a parameter describing how much the flow area, and thus the velocity, changes during the passage of the fluid through the regenerator. It turns out that velocity changes are very influential on the pressure drop / heat transferperformance ratio of a regenerator. With higher acceleration and deceleration of the flow the pressure drop is dramatically increased, the heat transfer is improved only slightly, if at all. The second important parameter is the ‘slimness’ of the single element, i.e. the thickness to length ratio. As can be seen in Fig. 8, there is a twofold influence of the slimness. For not so slim elements like cylinders, there exists a smallest possible ratio, i.e. acceleration and deceleration of the flow can not be avoided. In addition, slim elements have an inherently smaller pressure drop, especially if they are arranged in a staggered overlapping manner.
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Figure 11. Dependence of characteristic values on Re for shape 1.
Lenticular shaped elements with a slimness of 0.3 in an arrangement as shown in Fig. 6 and Fig. 7 have a ratio This means that the fluid sees hardly any acceleration and deceleration. In addition, it meets regularly new leading edges with very good local heat transfer. On the back side of the element the boundary layer is still quite thin because the following row of elements is preventing a separation of the flow. So there is still reasonably good heat transfer even at the back side of the element. The result is a ratio of NPH/NTU of about 2.4 at which is nearly as good as the parallel plate arrangement. And, as mentioned above, this value includes the longitudinal conduction in the matrix. The influence of the Reynolds number on interesting values f, j, NPH/NTU is depicted for “shape 1” in Fig. 11. The increase of NPH/NTU towards lower Re numbers is due to the axial thermal conduction. CONCLUSIONS
CFD calculations have revealed the main influence of the geometrical parameters on the pressure drop to heat transfer ratio of regenerator matrices. The best performance can be expected from arrangements which minimize the acceleration and deceleration of the flow. The optimum geometry consists of slim elements in flow direction in a staggered overlapping arrangement. Such regenerators could provide equal thermal performance as stacked screens with up to five times lower pressure drop. ACKNOWLEDGMENT
Financial support by a BMBF grant (13N6619/0) is gratefully acknowledged. REFERENCES [1] Ackermann, R.A., Cryogenic Regenerative Heat Exchangers, Plenum Press, New York, 1997. [2] Kroeger, P.G., Performance Deterioration in High Effectiveness Heat Exchangers due to Axial Heat Conduction Effects, Advances in Cryogenic Engineering, Vol. 12, Plenum Press, New York, 1967, p. 363. [3] Radebaugh, R., Louie, B., Proceedings of the Third Cryocooler Conference, NBS Special Publication 698, Washington DC, 1985.
[4] Heat Exchanger Design Handbook, Hemisphere Publishing Corporation, Vol. 1, 1983.
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[5] Kays W. M. and London A. L., Heat Exchangers Compact, 3rd Ed., New York: McGraw-Hill, 1985. [6] Taylor, M.A., Plate-Fin Heat Exchangers- Guide to their Specification and Use, HTFS publication 1990, p. 3.6. [7] J.S. Nigen et al., 3-D Flow Model for Cryocooler Regenerators, Cryocoolers 9, Plenum Press, New York, 1997, p. 451. [8] Schlichting, Grenzschichttheorie, Springer Verlag, Berlin, 1997. [9] Wilde, K., Wärme- und Stoffübergang in Strömungen, Bd. I, Dr. D. Steinkopff Verlag, Darmstadt, 1978.
NOMENCLATURE A
area specific heat at constant pressure hydraulic diameter
f j L NPH
NTU Nu P Pr
fanning friction factor Colburn modulus
length in flow direction or plate length mass flow rate number of pressure heads pressure drop number of transfer units Nusselt number
Re S
pressure Prandtl number Reynolds number wall surface
T U V
temperature velocity component in main flow direction Volume
J/(kgK) m -
m kg/s Pa K
m/s
Greek symbols heat transfer coefficient porosity viscosity thermal conductivity
kinematic viscosity shear stress
Pa
density difference
-
Subscripts C fluid H m
cold concerning the fluid constant heat flux
matrix
solid matrix
x
mean flow direction warm wall undisturbed value in surroundings
W 0
mean value
Pressure Drop in Pulse Tube Cooler Components H. E. Chen, J. M. Bennett, S. Yoshida, A. Le, and T. H. K. Frederking
Cryogenics Laboratory, Chemical Engineering Department, School of Engineering and Applied Sciences University of California, Los Angeles Los Angeles, California 90024
Abstract The present work investigates the usefulness of the perforated copper plate as multi-orifice plate for mass flow metering . Discharge coefficients have been measured, for Reynolds number 1 to 2000, using both liquid nitrogen and air. The consistencies of the results with theoretical models as well as classical orifice data encourage flow metering over a broad range of Reynolds numbers. Introduction
The heat exchanger studied here is a copper perforated plate with a geometric configuration described as a circular orifice system. The copper orifice plate is a novel flow component for which early data were reported1. The pressure difference across the plate was measured at various flow rates (primarily at low flow rates). The present and continuing study determined discharge coefficients with liquid nitrogen and air, and characterized the plate geometry by the Darcy permeability. The data obtained was compared to results from literature for orifice systems of similar geometry, and to theoretical predictions based on the characteristics of the plate. Theory Discharge coefficient. Many properties of the flow are described by the discharge coefficient The discharge coefficient is defined as follows2:
Experimental mass flow rates through the holes of the perforated plate is calculated as follows: Ideal mass flow rate, from Bernoulli & continuity equations, is calculated as follows: is a ratio defined as the velocity-of-approach factor.
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Flow Geometry Characterization. The Darcy permeability is employed as a reference cross section in order to characterize the flow geometry. Since “zero velocity” is inaccessible, the Darcy permeability was obtained by extrapolation of the data toward zero using the modified
Reynolds-Forchheimer equation: According to the above equation, a plot of is linear, and intercept. Darcy permeability is obtained as follows:
is the zero mass flow rate
From Darcy permeability, the relationship between the resistance across the orifice plate and the mass flow rate can be determined. Flow resistance ratio demonstrates such relationship
and is defined as follows:
The dimensionless velocities
and are used interchangeably. Reynolds number, based on the characteristic length (Lc), is defined as where Lc is defined as For comparison, the ratio of the two numbers is:
is
Experiment
The purpose of this experiment is to investigate the relationship between discharge coefficient and mass flow rate in order to evaluate the perforated plate’s potential as a mass flow metering device. Figure 1 shows the experimental setup for gas flow schematically. The system consists of a copper perforated plate joined by two pipes of equal diameter on either side. The tubes cover roughly 25% of the surface area of the plate while the remaining 75% is reserved for heat input (as is for the pulse tube). The air enters the upstream tube at low flow rates, then accelerates across the plate, and exits downstream into atmosphere. The pressure difference across the plate is recorded at various flow rates.
Figure 1. Schematic of Experimental setup and copper perforated plate dimensions.
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Figure 2. Plot of flow resistance versus mass flow rate. Results
Hydraulic (Darcy) permeability determination. A graphical analysis of the pressure gradient divided by the mass flow rate versus the mass flow rate (fig 2) leads to a determination of both the Darcy permeability and the two constants in the modified Reynolds-Forchhiemer Equation(eq 2). The resulting Darcy Permeability is
Figure 3 compares various Darcy permeabilities obtained from data sets and models. The experimental results were in agreement with previous work on a similar circular orifice system, but they were slightly lower than values predicted by models one and two. The variation between experimental results and models were due to certain assumptions made. The model results are upper bounds of the systems (Appendix). The lower Darcy permeability demonstrates that the actual resistance to flow is greater in this plate than predicted by the two models.
Flow Characteristics. To study the relationship between the flow resistance and the mass flow rate, the resistance ratio (eq. 6) is applied at various flow rates. A plot of resistance ratio versus Reynolds number is shown in figure 4. This plot compares our experimental data to that obtained for a similar circular orifice system4, and to that predicted for screen data3. Figure 4 shows that, for low flow rates, similar circular orifice systems behave like screen data. The data from Shull et. al4. follows the screen prediction very closely, and our data are also similar in magnitude to the screen prediction.
Figure 3. Comparison of the Darcy permeability to our models and a previous work.
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Figure 4. Flow resistance ratio versus Reynolds number.
Discharge coefficients. Figure 5 and 6 are linear plots of discharge coefficient versus Reynolds number. The dashed line in figure 5 represents the extrapolation of discharge coefficient at low flow rates. This line was determined by inserting the values for and into the modified Reynolds-Forchheimer equation. Discharge coefficients were obtained by inserting the same values for pressure gradient into equation (3). The plot shows that in the flow regime tested here the discharge coefficients remain, almost constant, independent of mass flow rate. As seen by our extrapolation, at low flow rates the discharge coefficients are not independent of Reynolds number. The advantage of the discharge coefficients being independent of flow rate is that the perforated plate can be useful as a flow-monitoring device. Figure 6 is a log-log plot of discharge coefficient versus Reynolds number in order to permit a comparison with “classical data. The liquid nitrogen data, taken with an area ratio are consistent with orifice data reported with an area ratio of 16. At the region where Reynolds numbers are around 100, the earlier work showed a weak dependence of discharge coefficient on Reynolds number. The theoretical values for the Boundary-Layer regime present the upper bound in the limit of incompressible flow.
Figure 5. Plot of discharge coefficient versus Reynolds number.
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Figure 6. A plot Discharge coefficients versus Reynolds number for comparison with literature.
Thus, the experimental data represented in figure 5 are slightly below the upper bound due to compressibility effects. The boundary layer assymtote1 results in the following discharge coefficient function: In the low flow regime(Re<35), Darcy-Stokes flow has the following discharge coefficient:
Conclusions The discharge coefficient data show consistency with fluid flow models and preceding works
over a wide flow range. The consistency encourages the use of this perforated plate for flow metering over a broad range of Reynolds numbers. Nomenclature a
D L N
base of unit cell cross sectional area of single hole in plate cross sectional area of inlet discharge coefficient diameter of one hole diameter of the entire plate thickness of plate mass flow rate number of holes Reynolds number based on Lc Reynolds number based hole diameter
VF
velocity through plate superficial velocity velocity-of-approach factor constants shear viscosity permeability density
Acknowledgements
We would like to acknowledge our industrial supporters and fellow students for their contributions.
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References 1. Yoshida, S., Ravikumar, K.V., and Frederking, T.H.K., “Friction Factors of Stacks of Perforated Plate”, Cryocoolers 8, Plenum Press, New York (1994), pp. 259-268. 2. Rohloff, T. J., Master Thesis, UCLA (1995).
3. Luna, J. et. al., “Compact Heat Exchanger Phenomena: Pressure Drops of Anisotropic Components and Related Performance”, AIChE Symposium Series, vol. 89, no 294 (1993), pp. 80-88. 4. Shull, C., Ravikumar, K.V., and Frederking, T.H.K., “Hydrodynamic Characterization of Perforated Plate Flow Passages”, Advances in Cryogenic Engineering, vol. 39, Plenum Press, New York (1994), pp. 1615-1623. 5. Timmerhaus, K.D., and Flynn, T.M., Cryogenic Process Engineering, Plenum Press, New York (1989), pp. 495-497.
Appendix: Experimental and Ideal model Darcy Permeabilities. With an externally applied pressure gradient, for a specific fluid the proportionality between the Darcy permeability and the volumetric flow rate (dV/dt), is:
Model: single duct value. The duct has a Darcy permeability value obtained from the Hagen-Poiseulle equation.
Two system permeability values are considered: 1. The hole number related value; 2. The void fraction related value
Model system 1. The reasoning behind this method is as follows. The product of the number of holes and the hole cross sectional area is the actual flow area, while the total plate area constitutes the approach area. The system permeability for this case is:
The resulting system permeability is Possible errors resulting from the use of this approach lie in the determination of N, which is uncertain at the outer rim.
Model system 2.
Noting the reduction factor used in model 1, it follows that the system porosity may be used as a reduction factor. The system permeability in this case is:
The resulting system porosity is two quantities used to model the entire system.
Again, there are uncertainties in the
Experimental Results on Inertance and Permanent Flow in Pulse Tube Coolers
L. Duband, I. Charles, A. Ravex, L. Miquet,1 and C. Jewell2 1
Service des Basses Températures /DRFMC/CEA 17 rue des Martyrs - Grenoble 38054 Cedex 9 - FRANCE 2 European Space Research and Technology Centre, PO Box 299 2200 AG - Noordwijk, Netherlands
ABSTRACT A development program on pulse tube coolers was initiated a few years ago at CEA/SBT. Both theoretical and experimental aspects have been studied. A fairly complete model used to design and optimise most prototypes has been successfully tested against experimental results. A wide range of pulse tube coolers have been developed, with performance ranging from 100 Watts at 72 K for the low frequency / large sizes, to 1 Watt at 80 K for the high frequency miniature coolers. In addition work is currently being performed on multistage pulse tube coolers. In the framework of an ESA contract we are developing for space purposes a pulse tube cryocooler associated with a compressor based on the technology of diaphragm springs and clearance seals. Various prototypes of the pulse tube cooler have been characterised. The preliminary experimental results presented at the last cryocooler conference (1996) showed a peculiar behaviour in double inlet mode : temperature oscillations were seen as a function of the setting of the impedance (V2). The reason for these oscillations, now clearly understood and experimentally validated, is due to a gaseous mass flow circulating through the regenerator and tube. Experimental results on high frequency (45 Hz) pulse tube coolers regarding this aspect are presented, as well as technological solutions to suppress or compensate for it. In addition work has been performed on the inertance aspect in orifice mode for high frequency pulse tube. The use of inertance for the impedance (V1) improves the phase shift between the pressure and the mass flow rate, and provides a significant gain in performance. PULSE TUBE TECHNOLOGY The pulse tube refrigerator discovered in 1963 by Gifford and Longsworth1 has been extensively described and the reader is refered to the relevant publications. As shown in Fig. 1 a compressor associated with a distributor, or a pressure oscillator, generates an oscillating pressure wave and a mass flow rate in a cold head which contains a regenerator. The cooling effect is obtained by the proper phase shift at the cold end between the mass flow rate and the pressure oscillation. The phase shift is adjusted by one or two impedances.
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Figure 1. Schematic of a pulse tube refrigerator.
Several configurations of the pulse tube cryocooler exist : basic, orifice and double inlet. Each configuration differ by the state of the impedances V1 and V2 (basic = V1 + V2 closed - orifice = V1 open + V2 closed - double inlet = V1 open + V2 open). A development program on pulse tube coolers was initiated a few years ago at CEA/SBT.
Both theoretical and experimental aspects have been studied. A fairly complete model used to design and optimise most prototypes has been successfully tested against experimental results.
A wide range of pulse tube coolers have been developed, with performance ranging from 100 Watts at 72 K for the low frequency / large sizes, to 1 Watt at 80 K for the high frequencies miniature coolers. In addition work is currently being performed on multistage pulse tube coolers. PERMANENT FLOW
General Considerations In 1994 ESA issued a competitive request for quotation (RFQ) to develop a space cryocooler based on pulse tube technology and using Oxford/RAL type compressors. This RFQ
was won by a collaboration composed of CEA-SBT, RAL and MMS-B. In the framework of this contract various prototypes of the pulse tube cooler have been characterised. During the development of these various coolers, we occasionally had some peculiar behaviour that could
not be related to any predicted performance. For instance in double inlet mode, temperature oscillations were seen as a function of the V2 settings : for a given V2 setting the temperature was stable and reproducible, but a slight change in V2 could lead to a temperature change of over 100 K, though a further opening of V2 could bring back the temperature to its previous value. A typical result presented at the last cryocooler conference (1996) obtained on high frequency coolers (30 and 45 Hz) is shown in Fig. 2. Note that the plotted line on the curve is a guide for the eyes, there could have been more oscillations between each data points. It is important to note that similar results have been obtained on low frequency multi stages pulse tube coolers. A fairly extensive work was then performed on each components of the cooler, which led to some improvement, in particular on the use of inertances in orifice mode (this aspect is
developed further in this paper), but did not address the problem. Since this problem was only seen in double inlet mode (V2 open) we finally came up with the idea that some DC mass flow
rate looping through the regenerator and tube via the by pass impedance could carry an additional thermal loss and consequently significantly raise the temperature.
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Figure 2. Typical experimental results obtained on high frequency pulse tube coolers, showing the "oscillations" problem in double inlet mode (valves are used for V2).
Moreover at the 1996 Cryocooler conference D. Gedeon2 presented theoretical calculations predicting this effect. This permanent flow is a micro flow which loops through the regenerator, the tube and the bypass impedance V2. This flow should not be seen like a real DC flow but rather like a disymetric flow; the mass flow rate which crosses a section in one direction during half a cycle is not the same as the mass flow rate in the opposite direction during the next half cycle. Indeed simple considerations show that this DC flow can significantly degrade the performance. If we assume the tube to be adiabatic, we can see that the DC flow which travels from the hot to the cold part results in a loss given by where m is the DC component of the mass flow rate, Cp is the specific heat of Helium, and are respectively the temperature of the hot and the cold exchanger. It is important to notice that only a very small value of the DC flow (typically a few % of the oscillating flow) may produce a parasitic heat input at the cold end, independently of the direction of this DC flow, of the same magnitude of the net cooling power. In practice the gas exchanges with the tube wall and the regenerator matrix, and this exchange affects the tube and regenerator wall temperatures2. In the regenerator for instance, without any DC flow and in steady state, the heat stored by the matrix during the compression phase (gas flowing in towards the cold end) and the heat taken away by the gas flowing out during the expansion phase lead to one temperature profile along the regenerator. If a DC flow is now added, more gas will cross the matrix in one direction than in the other, which will modify the temperature gradient along the regenerator and consequently the temperature of the gas and the wall, until a new equilibrium is found. Thus by measuring the temperature profile along the regenerator and tube walls, this effect can be demonstrated. Experimental Apparatus
In order to show evidence of the DC flow, the wall temperature profiles of the regenerator and the tube were measured on different devices and under various conditions. Measurements were made using thermometers thermally anchored along the length of both tubes. It should be noted that the goal is to compare various profiles. The DC flow results from the fact that two
non linear impedance (regenerator + pulsation tube and second orifice) are in parallel. It can be easily shown by calculation, that only if both impedances were following an identical pressure
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drop versus mass flow rate law no DC flow would appear. In practice DC flow is almost automatically present, and can be increased by a geometrical asymmetry. For the purpose of the demonstration we have selected here a worst case experiment. Thin wall drilled discs with holes ranging from 0.48 mm to 1 mm were used for the V2 impedance. Although they were expected to be symmetric, pressure drop measurement performed at room temperature on a dedicated set-up indicated large difference between the discs mounted in one direction and the other. For instance the most asymmetric disc ( hole) at 2 MPa and 0.5 produces pressure drops of 99 and 58 kPa respectively for the two flow directions. Thus for reasons given before we anticipated a permanent flow to circulate within the pulse tube but we also expected we could change the direction of this DC flow simply by turning over the discs. The experiment was performed on an ESA pulse tube prototype associated with a MMS pressure oscillator, and operated at 2 MPa and 45 Hz. The result displayed in Fig. 3 clearly show evidence of the DC flow : in one case the regenerator and tube are respectively cold and warm along two thirds of their lengths, and in the other case simply by turning over the disc the opposite result is achieved. In this figure we have plotted the temperature profile along the cooler and for clarity the temperature have been normalised according to so that this ratio is 0 and 1 respectively at the cold and hot end. This particular experiment, chosen to demonstrate the presence of the permanent flow, is spectacular although it lead to very poor performance since for both cases the DC flow is fairly large. It should also be noted that in both cases the measured ultimate temperature and pressure swings were essentially the same. Similar results were obtained with various type of valves.
Work has been performed on technological solutions to reduce or suppress this DC flow. Several solutions have been evaluated with more or less success such an asymetric by-pass orifice, laminar impedance, etc.... To date the most efficient solutions seem to be with the DC flow by injecting at some appropriate location an opposite DC flow to suppress or control the former.
In the case of low frequency pulse tube coolers using a compressor associated with a distribution valve, for which we had similar DC flow problems, a specific arrangement easily implemented because it relies on the existence of two distinct pressure reservoirs (LP, HP), is now used. We have shown that it is possible to totally suppress the oscillations problem3 and the performance have been substantially improved. For instance on a multi stage pulse tube the cooling power has been increased from 1 to 3 Watts at 21 K.
Figure 3. Temperature profiles - Operating conditions 2 MPa / 45 Hz - Impedance V1 : Orifice.
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Figure 4 . DC flow controller set-up.
For the high frequency pulse tube using a pressure oscillator, work is currently being performed on a DC flow passive controller, and as shown in Fig. 4 a technique which appears interesting for laboratory use has been tested. A one way valve is used at the outlet of the pressure oscillator to fill a high pressure reservoir which is then connected via an adjustable needle valve to the surge volume. In Fig. 5, the positive effect of this device is demonstrated. Two curves are shown. The first one is obtained with the V3 valve closed, i.e. standard double inlet configuration for which as expected temperature oscillations are obtained; then for each point, the valve V3 is opened and consequently a controlled DC flow is injected into the buffer volume, until an ultimate temperature is reached (second curve). Thus it is possible to compensate the DC flow effect which is increasing as the by pass impedance is opened. The magnitudes of the flow derived at the output of the oscillator and injected in the buffer volume are shown in Fig. 6. On this figure we have plotted the temperature profiles obtained for a V2 settings of 5 turns (ultimate temperature), and for various values of the controlled DC flow injected into the pulse tube.
Figure 5 . Operation in double inlet mode with and without DC flow controller (2 MPa - 45 Hz).
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Figure 6 . Influence of the controlled DC flow on the temperature profiles.
Indeed this flow corresponds to a few percent of the total mass flow rate. At the best performance (21.1 mg.s-1) a regenerator profile fairly close to the profile due to the thermal
conductivity of stainless steel is demonstrated. INERTANCES
In most mechanical coolers to obtain any efficient cooling effect requires schematically to gather the maximum amount of gas at the cold end of the cooler and then to expand this gas. This means that the pressure and the mass of gas are out of phase by 90° (the average mass is maximum while the pressure drops from high to low) and thus consequently the mass flow rate and pressure oscillations are in phase at the cold end. It is well known that in a Stirling or GM cooler the adequate phase shift is obtained by controlling the phase of the regenerator/displacer piston with the pressure oscillator piston. In a pulse tube this phase shift is obtained pneumatically, in the so-called orifice mode, by use of an impedance and a surge volume. Yet as shown hereafter this pneumatic "RC" type phase shifter may not provide the proper phasing depending on the nature of the impedance 4, 5, 6 : in most cases the zero phase shift is obtained at the hot end and not at the cold end, simply because the tube has a finite volume in which gas is alternatively "stored and restituted" at a finite speed, which indeed results in a phase shift between mass flow rate at both ends. It appears very helpful to develop a simple electrical analogy of the pulse tube cooler, which provides a straight forward demonstration of this aspect and gives an obvious solution to this problem experimentally verifiable. In this analogy the pressure and mass flow rate correspond respectively to the voltage and current, and the impedance is defined as This analogy is valid if we assume the behaviour is laminar. As in an electrical circuit, the impedance Z can have three contributions : a resistance (pressure drop), a capacitance (dead volume) and a self inductance. In the acoustic community, a self inductance is called an inertance and is typically a pipe. For a capillary of diameter and length 1, a tube of internal volume at temperature it can be easily shown that for a laminar flow the equivalent self L and associated resistance R, and the capacitance C are respectively given by:
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where and M are respectively the gas viscosity, density and molar mass, and is the ratio of isobaric to isochoric specific heats. In some cases (regenerator for example) the compression can
be assumed to be isothermal. An equivalent circuit for a pulse tube operated in orifice mode is shown in Fig. 7. The pressure oscillator is represented as a voltage source, the regenerator is represented as a resistance Rr (pressure drop in the regenerator matrix) in parallel with a capacitance (void volume) Cr. The pulse tube and buffer volume are pure capacitances and "A" features a current amplifier, to take into account the temperature difference between the cold and warm end of the PT, and is the impedance between the tube and reservoir (buffer). This circuit has been intentionally simplified for the purpose of the demonstration; the regenerator is subjected to a temperature gradient and thus it would be more accurately described by for instance a series of capacitances and resistances in parallel. In addition we are operating at high frequency and the buffer volume is fairly large so the term will be neglected in the following. Using this simple approach we can evaluate the phasing between the cold mass flow rate and the oscillating pressure. From the above equations and the electrical junction rules and using complex impedances, the current and voltage, or mass flow rate and pressure, are described as :
where is a coefficient which takes into account that the mass flow rates are at a different temperature and f = frequency of operation.
for respectively a resistance or a resistance associated in serie with an inertance.
Using the Fresnel geometrical construction and assuming P is on the x axis, we found the results displayed in Fig. 8. If the orifice is assumed to be a simple resistance
(orifice, valve), we can
see (Fig.8a) that the phase shift at the cold end of the PT is never the optimal value (zero). This is due to the capacitance effect in the pulsation tube This effect can significantly affects the performance. In order to counter balance the capacitance effect, it is necessary to introduce at the orifice level a self inductance L (inertance). As shown in Fig.8b a "zero" phase shift between the cold mass flow rate and the pressure oscillation can be restored. One can easily show that if the capacitance of the buffer volume, is taken into account, the phasing
due to the addition of the capacitances is further shifted and in this case the use of inertances is evenmore beneficial.
Figure 7 . PT orifice mode - equivalent electrical circuit.
Figure 8 a . Z1 purely resistive.
Figure 8 b . Z1 resistive + inertance.
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Figure 9. Effect of the type of impedance in orifice mode (2 MPa, 45 Hz).
It is now evident that the impedance to control the flow between the pulse tube and the surge volume must include an inertance term. In practice there is probably no equivalent to a pure self inductance; we use capillaries which are the association of a resistance (pressure drop) and a self inductance. In this case a compromise has to be found between the length 1 and the diameter to fulfil simultaneously the inertance and resistance values required by respectively the phase shift control at the cold end and the mass flow rate in the surge volume. All these aspects are clearly shown experimentally. We have reported for illustration in Fig. 9 the
results obtained in orifice mode with one prototype with various type of impedances : diaphragms (R), valves (R and R/C) sintered discs (R and R/C) and capillaries of various inside diameter(R/L).
Figure 10 . Example of inertance and resistance tuning (orifice mode - 2 MPa, 45 Hz).
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Figure 11 . Effect of the inertance term on the performance (orifice mode - 2 MPa, 45 Hz).
All these results were obtained at a frequency of 45 Hz and a mean pressure of 2 MPa. The equations given before show that it is possible for a capillary to independently adjust both the self inductance contribution to tune the phasing and the resistance term, by carefully choosing the length and diameter. This is illustrated in Fig. 10 where for one experiment we have plotted as a function of the capillary diameter, the resistance, the inertance and the corresponding ultimate temperature obtained for each diameter when varying the length. Fig. 11 displays the influence of the inertance on the performance for the same experiment. We wish to point out that these experiments were performed to study the inertance effect; the experimental set up, dead volumes, etc...were not designed with the objective to obtain the best performance.
The effect of the frequency of operation has also been evaluated. As expected from the previous simple electrical considerations, the results indicate that as the frequency is increased the corresponding inertance is reduced (larger diameter and/or shorter length for the capillary).
More spectacular results have been obtained with the last version of pulse tube developed in the framework of the ESA contract, for which an ultimate temperature of 74 K in orifice mode has been obtained with only 80% of the maximal stroke of the oscillator.
Indeed the use of self inductances (capillaries) provides a significant gain in performance. Although a helpful tool it should be noted that the electrical analogy does not take into account the inefficiency of the regenerator, and similarly temperature and cooling power are not represented. CONCLUSION
It has been experimentally demonstrated that the peculiar behaviour obtained in the double inlet configuration which substantially degraded the performance of earlier work7 is due to a DC gas flow circulating through the regenerator and tube. Instead of trying to suppress the DC flow which would in our opinion be difficult, technological solutions to balance it with a second controlled flow have been successfully tested.
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In addition work has been performed in orifice mode on the use of inertances as impedances. Inertances allow a better adjustement of the phase shift between the mass flow rate and the oscillating pressure and provide a significant gain in performance. From these two results the question now is whether or not the double inlet configuration is needed. Operation in orifice mode would by definition eliminate the DC flow and simplify the design. This requires additional work and in particular depends on the design of a very efficient regenerator. REFERENCES 1
Gifford W.E. and Longsworth R.C., "Pulse Tube Refrigeration", Trans. ASME J. Eng. Ind. (1964) 63, p. 264.
2
Gedeon D., "DC gas flows in Stirling and Pulse tube cryocoolers", Cryocoolers 9, Plenum Press, New York (1997) pp. 385-392
3
Charles I., Duband L. and Ravex A., "Permanent flow in low and high frequency pulse tube coolers - Experimental results", to be published
4
Zhu S.W., Zhou S.L., Yoshimura N. and Matsubara Y., "Phase shift effect of the long neck tube for the pulse tube refrigerator", Cryocoolers 9, Plenum Press, New York (1997) pp. 269-278
5
Gardner D.L. and Swift G.W., "Use of inertance in orifice pulse tube refrigerators", Cryogenics, vol. 37, no. 2 (1997), pp. 117-121
6
Roach P. and Kashani A., "Pulse tube coolers with an inertance tube : theory, modeling and practice", To appear in the proceedings of the 1997 Cryogenic Engineering Conference - July 97 Portland OR USA
7
Duband L., Ravex A., Bradshaw T., Orlowska A., Jewell C. and Jones B., "50-80 K pulse tube cooler cryocooler development, Cryocoolers 9, Plenum Press, New York (1997) pp. 213-221
Experimental Results of Pulse Tube Cooler with Inertance Tube as Phase Shifter
K.V. Ravikumar1 and Y. Matsubara2 1 Atlas Scientific, NASA Ames Research Center Moffett Field, CA 94035 2 Atomic Energy Research Institute, Nihon University
Funabashi, Chiba, Japan 274
ABSTRACT
Interest in miniaturization of pulse tube coolers necessitated its operation at higher
frequencies where fluid inertia effects start to show significant influence on the performance of the cooler. Commonly used means to achieve optimum performance of the cooler are orifice or
capillary tube and double inlet. Fluid inertia at higher frequencies provide another means to achieve optimum performance. This type of phase shifter is referred in literature as “neck-tube” or “inertance tube”. Ravikumar and Matsubara [Adv. Cryogenic Eng., Vol. 43] presented
experimental results of the cooler using fluid inertia as phase shifter. Their results showed that as frequency increased the “inertance tube” phase shifter enhanced the cooler performance in a region where orifice or double-inlet deteriorated the performance. For similar operating conditions inertance tube phase shifter resulted in higher pressure ratios inside the pulse tube. The fluid
inertia is a strong function of inertance tube cross sectional area and length. Further experimental results exploring the dependence of no-load refrigeration temperature and refrigeration load vs refrigeration temperature on frequency of operation, inertance tube diameter and length are
presented here.
INTRODUCTION
Performance of the pulse tube cooler (PTC) depends on the proper phase between the pressure oscillations and the fluid displacement, for example, at the cold end of the regenerator / pulse tube. The commonly used means for achieving the proper phase is an orifice - a flow restrictor placed between the warm end of the pulse tube and a reservoir. Further improvements in performance are achieved through adjustment of mass flow through the regenerator by
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connecting hot end of the regenerator to the hot end of the pulse tube through a by-pass valve which is commonly known as double-inlet valve. A good review of the Orifice Pulse Tube Cooler (OPTC) and Double-inlet Pulse Tube Cooler (DOPTC) is available in Ref. 1. The numerical simulations of pulse tube cooler conducted by Zhu et al2 investigated the effects of fluid inertia on the performance of PTC. It was shown in their work that it is possible using fluid inertia effects to control the phase and to achieve at the cold end of the pulse tube equivalent PVdiagrams similar to that of Stirling refrigerator. Tominaga3 and Gardner and Swift4 arrived at similar conclusions by invoking analogy between electrical circuit components and components of PTC. To a large extent works of Kanao et al5 and Haruyama and Inoue6 depended on fluid inertia in optimizing their high frequency miniature / compact coolers. Ravikumar and Matsubara7 reported experimental results of PTC based on fluid inertia and its comparison with OPTC and DOPTC. Experiments performed using same hardware and different phase shifters showed that the optimum performance of a PTC in the frequency spectrum could be moved to a higher frequency by simply changing from orifice phase shifter to a phase shifter based on fluid inertia henceforth referred to as “inertance tube phase shifter”. Their results also noted that as frequency increased the inertance tube phase shifter enhanced the refrigerator performance in a band width of frequencies where orifice and double-inlet phase shifters deteriorated the performance. The other important features observed in this frequency band width are: higher pressure ratios inside the pulse tube for similar operating conditions, and higher equivalent PV work at the cold end of the regenerator per unit of mass displaced through the regenerator. Mass displaced through the regenerator is based on either positive half or negative half of the cycle. The facts suggest the following benefits through the use of inertance tube phase shifter: operation at higher frequencies leads to miniaturization of refrigerator, and higher refrigeration capacity for the same regenerator or smaller regenerator for the same refrigeration capacity. Also compressor can be operated at smaller pressure ratio.
The fluid inertia, hence the function of inertance tube phase shifter, is a strong function of inertance tube cross sectional area and length. In the following sections results of further experimental investigations are presented including the dependence of no-load refrigeration temperature vs frequency of operation and refrigeration load vs refrigeration temperature on inertance tube diameter and length.
Figure 1. Scematic of experimental arrangement. a) pulse tube-orifice phase shifter, b) pulse tube-inertance tube phase shifter, and c) inertance tube phase shifter.
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PHASE SHIFTER CHARACTERISTICS
Figure 1. shows schematic of the experimental arrangement. The following three groups of experiments are conducted. Group-A. experiments are done with pulse tube and orifice phase shifter (Fig. 1-a) while the inertance tube phase shifter replaced the orifice in Group-B (Fig. 1-b).
In Group-C experiments pulse tube together with orifice are replaced by inertance tube (Fig. 1-c). The main idea behind such arrangements is to focus on the characteristic behavior of fluid at the cold end of the regenerator. Here the piston displacement simulates the gas particle displacement at the cold end of the regenerator in an actual cooler. Pressure at the compressor is measured in
order to calculate the PV-work and the pressure ratio. In an actual cooler this PV-work represents an equivalent PV-work at the cold end of the regenerator performed by a virtual piston and this work is an indicator of the cooler performance. Assuming isothermal condition at compressor, mass displaced in a half cycle is calculated which in actual cooler represents the mass displaced at the cold end of the regenerator. Buffer volume used in these experiments is 2.25 Liters and the charge pressure is 1.58 MPa. Figures 2 presents the experimental results which compare pressure ratio response of the three phase shifters to frequency variation. As for the pulse tube-orifice phase shifter, the pressure ratio did not change appreciably with frequency. Pressure ratio, however, increased initially and then decreased as orifice opening increases. On the other hand the other two phase shifters displayed a strong influence of frequency on pressure ratio. As frequency increased the pressure ratio increased initially and then decreased. At higher frequencies the pressure ratio is leveling off to a value which is a function of the compressor stroke volume when all else remains same. Inertance tube (Group-C) having 4m long and 4mm inner diameter (50.3cc) has achieved higher pressure ratios when compared to pulse tube-inertance tube combination (Group-B) having same inertance tube and 0.12m long and 19mm inner diameter (34cc) pulse tube. Also they have a common tangent at lower frequencies and the addition of pulse tube has lowered the pressure ratios and moved the maximum pressure ratio to a lower frequency. Compared in the same Figure are the results of inertance tube (Group-C) having 4m long and 6mm inner diameter (113.1cc) which points out that in addition to total volume the size and shape of the volume, i.e.,
Figure 2. Comparison of pressure ratio versus frequency characteristics of the three phase shifters. Inertance tube phase shifter (4m long, 4mm ID) with compressor stroke volumes of 13cc ( ) and 9cc ( ). Pulse tube (0.12m long, 19mm ID) + Inertance tube (4m long, 4mm ID) phase shifter with compressor stroke volumes of 13cc and 9cc Inertance tube phase shifter (4m long, 6mm ID) with compressor stroke volumes of 9cc Pulse tube (0.12m long, 19mm ID) + Orifice phase shifter with compressor stroke volume of 9cc and orifice valve openings of 11 turns (————— ), 7 turns (— —— —— ) and 4 turns ( - - - - - ) .
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Figure 3. Comparison of PV-work per unit of mass flowed through the regenerator versus frequency characteristics of the three phase shifters. (For notation please refer to the caption of Figure 2.)
flow cross section and flow length are the significant factors in determining the phase shifter characteristics. It may also be observed from Figure 2 that the pulse tube-inertance tube phase shifter (Group-B) and inertance tube phase shifter (Group-C) “amplified” the pressure ratio in comparison to pulse tube-orifice phase shifter (Group-A) for the same operating conditions and above a certain frequency. This phenomena clearly demonstrates the effects of fluid inertia due to inertance tube. It is known that the performance of OPTC is affected by its orientation in the gravity field. In the case of inertance tube phase shifter, however, inertia forces play dominant role and thereby leading to the possibility of suppression or elimination of gravity effects on
cooler performance. Similar observations also results from Figure 3. which presents the PV-work per unit of mass flowed through the regenerator, an indicator of performance of regenerator and also cooler as a whole, as a function of frequency. The dependence is linear and weak in case of pulse tube-orifice phase shifter where as the dependence is non-linear and strong for the other two phase shifters. The pulse tube-inertance tube phase shifter (Group-B) and inertance tube phase shifter (Group-
Figure 4. Influence of inertance tube length and flow cross section on pressure ratio versus frequency characteristics.
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Figure 5. Influence of inertance tube length and flow cross section on PV-work per unit of mass flowed through the regenerator versus frequency characteristics.
C) in comparison to pulse tube-orifice phase shifter (Group-A) display a frequency band in which the cooler performance is higher. This frequency band, as can be noted from the Figure 3., is strongly influenced by the volume of the phase shifter, its flow cross section and flow length. Further Group-C experiments are conducted by systematically changing the flow cross section and the flow length of inertance tube in order to understand their influence on inertance tube phase shifter characteristics. The results are presented in Figures 4 and 5. The general
observations are as follows. The location of both the maximum pressure ratio and the maximum PV work per unit of mass flowed through the regenerator are moved to the higher frequency as inertance tube length is made shorter or as inertance tube cross section is increased or both.
COOLER CHARACTERISTICS
The results of the experimental studies presented in the previous section suggest that use of inertance tube phase shifter allows us to obtain higher pressure ratios in the pulse tube and to enhance the equivalent PV work per unit of mass flowed through the regenerator. Both of these effects are beneficial in improving the cooler performance. In this section performance of the pulse tube cooler with inertance tube phase shifters of various flow cross section and flow length are presented. The results of inertance tube phase shifter are also compared with the results of
orifice phase shifter and double inlet phase shifter. For a realistic comparison all the experiments are performed using the same regenerator and pulse tube changing only the phase shifter. A rotary valve along with a GM compressor is employed as a pressure wave generator. The regenerator has inner diameter of 38 mm and length of 92 mm, and its matrix is made up of 200 mesh phosphor bronze screen disks. The pulse tube is a 120 mm long thin walled stainless steel
tube having 19 mm inner diameter. Comparison of the no load refrigeration temperature vs frequency of operation characteristics of pulse tube cooler with various phase shifters are presented in Figures 6 and 7. In case of PTC with inertance tube phase shifter, it can be noted from these figures, the optimum frequency at which the lowest no load refrigeration temperature occurs is inversely proportional with inertance tube length and directly proportional to flow cross section of the inertance tube. Comparing the results of Figures 2 and 3 with Figure 6 it may be said that the optimum
frequency can be predicted from the results of simple experiments described in the previous
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Figure 6. Influence of inertance tube length on no load refrigeration temperature versus frequency characteristics of pulse tube cooler with inertance tube phase shifter.
section. The results also show that inertance tube phase shifters become useful at higher frequencies where orifice and double inlet phase shifters loose their effectiveness. Another observation is that the benefits of higher pressure ratio in the pulse tube and high equivalent PV work per unit of mass flowed through the regenerator in case of inertance tube phase shifter in comparison to orifice phase shifter are more than offsetting the regenerator losses and other losses that occur at high frequencies. This can be said so because the same hard ware has been used in all the experiments. Figure 8. presents the refrigeration capacity vs refrigeration temperature characteristics of PTC with various inertance tube phase shifters. OPTC characteristics at optimum frequency and another at a higher frequency are also shown in the Figure for comparison purposes. It can be noted from the results that inertance tube phase shifter is able to reproduce the optimum
performance of orifice phase shifter at a higher frequency. Also, better performance has been
Figure 7. Influence of inertance tube flow cross section on no load refrigeration temperature versus frequency characteristics of pulse tube cooler with inertance tube phase shifter.
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Figure 8. Refrigeration capacity versus tefrigeration temperature characteristics of pulse tube cooler with inertance tube phase shifter.
achieved at higher frequencies by inertance tube phase shifters than orifice phase shifter which looses its effectiveness at higher frequencies.
CONCLUSIONS
Fluid inertia effects show significant influence at higher frequencies. In this work usefulness of fluid inertia as phase shifter has been experimentally explored further for use in pulse tube cooler. Inertance tube not only acted as a good phase shifter at higher frequencies, but also “amplified” pressure ratio in the pulse tube compared to orifice phase shifter. It also increased cooler performance by enhancing the ratio of equivalent PV work at the cold end to the mass flowed through the regenerator. Also, as frequency increased the inertance tube phase shifter enhanced the cooler performance in a region where orifice or double-inlet deteriorated the performance. Inertance tube characteristics are strongly dependent on inertance tube cross sectional area and length. Qualitatively, the optimum frequency for inertance tube phase shifter has been found to vary inversely with its length and proportionally with its cross section. Experimental results showing the dependence of no-load refrigeration temperature and refrigeration load vs refrigeration temperature on frequency of operation, inertance tube diameter and length are presented here. Dominant role played by inertia of fluid in the case of inertance tube phase shifter may effectively be used in suppression of gravity effects on cooler performance. Further work is needed for quantifying the optimum frequency of inertance tube
phase shifter and its dependence on its length and cross section, and to determine the influence of other cooler components such as regenerator, compressor and reservoir on the functioning of the inertance tube phase shifter. ACKNOWLEDGEMENTS
First author thankfully acknowledges the National Science Foundation for supporting the research with a fellowship under Center for Global Partnership program.
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REFERENCES
1. R. Radebaugh, Advances in Cryocoolers, Proceedings of the 16th International Cryogenic Engineering Conference/International Cryogenic Materials Conference, Editors: T. Haruyama, T. Mitsui and K. Yamafuji, Elsevier, Part-1, p:33 (1997). 2. S. W. Zhu, S.L. Zhou, N. Yoshimura and Y. Matsubara, Phase Shift Effect of the Long Neck Tube for the Pulse Tube Refrigerator, Cryocoolers 9, Editor: R.G. Ross, Jr., p:269 (1997).
3.
A. Tominaga, Phase Controls for Pulse-Tube Refrigerator of the Third Generation, Cryogenic Engineering, Vol. 27, No. 2, p: 146 (1992) (in Japanese).
4. D. L. Gardner and G.W. Swift, Use of Inertance in Orifice Pulse Tube Refrigerators, Cryogenics, Vol. 37, No. 2, p: 117(1997).
5. K. Kanao, N. Watanabe and Y. Kanazawa, A Miniature Pulse Tube Refrigerator for Temperatures Below 100K, Cryogenics, Vol. 34, ICEC Supplement, p: 167 (1994).
6. T. Haruyama and H. Inoue, Cooling Characteristics of a Modified Miniature Pulse Tube Refrigerator, Advances in Cryogenic Engineering, Vol. 41B, p: 1427 (1996). 7. K.V. Ravikumar and Y. Matsubara, Pulse Tube Refrigerator Based on Fluid Inertia, Advances in Cryogenic Engineering, Vol. 43 (1998) (in press).
Observation of DC Flows in a Double Inlet Pulse Tube V. Kotsubo, P. Huang, and T. C. Nast Lockheed Martin Missiles & Space Advanced Technology Center Palo Alto, CA, USA 94304-1191
ABSTRACT
Pulse tube cryocoolers using the double inlet configuration have demonstrated increased performance over the single orifice configuration. However, the double inlet has an internal closed loop fluid path that can allow a dc fluid current to flow unregenerated from ambient temperatures to the cold end, which can impose a significant heat load on the cold stage. As pointed out by Gedeon2, there are hydrodynamic mechanisms in pulse tube cryocoolers that generate these flows. These mechanisms include "streaming", where oscillating flow and pressure fields viscously interacting with boundaries lead to non-zero time-averaged mass flows, and asymmetric pressure drops due to asymmetric entrance/exit effects. Recently, several investigators have demonstrated that dc flows are present in double inlet pulse tubes. To investigate this further, we have intentionally introduced dc circulating flow into a double inlet pulse tube, controlling both the magnitude and direction of the flow. We observed shifts in the temperature profiles of both regenerator and pulse tube, with the directions and magnitudes consistent with the magnitude and direction of the dc flow. We have measured significant loss of cooling power as the dc flows increased. We have also observed that dc flows can be highly unstable under certain conditions, with sudden onsets of large flows with only minor changes in operating conditions. INTRODUCTION
There has been substantial improvement in pulse tube efficiencies over the past few years.
One of the breakthroughs has been the introduction of the double inlet configuration1, which has two degrees of freedom in the terminating impedance to allow optimization of both real and imaginary parts, leading to substantial efficiency improvements over the single orifice pulse tube. The double inlet configuration, however, has a closed loop fluid path, which can allow a dc circulating mass current to flow unregenerated from ambient temperatures to cold temperatures, and can result in a significant heat load on the cold stage. As first pointed out in a publication by Gedeon2, there are high order hydrodynamic mechanisms that generate dc biases wherever there is viscous pressure drop. Various researchers have reported unexplained behavior in the double inlet pulse tube which can be attributed to dc flows. These include work by Shigi et al.3, and Seki et al.4 More recently, Ju et al.5 performed experiments where anemometry was used to demonstrate dc flows in a multi-bypass pulse tube. Ravex has also performed experiments specifically to demonstrate the presence of dc flows6. He used asymmetric orifi in the secondary bypass to
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induce dc flows, and reported shifts in the temperature profiles of the pulse tube and regenerator, and loss of cooling power. In unpublished work, Radebaugh has performed experiments on dc flows where he used an adjustable check valve to adjust the dc flows7. In these measurements, the cooling power of a double inlet pulse tube was observed to increase with proper adjustment of the check valve setting.
We have extended the measurements by Ravex by intentionally inducing dc flows into a double inlet pulse tube. Using a novel dual needle valve secondary orifice, we controlled the dc flow both in direction and magnitude. The temperature profiles of both the regenerator and pulse tube shifted in a direction consistent with the direction of dc flow, and cooling power decreased as the magnitude of the flows increased. We have observed fluctuations in the wall temperatures of the pulse tube at high dc flow rates, which we attribute to fluctuations in the dc flow rates. In certain
situations, the dc flow is not well behaved, but can be unstable with respect to small changes in secondary valve settings.
MECHANISMS OF DC FLOWS Figure 1 shows a schematic of a double inlet pulse tube. As indicated by the arrows, the double inlet configuration has a closed loop fluid path that allows a dc mass flow to circulate through that loop. The mass flows unregenerated through the pulse tube and applies a heat load given by:
Figure 1. Schematic of the dc flow path in the double inlet pulse tube. Also shown is the dual metering valve configuration used to control the dc flow.
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where is the mass flow rate, is the specific heat, and are the ambient and coldstage temperatures. An on the order of only a percent of the AC mass flow rate can significantly degrade
the cooling power. There are three mechanisms that generate dc biases from the oscillating velocity and density fields, all requiring viscous interaction with solid boundaries. As pointed out by Gedeon2, the mass flow in the Ergun equation is density dependent for a given velocity, such that in general, the mass flow does not time average to zero, resulting in a small dc component. Pressure drops due to eddies at duct entrances and exits are also asymmetric with the direction of flow, with the degree of asymmetry dependent on the Reynolds number and geometric details. Finally, within the pulse tube itself, the interaction of the oscillating flow with the tube walls generates dc flows by a phenomena known as acoustic streaming, described, for example, by Rayleigh8. Recently, work by Lee et al.9 and Olson and Swift10 has calculated the streaming flows in a pulse tube. Thus, the regenerator, pulse tube, capillary lines, and orifi are all sources of dc biases. The dominant mechanism in the regenerator/pulse tube flow path is the non-zero time averaged mass flow from the Ergun equation as pointed out by Gedeon. Within the bypass line, duct entrance/exit pressure drop asymmetry, and asymmetric mass from the time averaged Ergun equation are the contributors, with the relative contributions depending on the details of the flow passages. Since the density-velocity phasing is nearly identical to the pressure-velocity phasing, and since the pressure-velocity phasing is such that the hydrodynamic work flow is away from the compressor, the dc mass flow bias is also away from the compressor such that the dc pressure is higher on the pulse tube side of the regenerator. We have observed this in all of the pulse tubes we have tested. Since both the regenerator and the bypass line tend to produce dc biases in the same direction, the resultant dc circulation is the imbalance between the two. These imbalances can arise from (1) differences in the Reynolds number dependence of the friction factors in the regenerator
or bypass capillary; (2) highly non-linear duct entrance/exit effects; (3) imbalances in the densityvelocity time averages due to differences in temperature oscillation amplitudes, (4) differences in viscosity because of the temperature difference between regenerator and bypass line, and (5) gas
density differences between regenerator and pulse tube due to the average temperature differences.
EXPERIMENTAL DETAILS The pulse tube used for these experiments was an in-line configuration comprised of a #400 mesh stainless steel screen regenerator with a volume of 3.5 cc and a pulse tube volume of 1.8 cc, both constructed with thinwall stainless steel tubing. We used our 1710C dual piston compressor11
with a swept volume of 3.36 cc to drive the pulse tube. The frequency of operation was about 45 Hz, with a charge pressure of 2.7 Mpa. This pulse tube reached a no-load temperature of 55 K, with a cooling power of 1.4 W at 80 K with 100 W of total compressor power. The coldstage temperature was monitored with a calibrated PRT, and heat was applied with manganin wire heater epoxy bonded to the coldstage. An array of thermocouples was epoxy bonded along the outer walls of the regenerator and pulse tube to observe the temperature profile. The coldtip was wrapped with a few layers of MLI and mounted in a vacuum can for thermal isolation. The vacuum feedthroughs and much of the ambient temperature wiring were ordinary copper wire such that some absolute thermocouple accuracy was lost; the temperature profile measurements are primarily qualitative in nature. We induced and controlled the magnitude and direction of the dc flow with a pair of metering valves in the secondary bypass line, mounted in parallel but with the valve stems oriented in the opposite direction12 as shown in Figure 1. These valves are known to have a pressure drop asymmetry. Our characterizations have shown about a 20% difference in pressure drop for the same mass flow but in opposite directions, with the lower pressure drop corresponding to flow in the direction of the needle. While we have not verified this for oscillatory flow, the results of the experiments described below strongly suggest similar behavior for ac flows. These valves served
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both as dc flow control and secondary orifice impedance adjustments. A metering valve was also used for the primary orifice. RESULTS
We operated the pulse tube at fixed frequency, fixed piston stroke, and fixed primary orifice setting, and observed the behavior of the pulse tube as we adjusted the dual secondary metering valves. We typically regulated the coldstage temperature to a constant value during these tests such that there should be no effects on the temperature profile other than that generated by secondary valve adjustments. Our first observation was the sensitivity of the temperature profile to the secondary settings, which shifted in a direction consistent with dc flow in the direction determined by the metering valve asymmetry. This is illustrated in Figure 2. The dashed line profile corresponds to the maximum cooling power, and was obtained by adjusting both secondary valves to about a 4 to 3 ratio in the number of turns opened. We believe this is the case where there is minimal dc flow. The open circles correspond to having opened only that needle valve which, if the dc flow is in the direction of the needle, would generate flow from right to left in the figure. The filled circles are data for the case where only the other needle valve is opened. As expected, if the dc flow is from the regenerator to the pulse tube, the warm mass flow from ambient imposes an additional load on the regenerator, increasing the local temperature, while in the pulse tube, the flow is cooled by the coldstage, flows into the pulse tube, and reduces the local temperature. While we have no direct measurement of the mass flows, we consider this compelling evidence for dc flows within the pulse tube. These results are similar to those presented by Ravex6. These temperature profile shifts were accompanied by loss in cooling power. Figure 3 shows the cooling power and the accompanying regenerator and pulse tube midpoint temperatures as only one secondary valve is opened, and as both valves are opened simultaneously, with the ratio adjusted to keep the midpoint temperature constant. Experimentally, each run is initiated with
Figure 2. Shifts in the regenerator and pulse tube temperature profiles from dc flows. Dashed line is with no dc flow; solid circles are with dc flow from left to right; open circles are with dc flow from right to left.
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Figure 3. Behavior of the midpoint temperatures of the regenerator (circles) and pulse tube (triangles) and cooling power (diamonds) as a single bypass valve is opened allowing dc flows (Figure 3a), and with both bypass valves opened, adjusted to maintain the midpoint temperatures nearly constant, preventing dc flows (Figure 3b).
both secondaries closed, the primary opened at a fixed number of turns, and the temperature regulated well above the no-load temperature such that there is measurable cooling power. As shown in Figure 3a, if only one secondary is opened, the regenerator midpoint temperature (open circle) remains constant, then abruptly increases once the secondary is opened past about 1 turn. The pulse tube midpoint temperature (open triangle) simultaneously decreases at this point. This shifting of temperature is precisely the behavior depicted in Figure 2. The cooling power (open diamond), begins to increase as expected from the double inlet effect, then begins to decrease at about the point where the regenerator and pulse tube temperatures begin to deviate from the initial state. We interpret this as the onset of dc flows. If the other secondary only is opened, the same
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behavior is observed except that the pulse tube temperature increases and regenerator temperature decreases. As shown in Figure 3b, when both needle valves are opened simultaneously, and the ratio of the number of turns adjusted to maintain the regenerator and pulse tube temperatures
(filled circle and filled triangle) near the initial point, the cooling power (filled diamond) reaches higher values at substantially larger secondary openings, suggestive of dc flow cancellation. While the experiments discussed above had dc flows which were well behaved and well controlled by the dual needle valve technique, in other instances the flow was highly unstable. Figure 4 shows an example of this, plotting the time trace of temperatures of the various thermocouples along the outside of the pulse tube for a particular experimental run. Starting at the left, the cooler is turned on and allowed to cool down. Once the bottom temperature is reached, the coldstage temperature is regulated, as indicated by the first arrow. Each successive arrow corresponds to an incremental increase in the opening of a single bypass needle valve. As expected, the temperatures start increasing, corresponding to dc flow from the warm end of the pulse tube toward the cold end. The striking feature is that, at large needle valve openings where the dc flow is high, the temperatures show large fluctuations. We interpret this as the dc flow showing highly erratic and unstable behavior. At the end of the run, the compressor is turned off, and the pulse tube is allowed to warm. Here, the temperature also fluctuates, and is likely due to natural convection in the pulse tube since the pulse tube was mounted vertically, with the cold end up.
We have also observed unstable response to needle valve adjustments in other pulse tubes. In one case, almost no change was observed in pulse tube and regenerator temperature profiles for a large range of needle valve settings until a critical setting was reached. Then, a small incremental adjustment produced a drastic change in temperature profiles. The setting at which this occurred depended on the operating conditions such as the setting of the other valve, or charge pressure or frequency. We have also observed instances where the direction of dc flow was in the opposite direction expected from the particular valve being opened. In these cases, we speculate that the pressure drops from entrance/exits effect or sharp bends in the secondary bypass lines dominate over the pressure drop through the metering valves, and hence strongly affect the dc flows. We
Figure 4. Behavior of the temperatures along the pulse tube as dc flow levels are increased. At high flow levels, the temperatures are unstable, suggesting the dc flow is also unstable.
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have also observed that contamination leads to similar unstable behavior. We do not understand the mechanisms that lead to these instabilities. SUMMARY
We have demonstrated that the double inlet pulse tube can have dc circulating flows that can reduce the efficiency of the pulse tube. We used a dual needle valve technique to control both magnitude and direction of the flow, and have observed shifts in the temperature profile of both regenerator and pulse tube, and loss of cooling power as the dc flow increased. The onset of DC flows can be highly unstable—in certain situations, slight changes in conditions, such as contamination buildup, may generate dc flows. The use of a permanent asymmetric orifice has been suggested by Gedeon as a method of dc flow control in a real device13. ACKNOWLEDGMENT
This work was supported by Lockheed Martin Internal Research and Development funding, and by a NASA AITP cooperative agreement, NCC5-117. REFERENCES 1. Zhu, S. W., Wu, P. Y., and Cheng, Z. Q., “Double-Inlet Pulse Tube Refrigerators: An Important Improvement,” Cryogenics, Vol. 30 (1990), p. 514.
2. Gedeon, D., “DC Gas Flows in Stirling and Pulse Tube Cryocoolers,” Cryocoolers 9, Plenum Press, New York, p. 385. 3. Shigi, T., et al., “Anomaly of One-Stage Double-inlet Pulse Tube Refrigerator,” 16th Int. Cryo. Eng. Conf., Japan (1996).
4. Seki, N., et al., “Temperature Stability of Pulse Tube Refrigerators,” 16th Int. Cryo. Eng. Conf., Japan (1996). 5. Ju, Y. L., Wang, C., and Zhou, Y., “Dynamic Experimental Investigation of a Multi-bypass Pulse Tube Refrigerator,” Cryogenics, Vol. 37 (1997). p. 357. 6. Ravex, A., Poncet, J. M., Charles, I., and Bleuze, P., “Development of Low Operation Frequency Pulse Tube Refrigerators,” to be published in Adv. Cryo. Eng.
7. Radebaugh, R. and Marquardt, E., private communication. 8. Lord, Rayleigh (Strutt, J. W.), Theory of Sound, Dover, New York (1945).
9. Lee, J. P., Kittel, P., Timmerhaus, K. D., and Radebaugh, R., “Higher Order Pulse Tube Modeling,” Cryocoolers 9, Plenum Press, New York (1996). 10. Olson, J. R., and Swift, G. W., “Acoustic Streaming in Pulse Tube Refrigerators: Tapered Pulse Tubes,” Cryogenics, Vol. 37 (1997), p. 769. 11. Nast, T. C., et. al. “Design, Performance, and Testing of the Lockheed-Developed Mechanical Cryocooler,” Cryocoolers 8, Plenum Press, New York (1994), p. 55.
12. This was suggested to us by Greg Swift of Los Alamos National Laboratory. 13. Gedeon, D., private communication.
Suppression of Acoustic Streaming in Tapered Pulse Tubes J.R. Olson* and G.W. Swift Condensed Matter and Thermal Physics Group Los Alamos National Laboratory Los Alamos, NM 87545 *Current address: Lockheed Martin Advanced Technology Center Palo Alto, CA 94304–1191
ABSTRACT In a pulse tube cryocooler, the gas in the pulse tube can be thought of as an insulating piston, transmitting pressure and velocity from the cold heat exchanger to the hot end of the pulse tube. Unfortunately, convective heat transfer can carry heat from the hot end to the cold end and reduce the net cooling power. Here, we discuss one driver of such convection: steady acoustic streaming as generated by interactions between the boundary and the oscillating pressure, velocity, and temperature. Using a perturbation method, we have derived an analytical expression for the streaming in a tapered pulse tube with axially varying mean temperature in the acoustic boundary layer limit. Our calculations showed that the streaming depends strongly on the taper angle, the ratio of velocity and pressure amplitudes, and the phase between the velocity and pressure, but it depends only weakly on the mean temperature profile and is independent of the overall oscillatory amplitude. With the appropriate tapering of the tube, streaming can be eliminated for a particular
operating condition. Experimentally, we have demonstrated that an orifice pulse tube cryocooler with the calculated zero–streaming taper has more cooling power than one with either a cylindrical tube or a tapered pulse tube with twice the optimum taper angle.
INTRODUCTION In a simplified view of the operation of the orifice pulse tube refrigerator,1 the gas in the pulse tube acts as a long (and slightly compressible) piston, transmitting pressure and velocity oscillations from the cold heat exchanger to the orifice at higher temperature. The gas in the pulse tube must thermally insulate the cold heat exchanger from higher temperatures. Unfortunately, convective heat transfer within the pulse tube can carry heat from the hot heat exchanger to the cold heat exchanger and thereby reduce the net cooling power Such convection can be steady or oscillatory, and has causes as mundane as gravity or as subtle as jetting due to inadequate flow straightening at either end of the pulse tube.
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Here, we consider convection driven by streaming. In the acoustics literature, streaming2 denotes steady convection which is driven by oscillatory phenomena. In the context of the pulse tube, this driving can occur in the oscillatory boundary layer at the side wall of the pulse tube; in this layer, both viscous and thermal phenomena are important. Lee et al.3 first suggested that tapering a pulse tube slightly, as illustrated in Figure 1, might suppress such streaming-driven convection. A general method for calculating boundary-layer streaming was published by Rott4 for the case of axially varying wall temperature, with standing wave phasing between pressure and velocity and with constant tube cross-sectional area. However, standing wave phasing is a poor assumption for the pulse tube in an orifice pulse tube refrigerator, because significant acoustic power flows along the pulse tube. We recently published5 a more general method for calculating streaming, following Rott's approach but incorporating variable cross section and arbitrary phase between pressure and velocity. This method yielded a prediction for the taper angle that suppresses streaming. We also presented measurements confirming these calculations. Here, we briefly summarize our previous work in order to bring it to the attention of a wider audience, and we elaborate on some aspects. SUMMARY OF OUR PREVIOUS WORK We restricted attention to ideal gases, so that
and
where p is the pressure,
is the density, is the gas constant, T is the temperature, a is the sound speed, and is the ratio of specific heats. The coordinates used are shown in Figure 1. The relevant variables were expanded to second order in the equations of hydrodynamics and heat transfer; for example:
where u is the axial component of the velocity, x is the axial position, y is the lateral distance
Figure 1 Schematic representation of pulse tube refrigerator with tapered pulse tube. The coordinate system used for the equations is shown. The angle of the conical pulse tube has been greatly exaggerated.
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from the wall, t is the time, is the angular frequency of the oscillations, and Re[z] denotes the real part of z. The lateral component of velocity v has the same form as u; the dynamic viscosity and the density have the same form as T. In this notation, the variables with subscript “m” are steady–state mean values, without time dependences; these represent the values the variables would have if there were no oscillating
pressure or velocity. The mean temperature profile
is assumed to be known, and leads to
the x dependences of and We omit the subscript “m” on constant properties (such as ) and on variables for which terms of order higher than mean are unimportant in this work (such as a). The subscript 1 indicates the first-order part of each variable, which accounts for
oscillation at angular frequency The first-order variables are complex quantities, having both magnitude and phase to account for their amplitudes and time phasing. The oscillating pressure and the lateral spatial average of the oscillating axial velocity are assumed to be known. Expressions for the other oscillating variables (temperature, density, etc.) in terms of and are well known.4,6 The tube radius is assumed to be much larger than the boundary layer thickness. Surprisingly, the oscillatory part of the viscosity cannot be neglected; we assume it to be independent of pressure and to depend on the oscillatory temperature via the temperature dependence of the viscosity, which takes the form Second–order, time– independent parts of variables are indicated by the subscript “2,0”. This includes the axial mass flux density which is of greatest interest here. With this notation, we found the streaming mass flux density just outside the boundary layer to be
where is the phase angle by which leads A is the pulse tube cross–sectional area, and is the Prandtl number. This mass flux (which is outside the thin boundary layer but still effectively “at” the wall compared to R) acts like a forced slip boundary condition, to drive the gas at larger y into a profile of steady streaming. For sufficiently slow streaming, and with pulse tube radius much larger than the penetration depth, the resulting velocity profile is essentially parabolic, as shown in Figure 2a, with the gas velocity near the wall equal to the drift velocity just outside the penetration depths, and the velocity in the center of the pulse tube determined by the requirement that the net mass flux along the tube must be zero. Thus, the streaming profile away from the wall is given by where r is the radial coordinate in the pulse tube. In turn, this parabolic streaming profile convects heat along the pulse tube. (To display the boundary-layer behavior clearly, the Figure does not have and hence the maximum of falls slightly short of in the Figure. However, our calculations are only valid for
when the maximum of
)
Setting Eq. (4) equal to zero, we found that streaming is eliminated in a tapered pulse tube if
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PULSE TUBE FLOW AND OPERATIONAL STABILITY INVESTIGATIONS
In Eq. (6), we have used numerical values for low-temperature helium gas: Note that for a conical tube, where is the taper angle
Figure 2 (a) The net drift near the wall, illustrated in (c) and (d), affects the entire pulse tube,
causing an offset parabolic velocity profile which leads to a second-order mass flux density For the signs chosen here for illustration, the gas moving downward in the center of the pulse tube is hotter than the gas moving upward around it, so that heat is carried downward. The coordinate r measures the distance from the center of the pulse tube, of radius R. The calculations are for R >>
, but, for clarity,
is exaggerated in the figure. (b) Mass flux density for the case where the streaming is suppressed by tapering the tube. Although there is still some streaming near the wall and a correspondingly small offset flux in the rest of the tube, they carry negligible heat, (c) Illustration of net drift caused by one process within a penetration depth
of the pulse tube wall. A parcel of gas is shown in three consecutive
positions. Here we imagine that the temperature dependence of the viscosity is the only important effect, and assume that the temperature is lower during upward motion than during downward motion. (d) Illustration of net drift caused by a different process, as discussed in the text.
ACOUSTIC STREAMING IN TAPERED PULSE TUBES
311
(see Figure 1) and R is the pulse tube radius. For this value of dA/dx, the parabolic part of the velocity profile shown in Figure 2a is eliminated; the only nonzero streaming occurs at distances
from the wall comparable to the penetration depths, as shown in Figure 2b. To verify Eq. (6), we tested an orifice pulse tube refrigerator with three pulse tubes: a right– circular cylinder, a truncated cone with the optimum angle determined by Eq. (6), and a second
cone with roughly twice the optimum angle. The double-angle cone should induce the same amount of streaming as the cylindrical pulse tube, but in the opposite direction, and so should exhibit roughly the same performance as the cylindrical pulse tube. The pulse tube refrigerator, shown schematically in Figure 1, was filled with 3.1 MPa helium and driven at 100 Hz. The optimum–angle cone pulse tube (see below) had a total included angle of with the cold end larger than the hot end. This is a small enough angle that it is barely discernible by eye. The double–angle pulse tube had an angle of 0.103 rad. In both cases, the internal diameter at each end was calculated in order to make the volume of the pulse tube the same as in the cylindrical case. The results of our measurements are shown in Figure 3. The data corresponding to the cylindrical pulse tube are represented by the circles, while those of the optimum–angle conical pulse tube are shown as triangles and those of the double–angle cone as squares. For all measured values of applied heat, the temperature corresponding to the optimum cone pulse tube was roughly 5–10 K colder than the temperatures with either the cylindrical or the double-angle pulse tubes, indicating that the optimum cone performed significantly better than the other pulse tubes. From an alternative point of view, it appears that the streaming–driven convective heat load on the cold heat exchanger was 3–5 W greater with the cylindrical and double–angle pulse tubes than with the optimally tapered pulse tube.
DISCUSSION The complexity of Eqs. (4) and (5) is due to many nonlinear interactions among oscillatory
Figure 3 Experimental results, showing cooling power as a function of the cold temperature for three different pulse tubes with equal volumes and lengths. The conical pulse tube (triangles), designed according to Eq. (6), had more cooling power than either a cylindrical one (circles), or one with twice the optimum cone angle (squares). The lines are only guides to the eye.
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pressure, velocity, and temperature in the fundamental equations of hydrodynamics and heat transfer. One such interaction is easily imagined and is illustrated in Figure 2c. Consider a small parcel of gas oscillating up and down along the wall, at a distance from the wall of the order of the relevant boundary layer: the viscous penetration depth On average, the gas between the parcel and wall will have a different temperature during the parcel's upward motion than during its downward motion, due to imperfect thermal contact with the wall's temperature gradient and to the adiabatic temperature oscillations with time phasing between oscillatory motion and oscillatory pressure. Since the viscosity depends on temperature, the moving parcel will experience a different amount of viscous drag during its upward motion than during its downward motion, and hence will undergo a different displacement during its upward motion than during its downward motion. After a full cycle, the parcel does not return to its starting point; it experiences a small net drift which contributes to This process is represented by terms proportional to the product of b, and in our starting nonlinear equations, and is responsible for the presence of b in Eqs. (4) and (5). The effect of a taper on streaming can also be imagined, with reference to Figure 2d. In general, a gas parcel close to the wall will be farther from the wall during, say, its upward motion than during its downward motion, due to, for example, the compressiblity of the gas in the
boundary layer and the phasing between oscillatory motion and pressure. Hence, again the moving parcel will experience a different amount of viscous drag during its upward motion than
during its downward motion, and so the parcel will again fail to return to its starting point after a full cycle. This effect is represented in our starting equations by terms proportional to the product
of and However, the boundary–layer continuity equation couples and while the yaveraged continuity equation couples to dA/dx. Hence the process shown in Figure 2d is controlled in part by the taper dA/dx, and is responsible for the
presence of dA/dx in Eq. (4). Including all such second-order streaming effects in our calculations allowed us to determine the conditions under which they all add to zero, represented by Eq. (5). Note that all the variables in the right hand side of Eq. (5) or (6) should be known during the design of a pulse tube refrigerator and are experimentally accessible. The most challenging of these variables are the velocity and phase difference but these can be calculated from other known or measured quantities: The pressure amplitude and phase in the compliance, its volume, and the adiabatic compressibility of the gas yield the volumetric velocity and phase of the gas entering the compliance, which is essentially that at the hot end of the pulse tube. The volumetric velocity and its phase at other x in the pulse tube can then be estimated from its geometry and the adiabatic compressibility of the gas. We generally use these methods to calculate the magnitudes and relative phases of and at both ends and in the middle of the pulse tube, substituting these values into Eq. (6) to determine the optimum dA/dx at the ends and in the middle of the pulse tube. In all cases we have considered to date, the values have been such that a simple cone is a reasonable approximation to the ideal shape, making more difficult fabrication unnecessary. The fact that streaming in a pulse tube can be suppressed so simply and conveniently is the result of a remarkable series of fortunate coincidences. First, there was no a priori guarantee that tapering the pulse tube would have a large enough effect on streaming to cancel streaming’s other causes. Second, it might have turned out that rather large taper angles (say, greater than 10 or 20 degrees) were required; in this case, flow separation of the high-Reynolds-number first-order velocity from the pulse tube wall would have invalidated the entire laminar, boundary-layer approach. Third, it is fortunate that most pulse tubes operate in the “weakly turbulent” regime of oscillatory flow, and with tube surface roughnesses much smaller than the boundary-layer thicknesses, so that laminar analysis is adequate in the boundary layer. Fourth, the perturbation expansion upon which this calculation is based is only valid for zero or extremely weak streaming—the very situation we are most interested in. This point is subtle.
ACOUSTIC STREAMING IN TAPERED PULSE TUBES
313
Strong streaming (but nevertheless with streaming velocity small compared to the oscillatory velocity) distorts the axial temperature profile of the pulse tube significantly, contradicting our fundamental assumption that the time-averaged temperature, density, etc in the boundary layer are well approximated by their zero-oscillation values etc. This fundamental
assumption requires that the streaming be so weak that the temperature profile in the pulse tube is unperturbed by the streaming—or, equivalently, that the streaming is so weak that it carries negligible heat! Hence, the calculation self-consistently predicts the conditions of zero streaming, but it cannot be used to accurately predict the magnitude of strong streaming. Fifth, there are numerous other fourth–order energy flux terms in addition to which would in principle have to be considered to obtain a formally correct fourth-order result. Fortunately, the only other large term, is zero at the same taper angle that makes zero, while the remaining terms, such as those involving products of first and third order quantities, are small for all angles. Hence, the suppression of is sufficient to suppress streaming heat transport. Sixth, as a practical matter, it is extremely convenient that the streaming–suppression taper is independent of oscillatory amplitude. Seventh, it is also convenient that the streaming– suppression taper is only weakly dependent on temperature gradient, so that streaming is suppressed over a broad range of and of Finally, at least in our experience so far, a simple conical section is an adequate approximation to the ideal streaming-suppression taper shape A(x). ACKNOWLEDGMENTS The authors thank Ray Radebaugh, for numerous helpful discussions about pulse-tube refrigerators and from whom we first heard the idea of shaping pulse tubes to reduce streaming. This work was completed with support from the Division of Materials Sciences in the DOE's
Office of Basic Energy Sciences. REFERENCES 1
Radebaugh, R., A review of pulse tube refrigeration. Adv. Cryo. Eng., 1990, 35, 1191-1205.
2
Nyborg, W.L.M., Acoustic Streaming. In Physical Acoustics, Vol. IIB, ed. W.P. Mason, Academic Press, New York, 1965, p265. 3
Lee, J.M., Kittel, P., Timmerhaus, K.D. and Radebaugh, R., Flow patterns intrinsic to the pulse tube refrigerator. In Proc. 7th Int. Cryocooler Conf., Kirtland AFB, NM 87117-5776, Phillips Laboratory,
1993, p. 125. 4
Rott, N., The influence of heat conduction on acoustic streaming. Z. Angew. Math. Phys., 1974, 25, 417. 5 6
Olson, J.R., and Swift, G.W., Cryogenics, 1997, 37, 769–776.
Rott, N., Damped and thermally driven acoustic oscillations in wide and narrow tubes. Z. Angew. Math. Phys., 1969, 20, 230.
Performance of a Tapered Pulse Tube G. W. Swift,1 M. S. Allen, and J. J. Wollan Cryenco Inc. 3811 Joliet Denver CO 80239
ABSTRACT The figure of merit of a pulse tube is the ratio of total power to pV power flowing up and out of the hot end of the pulse tube. The figure of merit is less than unity because of heat transport from the hot end to the cold end. Streaming-driven convection in the pulse tube can contribute to this heat transport. In a well instrumented pulse tube refrigerator having 1500 W of cooling power at 125 K, we have measured the figure of merit of a tapered pulse tube at several operating points. At operating points near those for which the taper should eliminate streaming-driven convection, the figure of merit is This is close to the theoretical optimum figure of merit 0.97 calculated for this pulse tube considering only two heat-transport mechanisms: heat conduction in the metal pulse tube wall and ordinary boundary-layer thermoacoustic heat transport in the gas close to the wall. At other operating points, the measured figure of merit is much lower, as streaming-driven convection adds a third heat-transport mechanism.
INTRODUCTION Lee, Kittel, Timmerhaus, and Radebaugh2 observed a toroidal steady flow in a pulse tube, realized that it must convect significant heat, explained that it might be a streaming phenomenon driven by the sidewall boundary layer, and proposed that tapering the pulse tube might reduce it. Olson and Swift3 developed this idea further, derived an expression for the taper that suppresses streaming, and presented measurements showing that this remarkably small taper indeed yields improved pulse-tube-refrigerator performance. Here, we present further experimental evidence that streaming-driven convection is suppressed in a pulse tube tapered according to the formula of Olson and Swift. Our measurements show that, with the correct taper, the heat load on the cold heat exchanger due to the pulse tube can be close to that due to ordinary heat conduction in the pulse tube wall plus ordinary “boundary-layer” thermoacoustic heat transport near the pulse tube wall.
APPARATUS AND INSTRUMENTATION The pulse tube refrigerator used for these measurements, shown schematically in Fig. 1, was part of Cryenco’s 1997 thermoacoustically driven orifice pulse tube refrigerator,4 which Cryocoolers 10, edited by R. G. Ross, Jr. Kluwer Academic/Plenum Publishers, 1999
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PULSE TUBE FLOW AND OPERATIONAL STABILITY INVESTIGATIONS
Figure 1. Schematic of the apparatus. “P” indicates locations of pressure sensors. The angle tapered pulse tube has been greatly exaggerated.
of the
has liquefied natural gas at rates as high as 140 gallons per day. This pulse tube refrigerator is much more powerful than those usually discussed at the International Cryocooler Conferences; its measured net cooling power at 125 Kelvin ranged from 1100 W to 1700 W for the
data reported here. This high power was achieved primarily through use of large-diameter components; for example, the pulse tube itself had a diameter of 10 cm. The refrigerator used 3.1 MPa helium at 40 Hz; for all measurements reported here, the pressure amplitude was 210 kPa at the pressure sensor near the hot end of the pulse tube (see Fig. 1). In a pulse tube refrigerator with such a large cooling power, the use of inertance5,6 is easier than in smaller systems. A large inertance has relatively little dissipation, so dissipative series and parallel orifice valves, as shown in Fig. 1, can be used with a large inertance to allow adjustments of the refrigerator’s operating point over a wide range.7 In this paper, we will identify refrigerator operating points by means of the complex acoustic impedance Z at the hot end of the pulse tube, at the location shown in Fig. 1, just below the
middle flow straightener. The magnitude of Z is the ratio of oscillatory pressure amplitude to oscillatory volumetric velocity amplitude; the phase of Z is the phase angle by which
pressure leads volumetric velocity. In the design of our refrigerator, calculations8,9 of viscous and thermal losses in the regenerator led us to identify an operating point with the best expected regenerator performance at the desired cooling power, which we will call the “design” operating point and
PERFORMANCE OF A TAPERED PULSE TUBE
317
Figure 2. Complex acoustic impedance Z at the hot end of the pulse tube. Re[Z] and Im[Z] are
the real and imaginary parts of Z, respectively. The open circle is the design operating point, and the line through this circle indicates Z’s for which the as-built taper should suppress streaming. The dashed line indicates Z’s which would have had zero streaming if the pulse tube had been built straight. The filled circles indicate the experimental operating points.
which is shown as the open circle in Fig. 2. Analysis using Eq. (13) of Olson and Swift (reproduced below for convenience) showed that a straight pulse tube would have zero streamingdriven convection at operating points on the line marked “straight” in Fig. 2. We judged that the design operating point was too far from this line, so we used Eq. (13) to find that a taper of (1/A) should result in zero streaming-driven convection at the design operating point. Our pulse tube was built with this taper, which is equivalent to a total included angle of 1.3° with the cold end larger than the hot end. Further analysis using Eq. (13) showed that all operating points on the line marked “as built” would have zero streaming-driven convection with this taper. The five experimental operating points, shown as filled circles and labeled “1” through “5,” were selected to be close to the design operating point and close to the “as-built” line of possible operating points. We predicted that experimental operating points farthest from this line would exhibit the greatest streaming-driven
convective heat transfer. To detect the presence of this convective heat transfer, we measured the total power flow up the pulse tube and the acoustic (or pV) power up the pulse tube. In the absence of any axial heat transfer in the pulse tube (due to conduction, convection, or radiation), so the deviation of from unity is a measure of such undesired heat transfer,
which directly reduces the net cooling power of the refrigerator. Hence, this ratio is often called the figure of merit for a pulse tube. In most previous pulse tube refrigerators, the pulse tube figure of merit has typically9 been in the range from 0.6 to 0.85. For our experiments, all parts of the pulse-tube refrigerator (including orifice valves,
inertance, and compliance) were thermally insulated, except for the fluid streams in contact with the heat exchangers. Hence, measurement of the heat carried away from the heat exchanger at the hot end of the pulse tube by its cooling water was a direct measurement of To measure this heat, we used an electric resistance heater in the water downstream of the heat exchanger, with three thermocouples: upstream of the heat exchanger, between
the heat exchanger and the electric heater, and downstream of the electric heater. Thus the desired heat could be obtained by multiplying the easily measured electric heater power by
a ratio of temperature differences, in ignorance of the exact flow rate and specific heat of the water (which included antifreeze).
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PULSE TUBE FLOW AND OPERATIONAL STABILITY INVESTIGATIONS
Figure 3. Figure of merit of the pulse tube at various operating points vs how far each operating point is from the “as-built” line on Fig. 2. Filled circles show experimental values. Open circle shows the design operating point; solid line shows calculated neglecting streaming; dashed line shows rough calculated estimate of including streaming.
Pressure sensors in the compliance and near the hot end of the pulse tube allowed measurement of acoustic power up the pulse tube. If the void volumes in the hot heat exchanger, inertance, orifice valves, and associated plumbing were zero, and Z could be obtained from such pressure measurements in a straightforward way7 known to most pulsetube experimentalists. However, in our hardware these void volumes were not negligible, so we used a complete numerical model8 containing the geometry of these components, with the two measured pressure amplitudes and phases, to obtain and Z. This more sophisticated
method yielded values of
and Z from 5% to 10% different from those which would have
resulted from the assumption of zero void volumes.
The uncertainty in the resulting values of is probably about 4%, which we obtain by assuming that the various uncertainties contributing to the result add in quadrature. (Worst case addition of these uncertainties is 9%.) The largest contributors are uncertainties
in the volume of the compliance, in the calibration of the ammeter used to measure the ac electric current in the water-stream heater, and in the temperature differences. Most of the uncertainties are systematic, so the most likely errors would shift all the data together, without changing their relative positions.
RESULTS Figure 3 displays the experimental values of the pulse tube figure of merit for the five experimental operating points. The horizontal axis shows how far those operating points are from operating points of predicted zero streaming-driven convection, i.e. how far those operating points are from the “as built” line in Fig. 2. To provide a quantitative measure of this “distance,” the horizontal axis of Fig. 3 is the difference between the “as-built” value
of (1/A) dA/dx and the value that Olson and Swift’s Eq. (13) yields for (1/A) dA/dx at the selected operating point. Operating points 2 and 5, which were closest to the “as built” optimal condition, had the highest value of approximately 0.96. Based on the lower values of for the other three operating points, we estimate that operating points 2 and 5 would have yielded an experimental value of near or below 0.85 if the pulse tube had been straight instead of being tapered according to Eq. (13).
The solid line in Fig. 3 near of
shows calculated values [see Eq. (4) below]
including only thermal conduction in the stainless-steel pulse tube wall and ther-
PERFORMANCE OF A TAPERED PULSE TUBE
319
moacoustic boundary-layer heat transport, but not including streaming-driven convection. (As in Fig. 2, the design operating point is indicated by the open circle.) Hence, within our experimental uncertainties of a few percent in
experimental operating points 2 and 5
exhibit no streaming-driven convection. The dashed lines in Fig. 3 are rough estimates of taking streaming-driven convection into account, using Eqs. (10) and (11) of Olson and Swift evaluated at the center (axially) of the pulse tube to estimate the streaming mass-flux density as a function of radial
position, and then calculating the heat load with the assumption that all the downwardstreaming gas is at room temperature and all the upward-streaming gas is at 125 K. These lines are in reasonable agreement with the data. They must be regarded as extremely rough estimates because many of the assumptions built into Olson and Swift’s work are only valid for streaming so weak that its effect on time-averaged temperatures is small.
MATHEMATICAL DETAILS In boundary-layer approximation, streaming is eliminated if a pulse tube of circular cross section is tapered according to Eq. (12) or (13) of Olson and Swift. We rewrite those equations in a more convenient form:
If streaming is thus eliminated, the heat flow
can be calculated using
Here, is the full taper angle (with positive signifying that the cold end is larger than the hot end), R is the internal radius of the pulse tube, A is its internal area, and is the cross-sectional area of its wall, x is the axial distance from the cold end of the pulse
tube, f is the frequency of oscillation and is the mean temperature, is the mean pressure, is the amplitude of the oscillatory pressure, is the amplitude of the oscillatory volumetric velocity and is the phase angle11 by which leads (with positive in the positive x direction), is the mean density of the gas, is its isobaric heat capacity per unit mass, is its ratio of isobaric to isochoric specific heats, K is its thermal conductivity, is its Prandtl number, where is its dynamic viscosity, is its thermal penetration depth, and is the thermal conductivity of the wall material. The second form of each equation incorporates numerical values for low-temperature helium gas: and In Eqs. (3) and
(4), the first two terms account for ordinary heat conduction, and the last, more complicated
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PULSE TUBE FLOW AND OPERATIONAL STABILITY INVESTIGATIONS
term is from Rott’s second-order (“thermoacoustic”) energy flux12 in the limit13 (Note that pulse tube end effects can cause to be significantly larger than where is the temperature difference between the two heat exchangers and
is the
length of the pulse tube.)
SUMMARY Streaming-driven convection can be eliminated in a pulse tube, either by tapering the pulse tube or by choosing an operating point for which a straight pulse tube generates no streaming. In a large pulse tube, a figure of merit is possible, as demonstrated
here. Even in smaller pulse tubes, for which boundary-larger thermoacoustic processes carry a larger fraction of the heat, figures of merit ~ 0.9 should be possible. Careful attention to detail is required to achieve such performance. The operating point Z of the refrigerator must be planned very accurately, or means to finely adjust it must be provided. Flow straightening at the ends of the pulse tube must eliminate jet-driven convection. The pulse tube wall roughness must be much less than the viscous penetration
depth, and the pulse tube must operate in the laminar or weakly turbulent regime.
ACKNOWLEDGMENTS This work has been supported by Cryenco, Inc. and by the Office of Fossil Energy in
the US Department of Energy.
REFERENCES 1. Permanent address: Condensed Matter and Thermal Physics Group, Mail Stop K764, Los Alamos National Laboratory, Los Alamos NM 87545. 2. Lee, J. M., Kittel, P., Timmerhaus, K. D., and Radebaugh, R. , “Flow patterns intrinsic to the pulse tube refrigerator,” Cryocoolers 7, Phillips Laboratory, Kirtland AFB, NM 87117-5776 (1993), pp 125-139. 3. Olson, J. R., and Swift, G. W., “Acoustic streaming in pulse tube refrigerators: Tapered pulse
tubes,” Cryogenics, vol. 37, (1997), pp. 769–776. 4. Swift, G. W, “Thermoacoustic natural gas liquefier,” Proceedings of the DOE Natural Gas Conference, Houston TX, March 1997.
5. Zhu, S. W., Zhou, S. L., Yoshimura, N., and Matsubara, Y., “Phase shift effect of the long neck tube for the pulse tube refrigerator,” Cryocoolers 9, pp. 269-278 (Plenum, New York, 1997). 6. Radebaugh, R., “Advances in cryocoolers,” Proceedings of the 1996 International Cryogenic Engineering Conference (Kitakyushu, Japan, May 1996).
7. Gardner, D. L., and Swift, G. W., “Use of inertance in orifice pulse tube refrigerators,” Cryogenics, vol. 37, pp. 117–121, (1997). 8. Ward, W. C. and Swift, G. W., “Design environment for low amplitude thermoacoustic engines (DeltaE),” J. Acoust. Soc. Am., vol. 95, pp. 3671–3672 (1994). Fully tested software and users guide available from Energy Science and Technology Software Center, US Department of Energy, Oak Ridge, Tennessee. To review DeltaE’s capabilities, visit the Los Alamos thermoacoustics web site at http://rott.esa.lanl.gov/. For a beta-test version, contact
[email protected] (Bill Ward) via Internet.
9. Swift, G. W., and Ward, W. C., “Simple harmonic analysis of regenerators,” Journal of Thermo-
physics and Heat Transfer, vol. 10, pp. 652–662 (1996). 10. Radebaugh, R., personal communication. 11. This sign convention for
is opposite that for the phase of acoustic impedance Z shown in Fig. 2.
12. Rott, N., “Thermally driven acoustic oscillations, part iii: Second-order heat flux,” Z. Angew.
Math. Phys., vol. 26, pp. 43–48 (1975). 13. Swift, G. W., “Thermoacoustic engines and refrigerators,” Encyclopedia of Applied Physics, Wiley (for American Institute of Physics), vol. 21, pp. 245–264 (1997).
Numerical Study of Pulse Tube Flow Yoshikazu Hozumi Chiyoda Corporation, 13, Moriya-cho 3, Kanagawa-ku, Yokohama Japan, 221-0022
Masahide Murakami Institute of Engineering Mechanics, University of Tsukuba, Tennodai
1-1-1, Tsukuba, Japan 305-8573
Teruhito Iida National Space Development Agency of Japan, 2-1-1 Sengen, Tsukuba Japan, 305-0047
ABSTRACT
Numerical simulation of the viscous compressible flow in a pulse tube is carried out to study the detail of the flow and the fundamental mechanism of refrigeration. Axisymmetric two-dimensional Navier-Stokes equations are solved numerically using a finite-volume method. The analytical solution of oscillating flow in terms of Bessel functions is applied for setting up the velocity profile at the pulse tube inlet as a boundary condition. The phase difference in oscillating flow between the tube center and the nearby tube wall is observed by the simulation. The temperature separation between the cold end and the hot end, and
also the time-averaged enthalpy flow, are evaluated in order to investigate the refrigeration mechanism. The simulation results suggest that the flow of steady refrigeration enthalpy depends also on the profile of the inlet velocity. Moreover, the internal velocity distribution induced by the inlet flow profile might be also an important factor for determining the optimum compressor driving frequency. INTRODUCTION
Several approaches of numerical investigation of the flow inside a pulse tubes have been carried out in recent years. The state of compressible viscous flow is expressed by the NavierStokes equations — the equation of mass conservation, momentum conservation, and energy conservation. A qualitative comparison among different approaches of modeling is made by Lee and Kittel, et al.1 Since the computation of descretized full three-dimensional Navier-Stokes equations is still beyond our current computer ability, several simplified governing equations have been investigated. One typical approach is the simplification from three-dimensions to two-dimensions, or to one-dimension, to carry out the calculations effectively. Fortunately, the phenomena inside of a pulse tube can be treated as an axisymmetric flow. Several two-dimensional approaches are reported. The other approach, reduction of nonlinear terms of the differential fluid equations to linear ones, is valid for the case with very small magnitudes of the nonlinear advection term. Standing on the assumption of small disturbances, the two dimensional linear set of equations is solved to Cryocoolers 10, edited by R. G. Ross, Jr. Kluwer Academic/Plenum Publishers, 1999
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investigate the enthalpy flow and steady secondary streaming1. Similar approaches have been
made based on linear partial differential equations for sinusoidal time dependent flow in the thermoacoustic theory2. Although there exists finite velocity flow and finite energy oscillation in realistic pulse tubes, the approaches based on linearized modeling must assume small disturbance flow. Since the practical highly efficient pulse tubes which are categorized as orifice type and double inlet type have a finite flow velocity field completely through the inside of pulse tube, the flow would be dominated by strong nonlinear advection phenomena. The nonlinear approach is required to quantitatively evaluate the performance and to optimize the pulse tube using direct simulation. In this study, axisymmetric two-dimensional Navier-Stokes equations are directly solved to predict the pulse tube flow. NUMERICAL APPROACH A finite volume method is applied to the present simulation. The phenomena inside of the pulse tube can be treated as an axisymmetric flow. Primarily, simple two dimensional NavierStokes equations are used to explain the procedure of differential formulation, then the formulated governing equations are extended to axisymmetric two-dimensional forms by using cylindrical coordinates. The two dimensional Navier-Stokes equations are written in Cartesian coordinates as follows:
where
where is density, P is pressure, u and v are velocity in x and y directions, is specific heat. In order to close the equations, the assumption of perfect gas is made. The perfect gas equation of state is is internal energy, where The properties of helium and air are used, depending on the simulation case. The above equation is transformed into a general coordinate system. An implicit flux split difference approximation is applied to the Navier-Stokes equations.
where V is the volume of a calculation cell, and suffixes + and - show transgressing flux direction in the characteristic equation. Finally, the cell volume and side stress term of the governing equation are modified because cylindrical coordinates are used to describe the phenomena. In order to avoid numerical dissipation by iterative calculation, the second-order Roe’s scheme is applied for flux calculation to keep the resolution of the simulation results. SIMULATION MODEL
Figure 1 shows the simulation model of the pulse tube. In order to focus this study on how enthalpy is transferred from the cold end to the hot end in the pulse tube refrigerator, only fluid motion inside the pulse tube is investigated. The regenerator is modeled as an energy buffer for the internal energy income and outgo through the pulse tube inlet. The heat exchanger is installed at the
NUMERICAL STUDY OF PULSE TUBE FLOW
323
Figure 1. Axisymmetric Two-dimensional model of pulse tube flow field.
hot end. A certain amount of heat is transferred depending on the temperature difference between gas and pulse tube wall. The wall is assumed to be adiabatic except the heat exchanger portion at the hot end. A 50 x 21 (length x radius) grid point mesh is generated to carry out the calculation. Viscous boundary layer thickness is given as a function of the magnitude of the internal flow velocity. For better resolution of the viscous shear stress on the wall surface, the mesh points are concentrated near the wall surface and approximately 8 mesh points are imposed within the viscous boundary layer. INLET CONDITION
A numerical simulation requires boundary conditions for solving the governing equations. Because of its great influence on the results of simulation, the practical velocity profile must be specified as the tube inlet boundary condition. The analytical solution of the Navier-Stokes equations for pulsating flow in an open-end tube is obtained in terms of Bessel functions3. The solution is expressed as a function of time and radius, and driving frequency. Nondimensional velocity is given as follows:
where r´ is the nondimensional radius, is the nondimensional driving frequency, and t´ is the nondimensional time. The dimensions of the simulated pulse tube are 288 mm x 15.6 mm (length x radius); consequently, the closed-end condition has a small effect on the inlet velocity profile, and the analytical solution can be used as the inlet condition. A comparison of inlet velocity profiles with respect to driving frequency is shown in Figure 2; here, each velocity profile is normalized by the centerline velocity. The shape of the velocity distribution is similar to Poiseuille flow at low driving frequencies. On the other hand, the location of the largest velocity moves from the center toward the wall as the frequency is increased.
RESULTS The simulation was carried out to investigate the mechanism of refrigeration so that, the time variation of the temperature
Figure 2. Comparison of inlet velocity profiles.
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Figure 3. Time variations of inlet pressure, and nondimensional velocity profiles for these frequencies.
Figure 4. Comparison of time-averaged temperature and time-averaged enthalpy flow (both for the same frequency of 10 Hz, and pressure ratio of 3.3).
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Figure 5. Temperature oscillation at both ends at 2 Hz.
that, the time variation of the temperature profile, time-averaged temperature, and time-averaged enthalpy flow are shown. Throughout the whole pulse tube simulation, the time variation of the inlet velocity profile is checked. Inlet pressure and velocities are plotted as a function of driving frequency as shown in Figure 3. The effect of the inlet velocity profiles on performance is investigated by comparing the time-averaged temperature distributions for two cases of inlet flow profiles: 1) Poiseuille-like flow, and 2) annular-like flow (See Figure 4). Significant temperature separation is observed only in the Poiseuille-like inlet flow case. The
time-averaged enthalpy flows from the cold end to the hot end, and is concentrated toward the heat exchanger at the hot end. However for an annular-like inlet flow profile, there is no significant temperature separation, so the pulse tube does not work as a refrigerator. Figure 5 shows the temperature oscillation at the cold and hot ends for operation at 2 Hz. The time variation of temperature distribution is compared for three driving frequencies in Figure 6. The simulation was carried out at a fixed pressure ratio of 3.3. The work done by the compressor is in proportion to the driving frequency, but the reachable lowest temperature at the cold end does not decrease in inverse proportion to the frequency. If the driving frequency is increased under the fixed pressure ratio, the inlet velocity will be increased. Since the pressure drop induced by fluid motion increases as the square of the velocity, the dissipative loss within the regenerator significantly grows as the drive frequency is increased. These effects are ignored in the present simulation by neglecting regenerator modeling. Therefore, the refrigeration effect might be worse in actual refrigerators. In order to explore the dependency of refrigeration on drive frequency, two different operational conditions were studied. One is operation at the same pressure ratio, and the other is operation at the same input power per unit time. The time-averaged temperatures and time-averaged enthalpy flows are compared in Figure 7. It seems that low frequency oscillation is more effective in refrigeration for unit input power.
CONCLUDING REMARKS A simulation program for axisymmetric pulse tube flow has been developed for evaluating the flow governed by nonlinear advection. Direct simulation to predict pulse tube flow is performed by the program. To establish practical inlet boundary conditions, the analytical solution of oscillating flow in terms of Bessel functions is applied. The simulation results suggest that inlet velocity profile dominates the flow of steady refrigeration enthalpy.
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Figure 7. Comparison of the refrigeration performance at three frequencies for two different operational conditions: (a) at same pressure ratio, and (b) same input power per time. The data are given for time averaged temperature (K).
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REFERENCES 1. 2. 3. 4.
5. 6.
J.M. Lee and P. Kittel, “Higher Order Pulse Tube Modeling,” Cryocoolers 9 (1997), p.345.
A. Tominaga “Thermodynamic aspects of thermoacoustic theory,” Cryogenics, 35 (1995), p. 427. A.F. D’Souza, and Oldenburger, Tran. ASME, D, 86-4 (1964-9), p. 589. M. Shiraishi and N. Nakamura, and M. Murakami, “Visualization Study of Velocity Profiles and Displacements of Working Gas Inside a Pulse Tube Refrigerator,” Cryocoolers 9 (1997), p. 355.
M. Shiraishi et al., “Visualization Study of Flow Fields in a Pulse Tube Refrigerator” Advances in Cryogenic Engineering, v.43 (1998). R. W. MacCormack and G. V. Candler, “The Solution of The Navier-Stokes Equations using GaussSeidel Line relaxation,” Computer & Fluid, v.17 No. 1 (1989), p.135.
7.
M. Shiraishi et al., “Start-up Behavior of Pulse Tube Refrigerator,” Advances in Cryogenic Engineering, v.41 (1996), p.1455
8.
W.E. Gifford and R.C. Longsworth “Surface Heat Pumping,” Advances in Cryogenic Engineering, v.11 (1965), p.171.
9.
A.S. Worlikar and O.M. Knio “Numerical Simulation of Thermoacoustic Refrigerator” Journal of Computational Physics 127 (1996), p.424.
Visualization Study of the Local Flow Field in an Orifice and Double-Inlet Pulse Tube Refrigerator M. Shiraishi, N. Nakamura*, K. Takamatsu* M. Murakami* and A. Nakano Mechanical Engineering Laboratory Tsukuba, Ibaraki, 305-8564 Japan *University of Tsukuba Tsukuba, Ibaraki, 305-006 Japan
ABSTRACT
The local flow behavior inside an orifice and a double-inlet pulse tube refrigerator has been observed during a cycle by using a smoke-wire and a tuff flow visualization technique. To clarify the difference between an orifice and a double-inlet configuration, the flow behavior in the bypass tube of the double-inlet configuration has been carefully observed. The results are classified into six flow patterns and percentile plots of those flow patterns during the cycle. From these plots, the effects of opening of the bypass valve on the pulse tube performance has been observed, and the optimum conditions defined. It is found that the flow in the bypass is similar to a typical viscous oscillating flow and its velocity is roughly one order faster than in the pulse tube. It is also found that there is a significant difference in the percentile plot between the orifice and the double-inlet configuration. For the double-inlet configuration, optimum conditions are distinguishable from other conditions by the fact that the duration times of some of the flow patterns increase significantly. INTRODUCTION
In recent years there is increased interest in using small low-cost cryocoolers to cool communications equipment and computer devices.1 Although Stirling and G-M cryocoolers are now widely used, these cooler types have the inevitable shortcoming of incorporating cold moving parts in the displacer; these makes them mechanically complicated and generally demand periodic maintenance. Therefore, the development of a high performance and reliable cryocooler with a long lifetime without maintenance is an important design goal. The pulse tube refrigerator, which has no cold moving parts, is one of the most promising cryocooler types. On the other hand, although its simplicity in structure promises high reliability and low cost, its performance is often inferior to other cryocoolers. Despite the fact that strenuous efforts have been made to develop a commercially practical pulse tube refrigerator, few systems are available. The main reason is that the refrigeration mechanism is not yet clearly understood, and useful realistic models based on experimental results do not exist. A few experiments concerning the flow behavior have been made,2,3 but very little has been concluded. Cryocoolers 10, edited by R. G. Ross, Jr.
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Flow visualization is one of the most useful methods for understanding flow behavior. Lee et al.4 have conducted flow visualization experiments and have observed secondary flow which they associate with a loss mechanism. Recently, we have also attempted to visualize the internal oscillating flow in both a basic and an orifice pulse tube refrigerator.5 Although only a bypasstube and valve were installed in the orifice configuration, the double-inlet pulse tube refrigerator achieved a lower attainable lowest temperature than the orifice configuration as long as the driving conditions were optimized. Recently, a temperature instability in the double-inlet configuration has been reported.6 Therefore, we believe that understanding the refrigeration mechanism in more detail is necessary to improve the capability and reliability of the double-inlet pulse tube refrigerator. The objective of this study is to investigate the flow behavior in a double-inlet pulse tube refrigerator with special emphasis on the flow behavior in the bypass tube. The goal is to clarify the flow behavior related to the refrigeration mechanism by comparing it with the flow behavior of the simple orifice pulse tube. EXPERIMENT
A schematic of the experimental apparatus is shown in figure 1. The pulse tube is made of a transparent acrylate tube having an effective length of 280 mm and an inner diameter of 16 mm. It has three tube fittings — at the middle and near both ends — to thread a smoke-wire. The regenerator is composed of a stainless steel housing and a Bakelite tube which is packed with 770 discs of 100-mesh stainless steel screen, has a length of 168 mm, and an inner diameter of 18 mm. The reservoir has a volume of about 5 times that of the pulse tube volume. To observe the inside flow behavior, a tee connecting at the hot end and a bypass tube, which connects between the warm side
Figure 1. Experimental apparatus.
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Figure 2. Pressure oscillations in the double-inlet configuration at 5 Hz. The supply pressure is obtained at the warm side of the regenerator. The numbers above the graph correspond to those of photographs in figure 4.
of the regenerator and the bypass valve, is also made of a transparent acrylate tube. Inner diameters of the tee connecting and the bypass tube are 6 mm and 4 mm, respectively. Tufts are installed at each section of the tee connecting to visualize the flow direction as shown in the figure. The bypass
tube has a tube fitting to fix a smoke-wire. The smoke-wire is made of a tungsten wire with 0.1 mm diameter. Both ends of the wire are connected to copper supports served as the electrical connections for the wire and as the fixing to the tube fittings. Pressure transducers are installed at three positions, the entrance of the pulse tube, the pulse tube and the reservoir. Mineral insulated thermocouples(CA) of 0.15 mm in diameter are also installed at the cold and hot ends to evaluate the performance from the difference of gas temperature. Air is used as the working gas. Pressure oscillation is generated by introducing the pressurized air of about 0.2 MPa into the rotary valve and releasing to the atmosphere. Figure 2 shows obtained pressure oscillations in the double-inlet configuration under the optimized conditions at 5 Hz. In this experiment 5 Hz is adopted as the operating frequency because the maximum temperature difference can be obtained at that frequency for the double-inlet and orifice configurations. After coating the smoke-wire with a paraffin thinly, it is suspended in the tube and connected to the high voltage pulse generator through the leads. The pulse tube is operated under the prescribed conditions of which the frequency and the compression ratio of the pressure oscillation are 5 Hz and about 1.2, respectively. By monitoring the gas temperature, the reaching to the steady state is confirmed and then, the observation and measurement are done. The change of emitted smoke-line and the movement of tuft is recorded by a high speed video camera and a still video camera. RESULTS AND DISCUSSION
Flow Behavior in the Pulse Tube Figure 3 shows the variations of the temperature difference for all pulse tube configurations at 5 Hz with openings of the orifice and the bypass valve. For the basic configuration, which is equivalent to opening of zero in the orifice valve, the temperature difference is very small. The
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OPENING OF BYPASS VALVE
Figure 3. The variations of the temperature difference between cold and hot ends with openings of orifice and bypass valves.
optimum opening of orifice valve is determined as about 35 because the temperature difference has a maximum value. For the double-inlet configuration, the measurement is conducted by changing the opening of the bypass valve under the condition that the orifice valve opening is fixed at the optimum opening of 35. The temperature difference increases slightly as the increase of the opening until about 8, but it decreases slightly for larger openings. The optimum opening is determined as in the range from 6 to 10. The temperature differences at the optimum opening for the orifice and double-inlet configurations are about 15 K and 18 K, respectively. It is confirmed that this experimental pulse tube works as the refrigerator with a significant temperature difference. Figure 4 shows the sequential close-up photographs taken at the middle point of the doubleinlet configuration during one cycle. The smoke-line is emitted at the onset of compression process and recorded after first one cycle. It should be noted that the shape of the smoke-lines is formed as an integration of movement of smoke particles from the emission so that it does not show the velocity profiles directly. Although the compressor pushes the gas toward the hot end after starting of the compression process, the gas flows toward the cold end in the beginnings of the process(see (1), (2)). This denotes that there is a phase difference between the oscillation of pressure and gas displacement Fundamental flow behaviors are very similar to that of the orifice one except for the smaller displacement7,8). Observation of Flow in the Bypass Tube and Tee Connecting
Figure 5 is a visualized flow in the bypass tube. The shape of smoke-line is concave. This means that velocity near the wall is larger than at the centerline. It becomes evident that the flow behavior in the bypass tube is also a typical viscous oscillating flow like the main flow in the pulse tube8). The deformation in the smoke-line near the wall is larger than that around the center axis, and the shape of the peaks is sharper than that seen in the pulse tube. The main flow is strongly
affected by the viscosity near the wall because of small radius of the bypass tube. The smoke-line moves with large velocity and displacement compared to that in the pulse tube. In this case the displacement of smoke-line reaches a distance of 400 mm. It is roughly one order larger than that in the pulse tube. After one cycle, in mis case the smoke-line returns to near the smoke wire so that a DC gas flow is not recognized clearly9). That is the subject for a future study.
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Figure 4. Sequential photographs of video frames recorded by the high speed video
camera at the middle point of the double-inlet configuration. The interval between photographs is 20 msec and the numbers correspond to those appeared in figure 2.
Figure 6 shows time variations of the velocity of the smoke-lines which are emitted at the cold and the middle point of the pulse tube, and in the bypass tube. Each velocity is obtained at the
center axis of each tube. The corresponding pressure oscillation is also plotted. The velocity in the bypass tube increases and decreases immediately with beginnings of the compression and expansion processes, respectively, but the variations of velocity in the pulse tube are delayed by the phase difference. The velocity in the bypass tube is roughly one order faster than that in the pulse tube. It is thought that the gas in the bypass tube reaches the hot end faster than that in the pulse tube in a compression process and the warm side of regenerator in an expansion process. This fact has
significant effects in the flow behavior in the double-inlet configuration compared with the orifice one.
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Figure 5. Composite photographs of two successive smoke-lines in the bypass tube recorded by the high speed video camera. The time interval between two smoke-lines is 5 msec.
Figure 6. The velocity of the smoke-lines during one cycle.
The flow behaviors in the tee connecting are visualized by using a tuft visualization technique. Figure 7 shows the visualized typical flow pattern under the conditions of optimum valve opening. Arrows indicate the local flow directions at that moment determined from the movement of tufts at each section. It is found that the flow behavior during one cycle can be classified into typical 6 flow patterns in the double-inlet configuration. A duration time of each pattern in a cycle changes by the opening of the bypass valve. Figure 8 shows summarized percentile plot of those patterns in a
cycle. The top figure shows the flow patterns and legends in the percentile plot. “Close” of the bypass-valve corresponds to the orifice configuration with the optimized opening of the orifice valve. For this configuration, the cycle is composed of two patterns, (C) and (F). The flow from or
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Figure 7. The flow direction visualized by tufts in the tee connecting of bypass tube.
Figure 8. The percentile plot of the flow patterns in the tee connecting during one
cycle. The top figure shows the flow patterns and legends. “Close” of the bypass-valve corresponds to the orifice configuration with the optimized opening of the orifice valve.
to the bypass valve in two patterns appears just after changing of a direction of the flow between the hot end and the reservoir, but its duration time is negligible short. “Optimum” and “Large” of the
bypass-valve correspond to the double-inlet configuration with the openings of optimized value of 10 and larger value of 20, respectively(see figure 3). It is remarkable that the patterns (A) and (D) appear as the beginning of compression and expansion processes. Appearance of these patterns characterizes the flow behavior in the double-inlet configuration. This is explained from the results in figure 6 that the flow in the bypass tube is faster than that in the pulse tube. For the larger opening
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of 20 , the duration time of patterns (A) and (D) decreases. It is observed that patterns (A) and (D) have a large effect on the optimum condition of the double-inlet configuration. SUMMARY We have observed the flow behavior in the bypass tube of a double-inlet pulse tube refrigerator as well as in the pulse tube by using a smoke-wire and a tuff flow visualization technique. The results are classified into six flow patterns and percentile plots flow patterns during a cycle. Some results are as follows: 1) The fundamental flow behavior in the pulse tube is very similar to that in the orifice pulse tube except for a smaller gas displacement. 2) The flow in the bypass tube is similar to typical viscous oscillating flow and its velocity is roughly an order of magnitude faster than that in the pulse tube. 3) There is a significant difference in the percentile plot of flow patterns between the orifice and the double-inlet configuration. For the double-inlet configuration, optimum conditions are distinguishable from other conditions by the fact that the duration times of some of the flow patterns increase significantly. 4) In this experiment, DC gas flow in the bypass tube and the pulse tube was not observed. This is a subject for future study. REFERENCES
1. Nisenoff, M., Patten, F. and Wolf, S.A., And What about Cryogenic Refrigeration?, Cryocoolers 9, Plenum Press, New York (1997), pp. 25-27. 2. Rawlins, W., Radebaugh, R. and Timmerhaus, K.D., “Thermal anemometry for mass flow measurement in oscillating cryogenic gas flows”, Rev. Sci. Instrum. 64(1993), pp.3229-3235. 3. Inada, T., Nishio, S., Ohtani, Y. and Kuriyama, T., “Experimental investigation on the role of
4. 5. 6.
7. 8.
9.
orifice and bypass valves in double-inlet pulse tube refrigerators”, Advances in Cryogenic Engineering, Plenum Press, New York, 41 (1996), pp.1479-1486. Lee, J.M., Kittel, P.,Timmerhaus, K.D. and Radebaugh, R., “ Flow patterns intrinsic to the pulse tube refrigerator”, Proceedings of the 7th International Cryocooler Conference, PL-CP93-1001, Santa Fe (1993), pp. 125-139. Shiraishi, M., Nakamura, N., Seo, K. and Murakami, M., “Investigation of velocity profiles in oscillating flows inside a pulse tube refrigerator”, Proceedings of the 16th ICEC/ICMC, Elsevier Science, Oxford(1997), pp. 255-258. Seki, N., Yamasaki, S., Yuyama, J., Kasuya, M., Arasawa, K., Furuya, S. and Morimoto, H., “Temperature stability of pulse tube refrigerators”, Proceedings of the 16th ICEC/ICMC, Elsevier Science, Oxford(1997), pp. 267-270. Seo, K., Shiraishi, M., Nakamura, N. and Murakami, M., “Investigation of radial temperature and velocity profiles in oscillating flows inside a pulse tube refrigerator”, Cryocooler 9, Edited by R.G.Ross,Jr, Plenum Press, New York, 1997, pp.365-374. Shiraishi, M., Nakamura, N., Seo, K. and Murakami, M., “Visualization study of velocity profiles and displacements of working gas inside a pulse tube refrigerator”, Cryocooler 9, Edited by R.G.Ross,Jr, Plenum Press, New York, 1997, pp.355-364. Gedeon, D., “DC gas flows in Stirling and pulse tube cryocoolers”, Cryocooler 9, Edited by R.G.Ross,Jr, Plenum Press, New York, 1997, pp.385-392.
Stability Study of Coaxial Pulse Tube Cooler Driven by Air Conditioning Compressor L.W. Yang, J.T. Liang, Y. Zhou, P.S. Zhang, W.X. Zhu, J.H. Cai
Cryogenic Laboratory, Chinese Academy of Sciences P.O. Box 2711, Beijing, 100080, China
ABSTRACT
A simple introduction is given to the development of a pulse tube cooler driven by a selfadjusted air-conditioning compressor in our laboratory. Based on our previous work, with input power increased to less than 650 W, the cooling power at 80 K increased from 4 W to 6 W and the lowest temperature slowly dropped from 51 K to 38.4 K. Some results related to stability when operating for a relatively long time are also given. They include the latest three steps of experiments which show that through continuous improvement, our pulse tube refrigeration system has been developed into a very stable system that can operate for a long time. This system will soon be available for practical applications. INTRODUCTION Pulse tube cryocoolers have developed quickly because they have no moving parts in the cold head. This results in long life and low vibration compared with other cryocooler types, and makes them particularly suitable for cooling HTS and other electronic devices.1 Such kinds of applications usually require a few watts of cooling capacity at 80 K, high reliability, and low cost. For such an application, long life low cost operation of the compressor is becoming even more crucial. A pulse tube refrigerator driven by a mature long-life commercial air conditioning compressor may satisfy this requirement. Since 1997, we have been working in this direction under the support of the Superconductivity Center of China.2,3 The following is the main work we have done. PULSE TUBE REFRIGERATION SYSTEM
The system is schematically shown in Figure 1. It includes three main parts: compressor, rotary valve, and pulse tube cooler. A water jacket outside the surface of the compressor is used to remove the heat of compression and to condense the oil vapor in the compressor. A specially designed oil filter is used to further eliminate oil vapor from the gas stream. The compressor and the oil filter are installed in a specially designed box with four wheels. The rotary valve connected to the compressor is made by our lab and its operation is quite stable. The operating speed of the motor can be adjusted at the beginning, but after the refrigeraCryocoolers 10, edited by R. G. Ross, Jr. Kluwer Academic/Plenum Publishers, 1999
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Figure 1. Pulse tube refrigeration system driven by air-conditioning compressor.
tion effect is confirmed, the motor is kept unchanged at the best speed. Generally, the best operating frequency is 3-5 Hz. When reaching a stable state, the fluctuation of frequency is within 0.2 Hz. The pulse tube is installed in the middle of the regenerator, which is filled with annular stainless steel screens. The cold tip is made of copper and the gas reservoir is made of aluminum alloy. The impedance for the orifice, double-inlet and multi-bypass are all integrated within the pulse tube and regenerator. The cold head can be easily coupled with a vacuum jacket designed according to the specifications of the user. MEASUREMENTS
In experiments, the main parameters measured include:
1. Cold tip temperature, using gold-iron thermocouples calibrated by the Cryogenics Temperature Calibration Station, Chinese Academy of Sciences. 2. Pressure wave amplitude, including average pressure, high pressure, and low pressure pressure wave. The pressure was measured using a pressure transducer; the pressure wave was shown using an HP54602B oscilloscope. 3. Operation frequency, recorded through an oscilloscope. 4. Cooling capacity, using HP6634A System DC Power Supply (0-100 V/0-1 A, 100 W). 5. Input power (including compressor and rotary valve), using AC power meter. The input power of the rotary valve is rather low, and generally below 20 W; its value changes little. 6. Vacuum level, generally 1.5-2 Pa. COOLING PERFORMANCE
In 1997, we begin to use a commercial air conditioning compressor readjusted by ourselves to drive a pulse tube cooler. In the beginning, a U-type pulse tube was used and achieved (see Fig. 2) a no-load temperature of 51 K and 4 W of cooling power at 80 K with an input power of 580 W.2 Later, a coaxial pulse tube was designed and fabricated to go with the compressor. Through improvement, a no-load temperature of 38.4 K was achieved (see Figure 3) together with 3 W of cooling power at 51 K, and more than 6 W cooling power at 77 K with an input power of about 640 W.3 Even this performance can satisfy some applications. Next we found that operational stability was a crucial issue in practice.
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Figure 2. Performance of U-type pulse tube cooler driven by air-conditioning compressor.
Figure 3. Performance of co-axial pulse tube cooler driven by air-conditioning compressor.
STABILITY EXPERIMENT RESULTS
Generally, the duration of the performance tests of our pulse tube cooler is limited to a few hours. The start-up process is recorded, which is generally within 1 hour, then the cooling load is added. When the temperature is stable, the data are recorded. Within such a period, the system is rather stable. Compared to this general condition, the time for our stability experiments is a little longer and includes two parts: the start-up process, and the later stable state. During the whole process nothing requires adjustment and the system can reach its lowest temperature automatically. Certainly, recorded parameters, such as input power, pressure ratio, frequency, changed during the start up process automatically. The following three figures show our progress in this area.
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Figure 4. First group stability experiment.
Figure 4 is the first group test data for a coaxial pulse tube cooler. It is evident that in 20 minutes the lowest temperature of 44 K was reached, and such a temperature remains unchanged for one and a half hours. But later, the temperature began to increase gradually, and in four hours the temperature increased from 44 K to 70 K. Because of this instability, the experiment was stopped after 6 hours of operation. After the experiment, we found oil existing at the regenerator inlet and rotary valve. That meant our oil filter system of the compressor needed to be improved. After improvement of the oil filter, further work on stability was continued as shown in Figure 5. Another coaxial pulse tube was used in these experiments; it has a no-load temperature of 44 K with an input power of 520 W and an average pressure of 1.5 MPa, and 4 W of cooling power at 74 K with an input power of 540 W. However, in this test the fill pressure was only 1.0 MPa. The test data show that the temperature dropped to 55 K in 50 minutes with an input of 420 W, rising to 56 K in three hours. Later, 1 W of cooling load was added and the cold tip
Figure 5. Second group stability experiment.
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Figure 6. Third group stability experiment.
temperature increased to 75 K and kept stable for one hour. When the cooling load was with-
drawn, the temperature dropped to 58 K, and this temperature remained unchanged until the experiment ended two hours later. This 9-hour test shows that, though a small fluctuation of input power and temperature is evident, the system is relatively stable. However, from the noted changes, we might predict that over a prolonged time, the temperature may increase. So further
improvement was continued. Through further experiments we found that our compressor oil filter system needed to be improved greatly. So, a new compressor was constructed with improved water cooling and oil filtration. Figure 6 gives the experimental results using this new compressor. These test data
show a very stable performance. In nearly 60 minutes, the temperature reaches 53 K with an input power of 484 W. In the following 9 hours, all parameters were stable except the temperature dropped to 52.5 K. When 2 W of cooling load was added to the cold tip, the temperature increased to 65 K quickly and kept stable, with the input power increasing to 500 W and then gradually decreasing to 486 W. About two hours later when the cooling load was withdrawn, the temperature dropped to 52 K and stabilized at 51.5 K for the remaining two hours of the test, and input power remained unchanged at 484 W. This 14-hour test shows a rather good stability in comparison to the former experiments. In fact, none of the recorded parameters became worse in this test.
The latest test improved our refrigeration system for practical applications. Longer experiments are scheduled for the future. CONCLUSION
Through continuous research and improvement the performance of pulse tube coolers has been improved. In this paper, with the quality of our self-adjusted air conditioning compressor being improved greatly, the stability of the refrigeration system was proved to be rather good. This system is being developed into a product. ACKNOWLEDGMENT
The author gratefully acknowledges the support of Superconductivity Center of China and K.C. Wang Education Foundation of Hong Kong.
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REFERENCES 1.
Thummes, G., Landgraf, R., Giebeler, F., Much, M., Heiden, C., “Pulse tube refrigerator for highTc SQUID operation”, Adv. Cryo. Eng., vol. 41B (1996), p. 1463.
2.
Yang, L.W., Liang, J.T., Zhou, Y., Zhang, P.S., Zhu, W.X., “Pulse tube refrigerator driven by commercial air-conditioning compressor”, Cryogenics and Superconductivity of China, vol. 25, no. 3 (1997), pp. 1-4.
3.
Liang, J.T., Yang, L.W., Bian, S.Y., Zhou, Y., Zhang, P.S., Zhu, W.X., “Test of a coaxial pulse tube refrigerator with air conditioning compressor”, ICCR'98, April 25-28, 1998, Hangzhou, China, pp. 19-24.
Gas Contamination Effects on Pulse Tube Performance J. L. Hall and R. G. Ross, Jr. Jet Propulsion Laboratory California Institute of Technology Pasadena, CA., USA 91109
ABSTRACT
Experiments were performed to quantify the effect of contamination in the helium working gas of a typical pulse tube cryocooler operating at 60 K with 1 W of applied heating load. A quadrupole mass spectrometer with a closed ion source was used to measure the gas compositions both before and after pulse tube operation. Five tests were performed with different combinations of water, carbon dioxide, nitrogen and argon gas contaminating the helium across a range of 20 to 420 parts per million (ppm). Except for hydrogen, no other contaminants were detected in any experiment either before or after running the pulse tube. A decrease in pulse tube refrigeration performance was observed in all five tests, with asymptotic steady-state temperatures reached on a time scale of 7 to 10 days. Under a 1 W applied load condition with a constant piston stroke, the cold head temperature rise ranged from 0.6 to 2.8 K depending on contamination level. Comparisons between cryopumped and non-cryopumped vacuum chambers appeared to rule out the possibility that the contamination was external to the pulse tube. Water and carbon dioxide were the most important contaminants at the nominal operating temperature of 60 K, although there was evidence for nitrogen condensation under some conditions. Warming up of the pulse tube to room temperature tended to restore the original performance, although this improvement degraded more rapidly the second time around to the asymptotic steady-state temperature. INTRODUCTION
One of the potential life-limiting factors for cryocoolers is internal contamination of the helium gas. Condensible species such as water and carbon dioxide will freeze on cold surfaces at
sufficiently low temperatures, leading to blockage of the flow and possibly increased parasitic heat conduction. For example, blockage problems have been observed in the small orifices of J-T cryocoolers and studies have been conducted to understand and mitigate the problem.1,2 Longterm testing of Vuilleumier coolers has also addressed the contamination issue.3 Finally, life test studies of the Oxford ISAMS prototype Stirling cooler, the Hughes/Raytheon ISSC BE Stirling
cooler and the TRW SBIRS-low mini pulse tube cooler have all showed some evidence of performance loss due to internal contamination. In order to address the problem of contamination in pulse tube cryocoolers, it was decided to conduct a set of experiments designed to identify the specific contaminant species and the levels at
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which they produce a given degradation of performance. The available equipment consisted of a small pulse tube that had no capability for flow visualization or any other means of measuring the formation and distribution of frozen contaminants inside the cryocooler. Therefore, contamination effects were deduced from measured refrigeration performance (cold head temperature and compressor input power) at constant operating conditions, along with measurement of the gas composition before and after pulse tube operation. This enabled us to quantify the contamination effect without being able to explore the specific internal mechanisms of ice formation and performance loss. A new quadrupole mass spectrometer was acquired to perform the composition measurements to a precision of 1 ppm across an atomic mass number range of 1 to 200. The nature of the test setup did preclude, however, the ability to sample the gas during pulse tube operation. The tests conducted in this study were of 7 to 10 days duration, usually with a follow up test of the same gas after a brief shutdown period. This is a short time scale compared to the many year lifetime requirement of space-based cryocoolers; however, a first order estimate of the mass diffusion time scales within the pulse tube of this test yielded 1 to 2 weeks, suggesting that the primary impact of initial gas contamination would be observed within the chosen time frame. This was confirmed by the experiments. The issue of longer term degradation due to outgassing within the pulse tube awaits further study. EXPERIMENTAL APPARATUS AND PROCEDURE
Pulse Tube
These experiments were conducted using a device that consisted of a TRW double-inlet orifice pulse tube driven by a Lockheed two-piston back-to-back linear compressor (Fig. 1). The
compressor had a maximum swept volume of ( per piston). The system was typically operated at 44 Hz with helium gas at a pressure of 1.8 MPa (255 psia). The entire assembly was mounted in a vacuum chamber and operated in the pressure range of less than torr. The nominal performance under these conditions was 1 W of cooling at 60 K cold tip temperature and 293 K heat rejection temperature. The power input to the compressors was
Figure 1. Schematic Layout of Pulse Tube and Compressor.
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approximately 62 W under these conditions, resulting in a nominal specific power of 62 W/W at 60 K and a Carnot efficiency of 6.3%. The cold tip temperature was measured with a Lakeshore DT 470 cryodiode connected to a Lakeshore 820 Cryogenic Thermometer readout. For some of the experiments, these data were also recorded on a portable PC computer using National Instruments LabView data acquisition software. Otherwise, the temperature data were recorded by hand. Compressor input power was measured by a Valhalla 2300L Programmable Three-phase Digital Power Analyzer. This measurement was taken between the compressor and the drive electronics and therefore does not account for drive electronic inefficiencies. The heat rejection temperature of 293 K was maintained with a water chiller located outside the vacuum chamber and connected to the fluid loop shown in Fig. 1.
The nominal test procedure was to operate the pulse tube continuously for a minimum of one week. Some of this time would be under no-load conditions, and some under a 1 W of applied load. Satisfactory operation would be verified periodically by inspection of the temperatures, power input, vacuum level, and by compressor piston waveform plots on an oscilloscope. After the first week, the pulse tube was usually turned off and allowed to heat back up to room temperature in order to reverse the effects of condensed impurities in the pulse tube. The pulse tube was then reactivated and performance monitored for a further period of hours or days. One test (G3) employed the use of a Gifford-McMahon cryopump during the second week in order to assess the possibility of condensation on the outside of the pulse tube. Further test details can be found in the Results section. Clean up of the pulse tube gas after a test was accomplished with a vacuum bakeout technique. The pulse tube was placed in an oven and heated to 50°C. Quarter-inch tubing connected the pulse tube inlet to a vacuum pump, either an Alcatel drag pump (to clean prior to Gl, G3 and G4) or a Leybold turbomolecular pump (to clean prior to G5). Typically a vacuum bakeout of 5 to 7 days was performed using up to five purges with 7 atm. (100 psi) high purity helium gas sequenced through the bakeout. Cleaning was acceptable for all cases except for G4, for which a large amount of excess nitrogen was measured after the pulse tube test. Test G2 was conducted with the original gas loaded by TRW. Mass Spectrometer Pulse tube gas compositions were measured using a Stanford Research System CIS 200 closedion source quadrupole mass spectrometer. This device can measure compositions to a precision of approximately 1 ppm across a mass range of 1 to 200 atomic mass units at an inlet pressure of torr. Accuracies are typically ± 1 ppm up to 20 ppm and ±3 % of the reading above that. A gas inlet system was designed and built to interface the high pressure pulse tube gas to the mass spectrometer inlet pressure requirements. It consisted of an intermediate 7 liter vessel into which a small gas charge from the pulse tube could be loaded. This gas was then flowed through a capillary tube into the mass spectrometer inlet. Hence, the inlet system functioned in a batch mode to achieve the torr pressure reduction. Two turbomolecular vacuum pumps were used to evacuate the system before a test and to provide the necessary differential pumping for the closed ion source operation. Three different calibration gases were used to verify proper operation of the mass spectrometer and to adjust calibration coefficients for the various gas components. These gases are listed in Tables 1a through 1c along with independent measurements made by the analytical chemistry group at JPL for gases 1 and 3. The third gas was acquired to check the nitrogen sensitivity given the vendor’s disagreement with gases 1 and 2. Calibration gas 3 confirmed the JPL measurements and therefore we concluded that the nitrogen concentrations reported by Matheson for gases 1 and 2 were slightly in error. Hydrogen gas and water vapour posed significant experimental problems for these tests. The detector in the CIS 200 is calibrated with nitrogen gas and therefore does not accurately measure the hydrogen level. Stanford Research Systems indicated in a private communication that the hydrogen signal can be in error by an order of magnitude in this mode. Unfortunately, if the
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detector is recalibrated for hydrogen, then the error shifts to the other higher mass components. We decided to leave the calibration as it was and essentially ignore the hydrogen signal in our data analysis. Since the hydrogen will not condense in the pulse tube under our operating conditions, this seemed like a reasonable approach. Water was a problem because of its ability to attach itself to the walls of the mass spectometer system. Even though the inlet plumbing of the mass spectrometer was kept at 80°C for all tests, significant amounts of water were absorbed on the walls when a new gas sample was injected into the system, thereby under-measuring the true water content. Even the use of electropolished tubing could not solve this problem. The solution was to pre-load the plumbing with the sample gas in order to accelerate equilibration between the gas and the wall. Typically this was done by flowing the pre-loaded gas for 5 minutes at a pressure of approximately 0.5 to 1 torr in the ionization chamber. The validity of this technique was verified by accurately measuring the calibration gases listed in Table 1. Nevertheless, the measurement accuracy for water was approximately ±2 ppm up to 20 ppm and ± 5 % of the reading above that, which is half as good as for the other gas species. RESULTS AND DISCUSSION
Table 2 lists all five gases that were tested in this study. They have been ordered from highest (G1) to lowest (G5) in contaminant level. Compositions were measured with the mass spectrometer both before and after running the pulse tube. Note that the “at finish” values listed in the table are averages derived from multiple mass spectrometer measurements made as the gas was removed from the pulse tube. Averages are required because the contaminant species in the pulse tube develop spatial concentration gradients during condensation, gradients that persist for some time after the pulse tube is named off. A typical mass spectrum is shown in Fig. 2 with the main peaks labeled. It corresponds to Calibration Gas 1. which is also test gas G1. The partial
pressure values in this plot have not been adjusted for the relative ionization sensitivities of the different gases. When this is done, the true partial pressure of helium rises to roughly torr while the other labelled species do not change appreciably. Therefore, a value of torr corresponds to one part per million. The time histories of all five pulse tube tests are shown in Fig. 3. The data are steady-state
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Figure 2. Mass Spectrum for Calibration Gas 1 and Test Gas G1.
cold tip temperatures for a nominal 1-W applied heater load at constant piston stroke. All tests show a temperature rise with time, which is indicative of a loss of performance due to contamination. The performance degradation is corroborated by a decrease in measured Carnot efficiency based on compressor power measurements. The temperature rise and Carnot efficiency data are presented in Table 3. Generally speaking, the different tests all start at roughly the same temperature and rise by 0.6 to 2.8 K at the end of the testing period of 7 to 10 days. This test duration was chosen because an asymptotic temperature appears to have been achieved in all cases. The most contaminated gas,
Figure 3. Time traces of all pulse tube test.
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G1, starts with a slightly higher temperature than the other cases, presumably due to immediate condensation of large amounts of and The reason for the low starting temperature of gas G4 is not clear. It was not the cleanest gas tested; therefore, either the presence of extra nitrogen somehow yielded a slightly more efficient compression process, or the measurement reflected an instrumentation offset error. The temperature increase data do, however, show a clear trend of greater rises for greater contamination levels. The clean gas test of G5 shows a distinctly different behaviour than the other cases, even though its level of contamination is only slightly greater than that of test G4. We therefore conclude that the presence of 70 ppm of nitrogen in test G4 was the cause of the difference, indicating that some nitrogen condensation occurred despite the fact that the 60 K is well above the freezing temperature of nitrogen at this concentration. Table 4 lists the freezing temperatures of the four contaminant gases at the 1, 10 and 100 ppm levels in the pulse tube. Based on this data, we would expect only and condensate to be present in the pulse tube during the 1 W at 60 K operation. The reason why the frozen nitrogen does not evaporate when the temperature is raised to 60 K is unclear. Perhaps there is some unusual surface chemistry associated with a combined and ice that retains even when the temperature rises above its freezing point. Alternatively, there could be a capillary condensation effect in the regenerator matrix that depresses the nitrogen vapor pressure. Or, perhaps the nitrogen is freezing in the region of the pulse tube downstream of the cold head where the temperature is lower than the nominal 60 K. Although nitrogen contamination is seen to be a factor, the similarity of results for tests G3 and G4, despite them having very different nitrogen concentrations, suggests that is not as strong a degrading influence as and at a given concentration. The source of gas for tests G3, G4 and G5 was commercially available “five-nines” pure helium. Concentration differences arose because of different cleanliness levels in the connecting plumbing and in the pulse tube itself. Even the cleanest gas in G5, however, did not avoid performance degradation due to contamination. This suggests that total impurity levels need to be lower than 10 ppm to completely prevent performance loss on a time scale 7 to 10 days. Note, however, that the performance decrease is relatively small. If the change in performance were due to an increase in parasitic heat load at 1 W nominal, rather than due to a decrease in second law efficiency, then the most contaminated gas (G1) equates to an increase of 40 mW after 10 days, while the least contaminated gas (G5) equates to an increase of only 12 mW. However, pulse tube warm-up tests (measuring the temperature rise of the cold head when the pulse tube was turned
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Figure 4. Time traces for cryopumped test of G4.
off) aimed at measuring this additional parasitic load did not detect any difference between the
clean and contaminated states. It is worth emphasizing that a contamination level of 100 ppm of water will form a sphere only 1.5 mm in diameter if all of it in the pulse tube were collected into a single frozen droplet. Note
that in this pulse tube approximately 85 % of the gas is on the backside of the compressor pistons, and that all of the water in this space must migrate through the piston-wall gap to form a droplet of this size. It seems somewhat surprising that such a small volume of water could produce the observed effects unless it were distributed across a thin (i.e. tens of microns) layer on a regenerator screen, for example. In such a scenario, significant blockage could occur despite the paucity of frozen material. However, our experimental set up did not allow us to measure the distribution of frozen material in the pulse tube, and therefore this issue remains unresolved. The possibility that the observed performance degradation was due to external, rather than internal, contamination was explored in an additional test on gas G4. The hypothesis is that contaminant species could condense on the cold outside surface of the device and compromise the effectiveness of the multilayer insulation. This issue was addressed by running the pulse tube with gas G4 for another week after the initial test and after first warming it back up to room temperature to drive off any frozen gas that had collected on the outside. In addition, a Gifford-McMahon cryopump was operated in the vacuum chamber for the second test to drastically deplete the partial pressure of the chamber. Measured vacuum levels were down a factor of 50 to torr. The result is shown in Fig. 4, where the second cryopumped curve is shown superimposed on the first nominal curve. Clearly the cryopump made no appreciable difference. The second curve rises more quickly to the asymptotic temperature, but this would be expected because the internal contaminants could not completely disperse throughout the pulse tube in the one day that it was inactive at room temperature. Past studies4,5 have indicated that multilayer insulation performance starts to degrade when a thickness of roughly 0.1 microns of or ice forms on it. In a vacuum system, such a thickness represents a vast amount of material, on the order of many hundreds of molecular monolayers. Simple first order analyses indicate that such large amounts of water are inconsistent with the observed vacuum chamber pressures of less than and the capabilities of our pumping system. This, combined with the cryopump test results lead us to the conclusion that the observed performance degradation was due to internal, not external contamination. For all gases except G2 it was observed that any brief shutdown and warm-up to room temperature of the pulse tube had the effect of restoring the pulse tube to its original performance. However, this original performance tended to be relatively short-lived, as seen in Fig. 4 for gas
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G4. We hypothesize that the condensed gases in the pulse tube diffuse away from their frozen
location during the few hours of shutdown, but do not get very far away. Therefore, upon reactivation of the pulse tube, it takes significantly less time for this gas to migrate to the cold, condensing part of the pulse tube. However, this behaviour was not observed for gas G4. This gas showed no improvement in performance after a two day shutdown, but did show almost complete recovery of the original performance after a six day shutdown. The reason for this difference in performance is not known, although it is interesting to note that this gas was the original gas loaded by TRW. As had been originally requested, they did not bake out the pulse tube prior to loading of this gas. Moreover, the gas sat in the pulse tube for almost two years before this study was conducted. CONCLUSIONS
Five different gases were tested in a TRW/Lockheed pulse tube to determine the effects of contamination in otherwise pure helium. Total contamination levels ranged from approximately 20 to 420 ppm of water, carbon dioxide, nitrogen and argon. No other contaminants were present above a level of 1ppm except hydrogen, but due to instrumentation restrictions the hydrogen levels were not accurately measured. A decrease in pulse tube refrigeration performance was observed in all tests, with asymptotic steady-state temperatures reached on a time scale of 7 to 10 days. Under a 1-W applied load condition with a constant piston stroke, the total cold head temperature rise due to contamination ranged from 0.6 to 2.8 K depending on contamination level. Comparisons between cryopumped and non-cryopumped vacuum chambers appeared to rule out the possibility that the contamination was external to the pulse tube. Water and carbon dioxide were the most important contaminants at the nominal operating temperature of 60 K, although there was evidence for nitrogen condensation under some conditions. Warming of the pulse tube to room temperature tended to restore the original performance, although this improvement degraded more rapidly the second time around and returned to the asymptotic steady-state condition. ACKNOWLEDGMENTS
The work described in this paper was carried out by the Jet Propulsion Laboratory, California Institute of Technology, and was sponsored by the NASA EOS IMAS TechDemo Project through an agreement with the National Aeronautics and Space Administration. The authors would like to acknowledge the laboratory assistance of Scott Leland and Gary Plett, and technical discussions with Dr. Simon Collins, all at JPL. REFERENCES 1. Wade, L. , Donnelly, C., Joham, E., Johnson, K., and Phillips, R. “An Investigation Into The Mechanics of Joule-Thompson Valve Plug Formation,” Adv. Cryo. Eng. 33, 1988, pp. 699-706. 2. Fieldhouse, I. B. and Porter, R. W. “Cryogenic Cooling of Infrared Electronics,” GACIAC SOAR-86-
02, IIT Research Institute, Chicago, Ill, May 1986. 3. Cranmer, D. C., Watt, E. J., To, G. A., Marquex, N. and Adams, P. M. “Analysis of Component Contamination from Accelerated Contamination Testing of Five-Year Vuilleumier Cryocoolers,” TR-
0086(6057)-2 (SDTR-86-100), Aerospace Corp., Feb. 1987.
4. Cunningham, T.M. and Young, R.L., “The Radiative Properties of Cryodeposits at 77 K,” Adv. Cryo. Eng., Vol. 8, 1963, pp. 85-92. 5. Caren, R.P., Gilcrest, A.S. and Zierman, C.A., “Thermal Absorptances of Cryodeposits for Solar and 290K Blackbody Sources,” Adv. Cryo. Eng., Vol. 9, 1964, pp. 457-463. 6. O’Hanlon, J.F., A User’s Guide to Vacuum Technology, John Wiley & Sons, N. Y., 1980.
Simple Two-Dimensional Corrections for One-Dimensional Pulse Tube Models
J. M. Lee and P. Kittel NASA Ames Research Center Moffett Field, CA 94035 K. D. Timmerhaus University of Colorado, Boulder Boulder, CO 80303 R. Radebaugh National Institute for Standards and Technology Boulder, CO 80303
ABSTRACT One-dimensional oscillating flow models are very useful for designing pulse tubes. They are simple to use, not computationally intensive, and the physical relationship between temperature, pressure and mass flow are easy to understand when described with phasor diagrams. They do not possess, however, the ability to directly calculate thermal and momentum diffusion in the direction transverse to the oscillating flow. To account for transverse diffusion effects, either parameter corrections must be obtained through experimentation, or solutions to the twodimensional differential fluid equations must be found. A linearized two-dimensional solution to the fluid equations has been obtained. The solution provides lumped parameter corrections for one-dimensional models. The model accounts for heat transfer and shear flow between the gas and the tube wall. The complex Nusselt number and complex shear wall are useful in describing these corrections, with phase relations and amplitudes scaled with the Prandtl and Valensi numbers. The calculated enthalpy flow ratio, between a two-dimensional solution of the oscillating temperature and velocity and a one-dimensional solution for the same shows scales linearly with Va for Va < 30. In this region that is, the enthalpy flow calculated with a two-dimensional model is 50% of a calculation using a one-dimensional model. For Va > 250, showing that diffusion is still important even when it is confined to near the tube wall.
Cryocoolers 10, edited by R. G. Ross, Jr. Kluwer Academic/Plenum Publishers, 1999
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INTRODUCTION A previous paper1 that examined the scaling parameters for pulse tubes is based on a twodimensional analysis of anelastic oscillating and compressible low Mach number flow of a gas contained in a tube of thin but finite wall thickness.2 Anelastic flows are characterized by low Mach numbers and oscillating frequencies much less than system resonance frequency. This approximation is appropriate when acoustic and shock wave energies are negligible relative to the energy needed to compress and expand the bulk gas.3 For a tube radius and a tube wall thickness of and smaller than the tube length, respectively, and for a tube with z axial coordinate scaled from 0 to 1, where the cold end is at the scaling reduces the problem to 8 dimensionless groups. These groups are shown in Table 1. The parameters and determine anelasticity. For an anelastic approximation holds. Linearization of the fluid equations applies with the added constraint The superscript starred ‘*’ quantities in Table 1 are dimensional quantities. The tilde ‘~’ quantities are complex and result from using complex embedding,
for the linear solutions where is the real part of the bracketed complex quantity, is the amplitude, and is the phase angle. The velocity phase angle between the locations and is which is scaled from 0 to 1 (corresponding to 0° to 360°). The phase angle of velocity at is taken to be zero. The velocity amplitude ratio is given by Three of the dimensionless groups are relevant to transverse (radial) diffusion: the Valensi number, Va, is the squared ratio of tube radius to viscous diffusion length; the Prandtl number, Pr, is the squared ratio of viscous diffusion to gas thermal diffusion lengths; and the Fourier number of the tube wall, Fo, is the squared ratio of thermal diffusion length in the wall to the tube wall thickness. This paper explores the use of Va, Pr and Fo in providing lumped-parameter corrections of transverse diffusion for one-dimension models.
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RESULTS OF MODEL
How Thermal Diffusion Affects Bulk Pressure and Temperature Phasors Diffusion shifts the phase and amplitude of local temperature and velocity within the diffusion layer near solid boundaries, such as near the tube wall. This can be important for the transversely integrated bulk oscillating temperature, mass flow, and pressure typically used for one-dimensional models. Figure 1 shows how thermal diffusion in the gas and tube wall affect the bulk pressure and temperature phasors, and respectively, for and 100; and and 30 and 100. The bulk temperature phasor is the integrated oscillating temperature profile over the tube radius, is the normalized velocity amplitude at and is the bulk oscillating pressure.
Figure 1. Column 1 shows the effect of Va, and Fo on bulk pressure, and bulk temperature, phasors for The reference velocity phasor is along the real axis with unit amplitude. Column 2 shows the oscillating temperature, Temperature profiles for are pinned at for Fo = 100 (identified with '*') float at
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The condition for an isothermal wall boundary is (thick wall relative to thermal penetration). The condition for an adiabatic wall is (thin wall relative to thermal penetration). For the tube inner radius is much larger than the thermal penetration distance in the gas. The shaded areas of the phasor diagrams in column 1 indicate the
corresponding velocity phasors at locations in the tube between and Figure 1, column 1, shows that for an adiabatic wall condition temperature and pressure phasors are in-phase as would be expected. The thermal penetration distance within the tube wall is much larger than the tube wall itself, thus there is no thermal time lag in the tube wall. The tube wall temperature closely follows the oscillating temperature; likewise, the tube wall does little to constrain the oscillating gas temperature amplitude. This can be clearly seen from the oscillating gas radial temperature profiles given in column 2, where the curves designated with a “*” are for As seen in column 2, the gas temperature is not “pinned” to zero at (the gas/tube wall interface), but instead “floats”, thereby having less effects due to thermal diffusion. Often for one-dimensional models it is assumed that adiabatic conditions on the gas are present This condition is not likely achieved for real for pulse tubes, however. For example, a stainless steel tube with a wall thickness of 0.01 cm and thermal diffusivity of filled with gas oscillating at 30 Hz will have A non-adiabatic condition in the gas can have profound effects on the gas temperature phasor. For the isothermal condition the oscillating bulk temperature, lags the oscillating pressure by about 15°. This is illustrated in Fig. 1, column 1. More importantly, the temperature phasors are shifted out-of-phase relative to the velocity phasors (shaded area). This reduces enthalpy flow, since enthalpy flow depends on the cosine of the phase angle between velocity and temperature. Even at large where the thermal diffusion layer of the gas is thin, an approach to adiabatic conditions might be expected. However, there is still a significant detrimental phase shift for out of the shaded velocity phasor area, that is, even for large Va, is never really in-phase with the gas velocities. The phasor diagrams given in column 1 would indicate that most pulse tubes probably do not operate as ideal adiabatic systems. Column 2 of Fig. 1 shows the corresponding oscillating temperature for isothermal wall conditions. Here, diffusion has a large effect on oscillating gas temperature. For small (or ), dominant diffusion tends to dampen the temperature amplitude of the gas. As Va increases, the dampening of the gas temperature lessens. However, diffusion still constrains the temperature oscillations near the wall. And since the area averaged bulk temperature scales with the square of the radius, this still constitutes a significant effect, even at large Va. A useful relation quantifying the relation between bulk temperature and heat transfer to the tube wall is contained in the complex Nusselt number,
where is the complex heat transfer to the tube wall and is the complex temperature of the 4 tube wall. Kornhauser and Smith have examined this for basic pulse tube (BPT) systems (velocity at is zero) and rectangular coordinates. For orifice pulse tube (OPT) systems (velocity at is non-zero) the complex Nusselt number is the same as for BPT-systems. This is seen in Fig. 2 where the phasor, is the same for both the BPT and OPT. The conclusion is that does not depend on nor thus the relation for proposed by
SIMPLE 2-D CORRECTIONS TO 1-D PULSE TUBE MODELS
Figure 2.
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Effect of PrVa and Fo on heat transfer amplitude and phase. Reference velocity phasor along real axis with unit amplitude.
Kornhauser for BPT-type systems for rectangular geometries can be used for similar OPT systems. At small PrVa, heat transfer correlations for steady flow in a tube can be used since the dominance of diffusion constrains and to no phase shift. At larger PrVa, deviations occur resulting in a phase shift of away from steady-state correlations. Figure 3 plots the amplitude and phase of for increasing PrVa. The plot can be used for corrections to one-dimensional models with the form for thermal modeling, and the form for shear (friction). These relations are useful when used in onedimensional relations such as for heat transport.
for momentum transport, and
Enthalpy flow comparisons
Normalized Enthalpy flux is given by the time averaged product of oscillating temperature and oscillating velocity. The flux integrated over the tube cross-sectional area gives the normalized enthalpy flow. Normalization is calculated using the enthalpy flow at the mean temperature (not bulk oscillating temperature). The solutions obtained from ref. 2 are twodimensional and so the enthalpy transport calculations given in ref. 2 contain the effects of both temperature and velocity diffusion.
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Figure 3.
Amplitude and phase of the complex Nusselt number, or complex wall shear factor,
Figure 4. The effect of heat transfer on enthalpy flow for and Column 1 is the enthalpy flux and column 2 is the enthalpy flow.
SIMPLE 2-D CORRECTIONS TO 1-D PULSE TUBE MODELS
Figure 5.
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Enthalpy flow correction factor, as a function of Va. The correction factor is the ratio between the 2D enthalpy flow calculations of ref. 2 and
that calculated for 1D flow using ref. 4, Eq. 2-38.
Figure 4 plots enthalpy transport for and (isothermal wall condition). The plots of in column 1 are shown for 0.5 and 1. For the case of and enthalpy flows in the reverse direction between the locations to This is also shown in the corresponding plot of enthalpy flow vs. z in column 2, where the enthalpy flow is negative. Also, the magnitude of the enthalpy transport is very small due to damping of the gas temperature. These plots show that operating with isothermal walls and small PrVa is not desirable for a pulse tube. The pulse tube should be operated where the thermal diffusion region is confined to a thin layer near the tube wall. This can be seen in column 1 for Enthalpy flux is positive throughout the tube, and the core is not dampened by diffusion at the tube wall. Also, there is an overshoot between the core and the diffusion layer. This corresponds to the oscillating temperature overshoot in Fig. 1. The oscillating velocity also exhibits an overshoot. Column 2 of Fig. 4, which plots enthalpy flow, shows how an adiabatic wall condition significantly increases enthalpy flow over an isothermal wall. The dashed-line is the near adiabatic condition of and the solid line is the isothermal condition Column 2 clearly shows that there are advantages of having large PrVa and large Fo. The steep slope in the isothermal condition indicates that enthalpy flow is decreasing along z, and that enthalpy flow is continuously being converted to heat flow. For the case of the present model, heat flows to a constant temperature sink. For pulse tubes with an axial temperature gradient, it would be converted to heat flow down the temperature gradient. This can be a significant loss, as shown by the plots of column 2. Figure 5 plots a correction factor, α, vs. Va for use in correcting one-dimensional calculations of ideal adiabatic enthalpy flow in pulse tubes. The correction factor is the ratio between the calculated enthalpy flow given by ref. 2 and the one-dimensional relation for calculating enthalpy flow given by Storch5 in Eq. 2-38. The plot of Fig. 5 is applicable for to and to The correction does not exhibit a strong dependency on Pr and within this range. Figure 5 is useful for quickly correcting adiabatic one-dimensional pulse tube models to account for transverse diffusion effects. DISCUSSION The calculated oscillating quantities for pressure, temperature, velocity and heat transfer, and the time-averaged enthalpy flow from a two-dimensional continuum theory gives an insightful understanding of the transport mechanisms for pulse tubes.
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The orifice pulse tube (OPT) is a cooling device that does not rely on diffusion to obtain the appropriate phase angles between velocity and temperature in order to produce enthalpy flow. The appropriate phase angles are obtained through the velocity boundary conditions at the tube ends. Thus, the OPT should be operated where the thermal diffusion region is confined to a thin layer near the tube wall so as to minimize the effects due to viscous and thermal diffusion. This condition requires Va and Fo to both be large. There is a practical limitation to having both PrVa and Fo very large, as these requirements lead to a system that must contain high pressures with a large diameter, thin-walled tube. Operating at small Va and small Fo is detrimental to an OPT because heat transfer between the gas and the tube wall: i) reduces the oscillating temperature amplitude near the tube wall, and ii) creates unwanted phase angles between velocity and temperature. Both of these effects will tend to reduce enthalpy flow. For small Va it can even reverse enthalpy flow. Heat transfer between the gas and the tube wall has an important effect on the pressure and temperature phasors. When there is significant heat transfer between the gas and tube wall, that is, for the pressure and temperature phasors move out-of-phase relative to each
other. Calculations indicate this to be as much as 20°. This is important because onedimensional models often assume adiabatic conditions on the gas and so there is a presumption that the temperature is always in-phase with pressure. Most pulse tubes operate at which is closer to isothermal wall conditions. The complex Nusselt number is found to be independent of Fo, velocity amplitude ratio and velocity phase angle at the tube ends. Thus, the relations of Kornhauser for BPTsystems with rectangular geometries can be used for similar orifice pulse tube geometries (though not typical). For cylindrical geometries, the plot of Fig. 3 can be used. When written in the form is about 4 for PrVa < 3 and is linear with PrVa for PrVa > 25. The phase angle for PrVa < 0.5 is and for PrVa > 500, The same relation for the complex shear wall factor exists, as function of only Va. The complex Nusselt number and shear wall factor can be used for one-dimensional linear oscillating flow in a tube to account for radial heat transfer or shear at the tube wall. The plot of Fig. 5 can be used to correct enthalpy flow calculations from adiabatic, one-dimensional models to account for transverse diffusion. REFERENCES
1. Lee, J. M., Kittel, P., Timmerhaus, K. D., and Radebaugh, R., Useful scaling parameters for
the pulse tube, Adv Cryo. Eng., Plenum Press, vol. 41B (1995), pp. 1347-1356. 2. Lee, J. M., “Steady Secondary Flows Generated by Periodic Compression and Expansion of an Ideal Gas in A Pulse Tube”, Ph. D. Thesis, University of Colorado, Boulder (1997)
3. Sherman, F. S. Viscous Flow, McGraw Hill, (1993), p. 82 4.
Kornhauser, A. A. and Smith, J. L., Application of a complex Nusselt number to heat transfer during compression and expansion, J. Heat Transfer, no. 116 (1994), 536-542.
5. Storch, P. J., Radebaugh, R. and Zimmerman, J. E. 1990 Analytical model for the refrigeration power of the orifice pulse tube refrigerator, NIST Tech Note 1343, National Institute for Standards and Technology, Boulder, CO.
Pulse Tube Development Using Harmonic Simulations H.W.G. Hooijkaas* and A.A.J. Benschop Signaal-USFA, Meerenakkerweg 1, PO-box 6034 5600 HA, Eindhoven, The Netherlands
*Eindhoven University of Technology, PO-box 513 5600 MB Eindhoven, The Netherlands
ABSTRACT Based on the experiences with our in-house simulation model for Stirling coolers, we are developing a similar model for pulse tube coolers. This model is built on first order harmonic approximations and time averaging of the energy flows. It predicts the pressure waves, the temperatures and the cooling power with its associated loss contributions. The main benefit of
this simulation model is the combination of short calculation times and sufficient accurate
predictions., which is established by means of analytical simplification. Simulations of a particular cooler are obtained from solving the model’s equations iteratively. The iterations are initiated with a chosen initial temperature profile throughout the whole system, which is used for simultaneous calculations of mass flows and pressure waves.
With the results of these calculations the average temperatures and enthalpy flows are computed. These subsequent calculation steps are repeated until sufficient stability is achieved. At that point the properties of the particular pulse tube cooler in its stationary operational mode are known. The simulation model is experimentally verified. Predictions of cooling powers, mass flows, pressure waves and temperature profiles have been checked. This has given promising results, although certain aspects of the model may need more attention and improvement. INTRODUCTION
Signaal-USFA and Cryotechnologies manufacture low power cryocoolers of the Stirling type, which have cooling capacities within the range of 0.25W to 2W at 77K cold tip temperature. Based on their experiences with the development of Stirling coolers, they are nowadays working on the development of pulse tube coolers within the same performance range. In developing this type of coolers the delicate, optimal combination of enthalpy flows contributing to the cooling effect and the inevitable accompanying loss terms must be determined. The relation of the desired enthalpy flows and the occurring loss terms is strong but diffuse, so an accurate model of the hydrodynamics and the thermodynamics in the system can be of great help in pulse tube development A lot of research on this kind of model has been carried out (e.g. Kittel et al. 1 and de Waele et al.2), generally yielding a set of non linear differential equations which can be solved
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numerically. However, solving these equations is very time consuming which makes these models an inappropriate development tool. This implies that, in order to obtain an adequate development tool, these models should be simplified without too much loss of accuracy. A widely used simplification method is the harmonic approach(e.g. Swift et al.3, Kittel et al. 1 and de Waele et al.4). In this method the pressure wave is assumed to be harmonic and have small amplitudes (less than 10% of the mean value), which makes analytical elimination of the model’s time dependency possible. The model resulting after this elimination has solely position dependent differential equations, which is a great benefit for the needed calculation times. Our in-house simulation program is built on these simplified equations. One of the main benefits of the program is its capability to perform rapid optimizations on dimensions and operational conditions of the pulse tube refrigerator, providing useful information in the pulse tube development. Experiments have been done in order to check the validity of our model in its present status. For small pressure fluctuations the predictions of our model show to be in fair agreement with the experiments. PULSE TUBE SIMULATION MODEL
Probably the most important feature of a pulse tube refrigerator is its cooling power. The ‘cooling power’ is the total amount of heat that can be removed from the surroundings at the desired cold temperature and it is the result of both productive, heat removing mechanisms and loss contributions within the system. In general, for a pulse tube cooler operating under normal conditions, these productive and loss terms are within the same order of magnitude, while the resulting cooling power is much smaller. This implies that an accurate and detailed simulation model is required to obtain good estimations on the pulse tube performance. Our present day simulation model is based on the behavior of the single orifice pulse tube refrigerator, as most of the other pulse tube configurations can be derived from this particular type by means of a model extension or restriction. In building the model of the orifice pulse tube refrigerator, it is first divided into several basic parts, such as the regenerator, the actual pulse tube and the heat exchangers. First a mathematical description for each of these elements is made, after that all parts are combined into a pulse tube cooler. Fundamental relations For all parts of the refrigerator the description of the hydrodynamics and the thermodynamics of the system is based on the conservation principles for mass, momentum and energy, de Waele et al. 2 . Next to these three principles, the state of the non-ideal gas is described with a generalized form of the gas law. Conservation of mass reads:
with
the density of the gas and u the flow speed of the gas. Next we have the equation of motion, given by:
where A is the cross sectional surface, perpendicular to the flow, m-dot is the mass flow and is a flow friction factor. This latter factor depends on the system’s geometry and the speed, viscosity and density of the gas. Each of the terms in this equation has its own physical meaning. The first term represents
the change of pressure forces acting on the gas, while the second one stands for the dragging forces on the gas and the third term accounts for the changes of the kinetic energy of the gas flow. Finally, the fourth term is the resulting change in momentum that is caused by the three effects stated before. In general only the contributions from pressure drop and dragging forces play a role of significance in the miniature pulse tube.
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The thermodynamic behavior can be formulated by the first law of thermodynamics, stating energy conservation for the considered system. Applying this law to the gas inside a control volume with fixed boundaries in the refrigerator gives:
where is the coefficient of heat exchange per unit of exchanging surface and F is the heat exchanging surface per unit of length and h is the enthalpy per unit of mass. According to Fourier’s conduction law, is the coefficient of thermal conduction. The left side term of this equation represents the change of enthalpy in the control volume. The first term on the right hand accounts for the heat exchange with the surroundings and the second one represents the heat flow due to conduction. The third term represents the net contribution of enthalpy flows into and out of the volume. Finally, the fourth term accounts for the changes in enthalpy due to pressure variations, which equals the work done on the gas in case of mechanical reversible compression. Besides in the above derived energy balance of the gas, energy conservation is also considered for the materials that are in thermal contact with the gas. This yields a much simpler energy balance, which reads:
Here the changes in mass, pressure and enthalpy flow contributions are absent. Further it must be noticed that the contribution of heat exchange with the gas, as it appears as first term on the right side, is reversed with respect to equation 3. The fifth and last basic relation in the system description is the generalized gas law, which takes the non-ideal gas behavior into account, Fokker5. This relation reads: where Y is a polynomial function of the density and pressure, which is determined empirically. With this relation we can eliminate the density as an independent variable of state. So the behavior of the system is completely determined by the mass flow, the pressure and the temperatures of the gas and the surroundings. These variables are all related by means of Eqs. (1), (2), (3) and (4), which form a set of four coupled differential equations. This set can be solved numerically when the appropriate boundary and initial conditions are imposed, but the data on the factors of flow friction, the coefficients of heat exchange and the properties of the gas and the other materials should be available from empirical relations or data sheets. However, solving these time and position dependent equations is very time consuming. Simulating a single orifice pulse tube with this model requires typically about half an hour. This implies that the model is not very appropriate for pulse tube optimizations, as for this purpose several hundreds of calculations may be needed. This is especially the case for the more sophisticated types of pulse tube refrigerators. Applying the simulation model for development purposes successfully, requires much faster calculations on the performance of one single pulse tube configuration. One way to achieve this acceleration is simplifying the model. However, it is desirable that the accuracy of the model is not degraded. Harmonic approximation The introduction of a harmonic solution is a widely and successfully applied simplification method. This approximation results in a enormous reduction of the required calculation steps. However, the type of pressure waves which can be simulated successfully are restricted. They must be sine-shaped and have relatively small amplitudes i.e. pressure ratio These harmonic pressure waves can be represented by:
where is the mean pressure and the pressure wave amplitude. is the phase angle of the pressure wave relative to a reference phase. We assume that the temperature oscillations and
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mass flow are also harmonic as they are caused by the pressure wave. The temperature is now
given by: and for the mass flow we have: where for convenience it is assumed that there is no average mass flow in the system.
For these purely harmonic time dependent functions, a differentiation with respect to time results in a multiplication by and an additional phase lag of 90 degrees. As the remaining relations should be valid at any time, the orthogonal terms of sine and cosine as well as the terms with different angular frequencies in the argument can be separated. Because the amplitudes of the input are small, the harmonic terms of second or higher order can be neglected. Yielding a set of relations, which have in common that each of those relations only has identical time dependent terms. So time dependence can be omitted in solving these equations,. Time averaged contributions to the cooling power follow directly from the DC-terms in enthalpy- and heat-flowexpressions. The resulting model can be solved much faster. Because it only deals with harmonic solutions which solely satisfy the homogeneous differential equations, this model is inappropriate for simulations of the cool-down of a pulse tube. Despite this deficiency, the harmonic approximation yields a powerful tool for pulse tube development, as the calculations of
the orifice pulse tube configuration now only take a fraction of a second and accuracy is maintained for the restricted class of pressure waves. The working area of the model must be identified experimentally for oscillations that do not completely fulfill the restrictions. Calculation scheme As stated previously, the model of the single orifice pulse tube is used as the basis for the final simulation program. Models of more sophisticated pulse tube coolers, such as multi-stage and double- or multi-inlet systems, are derived from extending the basic model. In the simulation program the resulting models are finally integrated in an optimization routine. This latter routine enables an automated search for optimal dimensions and operational conditions for the system.
So the simulation program is structured of several connected data processing routines. These data processing blocks and the occurring data flows are visualized in figure 1. The core of this structure is the so-called OPTR-loop, the iteration loop where the orifice pulse tube is simulated.
Figure 1. Block diagram of the optimization program for a double inlet pulse tube.
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Figure 2. Schematic view of the in-line pulse tube and the measurement features.
In this loop the pressure, mass flow and temperatures in the orifice pulse tube are derived. In each iteration step the non-ideal gas properties are calculated at the newly found pressures and temperatures of the gas. After the iteration in this loop have met the stability criterions, the program enters the second loop, which is the DIPTR-loop. In this loop the mass flow through the bypass is estimated first and after that the behavior of the single orifice pulse tube including this bypass flow is calculated. When these iterations meet the stability criterion, the various miscellaneous loss contributions such as conduction through the tube walls are taken into account. In case of an optimization particular given input parameters are changed, after which the program simulates this new configuration. This entire procedure will be repeated until the optimum system and its operation conditions are found. EXPERIMENTAL SET-UP
Two similar experimental single orifice pulse tube refrigerators were constructed for the validation of our simulation model. One of these pulse tubes is equipped with a set of thermocouples inside the cooler. The coolers are of the single orifice type, as this cooler is the bases in our simulations. However, the design allows easy modification into a double inlet system. The pulse tube coolers are driven by a linear compressor which has operational frequencies that vary from 20 to 60Hz. A schematic view of these single orifice pulse tubes is given in figure 2. The orifice is an adjustable needle valve and the heat exchangers are all of the slit-type and are made of copper. At the cold spot both a heater and a temperature sensor are mounted. Both coolers have
facilities for pressure measurements in the split-pipe, in the buffer volume and at the hot end just before the orifice. It is also possible to measure the position of the piston in the compressor. The most important overall dimensions of the coolers are summarized in table 1.
As mentioned before, a number of thermocouples are mounted in the second pulse tube. In total there are 22 Copper-Constantaan thermocouples, which have a wire diameter of 0.2mm. They are placed almost equidistantial in both the pulse tube and the regenerator. These thermocouples give the possibility of measuring the gas temperature in the actual pulse tube and the regenerator. The axial positions of the thermocouples are given in table 2. These positions are relative to the boundary surfaces between the cold heat exchanger and the regenerator and the pulse tube respectively. The dimensions of both pulse tubes are chosen equal, as this enables a judgement of the disturbing influences caused by the thermocouples.
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RESULTS
Several experiments have been done with the described pulse tubes. A distinction between both coolers was made in analyzing the results, since the disturbances on the hydrodynamics and the thermodynamics of the system arising from the presence of the thermocouples may be considerable. Therefore, the pulse tube with the thermocouples was used mainly for temperature profile measurements, while the other one was used for the measurements of the cooling powers and pressure waves. However, the impact of the thermocouples can be analyzed roughly from the differences in cooling power and pressure waves under equal operational conditions. The simulation program does not take the changes in hydrodynamic behavior into account, but it does take the heat flows through the connecting wires into account. However, care must be taken that this heat flow does not dominate the thermodynamics in the system.
Pressure waves and cooling powers The pressure waves and cooling powers are measured on the pulse tube without the thermocouples. They are measured at several combinations of filling pressure, drive frequency and flow impedance of the orifice. The input conditions of the simulations are matched with the pressure wave in the split pipe. Most of these experiments are performed at low input powers
Flow impedance dependency. The results for various flow impedancies of the orifice are depicted in figure 3. In these experiments the pulse tube was operated at 32.5Hz, 10bar filling
pressure and 140K cold tip temperature. The simulations of the pressure waves in both the actual pulse tube and the buffer show a good agreement with the ones measured, as can be seen from
this figure. The simulations and measurements on the cooling powers give profiles which have similar trends for both. However, between simulations and experiments a difference of about 350mW is found. This is partially caused by the dewar losses which are not taken into account.
Figure 3. Simulations and experiments on cooling power at 140K and pressure waves for the single orifice pulse tube as function of the flow impedance of the first orifice.
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Figure 4. Measured cooling power and difference compared to the simulations as function of the frequency. System operating at 140K and 10W electrical input power.
Frequency dependency. Other series of experiments are performed with varying frequencies, a cold temperature of 140K and a filling pressure of 20bar. These measurements are repeated for flow impedancies of respectively 32.6, 25.3 and The results are
shown in figure 4, where both the measured cooling powers as well as the difference between simulations and measurements are depicted. It can be seen that a change in orifice setting causes a shift of the optimal frequency with respect to the cooling power. It should also be noticed that
the error in the simulation shows a dependency of the operating frequency. This error decreases with the frequency for frequencies below the optimal value of the system. Input power dependency. Finally, experiments were performed with varying input power at a filling pressure of 29bar and a cold temperature of 155K. Both the measurements and simulations of the cooling power and pressure waves are presented in figure 5. In this case too, the pressure waves in the split pipe are used in triggering the simulations. The simulation errors of both the cooling power and pressure waves increase with increasing input powers. Although it is hard to see in this figure, the simulation errors of the pressure wave in the buffer show the same tendency.
Figure 5. Experiments and simulations as function if the input power. Tc=155K.
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Figure 6. Temperature profiles over the actual pulse tube at several frequencies and at an orifice impedance of
Temperature profiles
The gas temperatures in the pulse tube are measured with the other cooler. As expected, this cooler has less performance than the first cooler. The cooling performance of this system is about 800mW lower at a low temperature of 160K, a mechanical input power of 8W, a filling pressure of 10bar, a frequency of 32.5Hz and an optimal orifice setting of approximately Only about one third of this decrease in performance rises from the heat conduction through the thermocouples, as this conduction is less than 250mW, implying that the differences between both pulse tubes and the impact of the enlargement of regenerator imperfections and the flow disturbances in the actual pulse tube are not negligible. However, it is hard to make a distinction between the effects of these three causes. The pressure drop over the regenerator and the actual pulse tube deviates less than 5% compared to the other system, which lies within the accuracy of the measurements.
The measurements and simulations of the gas temperatures in the actual pulse tube are given in figure 6. The low temperature 32.5Hz was 160K and at 20 and 25Hz it was 150K, but the mechanical input power was 10W in these cases. Unfortunately, some thermocouples were broken. The thermocouples at 21mm and 28mm from the orifice side short circuited and the one at 49mm was broken. Hence the temperatures at these positions are unknown. From Fig. 6 it can be seen that the simulation model has two major lacks. At the positions close to the orifice the temperature profile is much flatter than predicted. This is caused by gas that flows from the buffer, through the orifice and into the pulse tube. This gas is approximately at the temperature of the buffer and penetrates 5 to 10mm into the tube and this effect increases indeed for lower orifice impedancies. The other striking difference between measurements and simulations occurs at the cold side. At this point the real gas temperature is significant lower than predicted. This is means that the heat transfer at the cold end is less than calculated. These two effects cause both an increase of the loss effects accompanying the gas flow through the tube, which can be noticed from the steeper gradient in the middle part of the pulse tube. The effect is a loss in net cooling power of the pulse tube in comparison to the simulations. DISCUSSION
A harmonic simulation program for pulse tube refrigerators is being developed at SignaalUSFA. The present day status of this model is discussed and its validity is verified in this article. Applying harmonic approximations on the equations describing the pulse tube behavior yields a simulation model that can be solved several orders of magnitude faster than the model
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without approximations. The main benefit of this particular approximation method is that it only causes a minor loss of accuracy for harmonic pressure waves with small amplitudes. These two properties make the harmonic simulation model suitable as a pulse tube development tool. Validation of the simulation model shows that it rather well describes both the
hydrodynamics of the pulse tube and the trends in the relations between the cooling power and both the frequency and the orifice impedance. However, there is a significant difference between simulated and measured cooling powers for all operational conditions. This difference seems to depend on the combination of the frequency and the orifice impedance (Figures 3 and 4). It is also shown that the accuracy of the model decreases with increasing input powers. This is directly the effect of the harmonic approximations and its restriction on the input signals. From the prediction of the temperature profiles it can be seen that the thermodynamics in the model should be improved for two main items. The first one is the boundary effect at the
orifice side, where be the influence of the gas at room temperature, which flows from the buffer into the pulse tube, should be emphasized more. The second point is the calculation of the heat transfer in the heat exchangers, from which the cold one is the most important. These two improvements will cause the simulation model to predict lower cooling powers and therefore explain part of the difference between measurements and simulations. However, it must be checked whether these two shortcomings can count for the dependency of the simulation error
and the combination of frequency and orifice impedance. ACKNOWLEDGEMENTS
This work is carried out as part of a research cooperation between Signaal-USFA and the Eindhoven University of Technology. Signaal-USFA and its affiliate Cryotechnologies, together form the SBU Cryogeny of Thomson-CSF. REFERENCES
1.
Kittel, P, Kashani, A, Lee, J.M. and Roach, P.R., “General pulse tube theory”, Cryogenics vol. 36 (1996), pp. 849-857.
2.
de Waele, A.T.A.M., ten Steijaart, P.P. and Gijzen, J., “Thermodynamical aspects of pulse tubes”, Cryogenics vol. 37 (1997), pp. 313-324.
3.
Swift, G.W. and Ward, W.C., “Simple Harmonic Analysis of Regenerators”, Journal of Thermophysics and Heat Transfer, vol. 10, No. 4 (I996),pp. 652-661.
4.
de Waele, A.T.A.M., Hooijkaas, H.W.G., ten Steijaart, P.P. and Benschop, A.A.J., “Regenerator Dynamics”, Article to be presented at ICEC (1998).
5.
Fokker, H., “Analytical Program Complex of a Stirling Engine”, Internal Report, Philips Nat,Lab. (1969)
Analysis of a Two Stage Pulse Tube Cooler by Modeling with Thermoacoustic Theory A. Hofmann* and S. Wild**,1 *Forschungszentrum Karlsruhe, Institut für Technische Physik D76021 Karlsruhe, Germany **Universität Karlsruhe, Institut für Technische Thermodynamik und Kältetechnik 1 Now with AEG Infrarot-Module GmbH, Heilbronn
ABSTRACT / INTRODUCTION
A two stage pulse tube cooler operated with magnetic valves has been tested. The no-load temperature is close to 6 K, and with powers of 4 W and 15 W at the first and second stage, the respective temperatures are 20 K and about 60 K. Additionally, temperatures and pressures have been measured at different positions. This system is being analysed by calculations based on a linearized thermoacoustic theory. Reasonable agreement is obtained when some parameters are fitted. Moreover, the calculations give valuable information on volume flow, pressure drop, phase angles, energy flows, and temperatures at all positions of the system. Those results will be discussed with respect to further optimisation of such systems. EXPERIMENT
The scheme of the two-stage pulse tube refrigerator, originally designed for 20 K operation, is shown in Fig. 1a). The first stage regenerator is made from 200 mesh (0.056 mm wire diameter and 67 % porosity) stainless steel sheets stacked in a 150 mm long stainless steel tube with 35 mm i. d. and 1 mm wall thickness. The first stage pulse tube is 200 mm long with 24 mm i. d. also with 1 mm wall thickness. Heat exchangers (flow conditioners) made from 15 mm long stack of 80 mesh Cu are at both ends. The second stage regenerator has two components, 0.2 mm diameter Pb spheres filled to a length of 100 mm into a 24 mm i. d. tube, and in addition, 50 mm filled with 0.2 mm Er3Ni spheres. The second pulse tube is 380 mm long with 11 mm i. d. It is also equipped with Cu mesh heat exchangers at both ends. The cooler has been operated with a 6 kW compressor (Leybold RW 6000) and a magnetic valve pressure wave generator with passive (double inlet type) phase shifter ( and
buffer volumes). The lowest temperature with
has been achieved for operation
with 1.8 Hz and with pressure swing between 0.11 and 0.22 MPa. The temperatures adjusting at the first and at the second stage, and respectively, when powers and are applied, are shown in Fig. 2. Fore more detailed information, the time dependant pressure signals have been
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Fig. 1. Scheme of the 2-stage PTR a) Experimental set-up, b) Model for the numeric simulation with
calculated results for energy flows (arrows).
measured at the warm ends of the regenerator and of both pulse tubes, and in addition, time averaged temperatures have been measured at 20 positions of sensors attached to the outer surfaces of regenerators and pulse tubes. More details and data for other operational modes have been published earlier /1, 2/. The refrigerator has been set-up from components partially existing from earlier experiments. It is expected that more refrigeration power can be obtained by further modifications. A numeric model which can describe those results will be helpful for the optimisation of such systems. NUMERIC SIMULATION
The thermoacoustic theory with the differential equations as developed by Xiao /3/ has proven to be a powerful tool for the study of pulse tube coolers. In an earlier paper /4/ we have shown that many characteristic features of single stage pulse tube coolers can be described by such calculations, when a few parameters are fitted empirically. Here, such calculations will be extended for the two stage system. Adaptation of parameters The existing theory has been developed for real gas with harmonic small amplitude oscillations of pressure and volume flow interacting with the wall of circular channels with the inner radius b and the wall thickness bs. The present regenerators with mesh and sphere packages
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Fig. 2 Power chart for thermal loads simultaneously applied to both stages. will be described by equivalent arrays of parallel tubes with the parameters b and bs chosen such that the resultant structure will have the same porosity and the same inner surface. Both quantities are known for the regenerators used in the experiment, and the equivalent parallel tube parameters are given by
and
for the mesh structure with porosity and wires of diameter d , and with
for the package of spheres with the diameter D (the equivalent wall thickness is also obtained from Eq.(2)). The equivalent regenerators are described by arrays of densely packed hexagonal tubes with the inner radius mm and the wall thickness mm for the mesh with 67 % porosity, and with mm and mm for densely packed spheres with 0.2 mm
diameters. The overall size is the same as in the experiment. The heat exchangers are modelled in the same way. All other components are of circular shape with given tube parameters (length, inner diameter, wall thickness and material). The basic set of linear differential equations with the complex variables of pressure, p , and volume flow oscillation, U , and with the real variables of the time averaged local temperature, and enthalpy flow, E(x), is being solved for the arrangement as shown in Fig. 1b). The 13 components as indicated by the numbers are being taken into consideration. The inlet parameters such as volume flow given by the swept volume of a fictive piston compressor operated with the frequency f, the pressure wave given by its mean value the compression ratio, and the inlet phase angle at the inlet of the supply tube (pos. 0), and in addition the fraction of volume flow going from the first regenerator into the first pulse tube together with its phase angle difference are modified until a solution going through the given target points is obtained. A list of all input parameters is shown in Tab. 1.
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By this procedure, solutions describing any of the operation states with first and second stage temperatures and powers as marked by the points in the power chart of Fig. 2 can be found. Only the lowest point with zero load at both stages could not be reached. But the best approach will be discussed below. It is however found that numeric solutions going through those target points with temperatures given by Fig. 2 and ambient temperature (300 K) at the main inlet (pos.1) and at the warm ends of both pulse tubes can be obtained with different sets of initial data. The range of those data becomes narrower when more target points are fitted. Hence, we try to find numeric solutions with reasonable fits also to the measured temperature distribution. This can be done by iterative modifications of mainly 5 more or less free parameters such as swept volume, compression ratio, and the volume flow ratio at the branching of stage 1 and 2 (bold numbers in Tab. 1). The enthalpy flows in both regenerators are adjusted such that the T(x) curves go to 300 K at the ends of both pulse tubes. A typical list of all input parameters is shown in Tab. 1. Most material parameters are taken from the CRYODATA(R) code. Only the specific heat of has been obtained from interpolation of a given data list. Its thermal conductivity, a not so critical parameter for low frequency operation, has been assumed half the value of Pb. In addition to that, non-isotropic thermal conductivity of regenerator beds with only 20 % of the bulk material conduction value in axial direction has been assumed. (This assumption proves also to be not very critical. Typically, the effect on the lowest temperature is in the range of 0.5 K). Numeric approach for the no-load operational state The most interesting question is to find out how good this numeric model can predict the lowest temperature obtained by the experiment. The result of such a study is shown in Fig. 3
where measured and calculated temperatures at different positions are compared. This calculation has been done with the only constriction that the T(x) curves must go through the points marked
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by the open circles, i. e. room temperature (300 K) at the warm ends of regenerator and pulse tubes, and at the cold and of the first stage. The mentioned set of the 5 input parameters has been modified stepwise with the aim to obtain a numeric solution with lowest possible temperature at the cold end of the second stage. The best result is shown by the solid curves describing the time averaged temperature distributions in the three regenerators and also in both pulse tubes. The symbols mark the temperatures measured with sensors attached to outer walls of those components. The same result is plotted with arithmetic ordinate (a) and with logarithmic ordinate (b). This is done for emphasising the discrepance in the low temperature range. For the given geometry of the system with the parameters listed in Tab. 1, the lowest temperature achieved by manifold modifications of the operational parameters is T2,min = 7.28 K whereas a temperature of about 6.2 K has been obtained by the experiment. Also the temperature distributions do not agree quantitatively, but there is surprisingly good qualitative agreement for some typical features. The temperature distributions in all components are more or less S-shaped.
Fig. 3. Measured and calculated temperatures in regenerators and in pulse tubes of a 2-stage
PTR operated without heat load at 1.8 Hz (linear and log T scale)
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This is most pronounced for the second stage pulse tube. But here, the main curvature is in the low temperature range for the experiment whereas for the calculation, it proves to be in the high temperature range. The opposed situation is found for the first stage regenerator. A better agreement with respect to those S-shapes can be obtained for numeric solutions with higher flow rate fed into the regenerator and smaller volume flow fed into the warm end of PT 2. But the resultant low end temperature proves to increase for such calculations. This might by explained by enhanced dissipative losses caused by higher flow rate in the regenerator. The fact that the prediction is worse than the experiment might by explained by the difference in pressure waves which is harmonic for the model and more or less rectangular for the experiment.
Considering the manifold simplifications used in the theoretical model, one should not expect much better agreement. Nevertheless, the model proves to be very helpful for parametric studies on such coolers. The effect of changes either in any of the many geometric parameters or of operational conditions such as flow rate and compression ratio can be studied very quickly. The response time is only a few seconds when the calculations are done our main frame RISC computer. Moreover, such calculations yield many parameters which are helpful for the understanding and for the design of such systems, but which can not be obtained easily from experiments. Some of such results will also be given for the above mentioned operational state. One surprising result is that in most cases the best performance is obtained with inlet conditions such that the pressure oscillation is delayed by
behind the volume flow. The phase angles and also their difference change at other positions of the system. This is shown in Fig. 4 where and are plotted over the unified lengths of the different components. The phase shift of the pressure wave proves to be rather small for operation at the low frequency, but appreciable phase shift of the volume flow is seen to occur in both regenerators and before all in the pulse tubes. In the experimental system, the phase shift at their warm ends are forced by the passive resonators. A comparison with the calculated result has not yet been possible. For completeness, also the amplitudes of both quantities, pressure and flow, are given (Fig. 5). It is worth mentioning that the flow rates at cold and at the warm ends of both pulse tubes do not differ much. There is a minimum in the mid of PT 2. This correlates with the position of zero
phase difference between pressure and volume flow. This condition must not necessarily be in the mid of the pulse tube, but the numeric analysis indicates that this is advantages for obtaining lowest temperatures. The main pressure drop is seen to be caused by the regenerators. Both are seen to give about the same contribution. It proves to be smaller than for the experiment. A rough comparision is given in Fig. 6.
Fig. 4. Phase angle of pressure wave and volume flow rate at different positions of the system
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Fig. 5. Amplitudes of flow and pressure oscillation at different positions
In some cases it proves to be more instructive to plot those data in phasor diagrams as it is done in Fig. 7 for flow and pressure in the cold part of the system. The situation at the position where the flow is splitted into the first and the second stage is also displayed. The flow results from the calculation, but its splitting into and with is a system inherent parameter. In the present model, this is handled by two additional free parameters, the ratio (Flow 1/2) and the phase angle between both. First calculations where done with the assumption that all three vectors are parallel. In this case cold end temperatures not lower than 8.5 K could be obtained. The introduction of the additional phase angle has lowered the calculated temperature to 7.3 K, but the phase angle of has proven to be the optimum. In the experiment, such a phase shift will be caused by the different acoustic impedances of the components adjacent to the branching point.
Fig. 6. Pressure waves at the inlet of the first regenerator and in both pilse tubes. Comparison of measurement and numerics.
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Energy flows When the complex quantities of pressure and volume flow, p and U respectively, are known, different forms of energy flows can also be calculated. Those data are displayed in Fig. 1b, namely the time averaged "pV-work", the axial heat flow, and the enthalpy flow describing the total of different kinds of dissipative losses. The no-load system with and K at first and second stage, respectively, is considered here. The work flow supplied by the compressor is (This is not the input power). Less than 10 % of this work, becomes available for refrigeration at the first stage where it is balanced by the enthalpy flow and the heat flow which is caused by dissipative effects in the first pulse tube. This term proves to be surprisingly high. It is mainly caused by "shuttle losses" due to the oscillation of the gas flow. This fact has been verified by additional calculations done for pulse tubes with 0.5 mm and with 1.0 mm wall thickness. The difference in heat flow proves to be in the 10 % range. In the calculation, it is assumed that all dissipative losses of Reg. 1 and PT 1 are removed by the work flow arriving at the cold end of PT1. The work W is being used to remove the dissipative losses of both second stage regenerators, and the heat flow caused by dissipation in PT2. The balance of the three terms determines the cold
end temperature Other simulations have been done with higher initial flow rate (higher swept volume of the compressor) and also with higher pressure amplitudes and with different initial phase angles between both. This can make much higher work flows at the cold ends, but this effect proves to be overcompensated by increase of dissipative losses. The fact that the measured second stage cold end temperature is lower than predicted by this model might be caused by the broader plateau of the pressure wave as it results from operation with magnetic valves. For comparison, the pressure cycle measured at the warm ends of the regenerator and of both pulse tubes are displayed in Fig. 6. together with waves of the present calculation where laminar flow is assumed
at all positions. This graph shows also that the calculated pressure drop in the regenerators is smaller than in the experiment.
Fig. 7. Inlet pressure and volume flows at different positions
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CONCLUSIONS
A two-stage pulse tube refrigerator even with a rare earth regenerator part can be simulated by using the linearised thermoacoustic equations derived by Xiao. Reasonable results are obtained when the structure of both, mesh type and sphere beds, are substituted by parallel channel structures with the same porosity and with the same heat transfer surfaces. No numeric instability is found to be caused by the peak in specific heat of the regenerator. The performance of a real system operated with square wave pressure generator proves to be some what better than predicted for harmonic waves. Such calculations give valuable information on manifold parametric interactions. This will be very helpful for the design of optimised systems. An analogue simulation of data obtained with a harmonic wave driven cooler would be very valuable. ACKNOWLEDGMENT
The Er3Ni probe has been gifted by U. Haefner, Leybold Co., and G. Thummes, University Giessen has given us the specific heat data. Both are greatly appreciated. REFERENCES
1. Wild, S., Oellrich, R.L., and Hofmann, A., "Zweistufiger Pulsrohrkühler mit Magnetventilen bis 6K", DKV-Tagung, 19.-21.Nov. 1997, Hamburg
2. Wild, S., "Untersuchung ein- und zweistufiger Pulsrohrkühler", Fortschritt-Berichte VDI, Reihe 19, Nr. 105, VDI Verlag Düsseldorf 1997, ISBN 3-18-310519-5
3. Xiao, J.H., " Thermoacoustic heat transportation and energy transformation, Formulation of the problem", Cryogenics 1995, Vol. 35, p. 15-19
Part 1:
4. Hofmann, A.,Wild, S., and Oellrich, R.L., "Parallel flow regenerator for pulse tube cooler application", Advances in Cryogenic Engineering, Vol. 43 (1998)
Modeling Pulse Tube Coolers with the MS*2 Stirling Cycle Code Matthew P. Mitchell
Mitchell/Stirling Berkeley, California, USA Luc Bauwens
University of Calgary Calgary, Alberta, Canada
ABSTRACT
The MS*2 Stirling Cycle Code implements a numerical model originally designed to analyze Stirling cycle engines and refrigerators on a PC. It can also be used to create usefully accurate models of pulse tube cryocoolers. The technique requires preparation of higher and
lower resolution files from which a hypothetical file of infinite resolution can be projected. By varying heat transfer area assigned to the cold heat exchanger, different cooling loads are applied. From the resulting equilibrium cold end temperatures, load curves are developed. Separate determination of load, including load represented by losses not modelled by the code (principally, conduction in pulse tube and regenerator housing) leads to an estimate of temperatures achievable at various assumed loads by reference to the load curve. Output of the code includes projections of pressure drop loss in the system, enthalpy flow in the regenerator, PV work input in the
compressor and temperatures, heat flows, mass flows as well as mean Reynolds, Nusselt and Mach numbers for each control volume. Material options include several working fluids and regenerator materials. Regenerator configuration options include screens, spheres and parallel plates of any density and element size. Phase shift between flows at the compression piston and orifice is an adjustable input. INTRODUCTION
Regenerative gas cycle machinery such as Stirling, Gifford-McMahon and pulse tube coolers is impossible to analyze rigorously in closed form. The constant interplay of flow, heat transfer, matrix temperature change and pressure change, which vary continuously in every part of the machine, makes a numerical approach attractive. The practical challenges in producing a useful, convergent numerical code arise from the limitations of computing power, which limit the number of discrete volumes that can be modelled and the number of time steps into which a cycle can be broken. In a convergent scheme, however, the exact solution attainable with an infinite number of control volumes and time steps can be approximated by running parallel calculations in higher and Cryocoolers 10, edited by R. G. Ross, Jr. Kluwer Academic/Plenum Publishers, 1999
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lower resolutions and projecting a result. The projected solution is usefully accurate even if the resolution of the high and low resolution cases is comparatively crude and the raw results of both high and low resolution cases are far from accurate. A technical limitation that affects explicit numerical models results from the CourantFriedrichs-Levy (“CFL”) condition for stability. (1) The CFL condition limits the ratio of time steps to space steps. The projection technique avoids the CFL condition by permitting a relatively
coarse grid of spaces that are too large to be overridden by a moving fluid boundary during a cycle. ADVANTAGES OF THE NUMERICAL APPROACH
In a real machine, no space, including the regenerator, is completely adiabatic or completely isothermal. While the center of the regenerator may approach an isothermal condition,
the temperature of the matrix floats to some extent over the cycle. At the ends of the regenerator, more significant temperature differences between fluid and matrix arise when fluid is entering at a temperature different from that of the matrix, and the float in matrix temperature is more
pronounced. However, when fluid is leaving an end of the regenerator, fluid temperature remains close to that of the matrix. After the regenerator, heat exchangers are typically the most-nearly-isothermal parts of the machine. Again, however, there is a temperature difference that varies over the cycle. Ducts, compression and expansion spaces and, in pulse tube coolers, the pulse tube itself, may be close to adiabatic, but may also generate considerable heat transfer depending upon their dimensions and flow conditions. A numerical approach can address the unique conditions in each control volume separately. THE MS*2 STIRLING CYCLE CODE
The MS*2 Stirling Cycle Code discretizes the spaces of a gas-cycle cooler as shown in Figure 1. Within each of these discrete spaces, varying numbers of control volumes can be
specified; the combined total of all control volumes cannot exceed 200 in the PC version. The cooler and freezer are modelled as shell-and-tube heat exchangers, with any number of tubes (including just one) and any lengths and diameters. Ducts are modelled as single tubes. "Area factors" that vary the heat transfer area (and thus heat transfer) can be specified for all spaces. While the model was originally designed to model Stirling machines, it can also be used to model orifice pulse tube coolers. To do so, the freezer and the duct between the expansion space
are modelled as the pulse tube itself. The duct between the freezer and regenerator becomes the cold heat exchanger. The expansion space becomes the orifice, controlling the flow at the warm
Figure 1. Discrete spaces modelled by MS*2 Stirling Cycle Code.
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Figure 2. Space designation for a pulse tube cooler.
end of the pulse tube. The phase between the compression piston and expansion piston is userspecified. The arrangement is illustrated in Figure 2. The MS*2 code currently assigns the same wall temperature to freezer, expansion space, and ducts connected to them. If the expansion space is assigned near-infinite area and the pulse tube and cold heat exchanger are assigned no area at all, the pulse tube and cold heat exchanger
become adiabatic spaces while the temperature of fluid in the expansion space is effectively the assigned wall temperature. In that configuration, periodic steady-state operation at no-load condition is modelled and the last segment of the regenerator at its cold end reflects the no-load temperature. If the cold heat exchanger is assigned significant area (and assuming that it is assigned an appropriate length and diameter) it will begin to absorb heat and the temperature of the cold end of the regenerator will rise. The heat lifted in that cold-heat-exchanger/duct (i.e. the cooling load) is reported by the code. Although neither cold end temperature nor load can be specified as an input, load curves can be derived from several cases with differing area factors, and thus varying heat transfer, at the cold heat exchanger. Because the MS*2 code models heat transfers in all spaces based upon finite temperature differences, the irreversibilities generated by those heat transfers are incorporated in the results. Similarly, the float in regenerator matrix temperature as a result of heat transfers in and out is modelled directly and its effects are included in the results. Pressure drop throughout the machine is calculated, accumulated and reported separately. CONVERGENCE
The concept of "convergence" as applied to the issue of grid-independence of a numerical code is to be distinguished from "convergence" as sometimes applied to describe the process of
adjustment by which the solution of a problem approaches a stationary periodic condition. "Convergence" in the sense considered here means that as the code uses an ever-finer mesh in space and in time, its output approaches the exact solution, in the limit of an infinite number of spaces and time steps. The implication is that too coarse a single-resolution model, constrained by speed and memory limitations of affordable computers, will not be accurate. Indeed, even at maximum resolution, the MS*2 code is often far off the mark in modelling cryocoolers in a single pass. This issue was addressed, for the MS*2 Stirling Cycle code, in earlier papers. (2, 3, 4, 5 ) In an earlier analysis using a Cray supercomputer, as many as 400 control volumes and 8 time steps per cycle were used. (2) The conclusion was that, for spatial and temporal resolutions achievable with the PC version of the MS*2 code, a linear projection of results from high- and low-resolution solutions to the same problem will closely approximate the exact solution. (That is because stability requirements limit the algorithm to first order accuracy). In applying the MS*2 code to pulse tube coolers, parallel cases were run with the following discretization:
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EXTRAPOLATION
Extrapolation of results of two cases of different resolution is employed to approximate an exact solution not obtainable in a single pass. The device modelled is comprised of large spaces such as cylinders, and smaller volumes such as heat exchangers and segments of the regenerator. The latter can be usefully represented by a one-dimensional model that resolves longitudinal
gradients. On the other hand, a qualitatively equivalent representation of the physical processes that occur in the cylinders is not doable. That is because these inherently three-dimensional processes are time-independent and often turbulent. At best, the global error will be limited by discretization of the processes in the cylinders. Additionally, the mass fluxes due to piston motion appear as fronts with occasional large gradients, which require a conservative treatment. Monotonicity issues result in instabilities in high-order accurate algorithms, and work-around schemes such as flux limiters would result in an algorithm that truly could only be first order accurate. Under these circumstances, any attempt at high order accuracy is more or less pointless. The MS*2 code is first order accurate, and this has been verified in a convergence analysis. (4) While with a resolution fine enough, reasonably accurate predictions of local variables such as temperatures, pressure, etc. can be obtained, obtaining well-converged global data such as efficiency or COP remains impractical with available resolution. However, it is possible to use the linear convergence property of the first order scheme and to extrapolate results with relatively coarse schemes to the convergence limit. As shown below, the results obtained are surprisingly
accurate when compared with experimental data. VALIDATION
To assess the potential usefullness of the MS*2 code in designing pulse tube coolers, it
was used to model a double-inlet pulse tube cooler with screen regenerator as reported by a team
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in Karlsruhe. (6, 7) The technique employed was to specify dimensions for pulse tube and regenerator as built and to specify other spaces as appropriate to generate the flow conditions and pressure wave observed experimentally. The results were then incorporated in a spreadsheet with adjustments for axial conduction losses calculated from reported dimensions and materials of pulse and regenerator housing walls. To arrive at a predicted solution for the internal operation of a pulse tube cooler, the high and low resolution results were projected for cold-end temperature and cold-end load. Results are shown in Figure 3. Although both high- and low-resolution cases were far from observed performance, the projected result closely tracked the observed performance from temperatures around 50 K upward. The divergence of the projected result from the observed result between no-load temperature and about 50 K may be due in significant part to the reduced heat capacity of stainless steel at those temperatures; the MS*2 code assumes stainless steel properties at ambient temperature. PRACTICALITY
Practical requirements for a useful model require that it be able to handle the operating conditions found in real machines, that it be numerically stable and that it deliver a solution within a reasonable time. Stability is a key issue. Pulse tubes in valved systems operate at pressure ratios typical of Gifford-McMahon machines. With substantial pressure swings come large changes in temperatures and flows. The primary cause of instability is violation of the CFL condition, in which flow during one time step crosses two control volume boundaries. When large temperature changes in conjunction with large pressure changes and large flows produce large heat transfers, density changes generate additional flows (in the model) and instability is easily triggered, causing the program to crash. Another potential source of instability is the overrelaxation scheme that speeds adjustment of regenerator matrix temperatures to achieve a periodic steady state condition. The MS*2 code requires a lengthy data file as starting input. Configuration, dimensions and operating conditions described in that file can be altered through a series of menus to create a new model to be investigated. Almost any change in dimensions or operating conditions can cause the regenerator matrix temperature profile to differ from that which would exist under periodic steady-state conditions. If the code is run for enough cycles, unbalanced heat transfers into and out of each
Figure 3. Comparison of computer output with test results.
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regenerator segment will slowly force the matrix temperature profile toward the periodic steady state condition. However, without an overrelaxation factor to speed that process, many hundreds or even thousands of cycles may be required to reach a state such that further change in regenerator temperature profile generates no significant change in load or temperature at the cold
heat exchanger. A crude overrelaxation scheme
uses the change in temperature of a segment of
regenerator matrix over one cycle as the adjustment factor for the next cycle. If the adjustment factor is moderated appropriately, no instability will be triggered, and the program will not crash. However, a numerical oscillation may develop in which the temperatures of the regenerator segments continually adjust but do not reach equilibrium.
The MS*2 code addresses both sources of potential instability. By increasing the number of time steps per cycle or reducing the number of control volumes specified, the user can insure that the CFL condition is not violated as a result of flows that occur during the cycle. Potential instability arising from adjustment of regenerator matrix temperature between cycles is dealt with separately. Instead of adjusting temperature of regenerator matrix segments in proportion to their heat-flow imbalance, the MS*2 code adjusts by a fractional exponential factor, typically the square root of a fraction of the "raw" adjustment factor. Both the scale of the raw adjustment factor and the exponent are user-specifiable. By selecting appropriate scales and exponents, regenerator matrix temperatures can be forced to periodic steady state faster than would be possible if the regenerator matrix temperature were allowed to slowly float to equilibrium. As
computer speeds have increased, the MS*2 code has become an increasingly practical tool for pulse tube design using a PC. CAVEATS Some caveats are appropriate. Modelling cryocoolers is a challenging task at temperatures at which very large power inputs are producing very little cooling. The MS*2 Stirling Cycle Code assumes that its specifiable working fluids (hydrogen, helium, nitrogen, mixtures) are perfect gases. It does not model the reduced heat capacity of regenerator materials at low temperatures.
However, the MS*2 code does permit specification of a conical regenerator. By specifying a small diameter at the cold end the thermal mass at the cold end can be reduced, simulating the effect of reduced heat capacity. The side-effects of that adjustment include an increase in pressure ratio (unless the compression space displacement is reduced) and an increase in reported pressure drop.
This adjustment also artificially increases the Reynolds number in the cold end of the regenerator and the corresponding Nusselt number for heat transfer purposes.
In a typical pulse tube cooler, most of the pressure drop occurs in the regenerator. The effect is to reduce the pressure amplitude in the pulse tube relative to the pressure amplitude delivered to the warm end of the regenerator by the compressor. The MS*2 code adjusts pressure to the same level throughout the machine at each time step. Thus a geometry that produces the correct pressure amplitude in the pulse tube must be used although the pressure amplitude in the
compressor of the real machine will be higher. The MS*2 code does not attempt to deal with losses that result from conduction through pulse tube or regenerator housing walls. It does not attempt to deal rigorously with convection between working fluid and pulse tube walls. It cannot foresee and adjust for elements of
workmanship, design or operation that generate turbulence or streaming in the pulse tube. To the extent that they can be calculated or estimated, these losses can be dealt with externally. The MS*2 code does not provide directly for "double inlet" or "inertance tube" effects. However, the phase angle of flows at the warm end of the pulse tube are adjustable relative to the phase angle of the compressor piston, and real-time graphic output permits detailed analysis of the effect of those adjustments on flow phasing.
The MS*2 code permits variation in the phase of an expansion space piston relative to the phase of the compression piston position. The interaction of the phases between pistons produces
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non-sinusoidal flow and pressure variations, but they do not necessarily correspond to the actual
pressue and flow conditions in real machines, particularly in valved pulse tube coolers where inflows and outflows occur rapidly. CONCLUSION
The integrity of the numeric scheme embodied in the MS*2 Stirling Cycle Code is demonstrated by its ability to model pulse tube coolers with useful accuracy although the code was designed for another purpose. Accuracy is obtained by projecting the results of two parallel calculations done in high and low resolution, respectively. The numeric scheme adjusts quickly, without triggering instabilities, through use of a fractional exponential overrelaxation factor that damps large corrections to avoid triggering the CFL condition. ACKNOWLEDGEMENTS
Dr. Albert Hofmann of Forschungszentrum Karlsruhe graciously furnished operating data on the screen-regenerator-equipped pulse tube refrigerator described in references 6 and 7, and
modelled as described above. REFERENCES
1. Courant, R., Friedrichs, K. O., Lewy, H., “Ueber die Partiellen Differenzgleichungen der Mathematischen Physic”, Math. Ann. 100, (1928) p. 32. 2. Bauwens, L., Mitchell, M. P. "Consistency, Stability, Convergence of Stirling Engine Models",
Proc. 25th IECEC, (1990), vol. 4, p. 352.
3. Mitchell, M. P. and Bauwens, L., "Validation of Empirical Models: Empiricism vs. the Laws of Physics", Proc., 25th IECEC, (1990), vol. 5, p. 424.
4. Bauwens, L. and Mitchell, M. P., "Regenerator Analysis: Validation of the MS*2 Stirling Cycle Code", Proc. XVIIIth International Congress of Refrigeration, Montreal (1991), p. 930. 5. Mitchell, M. P., "The MS*2 Stirling Cycle Code", Proc. Seventh International Cryocooler Conference, (1992), vol. 1, p. 290. 6. Wild, S., Oelrich, L. R., and Hofmann, A., “Two Stage Double-Inlet Pulse Tube Refrigerator Down to 10 K”, Cryocoolers 9, Plenum Press (1996), p. 255 7. Hofmann, A., Wild, S., and Oelrich, L. R., “Parallel Flow Regenerator for Pulse Tube Cooler Application”, CEC, Portland, (1997).
Experimental Verification of a Thermodynamic Model for a Pulse Tube Cryocooler J. Yuan and J.M. Pfotenhauer Applied Superconductivity Center University of Wisconsin - Madison
ABSTRACT In a recently submitted article1 a thermodynamic model is described which provides an explanation for the performance of a GM type pulse tube cycle, permits optimization of cooling power for a given pulse tube system through the control of valve timing, and allows the design of pulse tubes to achieve the optimized coefficient of performance for specific cooling capacities. The cooling capacity is shown to be correlated with the net work done by the cold end control volume over one cycle. This paper presents an experimental verification of that model for two different scenarios, optimizing cooling power for a given pulse tube system and optimizing COP for a given compressor. The influence of the important parameters, including two intermediate pressures, the timing of the isobaric processes, and the size of the pulse tube on the refrigeration performance are
intensively investigated. In addition, the results define a minimum necessary compressor capacity based on the pulse tube size and cycle frequency. Performance is characterized by the cooldown time, the minimum cold end temperature, and the cooling capacity at 60 K. The model predictions are compared both with experimental results obtained from systems driven by a GM compressor with an electrical input power of 1 kW, and with those reported for the Active Buffer system of Zhu2. INTRODUCTION Pulse tube refrigerators have attracted extensive interest in recent years due to their high potential for reliability and simplicity. Since the first renovation of configuration (adding an orifice and reservoir) was introduced in 1984 by Mikulin3, great progress has been made to improve the performance of this kind of device. A variety of configurations have been proposed and tested. However, the industrializing process of the pulse tube cryocooler has been much slower. A particularly prominent reason for this is the lack of an easily accessible tool to guide the design process. Over the last decade, a variety of approaches have been developed to model the operation of various pulse tube configurations in order to provide a reliable design tool. These include the 1D enthalpy flow model4, 2-D models5 and a variety of numerical methods6. In a recently submitted article1, a thermodynamic model is described which provides an explanation for the performance of a GM type pulse tube cycle, permits optimization of cooling power for a given pulse tube system through the control of two intermediate pressures (which can be realized by an active valve system such as the five valve system described elsewhere7), and allows design of pulse tubes to achieve the optimized COP for specific cooling capacities. Cooling capacity is shown to be correlated with the net work done by the cold end control volume over one
cycle. This paper presents an experimental verification of that model for two different scenarios; optimizing cooling power for a given pulse tube system and optimizing COP for a given compressor. The influence of the key parameters, including two intermediate pressures, the timing of the isobaric processes, and the size of pulse tube, on the refrigeration performance are Cryocoolers 10, edited by R. G. Ross, Jr. Kluwer Academic/Plenum Publishers, 1999
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intensively investigated. In addition, the results define a minimum necessary compressor capacity based on the pulse tube size and cycle frequency. Performance is characterized by cooldown time, the minimum cold end temperature, and the cooling capacity at 60 K. An optimized cooling capacity of 15.3 watts at 60 K is achieved with an input power of 1 kW, equivalent to an efficiency
of 6.12% of Carnot. In the following paragraphs, the key features of the theoretical simulations are provided, followed by a description of the experiment setup and a comparison of the experimental results with the theoretical predictions. An additional test of the theoretical predictions is provided by a comparison with the results from the active buffer system of Zhu2. THEORETICAL MODEL
A complete and detailed description of the theoretical model is presented elsewhere1, however the main features of the approach and simulation results are summarized here. The calculation of
cooling capacity is based on the thermodynamic analysis of the control volume at the cold end of pulse tube as shown in Fig. 1. The boundaries of this control volume are defined by the walls of the pulse tube on the side, by a moveable boundary between the cold end volume and the gas piston, and by a fixed boundary at the lower end of the cold end heat exchanger. The control volume can do work on the gas piston ‘p’, but exchanges no mass or heat with the gas piston. The side walls are adiabatic. The thermodynamic cycle can be described by considering the cold gas segment ‘c’. Fig. 2 illustrates the equivalent P-V diagram of the cold gas segment ‘c’. Initially, the system is at state 1 with low pressure The operating process can be depicted as six steps: (1-2) Pressurize the system through the warm end of the pulse tube to the first intermediate
pressure
high pressure
(2-3) Further pressurize the system through the cold end of the pulse tube to the (3-4) Isobarically shuttle the gas in the pulse tube toward the warm end. (4-5)
Depressurize the system through the warm end of the pulse tube to the second intermediate
pressure (5-6) Further depressurize the system through the cold end of pulse tube to the low pressure (6-1) Isobarically shuttle the gas in the pulse tube to its initial state thereby finishing one cycle. The energy balance for the control volume as show in Fig. 1 over one cycle is given by
where is the cooling capacity per cycle and is the net work which the cold space ‘c’ does on the gas piston ‘p’ in one period. The theoretical analysis shows1 that the work term shown in Eq. (1) is a function of the dead volume of regenerator, the total volume of pulse tube, the operating condition of system (including high and low pressures, high and low temperature), the first intermediate pressure the second intermediate pressure and the timing of isobaric process. To verify the theoretical model, three series of thermodynamic simulations have been carried
Figure 1. The control volume described by
the model.
Figure 2. An equivalent P-V diagram of cold space of pulse tube.
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389
out in the present investigation. In order to compare the computational results with the experimental data gathered by ourselves and found in the literature2, the pulse tube size and working conditions used in the simulation are the same as those of the experiments. Table 1 lists the three pulse tube dimensions, regenerator sizes and working conditions used in the simulations.
The sizes of pulse tube I, II and regenerator I are relatively small while the size of pulse tube III and regenerator II are relatively large. The total volumes of pulse tube I, II and III are and respectively. The void volumes of regenerator I and II are and respectively. The operating frequency for pulse tube I and II is 1.8 Hz while the working frequency for pulse tube III is 1.9 Hz. Both pulse tube I and II operate with regenerator I while pulse tube III operates with regenerator II. Simulations have been conducted by varying the two intermediate pressures for each of the refrigeration systems. The final results are illustrated as the ideal cooling power at 60 K, the COP and the required theoretical (ideal) compressor work vs. the two intermediate pressures. The simulation results for case 2 are illustrated in Fig.3. Similar results are obtained for case 1. Both results clearly demonstrate that there is an optimized operating point for the refrigeration power. and while The optimized intermediate pressures for the case 1 is and The the optimized intermediate pressures for the case 2 is model suggests that the higher volume-ratio system will have values of and closer to the system high and low pressures, respectively. The optimized gross (ideal) refrigeration powers at
60 K for case 1 and case 2 are 17.3 watts and 33 watts, respectively. The required theoretical compressor work for pulse tube I and II are 225 watts and 373 watts, respectively. The calculation results for pulse tube III are depicted in Fig. 4. In this case, a larger pulse tube and regenerator, and different high and low pressures are used in the simulation. Note that the volume-ratio of pulse tube III to the void volume of regenerator II is similar to that of case 2. The simulation results show similar behaviors for the cooling power and compressor work. Indeed if, the same high and low pressures were used for both cases 2 and 3, the optimum point for cooling capacity (in terms of and ) would be nearly the same, even though the pulse tube volumes and gross cooling capacities are quite different. Thus, the model suggests that the optimum intermediate pressures are a function of the volume-ratio (pulse tube volume to regenerator dead volume), rather than the pulse tube volume. The optimized value of for case 3 is 2.05 MPa while the optimized value of is 1.66 MPa. The optimized gross refrigeration power for pulse tube III is about 278 watts and the required ideal compressor work is about 3.0 kW.
Figure 3. Dependence of refrigeration power and compressor work on the two intermediate pressures for pulse tube II.
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PULSE TUBE MODELING AND DIAGNOSTIC MEASUREMENTS
Figure 4. Dependence of refrigeration power and compressor work on the two intermediate pressures for Pulse Tube III.
Figure 5. Dependence of COP on the two intermediate pressures for Pulse Tube II and Pulse Tube III. Fig. 5 displays the COP as a function of the two intermediate pressures for pulse tube II and pulse tube III. From Fig. 5 we can observe that the optimum intermediate pressures for the best COP are different from those for the maximum cooling production shown in Fig. 3 and Fig. 4. From the plots we can see that for both cases the first intermediate pressures which produce the best COP are slightly lower than those which result in the maximum cooling production. However, the second intermediate pressures which produce the best COP values are close to the system low pressure. EXPERIMENT DESCRIPTION
Two sets of refrigeration systems have been built to verify the theoretical predictions for pulse tubes I and II. Test results reported by Dr. Zhu for the active buffer system have been used to compare the theoretical prediction for pulse tube III. The five valve configuration as shown in Fig. 6 has been used for pulse tubes I and II while the active buffer configuration was used for pulse tube III. All pulse tubes used in the tests have the same dimensions as those used in the simulations. The regenerator used for pulse tubes I and II is fabricated from a thin wall G-10 tube 100 mm in length, 19 mm in diameter and with a 3.3 mm wall thickness. About 1000 pieces of 200 mesh bronze screens are packed inside the regenerator tube yielding a regenerator porosity of about 65%. The regenerator is seated in a thin wall stainless steel support tube which is arranged side by side with the pulse tube. This arrangement allows us to utilize the same regenerator for testing pulse tube I and pulse tube II. The regenerator support tube and the pulse tube are connected at the cold end via the heat exchanger assembly and at the warm end via the base plate. The cold end heat exchanger is fabricated from pure copper with high thermal conductivity. To insure that all connections are tightly sealed, the pulse tube and the regenerator support tube are
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Figure 6. Configuration of five valve pulse tube refrigerator.
welded to the base plate and silver soldered to the cold end heat exchanger. The warm end of the pulse tube is equipped with a water-cooled heat exchanger. Both ends of the pulse tube are filled with approximately 50 250-mesh copper screens to provide both a large heat exchanger surface area and flow straightening. The base plate also serves as the cover to the vacuum vessel which contains the whole system during operation. A detailed description of the test apparatus for pulse tube III can be found elsewhere2. A 30 watt heater is placed at the cold end to measure the cooling capacity for pulse tube I and pulse tube II. Three separate piezoresistive pressure transducers which allow measurement of the pressure oscillation are mounted on the warm ends of the pulse tube and regenerator and reservoir, respectively. The temperature at the cold end of the pulse tube is measured with a silicon diode thermometer manufactured by Lake Shore. Five solenoid valves, manufactured by CO-AX, are used as control valves. EXPERIMENTAL RESULTS AND DISCUSSIONS Pulse Tube I and II All tests for pulse tube I and pulse tube II are carried out under the low and high pressures of 0.8 MPa and 2.1 MPa, respectively. The operating frequencies for both pulse tube I and pulse II
are 1.8 Hz. A series of timings have been used for both pulse tube I and II to obtain various
combinations of the two different intermediate pressures. The intermediate pressures are measured by inspecting the pressure waves obtained with the Nicolet 4094C oscilloscope. The optimized
results for pulse tube I and pulse tube II are summarized in Table 2. Here the valve timings listed in column 3 and 5 are the number of degrees out of 360 (for a full cycle) for which the valve 3 and 4 are open. The results in Table 2 confirm two features predicted by the theoretical model for pulse tube I and II. First and foremost, both series of experiments indicate that the two intermediate pressures play a significant role in the system performance. Second, there is an optimized operating point
for a given refrigeration system. Our tests also show that the system with the larger volume-ratio will have optimized values of and closer to the system high and low pressures, respectively. This can be observed by comparing the optimized intermediate pressures displayed in columns 2 and 4. The optimized timings for valve 3 and valve 4 further confirm this expectation. As expected, the larger volume-ratio system will require a longer time for warm-end
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pressurization (thus a higher ) and warm-end expansion (thus a lower ). Comparing the refrigeration power of pulse tube I and II with those predicted by the simulation, one finds that for both pulse tubes the total system loss is about 52% of gross refrigeration power. In the present study, no efforts have been devoted to identify the individual loss terms.
Our experiments also indirectly confirm the required compressor work predicted by the theoretical model for pulse tubes I and II. The theoretically required compressor work for pulse tube I and II at optimized operating conditions are 225 watts and 374 watts, respectively. Taking the various inefficiencies such as mechanical, electrical, and compression into account, the actual required compressor input power should be around 450 watts and 750 watts for pulse tube I and II, respectively. Although the 1 kW GM compressor is used for both pulse tube I and II, in practice, a compressor with 500, or 750 watts input power can be used to drive pulse tubes I or II respectively. The theoretical predictions show that we can improve our system performance further by enlarging the pulse tube size to fully utilize the 1 kW compressor capacity. The cooling power at 60 K for pulse tubes I and pulse II are found to be 8 and 15.3 watts respectively. According to the model, this capacity could be increased to 25 watts using the same compressor. Pulse Tube III Experimental results from reference 2 are used for comparison with the third set of simulation results. Both intermediate pressures used in the experiment
result in a near-optimum COP according to the model and as illustrated by figure 5b. The active buffer data can also be used to illustrate the most appropriate use of the model for
producing optimum cooling power from a given compressor. In order to find the best pulse tube
size for the given 3.7 kW compressor, with an assumed 50% compressor efficiency, and therefore an ideal compressor power of 1.8 kW, a sequence of calculations, each with different pulse tube sizes, are conducted. First, for each chosen pulse tube size, note the values of and which give the best COP for that pulse tube size (e.g. fig. 5b). Second, determine the required compressor power corresponding to these values of and (e.g. fig 4b). Third, if the required compressor work is smaller (larger) than 1.8 kW, repeat steps 1 and 2 with a larger (smaller) pulse tube until the required compressor power is equal to 1.8 kW. Following these steps, we find that the model defines a required pulse tube size essentially the same as that used in the active buffer experiment. As shown in figure 5b, the net compressor power of 1.8 kW is consistent with the optimum COP associated with values of and of 1.6 MPa and 1.3 MPa respectively. With the exception of the value, these are the same results reported for the active buffer experiment.
CONCLUSIONS An experimental verification of a thermodynamic model has been presented. The experiments indicate that the model correctly described the thermodynamic characteristics of a single stage pulse tube refrigerator. The model provides a method for calculating and optimizing intermediate pressures for maximizing the cooling power of a given pulse tube. ACKNOWLEDGMENTS This study was funded in part by the U.S. Defense Nuclear Agency under DNA MIPR 92-719, work unit CD:00014, RCC: 7010.
REFERENCES
1. Yuan, J. and Pfotenhauer J. “ Thermodynamic Analysis of Five Valve Pulse Tube Refrigerator” submitted to Cryogenics. 2. Zhu, S., Kakimi, Y., Fujioka, K. and Matsubara, Y. " Active-Buffer Pulse Tube Refrigerator", Proc. Sixteenth Intl Cryogenic Engr. Conf. / Intl Cryogenics Matl. Conf. 1996, p.291 3. Mikulin, E.I., Tarasov, A.A., and Shkrebyonock M.P., "Low Temperature Expansion Pulse Tubes," Advances in Cryogenic Engineering, Vol.29, Plenum Press, New York, 1984, p.629.
EXPERIMENTAL VERIFICATION OF THERMODYNAMIC MODEL
4. Storch, P.J., Radebaugh, R. and Zimmerman, J.
393
"Analytical Model for the Refrigerator
Power of the Orifice Pulse Tube Refrigerator," NIST Technical Note 1343 , 1991. 5. J.M.Lee, and P. Kittel, K.D. Timmerhaus and R. Radebaugh, “ Steady Secondary Momentum and Enthalpy Streaming in the Pulse Tube Refrigerator” Proc. Eighth Cryocooler Conf., 1994. p 359. 6. Wang, C., Wu, P. and Chen, Z., "Numerical Modeling of an Orifice Pulse Tube Refrigerator," Cryogenics, Vol. 32, No.9, 1992, p.785-790. 7. Yuan, J and Pfotenhauer, J, “ A Single Stage Five Valve Pulse Tube Refrigerator reaching 32 K” accepted for publication in Advanced in Cryogenic Engineering, Vol. 43, 1998
Measurements of Gas Temperature in a Pulse Tube Using the Planar Laser Raleigh Scattering Method K. Nara, Y. Hagiwara, and S. Ito Advanced Mobile Telecommunication Technology Inc. 500-1 Minamiyama Komenoki-cho, Nisshin-si Aichi-ken, 470-0011 Japan
ABSTRACT In pulse tube refrigerators, the behavior of gas in the pulse tube is considered to influence the performance of refrigerator. In order to improve this performance, it is necessary to observe characteristics of gas behavior carefully. The gas temperature profile in the pulse tube is important in understanding the behavior of gas in particular. However, it has been difficult to directly measure the profile of gas temperature. This paper proposes to apply Planar Laser Rayleigh Scattering Method (PLRSM ) using an excimer laser to measure the gas profile. The Rayleigh Scattering is the elastic scattering of light quanta from molecules or small particles. The gas temperature can be calculated from the Rayleigh scattering light measured by PLRSM according to the ideal gas law, if the pressure of the gas is known. PLRSM can achieve two-dimensional measurements without disturbing the flow of gas. A prototype of the pulse tube refrigerator was developed with a quartz-glass pulse tube in order to utilize the laser system for PLRSM. It has been demonstrated that the gas temperature under static pressure can be measured with an error less than 0.89%. Further experiments have demonstrated that the PLRSM can visualize the temperature profile of gas under dynamic pressure . The experimental results on three types of pulse tubes, basic type, orifice type and double-inlet type, show that each has different fluid behavior. An example of gas flow visualization shows that the gas flow in the pulse tube has a twodimensional profile. And it can be estimated from the measured profile of the gas temperature that a difference in phase exists between the gas temperature and the pressure. INTRODUCTION In recent years, research and development of the pulse tube refrigerator have thrived. The pulse tube refrigerator has several advantages : it has high reliability because of no moving
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mechanical parts, it can be easily miniaturized, and it has high maintainability because of its simple structure. The authors have proposed the application of pulse tube refrigerator to the HTS filter system for advanced mobile telecommunications. As the development of the pulse tube refrigerator progressed from the basic type1) to the orifice type2),and on to the double-inlet type3), the cooling power has been improved remarkably This improvement is mainly the result of the use of phase shifters at the hot end of the pulse tube. It has been reported that the displacement and phase of the virtual gas piston4), which affect the cooling power, change with the phase shifters in use. The behavior of the gas in the pulse tube have been studied based on theoretical analyses, e.g. , thermoacoustic theory5), equivalent PV work6), and so on. However, few research groups have carried out experimental studies7~10). In the real pulse tube, the gas flow is not uniform as such the virtual gas piston because there are the circular and the turbulent flows in the pulse tube11). The gas temperature profile is affected by slight changes in flow because of the large temperature-gradient in the pulse tube. The heat flow in the pulse tube is affected by the gas temperature profile. Experimental studies are few in spite of the importance of the temperature profile as mentioned above. The purpose of this paper is to investigate the phenomenon in the pulse tube from the viewpoint of the temperature profiles and variation timing. The Planar Laser Rayleigh Scattering Method ( PLRSM), which we propose to apply to measure gas profile, can
achieve two-dimensional measurement without disturbing the gas flow, and can visualize gas temperature profile. MEASUREMENT METHOD
The principle of temperature measurement using PLRSM is first explained .The Rayleigh scattering is the elastic scattering of light quanta from molecules or small particles, where the particle is about one-tenth or less of a wavelength. The scattering light intensity of Rayleigh scattering is proportional to the product of incident light intensity, scattering cross-section and gas density12). The scattering cross-section is an the inverse fourth power of the wavelength of incident light13). Therefore, if the incident light intensity and the scattering cross - section are regular, and the Rayleigh scattering light is measured by PLRSM, the gas density can be calculated by comparing the scattering light intensity of the standard condition. As the gas is the ideal gas according to the ideal gas law, the density is proportional to the pressure and in inversely proportional to the temperature. Finally, the temperature profile is obtained, because the scattering light intensity and gas pressure are measured at the same time. In this experiment, an excimer laser with a wavelength of 193 nm is used. The incident light in the form of sheet is irradiated into the pulse tube through the upper side, so that scattering light is obtained from Rayleigh scattering of the gas in the pulse tube. The intensity of scattering light is affected by the undesired variation of the incident light. In order to compensate for the undesired variation, Rayleigh scattering of incident light caused by air outside the pulse tube is measured. If the gas pressure is uniform in the pulse tube, the following equation is satisfied:
where I is the scattering light intensity from the measured object, P is the gas pressure, Is is the scattering light intensity in the standard condition, Ps is the pressure in the standard condition, Ts is the temperature in the standard condition. Helium is usually used as an operating gas, but a nitrogen is selected . Helium does not
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Figure 2. Photograph of the pulse tube refrigerator with quartz-glass pulse tube.
Figure 1. Schematic diagram of measurement system.
Figure 3. Schematic diagram of pulse tube refrigerator.
obtain enough of the scattering light intensity, because the above-mentioned scattering crosssection of helium is only one-hundredth of nitrogen14). And the excimer laser is used to increase the scattering cross-section as the wavelength is short. EXPERIMENTAL APPARATUS
A schematic of the experimental apparatus is shown in Figure 1. The experimental apparatus consists of the pulse tube refrigerator, excimer laser, receiving light system called MakstovNewton telescope, ICCD camera (CCD camera with Intensifier), and a compressor with two valves to treat nitrogen gas with high pressure and low pressure. Those two valves are opened alternately in order to provide pressure wave for the pulse tube refrigerator. A photograph of the pulse tube refrigerator is shown in Figure 2. As the experiment is demonstrated in the atmospheric pressure, a heat insulator is wound on the regenerator. The schematic diagram of pulse tube refrigerator is shown in Figure 3 . It consists of the pulse tube made of a quartz glass in order to utilize the laser system for the PLRSM, regenerator and heat exchanger with a pressure sensor, and a thermocouple with a diameter of The pulse tube is constructed from quartz glass to allow the use of the excimer laser, because the absorption rate is small at the wavelength of 193nm. Four sheets of glass are fastened by a hoop with a force that overcomes the inner pressure. A seal is an indium foil. Control of the hoop is demonstrated by a screw with a diameter of 10 mm . Table 1 presents the construction parameters for the prototype of the pulse tube refrigerator. It has been demonstrated that the gas temperature under static pressure can be measured with an error less than 0.89%15,16).
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RESULTS
Measurement area and Pressure
The points measured by Rayleigh scattering in the experiment are shown in Figure 4 and Figure 5. Figure 4 shows the measurement area on pulse tube axis, and Figure 5 shows five points in the direction perpendicular to measurement area. Area II is the cross section on center line of the pulse tube. In the figures, for example, "Area A II " indicates area "A" and area " II ". We explained the result of two-dimensional measurement in the pulse tube. The measurement frequency is 19Hz. The measurement frequency is changed by the measurement range, as the processing time of the ICCD camera is limited . The pressure and timing of temperature measurement is shown Figure 6. Temperature profile of A,B and C
The experimental results are described in this chapter. The experiment was carried out in the current maximum area, as large as 15mm in width and 20 mm in length, to observe the detailed temperature oscillation in two-dimension: in the axial and radial directions. Figure 7 shows the temperature profiles in the basic pulse tube of A II ,B II and C II during one cycle ( with valve 1 and 2 closed ). Because the laser can be irradiated at only one area , measurements at three areas shown in Figure 7 were not carried out at the same time. However, because all experiments were carried out in a steady state, the temperature in those experiments
can be compared using the phases of the gas pressure as a reference. The cold-end temperature
Figure 4. Measurement area for Rayliegh Scattering
Figure 5. Cross section of measurement area (Top view, point B of Figure 4).
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Figure 6. Pressure and timing of the temperature measurement. is 277K. The numbers at the side of the picture correspond to the number of the pressure wave
as shown Figure 6. Figure 7 indicates that the temperature amplitude of the high-temperature side is larger than that of the lower-temperature side, and the temperature range of hightemperature side shifts to high temperature. However, the temperature in all three areas
increases during compression and decreases during expansion. There is hardly any phase difference between pressure and temperature at the three areas. Figure 8 shows the temperature
profile in the orifice pulse tube of A II, B II and C II during one cycle (with valve 1 optimized and valve 2 closed ). The cold-end temperature is 251K. The result shows that the orifice pulse tube has behaviors different from the basic pulse tube : (1) gas in the pulse tube has a large temperature gradient, (2) the temperature in all areas decreases while the pressure increases in
Figure 7. Gas temperature profiles in the basic pulse tube during one cycle .
Figure 8. Gas temperature profiles in the orifice pulse tube during one cycle.
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Gas temperature profiles of B I ,B II ,B III, B IV and B V in the orifice pulse tube during one cycle. Figure 9.
Figure 10. Gas temperature profiles in the
double inlet pulse tube during one cycle.
pictures 3 to 9, and (3) the temperature at the low-temperature side drops earlier than at the hightemperature side. It is estimated that these behaviors occur because of the existence of a phase shift between pressure and gas displacement, and because the regenerator works under near isothermal process. And Figure 8 also shows that the upper and lower sides of the picture have a gas temperature difference as shown in the picture 5, etc. It is estimated that this phenomenon occurs because gas of high temperature flows on the upper side, while gas of low temperature flows on the lower side due to a gravity. Figure 9 presents the measurement result of the area in the direction perpendicular to the measurement area at the part of B during one cycle. Compared with that, the upper and lower sides of the measurement area have a temperature difference of about 50K, and little variation in temperature is measured in the direction perpendicular to the measurement area. Experiments of the double-inlet type have been carried out by optimizing the opening of the orifice valve and by opening the double-inlet valve by degrees. Through the experiment, the optimal opening could not be found because the achieved temperature was high (we had no complete matching), and the achieved temperature increased when the opening of the valve increased. It can be estimated that the temperature increased because the optimized displacement and phase of the gas were shifted by the gas flow from the high-temperature side. The cause of this temperature increase can be estimated from the experimental results shown in Figure 10. The figure shows the temperature profiles in the double-inlet pulse tube of A II, B II and C II (with valve 1 optimized and valve 2 10-turn ). The cold-end temperature is 258K. The figure shows that the temperature gradient in the pulse tube is affected by the change of gas flow , namely the phase shifters. The phase shift between pressure and temperature in the
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401
double-inlet pulse tube is larger than that of the orifice pulse tube ( this phenomenon is clearly under expansion in particular ). It is also estimated that the gas piston in the pulse tube is
disturbed by the gas flow from the high-temperature side, and it affects the temperature increase.
Oscillation of mean temperature at a II ,b II and c II Secondly, the oscillation of mean temperature is examined because the phase shifts between temperature and pressure can be obtained clearly from the result. And the temperature amplitude can also be estimated quantitatively. It was examined in an axial direction with selecting the represented points of a II, b II and c II because there is hardly any temperature difference in the direction perpendicular to the measurement area, as mentioned above. Figure 11(a) shows the comparison of the temperature at a II, b II and c II in the basic pulse tube. The temperature is the mean of the measurement area. The two-dimensional measurement shows that there is hardly any phase difference between pressure and temperature. The low-temperature side oscillates only between 250K and 290K, while the high- temperature side oscillates between 255K and 305K. Figure 11(b) presents the temperatures of the heat exchangers (thermocouple). Tb (at the heat exchanger of pulse tube hot-end) has a clear phase difference to the pressure. And Tb has a large temperature variation ; however, Ta ( at the heat exchanger of regenerator hot - end ) has hardly any variation. Figure 12 shows the comparison of the temperature about a II, b II and c II in the orifice pulse tube. It shows that the phase difference between pressure and temperature of the high-temperature side is larger than that of the low-temperature side. The low-temperature side oscillates at the low range compared with the high-temperature side. And its amplitude is small like the basic type.
Figure 11. Temperature in the basic pulse tube.
Figure 12. Temperature in the orifice pulse tube.
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Figure 13. Opening turn of double inlet valve vs temperature variation.
The relation between the opening of the double-inlet valve and the temperature variation in another experiment is shown in Figure 13. It describes that the more the double-inlet valve is opened, the more the phase between the pressure and the temperature shifts. The result of mean temperature shows the phenomenon in the pulse tube more clearly. CONCLUSIONS
This paper describes the measurement of the gas temperature profile in the pulse tube by PLRSM using an excimer laser and the prototype of the pulse tube refrigerator with the a quartzglass pulse tube. The experimental results on three types of pulse tubes show that each has a different temperature profile. There is a large temperature gradient in the pulse tube and it is affected by the gas flow, namely the phase shifters. The results also show that a difference in
phase exists between the gas temperature and the pressure in each type. The temperature difference of about 50K was observed between the upper side and the lower side of the tube. Little variation in temperature is measured in the direction perpendicular to the measurement
area. Further research will be focused on quantitative analysis of pulse tube refrigerators by measuring temperatures at other points and the velocity in the vacuum , where the effect of gravity is canceled with the pulse tube perpendicular. REFERENCE
1. W.E.Gifford and R.C.Longsworth, " Pulse tube refrigeration", Trans.ASME Ser. B J. Eng. Ind., 63, (1964), p.264
2. E.I. Mikulin, A.A. Tarasov and M.P.Shkrebyonock, " Low temperature expansion pulse tube ", Adv. Cryog. Eng., 29, (1984), p.629 3.
S.W. Zhu, P.Y. Wu and Z.Q. Chen, " Double Inlet Pulse Tube Refrigerator - An Important Improvement", Cryogenics, 30, (1990), p.514
4.
Y.Matsubara, "Progress on Pulse Tube Coolers ", Jour.HTSJ, 36, (1997) , p.63
5. A.Tominaga, " Thermoacoustic Theory of Viscous Fluid , Part 1-Energy Conversion and Energy Flux of Small Cycles ", Cryogenic engineering, 27, (1992), p.543 6.
Y.Matsubara, J.L.Gao, K.Tanida, Y.Hiraasaki and M. Kaneko," An experimental and analytical
investigation of 4K pulse tube refrigerator", Proc. 7th Int. Cryocooler Conf., (1992), p.166
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7. W.Rawlins, R.Radebaugh, K.D.Timmerhaus," Thermal anemometry for mass flow measurement in oscillating cryogenic gas flows ", Rev. Sci.Instrum, 64, (1993), p.3229 8. Y.Hozumi, M.Murakami and Y.Yoshizawa, " Numerical Study of Pulse-tube Flow ", Cryogenic engineering, 33, (1998) , p.200 9.
M.Shiraishi, N.Nakamura, K.Seo and M. Murakami, "Visualization Study of Velocity Profiles and Displacements of Working Gas Inside a Pulse Tube Refrigerator ", Cryocoolers 9, (1997), p.355
10. S.Ito, Y.Hagiwara and S.Yatsuzuka," Visualization and Measurement of Gas Displacement in Pulse Tubes ", Cryogenic engineering, 31, (1996), p.481 11. S.Yatsuzuka, Y.Hagiwara and S.Ito, " Influence of Gas Displacement and Phase Difference on Pulse Tube Cold Side on the Performance of the Pulse Tube Refrigerator ", 33, (1998), p.242 12. N.Yoshikawa, T.Fujikawa, C.Niwa, K.Ohtake, " Raman and Rayleigh Thermometries in Flames ", J.JSASS, 31, 350 , (1983), p.30 13. Alan C.Eckbreth," Laser diagnostics for combustion temperature and species" ,Abacus Press, UK, (1988) , p.210 14. LaVision, LDS instruction Manual 15. S.Ito, Y.Hagiwara and K.Nara, "Study of Temperature Measurement in a Pulse Tube with Rayleigh Scattering", 33, (1998), p.225 16. Y.Hagiwara, K.Nara, S.Ito and T.Saito," Temperature Measurement in Pulse Tube with Rayleigh Scattering and Computation of Enthalpy Flow ", 33, (1998), p.233
Mathematical Model of a Wave Cooler V.N. Kukharenko
Department of Technical Cryophysics Kharkov State Polytechnical University Kharkov, 310002, Ukraine
ABSTRACT
A wave cooler with a dynamical system of cooling is considered in this work. A one-dimensional system of gas dynamic differential equations is used for describing the thermo-gas dynamic processes in the expansion tube and outlet manifold. These equations are solved by a method of finite differences with allowance for discontinuities. A mathematical model of gas flow in a branching manifold was developed. The model assumes that the sizes of the branches are commensurate with a step of the grid in the tube and that it is possible to average flow parameters throughout all the branches. We have conducted the analysis of a device consisting of a stationary nozzle system with a moving aperture, which is arranged co-axially with the expansion tube. SCHEMATIC OF A WAVE COOLER Thermo-gas dynamic processes that take place in a wave cooler with a dynamic system of gas distribution are considered in the work. Such a device1 provides a high adiabatic efficiency (65-
70%) and is shown schmatically in Fig. 1. There is one rotating element in such an apparatus — the gas distributor. Its rotational frequency changes within the limits of 20-100 Hz depending on the characteristics of the device. The gas to be cooled enters the gas distributor through the inlet pipe. In the inlet nozzels, part of the internal energy of the compressed gas is transformed into kinetic energy. As the gas
Figure 1. Schematic of a wave cooler where: 1- entrance pipe; 2- body case; 3- hollow axle with jets; 4- one of several expansion tube; 5- exit pipe; 6- encapsulation; 7- bearing.
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Figure 2. The wave cooler with rotating screen.
distributor rotates, gas is periodically injected into the half-closed expansion tubes. Gas periodically flows into the expansion tube, compressing and heating the gas in the tube. The incoming gas exchanges energy with the gas already in the tube. After the end of contact of the jet with the
entrance aperture of the cavity, gas flows to the distribution manifold with lowered temperature. The gas flows through the outlet pipe and is removed from the apparatus. The decrease of the outlet gas enthalpy is proportional to the amount of heat, which is rejected from the gas in the tube and into the environment through the walls. The cooling of gas in a wave cooler is defined by four items: 1) by the portion of the gas which flows into the tube and participates in the energy exchange, 2) by the gas which remains in the tube, 3) by the efficiency of the energy transfer from the inflowing gas into the pre-existing gas, and 4) by the portion of heat which is rejected to the outside environment. In this work we consider a device (Fig. 2) in which the nozzle, -6-, is stationary and placed co-axially with a stationary expansion tube, -3-. Flow of gas into the tube is carried out through the rotating screen, -1-, with eight apertures, -2-. The gap between the tube and the screen can be adjusted. For constructing a mathematical model, it is assumed that the device has a nozzle, an expansion tube, and a small inlet cavity. This cavity has an inlet aperture of fixed shape. Opposite the aperture, there is an expansion tube with a heat exchange surface at the far end. Gas outflow is carried through an outlet manifold. MATHEMATICAL MODEL
For the description of thermo-gas dynamic processes in the expansion tube an one-dimensional system of differential equations of gas dynamics is used:
The gas charge entering through the inlet aperture is defined by the formula for isentropic expansion. The process of expansion is complex: during expansion a longitudinal pressure wave
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407
occurs within the expansion tube as a result of the changing velocity vector. For a classical device, it is assumed that during the inflow to the expansion tube the area of the inlet aperture (nozzle) changes according to a predetermined function. The outflow of gas is controlled by a fixed set of parameters. The processes in the small expansion chamber can be described by the following system of differential equations2:
The thermal processes in the wall of a tube are described by the one-dimensional non-stationary equation of thermal conductivity:
with boundary conditions
The heat exchange correlation for a Hartmann-Sprenger device 3 is used in the expansion tube:
where parameters at a given point in the tube are averaged over a cycle. The following correlations are used for calculating the resistance factors:
For the solution of this system of partial differential equations a method of finite differences is used. A grid with time and space steps is chosen. Applying this grid to equations (1-3), a central difference scheme is built keeping the first order of approximation in t, time, and the second in h, space. After transformation it looks like:
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To maintain computational stability the following condition must be met:
where is the maximum pressure wave speed. The calculation of boundary values is carried out by coordination of characteristics of gas dynamics equations:
Using the method of finite differences according to the scheme of “predictor-corrector” ensures a second order of approximation on t. It is possible to present the filling-exhausting of the expansion tube in a device (see description of the second device below) as superimposed flows. This takes into account possible reflections of waves from the rotating screen which may occur during the gas outflow. Equations (47) are used for this purpose. The calculation cell for superposition is a small volume: where d is the diameter of the tube. To allow for breaks in the solution of the gas flow, a grid displaced on h/2 from the main grid is introduced. For definition of parameters for this grid, the Gugonio equation for problems with a discontinuity is used. RESULTS
The preceding algorithm was programmed in Pascal. Calculations of a device with the following parameters was made4: length of expansion tube - 2.26 m; diameter of the tube - 0.01 m; diameter of the nozzle - 0.01 m; working gas - air; maximum pressure - 0.24 MPa; minimal pressure - 0.1 MPa; frequency - 240 Hz. Two variants of inlet valve construction were considered. The first type has a valve, which is made by rotating a circular opening past another. The second type has a valve, which is made by rotating of a circulsr opening past a rectangular
opening.
MATHEMATICAL MODEL OF A WAVE COOLER
Figure 3. Gas temperature distribution along the length of tube for
409
Figure 4. Gas temperature distribution along the length of tube for
The calculated gas temperature along the length of the expansion tube is provided in Fig. 3 and Fig. 4, respectively, as a function of the working gas expansion ratio, and
Curves 1 and 2 indicate type 1 and 2 inlet valves, respectively. Experimental data4 are marked with dots. The shape of the calculated curves coincide with experimental data. With a small degree of expansion, the mathematical model of the device with the second type valve has lower error (8.3 % or 30 K instead of 11.1 % or 40 K). With an increase of expansion ratio, the error of the model increases (20.8% - for the second type valve, 25% - for the first type). This may be explained by the absence of reliable correlations for defining the non-stationary heat transfer coefficient in the tube. The influence of the heat transfer increases with increasing pressure ratio. The parameters of a second device with a type 2 valve are the following: length of the expansion tube - 0.09 m; diameter of the tube - 0.002 m; diameter of nozzle - 0.02 m; working gas - air; maximum pressure - 0.4 MPa; minimum pressure - 0.1 MPa; frequency - 800 Hz. Pressure and temperature changes in this device with a 0.5 mm gap are presented in Figs. 5 and 6. At the beginning of gas flow into the expansion tube, a compression wave (curve 1) is formed.
Figure 5. Pressure changes in the wave cooler with rotating screen.
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PULSE TUBE MODELING AND DIAGNOSTIC MEASUREMENTS
Fig. 6. Temperature changes in the wave cooler with rotating screen.
With the closing of an input aperture, the wave advances inside the tube. At the moment the outflow starts, the compression wave reaches the mid point of the expansion tube (curve 2). As the gas flow continues the generated compression wave reaches the closed end (curve 3). At the same time, at the inlet to the tube, an expansion wave forms. It causes a local downturn of temperature. After reflecting from the closed end, the compression wave starts moving towards the tube inlet (curve 4). On reaching the latter, it causes an increase in temperature of the outflowing gas (curve 5). After reflection of the compression wave from the rotating screen its intensity falls (curve 6). The amount of reduction of wave intensity depends on the gap size. The temperature profiles corresponding to these pressure profiles are presented in Fig. 6. To estimate the influence of the gap between a tube and a rotating screen on wave propagation in the expansion tube, calculations for various gaps in the 0.3-1.5 mm range were carried out. If the size of gap is big (1.5 mm), the inflow of gas into the tube occurs at a pressure close to the initial pressure. This means, that during outflow the pressure in the expansion tube has time to return to an initial condition. As gas exits, the intensity of the wave quickly falls. This indicates that the gas flow to the outlet manifold predominates over the reflection from the rotating screen into the tube. Reducing the gap size (0.3 mm) changes this relation. In this case, at the initiation of gas inflow, the pressure in the tube far exceeds the pressure in the outlet manifold (about three times). During gas outflow, the intensity of the reflected compression wave from the rotating screen is not less than the intensity of the falling compression wave. Thus, the residual pressure in the tube is much higher than the minimal pressure. The temperature at the closed end of the expansion tube is higher if the gap size is small. This is explained by the increase of the average temperature of the gas in a tube as the gap size is reduced. This increase is caused by the increase of residual pressure in the tube. On the other hand, with the growth of residual pressure in the expansion tube, the ratio between pressures in the expansion tube as a function of the phase decreases. This results in a reduction of intensity of the wave. Hence, in a real device, two opposite processes work: formation of a gradient of temperature, provided by the wave process, and growth of the average level of temperature as the result of the residual pressure in the tube. In conclusion, there is an optimum gap size that ensures the maximum adiabatic efficiency. The calculations show that the maximum adiabatic efficiency is observed with the gap size - 0.4 mm and reaches - 0.25.
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REFERENCES 1. Jian Shao, Yudi Bao, Yongnian Shen, Yangpu Feng, “Experimental Investigation of a New Type Expander,” Advances in Cryogenic Engineering, 1986, v.31, pp. 685-692. 2. Gogrichani, G.V., Shipilin, A.V., “Transitional Process in the Pneumatic Systems,” Moskow: Engineering, 1986, 160 p. 3. Brosher E., Maresca G., “L‘etude des phenomenes theriques dans in tube Hartmann-Sprenger,” Int. J. of Heat and Mass Transfer, v.16, 1973. 4. Bobrov, D.M., Vasilev, Y.N., Laukhin, Y.A., Sirotin, A.M. and Chelikidi, L.M., “Using the Gas
Pulse Devices in Gas Industry,” Summary Ser.: Preparing and Utilization of Gas and Gas Condensate - Ì: VNIIE gasprom, 1985, v.7, 58 p.
Pulse Tube Modeling as a Means of Teaching the Design of Cryogenic Refrigerators V. N. Kukharenko
Department of Technical Cryophysics Kharkov State Polytechnical University Kharkov, 310002, Ukraine
ABSTRACT
A design manual titled “Low Temperature Pulse Tube Coolers” has been developed for use in a university course on the design of cryogenic machines. It allows students to be involved in the entire design process by using simple mathematical solutions of the rather difficult physical processes that exist in cryogenic gas machines with pulse tubes. Using the manual, students fulfill a real design of a pulse tube cooler. The manual consists of two parts. The first part treats a variety of topics including: 1) details of the thermal-physical processes involved and their mathematical modeling, 2) the "method of characteristics" for solution of the equations of gas dynamics, and 3) the definition and discussion of the principal items of cooler losses. For performing calculations, the students use object oriented programming, since examining pulse tube modifications parametrically has group properties. The construction of a model for optimization permits the students to find the optimum solution among all device modifications. The check of a chosen optimum device with appropriate parameters can be conducted with a more complex mathematical model, which is presently under development. Skill at reviewing and understanding the technical literature and, in particular, the patent literature has a large significance in designing. In the second part of the design manual one can find a literal description of the existing patents and tasks for the development of skills for the analysis of such materials. The design process for a given technology can last several semesters and includes elements of scientific research work. To reduce the designing course length, "design teams" can be formed; here a group of students with various participating roles fulfill the work. The group can consist
of the students of 3rd-5th years of learning. INTRODUCTION
Designing of systems is a multi-discipline process in which all qualities of a designer are displayed. He (or she) should understand the physical process quite well, know how to construct the whole series of mathematical models of various degrees of complexity, choose a method of calculation, be sure of the adequacy of the mathematical model, etc. During the process of training, a student should be involved in all design stages, and his qualification should depend on learning each stage. Wide application of computers during the Cryocoolers 10, edited by R. G. Ross, Jr.
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design process can be detrimental if it is not also accompanied with a deep understanding of all aspects of the process. The student can acquire an illusion that it is enough for him to know how to use a computer and to find the appropriate programs — and that this fully determines the design process. In this work, we propose to teach design through the use of a design project involving a Low Temperature Cooler with Pulse Tube (LTCPT). The Purpose of the training includes: • To study the peculiarities of the physical processes involved with LTCPT modifications. • To construct mathematical models describing the thermal-physical processes. • To know how to determine the main characteristics of the device through study of the analytical results. • To master numerical methods for the solution of mathematical-physics equations. • To learn to use object-oriented programming for drawing up the programs for calculation of the characteristics of LTCPT with various modifications. • To carry out computing experiments under the developed programs, and to do analysis. • To learn to build mathematical models that contain the necessary parameters. • To learn to analyze patterns and to build elementary mathematical models for the analysis of integrated characteristics. • To carry out drawing work in an environment like AutoCAD, observing necessary standards. The defined program's use of simple mathematical solutions of the rather difficult physical processes of various LTCPT designs allows a student to progress through the complete design process in a rather short term. MODEL PROBLEM
To become acquainted with the pulse tube, the student begins with a modeling problem describing the filling of a long cylindrical cavity under adiabatic conditions. The process is described by the equations of gas dynamics including the equation of preservation of mass:
and the equation of energy:
where gas density; w = velocity of gas; p = pressure; k = ratio of specific heat; t = time; and x = length coordinate. At the initial moment of time, and the temperature of the gas in the tube is constant and equal to With the change of pressure from at up to at the temperature at the entrance remains equal to The solution of this system of equations is by the "method of characteristics", where an equation of characteristic, and is ratio of characteristics. The solution shows that upon filling of the tube there is a gradient of temperature over an interval from up to where length of tube and The temperature in the remainder of the tube is equal to The change of temperature in the first stage is equal to This effect can be used for achieving low temperatures in cryogenic gas machines with a
pulse tube. The degree of cooling is determined by the quantity of energy transferred to the environment from the heat exchanger. To achieve a given temperature level, a regenerator is used. It is possible to achieve the pulsation of pressure in such a system using either an internal or external compressor.
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Figure 1. Basic low-temperature pulse tube cooler with: 1) compressor, 2) heat exchanger, 3) regenerator 4) cold heat exchanger, 5) pulse tube, and 6) hot heat exchanger. BASIC PULSE TUBE
W.E. Giffold was first to offer such a device in 1961 (Fig. 1). The study of the device begins with the construction of a physical model and determination of the ideal processes. Therefore, the process of transferring heat from the cold end of the tube to the warm end, and then to the heat exchanger is considered here in detail. Two ideal processes at work are identified as gas heat pumping and surface heat pumping. In the first case, heat pumping is conducted by the oscillatory movement of gas between the cold heat exchanger and the hot heat exchanger. It is possible to consider this process, in the ideal case, as adiabatic. In the second case (surface heat pumping) the process is conducted by oscillatory movement of gas within the cold heat exchanger — tube-and-tube — and hot heat exchanger with participation intermediate heat capacity body — wall of tube. In this ideal model the cycle consists of four periods: 1) a period of compression with adiabatic flow of gas in the tube; 2) a period of cooling, which proceeds at constant pressure, the gas accepts the temperature of the wall, 3) a period of expansion of the gas involving adiabatic flow of gas in the tube, and 4) a period of heating at
constant pressure as the gas accepts the temperature of the wall.
The simplified models described above let one analyze the processes associated with various LTCPT modifications, including the various ways that energy transfer to the environment can
take place. The next stage of the design process is construction of mathematical models to allow modifications to be explored, and choice of a method of solution. MATHEMATICAL MODEL
It is known that the processes in the LTCPT are relatively well described by one-dimensional equations of gas dynamics for regenerators and heat exchangers, and by a system of ordinary differential equations for elements with concentrated (average on volume) parameters. However, such a system of equations requires the use of rather difficult numerical methods and their use in the educational process causes certain complexities. For the development of a mathematical model describing the processes in the LTCPT, it is
assumed that the process of heat exchange in the regenerator is ideal. This lets us exclude from consideration the compressor/regenerator part of the system. The processes taking place in the left part of the system are considered with the following assumptions: 1) the gas is ideal, 2) the hydraulic losses in the tube are absent, 3) the tube has a constant area of cross section, 4) heat exchange in the cold heat exchanger and the hot heat exchanger is ideal, and 5) there are no turbulent processes in the tube. In this case, it is possible to use the following system of equations.1,2 The processes proceeding in the adiabatic receiver, installed behind the hot heat exchanger are described by equations of balance of mass and energy
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where = = = = = =
mass of gas in receiver internal energy of the gas in receiver mass flow rate at the input to the hot heat exchanger specific heat at constant pressure gas temperature at the input to the hot heat exchanger temperature of the hot heat exchanger
It is thus assumed, that the volume of the hot heat exchanger, located between the tube and the reservoir, is very small. In the case of no reservoir, the processes in the hot heat exchanger are described by
where volume of a hot heat exchanger, and density of gas in the hot heat exchanger. At the interface location between the tube and the hot heat exchanger, the pressure of the gas in the hot heat exchanger is related to the pressure in the tube, during the introduction of
gas into the hot heat exchanger, by
and at the exhausting of gas from the hot heat exchanger, by
where = local resistance, = gas pressure in the hot heat exchanger, = gas velocity in the tube at the input to the hot heat exchanger. The temperature of the wall of the tube in the first approach changes linearly with length1
where = temperature of cold heat exchanger; The refrigeration power of the LTCPT is determined as the quantity of energy transferred to the environment
where = period time. For an estimation of the efficiency of the cycle, it is possible to use adiabatic efficiency.
Reduction of the equation of energy to a nondimensional form gives
where allows a choice of the theoretical cycle.
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Figure 2.. Area of model application, where 1) 2) 3) experiment 2; 4) experiment 3 5 4 ; 5) experiment ; 6) experiment . Area between curves 1 and 2 corresponds to an adiabaticisobaric model; below curve 1 is the adiabatic model; above curve 2 is the isothermal model.
At
it is possible to neglect heat exchange in the tube, at
the process in the
tube is isothermal. An estimation of the modes of device operation (a laminar mode of gas flow
in the tube, shows, that and for the description of processes in the device it is necessary to use an adiabaticisobaric cycle. The condition of applicability of different models for the description of processes in the LTCPT can be estimated according to the expression
where
R = gas constant;
= hot
heat exchanger volume; = pulse tube volume; = viscosity. The results of calculations using this correlation equation and data of experimental devices are presented in Fig. 2. SOLUTION
For the solution of an adiabatic-isobaric model the method of characteristics is used. Equation (1) has one family of the characteristics, on which the following ratio is observed: The equation of the characteristics of an obvious type can be achieved by integration of the equation together with Eqs. 2 and 5
where
= beginning of characteristic.
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The characteristic, taking place through the beginning of coordinates, divides the area of the solution of the problem with filling on two parts: area of influence of initial and boundary conditions
where If the refrigeration power of the device is determined only by surface heat pumping. In this case the solution for all modifications is of a rather simple form
where for the device with a membrane; for the other devices Z, M are constant, which depend from geometrical and operational parameters LTCPT with a membrane.
The last component in brackets determines the quantity of mechanical work performed. In the case the refrigeration power of the device is determined by both surface pumping, and gas heat pumping. For determination of the gas temperature at the input to the hot heat exchanger, it is necessary to solve an ordinary differential equation concerning
with initial condition It is necessary to take into account the solution obtained from the calculation of the integral in the expression for determination of cold production. Using this generalized approach, it is possible to determine the correlation equations for calculating the theoretical refrigeration power associated with LTCPT modifications. For example, for the Gifford’s device at the theoretical refrigeration power is determined by the approximated correlation equation:
Characteristics of the offered method allows one to use object-oriented programming. Emergence of a new device results in creation of new object with several changed methods. It lets students observe the advantages of the object-oriented approach to programming.
In addition, while acquiring LTCPT designing experience, a student gets acquainted with finite difference methods. The difference scheme (implicit left- and right-handed) appears to be rather simple and takes into account direction of movement of gas flow. The solution lets a
student conduct research on stability, approximation level (choice of steps in time and space), and to determine a condition of solution convergence. For fulfillment of research, it is necessary for a student to create a database on the available literature, including all presently available experimental data. The real refrigeration power of any LTCPT, as researchers have shown, can be determined as a kind of a difference of a theoretical refrigeration power and sum of its losses. Among the most important losses it is possible to single out losses in the regenerator, the influence of nonideal conditions of heat exchange in the hot heat exchanger, and nonlinearity of the distribution of temperature along the length of the wall of the pulse tube. Although the course manual presents all necessary correlation equations, it is desirable to search for additional information in the scientific literature. The presence of rather simple correlation equations for determination of
refrigeration power lets one construct a mathematical model that allows understanding of the parameters of the device and optimization.
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As the last stage of the process of designing, the student is asked to conduct an analysis of the patents pertaining to the LTCPT, from which there is little experimental data. For this purpose it is necessary to use only patent descriptions, the text of which, as a rule, differs significantly from that accepted in the scientific literature. The analysis of the new device assumes construction of a mathematical model of the theoretical cycle, achieving analytical correlation equations (if it is possible), construction of a new object on an algorithmic language, fulfillment of computational experiments, and comparison with existing devices. DESIGN TEAMS
The closing stage of the design process is generating drawings of the AutoCAD type of the device. The above considered LTCPT design process can be implemented as a team or group activity as this is consistent with the modern use of product development teams, and promotes development of designing skills, and takes into account the level of preparation of students and their available time. One of the main principles of scientific organization of designer’s work is division and specialization. With reference to designing through teams, organization of creative design groups allows the fulfillment of complex projects of increased complexity. The qualified specialization of labor, based on assigning work of a certain complexity to a separate executor, is made inside the student group taking into account the level of preparation of each student and his aptitude to fulfil a certain kind of job. With team based designing there is also opportunity for parallel development of separate parts of the project, increase of a level of knowledge, and expansion of the outlook of developers. A key essential is that time spent on technical, routine work decreases, thus releasing time for creative processing of important parts of the project, and acquiring deep knowledge on the most detailed questions of the project. With team based designing it is necessary to nominate the minimum list of jobs included in the structure of the project, to divide them into separate stages, and to identify key milestones. The list of problems for solution is made out as a kind of technical project on designing, and theme of work as a network diagram, based on the algorithm of designing through teams. It can be seen that the work of educational design teams is built on principles similar to the activities of industrial organizations engaged in the design process. Successful designing through teams essentially depends on the level of preparation of the students and the motivation in the group. For creation of the necessary level of motivation, voluntary formation of groups and distribution of roles is used. REFERENCES
1. Narayankhedkar, K.G., Main, V.D., "Investigation of Pulse Tube Refrigerator," Trans. of ASME, vol. 95 (1973) pp. 300-307.
2. Gifford, W.E., Longsworth, R.S., “Pulse Tube Refrigeration,” Advances in Cryogenic Engineering,
vol. 10, Plenum Press, New York, 1964, pp. 69-79. 3. Gifford, W.E., Longsworth, R.S., “Surface Heat Pumping,” Advances in Cryogenic Engineering, vol. 11, Plenum Press, New York, 1966, pp. 171-181. 4. Tarasov, A.A., Mikulin, E.I., Kuznetsov, B.G., “Low Temperature Pulse Tube,” Chemical and Oil Engineering, 1973, pp. 43-44. 5. Mikulin, E.I., Shkrebenok, M.P., Tarasov, A.A., “Process Investigation in Work Area of Pulse
Expansion Machines,” Machines and Apparatus of Cold and Cryogenic Technique and Air Condition, Leningrad, 1986, pp. 74-80.
Design and Test of Low Capacity Reverse Brayton Cryocooler for Refrigeration at 35K and 60K J. A. McCormick, G. F. Nellis, W. L. Swift, and H. Sixsmith Creare Incorporated Hanover, NH, USA 03755 J. Reilly Air Force Research Laboratory Kirtland AFB, NM, USA, 87117
ABSTRACT A single stage reverse Brayton cryocooler being developed for the cooling of spaceborne surveillance sensors will provide 1 W of cooling at 35 K with less than 100 W of input power. This application is an extension to lower power levels of the technology of the 5 W, 65 K reverse Brayton cryocooler presently under life test at the Air Force Research Laboratory. The reverse Brayton technology features miniature high performance turbomachines running on self-acting gas bearings. These machines provide high reliability, long life, and negligible vibration. The critical components of the reverse Brayton cooler are a centrifugal compressor, high effectiveness recuperator, and cryogenic expansion turbine. The 35 K cooler uses advanced designs for the compressor and turbine in which high energy permanent magnets in the rotating elements provide electromechanical energy conversion. A turboalternator operating at cryogenic temperature converts the turbine power from the expansion of the gas to electric power that is dissipated through a resistive load at room temperature. The miniature centrifugal compressor is driven by a high efficiency three-phase AC motor using features that are similar to those of the turboalternator. The recuperator will use either the slotted plate heat exchanger of the 5 W, 65 K SSRB cooler, or an advanced lighter weight parallel plate heat exchanger. This paper presents the component and system designs and performance predictions for the 35 K cooler. Results of recent system tests with brassboard components are also presented. INTRODUCTION There are a number of critical requirements which drive the design of mechanical coolers for long term space applications. The more important ones are: • high reliability, • long life, • high thermodynamic efficiency, • very low vibration, • flexibility in packaging and integration, Cryocoolers 10, edited by R. G. Ross, Jr. Kluwer Academic/Plenum Publishers, 1999
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• low weight, and • simplicity and robustness. Among the candidate systems currently being pursued, the turbomachine-based reverse Brayton cryocooler (TRBC) offers the promise of meeting each of these requirements over a broad range of temperatures and loads. For cooling loads in excess of 5 to 10 W at temperatures above 4 K, TRBCs that achieve overall performance levels of 10% of Carnot efficiency or more can be built with today’s technology. Existing gas bearing technology provides for no-wear, vibration-free operation without the need for auxiliary counter balancing. Simplicity is inherent in these systems in that there is only one moving part in each machine, a very low mass rotating shaft, which rotates continuously at very high speed. Because of the high degree of inertial balancing that is required for successful high speed operation, there is no detectable mass vibration. Depending on the particular hardware and performance requirements, the required control system can be very simple, increasing the reliability of the primary life limiting element in these systems, the electronics. The high power density of turbomachines makes them very attractive in terms of system weight and packaging, and an important attribute of this system is that it is component based. Each of the components in the system is connected within a fluid loop. This modularization of the system allows for development and optimization at the component level, flexibility in packaging, and integration of individual components or subsystems with other reverse Brayton components or with cooling stages consisting of totally different systems. The principal challenge in adapting this technology to very low power space applications has been that of cycle efficiency. This is a consequence of the fact that the relative effect of parasitic losses in turbomachines increases as cycle capacity and size are reduced. In order to improve the performance of TRBCs so input power meets acceptable levels for very low cooling capacities, the sizes of the components must also be reduced without sacrificing component efficiencies. The four major components that are used in the single stage TRBC include: • An electronic controller and converter drawing unregulated DC power from the spacecraft bus and converting it to a controlled power output to drive a compressor, • A compressor that consists of a drive motor and centrifugal compressor to pressurize and circulate the flow through the cycle, • A counterflow heat exchanger or recuperator that pre-cools the high pressure gas flowing to the cold end of the cooler, and • A turbine that provides the net refrigeration for the cycle. Their arrangement in the cycle is shown schematically in Figure 1. The performance of each component can be characterized by an efficiency or effectiveness that is a function of the component’s size, overall design configuration, materials used, and the operating parameters such as pressure, temperature, and flow rate. The performance of the compressor and its electrical drive system is characterized by the power train
Figure 1. Schematic arrangement of a single stage turbomachine-based reverse Brayton cryocooler (TRBC).
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efficiency, which is the ratio of the isentropic compression power to the electrical power drawn from the spacecraft bus. This is the product of individual efficiencies for the compressor, motor, and electrical drive. The performance of the recuperator is characterized by its thermal effectiveness, which is the ratio of the isentropic compression power to the electrical power drawn from the spacecraft bus. This is the product of individual efficiencies for the compressor, motor, and electrical drive. The performance of the recuperator is characterized by its thermal effectiveness, which is the ratio of the actual to the maximum possible heat transfer between the high pressure and low pressure streams. The performance of the turbine is determined by its net efficiency, which is the ratio of its actual cooling power to the power available in an isentropic expansion. The power train efficiency, recuperator effectiveness, and turbine efficiency, along with the temperature and pressure ratios of the cycle and gas properties, determine the input power requirement for a given cooling load.
A useful benchmark for the desired low power system is a single stage TRBC presently under life test at the Air Force Phillips Laboratory.1, 2 This cooler provides 5 W of cooling at 65 K with approximately 215 W of input power at a 280 K heat rejection temperature. This corresponds to a cycle efficiency of approximately 7.5% of Carnot. The recuperator is a slotted plate heat exchanger; the compressor uses a solid rotor induction motor and the turbine consists of a turboexpander unit, in which the power transferred to the shaft by the expanding gas is absorbed by an aerodynamic brake at the heat rejection temperature. Because of the nature of the losses in the induction motor and the turboexpander, the input power to the system does not scale down proportionally with a decrease in refrigeration load. The electrical efficiency of the induction motor decreases as it is scaled down in size and power, resulting in an unacceptable decrease in power train efficiency. As the turboexpander is scaled down in size and power, the conductive heat leak from the brake to the turbine, which is approximately 2.5 W for the 5 W cooler, could decrease only slightly, actually becoming larger as a fraction of the turbine power. This would unacceptably decrease the net turbine efficiency. As a result, the input power to the system at a refrigeration load of 2 W would decrease only to about 165 W corresponding to a cycle efficiency of about 4% of Carnot.
To address these limitations, smaller components are being developed for the system. These components are being designed so as to provide efficiencies that are comparable to the versions used in the larger 5 W cryocooler. Turboalternator
A turboalternator is being developed in which the turbine power is absorbed by a miniature electrical generator. Current from the generator flows through leads of low thermal conductance to a resistive load at the heat rejection temperature. The entire turboalternator assembly operates at cryogenic temperature. This eliminates the conductive heat leak that is present in the housing of the turboexpander, leaving a smaller parasitic loss comprised of bearing and rotor friction and electrical loss. The parasitic loss is on the order of 1 W, versus 2.5 W for the turboexpander. The rotating assembly in the turboalternator is comprised of a radial inflow turbine rotor and a permanent magnet generator rotor. The permanent magnet is a solid cylinder of high energy material, magnetized across a diameter. It is contained in a hollow titanium shaft which is machined in one piece with the turbine rotor. A toothless stator provides high power density and
avoids radial magnetic forces on the gas bearings. Figure 2 shows an isometric view of this machine configuration. The radial turboalternator rotational speed is less than one-half that of the turboexpander, which improves the stability of the gas bearings at low temperature and reduces their drag.
Compressor
If the stator, in a machine mechanically identical to the turboalternator but somewhat larger, is driven as a synchronous motor, the machine becomes a centrifugal compressor well matched to the requirements of the desired low capacity TRBC. Importantly, the permanent magnet synchronous
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motor has a dramatically higher electrical efficiency than the induction motor at the low power levels of interest. It achieves higher efficiency by avoiding the induced rotor currents and iron teeth that are responsible for major losses in the induction motor. It also provides higher power density than the induction motor, allowing the shaft diameter to decrease, and, consequently, bearing drag loss to decrease, further increasing the power train efficiency. Recuperator
Although the slotted plate heat exchanger used in the 5 W system provides equally high effectiveness at lower power levels (approximately 0.99), its large size and weight can be a disadvantage in some applications. Work is underway on a novel parallel plate configuration that will be substantially smaller and lighter for the same effectiveness. Preliminary studies and subscale tests on the parallel plate concept indicate that the weight may be one half that of the slotted plate heat exchanger. An added attraction of the parallel plate concept is its potential scaleability to systems of even lower capacity or lower temperature. The following sections of this paper describe the development and test of a brassboard low capacity TRBC system that uses a turboalternator and permanent magnet motor compressor (PMMC). The brassboard versions of these machines have evolved from previously reported demonstration models3.
BRASSBOARD LOW CAPACITY COOLER Figure 3 shows a schematic of the single stage reverse-Brayton system indicating the major energy flows and thermodynamic state variables, all of which are measurable in the brassboard facility. Important measurement parameters are the load temperature the net refrigeration the input power and the heat rejection temperature The nominal performance target for the system is 1 W of refrigeration at 35 K with less than 100 W of input power and heat rejection at 300 K. Recent developments on the PMMC and turboalternator, with their integration in a brassboard test system, have demonstrated performance levels that are close to meeting these goals, and we expect to surpass them with refinements that are presently being implemented.
Figure 2. Mechanical assembly of turboalternator.
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Figure 3. Schematic of reverse Brayton test assembly.
Reference 3 described early test results on a brassboard turboalternator and a subscale PMMC. Since that time, a full scale brassboard PMMC has been developed and major improvements have been made in the turboalternator. These two turbomachines have been tested at the component level to evaluate shaft/bearing and electrical performance at cryogeic temperatures in the configuration shown in Figure 3. The turboalternator features aerodynamic rotor channels cut by electrical discharge machining (EDM) and an advanced alternator design. The compressor has similar EDM rotor channels and stator, but in a slightly larger size. The two machines have proven
to be well matched aerodynamically at system input powers in the 90 to 150 W range. An earlier version of the turboalternator was matched to a cooling capacity of about 7 W at Figure 4 shows a wound alternator stator and typical shaft with the rotor sectioned to show the EDM channels. The turboalternator has a rotor diameter of 0.30 inch and a shaft/bearing diameter of 0.14 inch. It operates at a range of speeds between 3000 to 5000 rev/s. The PMMC has a rotor (impeller) diameter of 0.50 inch and a shaft/bearing diameter of 0.17 inch. Its operating speed is from 7000 to 9000 rev/s. Figure 5 is a photograph of the turboalternator installation at the cold end of the brassboard cryogenic test facility This was taken with the vacuum vessel removed and prior to wrapping the cold end with MLI. The large vertical cylinder is the slotted plate recuperator. The turboalternator is oriented horizontally next to the recuperator. Three instrument flanges containing platinum resistance thermometers (PRTs) are visible in the tubing between the turboalternator and the recuperator. These sensors measure temperature of the gas at the turbine inlet, turbine exit, and at the exit of the load heater. Electrical leads penetrating the turboalternator housing are three generator power leads near the aft end and two shaft sensors on the top. The power leads run to a three phase load rheostat located outside the vacuum vessel.
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Figure 4. Turboalternator stator winding and shaft with EDM channels.
Figure 5. Turboalternator in cryogenic test installment.
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Performance tests on a turboalternator consist of measurements of temperatures, flow rate, and pressures over a range of speeds and flow rates. Speed is fixed by the setting of the alternator load
resistance, and flow rate is adjusted by changing system pressure or by throttling the flow from the compressor. Both quantities are easily adjustable in the facility. Net turbine power is calculated from the voltage
and current
measured at the load resistance according to:
where and are peak to peak readings. Turboalternator efficiency is the ratio of isentropic power which is calculated from measurements of the inlet temperature flow rate m, and pressure ratio Pr, according to:
to the mass
where
is the heat capacity at constant pressure and is the ratio of specific heat capacities. This efficiency includes all losses within the machine, but does not account for parasitic losses from heat exchange with ambient radiation and conduction down the leads which must be accounted for separately. Table 1 lists two sets of test results from an initial build of the turboalternator. The first column gives results for a configuration that was designed to replace the turboexpander used in the 5 W, 65 K cryocooler. Following the successful demonstration of this machine, minor modifications were made to the assembly to evaluate performance at a lower flow rate. The second column lists initial results for this machine. Further development is being performed to optimize the turbine at lower flow rates to match a design cooling condition of 1 W at 35 K. Performance tests were also conducted on a PMMC in the brassboard facility. Tests were performed at several speeds. Flow rate was throttled to produce curves of pressure ratio as a function of flow rate. These data were reduced to show the relationship between power train efficiency and dimensionless flow coefficient
where is the flow velocity at the inlet to the compressor, is the tip velocity of the impeller, is the isentropic increase in enthalpy across the compressor, computed from the inlet temperature and the pressure ratio, m is the mass flow rate, and I and V are the DC current and voltage to the electronic power supply. Figure 6 shows the results of these tests on a recent version of the compressor. The maximum power train efficiency is about 0.42 at a flow coefficient of 0.19. The two sets of data are derived from tests at two speeds – 8000 rev/s and 8500 rev/s. The peak efficiency value is greater than the maximum power train efficiency that had ever been achieved on the larger 5 W, 65 K cooler. It occurs for two compressor power levels that were tested. For the flow coefficient of 0.19, these power levels were about 125 W for the 8500 rev/s case and 110 W at 8000 rev/s. This result verified the major improvement in efficiency of the motor at the lower power condition.
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Figure 6. Permanent magnet motor compressor (PMMC) test results
Work is in progress to optimize the geometry of the flow passages in the turbine and the compressor so that the maximum efficiency of each machine is matched to a common flow rate corresponding to a 1 W load at 35 K. System modeling calculations based on the test results above
and the geometry modifications show that the electric input power to the cooler at this condition will be less than 100 W.
CONCLUSIONS Major advances have been achieved in reducing the size of turbomachines used in cryogenic refrigerators. A new machine construction, employing permanent magnet rotors operating in gas bearings at cryogenic temperatures, has facilitated the achievement of efficiencies in these devices that are required for TRBCs for lower cooling loads. Test results have confirmed that
these machines can be successfully operated and that efficiencies can be achieved that are comparable to machines twice their size and capacity. This new technology will extend the practical use of TRBCs to refrigeration applications where input power levels below 100 W are desired with relatively high cycle efficiency. Because this basic machine construction is suitable for cryogenic temperatures, and for operation as a turbine or compressor, new cycles may be practical where cryogenic circulators or cryogenic compressors are contemplated.
ACKNOWLEDGMENTS We would like to acknowledge support from NASA/GSFC, the Air Force Research Laboratory, and BMDO in the development of these components.
REFERENCES 1. Swift, W. L., “Single Stage Reverse Brayton Cryocooler: Performance of the Engineering Model in the 8th International Cryocooler Conference”, Cryocoolers 8, Plenum Press, New York (1995), pp.
499-506. 2. Dolan, F. X. et al. “A Single Stage Reverse Brayton Cryocooler: Performance and Endurance Tests
on the Engineering Model in the 9th International Cryocooler Conference”, Cryocoolers 9, Plenum Press, New York (1997), pp. 465-474.
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3. McCormick, J. A., et al. “Progress on the Development of Miniature Turbomachines for Low Capacity Reverse Brayton Cryocoolers in the 9th International Cryocooler Conference”, Cryocoolers 9, Plenum Press, New York (1997). 4. Nellis, G. F., et al. “Design and Test of Reverse Brayton Cooler for NICMOS Instrument Dewar on HST”, Presented at the 10th International Cryocooler Conference, Monterey, CA. May 26-28, 1998.
Reverse Brayton Cryocooler for NICMOS G. Nellis1, F. Dolan1, J. McCormick1, W. Swift1, H. Sixsmith1,
J. Gibbon2, and S. Castles2
1 Creare Inc. Hanover, NH, USA 03755 2 NASA Goddard Space Flight Center Greenbelt, MD, USA 20771
ABSTRACT
A Single Stage Reverse Brayton (SSRB) cryocooler is planned for installation on the Hubble Space Telescope (HST) during the third servicing mission in the year 2000. The cooler will
provide refrigeration to the NICMOS (Near Infrared Camera and Multiple Object Spectrometer) instrument cryostat which is presently cooled by the sublimation of solid nitrogen. A thermal short is causing the solid nitrogen to be consumed more rapidly than expected, reducing the useful life of the instrument to about two years or less. Implementation of the cryocooler will extend the life of the instrument for a period of five to ten years. The design of the cooler is based on the 5 W 65 K engineering model version presently under life test at the Air Force Research Laboratory at Kirtland AFB. The NICMOS cooler will provide approximately 7 W of refrigeration at 70 K with 300 W of electrical input power. Key components of the cooler are a centrifugal compressor, high effectiveness recuperator, and an expansion turbine. Except for external structural features, the compressor and recuperator are identical to their engineering model versions. The expansion turbine is a turboalternator that uses a permanent magnet generator to absorb power from the turbine. A miniature centrifugal circulator pumps cryogenic neon between the cryocooler and the
NICMOS cryostat. The neon absorbs approximately 7 W of heat, including parasitic loads, which is transferred to the cryocooler through a heat exchanger called the cold load interface (CLI). At the warm end of the system, the 300 W input power is rejected to a capillary pumped loop and external radiator. This paper presents a description of the cooler, initial test results, and a summary of planned pre-launch tests. INTRODUCTION
The Near Infrared Camera and Multi-Object Spectrometer (NICMOS) was built by Ball Aerospace and was installed on the Hubble Space Telescope (HST) during Service Mission # 2 in February 1997. It provides the HST with near-infrared and limited spectrographic capabilities in the 800 nm to wavelength range1. As originally designed, the instrument uses the sublimation of solid cryogen to maintain the detectors at a temperature of about 58 K. The solid nitrogen is enclosed in a dewar in such a way that the detector is cooled by the nitrogen Cryocoolers 10, edited by R. G. Ross, Jr. Kluwer Academic/Plenum Publishers, 1999
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through an aluminum conductive matrix. This conductive matrix also incorporates cooling tubes that were used during pre-launch operations to maintain the nitrogen in a solid condition. Bayonet connections are available at a valved interface manifold on the cryostat to permit the connection of helium cooling lines during pre-launch cooling. A thermal short has developed in the dewar assembly that has increased the rate of nitrogen depletion. This short will decrease the operating life of the instrument. Cryogen depletion is expected in 1999. In order to extend the life of this instrument, NASA has embarked on a development effort to install a mechanical cryogenic cryocooler on the HST that will provide cooling to the NICMOS instrument. The cryocooler is a closed loop reverse Brayton system using turbomachines for compression and expansion. The system also incorporates a circulation loop that will be pressurized on orbit and will interface to the cryostat at the existing bayonet connections. The system has been designed, fabricated and tested in a 14 month time period. It will be further tested on a shuttle flight demonstration. This flight demonstration is called the Hubble Orbital System Test (HOST) and is scheduled for late October 1998 (STS-95). Following the flight demonstration, the system will be installed on the HST in early 2000.
NICMOS COOLING SYSTEM (NCS) The NICMOS Cooling System includes three fluid loops that transport heat from the NICMOS instrument to space. A schematic of the system is shown in Fig. 1. The circulating loop includes a centrifugal circulator that pumps neon between the cryocooler and the NICMOS instrument. Neon is pumped in a continuous steady flow through the cryocooler’s Cold Load Interface (CLI) to the NICMOS cooling coil through flexible tubes that are to be connected to the dewar at bayonet fittings on an existing bayonet coupling interface. The circulating loop will be pressurized with gaseous neon on orbit following the connection to the bayonet couplings by the astronauts. The cryocooler loop consists of a centrifugal compressor, a recuperator and a turboalternator that conveys heat from the CLI to the Heat Rejection Interface (HRI) at the warm end of the cooling system. This portion of the system is pressurized with neon and hermetically sealed prior to launch. Purge and fill valves facilitate charging the system and cleaning it using a circulating high temperature bakeout process. The accumulator in the cryocooler loop allows the system to be charged to relatively low pressure when warm. Heat is removed from the system by means of a Capillary Pumped Loop (CPL) supplied by Swales Aerospace Inc.2 The fluid used in the CPL is ammonia. The NCS is designed to meet the specific constraints imposed by the HST, the NICMOS instrument and the flight demonstration. These constraints include thermal loads, packaging, mechanical support and heat rejection limitations. The NCS is designed as a package that can be installed by astronauts in the aft shroud of the HST. Because of the existing HST configuration, the NCS package will be separated from the bayonet coupling interface on the NICMOS dewar by the two flexible tubes, each of which is about 1 m long. In total, the circulating loop includes about 5 m of tubing in the flexible and fixed portions of the loop. The heat load at the NICMOS instrument is about 400 mW. However, the parasitics at the cold end of the system are substantial. At least 7 W of cooling must be provided by the cryocooler to overcome conductive losses from the bayonets at the bayonet coupling interface, radiation from the flexible lines, and conductive losses through the mechanical support structure for the cooler in its frame.
Circulator Loop The circulator loop consists of the centrifugal circulator, the cold load interface (CLI) with the cryocooler loop, stainless steel flexible tubes that interface between the cryocooler and the NICMOS dewar, bayonet couplings at this interface, and the internal cooling coil that transports fluid between bayonet couplings and the NICMOS detector. The loop contains high pressure and low pressure fill tanks. The low pressure tank is used to initially charge the loop. The high pressure tank is designed to compensate for long term leakage from the loop. The valves and
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Figure 1. The NICMOS Cooling System (NCS).
bayonet couplings at the NICMOS dewar ate physically attached to a surface at nearly ambient temperature, and account for a major portion of the parasitic heat penalty at the cold end. The centrifugal circulator is driven by a permanent magnet synchronous motor at a nominal rotational speed of 1500 rev/s. Mechanically, it is a derivitive of the turboalternator that is used in the cryocooler3,4. The centrifugal impeller is 7.6 mm diameter and the shaft is 3.6 mm diameter. Radial support for the shaft is provided by two self acting tilt pad gas bearings. Axial loads are carried by spiral groove thrust bearings on opposing surfaces of the impeller shroud. The circulator operates at cryogenic temperature and provides a pressure rise of about 6 kPa at a neon flow rate of about 0.4 g/s. The internal circulating loop pressure at design conditions is 3 atm. The input power to the circulator is about 0.75 W. This heat load is part of the total parasitic load which must be carried away by the cryocooler loop. The circulator housing is a welded assembly that is integrally connected to the CLI. The CLI is a stainless steel counterflow heat exchanger designed to efficiently transfer heat from the circulating loop to the cryocooler loop with minimal pressure loss. At design point, the pressure loss in the CLI on the circulator side is about 90 Pa. The thermal effectiveness of this
component is 0.963. The CLI consists of identical arrays of slotted flow passages on opposite sides of a conductive plate. On each side of the plate, the flow passes through twenty rows of fins. In each row there are 65 fins. The flow area between adjacent fins is 0.2 mm wide by x 4 mm. The CLI is the support structure for both the circulator and the turboalternator. It is a brazed assembly that is welded to the cold end of the cryocooler recuperator. It serves as a
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structural support for the cold end mechanical assembly. Kevlar straps attached between the CLI and the frame support the cold end of the cryocooler against launch loads.
Braided stainless steel flexible tubes are welded to the tube ends attached to the CLI on the circulator loop side. Special titanium alloy bayonets are welded to the ends of these tubes for interfacing at the NICMOS cryogenic manifold. Each flex tube is about 1m long. Radiation and conductive thermal loads in the tubes and at the bayonet connections account for a major portion of the heat load on the cryocooler. NICMOS Cryocooler (NCC) The NICMOS cryocooler is a single stage reverse Brayton (SSRB) cooler that employs turbomachines for the compression and expansion functions. The machines contain low mass shafts that rotate at high speed with negligible vibration. The lack of vibration in the system is critical to meeting the requirements of the HST. The cooler is derived from an earlier Engineering Model5,6 (EM) of the cooler that was designed for 5 W of cooling at 65 K. The EM version has been undergoing life tests at the Air Force Research Laboratory at Kirtland AFB since April 1995. It has accumulated about 23,000 hours of operation in these tests. Many of the features used in the earlier EM version of the cooler have been adapted to the NICMOS cooler. A few components were modified to take advantage of technology improvements or to accommodate the particular constraints imposed by NICMOS requirements. Table 1 summarizes the important operating parameters for the cryocooler at design point. Figure 2 shows views of the cryocooler assembly and its packaged configuration. At the warm end of the cryocooler, the HRI provides a conductive coupling between the cryocooler, the electronics and the CPL. The heat of compression, the losses in the electronics, and motor losses are each rejected to the CPL loop through this single thermal interface. The electrical drive for the cooler is a Pulse Width Modulated (PWM) design supplied by Orbital Sciences Corp.7 The Power Conversion Electronics (PCE) package includes a variable frequency ac drive for the centrifugal compressor in the cryocooler, a variable frequency ac drive for the circulator and an adjustable three phase resistive load to control the speed of the turboalternator. The PCE also provides DC/DC regulation. The power electronics are contained in a compact package with a 165 mm x 190 mm footprint on the HRI. The input power to the PCE is nominally up to 450 W. Up to 67 watts of heat are dissipated in the PCE and rejected to the HRI. The compressor is a centrifugal type machine with a 15.2 mm diameter impeller driven by a three phase induction motor. The shaft is 6.35 mm dia. and is supported radially by self acting tilt pad bearings. Axial loads are carried by spiral groove thrust bearings acting on opposing faces of the impeller shroud. The maximum operating speed for the compressor is 7,500 rev/s.
The housing for the compressor has been modified from the EM version of the machine. An integral aftercooler has been incorporated into the base plate. The housing is a brazed assembly
primarily of copper including a fin-type aftercooler. It is bolted to the copper HRI. The heat of compression and motor losses are conducted through the compressor housing to the HRI. The temperature difference across this interface at nominal operating conditions is about 15°C.
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Figure 2. The NICMOS Cryocooler system. View on left shows the cryocooler components and
portions of the circulator loop attached to the assembly. View on right is the packaged cooler without side panels or multilayer insulation.
The recuperator provides for efficient heat transfer between the high and low pressure streams of gas. The recuperator consists of 300 slotted copper disks that are equally spaced axially along the length of a stainless steel outer shell. The high and low pressure gas streams are separated by stainless spacer rings between adjacent pairs of disks. The design is quite similar to that used in the EM 5 W, 65 K cooler except that the wall thickness of the outer shell has been increased somewhat to meet launch vibration requirements. The thermal effectiveness of the heat exchanger is 0.992 and the total pressure loss in both streams is about 1 % of the system pressure. The warm end of the recuperator is mounted directly to the frame/support structure. The cold end is supported through the CLI by three Kevlar straps that are attached to the frame and pretensioned to meet launch requirements with low heat leak. The frame and structural support system for the cryocooler was designed and supplied by Swales Aerospace Inc. At the cold end of the cryocooler, a turboalternator provides the necessary refrigeration. Work is extracted by expansion of the gas across the turbine rotor. The resulting shaft work is then converted to electric power by the alternator and conveyed to a resistive load in the PCE assembly where it is dissipated as heat and rejected to the HRI. The turboalternator used in the NCC replaces a turboexpander used in the prior EM 5W, 65K cooler in order to simplify the system. The turboexpander used in the EM cooler incorporated a gas brake to control speed and transport heat to the heat rejection interface. This adds a bit of complexity to the cryocooler fluid loop. Pressure and flow controls are required to maintain the appropriate turbine speed and to provide for thrust balancing. Control of the turboalternator is accomplished by adjusting the load resistance to the alternator. This reduces the number of fluid lines and the supports for them. Performance tests at the component level demonstrated that the turboalternator also gives slightly higher efficiency for the same operating conditions. The turboalternator was adapted from a machine being developed for a 1 W, 35 K load application4. Mechanically it is virtually identical to the circulator. The turbine rotor is 7.6 mm diameter. The turbine rotor and hollow shaft are integrally machined from a single titanium alloy
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blank. A rare earth permanent magnet is positioned inside the hollow shaft. Radial support for the shaft is provided by two self tilt pad bearings. Axial support is provided by self acting spiral groove thrust bearings acting on opposed shroud surfaces on the rotor. The nominal design speed for the turbine is 4,500 rev/s and the net power output is about 10 W. During cooldown of the system, the turbine power may be raised to as high as 25 W. The turboalternator assembly is brazed into the CLI. At design conditions, flow leaving the turbine is at about 63 K and rises in temperature to about 67 K at the exit from the CLI where it
reenters the cold side of the recuperator. Packaging and Physical Characteristics
The NCC is packaged in a frame designed to facilitate electrical and fluid interfacing with the NICMOS instrument in the Aft Shroud area of the Hubble Space Telescope. The frame is also designed for EVA handling. The package includes the cooler, accumulator tanks, the PCE, instrumentation, a separate gas supply for the circulator loop, flexible lines and structural supports for all elements. The assembled package is about 400 mm x 400 mm x 760 mm and weighs about 55 kg. The cooler components make up for only about 35 % of the package mass.
Individual component features are listed in Table 2. TESTS
The NCS has completed several series of performance and flight qualification tests. These include vibration testing and vibration emittance tests, mass properties, EMI/EMC tests, thermal vacuum characterization tests and performance tests. Among all tests to date, only the EMI/EMC tests have required modifications to the design in order to meet requirements. The initial conducted emissions levels from the PCE are unacceptable for the HST. They are acceptable for the flight demonstration. Modifications to the PCE to reduce conducted emissions to acceptable levels will be implemented for the HST mission. Performance tests on the system are being conducted in several stages. Initial performance tests on the cryocooler were performed in March 1998. The cooler was manually controlled and driven by an engineering model inverter during the initial tests. A small resistive heater element was used to simulate the total refrigeration load to the cryocooler and circulation loop. A cooldown and parametric steady state tests were performed to establish the cooling capacity of
the system. Heat rejection temperature was controlled by a commercial chiller. Refrigeration loads and temperature were controlled by the electric power to the resistive heater on a short section of tubing in a simplified circulator loop. The bayonets and flexible tube sections were included in the test. At the nominal design capacity corresponding to the design load and rejection temperatures, the electric input power to the cooler was 8 W (3 %) higher than target. Additional pre-flight performance tests are being conducted to characterize the system performance as subsystems are added to the cooler. These tests will include flight versions of the PCE and the control electronics, tests incorporating a cold load simulator that replicates the thermal mass and pressure loss characteristics of the NICMOS dewar, and finally, tests that include the CPL and radiator loop. Following ground tests, the system will be installed on a pallet for the HOST mission that is scheduled for STS-95 in late October, 1998. The flight test will provide final operational characterization for all subsystems to be used in the HST. PERFORMANCE
The performance of the NCS is subject to characteristics of the cooler, the uncertainties in the parasitic heat loads, characteristics of the performance of the CPL and radiator, and the radiation environment imposed on the system by the flight demonstration carrier and by the HST aft shroud. Depending on the orientation of the shuttle during the flight demonstration, the radiation environment causes fairly large changes in the relationship between heat rejection capacity of the CPL and the temperature of the HRI.
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There are similar variations in radiation environment present on the HST. However, they are less extreme than those for the HOST mission. The relationships between radiation orientation and the capacity of the cooler and the CPL result in variations in the power that is required to maintain the NICMOS instrument (at the cold well) within acceptable temperature ranges. Figure 3 shows the relationships between predicted input power to the NCS for various extremes in thermal rejection environment for both missions, HOST and HST. The CLI refrigeration capacity reflects the net amount of cooling available from the cryocooler. This capacity must account for the parasitics associated with the circulator loop, the dewar and the load on the NICMOS instrument itself. The difference between the CLI temperature and the NICMOS instrument (identified in the figure as cold well) is about 2 K. CONCLUSIONS
A flight version of a single stage reverse Brayton cryocooler has been developed. This development has demonstrated that a turboBrayton cryocooler can be successfully adapted to use for space applications where low vibration and high performance are critical. The system development to date has shown that the basic features of the turboBrayton approach will meet flight and launch requirements and modest cryogenic refrigeration loads can be accommodated at relatively large distances from the cryocooler. This effort has further demonstrated the usefulness of a cryogenic circulator in providing a means for transporting refrigeration over significant distances.
Figure 3. Predicted NCS power requirements for the flight demonstration test (left) and the HST. Hot and cold cases represent extremes in radiation environment for the radiator/CPL loop. Curves are shown for varying cryocooler cooling capacity referenced to the refrigeration produced at the CLI.
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ACKNOWLEDGEMENTS This work was performed under contract to NASA Goddard Space Flight Center. The authors would like to acknowledge the collaborative support of other contractors that participated in this effort including Swales Aerospace Inc., Orbital Sciences Inc., LMMS, and Jackson and Tull. We would also like to thank BMDO and the Air Force Phillips Laboratory for their ongoing support of this technology. REFERENCES 1. Cheng, E. S., et al. “The Near Infrared Camera and Multi-Object Spectrometer Cooling System”, Presented at the SPIE 1998 Symposium on Astronomical Telescopes and Instrumentation, Kona, HI,
March 20-28, 1998 2. McIntosh, R., et al. “A Capillary Pump Loop Cooling System for the NICMOS Instrument”, Presented at the 28th International Conference on Environmental Systems, Danvers, MA, July 1316,1998.
3. McCormick, J. A., et al. “Progress on the Development of Miniature Turbomachines for Low Capacity Reverse Brayton Cryocoolers in the 9th International Cryocooler Conference”, Cryocoolers 9, Plenum Press, New York (1997), pp. 475-483. 4.
McCormick, J. A., et al. “Design and Test of Low Capacity Reverse Brayton Cryocooler for
Refrigeration at 35 K and 60 K”, Presented at the 10th International Cryocooler Conference, Monterey, CA, May 26-28, 1998.
5. Swift, W. L., “Single Stage Reverse Brayton Cryocooler: Performance of the Engineering Model in
the 8th International Cryocooler Conference”, Cryocoolers 8, Plenum Press, New York (1995), pp. 499-506.
6. Dolan, F. X. et al, “A Single Stage Reverse Brayton Cryocooler: Performance and Endurance Tests on the Engineering Model in the 9th International Cryocooler Conference”, Cryocoolers 9, Plenum Press, New York (1997), pp. 465-474. 7.
Konkel, C., et al. “Design and Qualification of Flight Electronics for the HST NICMOS Reverse Brayton Cryocooler”, Presented at the 10th International Cryocooler Conference, Monterey, CA, May 26-28, 1998.
Design and Qualification of Flight Electronics for the HST NICMOS Reverse Brayton Cryocooler C. Konkel and W. Bradley
Orbital Sciences Corporation Greenbelt, MD 20770 R. Smith
NASA/Goddard Space Flight Center Greenbelt, MD 20771
ABSTRACT
The NASA Hubble Space Telescope Project is developing an active cryogenic cooling system to replace the expendable cryogen used in the NICMOS instrument. A key element of the system is
the Power Conversion Electronics which converts spacecraft DC to 3-phase AC to drive a reverseBrayton cycle refrigerator and a circulator. Additional functions include DC power distribution, health and status telemetering, and safety monitoring. A flight unit meeting these requirements while remaining within a 6.5x7.5-inch footprint and 67W power dissipation constraint, was successfully designed, built and tested within 10-months of go-ahead. State-of-the-art electronics techniques, including pulse-width modulated inverters, laminated metal core printed wiring boards, and Faraday cage containment methods were utilized. Electrical efficiencies and performance parameters are described.
BACKGROUND The NICMOS Cryocooler (NCC) is a 70 Kelvin refrigerator based on the reverse-Brayton cycle (Nellis, et al1). It is being built to extend the lifetime of the cooling system presently in use on the Hubble Space Telescope (HST) NICMOS instrument. To provide this cooling, the NCC utilizes
three turbomachines including a compressor driven by a 3-phase induction motor, a circulator driven by a 3-phase synchronous motor, and a turboalternator. The PCE drive electronics generates 3phase AC to drive the two motors, using the spacecraft 28VDC bus input voltage. To provide maximum control flexibility, the DC input voltage and the fundamental drive frequency for both motors must be externally adjustable. This dictates the requirement for a variable voltage DC/DC converter and variable frequency DC/AC inverter. In addition to the PCE, a separate Turbo Resistor Assembly (TRA) was designed to house a series of load resistors to dissipate the electrical energy generated by the Turboalternator. Cryocoolers 10, edited by R. G. Ross, Jr. Kluwer Academic/Plenum Publishers, 1999
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In order to demonstrate this technology prior to deployment on HST, NASA has scheduled a test flight on the Hubble Orbital System Test (HOST) mission in late 1998. The Power Conversion Electronics was designed and built within 10 months of go-ahead in order to meet this schedule. The PCE is presently installed in the NICMOS Cryocooler and is undergoing system level testing at NASA/GSFC. PHYSICAL DESCRIPTION
The Power Conversion Electronics and Turbo Resistor Assembly together form the electronic system required to operate the three turbomachines used in the NICMOS Cryocooler. The electronic units are located within the NCC enclosure and are mounted to a cold plate which is a part of the Capillary Pumped Loop (CPL) system for temperature control. Figure l(a) shows the PCE and TRA as configured for box-level thermal vacuum testing. Figure l(b) shows the PCE mounted within the NCC enclosure, after Multi Layer Insulation (MLI) has been installed. The TRA is located behind the PCE and is not visible in this photograph. Housing components were fabricated from 6061 aluminum alloy and bolted together to provide maximum flexibility if any late modifications were required. The 6.5 X 7.5 - inch footprint constraint required a unique shape as seen in the figures. Access to connectors as well as avoidance of NCC internal components dictated the cutouts and overhangs shown. The separate TRA is
located behind the PCE. Wall thickness was driven by thermal considerations, resulting in a very robust structure that met all NASA environmental test specifications. Qualification results are discussed later. Internally, the PCE consists of three printed wiring modules. • Module A1 is the Compressor DC/DC Converter and is located at the bottom of the box as close as possible to the cold plate for maximum heat dissipation
Figure 1. Power Conversion Electronics and Turbo Resistor Assembly shown configured for thermal
vacuum testing (a) and as installed in the NCC (b).
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• Module A2 functions as the Circulator Converter/Inverter and the Compressor Inverter and consists of two multilayer printed wiring boards laminated to either side of an
aluminum heatsink • Module A3 is the Signal Conditioner consisting of one multilayer printed wiring board
and heatsink located on the “quiet” side of a Faraday cage Module A1 is bolted directly to the baseplate while A2 and A3 use card retainers and can be slid out for maintenance. All three modules consist of one or more Kevlar printed wiring boards bonded to an aluminum heatsink using thermabond (Arlon) sheet adhesive. The chassis volume above the modules houses six magnetic latching relays, two EMI filters, a Hall effect current sensor and harnessing. Outline dimensions are 6.5 x 7.5 x 6.8 inches. Component density is very high and particular care was taken to ensure proper clearances were maintained during assembly. Figure 2 shows the module arrangement and basic box outline. REQUIREMENTS
Figure 3 is a block diagram of the required PCE functions, and reflects the configuration as delivered to NASA/GSFC. A summary of requirements imposed on the PCE is shown in Table 1.
Figure 2. PCE module arrangement and box outline.
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Figure 3. Power Conversion Electronics block diagram.
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Based on these requirements, a design evolved that focused on electrical efficiency and volume. As the layout of the NCC enclosure matured, volume became the primary PCE design driver, ultimately pushing the converter and inverter design toward power densities beyond the “low risk” threshold rule-of-thumb of 6 watts per cubic inch. It was at this point in development that alternative approaches to converter and inverter design were identified and compared, with emphasis on volume reduction. The remaining discussion will focus on these modules.
DESIGN METHODOLOGY As a starting point, proven spacecraft switched-mode power supply design techniques and components that inherently address the critical design drivers were used for initial parts count and circuit analysis. These included the use of MOSFETs with low resistance and minimal gate capacitance, stacked ceramic and metal polypropylene capacitors with low equivalent series resistance (ESR), and multifunction microcircuits to minimize parts count and board space. Inverter Topology Various approaches to the generation of 3-phase AC using a solid state inverter are available. These approaches vary based on the motor characteristics, source voltage available and operating
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environment. For the low voltage-high current compressor motor used in the NICMOS cryocooler, and the DC voltage available from HST, a DC/AC inverter is required. The first design decision was to select the inverter topology, and the voltage waveform applied to the induction motor. Inverters
can be classified in terms of these waveforms, which include an amplitude adjustable sine wave, amplitude adjustable square wave, or amplitude adjustable, sinusoidal-driven pulse-width modulated (PWM) square wave. An example time-domain profile of these waveforms is shown in Figure 4.
Figure 4. Example waveforms for various inverter outputs.
An implementation for each of these approaches was evaluated in terms of the critical design parameters. A brief discussion of each technique follows. Sinewave Drive. An innovative design of a transformer-coupled inverter using six toroidal transformers and twelve MOSFET switches connected in a ring topology was built and tested by Creare Inc. (McCormick, et al2). In this approach, a sinusoidal AC voltage is generated by transformer secondary coils from square waves applied to transformer primary coils. The coils are arranged such that a 3:1 phase delay is maintained between opposite pairs of transformers. The square waves are clocked by an n-stage Johnson counter where resulting in a rotating pattern of primary coil excitation that sequentially excites the three secondary phases. Critical to the design is the correct selection of transformer turns ratios, thereby assuring the best approximation to a sine wave using a stepped waveform input. This optimum turns ratio N can be calculated as a function of output voltage and input DC voltage
Using this approach, the lowest harmonic of the resultant sine wave is (2n - 1) with an amplitude equal to l/(2n -1) times the fundamental. The only significant harmonics present are the eleventh (1/11), thirteenth (1/13), twenty-third (1/23) and twenty-fifth (1/25). Carrier Modulated PWM Drive. A pulse-width-modulated version of this inverter can also be conceptualized. In this case, a resistor network as shown in Figure 5 will be used to generate a
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stepped sine wave instead of transformers. The PWM output will also be direct coupled to the motor, requiring that the neutral be floating. A 6-stage Johnson counter can be used to clock the stepped sine waveform, but instead of the MOSFET switches operating at the fundamental frequency (in this case 8 kHz), they will operate at the PWM carrier frequency
Figure 5. Resistor network used to generate stepped sine wave for PWM carrier modulation.
Three resistor networks consisting of six resistors each will be required to generate the 0°, 120° and 240° portions of the 3-phase voltage. In a manner similar to the transformer synthesized sine wave, the ratio between resistors can be calculated to approximate a sine wave:
The true (uncomplemented) outputs from the Johnson counter are tied directly to each resistor to generate the 0° fundamental waveform, while a combination of true and complement outputs are tied to the remaining two networks to generate the 120° and 240° waveforms. Each network is followed by a combination summing amplifier and low pass filter. This is then used to modulate the pulse width of the PWM carrier. The output power MOSFETs, arranged as a half bridge, are driven by this square wave. All harmonics below the eleventh are canceled as before. Square Wave. The third option for 3-phase drive is square wave drive (Konkel, et al3). This approach is the simplest of the three, consisting of an 8 kHz square wave driving six MOSFET switches in a half bride configuration. As in the PWM case, the motor neutral must be floating for it to operate correctly. No specific harmonic neutralization technique is included. However, electrical isolation of the neutral inherently removes the third harmonic and all its multiples from the motor
voltage waveforms, significantly decreasing motor loss.
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Inverter Testing
All three of these applied voltage approaches have been tested using very similar motors and load conditions (7500 rps, 325W). The approach with the most operational time logged to date uses the sinewave scheme and transformer-coupling. A breadboard using the direct-coupled PWM technique was built and tested, including a phase shift network to correct for the 0.5 power factor. A square wave approach was also tested. Results are summarized in Table 2.
Based on these results, a PWM sinewave drive system was implemented for the NICMOS cryocooler. As stated earlier, the volume constraint severely limited the available PCE volume and six additional transformers and six MOSFET switches for the transformer sine wave inverter exceeded the volume available. The square wave inverter waveform exhibited additional harmonics and had as a result a lower efficiency, although its performance was impressive for such a simple drive technique. In addition, the option to move bulky components, such as the inductors and capacitors for the phase shift network, outside of the box and into an external harness was an option
available with the PWM approach. A block diagram of the PWM inverter is shown in Figure 6. A similar inverter was implemented for the circulator motor. Even though it is a synchronous motor with lower current and voltage requirements, the basic 3-phase drive technique used for the compressor is compatible.
Figure 6. PCE Compressor PWM inverter block diagram.
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Converter Topology
Compressor Converter. The Compressor DC/DC Converter requirements shown previously in Table 1 drove the design towards a standard forward converter topology with current feedback. It provides isolated and controlled DC voltage to the Compressor Inverter at power levels of more than 500W at up to 45VDC. Input to output isolation was met without the use of optocouplers using an op amp source impedance to meet the 1 megohm requirement. The primary side topology is a full wave bridge with transformer coupling to provide voltage step up at a switching frequency of 35kHz. The primary MOSFET switches are driven by IR2110 drivers which are in turn controlled by a current mode controller. A current transformer samples the primary current which is compared to the feedback error signal. The primary-referenced drive and control circuitry is powered by ±12Vdc derived from the input 28V bus. Secondary referenced drive and control circuitry is powered by a secondary ±15V from a separate off-the-shelf hybrid. The secondary side uses a synchronous rectification topology. The MOSFET switches are again driven by IR2110s which have their inputs isolated from and driven by the same controller that drives the primary MOSFETs. The rectified output voltage is filtered, switching spikes are both damped across the transformer secondary and clamped with a power zener, and the voltage is sampled and compared with the voltage command input. The feedback includes an integrator which gives very precise output control and a slow loop bandwidth. The Compressor Converter design goal of 94% power conversion efficiency was difficult to meet given the parts constraints such as hermiticity, radiation hardness, reliability screening, and volume limitations. A combination of RC damping, active zener clamping, and higher MOSFET ratings was used to achieve an efficiency of 89.5% which is constant over the input voltage range.
Circulator Converter. The Circulator Converter was originally a forward converter which provided primary to secondary isolation and voltage control of the power going to the Circulator Inverter. However, since the Circulator only required about 1W of power at 8 Volts, the supply topology was changed to a linear supply with peak current limiting of 0.75 Amp. The ±15V input power to the linear regulator is provided by the same ±15V hybrid which provides low voltage power to all of the PCE circuitry. Using this secondary referenced input source simplified control and telemetry circuits which would otherwise have to span the 1 megohm isolation barrier. The current limiting feature of the supply allows the circulator motor to start and to accelerate smoothly to full 1500 rps synchronous speed. The speed command is controlled by a 10 second RC time constant circuit which starts at zero and increases the motor speed at a rate which is compatible with the motor dynamics and input power. This circuit was combined with the Circulator Inverter and placed on one printed wiring board (bottom side of A2). An overall (Converter plus Inverter) electrical efficiency of 66% was achieved, which at the low power level amounted to a minimal 0.6 watt loss. QUALIFICATION TESTING
The PCE was tested per the NASA/GSFC General Environmental Verification Specification for Payloads (GEVS-SE). This included thermal vacuum, vibration, and EMI. All tests were passed with the exception of EMI conducted emissions, which was deemed acceptable for the HOST mission. Thermal Vacuum Test. The PCE and TRA were subjected to four hot/cold cycles at vacuum. Functional and performance tests were run at temperature extremes of -50°c to +70°C. The cold plate mounting interface was maintained at 45°C. Board level temperatures were within 3° of prediction. The hottest component case temperature measured was 69°C on the Compressor
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Inverter (top of A2). All component temperatures remained within GSFC Preferred Parts List (PPL-21) derated guidelines. Vibration Test. The PCE and TRA were subjected to sine and random vibration environments per GEVS-SE. An abbreviated functional test was run following each axis, and a full functional test was run following completion of all axes. All tests were passed. The PCE and TRA first natural frequency was well above the 50 Hz requirement. Electromagnetic Interference/Compatibility. The PCE and TRA were subjected to EMI/EMC testing while installed within the NICMOS cryocooler enclosure. All tests were passed except conducted emissions CE03. The levels measured were determined to be acceptable for the HOST mission. A modification to the power line filter within the NCC is being evaluated. FUTURE RECOMMENDATIONS
The success to date of the Power Conversion Electronics has resulted in increased confidence in the design approach taken. Areas in which additional investigation is warranted in order to enhance existing or future versions of this electronics unit include the following: • Further reduction in the size of the converter • Completion of a reliability and Mean Time Between Failure (MTBF) calculation • Incorporation of programmable functions such as start up profiles and fault monitoring • Further characterization of the square wave inverter and its mission applicability IN MEMORIAM
The authors dedicate this paper to the memory of Mr. Leland Van Allen, a dedicated engineer and polite coworker who will be sorely missed. REFERENCES 1.
Nellis, G., Dolan, F., Swift, W., Sixsmith, H., McCormick, J., Castles, S. and Gibbon, J., “Reverse Brayton Cooler for NICMOS”, Cryocoolers 10 Conference (1998).
2.
McCormick, J.A., and Valenzuela, J. A, Three Phase Inverter for Small High Speed Motors, Report TM-1499, Creare Inc., Hanover, NH (1991).
3.
Konkel, C.R., McCormick, J.A., Conceptual Design of the Flight Electronics to Operate a Single Stage Reverse Brayton Cryocooler, Task 005, Fairchild Space, Greenbelt, MD and Creare Inc., Hanover, NH (1994).
Flight Demonstration of the Ball Joule-Thomson Cryocooler R. Fernandez and R. Levenduski Ball Aerospace & Technologies Corp. Boulder, Colorado, USA 80306
ABSTRACT The Ball Joule-Thomson (J-T) cryocooler has been under development for over 10 years and has achieved a high level of technical maturity. The Cryogenic On-Orbit Long-Life Active Refrigerator (COOLLAR) program developed two protoflight versions of the J-T cryocooler. One version is the Engineering Development Model (EDM). This cryocooler consists of an oillubricated compressor with a gas purification system that provides ultra-pure nitrogen to the cold head. The cold head produces liquid nitrogen to provide two stages of constant temperature cooling for varying loads. The cold stage provides 3.5 W of cooling at 65 K, and the intermediate stage provides 5 W of cooling at 120 K. The COOLLAR Flight Experiment (CFE) program flight tested the EDM J-T cryocooler to verify microgravity performance. The cryocooler was flown on STS-85 aboard the space shuttle Discovery on the Technology Applications and Science 1 (TAS-1) experiment bridge. This paper presents the results of the flight experiment. All of the experiment objectives were accomplished, and all operating principles were demonstrated in this mission. The data shows that there are no functional performance differences between ground and flight operation. On-orbit performance robustness and adaptability useful to future applications have been demonstrated. BACKGROUND Many technical issues have been addressed during the development of the Ball J-T cryocooler. These have included contamination control, long life, and cryogen liquid management for precise temperature stability with varying loads. The Ball J-T cryocooler was flown aboard the space shuttle Discovery during a 12-day mission that began on August 7 and ended on August 19 of 1997. The goal of the mission was to demonstrate performance of all cryocooler systems in the microgravity environment. All of the experiment objectives were accomplished. The cryocooler was de-integrated from the bridge at NASA’s Goddard Space Flight Center and returned to Ball for post-flight tests. Post-flight testing has been completed and work is underway to put the cryocooler on extended life testing. Post-flight performance matched flight performance. Figure 1 shows the J-T cryocooler that was flown. EDM cryocooler details can be found in Reference 1.
Cryocoolers 10, edited by R. G. Ross, Jr. Kluwer Academic/Plenum Publishers, 1999
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Figure 1. EDM J-T cryocooler on the test bench prior to CFE integration.
CFE DESIGN OVERVIEW The CFE consists of the EDM cryocooler and additional hardware to allow for space flight and data acquisition. The major cryocooler components are shown in Figure 2. The compressor assembly is located near the center of the experiment test bed (midplate). Heat pipes attached to the compressor cooling ring and good thermal attachment to the midplate were used to dissipate the compressor heat. The cold head is attached near the radiator side of the midplate. The cold head was contained in a small vacuum chamber. Vacuum was acquired using a small vacuum getter. The electronics subsystem is flightworthy but is not the final flight design, since many of its components support test telemetry that would not be used in a typical flight cryocooler. The CFE assembly is shown in Figure 3. To fly the cryocooler, several additional components were required. Shuttle man-rated requirements drove us to the use of a dome enclosure. The cryocooler was heavily instrumented to obtain engineering data. This required an extensive set of electronics. The electronics boxes were the Power Conditioning Unit (PCU), Data Acquisition and Control Electronics (DACE), and Temperature Module Assembly (TMA). Thermal management was a critical issue since temperatures at the heat dissipation interfaces had to be controlled for experimental repeatability. We used a heat pipe radiator as the main heat dissipation device. In addition, heat pipes were used to remove heat from the compressor, and the midplate was designed to be a good conductor. OBJECTIVES OF THE FLIGHT EXPERIMENT
The objectives of the flight experiment were primarily aimed at demonstrating spaceflight and microgravity performance. The objectives were:
1. Characterize the J-T cryocooler’s operation in microgravity, focusing on the liquid nitrogen distribution system and the oil control system 2. Correlate flight performance data with ground performance data
3. Demonstrate the capability to maintain desired interface temperature stability with a variable heat load 4. Identify flight-related operational issues or constraints
All of the mission objectives were achieved, and all operating principles were demonstrated. There were no flight-related operational constraints. Comparison of flight data to ground test data shows good performance correlation.
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Figure 2. EDM J-T cryocooler located on the CFE midplate.
TEST RESULTS
The flight experiment objectives were met by running the J-T cryocooler through all of its operating modes for the full 12-day mission. Tests included two cooldowns, three start/stop cycles at different heat-rejection temperatures, and multiple load profile tests. Additional tests included two start/stop cycles with cold cryogenic interfaces. Also, tests were done with changes to the cold- and intermediate-stage heat loads and operating temperatures. The test results showed:
1. Temperature stability of better than ±0.25 K at 65 K with a 1.25 to 4.0 W load change Liquid nitrogen acquisition and storage at 65 K Load profile capability (duration at high load) Liquid nitrogen acquisition and storage at 120 K Temperature stability of better than ±2 K at 120 K Oil acquisition within the compressor Oil/gas separation in the compressor Additional cryocooler capabilities were also demonstrated: 1. Shutdown and start-up with cryogenic interfaces at operating temperatures 2. Steady-state performance during all shuttle maneuvers 3. Nominal operation over a heat-sink temperature range of 285 K to 315 K 4. Survival of launch environment All of the operating principles of the J-T cryocooler were proven.
2. 3. 4. 5. 6. 7.
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Figure 3. CFE assembly showing major components.
Liquid Nitrogen Acquisition and Storage
One of the primary mission objectives was to verify that liquid nitrogen could be acquired and stored. The J-T cryocooler produces a steady flow of liquid nitrogen that is captured and
stored for later use. The intermediate stage (T2) tank stores 120K liquid, which is near the critical temperature of 126.1 K. At this temperature the properties of the gas and liquid become less distinct, making it difficult to separate one from the other. The cold stage (T1) tank stores liquid at 65 K, which is near the triple point of nitrogen. The data shows that the J-T cryocooler was able to acquire and store liquid at both 65 K and 120 K. Figure 4 shows data from cooldown and T1 tank liquid accumulation during flight test on top and ground test on bottom. The T1 tank temperature can be seen leveling out in both cases at 65 K, indicating that saturated liquid at constant pressure (2.5 psia nominally) is stored in the tank. The slower fill rate during the flight is caused by having a higher heat leak during this time. Figure 5 shows data from cooldown and T2 tank liquid accumulation during flight test. The T2 tank temperature (T10) can be seen leveling out at 120 K, indicating that saturated liquid at constant pressure (350 psia nominally) is stored in the tank. The same performance was seen during the ground tests. We noticed that the T2 tank stored more liquid in micro-g. In one g, the storage capacity is limited because the wick material only has a wicking capability of about 0.5 in., but in micro-g this does not apply because storage is not wick “height” dependent since gravity is not fighting the capillary wicking capacity. This operational difference at the component level did not affect the functional performance. These data combined with previous ground testing show that liquid acquisition and storage over the full temperature range of nitrogen has been proven in all orientations in one g and in micro-g. The fundamentals of liquid capture and storage are well understood and the J-T technology can provide liquid nitrogen cooling over the full temperature range. It also confirms that multi-orientation testing in one g strongly indicates successful zero-g performance will be achieved.
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Figure 4. Graphs show data from cooldown and T1 tank liquid accumulation during flight test
on top and ground test on bottom. Data shows similar performance during flight and ground test.
Figure 5. Data from cooldown and T2 tank liquid accumulation during flight test. Ground test performance was the same.
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Figure 6. T1 interface temperature stability during load profile tests. Liquid is depleted at the end of each high-load period and re-accumulated during the low-load period.
Temperature Stability at Cryogenic Interfaces
Another primary mission objective was to demonstrate stable cryogenic interface temperatures. The T1 (65 K) interface temperature stability was better than the required ±0.25 K specification with a changing load. This was demonstrated with a heat load variation of 2.75 W. The low load was approximately 1.25 W, and the high load was 4.0 W. Figure 6 shows data obtained from a section of a load profile test. This graph shows the temperature stability of better than ±0.20 K as the load changes from low load to high load during flight. The increase at the end of the high-load period occurs because liquid is no longer available to absorb the load. The load cycles back down to low load when the T1 interface temperature exceeds 67 K. During steady state, low load conditions, the temperature stability was better than ±5 mK. The flight and ground data exceeded the ±0.25 K stability requirement. These results prove the heat transfer fundamentals at the interface are well understood and successfully implemented in the design. These results also improve confidence that the technology can be successfully applied to cool distributed interfaces. The temperature stability requirement of ±2 K was demonstrated at the T2 (120 K) interface. The performance of this interface was better than ±0.5 K while the T1 interface was going through the load profile. These tests prove the J-T cryocooler can simultaneously maintain two cryogenic interfaces at the proper temperatures. Load Profile Capability Demonstrating the load profile capability was another major accomplishment of the flight experiment. Testing proved the J-T cryocooler could store liquid during low load periods and use this liquid to provide constant temperature and peak cooling during the high load periods. Five load profile tests were done. Each load profile test consisted of eight high-load period, with lowload periods in between. Data from one of these tests is shown in Figure 6. These tests demonstrate consistent load profile performance under nominal operation, much like the ground tests. Oil Management Proper oil management was another of the primary mission objectives. Two critical aspects of the oil management system were verified in space. The first aspect verified was the micro-g collection and distribution of oil within the compressor. The oil collection system collects oil and
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feeds it to the oil pump for distribution. The pressurized oil is used to lubricate moving parts, aid in heat transfer, and provide piston sealing. This operation can be verified by looking at the oil
pressure levels, particularly during start-up. This is shown in Figure 7. Oil is collected at the pump, and pump pressure builds immediately upon start-up. The high-pressure oil feed builds up to about 875 psia, and the low-pressure oil feed builds to about 175 psia. This performance was very repeatable and can be seen in the data from the flight start-up tests. These start-up tests were done multiple times with a range of initial cryocooler heat sink temperatures from about 270 K to 320 K. Achieving the correct gas pressure indicates proper piston sealing. This is also shown in Figure 7. The only difference between the two graphs is in the gas pressure levels. The system high pressure was higher, nearly 1650 psia, in the ground test because there was a higher charge mass
during this test than during flight. The figure shows similar flight and ground performance indicating that there were no zero-g effects on the collection and distribution of oil within the compressor. The second oil management aspect verified was zero-g separation of the oil from the working gas by the oil scavenger. Because the working gas entrains oil as it moves through the compressor, it must be removed to prevent contaminating the cold head. Two indicators were available to assess the oil scavenger’s performance: a differential pressure transducer and performance degradation from oil migrating to the cold head. The pressure transducer indicated nominal operation throughout the mission and no cold head contamination was detected. Testing has now proven the scavenger works in all orientations in one-g and in zero-g. This breakthrough technology now makes commonly used oil-lubricated machinery adaptable to space
use.
Figure 7. Flight vs. ground test comparison of oil and gas pressures during start-up.
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Figure 8. Video images showing shuttle payload bay and CFE during STS-85 mission. SUMMARY AND CONCLUSIONS
All the mission objectives were successfully accomplished, and, in doing so, critical questions regarding zero-g operation of the J-T cryocooler have been answered. There are no
significant performance differences between ground and flight operation. The on-orbit test data shows that there are no zero-g issues. Oil and liquid nitrogen management has been proven, as have critical performance parameters such as temperature stability with varying heat load. In addition, on-orbit performance robustness and adaptability useful to future applications have been demonstrated. The only performance change attributable to zero-g was improved liquid acquisition and storage capability of the T2 tank. All other differences between flight and ground data were reconciled by post-flight testing and proven not to be related to zero-g. The cryocooler will now be put on extended life testing with the objective of gaining information about potential life-limiting factors of the cryocooler, including mechanical and
electrical components and contamination. Because the cryocooler has gone through extreme conditions that typical flight hardware would not experience, this is not a true life test. Nevertheless, there is substantial benefit to be gained by long-term operation of the cryocooler in a cost-effective manner such as this. Important information will be gained about the mechanical integrity of the design, the implementation of the electronics design, and the effectiveness of the gas purifier in controlling contamination. Information gathered from this test will be used to advance the design and increase the reliability of future builds. ACKNOWLEDGMENTS
The authors wish to thank the dedicated members of the CFE team for their outstanding support of this work. This work was sponsored by the Air Force SMC/TE-STP. We would also like to acknowledge the support from AFRL, the Aerospace Corporation, and the Goddard Space Flight Center. REFERENCES 1.
Levenduski, R., and R. Scarlotti, “Joule-Thomson Cryocooler Development at Ball Aerospace,” Cryocoolers 9, Plenum Press, New York (1997), pp. 493-508.
Design Optimization of the Throttle-Cycle Cooler with Mixed Refrigerant M. Boiarski, A. Khatri, and V.Kovalenko*
IGC-APD Cryogenics, Inc./ Allentown, PA 18103 USA *Moscow Power Engineering Institute/ Moscow, Russia ABSTRACT
Methodology and computer software were developed for an optimal refrigeration system design. The analysis takes into account mixed refrigerant properties, features of a given compressor and a counter-flow heat exchanger. This model combined with limited experimental data allows prediction of the refrigeration performance of a cooler with good accuracy. It is also possible to further optimize the system. To compare different configurations of counter-flow heat exchangers, heat exchanger effectiveness was used as a parameter. It was calculated as a ratio of a given cooler performance to the performance of an idealized cycle which takes into account only properties of the mixed refrigerant. This methodology was used in developing mixed refrigerants based coolers using an oillubricated, single-stage compressor. These coolers meet the requirements in electronic cooling, water-trapping technology and many others applications. INTRODUCTION
During research starting in 1965-1970 it was found that a throttle refrigeration system may be highly efficient by running mixed refrigerants ( MR). Retrospective and evolution of the mixed
refrigerant technology is given elsewhere[1]. The simplicity of a system with a single-stage oillubricated compressor and flexibility of the mixed refrigerant technology provide opportunities to build relatively low cost and highly reliable cryocoolers for different applications. The flow schematic of the cooler and the throttle-cycle parameters in the temperature-enthalpy diagram are presented in Fig.l. Small-scale coolers of different configurations are available in the market[2, 3]. This technology is also efficient in a semi-cryogenic temperature range of 120 K to 220 K to build a low cost and reliable water traps combined with turbo-molecular pumps[4, 5]. Earlier investigations of the MRs technology were mostly devoted to MR thermodynamic properties and the throttle-cycle analysis and optimization[1]. Cycle analysis is a very important step in developing a cooler for a certain application. It may be conducted based only on the information obtained from thermodynamic models combined with limited experimental data on the MR properties. This step also involves some assumptions on the hardware characteristics of the cooler. For a given ambient temperature refrigeration temperature and the MR thermodynamic properties (Fig. 1 ) a refrigeration capacity and power consumption of the compressor may be estimated by assuming values of temperature difference and pressure drops in the supply and return lines. Cryocoolers 10, edited by R. G. Ross, Jr. Kluwer Academic/Plenum Publishers, 1999
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Figure 1. Flow schematic of the cooler and temperature-enthalpy diagram.
In addition, the compressor volumetric efficiency (CVE) and power efficiency ( CPE ) should be given. The results of the cooler performance mostly depend on the accuracy of these assumptions. That is why some additional experimental work has to be done to optimize a MR cooler design and performance. System analysis and simulation of an optimal cooler design should take into account characteristics of either a real or an assumed hardware. Such an analysis is well developed in refrigeration technology for a vapor-compression cycle with a pure refrigerant[6]. In this case a simple system schematic, which does not include a counter-flow heat exchanger, makes it easy to correlate and generalize the experimental data. For the MR cooler, developing a system analysis is still new[7]. Such an analysis is much more complicated for two reasons. First, the transport properties of the MRs should be known within wide temperature range from to A MR may exist in the system at a different phase state- Fig. 1: homogeneous vapor ( V ) - at relatively high temperatures, vapor-liquid ( V-L ) or vapor-liquid-liquid ( V-L1L2 ) - in the low-temperature region. Appropriate models for predicting viscosity and heat conductivity of MR are available only for a rather narrow temperature range. Second, to define a thermal efficiency of the counter-flow heat exchanger, it is necessary to operate with models describing local heat transfer coefficients. Although the equations are available for the enforced convection, appropriate models for condensation and evaporation are not generic[8]. They need some limited experimental data to be adjusted for a certain configuration of the apparatus. Experimental data on the local heat transfer and pressure drop coefficients are difficult to obtain from testing small-size coolers. Therefore it is important to develop a methodology which would provide an optimal design based on the information obtained from testing cooler performance. Computer software must be available to simulate variations of the parameters at specific points of the cycle (Fig. l ) at different conditions. A preliminary comparison of calculated and experimental data shows [ 2 ] that the developed software allows an acceptable accuracy based on limited data obtained from testing of MR coolers. This software was used for further development of the optimal system design. Discussion of the methodology and results is the goal of this paper.
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MODELING PRINCIPLE FOR A SYSTEM PERFORMANCE
Optimal design of the MR cooler is based on a general numerical optimization concept[9],
where an objective function to be optimized must be identified. Independent variables and constraints should be selected according to a current task to be solved. Then it is possible to use one of the iterative optimization procedure if the initial variables were defined. Let us assume that for given temperatures and we design the cooler based on a singlestage compressor which has a displaced volume a nominal motor-driver power and a maximal discharge pressure The configuration of the heat exchanger (HE ) should be given as well. In this case two different objective functions may be defined: (a)
or (b)
where
is a heat load of the object to be cooled. Sometimes other objective functions - such as cooldown time, size of the cold, etc. - may be selected. For the optimal design of the cooler we selected as the objective function. To reduce the number of active independent variables let us assume that the MR composition is given as Based on the thermodynamic properties obtained from the equation of state one can make an estimate of the pressure values at point 1 and 4 ( Fig. 1 ) : For this set of variables the following system of non-linear equations should be solved to find
where is the compressor flow rate, different ways to solve these equations.
:
- a heat-transfer coefficient in the HE. There are
Cycle analysis may be conducted by estimating the values of and based on the design experience. In addition, the CVE value should be specified. Since the MR specific heat capacity depends on temperature, it is necessary to use a computer software to check that the minimal temperature difference in the counter-flow HE is positive Quality of different refrigerant may be evaluated with the cycle analysis. For practical applications it is important to know how much refrigeration capacity may be obtained with a given compressor. A specific refrigeration capacity which may be achieved with a single-stage compressor having at different are presented in Fig. 2. Data are given for in the range of 80 K to 240 K. A composition of the MR has been selected to provide Values of across the HE was taken as following:
It was also assumed for these calculations that Data for the pure refrigerants are presented for comparison. All the selected MR can provide refrigeration with pressure patio At the pure refrigerants can work if That is difficult to achieve with a single-stage compressor. This comparison displays an efficacy of the at
MR technology. To realize these opportunities to build highly efficient coolers optimization of
the design should be conducted based on the system analysis.
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Figure 2. Refrigeration capacity for a single-stage compressor with
System analysis needs the specifications of the effectiveness for both the compressor and the HE to predict for the cooler. A linear function may be used: to describe the CVE, where A and B are empirical coefficients. The heat exchanger effectiveness depends on both and values. Appropriate
models should be developed to solve the equations ( 1 ) - (7 ) for the cooler with the HE of a given configuration. Analysis of these models is given below. It is evident that and ( for the return line ) influence value in opposite directions. To increase the value of value must be higher (in order to reduce ), but for a given the must be lower to increase and . Calculated data presented in Fig. 3 for the same MR and HE show that depends more on at lower values of and it is almost independent of at higher values of During the iterative optimization procedure different configurations of the HE have to be compared to select the best one. We propose a single criteria in order to identify the heat exchanger effectiveness (HEE) as a ratio of (a real cycle) and (an idealized cycle):
where - refrigeration capacity of a real cycle with a given heat exchanger , - the refrigeration capacity of an idealized cycle. Values of this dimensionless function are in a range from 0 to 1, and allows an easy way to compare effectiveness of different HEs. This function also helps us to find conditions for the effective application of a heat exchanger. For example, Fig.4 presents the calculated data for a tube-in-tube heat exchanger having 7 tubes for the return MR flow. HEE values were obtained as a function of the HE length for two different refrigerants. One of them is flammable comprising nitrogen and hydrocarbons. Another one -
is a nonflammable blend based on nitrogen and perfluoro-hydrocarbons. The refrigerants have different molecular mass and Idealized cycle calculations (which are independent of the
predict for a compressor with
at
for the
and for the
at
for the
and for the
the following performance:
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Figure 3. Refrigeration capacity dependence on pressure drop in the return line.
Figure 4. Heat exchanger efficiency for different MR.
Thus using the computer software and the proposed HEE, we can select such a configuration of the tube-in tube HE which provides However this software needs models to calculate both heat transfer coefficients and DP.
COUNTER-FLOW HEAT EXCHANGER CHARACTERISTICS The software for the MR cooler optimization uses overall heat transfer coefficients calculated based on models available from the literature [8] for local heat transfer coefficients. These models
allow to generalize the experimental data for a certain type of the HE. Although different approaches have been developed to calculate DP and local heat transfer coefficients for evaporation under the enforced convection, we selected a simple equations based on the assumption of the homogeneous hydrodynamic structure of two-phase flows. It is known that this method may be used successfully if a proper correlation for viscosity of two-phase flow has
been found. We developed such a correlation by testing the MR coolers with different configurations of the tube-in-tube heat exchanger. The test set up allows to measure temperature and pressure
values at all the points of the cycle presented in Fig.l. The refrigerant mass-flow rate was measured with the pressure-differential flow meter. These data combined with the models for a local heat transfer coefficient and the original software allow us to define an overall heat transfer coefficient as a function of temperature Proper instrumentation makes it possible obtaining the data with estimated uncertainty of 10 %. Based on this information an average - overall heat transfer coefficient may be obtained, which is an integral characteristics of the given heat exchanger. There is no information
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in the literature which shows how to correlate such data. We found that these data may be plotted against a function which takes into account not only mass flow rate but the MR properties at the suction point as well. This function is in direct proportion to the value. Tests were performed with different types of MRs. Data are presented in the dimensionless form relatively to a basic point in which the measured (a circle) and calculated (a cross ) values are in good agreement. The mass flow rate values correspond to those which are typical for 1 and 2 cfm compressors working in a pressure range of 15 to 25 atm for the supply flow and of 1.5 to 6 atm - for the return flow. Figure 5 presents data for two types of the counter-flow heat exchangers a soldered pair-tubes and a tube-in-tube (1x1) obtained with the MR of a different composition: and Figure 6 presents data obtained for a tube-in-tube (7 in 1) type heat exchanger with the MR based on: and and 6, which have different contents of the components. Comparison of experimental and calculated data which were obtained only with the model without any experimental adjustment shows that they are in a good agreement. Even at the relatively high flow rate, the discrepancy of 25 % is quite acceptable for predicting a cooler performance. The methodology allows the use of empirical coefficients for further improvement of the accuracy. We were not able to derive a model for hydraulic pressure drop values which is in a good agreement with experiments in a wide range of the tested parameters. Therefore, this model includes a correction coefficient which is defined from the tests. It was found that such a coefficient is essentially constant for the same type of the heat exchanger. Thus the developed models allow further system analysis to take into account the heat exchanger influence on the cooler performance. THE THROTTLE-CYCLE COOLER PERFORMANCE
The computer-software which was developed with the methodology described above has predicted the cooler refrigeration capacity with an acceptable accuracy at different test conditions and composition of the MR. For example, Fig. 7 shows the influence of HE length on the value. The data are presented relative to a maximum value of which was 50 W and 80 W for the heat exchanger with 2 m and 4 m lengths, respectively. The MR was a threecomponent hydrocarbons blend. Predicted and experimental data are in a good agreement, especially at a relatively high refrigeration capacity. For this test a correction coefficient for DP calculations was adjusted at the point corresponding to
Figure 5. Overall average heat-transfer coefficient for paired-tube heat exchangers.
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Figure 6. Overall average heat-transfer coefficient for the tube-in-tube (7x1) heat exchanger.
Figure 7. Influence of the heat exchanger length on the cooler refrigeration capacity.
The predicted refrigeration map and the experimental data are presented in Fig.8 for the modified cooler with the heat exchanger optimized to increase the refrigeration capacity. A correction coefficient for was adjusted at one experimental point with MR1. Comparison shows that calculated and experimental data are also in a good agreement. The system optimization process described above was used in development of practical cryocoolers that are being used in several applications such as detector (infra red, FTIR, gamma etc.) cooling, electron microscopes, portable MRI systems, telescopes, gas chromatography, nondestructive testing, SQUID, water trapping, etc. A typical cryocooler cold end is shown in Fig. 9. This cold end was developed to work with an oil lubricated 1 cfm compressor [2] . This cooler has the heat exchanger design-A that was optimized to work with a set of mixed gas refrigerants in the temperature range of 70 K to 120 K. The optimization was carried out with some constrains such as the physical size of the cold end, ability to run multiple cold ends with one compressor and to separate the each cold end up to 100 ft from the compressor. To satisfy these constrains the MR should be in gaseous state beyond the aftercooler.
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Figure 8. Refrigeration capacity of the modified cooler.
The system performance was further optimized by proper selection of the cold plate design to reduce the temperature difference between the heat load source and the MR with a heat flux up to
Another heat exchanger design-B was also developed to increase the refrigeration capacity of the system without changing the compressor size. The computer analysis was useful
in designing the heat exchanger to obtain 15%-30% higher without changing the system design. This heat exchanger would still fit within the same diameter vacuum container but is slightly longer as shown in Fig. 9. The same types of the heat exchanger (A & B ) were used to design water traps[3] (AquaTrap®). In this application the major constrain was not to cooldown below 110 K temperature at minimum load condition and operate within 130 K temperature with of applied load. This was achieved with the unique MR to fit the specialized cold tip design. Same heat exchanger was also used to produce about 100 watts of refrigeration at about 175 K temperature with a different refrigerant blend. Experimental point are shown in Fig. 2 by stars. This was achieved with some minor system modifications. It was also concluded using computer analysis that the heat exchanger length could be reduced by 30% and still same performance could be achieved.
Figure 9. Heat exchanger outline.
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SUMMARY
Optimal design of the MR throttle-cycle cooler may be based on the model which describes interaction of the compressor and the counter-flow heat exchanger for the selected mixed refrigerant composition. Models based on the assumption of the homogeneous structure of the two-phase flow may be used for describing overall-average heat transfer coefficients and hydraulic pressure drop with an acceptable accuracy to conduct an optimal system design. The proposed criteria for the counter-flow heat exchanger efficiency helps to compare heat exchangers of different types and customize a cooler for the customer needs. The second generation of the throttle-cycle coolers improve refrigeration capacity by as much as 15 - 30 % in the temperature range of 70 K to 175 K running different mixed refrigerants. ACKNOWLEDGEMENT
R. Longsworth/ S. Harold of IGC-APD Cryogenics and B.Yudin of IGC-InterCool Energy are thanked for their help and contributions. REFERENCES 1. Boiarski M.; Brodianski V.; Longsworth R., “Retrospective of Mixed-Refrigerant Technology and Modern Status of Cryocoolers Based on One-Stage, Oilk-Lubricated Compressors”, Pap.DDS presented to Cryogenic Engineering Conference,
1997/ Portland, Oregon. 2. Khatri A.; Boiarski M., “A Throttle Cycle Cryocooler Operating with Mixed Gas Refrigerants in 70 K to 120 K Temperature Range”, Cryocoolers 9, Edited by R.G.Ross,Jr./ Plenum Press, New York, 1997. 3. Little W.; Sapozhnikov I., “Low-Cost Cryocoolers for Cryoelectronics”, Cryocoolers 9, Edited by R.G.Ross,Jr./ Plenum Press, New York, 1997. 4. Khatri A.; Boiarski M.; Nesterov S., “Water Trap Refrigerated by a Throttle-Cycle
Cooler with Mixed Gas Refrigerant, ICMC/ CES, 1997. 5. Missimier D., “Fast-Cycle Pump Forecast: Cold & Dry”, Photonic Spectra, 1984, The Optical Publishing Company, Pittsfield, MA. 6. Grodent M., et al. “Modeling Compressors and Chilling Systems”, Proc. Int. Conf. Energy efficiency in refrigeration and global warming impact, I.I.R. Com. B1 /2, Belgium, 1993. 7. Grezin A.; Landa J., “Design Method for J-T Microcooler Heat Exchangers Applying
Multicomponent Refrigerant”, Proc. V.III b, 19th Int.Cong. Refr. 1995, Equipment and Processes, The Netherlands, The Hague, 1995. 8. Heat Exchanger Design, Handbook, V.1,2. 1983 Hemisphere Publishing Corp. 9. Numerical Optimization Techniques for Engineering Design, by Garret N. Vanderplaats, McGraw-Hill Book Company, 1984.
Long-Life Cryocooler for 84-90 K V.T. Arkhipov, A.V. Borisenko, V.F. Getmanets, R.S. Mikhalchenko, and L.V. Povstiany
Specialized R&D Bureau for Cryogenics Technologies Kharkov, Ukraine
H. Stears Orbita Ltd, Kensington, MD
ABSTRACT A combined cooling system has been developed based upon a Joule-Thomson (J-T) gas mixture refrigerator and a melting-type thermal storage device. The system provides for high stability of the cryogenic load temperature within the 84 to 90 K temperature range for loads between 0.1 and 2.0 watts. To achieve a 1 to 5 year operational life, the J-T refrigerator uses a double-stage lubrication-free compressor with permanent lubrication of its bearings, and pistonto-bore sealing via non-wearing clearance seals. The cryocooler is designed primarily for stationary or mobile machinery that requires a long-life, high-reliability, cryogenic cooling capability. It was assumed in the design that the most critical factors were provision for a high degree of temperature stability, operational availability, and long-term reliability; there was no strict limitation on weight or power consumption for the cooling system. The developed scheme is simple in concept, provides relatively high efficiency, and additionally achieves low levels of mechanical and electromagnetic interference. It is suitable for supporting either scientific research or commercial applications, and with modification, could also be used on a spacecraft. INTRODUCTION
The task of achieving continuous, stable cooling at cryogenic temperatures, free from mechanical vibration and electromagnetic interference, has been an important refrigerator design goal for many years. For such applications, J-T type coolers have particular advantages over other high-efficiency coolers like integral Stirling coolers because the compressor of a J-T cooler can be located a considerable distance from the cooled object. In contrast to regenerative type coolers, J-T systems are also unique for their ability to achieve high theoretical efficiencies; the main losses of real machines are generally related to compressor efficiency. Utilization of gas mixtures1 as the working fluid for J-T systems makes their efficiency very competitive with other coolers, particularly within the 80 - 110K temperature range. Substitution of nitrogen-hydrocarbon or nitrogen-freon mixtures for pure nitrogen (a common working fluid for J-T cryocoolers) has increased the reliability and life of compressors and hence, of the system as a whole, thanks to a substantial decrease of the required working pressures in the cycle. Cryocoolers 10, edited by R. G. Ross, Jr. Kluwer Academic/Plenum Publishers, 1999
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This allows the option of using single- or dual-stage compressors in lieu of three- or four-stage units as required for pure nitrogen. To increase cryocooler life from 2,000 - 3,000 hours up to 30,000 - 50,000 hours, we have adopted the concept of a cryocooler combined with a phase-change thermal storage device, which we refer to as a thermoaccumulator (TA). In this concept, a J-T cryocooler using a mixed-gas working fluid for enhanced cooling efficiency is operated with low duty cycle to periodically refreeze the thermoaccumulator thermal storage material which passively maintains the temperature of the cooled load. The main technical problem with the proposed system is in achieving: 1) a decent mechanical lifetime for the compressor unit, 2) a high reliability for the throttle device, and 3) a high efficiency thermoaccumulator. The main difficulty in ensuring adequate lifetime for the compressor is achieving effective lubrication of the compressor drive mechanism. In the development of the compressor drive mechanism, we have applied the experience we have gained at engineering long-life piston compressors with permanent lubrication. With respect to eliminating blockage of the J-T throttle components, with say, oil, water, or wear products, it is necessary to carefully manage the intrusion of these elements into the gas stream. For this reason, we use special lubricants that are soluble in the working fluid at low temperatures, or use multistage cleaning systems to remove the contaminants from the gas stream. However, the selection of such special lubricants for temperatures below 80 K is still rather complicated. The third technical problem is in creation of a thermoaccumulator to ensure high temperature
stability as well as minimum parasitic heat inflows. SR&DB has gained considerable experience in engineering integrated cryostats3, both with passive and active thermal protection systems. Still, small-size cryostats are always rather complicated individual devices. SYSTEM TECHNICAL APPROACH
A schematic of the chosen cryocooler concept is shown in Fig. 1.
Figure 1. Schematic of the J-T cryoocooler concept: 1-compressor unit; 2- heat exchanger 3- finite heat exchanger; 4- combined filter; 5- throttle counterflow heat exchanger; 6- cryogenic-load heat
exchangers; 7- receiver tank; 8- electro-controlled valves; 9- pressure transducers; 10- temperature sensors; 11- shutoff valves; 12,13- valves; 14- body; 15- working capacity; 16- cold-duct; 17- drainage/filling
pipes; 18- vacuum pumping valve; 19- cryo-adsorption pump; 20- receiver-tank; 21- shut-off valves; and 22- pressure transducers.
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The cryocooler is composed of three units: • The J-T compressor and gas mixture supply
• The thermoaccumulator with J-T cryostat • The electric controls and power supply
Referring to the numbered items in Fig. 1, the main components of the compressor unit are the compressor itself (1), which accomplishes compression of the gas mixture circulating within the J-T system, a service heat exchanger provided for cooling the compressor itself, and its interstage heat exchanger (2). These compressor heat-rejection heat exchangers, jointly with the J-T loop precooling heat exchanger (3), are cooled by an exterior fluid loop using water or antifreeze. Also provided in the compressor unit, is a combined filter (4) to ensure thorough separation of any oil or mechanical contamination from the working gas.
Both the J-T counter-flow heat exchanger (5) and cryogenic-load heat exchangers (6) are housed inside the thermoaccumulator unit to minimize parasitic heat loads. Receiver-tank (7)
and electro-controlled valves (8) are provided to restore and to adjust the gas mixture proportions, and to maintain the J-T thermodynamic-cycle parameters. Monitoring and control of the J-T cryocooler are carried out via pressure and temperature transducers (9) and (10), respectively, located within the compressor unit and within the TA. To enable refilling and storage of
the system, shut-off valves (11) are provided. The thermoaccumulator unit, generally speaking, is a cryostat whose phase-change material (15) is thermally insulated and supported within a vacuum-tight housing (14). The phase-change material is imbedded with a system of heat removing fins that are interfaced via thermal contact with the cryogenic-load heat exchanger (6) and the cold-duct of the cooled-object interface (16). The phase-change material is filled through drainage/filling pipes (17). Vacuum in the insulation space is maintained with a cryo-adsorption pump (19). The thermoaccumulator may also have an
exterior expansion receiver-tank (20) for storage of the phase-change material when in its gaseous or expanded state at elevated temperatures. Other variants are possible for storage of the phasechange material at room temperature within its working volume, particularly if propane is used as the phase-change material. For control and service of the thermoaccumulator, shut-off valves (21) and pressure transducers (22) are provided. The required dimensions of the thermoaccumulator are determined by: • Cooling capacity of the J-T cryocooler • Cooling load for the cooled object • Parasitic loads into the TA during compressor switch-ons/offs.
The TA-unit dimensions may also depend on requirements for a guaranteed cold holding period for the cooled object during a machine emergency shutdown or auxiliary system breakdown. In Eq. 1, the parameter r is the phase transition (melting) heat, and phase-change material. The shut-off pause time length mance time-length:
where
is the density of the
in cryocooler operation, correlates with the active perfor-
is the time required for the cold-generator to cool down to its operating temperature.
The longevity, or lifetime (L), of the overall system depends on the active cryocooler lifetune as follows:
where k is a relative time function for TA initial cooling. It follows from Eq. 1 to 3, that, for a significant increase in system lifetime, it is necessary to use a cryocooler with a large cooling capacity relative to the cryogenic load, i.e.
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In addition, it is necessary that the parasitic heat inflows to the TA cryostat be minimal, i.e. Since increased switched-on time also brings enlargement of the required TA capacity, and increased parasitic heat inflows and weights as well, it is possible to estimate optimal cryogenerator operation regimes from a perspective of weight, lifetime, and energy consumption rate. In our systems, the duration of the compressor active operation is usually about 1–1.5 hours for every cycle, and the length of a passive cool-down period is about 5–10 hours, depending on the cryogenic load Also critical is the time of initial cooling and freezing of TA phase-change material. Still, by operating the cryocooler continuously at start-up, this period can be brought down to 5 - 40 hours for different combined cooling systems. LUBRICATION-FREE COMPRESSORS
There are two directions in engineering of compressors for J-T cryocoolers: 1) Application of conventional-lubrication compressors with special low-temperature lubricants together with a multistage filter system for separation and cleaning of the working fluid from traces of oil 2) Usage of “clean,” particularly, membrane-type, compressors4 In the SR&DB, a series of sealed compressors has been developed and tested (with positive results) for operation with gas mixtures; these compressors operate without the need for liquid lubrication of the crankcase and crankshaft mechanism. For purposes of continuous cooling (with the use of thermoaccumulators), a series of compressors has been engineered with capacities ranging from 1.5 to for gas compression from 0.05–0.1 to 4-6 MPa. These compressors have been designed to be driven by both hermetically sealed phase-asynchronous 0.5 to 2 kW electric motors with a speed of 1,470 rpm, and brushless DC motors. Representative compressors are shown in Fig. 2. The compressors have undergone a complete test cycle and have been fabricated in experimental-quantity batches. Operational life has proved to be 5,000 to 10,000 hours for an operating regime involving cyclic shut-offs every one to three hours. All of the compressors are provided with clearance-type seals for the piston/bore sliding surfaces; the fixed-clearance dimension is maintained by ball guides. Two-stage, two-cylinder machines of "V" or linear orientation are used to provide for compression pressures up to 4 - 4.5 MPa. For pressures over 6 MPa, three-stage, three-cylinder machines in a star-configuration are used. To compensate for the starting moment, a balancing scheme is used, driven by two oppositely rotating motors. The compressors are liquid-cooled from an exterior fluid-loop system. An air-cooled version is also available.
Figure 2. Compressors developed by the SR&DB.
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471
Figure 3. J-T cryocooler cooling capacity, depending on cryostat temperature, for different fixed intake pressures: 1- 0.06 MPa, 2- 0.08 MPa, 3- 0.1 MPa, 4-0.12 MPa, 5 and 6- at permanent mass values of gas mixture in the J-T cryocooler; the capacity of the compressor unit is 3 m3/hr.
J-T cryocoolers containing compressors are furnished with a combined filter-adsorber for separating mechanical particles and oil vapors from the working fluid. To stabilize the temperature level and maintain high efficiency, each system is supplied with pressure transducers and automatic valves to keep a preset pressure within the crankcase. Delivery pressure is not as critical for gas mixtures5; still, strict pressure control is available. Fig. 3 displays the system characteristics for regimes of: 1) a fixed intake pressure, and 2) preservation of the mixture amount within the system (at a fixed delivery pressure). It is apparent that while selecting this or that type of regulation system, it is possible to provide a variety of active control modes. The developed system has proven its effective operation with 3 to 5-m long connection piping between the compressor and the cryostat. Operational life of such a combined system, with no servicing or gas-mixture change out, has proven to be 3-5 averaged years. The system is run via automatic controls and interlock protection throughout the entire operational cycle; it is also equipped with emergency shut-off means to protect against emergency compressor-cooling or power-supply shutdowns. THERMOACCUMULATORS
The design of thermoaccumulators, TA, involves competing system-level requirements in the areas of dimensions, weight, thermal hold period, and service and maintenance feasibility. In addition, technical issues must be addressed involving the properties of the phase-change material, cryocooler parameters such as capacity and efficiency, and other technological and design constraints6. The introduction of a TA as a cryogenic device with its own piping, support mounts, insulation (and hence added parasitic heat inflows), brings about an overall reduction of the total cryocooler efficiency (Eq. 1 to 3). Such disadvantages are partly compensated by achieving close integration with the cooled object within the cryostat, by accruing reliability advantages associated with having the stored cold reserve, and by the improved efficiency that generally accompanies a higher capacity cryocooler. The thermoaccumulator has been optimized to enhance the heat transfer processes by selecting between substances characterized by saturated-vapor high pressures (such as argon or methane), as compared to different substances and their mixtures or solutions. Various thermoaccumulators have been developed and tested with or without an exterior receiver3 used for storage of the whole amount of the phase change material (gaseous or liquefied) within a cryostat under increased pressure. The latter variant is expedient with phase change materials whose critical temperature exceeds room temperature. In particular, propane has been used as such a material, with special additives to avoid its over cooling. The working pressure
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J-T AND THROTTLE-CYCLE CRYOCOOLER DEVELOPMENTS
Figure 4. J-T cryocooler 84 – 200 K.
inside the TA-volume does not exceed 2.0 MPa, and the unit itself features a compact and operation-feasible design. However, for applications desiring increased thermal storage that also have weight and parasitic heat inflow constraints, it is necessary to opt for systems with an exterior receiver bottle. It is recommended that inexpensive and widely available gases such as argon, nitrogen, or methane be used to ensure the possibility of operating the systems with a single-time usage of the phase change material, and to utilize a standard-make dry clean gas bottle as the receiver. Argon has very favorably justified itself with ground-based cryostatic systems thanks to the intensive heat-mass transfer5 that occurs due to vapor condensation upon the cooled surfaces of the cryostat walls and fins; this provides minimal temperature drop during the thermoaccumulator
cool-down period. Common constructional materials for the TA include aluminum alloys, copper, and stainless steel. Support components are fabricated of fiberglass, with high-vacuum, or screen-vacuum type insulation. In total, the SR&DB experts and specialists have engineered and tested TA units from 1 to capacity ensuring the passive cooling of objects with plus-minus 0.5 K stability, for periods ranging from 3-5 hours up to 1-2 days. Large thermal masses, together with prolonged turnon/shut-off cycle periods, has proven the feasibility and reliability of the stand-alone automatic control operation. Characteristic values of parasitic heat inflows to TA-cryostats are 0.25 to 0.4 W in the passive mode, and 0.2 to 0.3 W with the J-T operational. 84 – 90 K CRYOCOOLER
The exterior appearance of a complete representative J-T cryocooler with cryostat is presented in Fig. 4. The J-T cryocooler, which uses propane as the phase change material, displays the following characteristics: • operational temperature range, K: 84 to 200 • cooling capacity, W: 0.5 to 20 • compressor capacity, 1.2 • gas delivery pressure, MPa: 4.0 • input pressure, MPa: 0.1 • coolant: pure gas or gas mixtures • weight, kg: 30 • power consumption, W: 500 • overall dimensions, mm: 500 dia. by 700 • lifetime, years: 1 to 5
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473
The compressor unit (a 27 VDC – circuit drive variant) is mounted within a sealed cylindrical container that houses the electric motor commutation circuitry and monitor and control electronics. A cylindrical shaped shroud covers the V-shaped compressor unit, the receiver, and their relevant armature. The thermoaccumulator unit, furnished with a technological pressure valve, Fig. 4, is located apart from the other components. The throttle-type heat exchanger is separate from the cryostat for ease of feasibility testing and mock-up preparation. The TA is designed with elongated piping outside the cryostat unit. The weight and dimensional features of the custom-made TA have been made quite compact thanks to the usage of an exterior receiver. Modifications have been examined that would allow the system to be operated in the 70 to 75 K temperature range. ACKNOWLEDGMENT
The authors express their gratitude and acknowledgment to the U.S. Naval Research Laboratory and to the U.S. Air Force Research Laboratory for their assistance in task definition. REFERENCES 1.
V.T. Arkhipov, et al., “Multicomponent Gas Mixtures for J-T Cryocoolers,” Cryocoolers 10, Plenum Press, New York, 1999.
2.
A.K. Grezin and V.S. Zinoviev, “Microcriogennaia Technika,” Moscow, 1977, Maschinostroenie.
3.
R.S. Mikhalchenko, V.T. Archipov, et al., “Teplovye processy v kriogennych sistemach,” Kiev, 1986, Naukova Dumka, USSR.
4.
D.C. Bugby, “Development of 60 K Thermal Storage Unit,” Cryocoolers 8, Plenum Press, New York, 1995, pp. 313-320.
5.
V.T. Arkhipov, et al., “Nizkotemperaturnye processy i sistemy,” Kiev, 1977, Naukova Dumka, USSR, pp. 74-81.
6.
B.I. Verkin, et al., “Teplofizika nizkotemperaturnogo sublimatcionnogo ohlajdeniya,” Kiev, 1980, Naukova Dumka, USSR.
Mixed Gas J-T Cryocooler with Precooling Stage A. Alexeev, Ch. Haberstroh, and H. Quack
Lehrstuhl fuer Kaelte- und Kryotechnik Technische Universitaet Dresden 01062 Dresden, Germany
ABSTRACT
A mixed gas Joule Thomson refrigerator offers decisive advantages for several applications. Requirements such as an extreme long MTBF and long life, low levels of vibration and noise can be satisfied by such a refrigerator. However, the known coolers of this type have a relative low thermodynamic efficiency. An improvement of the efficiency is possible by introducing a precooling stage. This method will be discussed in the paper. A mixed gas JT refrigerator with precooling stage has been tested. The efficiency proved to be 1.5 times better compared with a Gifford-McMahon refrigerator. INTRODUCTION
The object under discussion here are coolers with a cooling capacity of more than 50 W at 90-110K. Mixed gas JT-coolers have a good chance to be introduced to the market in this capacity range, because of their low cost. However, to be really competitive, e.g. compared with Gifford-McMahon refrigerators, an improvement of the efficiency of mixed gas JT coolers is necessary. We have tried this by varying of the mixture composition but found that the greatest improvement can be obtained by the introduction of a precooling stage. In the large scale refrigeration the propane precooled mixed refrigerant process was introduced successfully by Lee S.Gaumer1 for natural gas liquefaction. MIXED GAS JT COOLER
The main components of a Joule Thomson system are compressor, counterflow heat exchanger (JT heat exchanger), throttle valve and evaporator.
The compressor supplies the high pressure flow of the mixed refrigerant, which flows through the JT heat exchanger, expands through the throttle valve and passes the evaporator, where it absorbs heat from the object to be cooled. The low pressure stream is then circulated back through the counterflow heat exchanger into the compressor. The thermodynamic cycle has been discussed in detail earlier2,3.
Cryocoolers 10, edited by R. G. Ross, Jr.
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J-T AND THROTTLE-CYCLE CRYOCOOLER DEVELOPMENTS
Figure 1 JT cycle with precooling .
Figure 2 T-h-diagram for the mixed gas JT cycle with precooling.
WHY PRECOOLING?
The refrigeration capacity of a mixed gas JT refrigerator is determined by the enthalpy difference between low and high pressure stream at the pinch point in the heat exchanger. For optimum mixtures the pinch point is situated at the warm end (similar to a JT cycle with pure refrigerant). The pinch point enthalpy difference depends on the temperature. At temperatures of
about 240 K it is approximately 1.5...2 times larger than at 300 K. Consequently the cooling capacity of the cooler can be 1.5...2 times higher. In order to use this effect the ,,ambient temperature“ should be shifted down to about 240 K. An additional cold source (precooling stage) can be used for that. Fig. 1 illustrates this principle. The cycle consists of compressor, 3-stream heat exchanger (precooler) and the conventional JTstage. The high pressure stream is cooled by an additional cold stream in the precooler before entering the JT-stage. The cooling capacity of the low pressure stream, which remains after the
JT stage, is used in this heat exchanger, too. The specific cooling capacity of such a cycle is given by: with - enthalpy difference between high pressure and low pressure stream at ambient temperature, - specific cooling capacity of the precooling cycle. Fig. 2 shows a typical temperature-enthalpy diagram for a mixed gas JT cycle with precooling stage. The high pressure mixed refrigerant (optimized for the cycle with precooling stage) is totally gaseous at ambient temperature. Therefore is relatively small. This value is considerably smaller than Consequently the necessary cooling capacity of the precooling cycle has to be in the same order of magnitude as the cooling capacity of the main cycle:
The COP of refrigerators used for temperatures of about 235-245 K is in the order of 1. Therefore the power consumption for the precooling refrigerator is numerically equal to the cooling capacity of the precooling cycle and is about 10 times smaller than the power consumption of the mixed gas compressor. Thermodynamically this small additional energy investment is profitable, because the cooling capacity of the mixed gas refrigerator is increased by a factor of 1.5-2.
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PROCESS DESIGN A mixed gas JT cooler with precooling
stage was developed at the Technische Universitaet Dresden. Fig. 3 shows the flow diagram of this refrigerator. It includes two
separate cycles: a mixed gas cycle and a precooling cycle.
Fig. 3 Flow diagram of the mixed gas JT refrigerator with precooling stage.
Mixed Gas Cycle The mixed gas is discharged at high pressure from the main compressor. The gas flows through the aftercooler and the oil separation unit. The separated oil is returned to the main compressor. Instead of the 3-stream exchanger shown in Fig. 1 two separate counterflow heat exchangers are provided. In the first heat exchanger the high pressure stream is cooled down by the low pressure stream from the JT stage. The further precooling takes place in the second heat exchanger by the cold stream from the precooling cycle. The high pressure stream than flows into the JT-stage to produce the cooling power at the required temperature.
A mixture of nitrogen, methane, ethane and propane is used. The composition was optimized* to achieve the maximum specific cooling capacity at 95-100 K. Oil Removal The use of an oil lubricated compressor has the disadvantage of oil contamination in the high pressure stream. If the oil contamination in the mixture exceeds compatible quantities, this would result in clogging in the cold box at low temperatures. Therefore an oil separation for the high pressure gas after the compressor is required. The required purity of the mixture can not be guarantied by conventional filter technology under all circumstances. In order to overcome this, we condense the vaporous oil by cooling of
the high pressure stream to below room temperature**. This can be realized by introducing an additional heat exchanger (oil condenser) before entering the oil filter. The high pressure gas is cooled to approximately 0°C, the vaporous oil is condensed and removed from the refrigerant by a conventional oil filter. This way the necessary purity of the refrigerant for failure-free operation can be achieved. Precooling Cycle The precooling cycle has two functions: precooling of the high pressure stream and cooling of the oil containing high pressure stream in order to condense vaporous oil. We use a simple throttle cycle (similar to domestic refrigerators) for precooling. The refrigerant is discharged at high pressure from the compressor and liquefied in a condenser at * Patent pending **
Patent pending
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J-T AND THROTTLE-CYCLE CRYOCOOLER DEVELOPMENTS
ambient temperature. A portion of the liquid refrigerant expands in a first throttling device. The cold refrigerant flows via precooler back to the compressor. The rest of liquid refrigerant expands in additional condenser, Refrigerant cycle.
a second throttle valve, producing cold. This is used in the oil and returned to the compressor. R507 was used for the precooling
Further Advantages The process design with a precooling stage brings additional advantages. The refrigerant
mixture can consist of four components only: high-boiling components like butane are not longer needed, the percentage of components like propane can be smaller. Therefore mixture design and composition control are more simple. A further effect is that solidification of these components at the lowest temperatures
can be prevented better this way. HARDWARE Our mixed gas JT cooler consists of two separate modules: the compressor unit and the cryostat, which are connected by gas lines. Figure 4 Cryostat interior The compressor unit consists of two compressors. The mixed gas compressor is a single stage oil lubricated rolling piston compressor (power consumption about 1 kW) with aftercooler and oil removal unit. The R507 compressor is
a smaller hermetic compressor with a condenser for the precooling loop (standard Danfoss unit). A multitube heat exchanger was used. For the high pressure stream eight small 4 mm copper tubes are provided, which are placed in a larger 16 mm tube. The low pressure stream flows in the larger tube in the space between the small tubes. The arrangement is wound into two spirals (with approx. 120 mm and 180 mm diameter). Figure 4 shows the cryostat interior. TEST RESULTS
The tests of the system started in the end of January 1998. So far more than 400 operating hours of the refrigerator have been accumulated. The cooling capacity of the system depends on the composition of the mixture. With an optimum mixture composition a cooling capacity of approximately 100 W was found at 100 K (Fig. 5). The energy consumption of the system amounts to approximately 1.1 kW. With the same mixture a minimum cooling temperature of about 83 K was achieved. This refrigerator is more effective compared with known mixed gas JT systems. At temperatures above 95 K this unit is more than 1.5 times more effective than a comparable Gifford McMahon refrigerator. RELIABILITY
The reliability of the system will be only slightly affected by the introduction of the precooling stage, because all additional components are as reliable as the original components of the mixed gas refrigerator.
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COSTS
The addition of the precooling stage does not cause a considerable increase in costs of the system, because all additional components are ,,low cost“ elements, available from conventional refrigeration. The costs of the compressorcondenser-unit (used as precooling stage) are less than 200$. Even including the costs
of the necessary periphery (heat exchangers, connection lines and throttling devices) and some additional expenditure for installation of these elements, the excess costs are lower than 1/10 of the cost of the complete system. SUMMARY
Figure 5 Cooling capacity of the mixed gas JT refrigerator with precooling stage .
A mixed gas JT refrigerator with precooling stage was developed, built and tested. This system is more effective (at least 1.5 times) compared with a Gifford McMahon refrigerator. The reliability of the system will not be essentially affected by the introduction of the precooling stage. The additional investment is negligible in comparison to the achieved improvement in efficiency. REFERENCES
1. Gaumer L.S., Newton Ch.L., US Patent # 3763658, Oct.9, 1973 2. A.Alexeev, Ch.Haberstroh, H.Quack, Low cost mixture Joule Thomson refrigerator,
Proceedings of the 16th International Cryogenic Engineering Conference, Kitakyushu, Japan, 1996 3. A.Alexeev, Ch.Haberstroh, H.Quack, Further development of a mixed gas Joule Thomson refrigerator, presented at the Cryogenic Engineering Conference, Portland, USA, 1997
Experimental Comparison of Mixed-Refrigerant Joule-Thomson Cryocoolers with Two Types of Counterflow Heat Exchangers E.C. Luo, M.Q. Gong, Y. Zhou, J.T. Liang Cryogenic Laboratory of Chinese Academy of Sciences Beijing 100080, China
ABSTRACT A miniature, closed cycle Joule-Thomson cryocooler with gas mixtures operating between 67.5K to 100K was investigated experimentally. In the experiment, two types of miniature counterflow heat exchangers were tested. One is so-called classic Hampson type heat exchanger, and another is a modified perforated plate heat exchanger which was recently developed by the authors. In the modified perforated heat exchanger, a thin stainless tube was used as a partition which was traditionally formed by diffusion welding method. The two types of heat exchanger have the same outline sizes: their diameters are 14mm, and their lengths are 120mm. With the same mixture and the same discharging pressure, the Hampson type of Joule-Thomson cryocooler achieved 70K, but the modified perforated type Joule-Thomson cryocooler achieved 67.5K. The former one obtained 2W cooling capacity at 80K, and the later one obtained 2.4W cooling capacity at 80K. Due to the better performance and simpler structure, the modified perforated plate Joule-Thomson cryocooler may become a good alternative to the classic Hampson type Joule-Thomson cryocooler in the future. INTRODUCTION Many electronic devices operating in the range of around liquid nitrogen temperature are in the phase of practical application out of laboratory. However, an obstacle to this application is the lack of a reliable and low-cost cryogenic source. Thanks to recent development of a closed cycle mixed-refrigerant Joule-Thomson cryocooler driven by a commercialized single stage oillubricated compressor, this deadlock seems now to be overcome. Using mixtures for liquid nitrogen temperature range began in the seventies in the former Soviet Union.1 However, before the nineties, a commercialized, single stage, oil-lubricated compressor had not been used in the cryocooler, thus a commercial product of mixedrefrigerant Joule-Thomson cryocooler with high reliability and low cost had not yet been
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available. From the beginning of the nineties, mixed-refrigerant, throttle refrigerators driven by a single stage oil-lubricated compressors for obtaining liquid nitrogen temperatures have been
studied, and today this technology is getting more and more attentions of cryogenic scientists.2,3,4,5,6 Two types of compact Joule-Thomson cryocoolers using a special mixed-refrigerant which can operate in the temperature range from 70K to 100K have been developed successfully in our laboratory. The thermodynamic performance and operating parameters of this cryocooler are reported here. MIXED-REFRIGERANT JOULE-THOMSON CRYOCOOLER
Figure 1 is a schematic of a typical mixed-refrigerant Joule-Thomson cryocooler driven by a single stage oil-lubricated compressor. The compressor unit generally consists of a single stage oil-lubricated compressor, an aftercooler with fan and an oil separation unit. Two gas lines
connect the compressor to the Joule-Thomson cryostat. The cryostat is made up of a counter-flow heat exchanger, a throttle valve and an evaporator. The gas is typically discharged at 1.2MPa to 2.5MPa with suction pressure from 0.1 MPa to 0.4MPa and then flows through the aftercooler and the oil separation unit. The oil from oil separation unit is returned to the compressor. The high pressure gas then goes to the cryostat. In the cryostat the gas passes through the counter-flow heat exchanger, where the high pressure mixture gas is cooled by the returning low pressure gas mixture. The high pressure gas mixture expands on when passing through a throttle valve and produces the cooling effect in the evaporator. COUNTERFLOW HEAT EXCHANGERS
The counter-flow heat exchangers of the cryostat in the experiment are very compact. Two types of heat exchanger were tested. One is a classic Hampson type, miniature heat exchanger. This type of heat exchanger are widely used in miniature fast-cool-down Joule-Thomson cryocoolers. In the beginning round of our experiments, our mixed-refrigerant Joule-Thomson cryocooler used this kind of heat exchanger, and it can successfully achieve temperatures below liquid nitrogen. But it requires a special technique for fabricating it. Therefore, later, we developed
Figure 1.
Schematic of a typical single-stage, oil-lubricated, mixed-refrigerant J-T cryocooler.
COMPARISON OF MIXED-REFRIGERANT J-T CRYOCOOLERS
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another new type of counterflow heat exchanger, the perforated plate heat exchanger with a very simple structure. Unlike a traditional perforated plate heat exchanger with a partition by diffusion welding technique, our new perforated plate heat exchanger used a very thin stainless tube as the partition between high and low pressure passages instead, so it can definitely eliminate the trouble of fluid leakage from high pressure passage to low pressure passage. Moreover, compared with a Hampson counterflow heat exchanger, our new perforated plate heat exchanger has two advantages. One is that it can effectively make use of full space and consequently has a bigger heat exchange surface area, and the other is that its fabrication is much more simpler. The experiment shows that the new perforated plate heat exchanger can be more compact, more effective and cheaper than the Hampson one. Figure 2 is a schematic diagram of the new structure perforated plate heat exchanger. Table 1 gives some structure parameters of the two types of counterflow heat exchangers.
Figure 2.
Schematic diagram of the new perforated plate Joule-Thomson cryocooler.
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CRYOGENIC MIXED REFRIGERANTS
The thermodynamic performance of a mixed-refrigerant Joule-Thomson cryocooler heavily depend on the used mixtures. There are three kinds of throttling effects: differential effect, integral effect and isothermal effect. Among them, the isothermal effect is the most important parameter for analyzing Joule-Thomson cycle. The isothermal effect is determined by the
working substance, the operating pressures and the temperatures. To improve the thermodynamic performance of a mixed-refrigerant Joule-Thomson cryocooler, it is necessary to increase isothermal throttle effects over the whole operating temperature range. Thus, the first thing is to select components of mixtures, and the second thing is to optimize the mixture to maximize their isothermal effects over the whole operating temperature range. Based on thermodynamics, the isothermal effect of a mixture is greatly relative to the specific heat capacity, the vapor-liquid phase change latent of each component. Among them, the vapor-liquid phase change latent of the component possibly makes a bigger contribution to the isothermal effect. Therefore, the preferable components are selected according to their boiling temperatures. There are a large number of substance from liquid nitrogen temperature to room temperature (e.g, 300K). However, considering the solubility limit of the mixture components, one can further determine the possibly selected components. Furthermore, if the other factors (e.g., safety factor and economic factor) are considered, only a few number of pure substances are the preferred candidates. These components are classified into seven groups shown in Table 2
according to their boiling temperatures. To form an efficient mixed refrigerant, we can choose one component from each group in Table 2. Group 5 and group 6 both include more components than the other groups. The component in the same group means that they are almost alternatives each other. Figure 3 shows the isothermal effects over the whole operating temperature range from 80K to 320K for several pure substance at the low pressure of 0.lMPa and the high pressure of 2.5MPa. According to this
figure, one can easily find the fact that each component has a big isothermal effect corresponding to its vapor-liquid phase change region. In order to form an efficient mixed refrigerant, the proper and adequate components should de included in the mixture. If the components of the mixtures
are not proper or the number of the components in the mixtures are not adequate, an efficient mixture can not be formed no matter how one adjust the compositions of the mixture. Because of an oil-lubricated air-conditioning compressor being used and a better solubility of hydrocarbons in oil, the hydrocarbon-based mixture was preferably used in our experiment. In addition, neon is also included in the used mixture, and the using of neon gas has two functions. One is to achieving a temperature lower than liquid nitrogen temperature, and the other is to increase the suction pressure of the single stage compressor and consequently improve the efficiency of the compressor.
COMPARISON OF MIXED-REFRIGERANT J-T CRYOCOOLERS
Figure 3.
485
Isothermal effect of pure substance.
EXPERIMENTAL RESULTS
Table 3 gives the performance and operating parameters of the two types of Joule-Thomson cryocoolers. To compare the performances of these two kinds of cryocoolers, the same mixture, the same discharging pressure of compressor and the same opening of throttling valve were given. From Table 3, we found that our new structure perforated plate Joule-Thomson cryocooler has a better performance than that of the classic Hampson Joule-Thomson cryocooler. Our new perforated plate Joule-Thomson achieved a no-load temperature of 67.5K and 2.4W at 80K while our classic Hampson one only obtained a no-load temperature of 71.5K and 2W at 80K. In addition, the authors found that the pressure drop of the low pressure passage of our new perforated plate heat exchanger is lower than that of the Hampson type. Finally, both these two kinds of Joule-Thomson cryocoolers can achieved 80K within 35 minutes. Up to now, we built several prototypes of such closed cycle mixed-refrigerant JouleThomson cryocoolers. The life time testing is also ongoing, and the perforated plate JouleThomson cryocooler have been running for about one thousand hours continuously. These results are encouraging, and we will use a more efficient mixture and optimize operating parameters to improve the efficiency of the whole refrigeration system further.
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CONCLUSIONS AND FUTURE WORKS
1. To increase the thermodynamic efficiency of Joule-Thomson cryocooler, the isothermal effects over the whole operating temperature range should be increased. Different component plays a unique role in the different temperature ranges. 2. At the same operating conditions and the same sizes, the perforated plate type JouleThomson cryocooler obtained better experimental results. The lowest temperature of 67.5K was achieved. The two types of Joule-Thomson cryocooler can produce the cooling capacity of 2 W at 80K. 3. Because of its high reliability and low cost, the mixed-refrigerant Joule-Thomson cryocooler driven by a single stage oil-lubricated compressor has good potential for application
in the temperature range around 80K. ACKNOWLEDGMENTS
This research work is financially supported by the National Natural Sciences Foundation of China under contract number 59776002. Also, the authors gratefully acknowledge the support of K.C. Wang Education Foundation , Hongkong. REFERENCES
1.
V.M.Brodianski, V.M. Yagodin, V.A.Nikolsky, and A.V.Ivantsov, “The Use of Mixtures as the Working Gas in Throttle Joule- Thomson Cryogenic Refrigerator”, The 13th International Refrigeration Conference, New York( 1973), pp.43-45.
2.
R.C. Lonsworth, M.J.Boiarski, and L.A.Klusmier, “80K Closed Cycle Throttle Refrigerator”, Cryocooler 8, Plenum Press, New York(1995), pp.523-527.
3.
W. A. Little, I. Sapozhnikov, “Low cost cryocoolers for electronics”, Cryocooler 9, Plenum Press,New York(1997), pp.509-513.
4.
A. Alexeev et al., “Low cost mixture Joule Thomson refrigerator”, Proc. of the 16th Int.Cryogenic Engineering Conf., Okalama, Japan(1996), pp.435-439.
5.
E. C. Luo, Y. Zhou and M.Q. Gong, “80K mixed-refrigerant Joule-Thomson cryocooler driven by R22/R12 compressor”, Advance in Cryogenic Engineering, Vol.43, Plenum Press, New York (1998) (in press)
6.
E. Marquardt et al., “A cryogenic catheter for treating heart arrhythmia, Advance in Cryogenic Engineering”, Vol.43, Plenum Press, New York(1998) (in press)
Multicomponent Gas Mixtures for J-T Cryocoolers V.T. Arkhipov, V.V. Yakuba, M.P. Lobko, and O.V. Yevdokimova Specialized R&D Bureau for Cryogenics Technologies Kharkov, Ukraine H. Stears Orbita Ltd. Kensington, MD, USA
ABSTRACT
This paper presents results of analytical and experimental research performed on multicomponent gas mixtures for closed-cycle Joule-Thomson (J-T) cryocoolers. It is shown that nitrogenhydrocarbon mixtures are capable of increasing J-T cryocooler efficiencies up to 20 to 40 W/W for temperatures in the 80 to 85 K range. The addition of inert neon to the working gas mixture enables lowering of the temperature down to 73 K, and to even as low as 65 K. The influence of separate cycle parameters upon system efficiency is explored for a gasmixture that provides a refrigeration temperature of 80 to 82 K. The formation of a two-phase fluid within the evaporation heat-exchanger provides stability to the refrigeration temperature. Also, partial liquefaction of the mixture prior to entering the counterflow heat-exchanger provides increased cooling efficiency. The use of gas mixtures in combination with lubrication-free compressors has provided the Specialized R&D Bureau (SR&DB) with a series of highly effective J-T cryocoolers with cooling capacities of 10 to 100 W. INTRODUCTION
Closed-cycle J-T cryocoolers are generally not as efficient as other mechanical coolers such as Stirling units for temperatures between 70 and 120 K. However, J-T coolers are widely used for cooling due to their simplicity and ease of integration with infrared, gamma-ray and microwave detectors. They can have very low vibration and electromagnetic interference because the compressor can be located several meters from the cold head. The relatively low efficiency of J-T coolers can often be compensated for by the low parasitics associated with their cold head. Various ways to improve J-T cryocoolers include:
1) Improvement of system designs through miniaturization of heat exchangers and J-T valves, creation of high-efficiency compressors, and incorporation of automatic controls for temperature regulation. In particular, the SR&DB has developed a series of medium-output compressors that does not require liquid lubricants and has long life and very low friction. 2) Usage of a gas mixture in lieu of a pure substance as the working fluid; this can achieve the same thermodynamic advantages as multiple cold stages, but with greater simplicity. Cryocoolers 10, edited by R. G. Ross, Jr.
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J-T cooling with gas mixtures has been universally used in the gas industry, and since the 1970s, has also been used in micro-cryogenic systems. The use of multicomponent mixtures has been widely shown in the works of Brodyansky, Yagodin, Lavrenchenko, Zakharov, Boyarsky, et al.1-4 Initial investigations started with mixtures that contained nitrogen and hydrocarbons, and later, more promise has been shown with freon-containing mixtures. Gromov3 showed that with nitrogen-freon mixtures, the efficiency of a J-T cryocooler with 1.75 W of cooling is 2.5%, as compared to 0.81% with pure nitrogen. Limitations on the use of freons due to the Montreal Protocol, and the low efficiency of freons below 85 K, have contributed to increased interest in nitrogen-hydrocarbon mixtures. The flammability of such mixtures has not significantly limited their use with J-T coolers because of the small amount of gas in J-T systems and the strict leakproof requirements that are applied to cryocooler system designs. The SR&DB has used gas mixtures with J-T cryocoolers since 1978, and conducted design work on systems with capacities of 15 to 100 W at cooling temperatures from 80 to 90 K. In this work, we have developed methods to calculate and experimentally optimize gas mixtures, and have developed a new series of working mixtures to provide cooling at 90, 80, and 73K. CALCULATION AND EXPERIMENTAL OPTIMIZATION OF MULTICOMPONENT WORKING FLUIDS The general scheme of selection of individual refrigerants for a multicomponent working
fluid (MWF) is as follows6,7: • Select a combination of low-temperature and high-temperature refrigerant fluids considering the lowest refrigeration temperature the available precooling temperature and the operational conditions and available compressor equipment • Choose preliminary values for the upper and lower pressures and in the system • Make preliminary choices of intermediate refrigerants that will provide for effective heat exchange of the high pressure and low pressure flows • Select initial component concentrations; if possible, the concentration of components should be determined by a computer software program • Determine the thermodynamic characteristics of the selected mixture composition in the context of the design for the specific cryocooler application • Experimentally confirm the system characteristics over a broad range of environmental conditions. A specific area of emphasis should be examining possible MWF-composition changes during initial cooldown or in the steadystate operating regime • Define composition and concentration parameters for the working mixture; there should be an allowable range for MWF-concentration that will give acceptable parameters for J-T cryocooler operation; one must account for changes associated with the accuracy of the initial mixture preparation and for possible leakage during operation To insure stable cooler performance, the following conditions should be observed when selecting the components and concentration parameters for mixtures: • No solid phase formation at the refrigeration temperature is allowed • No liquid phase formation within the compressor is allowed • Intermediate temperatures should allow for effective heat exchange within the entire length of the counterflow heat exchanger Various programs are useful for computation of the thermodynamics properties of a MWF and for optimization of its composition; these programs involve an equation of state formulated by Redlich-Kwong (as modified by Soave, Wilson, and Barne-King):
where P, V, and T are working pressure, volume, and temperature, R is the universal gas constant, and b and F are parameters that depend on properties of the individual refrigerants, and
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factors of intermolecular interaction. It should be noted that the resulting thermodynamic data from various computer programs can vary up to 30 percent or more. For analyzing nitrogen-hydrocarbon MWFs, the authors prefer a program package by M.J. Boyarsky et al., Moscow Energy Institute (MEI), due to its favorable agreement with our experimental results. However, when using computed MWF data, the results should also be experimentally proven and rechecked. It should be noted that accuracy depends upon the availability and precision in determining interaction coefficients for the individual refrigerant gasses being used. The enthalpy difference should satisfy the following condition for an optimal composition mixture (for direct and counterflow isobaric processes, and over the entire temperature range): at the maximal value of The value of is generally determined by the amount of the highboiling component. The value of is found by the amount of low-boiling component. To provide for good heat exchange at intermediate temperatures, additional fluids are introduced with normal boiling points between temperatures and An important factor to stabilize the actual refrigeration load temperature is the presence of two liquid phases, one of which boils at the refrigeration load temperature. An experimental check of the mixture viability involves the basic components of the J-T cryocooler, such as compressor and heat exchanger units, etc., along with a simulated heat load. TEST APPARATUS AND TECHNIQUES FOR MULTICOMPONENT GAS MIXTURES
A schematic of the SR&DB experimental apparatus used for testing mixed gas J-T systems is shown in Fig. 1. Key features include: • A unique two-stage compressor with piston clearance seals that enable operation without the
use of lubricants on the sliding surfaces; the compressor capacity is The entire cold head of the unit is within a vacuum insulated cryostat Throttle device (J-T valve) with control of shut-off pressure Evaporation-type heat exchanger with multi-sectional heat-load simulator Control of the pressure after the cryo-load heat-exchanger Auxiliary heat exchangers for controlling the temperature of the incoming mixture into the counterflow heat exchanger • Capability to take samples of the circulating gas to provide for analysis of the mixture composition
• • • • •
Figure 1. Schematic of experimental test apparatus for mixed fluid testing. C- compressor; Bbottle with pre-mixed fluid; HE1- counterflow heat exchanger; Th- J-T valve; HE2- water heat exchanger;
HE3- additional heat exchanger; HE4- heat exchanger for load; H1- load heater; H2- additional heater; F1- adsorption filter; F2- mechanical filter; R- flow meter;
pressure gauges.
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The test apparatus allows for a closed J-T cycle with compression of various composition mixtures within a range of 0.05 to 0.3 MPa, up to 2 to 6 MPa, at gas flow rates from 0.7 to
The apparatus is also capable of regulating the precooling gas temperature at the inlet to the counterflow heat exchanger over a temperature range from 223 K to 333 K. RESULTS OF CALCULATED AND EXPERIMENTAL STUDIES OF OPTIMAL MIXTURES Mixture for Operation at 82 K As an example of optimized selection and analysis of MWF characteristics, consider an opti-
mum Nitrogen-Hydrocarbon Mixture (NHCM) to provide maximal cold production at a refrigeration temperature of 80 to 82 K and with a precooling gas temperature of 253 to 323 K. Such a mixture starts with the use of nitrogen as the low-boiling component that will provide for the selected refrigeration temperature. As candidate high-boiling components to
enable a high throttle-effect at various temperatures, there are propane Intermediate components include methane
and ethane
and iso-butane However, phase-
equilibrium studies8 have proved that a liquid mixture with such a composition will disintegrate into two separate fluids at low temperature, wherein one of the liquids will be virtually pure nitrogen with minor amounts of hydrocarbons.
On reviewing the results of a large number of analytical and experimental studies10,11 of five-
and six-component mixture compositions, the most preferable mixture for found to be a four-component mixture, which we refer to as NHCM-1.
has been
Figure 2 shows enthalpy isobars for the NHCM-1 mixture in T-H coordinates. It is clear from the figure that the isobaric curves of 0.08 to 0.12 MPa from 200 to 350K virtually coincide, and thus the isothermal throttle-effect, increases with pressure, and decreases with temperature. It is optimum, therefore, to operate within the vapor-liquid zone, where is of
maximum value. At temperatures it is necessary to add iso-butane to this mixture to enhance the zone of vapor-liquid in the mixture. Isobaric curves of the high-pressure and low-pressure flows are actually parallel in the range of refrigeration temperatures, thus in the zone under consideration, is only weakly affected
by pressure and is completely determined by the quantity and the phase status of the lowboiling component. The gas pressure defines the temperature of the liquid-mixture boiling inside the evaporator, and, hence, the refrigeration temperature as well.
Figure 2. Calculated T-H diagram for mixture NHCM-1 for isobars: 1-0.08 MPa, 2- 0.10 MPa, 3- 0.12 MPa, 4- 3 MPa, 5- 4 MPa, 6- 5 MPa, 7- 6 MPa.
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Figure 3. Calculated values of cooling power for mixture NHCM-1 as a function of for (3 MPa), 2 (4MPa),
Figure 4. Measured specific cooling capacity W* as a function of for varying (3 MPa), 2 (4 MPa), 3 (5 MPa), and
3 (5 MPa), and 4 (6 MPa).
4 (6 MPa).
The calculated cooling power
is determined by the magnitude
but will never exceed
the value of
Figure 3 displays the dependence of calculated cooling power (no loss factor specified) of the NHCM-1 mixture for inlet temperature into the counterflow heat exchanger and for pressures of 4.0, 5.0 and 6.0 MPa at the refrigeration temperature It is observed that, up to temperature the parameter preserves its maximum magnitude of defined by the factor At higher values of a higher inlet temperature can maintain the maximum cooling power. At pressure this value is a maximum up to a temperature
Experimental Measurements of NHCM-1 Mixture The characteristics of the NHCM-1 mixture were experimentally measured on the test apparatus for mixture pressures of 3 to 6 MPa, pressures of 0.08 to 0.12 MPa, and temperatures at the entrance to the counterflow heat exchanger of 262 to 323 K. Using a compressor unit with a flow rate of and an optimal pressure, we obtained a cooling power of 17 to 32 W for and 35 to 65 W for for inlet temperatures, of 323 and 253 K, respectively. The compressor input power was 880 to 1200 W (at Of interest is the fact that at the optimum pressure the maximum cooling power occurred for a of 3.5 to 4.5 MPa as pressure increased from 0.08 to 0.12 MPa. The maintaining of maximum values for cooling power could be explained by the simultaneous effect of two counteracting processes: the increase of mixture cooling power along with the decrease of compressor efficiency at rising pressure and constant pressure Using the results obtained with load W and gas flow rate Q, the authors have computed values of the expected specific cooling power W* (see Fig. 4). As it was predicted by calculations, Fig. 2, the specific cooling power does not depend on gas pressure in the back flow, within allowed experimental-error limits. One small exception is the fact that at the lowest temperature studied, the parameter W* increases slightly and pressure rises, presumably due to a change of the circulating mixture composition from the initial tanked-up proportions. As is evident from Fig. 4, the specific mixture cooling power W* substantially increases with increasing pressure and substantially decreases with increasing precooling temperature This agrees fairly well with the computations. The calculated and experimental values of specific cooling power at similar valves of and agree well with measured values of W*, but always lie below the computed values of This is reasonably well explained by the fact that the predictive calculations did not account for all possible losses. Still, at temperature and at pressure the parameter, W*
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exceeds the predicted value of Such an effect may be caused by a circulatingmixture composition change which could be confirmed by mixture-probe analysis taken during test bed operation. An increase of the nitrogen content within the circulating mixture defines a value of cooling power at the refrigeration temperature, along with the simultaneous decrease of temperature that results in the rise of cooling power at high temperatures. This brings about increased cooling power as compared to the calculated values for the initial tanked-up mixture composition. A computation based on analytically derived estimates of the circulating mixture composition at and has predicted the magnitude of cooling power, as The most important parameters that allow for evaluation of the total J-T cryocooler system are cooling efficiency and Carnot efficiency The cooling efficiency, is computed as the ratio of the useful cooling capacity W to the expended-work, which, in general, is equal to the input power to the compressor, i.e.
The Carnot efficiency, is defined as the ratio of the cooling efficiency, Carnot-cycle cooling efficiency, Thus,
divided by the
Figure 5 shows typical characteristic curves of the specific power and as a function of and at On analyzing the results, one finds that cooling efficiency goes up as the temperature drops, and as increases. The function at the actual values of and is at its maximum. Decreasing the temperature results in lowering the optimal pressure within the flow from 6 MPa to 3.5 MPa as the temperature varies from 323K to 263 K, respectively. The presence of a maximum or minimum energy-consumption can be explained by the above factors and limitations. The measured values of for the NHCM-1 mixture within the investigated pressure and temperature range, amount to 10.5% at and and 7.5% at and The specific energy, increases from 21.7 to 41.7 W/W, respectively, at these same conditions. The experimental results show that to gain the highest J-T cryocooler performance efficiency, it is useful to take advantage of capabilities available for cooling the gas in the direct flow, to operate under elevated back-flow pressures (assuming the calculated values of the -parameters can be maintained), and to operate with optimal gas pressures under direct flow conditions.
Figure 5. Specific energy consumption
various values of
and Carnot efficiency
as a function of pressure
for
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MIXTURE FOR OPERATION AT 70 TO 80 K
NHCM-type mixtures that form two liquid phases when cooled are used as a basic composition for obtaining temperatures in the range of 70 to 80 K. Since the nitrogen-saturated liquid phase contains some amount of methane, its boiling temperature will somewhat exceed the boiling temperature of pure nitrogen. It is possible to decrease the boiling temperature for this mixture either by lowering the back-flow pressure, or by adding some amount of a non-condensable gas such as neon. Considering the above, about 20 neon-nitrogen-hydrocarbon (NNHCM) mixtures were selected for analysis. It was found, analytically, that mixtures with less than 6% neon would not allow the required refrigeration temperature to be attained, and hence, these mixtures were excluded from further study. Three mixtures, denoted as #3, #5 and #15, have been found to be of particular interest, and were experimentally evaluated at pressure to 0.1 MPa, and to 295 K. The #3 mixture has a lesser level of neon for use with a lower back-flow pressure. Figure 6 presents curves that show the typical dependence of and (temperatures after the J-T valve, and at the outlet of the useful-load heat exchanger, respectively) on specific cooling capacity for neon-containing mixture #5. At load values approaching zero, the temperature goes down to values lower than expected. This fact may be explained by the increase of liquidphase (mostly nitrogen-containing) and, hence, by an increase in the neon-concentration within the circulating mixture. This effect disappears at W* = 5 to At higher values of W*, Tx increases sharply at some critical point. Analysis of the experimental results reveals that: • The temperature difference at the entrance and exit of the heat exchanger amounts to 1.5 to 4.5 K as a result of the heat load. • With increase of neon at constant pressure the refrigeration temperature will fall, and this is followed by a decrease of along with an increased dependence of on the load. • As the temperature decreases from 295 K to 263 K, W* increases by about 30 to 40 percent. Furthermore, W* becomes even greater than predicted, which may be due to the nitrogen-saturation effect. The test data also demonstrate that a reduced back-flow pressure reduces the refrigeration temperature. From the viewpoint of system efficiency, however, it is better to increase the neon content rather than to reduce the back pressure. The efficiency of J-T cryocoolers with a range of neon-concentrations from 4 to 12 percent varies little for small heat loads, but at lower neon concentrations, even larger values of and can be obtained.
Figure 6. Experimentally measured temperature of mixture NNHCM-5 after the J-T valve (1) and at the exit of the cryo-load heat exchanger (2) as a function of specific cooling capacity W* for various precooling temperatures (pressure of the mixture in the inverse flow, ).
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Figure 7. An experimental measurement of refrigeration temperature as a function of specific cooling capacity W* for mixtures with various amounts of neon: 1-0.5%, 2-4%, 3-7%, 4- 15% at and 5-15% at
Figure 8. An experimental measurement of the magnitude of cooling capacity W (1) and specific power consumption (2) as a function of refrigeration temperature for a J-T cryocooler using nitrogen-hydrocarbon mixtures with various amounts of neon.
MIXTURES FOR OPERATION BELOW 70 K
Since there are no pure gases with normal boiling temperatures between 77.3 K and 27.1 K (nitrogen and neon, respectively), the researchers have attempted to exploit nitrogen-hydrocarbon mixtures with increased content of neon. A series of computations has been made on the thermodynamic properties of several NHCM mixtures with neon contents up to 18 percent by volume. As a result of these studies, the authors have generated data on: a) the minimum refrigeration temperature, b) temperature boundaries for a heterogeneous liquid phase, and c) the theoretical cooling capacity of the mixtures. Tests have shown that within the investigated concentration range, the minimum temperature has a linear dependence upon the amount of neon, and at 18% neon, almost approaches the solid nitrogen melting temperature. A further increase in neon content does not seem reasonable since this may cause the formation of a solid phase of nitrogen. Calculations for 18% neon mixtures have shown that the predicted cooling efficiency will decrease more than four times, from 22.8 to 5.4 W.h/m3, as the temperature decreases from 80K to 65 K. This shows that, no matter how high the neon content is, the refrigerator will operate above 65 K; the J-T cryocooler specific power may reach 100 W/W. Figure 7 shows experimental test results of cooling efficiencies of mixtures with various amounts of neon. With 0.5 % neon operating at 78 K, it is possible to obtain cooling efficiency, while with 15 % neon operating at 65 K, only cooling efficiency is achieved. For stable operation of a J-T cryocooler at 65 to 70 K, it appears best to decrease the pressure to about 0.08 to 0.07 MPa. In contrast, for more effective usage of the system’s cooling capability, it seems reasonable to divide the net load into 2 or 3 levels, for example, 65K and 70 K. Figure 8 shows curves of J-T cryocooler efficiency for mixtures of nitrogen-hydrocarbons and neon-nitrogen-hydrocarbons at to 85 K, and CONCLUSIONS
1. A series of analytical and experimental multicomponent nitrogen-hydrocarbon mixtures involving such gases as methane, ethane, propane and iso-butane, have shown that it is possible to obtain efficient J-T cooling systems whose thermodynamic characteristics at 80 to 85 K are approaching the best ever achieved for gas-operated cryogenic cooling systems. 2. The addition of neon to a nitrogen-hydrocarbon mixture enables a reduction of temperature to 65 K to 70 K. Due to the reduction of temperature stability within the temperature range, it
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appears reasonable to use the above types of Joule-Thomson cryocoolers in combination with large heat capacitance cold heads or as precooling cascades within two-cascade-type J-T cryocoolers. 3. Upon starting cooldown of mixture-operated J-T cryocoolers, one should account for the fluid mixture composition change as part of the fluid liquefies in various parts of the cooler. 4. The use of gas-mixtures is especially advantageous in combination with state-of-the-art, highefficiency compressors that operate without a need for lubrication. ACKNOWLEDGMENT
The authors are grateful to the U.S. Air Force Phillips Laboratory for support with this work, and for their help in providing information for this report. REFERENCES
1.
Brodyansky, V.M., Groom, E.A., Gresin, A.K., et al., “Effective Throttle Cryogenic Refrigerators Working on Mixtures,” Chemical and Petroleum Mechanical Engineering, n. 12 (1971), pp. 16-18.
2.
Alfeev, V.N., Nikolsky, V.A., Yagodin, V.M., “Throttle Cryogenic Systems on Multicomponent Gas Mixtures,” Electronic Engineering, ser 15. A Cryogenic electronics engineering, Release 1(3) (1971), pp. 95-103. Gromov, E.A., Gresin, A.K., Zakharov, N.D., “Low Temperature Engineering”, in Materials of
3.
Republican Scientific Conference, LTIChP, (1971), 87 p. 4.
Brodyansky, V.M., Gresin, A.K., “A Raise of Effectiveness of Low Temperature Refrigerating
Machines,” Refrigerating Engineering, Release 13, (1973), pp. 1-6. 5.
Dudar, B.G., Yevdokimova, O.V., Lobko, M.P., Mikhailenko, S.A., “A Phase Equilibrium in
Multicomponent Nitrogen-Hydrocarbon Mixtures and H(T)-diagram in Three-phase Region,” Preprint, Ac. Sci UkrSSR, ILTPh&E, Kharkov, 28-88, (1988), 28 p. 6.
7.
8.
9.
Arkhipov, V.T., Vihodtsev, I.S., Dudar, B.G., Yevdokimova, O.V., Lobko, M.P., Mikhailenko, S.A., Yakuba, V.V., “On the Problem of the Influence of the Nitrogen-Hydrocarbon Mixtures Composition on Performance Parameters of Compression-Throttle Cycles,” Preprint Ac. Sci. UkrSSR, ILTPh&E, Kharkov, 36-90, (1988), 32 p. Arkhipov, V.T., Vihodtsev, I.S., Dudar, B.G., Yevdokimova, O.V., Lobko, M.P., Mikhailenko, S.A., Yakuba. V.V., ‘About the Influence of Nitrogen-Hydrocarbon Mixtures Composition on Magnitude of a Joule Thomson Effect and Performance of Compression-Throttle Cycles,” High-temperature Superconductivity, VNIIMI, release.3-4, (1990), pp. 17-26. Grezin, A.K., Gromov, E.A., Zakharov, N.D., “Forming and Optimization of a Structure of Cooling Agents for Throttle Cryogenic System,” Chemical and Petroleum Mechanical Engineering, n.9, (1975), pp. 7-8. Boyarsky, M.Ju., Hodgaev, D.N., Mogorichny, V.N., et al., “Problems of the Definition of Thermodynamic Properties of Multicomponent Cryogen’s with Heterogeneous Liquid Phase,” Book of Reports of 3rd All-Union Scientific and Technical Conference on Cryogenic Engineering, Cryogenmash, Moskov (1983), part 2, pp. 192-206.
An Experimental Study and Numerical Simulation of Two-Phase Flow of Cryogenic Fluids through Micro-Channel Heat Exchanger W.W. Yuen and I.C. Hsu University of California, Santa Barbara Santa Barbara, CA 93106 and Lockheed Martin Missiles & Space Palo Alto, CA 94304-1191
ABSTRACT
The design, fabrication, testing and analysis of a micro-channel heat exchanger, a key com-
ponent for a microminiature Joule-Thomson Cryogenic Refrigerator, is described. Results show that the heat exchanger can be fabricated efficiently and economically with the existing manufacturing technology. The heat exchanger was tested and found to be mechanically robust and durable under high pressure operating conditions. “Choking” did not occur in the micro-channels of the heat exchanger. The thermal performance is excellent, providing the rapid cooling as designed. A numerical code is developed both to interpret the data and to provide some sensitivity study and assessment on the performance characteristics of the heat exchanger. Furthermore,
this code will be utilized as a design tool for optimizing the performance of next generation planar heat exchangers. INTRODUCTION
A microminiature Joule-Thomson cryogenic refrigerator is a highly compact and efficient refrigerator which has a wide range of potential applications in the cooling of laboratory apparatus and low-noise electronic devices under cryogenic conditions. The objective of this work is to develop the fundamental understanding of the thermal and mechanical performance characteristics of a planar micro-channel heat exchanger, a key component of the refrigerator. In this paper, results of the first-year experimental and analytical studies are presented. EXPERIMENTS
Design and Fabrication of the Heat Exchanger While the need to operate in microminiature size is essential for the success of the refrigerator [1], the exact size of the heat exchanger and the geometric dimensions of the micro-channels (depth and width) depend on the available fabrication technology. Three micro-channel heat Cryocoolers 10, edited by R. G. Ross, Jr. Kluwer Academic/Plenum Publishers, 1999
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Figure 1. Schematic of the silicon layer of the micro-channel heat exchanger.
exchangers with similar designs were fabricated by the Center for Microelectronics and Optoelectronics of Lawrence Livermore National Laboratory. The dimensions for each heat exchanger unit are 2.54 cm by 3.04 cm and 0.4 cm thick. Each unit is comprised of two halves made of silicon and Corning 7740 borosilicate glass, whose coefficient of thermal expansion closely matches that of silicon. The silicon layer is 1.016 mm thick with multiple micro-channels parallel to the long dimension of the heat exchanger. The first heat exchanger unit has an aspect ratio, between the depth of the micro-channel to its width, of 1.6:1 but the second and third unit
have been fabricated with an aspect ratio of 3.4:1 in order to increase the surface area for heat transfer. All of the micro-channels are cut into the silicon layer using a computer controlled dicing saw. The dimensions for the micro-channels in the silicon layer are deep and either or wide, depending on the aspect ratio for the particular unit. All the wall thicknesses for the micro-channels are thick. The first heat exchanger unit, with an aspect ratio of 1.6:1, has 36 parallel channels per silicon layer and the other two heat exchangers, with an aspect ratio of 3.4:1, has 65 channels per layer. A schematic of the silicon layer with the micro-channel is shown in Figure 1.
The borosilicate glass layer is 3.05 mm thick with two flow manifolds of 3.81 mm wide cut into it, parallel to the short dimension of the heat exchanger. The depth of each manifold is 2.03 mm with two circular holes of 3.30 mm diameter bored through the glass layer acting as inlet and
exit for the coolant flow. A schematic of the glass layer is shown in Figure 2.
Figure 2. Schematic of the borosilicate glass layer.
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Figure 3. A close-up photograph of the micro-channel heat exchanger.
Finally, the finned silicon layer is contact bonded to the glass flow manifold layer to form the complete planar heat exchanger. A photograph of the micro-channel heat exchanger is shown in Figure 3.
Preliminary Test Results
To test the mechanical and thermal performance characteristics of the heat exchanger, a series of blow-down experiments have been performed. The heat exchangers are instrumented so that cooling can be achieved and monitored. To enable the high-pressure coolant to flow through the heat exchanger, a stainless steel flat washer was brazed onto one end of a 3.175 mm-OD stainless steel tubing with a diameter hole drilled through this washer to act as a JouleThomson expansion orifice. This tubing is then epoxy bonded to the inlet manifold of the planar heat exchanger and supports the heat exchanger inside the test chamber. A rectangular focal plane substrate made of alumina is epoxy bonded to the silicon side of the planar heat exchanger to represent the approximate thermal mass of a planar infrared focal plane assembly. Then a platinum resistance temperature sensor is epoxy mounted onto this alumina substrate and utilized to measure its temperature during Joule-Thomson expansion of coolant flowing through the planar heat exchanger. A schematic of the experimental setup is shown in Figure 4. A photograph of the “instrumented” micro-channel heat exchanger is shown in Figure 5.
Figure 4. Schematic of the blow-down experiment for the micro-channel heat exchanger.
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Figure 5. Close up photograph of the “instrumented” micro-channel heat exchanger.
Blowdown experiments are performed using krypton compressed to high pressure (170 to 449 bar) with two heat exchangers of different channel width. For a 100 cc volume krypton tank and initial substrate temperature of 288 K (15°C), the minimum temperature achieved for the two heat exchangers with different channel widths are shown in Table 1. Typical temperature transient data for the channel-width heat exchanger (at 272 bar) and the channelwidth heat exchanger (at 449 bar) are shown in Figures 6. To better compare the thermal performance of the heat exchangers, the flow rate of the coolant through the Joule-Thomson expansion nozzle has to be set to a standard flow rate. Thus a second planar heat exchanger with channel-width has been retrofitted with a capillary tubing instead of a Joule-Thomson expansion orifice in order to facilitate the adjustment of flow rate. A short length of fine stainless steel hypodermic tubing with an internal diameter of was brazed into the 3.175 mm-OD coolant supply line and utilized as a restriction to the flow. The flow rate through this restriction was adjusted to 4.5 standard liters per minute (slpm) of nitrogen by trimming the length of the hypodermic tubing. The nitrogen gas was supplied from a gas bottle with its outlet pressure regulated to 69 bar (1000 psig). Due to a factor of 3.5 reduction in the diameter of the flow restriction, compared with the first heat exchanger unit, the
Figure 6. Typical temperature transient data for the two micro-channel heat exchangers.
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flow rate of krypton through the heat exchanger has been proportionally reduced. This accounts
for the slower cooling rate observed for the channel-width heat exchanger shown in Figure 6. In all cases, “choking” did not occur in the micro-channels and the heat exchanger remained mechanically stable for all tests. It is interesting to note that both heat exchangers yield approximately the same minimum temperature at the high pressure limit. Additional tests are currently under consideration to characterize more completely the thermal and mechanical performance characteristics of the heat exchanger. ANALYSIS
The focus of the analytical effort is to develop a computational capability to perform two phase thermal and flow analysis in arbitrary three-dimensional geometry. The computer code will be used first to analyze performance data of the micro-channel heat exchanger. It will also be the basic design tool used to determine optimal design parameters (e.g. gas mixture ratio, geometric dimensions, etc.) for the microminiature Joule-Thomson cryogenic refrigerator. The code is based on a multifield Eulerian treatment [2]. To facilitate the simulation of flows in complex geometry, it assumes three continuous fields (solid, liquid and gas). Complex flow geometry, such as those in a micro-channel heat exchanger, is simulated by specifying a solid fraction distribution in a 3-d flow field. Heat transfer and friction loss boundary conditions can also be simulated with the specification of appropriate solid/liquid and solid/gas interfacial drag and heat transfer constitutive relations. Because of the three-fluid formulation, the code can also be used to simulate the flow of solid particles in liquid/gas two phase mixture. The mathematical formulation is similar to an existing code developed for the analysis of steam explosion [2]. The detail is given in the same reference and will not be repeated here. Due to the relatively large uncertainty in flow conditions associated with the current set of preliminary experiments (for example, the friction loss associated with the various inlet and outlet restriction/orifices are not known), the initial effort of the numerical simulation will focus mainly on the qualitative performance characteristics of the micro-channel heat exchanger. These results are useful to demonstrate the effectiveness of the micro-channel concept and to provide improvements for future designs.
Numerical Parameters of the Simulation A schematic of the cross-section vertical plane including the inlet from the high pressure tank is shown in Figure 7. Note that this vertical section is closed to the ambient because the venting orifice is located in a different plane. The flow volume in the micro-channel heat exchanger is modeled by a 28 mm by 22 mm by 3 mm computational domain with a grid size of
Figure 7. Schematics of the vertical plane including the inlet from the high pressure tank.
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The silicon layer is modeled as a two-dimensional plane layer within which the fluid is restricted to flow in the y-direction only (the “long (28 mm)” side of the layer). The thickness of the layer is one computational cell, corresponding approximately to the actual depth of the microchannel
The flow volume in the channel is preserved by specifying a solid
fraction of 0.167 in the layer. The heat transfer across channels is calculated based on fullydeveloped pipe flow correlations [3]. The two manifolds are modeled as 4 mm wide, 22 mm long and 2 mm deep rectangular volume located on top of the silicon layer. The time step is adjusted automatically by the code to maintain numerical stability. For the results presented in this work
(with an inlet pressure of 300 bar), the average time step is
sec.
It is important to note that a direct simulation of all the geometric characteristics of the
micro-channel heat exchanger (e.g. taking a grid size of 0.5 mm) is computationally intensive and is probably unnecessary for the current objective of illustrating the qualitative behavior of the heat exchanger. More detailed geometric simulation, together with better characterization of the inlet and outlet friction loss, will be the objective of future calculations. RESULTS AND DISCUSSION Numerical results generated for the blowdown of krypton at a tank pressure of 300 bar is
presented. The thermodynamic properties are generated by the program GASPAK [4]. To demonstrate the condensation and flow behavior, the transient void fraction distribution at four different times at the silicon layer is shown in Figure 8. The corresponding distribution in the vertical plane of the inlet orifice in the direction of the microchannel is shown in Figure 9 and the
distribution in the vertical plane in the direction of the manifold is shown in Figure 10. The average void fraction in the silicon layer is shown in Figure 11. The void fraction distribution shows clearly that the micro-channel is effective in generating quickly a uniform liquid zone across the full heat transfer surface. The design objective of generating a uniform temperature surface (for example, in infrared detector application) is thus met. The condensation occurs almost instantaneously in the micro-channels near the inlet orifice as demonstrated in Figure 9, and is much less efficient the direction on the manifold, as illustrated in Figure 10. This difference in condensation effect can be attributed to the incompatibility between the inlet
Figure 8. Void fraction distribution and liquid volumetric flux at the bottom silicon layer.
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Figure 9. Void fraction distribution and liquid volumetric velocity at the vertical plane of the inlet orifice in the direction of the micro-channel.
Figure 10. Void fraction distribution and liquid volumetric velocity at the vertical plane of the inlet orifice in the direction of the manifold.
flow from a single orifice and the flow direction restricted by the micro-channels. At the channels near the inlet orifice, the stagnation pressure built up locally forces the flow into the micro-channel naturally. The fluid which is directed in the direction of the manifold, on the other hand, must make two 90-degree turns (one downward toward the silicon layer and then sideways into the microchannel) to enter the micro-channel. This requires significant pressure buildup and thus delays the cooling and condensation of the middle section. This result suggests that the current design can be improved by either eliminating the micro-channel at the area directly under the two manifolds, or
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Figure 11. Average void fraction at the silicon layer.
providing a distributed inlet flow across the inlet manifold. The condensation and cooling can then be achieved more quickly and efficiently. In general, the numerical prediction of temperature and pressure distributions is consistent
with experimental data. Discussion of the numerical data, however, is meaningful only if they are presented with a full parametric study on the effect of frictional loss along the orifice. These numerical data are currently being generated and they will be presented in future publications.
CONCLUSIONS
The important basic component of a microminiature Joule-Thomson cryogenic refrigerator, a micro-channel heat exchanger, has been designed and fabricated. Preliminary experiments show that the heat exchanger is mechanically robust and has excellent thermal performance characteristics. A computer code has been developed both to interpret the performance characteristics of the heat exchanger and to serve as a design tool for the refrigerator. Results of numerical simulation show that the micro-channel heat exchanger is effective in generating a uniform liquid fraction over the whole base surface. Analysis of the predicted flow dynamics suggests a modification in design to improve the condensation efficiency. ACKNOWLEDGMENT
This work is supported by Lockheed Martin and the U.C. MICRO program. REFERENCES
1. 2. 3. 4.
Little, W.A. , Proceeding of the NBS Cryocooler Conference, Edited by J.E. Zimmeman and T.M. Flynn, NBS Spec. Publ. No. 508, p. 75, April 1978. Yuen, W.W. and Theofanous, T.G., PM-ALPHA: A Computer Code for Assessing the Premixing in Steam Explosion, DOE/ID-10502, April 1995. Incropera, F. P. and DeWitt, D. P., Fundamentals of Heat and Mass Transfer, 4th Ed., John Wiley and Son, Inc., 1996. McCarthy, R. D. and Arp, V., User’s Guide to GASPAK, Version 3.1, CRYODATA, P.O. Box 558, Niwot, CO, 80544, 1992.
Hybrid 10 K Cryocooler for Space Applications R. Levenduski, W. Gully, and J. Lester Ball Aerospace & Technologies Corp. Boulder, Colorado, USA 80306
ABSTRACT Ball Aerospace is combining its flight proven Joule-Thomson (J-T) and Stirling cryocooler technologies to create an efficient, low-mass, and reliable 10 K cryocooler. The hybrid machine consists of an enhanced J-T cooling stage coupled to a split Stirling precooler. The J-T loop incorporates a low-pressure compressor to provide cooling at the sensor interface and a thermal capacitor to provide peak load cooling capacity. The J-T loop is precooled by a modified version of a previously developed three-stage Stirling cryocooler. This paper describes the concept, shows system-level estimates for mass and power, and presents related test data. INTRODUCTION Future space systems may require cooling at 10 K and below. This has sparked interest in developing an efficient, long-life 10 K cryocooler for space applications. Several technologies are available that could potentially provide cooling at 10 K. These include Joule-Thomson (J-T), reverse-Brayton, and Stirling-cycle technologies. The Air Force Research Laboratory (AFRL) is actively developing 10 K cryocoolers and has recently awarded a contract to Ball to develop a 10 K cryocooler using a combination of J-T and Stirling technologies with a thermal storage unit (TSU). This patented combination of technologies creates an efficient 10 K cryocooler with variable load capability that is easily integrated to a sensor. 10 K CRYOCOOLER DESCRIPTION System Description The Ball 10 K Cryocooler is a hybrid machine consisting of an enhanced J-T cooling stage coupled to a split Stirling precooler via a TSU. The J-T loop consists of a single-stage compressor with drive electronics and a cold head. The cold head houses heat exchangers, a J-T valve, the TSU, and a simulated 10 K sensor. The Stirling precooler consists of opposed linear drive compressors, a three-stage displacer, and an electronics module. The TSU for this device is a simple lead mass of approximately 200 g. Figure 1 shows a CAD rendering of the 10 K Cryocooler. Operation The hybrid approach optimally combines Stirling and J-T cycles to circumvent their limitations and take full advantage of their capabilities. The cryocooler is designed so most of the
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Figure 1. CAD rendering of 10 K Cryocooler.
cooling is done by the Stirling precooler. The Stirling cycle is more simple and efficient than the J-T cycle in providing low-temperature, steady state cooling. However, as the cooling temperature approaches 10 K, the heat capacity of most materials drops too low to provide the efficient regeneration needed for the Stirling cycle. Fortunately, the relative efficiency of a helium J-T cycle improves as the temperature approaches 10 K so there is a natural synergism in combining the two cycles. This combination has long been known to be an effective means of achieving cryogenic refrigeration at and below 10 K.
A system trade study determined the most power efficient precooling temperature for this concept to be 15 K. This temperature allows the Stirling-cycle cryocooler to operate in a region where lead still has significant heat capacity for regeneration and allows the J-T cryocooler to operate at relatively low absolute pressure and pressure ratio. This approach allows an existing Stirling cryocooler with a modified regenerator cold stage and an efficient compressor with
demonstrated long-life features to be used. A three-stage Stirling cryocooler is used as the precooler. The cold stage provides the primary precooling at 15 K for the J-T loop. The middle stage provides cooling at 40 K to also precool the J-T flow and to cool a thermal shroud within the cold head vacuum shell to minimize parasitic heat loads. The warm stage operates at 170 K and minimizes internal parasitic heat loads. The J-T compressor circulates low-pressure helium through three small counterflow heat exchangers and two precooler heat exchangers in the cold head. The gas gets progressively colder
as it passes through each heat exchanger until it reaches the J-T valve. The J-T valve (capillary tube) expands the helium to lower pressure and drops the gas temperature the last few degrees to achieve 10 K at the sensor interface. No liquid helium is produced since the low-side pressure remains above the critical pressure. The cold gas absorbs the sensor heat load and cools the incoming gas as it returns through the counterflow heat exchangers on its way back to the ambient temperature compressor. The system provides constant temperature cooling with a varying sensor heat load by using a temperature control algorithm that links the compressor speed to the sensor temperature. If the
sensor temperature rises, the compressor speed will increase. The speed increase will increase the high-side pressure, causing more flow, and will decrease the low-side pressure, causing a decrease
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in gas temperature. These effects combine to hold the sensor temperature constant. The TSU aids the process by cooling the increased gas flow to the proper temperature before it enters the J-T valve. Although the TSU slowly warms, it is sized to provide enough cooling to minimize the required dynamic range of the compressor. When the sensor load decreases, the compressor
speed and gas flow decrease, which allows the Stirling precooler to return the TSU to its nominal operating temperature. The system is then ready for another high-load cycle. Features and Benefits The hybrid arrangement has many benefits. The Stirling precooler has multistage capability that can be used to cool optics and thermal shields to minimize parasitic heat loads to the cold stage. This inherent capability becomes critically important when considering cryocooler redundancy. The J-T cold head, consisting of small, coiled tubes, has almost unlimited configuration flexibility so it can meet virtually any integration requirement. The small tubes also mechanically and thermally isolate the sensor from the surroundings, which minimizes vibration and parasitic heat loads. The Stirling precooler and J-T compressor can be located up to a meter
away from the sensor to minimize EMI at the sensor and ease heat rejection. The system can be easily adapted to different temperatures and heat loads without losing heritage. The J-T loop can operate at different temperatures by changing operating pressure and
working fluid. For example, the J-T loop was tested with nitrogen at 2.5 psia to produce cooling at 65 K. The compressor has also been tested with helium, which will be used to provide cooling at 10 K. Other gases are available that enable operation over virtually the entire cryogenic range.
The TSU enables the J-T loop to provide variable cooling capacity at constant temperature while the Stirling precooler, which consumes most of the power and comprises most of the weight, operates at the average load. This results in maximum system efficiency and minimum system weight.
Performance Projections The projected cooling performance of the 10 K Cryocooler is shown in Figure 2 at 87 W of input power. The precooler is expected to provide the required precooling with the compressor
operating at 80% stroke. The J-T compressor will operate with a compression ratio of nearly two and produce the flow necessary to achieve 10 K. The performance curve at 183 W of input power shows the maximum performance that can be achieved with the existing precooler
operating at full stroke. A greater capacity J-T compressor (larger displacement or faster speed) would be required to achieve this performance. A sensitivity analysis showed there is a relatively large “sweet spot” for precooling temperature that gives good overall efficiency. The results are shown in Figure 3. This sweet spot adds confidence that the required heat lift can be achieved for the desired input power. Table 1 shows the input power breakdown. The J-T cold head consumes no power. Table 1 also shows the cryocooler mass breakdown. The J-T compressor and cold head comprise 14% of the total weight. SUBSYSTEM DESCRIPTIONS
Joule-Thomson Loop
A new compressor is needed for the J-T loop. To make the system power efficient for 10 K cooling, a compression ratio of about two to one is needed with a low-side pressure of 60 psia (414kPa). The compressor must have low-vibration output, be light weight, be efficient over a speed range, and be reliable over a 10-year operating life. Early development projects at Ball led to the selection of a commercial compressor concept. Tests were conducted on modified commercial compressors to determine if the operating life of these machines could be extended. Figure 4 shows wear rate data for the modified commercial units. After undergoing special processing, the machines were operated with dry helium for
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Figure 2. Cooling performance of 10 K Cryocooler.
Figure 3. Efficiency of 10 K Cryocooler at 100 mW of cooling.
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Figure 4. Wear rate data for modified commercial compressors.
periods ranging from hours to days between wear rate measurements. At first, the wear rates were relatively high, but as the components became polished the wear rates fell to a very low level. The upper curve is for the factory stator surface finish. The lower two curves show wear rate reduction as a result of fine honing of the stator bore. Wear continued to decrease up to the end of the tests. The data indicates a potential life of over 17 years. These results give confidence that a compressor with at least 10 years operating life with dry helium can be developed. A compressor meeting the specific requirements of the 10 K Cryocooler is currently under development. This task involves converting the commercial device to an aerospace quality machine that includes hermetic seals, long-life motor, and efficiency-enhancing features. These compressors will be relatively inexpensive even with the added aerospace features. This creates the possibility that a full reliability evaluation, involving many units, could be carried out. An efficient cold head for the specific requirements of the 10 K Cryocooler is also being developed. The Stirling precooler cold stage will attach just upstream of the final heat exchanger at 15 K to provide precooling at the most power-efficient temperature for the system. An intermediate heat exchanger in the cold head will attach at the 40 K precooler stage for the purposes of reducing the thermal load on the 15 K stage and absorbing parasitic heat leak. The cold head contains no moving parts. Included within the vacuum shell are the heat exchangers, J-T valve, thermal storage unit, and insulation. All of these components are well developed at Ball and are now reduced to engineering tasks. We have built small tube-to-tube heat exchangers for helium with effectiveness near 99%. J-T valves are highly developed as a result of the Cryogenic On-Orbit Long-Life Active Refrigerator (COOLLAR) program. These valves are based on capillary flow and incorporate defrost heaters should it ever be necessary to eliminate frozen gases. An independent research and development project has proven the concept of the built-in thermal storage unit Gas purification techniques were also developed on the COOLLAR program. Multilayer insulation has been used for years at Ball for cryogenic systems. Stirling Precooler A Stirling-cycle mechanical cryocooler will be used as a precooler for this hybrid system. The Stirling will be required to absorb the heat load consisting of the enthalpy imbalance in the flows. This imbalance is caused by the J-T heat exchangers’ inefficiencies, which result from the parasitic radiation loads. The system is designed to use the Stirling cryocooler to cool into a range where only a single-stage J-T compressor is required to make the final step to 10 K. A threestage brassboard Stirling cryocooler developed for the NASA 30 K program will be used.1 A three-stage Stirling cryocooler for space systems requiring refrigeration at two separate temperatures was developed for AFRL and NASA. It employs a standard linear compressor and
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Figure 5. Three-stage Stirling brassboard displacer.
a special displacer with two copper interface attachment points for connection to the external
loads. Each interface plate has an integral heat exchanger that intercepts the flow between stages of the annular regenerator. This is an advantage of the fixed regenerator design. A mature version of this cooler, described in another paper at this conference,2 meets typical flight requirements. Brassboard development hardware will be used for this effort to perfect the thermal operation of the design. The brassboard is shown in Figure 5. The brassboard has demountable cold stages, which simplifies the task of modifying the cooler for different operating conditions. The dualtemperature cooler was designed to operate at 60 K and 35 K, but the unit must operate at a much lower temperature to support the J-T work. The brassboard will be re-optimized to operate at 15 K. The focus will be on adapting the final stage of the regenerator to lowtemperature operation. As is well known, the regenerator matrix heat capacity drops precipitously between 35 K and 15 K, and the fluid density more than doubles. To keep the regenerator
losses in check, the regenerator will be converted from bronze screens to lead spheres in the last stage. The performance of the modified cooler will be explored over a range of interest, and a suitable operating point will be selected for the final system. Thermal Storage Unit
A thermal storage unit is provided to allow the system to absorb thermal loads higher than the average load for short periods. This TSU is installed and operated differently from the conventional method. Typically, a thermal storage unit is mounted directly on the load to be cooled. When the load is lower than the average, the material in the TSU freezes at constant temperature and stores the excess cooling. When the load is higher than the average, the material in the TSU melts at constant temperature and releases the excess cooling as needed to hold the
desired temperature. The TSU in the 10 K Cryocooler is arranged differently for two reasons. First, when cooling to 10 K, there is no TSU material that will undergo a phase change to store and release heat at constant temperature. Second, there is no practical material that can be used as a sensible heat capacitor that could perform the load-leveling feature for a small temperature change. Heat capacities are so small at 10 K that enormous quantities of material would be needed to limit the load temperature change to a small value. For the 10 K Cryocooler, the TSU is mounted on the warm side of the final cooling stage of the J-T loop at 15 K and not at the 10 K load. In this system, the J-T loop does the job of controlling the load temperature and the TSU does not. The mass flow of the J-T loop is increased to hold the load at 10 K during a high heat load period. The TSU maintains the
temperature of the helium in the loop nearly constant to make up for the shortfall of precooler capacity. This method of operation allows the use of TSU materials that do not change phase,
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and allows a moderate rise in TSU temperature. The TSU operating temperature is nominally 15 K where thermal capacities of TSU materials are much greater than they are at 10 K, so only small quantities of TSU material are needed. Approximately 200 g of lead will be used in the
first-generation 10 K Cryocooler. When the high heat load periods become longer, the TSU material can be changed from lead to helium that is contained in a small bottle at the cold stage of the Stirling precooler. Helium is more effective than lead as a TSU material because it has a much greater heat capacity. This approach is preferable for large systems. 10 K CRYOCOOLER PROGRAM OVERVIEW
Program Description The AFRL 10 K Cryocooler program will develop an Engineering Design Model (EDM) cryocooler that produces 100 mW of continuous cooling at 10 K. The requirements and goals are shown in Table 2. The program will span 33 months. The Stirling regenerator and J-T compressor will be developed during the first year of the program. The J-T cold head and electronics will be developed during the second and third years. Acceptance testing and delivery
will occur in late 2000. The cryocooler is tentatively planned to undergo thermal vacuum testing at AFRL upon delivery.
CONCLUSION
Ball Aerospace is developing a 10 K cryocooler that produces 100 mW of steady state cooling. Variable load capability at constant temperature will also be demonstrated. The hybrid cryocooler that combines Stirling and Joule-Thomson technologies with a thermal storage unit will be delivered to the Air Force Research Laboratory in December 2000. REFERENCES 1.
Gully, W.J., H. Carrington, W.K. Kiehl, and K. Byrne, “A Mechanical Cooler for DualTemperature Applications,” in Proc. of STAIF-98, Am. Inst.Phys., CP420 (1998), pp. 205-210.
2. Gully, W.J., H. Carrington, and W.K. Kiehl, “Qualification Test Results for a Dual-Temperature Stirling Cryocooler,” to be published in Cryocoolers 10, Plenum Press, New York (1999).
Design and Development of a 4K Mechanical Cooler S.R. Scull1, B.G. Jones1, T.W. Bradshaw2 A.H. Orlowksa2, and C.I. Jewell3 1
Matra Marconi Space, Filton, Bristol, England Rutherford Appleton Laboratory, Chilton, Oxfordshire, England
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3
ESA - ESTEC, Noordwijk, The Netherlands
ABSTRACT
This paper describes the design and development of a closed cycle Joule-Thomson 4K Cooler aimed at the ESA Far Infra-Red Space Telescope (FIRST) and PLANCK astronomy missions. A cooler of this type has previously been developed by the Rutherford Appleton Laboratory (RAL) under ESA contract, but the criticality of the cryogenic cooling sub-system to the mission is such that ESA identified the need to prove and qualify the cooler against project specific requirements before the start of the spacecraft development. The environmental and performance requirements (>9mW heat lift at 4.2K.) of the cooler are challenging and the FIRST 4K Cooler includes many features to enhance performance over previous designs. In making these changes the heritage gained on other MMS coolers has been retained and the cooler has been designed with a high degree of modularity with respect to MMS Stirling coolers that are in batch manufacture. One key area of the design has been the ruggedisation of the Joule Thomson stage to meet
the Ariane 5 launch environment. This ruggedisation, which has to meet the conflicting thermal and structural requirements, uses pre-tensioned Kevlar threads and snubbers to achieve launch survivability. INTRODUCTION
In December 1994 ESA initiated a programme of work at MMS to qualify the critical cryocooler technology required for the FIRST/PLANCK missions. The programme initially looked at a 20K two-stage Stirling cycle cooler1 for use both as a stand-alone cooler and as a pre-cooler for a 4K Cooler. The design, development and qualification for the 4K, closed cycle, Joule-Thomson cooler described in this paper was initiated in September 1995. The design of the cooler is now complete and the manufacture of the two qualification models required under the contract is well advanced. Presented within this paper are the key requirements and design drivers, together with an overall description of the cooler with particular reference to the ruggedisation of the J-T Assembly to withstand launch vibration.
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KEY REQUIREMENTS
The key requirements of the FIRST 4K Cooler are given in Table 1. It should be noted that these are operating values and that the usual factors are taken for qualification and design limits. In addition to the performance requirements given, certain configuration constraints were also specified. In particular, the cooler must be capable of operating with a 2m separation between the J-T compressors and the Pre-cooler/Cold Stages in order to give greater flexibility in the spacecraft installation. Also, the cooler must be capable of tolerating ±7.5mm displacements, in all 3-axes, between the cooler and the instrument.
OVERALL DESIGN DESCRIPTION
The overall layout of the qualification model 4K Cooler is shown in Figure 1. It comprises the following major items: • 20K, 2-stage, Stirling Pre-Cooler • J-T Heat Exchanger Assembly
• Two J-T Compressors • Flight Support Structure • Ancillary Items e.g. Control Valve, Gas Purifier and Filters The Pre-Cooler specified for this programme is that being developed for the FIRST
programme1. It comprises two compressors and a two-stage displacer (with integral momentum balancer). The item is being qualified complete with flight support structure and instrumentation. In order to meet the stringent exported vibration requirements, the cooler is controlled by a Low Vibration Drive Electronics which monitors the cooler out of balance via force transducers and adjusts the cooler drive current to compensate.
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Figure 1. FIRST 4K Cooler - Qualification Model.
The J-T heat exchanger assembly (see Figure 2) is mounted on the Pre-Cooler cold finger.
The heat exchanger together with the J-T Compressors and ancillaries panel form the closed-loop J-T Cooler.
The pressurised helium from the high pressure compressor passes, via a gas purifier mounted on the Ancillaries Panel, through heat exchangers that are cooled by the Pre-Cooler to nominally 150K and 20K before expanding at the J-T orifice.
The heat exchangers at 150K and 20K are attached to the Pre-Cooler and comprise, porous, sintered, stainless matrices. Heat exchangers between these stages and between the 20K and 4K stages comprise 1m coiled pipes in which the high pressure gas passes through a small bore tube within a larger bore tube through which the cold return gas flows. In addition to the J-T circuit, a “by-pass” route is incorporated in the design to increase the gas flow at higher temperatures and thus speed up the initial cool down. The design of the 20K to 4K coiled pipe heat exchanger is of particular interest as it acts as a flexible link between the cooler and the instrument; the 4K stage of the cooler being attached directly to the instrument cold bus bar. Also, much care has been taken over the launch support system for the J-T Assembly. These features are described in more detail later in this paper. The two (high and low pressure) J-T Compressor units are mounted as a balanced pair in a flight support structure. The design of the J-T Compressors is virtually identical to that of the Stirling cycle units. The exceptions being the size of the piston in the HP Compressor (20mm) and the inclusion of, non-return, reed valves to create the “d.c.” pressure across the J-T orifice. Other items within the J-T Cooler system are mounted on the “Ancillaries Panel” and include: – Solenoid Operated Valve (used to select the by-pass circuit during initial cool down
– 2-Stage Gas Purifier – Particulate Filters – Buffer Volume (used to balance the pressure between the J-T Compressors)
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Figure 2 4K Cooler J-T Assembly.
The Gas Purifier is heated to 400°C during ground testing in order to remove any gaseous impurities in the system. Since the J-T system is a closed loop it is very unlikely that the purifier needs to be heated in-orbit although it is thought prudent to include heaters at this time in case of blockage of the J-T orifice by condensables. COOLER PERFORMANCE
Improvements have been made to the performance of the 4K Cooler from previous developments by both improving the pre-cooler performance and the efficiency of the J-T heat exchangers. Pre-Cooler Performance The temperature achieved by the pre-cooler has a direct bearing on the heat lift capability at 4.2K. The 20K pre-cooler performance has been significantly improved by optimisation of the cold finger configuration and regenerator materials. Initial improvements have already been reported3 and the performance achieved has been further improved by the introduction of rare earth materials in a form that has a high surface area, acceptable pressure drop and is robust enough for high reliability space flight application. The performance of the cooler has been evaluated in development model testing at RAL and MMS and the performance achieved with a cold finger in flight configuration is shown in Figure 3.
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J-T Cooler Performance Predictions Increasing the length of the coiled pipe heat exchangers from 750mm to 1m has improved efficiency and reduced the heat load on pre-cooler. Using correlated RAL modelling data4 a worst case analysis has been performed on the cooler performance that shows that with a precooler lifting 120mW of heat at 20K the ESA requirements can be met with an input power of 140W. Mass (Includes Flight Structure) Pre-Cooler 20.9kg J-T Cooler 16.4kg Ancillaries Panel 5.8kg Total 43.1kg
SUPPORT OF J-T HEAT EXCHANGERS FOR LAUNCH SURVIVAL
A key area of the design of the FIRST 4K Cooler is the support of the J-T Heat Exchangers to meet the launch vibration case. The J-T Assembly is very delicate, with the coiled pipe heat exchangers in particular having a low natural frequency and high Q. Also, the support system needs to limit loads imparted on the Pre-Cooler cold finger to acceptable levels. In order to meet the conflicting thermal and strength requirements a support system has been devised using a combination of non-contacting snubbers (bump stops) and Kevlar thread, see Figure 2. The primary support structure comprises a machined titanium tube that is gold plated to reduce radiative coupling to the cold stages. From this primary structure the individual areas of the J-T Assembly are supported either by snubbers or threads of Kevlar 29. Each stage of the
support system is described below. The 300K to 150K Heat Exchanger is supported by a stainless steel, threaded, ring that is wound into the coiled pipe and bolted to the base-plate. The thread is machined away such that only four thin, threaded, pillars remain. This has the double benefit of reducing the contact area on the pipe and thus the parasitic heat load and also reducing the mass. The 150K Heat Exchanger stage is bolted to the aluminium ring of the Pre-Cooler cold finger launch support tube. The launch support tube comprises a thin wall (0.7mm) GFRP tube
Figure 3 Development Model Pre-Cooler Performance.
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with aluminium fittings. The top ring of the launch support tube is connected to the mid-stage of the Pre-Cooler by a thermally conducting adhesive. During launch vibration, larger displacements are prevented by snubbers mounted on the primary structure. The snubbers are nominally clear of the stage plate in normal operation thus imparting minimal parasitic heat load. The coils of the 150K to 20K Heat Exchanger are locked by a GFRP threaded ring of similar design to the stainless steel ring used for the 300K to 150K Stage. This GFRP ring is supported by four GFRP clips that are bolted to the 150K stage plate. The 20K Heat Exchanger is bolted directly to the Pre-Cooler cold finger. To provide support when the cooler is operated horizontally in a 1g environment, the stage is supported by Kevlar threads to the J-T Support Structure. This does not provide adequate support for launch and to provide enough Kevlar to do so carries an unacceptable heat load penalty. The Kevlar threads are, therefore, backed up by snubbers of a similar design to the 150K stage. The 20K to 4K Heat Exchanger provides the flexible link between the cooler and the cold bus-bar. It is supported to increase its stiffness by Kevlar threads to two GFRP posts that are attached to the 20K plate. The 4K Stage Plate is attached directly to the spacecraft/instrument cold bus bar. Within the 4K Stage the J-T orifice requires additional support and this is achieved via Kevlar threads to four stainless steel pillars The structural and thermal analysis of the support system and the development testing necessary to verify adhesive performance, Kevlar pre-tensioning techniques and overall structural integrity has been successfully completed.
4K COOLER FLEXIBLE LINK
In general, all mechanical coolers are connected to the relevant instrument/sensor via a separate, flexible, thermal link that is capable of accommodating relative displacements between the Cooler and the payload due to alignment/integration tolerances, contraction at cool down and dynamic displacements during launch. The thermal link has to meet all of these flexibility requirements whilst providing a good thermal path. Such a concept was originally envisaged for the FIRST 4K cooler. Considerable difficulties were experienced in providing the support to the 4K Stage and a review of the design with ESA identified a change of approach that would resolve many of these problems. It was proposed that the 4K Stage of the Cooler be connected directly to the cold bus bar and that the flexible link be provided by the 20-4K Heat Exchanger. Much analysis was undertaken to optimise the design of the heat exchanger such that all of the flexibility requirements were met and yet the heat exchanger remained stiff enough to withstand launch. The shape shown in Figure 4 was finally determined. The majority of the 1m of free pipe is “lost” in a tightly wound coil which contributes little to the flexibility requirements but ensures, that once supported via the Kevlar, launch vibration requirements can be met. The remaining pipe comprises the flexible arm. This arm permits, at qualification level, ±7.5mm of relative movement in all axes. In tests carried out on a fully representative link a number of cycles representative of a factor of 4x greater than that seen in combined ground testing and flight has been successfully completed. This test will be repeated as part of the cooler qualification programme. CURRENT STATUS
Key dates are given below: • J-T Cooler Assembly Manufacture Complete - May ‘98 • Pre-Cooler Qualification Complete - Nov ’98 • Initial Testing of 4K Cooler Complete - March ‘99 • Qualification Programme Complete - July ‘99
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Figure 4 20K to 4K Flexible Heat Exchanger.
CONCLUSIONS
A mechanical cooler based upon RAL developments and capable of lifting 9m W of heat at 4.2K has been designed for the ESA FIRST/PLANCK programmes. Two qualification model coolers are presently in manufacture at MMS, Bristol. These coolers will be qualification tested during 1998/9. The specification of the cooler is such that it will be of interest to other similar applications. REFERENCES 1. Scull, S.R. et al, “Design and Development of a 20K Stirling Cooler for FIRST”, Cryocoolers 9, Plenum Press, New York, 1997, pp 89-96.
2. Bradshaw, T.W. and Orlowska, A.H., “A Close-Cycle 4K Mechanical Cooler for Space Applications”, Proceedings 9th European Symposium on Space Environmental Control Systems, Florence, Italy, 1991 3. Bradshaw, T.W. et al, “Improvements to the Cooling Power of a Space Qualified Two-Stage Stirling Cycle Cooler”, Cryocoolers 9, Plenum Press, New York, 1997, pp 79-88. 4. Bradshaw, T.W. and Orlowska A.H., “Life Test and Performance Testing of a 4K Cooler for Space Applications”, Cryocoolers 10, Plenum Press.
Life Test and Performance Testing of a 4K Cooler for Space Applications T. W. Bradshaw1, A. H. Orlowska1, and C. Jewell2 1
Rutherford Appleton Laboratory, Chilton, Didcot, UK, OX11OQX European Space Research and Technology Centre, PO Box 299, 2200 AG Noordwijk, Netherlands 2
ABSTRACT
At the Rutherford Appleton Laboratory we have developed a 4K closed cycle cooler suitable for space applications. This cooler is based around a 20K two-stage Stirling cycle cooler that is used to pre-cool a 4K Joule-Thomson (JT) system. For the purposes of these studies the 20K Stirling cooler was replaced by a commercial two stage Gifford McMahon refrigerator. The JT system relies on the expansion of helium gas through a fixed orifice. There is concern that this could block with contaminants over a period of time. The cooler is designed to try and overcome this problem in several ways; The gas is cleaned and purified before introduction into the cooler, there is a gas purification system based around a hot reactive getter and there are filters on each stage of the pre-cooler. Contamination was introduced directly into the working gas to assess the efficacy of the purification system. The results of these tests are presented together with a strategy for operation of the cooler in orbit. As the maximum flow through the orifice is a strong function of temperature the full cooling power is only attained at base temperature. This could be a problem if there is a constant temperature independent heat load. The flow through the orifice was measured as a function of temperature in order to aid the modeling of the system. Tests were made on the ability of the cooler to cool from the temperature of the pre-cooler with various applied heat loads.
INTRODUCTION
At the Rutherford Appleton Laboratory (RAL) we have developed a range of coolers for 1,2,3,4 . These coolers are based on the “Oxford” type diaphragm spring suspension system that was pioneered for space use by RAL and Oxford University5,6. One of our current developments is a long life 4/2.5K cooler based the JT expansion of helium through a fine nozzle. This cooler has been baselined as a component of the closed cycle cooling system for the High Frequency instrument on the Planck explorer mission7. space that cover the temperature range 2.5-80K
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This instrument is to measure the cosmic microwave background and an operational lifetime of about 5 years is required. Of concern in this type of cooler is the susceptibility of the system to condensible impurities. These will eventually migrate to the cold end causing blockages in the orifice and loss of cooling power. The susceptibility of the device to contamination has been investigated in this study.
The cooling power as a function of temperature is a required input to any thermal model. This is not a straightforward function of temperature as the flow through the orifice with temperature is highly non-linear. The maximum cooling power is achieved at base temperature with the cooling power falling off towards the temperature of the pre-cooler. In some instrument configurations it will not be possible to achieve base temperature unless the cooling power at intermediate temperatures is greater than the external load. This has been investigated by measuring the flow through the orifice as a function of temperature (which is useful in thermal analysis programs) and by measuring the cooling power at base temperature and at the temperature of the pre-cooler. EXPERIMENTAL LAYOUT
The layout of the cooler is shown in Figure 1 and a photograph of the cold stages without the radiation shields, vacuum can and insulation is shown in Figure 2. Two compressors are used that produce approximately 10 bar on the high pressure side of the JT system and 1 bar on the low pressure side. The compressors are identical to those used in the Stirling coolers that we have developed with the addition of reed valves in the head. These regulate the alternating pressure pulse to give a d.c. flow. Two stages of compression are used. The high pressure gas passes through a hot reactive getter8 and then splits into two streams, one of which is controlled by a room temperature valve. The two high pressure lines are soldered together on the inside of two of the three concentric tube heat exchangers. After the 20K stage the high pressure line controlled by the valve by-passes the top JT heat exchanger and enters the liquid reservoir on the 4K plate. This allows the 4K plate to be rapidly cooled to the pre-cooler temperature by the operation of a single valve at room temperature. As soon as the 4K plate is close to the pre-cooler temperature the room temperature valve is closed and the gas is diverted to the JT orifice9. The heat exchangers between 300 and the 150K stages and between the 150 and 20K stages reduce the heat load on the pre-cooler. The high pressure gas passes through small filters/heat exchangers consisting of a small cell containing fine sintered material. These act to purify the gas as well as promote heat exchange with the stage of the pre-cooler. After expansion of the gas through the JT orifice the liquid collects in a small reservoir containing a quantity of sintered material. The JT orifice consists of a simple crimped tube with a gas exit hole of approximately 12µm. The photograph shows the cold finger of the GM machine to the right with the 4K heat exchangers to the left. The JT expansion orifice is in the small chamber at the top of the picture. The thermometers, heaters and strokes on the compressor pistons are monitored with a data acquisition system based around a personal computer. This reads the voltages from the sensors, performs the necessary calibrations and outputs the results to file and screen. A Allen Bradley resistor was used to monitor the 4K plate temperature. Platinum resistance thermometers were used on the 20K and 80K plates to monitor the temperature. The temperatures of each of the stages were regulated with proportional integral differential controllers that switched heaters on each of the stages of the pre-cooler. Outgassing, and consequent blockage of the orifice by condensable impurities, is a potential drawback of all JT coolers. The RAL 4K cooler has been designed to contain a minimum of nonmetallic parts to reduce the problem, but some, such as the compressor drive coils, magnets and cylinder liners, remain. There is a large surface area of stainless steel tubing in the heat
exchanger system which will outgas hydrogen. In order to reduce the level of contaminants, the coils and magnets are baked before compressor assembly, but at a relatively low temperature (at about 60 C) to prevent deterioration. Off-line tests have measured the levels of outgassing.
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Figure 1 The Layout of the Cooler.
Figure 2 A Photograph of the cooler.
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GETTER EFFECTIVENESS TESTS
A JT system can never be entirely free of contaminants, but there will be a level of contamination that can be tolerated by the system with no deterioration in performance. The design of the heat exchanger system incorporates stage heat exchangers/filters mounted on the two refrigeration stages of the pre-cooler. Any water in the system should be trapped at the intermediate stage (at 120 - 150K), while nitrogen, oxygen and argon will be trapped at the 20K stage. It is known that these condensed impurities may migrate over time so a hot reactive getter is incorporated into the JT circuit before the gas enters the cold part of the system. The contaminants in such a case will be frozen out on the cold surfaces within the cooler, particularly in the stage heat exchangers, or will be of such small size that they will pass through the orifice without clogging. The aim of these tests was to determine this level, if measurable,
and to find the length of time that the cooler could be run without a hot getter. The system was deliberately contaminated with air in the following way. The volume of air used was at atmospheric pressure, which was put into the JT system (total volume When frozen this would have a volume of more than enough to block the system. The JT system was then filled to a total pressure of 5 bar with clean helium through a nitrogen trap. The gas was passed through the cooler by-pass and the hot reactive getter at a fixed flow rate of 3mg/s. A mass spectrometer was used at various stages to monitor the impurity level in the working gas. The analysis of the gas samples showed that although some reduction in the contaminants could be seen in the first few hours of running through the getter, the decrease was not dramatic. This may be due to difficulty in measurement of impurity levels of about 0.1%, i.e. at and that there is scatter in the data, or that levels are so low that results are dominated by contamination in the gas sampling system. These tests were followed by a life test in which the cooler was filled with contaminated gas, and allowed to run with the getter turned off, to show how sensitive the cooler is to known levels of contaminant. It was anticipated that the system would not run for long without blockage since air had been introduced. The effectiveness of the getter when cold had not been appreciated when these tests began. A further cold test was performed as follows. A volume of of air at atmospheric pressure was allowed into the evacuated system which was then filled to a total pressure of 5 bar with clean helium through a nitrogen trap. The cooler was run in the normal way. The pre-cooler was turned on and when the stages were below the ice point the JT compressors were turned on with the by-pass open and a flow of about 2mg/s. When the pre-cooler reaches approximately 20K the by-pass is closed and the JT compressors turned up to give a pressure on the orifice of about 10 bar. During the initial run the JT stage could not be cooled below 60K, as the bypass blocked completely. The cooler was warmed, the bypass unblocked, and the cooldown was repeated. Again, the JT stage could not be cooled below 60K. The getter was switched on and gas was circulated at about 2.5mg/s for 1 hour. The getter was then turned off. After 1 hour of cleaning The cooler was restarted and the JT stage reached a temperature below 20K. The bypass was closed and the compressors turned up to increase the pressure to about 9 bar. The JT stage reached 5.1K (the exhaust pressure was about 1.8 bar). The cooler ran for between 26 and 33
hours before blocking and warming. The getter was switched on and gas was circulated at about 2.5mg/s for 2 more hours. The getter was then turned off. After a further 2 hours of cleaning
The cooler was restarted and the JT stage reached a temperature below 20K. The bypass was closed and the compressors turned up. The JT stage ran between 4.4K and 5.3K (the exhaust
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pressure was varied by changes in the compressor amplitudes). The cooler did not block for
almost six months. During this time the cooler attitude was changed twice, to run “on its side” and “upside down”. There was only a slight difference in performance of the cooler with respect to its orientation. This was found to be because of a slight variation in the temperature of the compressors. This run became the lifetest for this cooler. RESULTS Life test data. A set of results taken over 17 days during the lifetest is shown in Figure 3. The spike in the data at 2040 hours was due to a temporary power failure to the laboratory. The
temperature stability of the system is good with approximately 30mK variation over a 24hr period. This is due mainly to the variation in laboratory temperature causing fluctuations in the pressure drop down the low pressure side of the heat exchanger. The mass of the cold end was approximately 29g although the copper contributes very little to the overall heat capacity of the system. The heat capacity of the volume of gas in the 4K region is approximately 40 times greater than the copper. Cooling power. The cooling power of the cooler was measured in two ways. The first measurement was made with a constant heat load applied at the temperature of the pre-cooler.
Figure 3 The temperature variation of the 4K stage and the pre-cooler.
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Figure 4 Cool-down curves with a constant applied heat load.
Figure 5 The cooling power around base temperature.
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Figure 6 The mass flow through the JT orifice as a function of temperature.
The cooler was then allowed to cool down to base temperature. The results of applying successively increasing heat loads is shown in Figure 4. It can be seen from this figure that base temperature was achieved in all cases up to a load of about 2.5mW. The second measurement was made around the base temperature of the cooler. The results from this are shown in Figure 5. In order to measure the cooling power at the base temperature the liquid in the reservoir has to be
boiled off. This is done by applying a large heat load until the temperature is seen to rise above the base temperature. The heat load is then reduced to a specific value and the temperature of the cold stage observed. If the temperature returns to the base then the cooling power must be greater than the applied load. This process is repeated with successively higher applied heat loads until the cooler is unable to return to the base temperature. It can be seen from the figure that the cooling power at base temperature is around 6mW. The cooling power at base temperature is therefore significantly higher than at the temperature of the pre-cooler. The reason for this is that the mass flow through the JT orifice is a strong function of temperature. This can be seen in Figure 6 which shows the mass flow through the JT orifice as a function of the 4K stage temperature. The data was collected from runs made over two days. The difference in the two sets of data is due to a slight shift in the calibration of the mass flow meter. The mass flow
through the orifice is over three times lower at 20K than at 4K. At base temperature the cooling power of the JT system is compromised slightly by the heat exchanger ineffectiveness so the cooling power at base temperature does not quite scale with the mass flow. CONCLUSIONS
The 4K system is tolerant of contamination. Even with gross contamination the system was able to clean itself up after blockage of the orifice. The getter appeared to work well even when cold. The use of plastics in the construction of the cooler did not cause a major contamination problem. The cooling power of the 4K system is a strong function of the temperature of the 4K stage with a factor of about 2.4 between the cooling power at the temperature of the pre-cooler (in this instance 22-23K) and at base temperature. The flow through the orifice as a function of temperature was measured and found to be four times lower at the pre-cooler temperature than at base temperature. The temperature stability of the 4K stage was found to be of the order of 30mK over a 24hr period. The fluctuations were found to be synchronous with the laboratory temperature and thought to be due to changes in the pressure drop along the low pressure side of the cooler.
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As soon as the gas cleanliness is below a certain level it is not necessary to operate the getter hot. This will save electrical power. It would be prudent, however, to leave the getter in the circuit as it is effective even when cold. Any long term build up of contaminants can be removed by the simple expedient of warming the orifice and circulating the gas through the hot getter. ACKNOWLEDGEMENTS
This work was supported by the European Space Agency under contract number 11342/95/NL/FG. The assistance of R Wolfenden and W Blakesley in the construction of the unit is gratefully acknowledged. REFERENCES 1. Bradshaw, T. W. and Orlowska, A. H., “A Closed Cycle 4K Mechanical Cooler for Space Applications”, Proceedings of the fourth European Symposium on space Environmental and Control systems, Florence Italy, 21-24th October 1991, published in ESA SP-324/ISBN 92-9092-138-2. 2. A Orlowska, T W Bradshaw and J Hieatt, Development Status of a 2.5 - 4K Closed Cycle Cooler Suitable for Space Use, presented at the International Conference on Environmental Systems in Friedrichshafen in June 1994 and at the 8th International Cryocooler conference, Vail, Colorado. 3. T W Bradshaw and A H Orlowska, “Technology Developments on the 4K Cooling System for COBRAS/SAMBA and FIRST”, presented at the 6th European Symposium on Space Environmental
Control Systems, Noordwijk, the Netherlands, 20-22nd May 1997, ESA SP400, vol 2, p465-470. 4. Orlowska, A. H., Bradshaw, T. W. and Hieatt, J., "Closed Cycle Coolers for Temperatures below 30 K", Cryogenics vol. 30, (1990), pp. 246-248. 5. S.T. Werret et al., "Development of a Small Stirling Cycle Cooler for Spaceflight Applications",
Adv. Cryo. Eng. vol. 31, 791-799, (1986). 6. T W Bradshaw, J Delderfield, S T Werrett and G Davey, Adv. Cryo. Eng. vol. 31 801-809 (1986),
Plenum " Performance of the Oxford Miniature Stirling Cycle Refrigerator". 7. “The FIRST/Planck Mission. Cryogenics Systems - Current Status”, B Collaudin and T Passvogel, Proc. SPIE 1998 Symposium on Astronomical Telescopes and Instrumentation, Space Telescopes and Instruments V, 1998. 8. SAES Getters, 1122 E. Cheyenne Mtn Blvd., Colorado Springs, CO 80906, USA. 9. Joule - Thomson By-pass for Cryogenic Cooler UK patent GB 2241565 27/5/94 and USA 5317878
7/6/94.
Long-life 5-10 K Space Cryocooler System with Cold Accumulator Arkhipov V.T., Getmanets V.F., Levin A. Ya Special R&D Bureau (SR&DB) in Cryogenic Technologies Kharkov, Ukraine Stears H. Orbita Ltd, Kensington, MD, USA
ABSTRACT The SR&DB in a joint project with the U.S. Air Force Research Laboratory and Orbita Ltd.
(USA) has developed a concept for a periodically-operating long-life (5 to 10 years) refrigeration system for providing periodic cooling at 5-10 K. The design goal for the system cooling capacity is 0.5 W during 20% of the time. Conducted analyses have shown that the task is best met with a combined system using a two-stage Stirling cryocooler upper stage together with a helium JouleThomson (J-T) throttle cooler for the lower stage. The system also incorporates cold accumulators at 65-85 K (using the melting of nitrogen or argon) and at 20-26 K (using the melting of deuterium or neon) to precool the helium in the J-T system. The cold accumulators are cooled
periodically (or permanently) by a two-stage split-Stirling cryocooler. The periodic 5-10 K load is met by operating the helium Joule-Thomson throttle cooler on demand. The throttle cooler is based on a lubrication-free piston compressor with clearance seals that was developed by the SR&DB. The problem of continuous cooling at 5-10 K was also examined based on the use of an appropriate cold accumulator. Helium is a candidate material for the cold accumulator; or other substances having appropriate physical characteristics (various kinds of phase transitions in the 410 K temperature range) could be used. INTRODUCTION
An important need of modern space instruments is cryogenic refrigeration systems with a lifetime of 10 years and more. For cooling in the 35-80K temperature range there are several available space-qualified cryocoolers. However, for lower temperatures, and in particular for the 4-10 K temperature range, there are only a few candidates (Table 1).1-5 One of the first low-temperature coolers to be flown was tested on board the “Salyut-6” space station. This system, which was quite complex, contained two Stirling cryocoolers (providing 20 K and 80 K cooling) and a continuously operating helium throttle system at 4.2 K. This system had relatively good thermodynamic efficiency (3.0 kW/W) and mass (254 kg/W), but the projected lifetime was limited by the use of the rotary Stirling cryocoolers. As far as we know the life did not exceed 3,000 hours, which is insufficient for long-life applications.
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Another system of a similar type (but with a very low cooling capacity of 0.005 W) has been developed by the Rutherford Laboratory.5 Its projected lifetime is 10 years, but its specific power and mass characteristics are an order of magnitude worse (correspondingly, 26.4 kW/W and 600 kg/W). A third long-life 10 K cryogenic system uses a hydrogen metal-hydride compressor.2,3,4 It has much worse specific mass characteristics (13,300 kg/W), but better power consumption characteristics (6.7 kW/W). This periodic cooler system is designed as a hydrogen liquefier, with the subsequent solidification and sublimation of solid hydrogen at 10 K in the cold accumulator. Precooling of the hydrogen in the J-T stage to 60 K is fulfilled through the use of Stirling cryocoolers.
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Because of the need for long-life 4-10 K cryocoolers having better specific mass and power
consumption characteristics, we have taken another look at possible system configurations addressing the 4-10K temperature range. The work reported here examines one such possible system, the concept of which was developed by us under contract to the U.S. Air Force Research
Laboratory and Orbita Ltd (USA). 5-10 K REFRIGERATION SYSTEM WITH COLD ACCUMULATOR
The overall requirements for the cryocooler system concept described here include: • Cooling capacity of 0.5 watt at 5-10 K with a duty cycle of 20% • An additional cooling capacity of 2 W at 20-25 K (this is equivalent to 0.4 W at 4-6 K) • Projected lifetime of 5 to 10 years
In our opinion, the primary cooling load at 5-10 K can be most effectively met with a helium throttle-cycle system making use of a rotational piston compressor without liquid lubrication (with a clearance seal of the piston-cylinder pair).6 The combination of a helium J-T cryocooler
with a melting-freezing cold accumulator at 18 K (deuterium triple point) or at 24 K (neon triple point) gives the possibility of creating a cryostat system with a 5-10 year lifetime. Such a lifetime can be reached with a compressor lifetime of 1-2 years if it operates only 20% of the time. To lower the power consumption, the helium throttle system is configured with three precooling levels: 18-24 K, 63-83 K and 170 K. In the proposed cryogenic system the cooling level at 170 K would be supplied by a passive spacecraft radiator. The 63 K (83 K) and 18 K (24 K) levels would be supplied by two cold accumulators using solidified nitrogen (or argon) and deuterium (or neon). In turn, these cold accumulators would be cooled by a two-stage split-Stirling
cryocooler (or two one- and two-stage split-Stirling cryocoolers). Preliminary technical requirements on these upper-stage cryocoolers are listed in Table 2. Using the melting-freezing cold accumulators between the Stirling cryocooler and the helium J-T pre-cooling stages provides the following benefits:
• Lower heatload for the second stage of the helium J-T cooler by 3-5 times during the helium throttle system operation
• Increased temperature stability at 20-25 K and 65-85 K • Rapid temperature stabilization at 5-10 K during periodic operation • Maintenance of a high vacuum for the cooled objects, even during periodic warming-ups (up to 20-25 K)
• Maintenance of continuously stable and constant heat loads at the 20-25 K and 65-85 K temperature levels
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Figure 1. Schematic of the proposed 5-10 K periodic refrigeration system.
A schematic of the proposed system is shown in Fig. 1. Its key performance parameters are listed in Table 1, while Table 3 lists the approximate masses of the cryosystem's main components. As seen in Table 3, the mass of such a system is about 65 kg when using a cold accumulator, or 90 kg when using a Ne accumulator. With respect to its figures of merit (specific mass: 130-180 kg/W and specific power: 3 kW/W) the proposed concept proves to be 1.5-2 times more efficient than the other known 5-10 K cooler systems (see Table 1). With respect to efficiency, it is close to the PO “Energija” system, and twice as efficient as the solid hydrogen system; it is more than 60 times better than the hydrogen system on specific mass. A disadvantage of the proposed scheme is that it only provides the possibility of cooling the load 20% of the total time. That is why we have additionally considered an advanced variant of our scheme that can ensure continuous cooling of an object with a thermal load of 0.1 W at 5-10 K during a 5-10 years operational period. The continuously operating cryogenic system (shown schematically in Figure 2) differs from the periodic system shown in Fig. 1 by the presence of an additional helium cold accumulator (working on the evaporation-condensation phase transition),
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Figure 2. Schematic of 5-10 K refrigeration system with helium accumulator.
a liquid-vapor phase separator, a manostat with a pressure of about 1.5 bar, and a bellows gasholder with a volume of 150-200 liters. In this system, continuous cooling is implemented by the helium evaporation in the cold accumulator. The compressor of the helium throttle system is still operated periodically with about a 20%-on, 80%-off duty cycle. Naturally, such a cryogenic system has a 25-30% increased mass as compared to the system with periodic cooling. If the requirement for the additional thermal load of 2 W at 20-25 K was eliminated, then its specific thermal and mass parameters would remain at about the same level as with the periodic system. STATUS OF THE SYSTEM DEVELOPMENT AND REALIZATION OF A 6-10 K CRYOGENIC SYSTEM
At the present time only a conceptual design of the cryosystem has been developed, together with estimates of the overall cryosystem power consumption and mass. Nevertheless, the quality of the estimates is quite high. All calculations and estimates have been based on existing and experimentally verified system components: split-Stirling cryocoolers (made in USA), helium compressor, cold accumulators and their thermal protection systems, counterflow heat exchangers, etc. In particular, SR&DB possesses its own 30-years experience on thermal accumulators and their thermal insulation development,1 has solved the problems connected with excluding contamination from the cold objects and superinsulation, and has developed various kinds of J-T cryocoolers with lubrication-free piston compressors.1,8 CONCLUSIONS
1. A concept has been proposed and estimates made for both a periodic and continuous operating refrigeration system for use at 5-10 K. The system is based on a helium throttle J-T cryocooler. Precooling of the J-T fluid is accomplished at intermediate temperature levels using a cold accumulator at 18-24 K (melting-freezing of solid nitrogen or argon) and at 6586 K (melting-freezing of solid deuterium or neon). These cold accumulators are in-turn cooled by a two-stage or two (one- and two-stage) split-Stirling cryocoolers. 2. It is shown that using this scheme of a periodically operating throttle system in combination with the cold accumulators allows increased system lifetime of up to 5-10 years. 3. Preliminary calculations reveal that the tentative specific power and mass parameters for these new 5-10 K cryogenic systems are better than those of existing systems.
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4. It is shown that in order to ensure continuous cooling at 5-10 K it is necessary to incorporate an additional helium cold accumulator with a liquid-vapor phase separator, manostat, and gasholder; these additions worsen the specific mass parameters by 25-30%. REFERENCES 1.
Arkhipov, V.T., et al., “Cold Accumulators as a Way to Increase Lifetime and Cryosystem Temperature Range,” Cryocoolers 10, Plenum Press, New York, 1999.
2.
R.C. Bowman, Jr., D.R. Gilkinson, R.D. Snapp, et al., “Fabrication and Testing of Metal Hydride Sorbent Assembly for a Periodic 10 K Sorption Cryocooler,” Cryocoolers 8, Plenum Press, New York, 1995, pp. 601-608. S. Bard, J. Wu, P. Karlmann, P. Cowgill, et al., “Ground Testing of a 10 K Sorption Cryocooler Flight Experiment (BETSCE),” Cryocoolers 8, Plenum Press, New York, 1995, pp. 609-621.
3. 4.
P. Bhandari, J. Rodriguez, S. Bard and L. Wade. “Dynamic Simulation of a periodic 10 K Sorption Cryocooler,” Cryocoolers 8, Plenum Press, New York, 1995, pp. 581-600.
5.
B.G. Jones and D.W. Ramsay, “Qualification of a 4 K Mechanical Cooler for Space Applications,” Cryocoolers 8, Plenum Press, New York, 1995, pp. 525-535.
6.
In: Autonomous cryogenic refrigerators of small capacity, Ed. Brodyanskiy, V.M., Moscow, Energoatomizdat, pp. 124-128 (in Russian). Kabankov, A.I., Murinets-Markevich, B.N. “An analysis of the mass and energetic characteristics of the existing machine closed-cycle cryogenic devices,” Low Temperature Engineering, Naukova Dumka, Kiev, 1979, pp. 99-107 (in Russian). Bondarenko, S.I., Getmanets, V.F., “Development of Cryogenic Cooling Systems at the SR&DB in the Ukraine”, Cryocoolers 10, Plenum Press, New York, 1999.
7.
8.
Periodic 10 K J-T Cryostat for Flight Demonstration R. C. Longsworth, A. Khatri, and D. Hill APD Cryogenics, Allentown PA
ABSTRACT A 10 K JT cryostat was built for the Jet Propulsion Laboratory (JPL) which was incorporated in the Brilliant Eyes Ten-Kelvin Sorption Cryocooler Experiment (BETSCE) that flew on STS-77 in May 1996. The JPL system used three separate hydride beds, one to sorb H2 at a pressure of about 400 kPa during cool down and filling of a reservoir in < 2 minutes, and a second to sorb H2 at a pressure low enough to solidify the collected liquid and maintain a sublimation temperature below 11 K for more than 10 minutes. The sorption beds were then sequentially heated to transfer the H2 to a third that was then heated to repressurize the H2 storage bottle. Small Stirling cryocoolers were used to keep a thermal mass and the cold section of the cryostat at a temperature of about 65 K prior to a cool down. The Shuttle environment imposed requirements of designing for launch vibrations, limitations on radiant heat rejection, and the need to design the cooling system for a maximum pressure of 13.6 MPa. This paper describes many of the design concepts and details that solved problems which were unique to this system. These include the reservoir which collects liquid and retains it when it is rapidly depressurized then transfers heat through a heavy wall to the solid H2, the heat exchanger which has high efficiency but has small thermal mass, the dynamics of the thermal storage device, the support, cooling, and flexible coupling of the Stirling coolers, the structural support mechanisms, and the radiator design. INTRODUCTION A periodic 10 K sorption cryocooler with low average power consumption was first conceived in 1991 by Johnson and Jones and described in 1993 1. A proof of principal unit was demonstrated in 1992 2 followed by design and testing of the unit that flew on the shuttle, STS77, in May 1996 3,4. The Brilliant Eyes Ten-Kelvin Sorption Cryocooler Experiment was carried out by the Jet Propulsion Laboratory and incorporated a sorption compressor assembly that was built by Aerojet and the cryostat assembly that was built by APD Cryogenics. A simplified schematic of the system is shown in Figure 1. The sorption compressor is part of the tank and valve assembly that was built by JPL It has, a fast bed which adsorbs H2 at about 400 kPa during cool down from 65 K to 28 K and while the reservoir is filling with H2, a low pressure bed to adsorb H2 to cool it from about 25 K to < 11 K and hold it there for more than 10 minutes. These two beds are sequentially heated to transfer the H2 to the high pressure bed which
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Figure 1 Simplified Schematic of BETSCE System
is in turn heated to repressurize the storage tank at about 10 MPa. Solenoid valves are used to initiate the flow of gas to the cryostat and direct the return flow to the appropriate adsorber.
H2 is liquefied by Joule Thompson expansion in the cryoatat assembly which includes a vacuum housing, three Stirling coolers connected by flexible couplings to a thermal storage mass that is cooled to about 65 K, a counter flow heat exchanger, a reservoir to collect liquid H2 and retain it while it is pumped to a solid, and support structures that enable the system to survive shuttle launch loads. Not shown are radiators that reject heat from the sorption compressors and the Stirling coolers. DESCRIPTION OF CRYOSTAT ASSEMBLY
Figure 2 is a simplified drawing of the cryostat assembly that identifies the major components. Requirements
H2 is supplied from the sorption compressor assembly which has a 3.8 L storage bottle charged to 9.36 MPa at 300 K, a fast adsorber bed that can absorb 7.63 g of H2 at 375 to 410 kPa, and a low pressure bed that can absorb 1.52 g of H2 at < 0.2 Pa. A thermal storage device, TSD, consisting of < 3.5 kg of A1 has to be kept below 70 K by two Stirling coolers (Hughes model 7044H) which have a rated capacity of 3.5 W each at 80 K and produce about 2.5 W at 65 K. From the time when gas flow is initiated to the cryostat the cold plate has to be cooled to < 11 K in < 120 s and held at < 11 K for > 10 m with an applied heat load of 100 mW.
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Figure 2 Simplified Drawing of Cryostat Assembly.
1 2 3 4
5 6 7
8 9 10 11
10 K Cold Plate Reservoir
Reservoir Support Tube JT Heat Exchanger, Lower JT Heat Exchanger, Upper Thermal Storage Device, TSD Thermal Shunt Vent Tube Vent Gas Warm Up Thermal Coupling Stirling Expander
12 13 14 15 16 17 18 19 20 21
Stirling Compressor Cold Finger Support TSD Support Warm Flange Cold Radiation Shield Vacuum Housing Radiator H2 In H2 Out to Fast Sorbent Bed H2 Out to Low Pressure Sorbent Bed
The structure has to be able to survive the shock and vibration loads of the shuttle launch then operate in near earth orbit with exposure to the sun part of the time. 200 W of heat from the compressors has to be rejected by radiation while keeping the compressor housings below 60 C. Because the shuttle is manned, the H2 pressure circuit is designed for a maximum pressure of 13.6 MPa. Maximum system weight was set at 68 kg.
Cold Plate, Reservoir and Support During the period when H2 is flowing from the storage bottle it emerges from the high pressure tube down stream of the JT restrictor and impinges on the backside of the cold plate. It then flows through the matrix in the reservoir that has layers of wicking material which retain the liquid and porous conductor material 5,6 which transfers heat from the cold plate to the gaseous, liquid, or solid H2. The cold plate is Cu which is brazed into the heavy walled SS housing. The reservoir is supported by a long vent/support tube which is connected by a tee to two branches of the vent tube that are clamped to the TSD. The main purpose of the long support tube is to
assure that the lower end is cooled to < 30 K while H2 is flowing so that the amount of liquid H2 that collects which is used to cool the support tube is minimized. This support tube with the
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supply tube in it were not designed to be a heat exchanger but during the flight test the H2 supply valve leaked and fed H2 to the reservoir while it was below 11 K. The thermal loss from this flow was reduced by the heat transfer in this support tube between the supply tube and the vent gas. JT Heat Exchanger and TSD The JT heat exchanger is a bundle of 7 SS tubes of equal diameter that are soldered together with the high pressure gas in the center tube. The upper heat exchanger has 1.5 mm OD tubes and is 2.4 m long while the lower heat exchanger tubes are .8 mm OD by 1.7 m long. The upper heat exchanger is coiled and supported by Kevlar threads under the TSD and the lower heat exchanger is coiled inside the cold radiation shield and tied to support posts. One of the key design concepts to minimize cool down time and gas consumption when cooling from 70 K to 30 K is to have a thermal shunt from the TSD to a point that is about one third the distance from the warm end of the heat exchanger. Having most of the heat exchanger at 70 K when gas flow is initiated results the upper heat exchanger warming up rather than cooling down during this transient period. The TSD was sized on the assumption that it would warm uniformly from about 65 K to 72
K with about 55 % of the load coming from precooling of the H2 and the balance from heat exchanger losses. In practice it was found that the heat exchanger losses were less than allowed
for and the way that the precooling tube was wrapped around the TSD resulted in the A1 near the inlet end of the tube warming a lot more than the A1 near the outlet end. Gas only flows for about
50 s so temperature differences of several K developed in the type 6063 A1. Vent Tube and Vent Gas Warm Up
The vent tube was sized on the assumptions that H2 would sublime at a rate corresponding to 250 mW at a pressure of < 0.35 Pa (10.5 K) and leave the outlet port at 0.1 Pa. It has an effective
diameter of 11.3 mm and length of 350 mm. There was a concern that the high flow rate during depressurization of the reservoir would carry cold H2 through the solenoid valve to the sorbent bed and effect one or both of these. A warm up heat sink was thus designed into the vent tube. Approximately 280 mg of H2 flows in about 5 s when the valve to the low pressure bed is opened and it was estimated that 100 g of A1 would be warmed < 8 K. After several studies it was decided that the warm up sink was not needed so the housing was left open. Subsequent tests confirmed that the cold H2 that vented during depressurization did not affect either the valve of the sorbent material. The pressure drop was thus much less than allowed for during the hold period. Stirling Coolers, Thermal Couplings, and Cooler Support
Test data for the Stirling cooler showed that a cooling capacity of 1.8 W could be expected at 60 K with the compressor case at 60 C. Parasitic heat losses with all 3 Stirling coolers operating were calculated to be 2.0 W and the goal was to have the TSD at < 63 K at the start of each cool down. If one of the coolers is off it was calculated to impose a parasitic load of about 0.4 W on the other coolers. The cold finger of the cooler can withstand a deflection of up to 0.2 mm during launch when it is not operating and during operation. In order to stay within these limits a support bracket was made to fit around the cold finger with Kevlar thread tied from the end of the bracket to the cold finger. Flexible thermal couplings were then designed and built to transfer heat from the TSD to the coolers while reducing the deflection of the TSD transferred to the coolers. Each of the thermal couplings have 20 stranded Cu wires which in turn have 19 wires that are 0.5 mm diameter, 40 mm long, and connected to Cu end plates such that they have a 13 mm radius. Studies showed that the best ratio of thermal conduction to stiffness was obtained by minimizing the diameter of the individual strands of Cu wire. Each coupling has a conductance
of 1.3 W/K so it was calculated that the temperature difference across the couplings with 2 coolers operating would be 1.6 K. This is well within design margins.
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TSD Support
Initially it was assumed that small diameter triangulated G10 tubing would provided the lightest weight support for the TSD assembly but studies of other options led to the use of a single support tube as being a more practical choice. The tube is 88.9 mm OD, 84.3 mm ID, by 150 mm long. It is epoxy bonded to A1 flanges at each end which in turn are bolted to the warm flange and TSD. Conduction loss is calculated to be 580 mW when the warm flange is at 323 K. Warm Flange, Radiator, and Vacuum Enclosure
The warm flange is machined from A1. It has ears for attaching mounting brackets to the side wall of the shuttle, threaded holes on top for attaching the radiator, threaded holes on the bottom for attaching the Stirling coolers and TSD support, and a skirt with through holes for all of the gas and electrical lines. The Stirling cooler compressors and expander heads have A1 brackets clamped around them that conduct heat from the surfaces where heat is being dissipated to the
warm flange. The radiator is 725 mm wide, 1,145 mm long, and 190 mm deep. It is made of T6061 A1 with a silver Teflon coating and is dished to provide structural rigidity with minimum weight. Natural frequency was calculated to be 53 Hz. When radiating 200 W of heat, with the radiator looking away from the sun, the base of the radiator is at 50 C and the heads of the Stirling coolers are at 60 C. The vacuum housing is A1 consisting of a spun dish welded to a flanged sleeve. Cryostat Assembly Weight
Table 1 lists the weights of the cryostat assembly components.
H2 COOLING TRADE STUDIES
The H2 cooling circuit was designed to meet the performance requirements even if there was degradation in some of the component performance. The most significant ones are an increase in the TSD temperature, reduced H2 supply pressure, and increased sorption pressure in the fast sorption bed. Figure 3 shows the calculated effect of increasing the average TSD temperature on the time it takes to cool and fill the reservoir, and the amount of H2 that flows to the fast sorbtion bed. A time to fill of 70 s is acceptable but the fast sorbtion bed is designed for 7.6 g so the average TSD temperature has to be below 70 K if the initial pressure can be kept above 9.25 MPa. If the initial pressure is 8.5 MPa then the TSD has to be below 68.5 K to fill the reservoir.
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Figure 3 Calculated effect of average TSD temperature on time to cool down and fill the reservoir, and amount of H2 that flows to the fast sorbtion bed. Initial pressure of 8.50 MPa, upper, 9.25 MPa, lower.
In practice it was found that the temperature gradients in the TSD resulted in lower exit gas temperatures so the average TSD temperature could be several degrees higher.
The pressure in the sorbtion bed during the fill period has an affect on the hold time at 10 K. As the temperature at which liquid collects is increased there is an increase in the amount that vaporizes in cooling the reservoir to 10 K and solidifying the remaining H2. The calculated relationship between liquid collection temperature and hold time at 10 K is shown in Figure 4. A value of 28 K was used as the design point based on the sorbent bed being at 400 kPa and designing the heat exchanger for a pressure drop of 175 kPa. The design hold time of 14.5 m compares with a required hold time of 10 m. TESTING AT APD CRYOGENICS
Figure 5 is a schematic of the gas supply panel and cryostat assembly which shows the location of some of the pressure and temperature sensors. The gas panel has separate vent lines for the H2, one with a pressure relief valve to set the vent pressure during cool down, the other with a valve that opens direct to a vacuum pump. The cryostat was built with LN2 cooling coil attached to the TSD so that tests could be run without operating the Stirling coolers. Test data from a run that was made with an initial gas pressure of 9.4 MPa and TSD temperature of 68 K is shown in Figures 6, 7, and 8. Figure 6 shows the temperatures at the cold plate, the top of the reservoir support tube, and the TSD. The TSD was cooled by LN2 which was pumped to reduce its temperature. Figure 7 shows the H2 supply pressure and the pressures entering and leaving the return side of the JT heat exchanger during cool down. Figure 8 shows key temperatures during cool down. This test was run with the cold end up. A heat load of 100 mW was applied.
Figure 4 Calculated relation between hold time at 10 K for 0.25 W total load and LH2 fill temperature.
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Figure 5 Test set up at APD Cryogenics.
Figure 6 Test at APD Cryogenics with 9.4 MPa initial H2 pressure, TSD cooled by LN2 to 68 K.
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Figure 7 Pressures during cool down for test with 9.4 MPa H2, TSD at 68 K.
Figure 8 Key temperatures during cool down for test with 9.4 MPa H2, TSD at 68 K.
The results of the test shown in figures 6-8 compare very well with predictions. From Figure 3 we would expect to fill the reservoir in 55 s for these test conditions but the flow was stopped after 62 s. The change in H2 supply pressure is proportional to this extended flow period. Figure 4 predicts a hold time of 14.5 m (870 s) for the fill temperature of 28 K while the measured hold time was 17.9 m (1,076 s). The extended hold time is due in part to extending the fill period but also because the reservoir exceeds its nominal design capacity and the heat losses are less than the 250 mW total design load. The pressure data shows that the pressure drop in the return side of the heat exchanger is greater than predicted which is partly due to the additional pressure drop in the test panel gas lines and valves. The pressure relief valve setting was reduced for this test to set the pressure at the reservoir near 575 kPa during the fill period. The temperature patterns in the cryostat show that some of the important design concepts were realized. Sensor T218 which measures the temperature of the high pressure H2 to be precooled by the TSD starts out colder than after flow is established because the heat exchanger has been precooled by the thermal shunt. Data from the warm end of the upper heat exchanger which is not shown also shows cold gas leaving the heat exchanger for about 10 s then approaching with in a few degrees of the incoming H2. The upper heat exchanger is more efficient than assumed when sizing the TSD and as a result the TSD only warms 3.5 K rather than the 6 K that was assumed.
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Sensor T232 at the top of the reservoir support tube shows that the return flow of H2 up the tube is effective in cooling the tube by the time the reservoir has been filled. As a result the parasitic heat flow into the reservoir is low right from the start of the hold period. The temperature dropped below 11 K within 12 s after opening the vent valve and stabilized below 10 K 15 slater. MODIFICATIONS AFTER DELIVERY Extensive testing was done at JPL3 to verify operation with the sorption compressor system, study sensitivity to off design conditions, and subject the system to the shock and vibration requirements of the shuttle launch. Three problems that required changes in the cryostat were identified and corrected. First, testing with the sorption compressors showed the need for a cold adsorber to filter the gas as it is being charged into the supply bottle. A simple paired tube heat exchanger was added in parallel with the LN2 cooling tubes and connected to a small adsorber mounted on the TSD. Second, the shock and vibration test caused several of the Kevlar ties that held the lower heat exchanger to break because they did not remain in tension. The ties were redone in such a way that they would be kept in tension. Third, a particle carried down to the JT capillary and partially restricted the flow. A filter was added ahead of the JT capillary tube and the entrance to the capillary was put inside a housing that provided room for particles to collect before entering the capillary tube. Another modification that was made at JPL was to replace the superinsulation. Their technician was able to reduce the parasitic heat losses to the TSD by about 1.5 W and thus achieve the loss rate that was predicted.
SUMMARY The BETSCE cryostat assembly was designed without the benefit of a lot of prior experience with many of the technologies that were incorporated. As a result the design was done with margins built in to allow for uncertainties. The JT heat exchanger was more efficient than it had to be and the reservoir had excess capacity. The concepts of the thermal shunt that kept the upper heat exchanger cold and the long reservoir support tube with cold return flow in it proved to work well. The margins paid off in the flight test4 when the supply valve did not fully close and bled H2 through the reservoir. Despite the leak the reservoir stayed below 11 K for the required 10 m and the test was considered a success. REFERENCES Johnson, A.L. and Jones, J.A., “Evolution of the 10 K Periodic Sorption Refrigerator Concept,” 7th International Cryocooler Conference Proceedings, Air Force Phillips Laboratory Report PL-CP-93-1001, Kirtland A.F.B./NM (1993), pp. 831 - 853. 2. Wu, J.J.; Bard, S.; Boulter, W.; Rodriguez, J. and Longsworth, R., “Experimental 1.
Demonstration of a 10 K Sorption Cryocooler Stage,” Advances in Cryogenic Engineering,
Vol. 39, Plenum Press/NY (1994), pp. 1507-1514. 3. Bard, S.; Wu, J.J.; Karlmann, P.; Cowgill, P.; Mirate, C. and Rodriguez, J., “Ground Testing of a 10 K Sorption Cryocooler Flight Experiment (BETSCE),” Cryocoolers 8, Plenum Press/NY (1995), pp. 609 - 621. 4.
Bard, S.; Cowgill, P.; Rodriguez, J.; Wade, L.; Wu, J.J.; Gehrlein, M. and Von Der Ohe, W.,
“10 K Sorption Cryocooler Flight Experiment (BETSCE),” 7th International Cryocooler
Conference Proceedings, Air Force Phillips Laboratory Report PL-CP-93-1001, Kirtland A.F.B./NM (1993), pp. 1107-1119. 5. 6.
Longsworth, R. C. "Cryogen Thermal Storage Matrix", US Patent 5,012,650 May 1993
Longsworth, R. C. "Method and Apparatus for Collecting Liquid Cryogen", US Patent 5,243,826 Sept. 1993
Characterization of Porous Metal Flow Restrictors for Use as the J-T Expander in Hydrogen Sorption Cryocoolers Alan R. Levy1 and Lawrence A. Wade2 1
University of California at Santa Barbara Department of Physics Santa Barbara, C A 93106 2
Jet Propulsion Laboratory California Institute of Technology
Pasadena, C A 91109
ABSTRACT
A program has been completed to measure the flow rate of hydrogen at pressures up to 10 MPa through commercially manufactured porous metal flow restrictors. Such flow restrictors offer substantially greater resistance to contamination induced plugging or changes in flow when compared
with conventional orifices and capillary tubing. The primary goal of the program was to find the appropriate flow restrictor to serve as the Joule-Thomson expander in the cryostat for the UCSB continuous operation hydrogen sorption cooler. This was accomplished by finding a porous plug with the desired pressure drop and mass flow combination at the expected restrictor inlet temperature of about 35 Kelvin. The second goal of the program was to develop the means to accurately predict hydrogen mass flow at cryogenic temperatures from the flow rating given by the manufacturer. This ability will greatly reduce the amount of time and effort required in finding the correct flow restrictor for future sorption cryo-
coolers.
A summary of the cryostat design and test procedure is given. Also discussed are three designs for flow restrictors that were tested and how the final design was chosen. Finally, flow test results are presented and the predictability of cryogenic high pressure hydrogen flow rates from the manufacturer’s rating and room temperature hydrogen flow tests is considered. INTRODUCTION
Sorption cryocoolers offer long-life, vibration-free, reliable refrigeration. A sorption cooler is comprised of a sorption compressor and a Joule-Thomson (J-T) cryostat. The sorption compressor
pressurizes the refrigerant by adsorbing refrigerant at low pressure and desorbing at higher pressure through heating of the sorbent material. The choice of refrigerant depends on the desired cold end temperature and the refrigerant determines the sorbent material. For cooling from 30 K down to 8 K the appropriate refrigerant is hydrogen and the sorbent material is metal hydride1.
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A 25 K continuous operation sorption cryocooler2, shown in Figure 1, has been built and is being tested to fly on the University of California at Santa Barbara (UCSB) long duration balloon (LDB) payload to map the Cosmic Microwave Background Anisotropy at the 0.3 degree angular scale. The High Electron Mobility Transistor amplifier detectors on the payload will be cooled to around 25 K in order to take advantage of their low noise properties. A sorption cooler is desired for cooling the detectors to replace the large (at least 250 liters for a 10 day flight) liquid helium dewar that is otherwise required. Part of the program to build the UCSB LDB sorption cooler was to build and test the J-T cryostat that attaches to the sorption compressor to make the cryocooler. The most time consuming aspect was the testing and characterization of the J-T expander to be used in the cryostat. Because of the low refrigerant mass flow desired, 3.3 mg/s, at a pressure drop of 10 MPa, we decided to use
porous metal flow restrictors instead of the more commonly used capillary tubes or orifices. Porous metal flow restrictors are expected to be more resistant to contamination induced plugging or
changes in flow than the very small orifice or small inner diameter, long length capillary tube that would be used instead. Porous plugs also offer a wide range of possible flow rates at high pressure and they are small in size and easy to implement. Presented in this paper are the results of the porous metal flow restrictor characterization effort as well as a summary of cryostat design, test procedure, J-T design, and to what extent mass flow rates of hydrogen through the porous plugs can be predicted from the manufacturer’s rating and room temperature hydrogen flow. CRYOSTAT DESIGN AND TEST SETUP The test cryostat, which is almost identical to the flight cryostat, is constructed completely of high purity 316L stainless steel weld fittings, assemblies, and electropolished tubing. All joints are
welded with the exception that there are some VCR fittings located on the cold end to allow flow restrictors to be easily removed and replaced. The cryostat consists of a 1.0 m long tube-in-tube
“warm” counterflow heat exchanger, a 15 cm long precooling heat exchanger, a 1.6 m long “cold” heat exchanger, a porous metal contamination trap, the J-T expander, and a liquid refrigerant reservoir. Silicon diode temperature sensors are placed on the cryostat to take measurements at twelve
locations. Precooling of the hydrogen refrigerant at 50 to 65 K and radiation shielding is provided by a Gifford-McMahon cooler. A temperature controller and heaters are used to keep the precooling temperature constant and to provide thermal load to the cold end for determining the net refrigeration capacity of the cryostat for a given mass flow and pressure drop. A pressure transducer is connected to the cryostat to measure the inlet high pressure and a flow meter is connected to the outlet of the cryostat to measure mass flow. To the right in Figure 1 is the flight cryostat without the contamination trap, J-T expander, and reservoir. To reduce the overall length of the cryostat, in addition to coiling the heat exchangers, the warm, high pressure gas flows through the heat exchangers in the annulus between the outer tube (0.635 cm outer diameter, 0.457 cm inner diameter) and the inner tube (0.318 cm outer diameter, 0.216 cm inner diameter). Having the gas flow in this manner allows the refrigerant to be precooled without having to weld extra components at the ends of the precooler. It is likely that this unusual heat exchanger configuration leads to a slightly reduced cold heat exchanger effectiveness as com-
pared to what would be expected for a heat exchanger with the warm gas inside the inner tube. Based on the temperature data taken during the flow tests, this reduction in heat exchanger effectiveness is probably not extremely large and could easily be remedied in future cryostats, if necessary, by adding more length to the cold heat exchanger.
The contamination trap is a point-of-use-filter designed to remove particles larger than 0.01 micron from the refrigerant stream at the part per billion level3. It is located just before the J-T assembly in the cryostat and thus is at a temperature around 35 K. Any contaminants in the refrigerant, such as nitrogen, oxygen, or argon should freeze out of the gas stream and be trapped in the filter before reaching the J-T assembly. In the flow tests, research grade hydrogen (99.9995 percent pure or better) was used to avoid having to pre-filter the refrigerant. During the actual tests, the cryostat plugged very rarely, even after being cold for as long as 24 hours and even though the
cryostat was never baked out. The three times that the cryostat did plug during all of the testing
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Figure 1. Picture of the sorption compressor and flight cryostat. The flight cryostat, seen on the right, does not yet have the point-of-use filter, flow restrictor, or liquid reservoir attached. The sorption compressor is shown just after assembly and before electrical wiring was completed.
could be attributed to errors during switching gas bottles. The liquid reservoir is made from approximately 20 cm of 0.635 cm outer diameter electropolished tubing bent into a U-shaped configuration. A copper clamp (6.35 cm by 7.62 cm by 1.27 cm) is attached to the reservoir to simulate the UCSB LDB focal plane and has temperature sensors and the heater mounted to it. 200-mesh copper screen is placed inside the liquid reservoir to serve as wicking material to separate the liquid hydrogen from the refrigerant stream and retain it in the simulated focal plane. Also attached in the clamp is a 0.318 cm outer diameter tube through which liquid nitrogen is allowed to flow to precool the cold end from around 295 K to 80 K. JOULE-THOMSON EXPANDER DESIGNS
Figure 2 shows three tested designs for the flow restrictor holders. At first, several standard flow restrictors were ordered from a commercial manufacturer3. These standard flow restrictors measured 0.635 cm in diameter by 0.635 cm long and were press-fit into the shaded region in the cylindrical receptacle depicted in the uppermost drawing in Figure 2. The UCSB Physics Machine Shop used 316L VIM/VAR material to make, the flow restrictor receptacles. The wall thickness around the flow restrictor was reduced to allow for welding the restrictor in place in case leaks occurred between the porous plug and the inner wall of the receptacle. The flow restrictors consisted of an approximately 0.2 cm diameter porous metal cylinder encapsulated by a stainless steel sheath. The standard flow restrictors were rated at 1, 10, 25, 50, 100, or 250 SCCM for room temperature nitrogen at 30 psid. After testing, as discussed in reference 2 (Wade and Levy, 1997), it was discovered that none of the standard porous plugs were restrictive enough to match the required cryogenic flow rate at the desired pressure drop. As many as five of the most restrictive (1 SCCM) porous plugs were tested in series after welding their receptacles together. The ratio of mass flow, when cold, to room temperature mass flow was much higher for multiple flow restrictors in series than for one flow restrictor. This ratio of cold flow to room temperature flow was also larger for five flow restrictors than for three. Perhaps it is the case that the hydrogen liquefies in the middle of the flow restrictor chain and the final flow restrictors are, in effect, not as restrictive as the first couple in the chain. Welding the flow restrictors in place reduced the mass flow significantly and unpredictably. It is possible that welding might damage the flow restrictors either through melting or cracking the
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Figure 2. Three designs for J-T expanders. The shaded regions show the location of the flow restrictor element. The top drawing shows the first design that was used with a standard press-fit flow restrictor. The middle drawing depicts the second design with the porous plug pressed in by the manufacturer. The bottom drawing shows the final design in which the receptacles were
manufactured at UCSB and then sent out for insertion of the porous media. Dimensions are in millimeters.
stainless steel sheath around the porous media. During the testing of the standard flow restrictors, some custom flow restrictors were obtained.
Originally, more press-fit type flow restrictors with flow ratings lower than the standard ones were going to be ordered. It turned out, however, that the company could make flow restrictors already encapsulated into tubing using the press-fit receptacles. This way we could avoid the press-fit
operation and also avoid welding reducing unions to the ends of the receptacles in order to connect the flow restrictor to the cryostat. So we bought 0.1,0.2, and 0.5 SCCM flow restrictors made as depicted in the middle drawing in Figure 2. Unfortunately, we did not realize the shortcomings of the design until after testing. One problem was that the 316L stainless steel porous media was encapsulated in a 304 stainless steel shell, which was in turn installed in 316L stainless steel tubing. This meant that a leak between the flow restrictor and the tubing could form due to differential thermal contraction when cycling between room temperature and cryogenic temperature. The other problem was that the porous plug was merely press-fit into place with no way of preventing the plug from slipping due to the 100 atmosphere pressure difference across the porous plug. The result was that the flow rate of hydrogen through the flow restrictor during the second cool down was much higher than during the first cool down. The room temperature mass flow was also much higher after the second cool down than before the first. Thus, the flow restrictor broke and a new design was needed. The third, and final, design is shown in the bottom drawing of Figure 2. The UCSB Machine Shop constructed new receptacles using strain hardened, implant quality 316L stainless steel. These receptacles were sent to the manufacturer where the porous media was inserted, without any encapsulating material, into the shaded area as shown in the bottom drawing of Figure 2. Each porous plug was then high temperature sinterbonded and staked into place, rated, and labeled. A small
ledge was incorporated into the design to insure that the flow restrictor would not be able to slip. One of these flow restrictors was cooled down multiple times and did not show any changes in the flow rate when at room temperature or at 35 K. TEST PROCEDURE
During a standard flow test, once the dewar is evacuated and after taking room temperature flow measurements, cool down is initiated by starting the G-M cooler and starting liquid nitrogen
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Figure 3. Schematic of the test cryostat. The liquid nitrogen precooling line is not shown.
flow through the cold head precooling line. Hydrogen at moderate inlet pressure, about 6 MPa, is allowed to flow through the cryostat to help cool the heat exchangers and the J-T assembly. This also helps to blow out any residual contamination remaining in the cryostat from when it was opened to air in order to change the flow restrictor. The temperature controller is set so that the hydrogen precooler stays at around 50 Kelvin. The temperature, inlet pressure, and mass flow data are displayed on a computer screen and logged using Lab VIEW software. Figure 3 shows a schematic of the test cryostat system. After 12 to 18 hours, the temperature of the cryostat is around 80 K. Some parts of the cryostat might be at temperatures below 77 K because the flow rate of hydrogen has increased enough to provide significant refrigeration. At this point, the liquid nitrogen flow is stopped and the cold end precool line is capped. The inlet pressure to the cryostat is increased to at least 10 MPa to provide enough mass flow to cool the cold end down the rest of the way to liquid hydrogen temperature. The cool down time for the cold end from 70 K to 20 K is about six hours and depends, of course, on the inlet pressure and related mass flow, the mass attached to the cold end, and the refrigerant precool temperature. Once the cryostat is cold, the inlet pressure is set to 10 MPa and the precooling temperature is raised to 65 K (the expected precool temperature when the sorption cooler is operating). The net refrigeration of the cryostat is measured by slowly adding increasing amounts of heat until the cryostat begins to heat up. Flow is measured at the net refrigeration level. The inlet pressure is reduced to measure the net refrigeration and flow at multiple inlet pressures. Measuring net refrigeration and flow as a function of inlet pressure at different precool temperatures is desirable for more complete cryostat characterization, but most of the time the hydrogen runs out before this can be accomplished. TEST RESULTS A summary of the data from tests of restrictors of the second and third designs is shown in
Figure 4. On the right hand side of Figure 4 is a plot of cryogenic hydrogen flow data for two different flow restrictors, which were rated at 0.5 SCCM and 0.72 SCCM by the manufacturer. The 0.5 SCCM flow restrictor is the one that eventually broke while the 0.72 SCCM one is of the third design. The 0.72 SCCM porous plug was tested because, based on the previous data, it was expected that this flow restrictor would give the desired flow of 3.3 mg/s or around 2.4 SLM at a pressure drop of 10 MPa. The data to the left in Figure 4 shows the room temperature hydrogen
flow rates through several different flow restrictors. The 0.1, 0.2, and 0.5 SCCM flow restrictors are of the second design while the 0.29, 0.50, 0.72, and 1.01 SCCM flow restrictors are of the third design.
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Figure 4. Flow rate as a function of pressure for several flow restrictors. Data was taken at room temperature and at cryogenic temperature when the flow restrictor is at about 35 Kelvin.
It is quite striking how linear the flow at high pressure is as a function of inlet pressure (the outlet pressure is about one atmosphere). It is also interesting to note that, although the pressure versus flow rate slope gets steeper as the flow restriction increases, the room temperature flow does roughly scale with the restrictor rating. In addition, it can be seen that the flow rate increases more rapidly with increasing pressure for cryogenic flow than for flow at room temperature. This change in slope is probably due to a higher fraction of the refrigerant becoming liquefied, with a corresponding reduction in flow resistance, as the pressure drop increases. Table 1 gives the numerical flow data for the 0.5 SCCM and 0.72 SCCM flow restrictors.
Figures 5 and 6 show representative cryogenic flow data taken during tests of the 0.72 SCCM flow restrictor. When no thermal load is added to the cold end, the flow reading fluctuates wildly, most likely due to liquid hydrogen pool boiling in the reservoir. Once enough heat is added to the cryostat to match the net available refrigeration, the flow settles down and fluctuates very little. In
order to find the net available refrigeration, the heat load has to be turned up slowly. Too little heat eventually causes a build-up of liquid hydrogen in the reservoir and fluctuations in the flow rate. Too much heat causes the cryostat to heat up rapidly once all the liquid hydrogen in the reservoir
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Figure 5. Cryogenic hydrogen mass flow at 10 MPa.
Figure 6. Cryogenic hydrogen mass flow at 8 MPa.
has evaporated. In practice, the net refrigeration was, at best, measured to 50 mW, which is about 10% ofthe net refrigeration. At this level of precision it takes up to an hour to determine whether the heat input is higher or lower than the net refrigeration. ABILITY TO PREDICT CRYOGENIC FLOW
Using porous metal flow restrictors for future Joule-Thomson cryocoolers would be considerably more convenient if the manufacturer’s rating and room temperature flow data could be used to predict cryogenic mass flow. Then a flow restrictor could be chosen that would be pretty close to what is required without conducting the many cryogenic flow tests that were initially carried out.
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After conducting room temperature tests on the final design flow restrictors, it was discovered that mass flow through them was consistent with the flow through the restrictors of the second design. This gave us hope that the cold flow rates of the new flow restrictors would be the same as that of the older ones. In attempting to predict the cold flow of the 0.72 SCCM restrictor we took the ratio of the manufacturer’s ratings (0.72 divided by 0.5) and multiplied this by the cold flow of the 0.5 SCCM restrictor to yield an expected flow of about 2.4 SLM. As can be seen in Table 1, though, the ratios between cold and warm flow at 10 MPa for the 0.5 SCCM flow restrictor (before it broke) and the 0.72 SCCM flow restrictor are both around 4.6. So it seems to be much more accurate to use the ratios of room temperature to cryogenic flow to make predictions. In any case, to within ten percent at least, we were successfully able to predict cold flow. CONCLUSION
A program has been completed to characterize porous metal flow restrictors for use as the Joule-Thomsom expander in the cryostat for a hydrogen sorption cooler. Three different flow restrictor designs were tried with the final design being the best choice for future use. This design, developed in collaboration with the Mott Industrial Division of Mott Corp.3, is a robust solution that gives predictable results. One of these flow restrictors was tested and it gave the desired flow rate and pressure drop characteristics at cryogenic temperature. The results of this test show that hydrogen flow rates at cryogenic temperature can be predicted to better than ten percent using the
manufacturer’s rating for room temperature nitrogen at 30 psid, room temperature hydrogen flow data, and the results of previous cryogenic flow tests. In the future, as these porous metal flow restrictors are further tested and used for future J-T coolers, more data will be collected and the ability to predict cryogenic refrigerant flow rates should improve. ACKNOWLEDGEMENTS
This effort was completely funded by the National Aeronautics and Space Administration through NASA grants NAGW-1062, NAG5-4078, and NAG5-4185 at UCSB and through NASA Technology Development Funding and NASA Advanced Technology Development Funding at the Jet Propulsion Laboratory (JPL). The work described in this paper was completed at JPL and UCSB. We would like to thank Bill Boulter, Steve Elliot, Bob Losey, Monica King, and Mike Schmelzel at JPL who helped make this program successful. We would also like to express our appreciation to the UCSB Physics Machine Shop, A1 Anderson at A.G. Anderson Co., Inc., and Ed Napiersky at Mott Metallurgical Corp. for their important contributions to this project. Finally we would like to thank Johannes Burger of the Applied Physics Department at the University of Twente in the Netherlands for helpful discussions and his hard work during one phase of the flow testing.
REFERENCES 1.
Wade, L.A., “Advances in Cryogenic Sorption Cooling,” Recent Advances in Cryogenic Engineering–1993, American Society of Mechanical Engineers, New York (1993), pp. 57-63.
2.
Wade, L.A., and Levy, A.R., “Preliminary Test Results for a 25 K Sorption Cryocooler Designed for the UCSB Long Duration Balloon Cosmic Microwave Background Radiation Experiment,” Cryocoolers 9, Plenum Press, New York (1997), pp. 587-596.
3.
Mott Metallurgical Corp., 84 Spring Lane, Farmington, CT 06032-3159
Thermodynamic Considerations on a Microminiature Sorption Cooler J.F. Burger, H.J. Holland, L.A. Wade*, H.J.M. ter Brake, and H. Rogalla
University of Twente, Faculty of Applied Physics P.O. Box 217, 7500 AE Enschede, The Netherlands * Jet Propulsion Laboratory, California Institute of Technology Pasadena, California 91109, USA
ABSTRACT
The sorption/Joule-Thomson cycle is a promising cycle for microscale cooling of lowtemperature electronic devices because the cycle lacks moving parts. This facilitates scaling down to small sizes, eliminates interferences, and contributes to achieving a long life time. A thermodynamic analysis is presented in which the behaviour of compressor and cold stage are analysed separately, leading to a better understanding of sorption coolers. Some fundamental possibilities to improve the thermodynamic efficiency are discussed, and as a part of this a novel two stage compressor concept is proposed. INTRODUCTION
Highly reliable cryogenic refrigeration is the enabling technology for the succesful commercialisation of low temperature electronics1. Some low temperature applications require very little cooling power, such as a single chip with a low noise amplifier or a superconducting SQUID magnetometer. If energy-efficient cryogenic packaging is used for such applications, a cooling power in the low milliwatt range should be enough to operate these systems. A range of cooling techniques is available for cooling such devices, but these are often largely oversized2. The sorption/Joule-Thomson (JT) cycle was identified as a potential candidate for the development of a microminiature cooler aiming at a cooling power in the range of 10 mW at 80K2. The advantage of this cycle is the absence of wear-related moving parts, except for some check valves. This facilitates scaling down of the system to very small sizes, it minimizes electromagnetic and mechanical interferences (which is important for many applications), and it offers the potential of a long life time. A sorption cooler consists of a compressor unit, a counterflow heat exchanger, and a JT expansion valve, see figure 1. Compressed gas coming out of the compressor unit is cooled to the environmental temperature after which it is fed into the recuperative heat exchanger. The compressed refrigerant is expanded in the JT valve to provide refrigeration. The low pressure refrigerant then returns through the recuperative heat exchanger to the compressor unit. The compressor unit contains four sorption cells and several check valves to control the gas flows. Low and high pressures are generated by the cyclic ad- and desorption of a working gas on a sorption material, which is accomplished by cooling and heating of the
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Figure 1. Sorption cooler set-up.
Figure 2. Schematic of compressor cycle.
sorption material. The gas can either be physically adsorbed onto or chemically absorbed into various solids. Usually, heating occurs with an electrical heater and cooling is done with a heatswitch between the sorption cell and a heat sink on the outside (typically a gas-gap switch). A compressor cycle of one cell is schematically shown in figure 2. The cell is heated during sections A and B, and cooled during C and D. During sections A and C both valves of the cell are closed, and the cell is in a regenerating phase. During sections B and D one of the valves is opened; the cell generates a high pressure gas flow out of the cell during B, and a low pressure gas flow into the cell during D. In our application we plan to cool from 300 K down to 80 K. For this temperature range the major drawback of sorption coolers is their limited efficiency. However, the coolers being investigated are intended to supply very little cooling power (range: 10 mW – 50 mW) and for
such a small cooler efficiency is a misleading parameter to compare with the established performance of other cooler types. This is because a cooler with a small cooling power can be very attractive, despite a limited efficiency, if the input power is below a certain limit (e.g. 10 W). On the other hand, it is important to notice that for cooler stages operating at lower temperatures (< 40 K) sorption can become very competative in comparison with other cooling cycles. One reason for this is that highly efficient chemical absorbers can be used in combination with hydrogen gas3. The second reason is that competative regenerative cycles experience substantial losses below 40 K because of decreased regenerator effectiveness. In this paper the thermodynamics involved in a sorption cooler are systematically categorized so as to investigate the physical limitations to its overall efficiency. This
thermodynamic analysis assumes quasi-static conditions, in which the system is considered in thermal equilibrium. In a practical design, however, dynamical effects can occur that lower the performance of the cooler such as temperature profiles in the sorbent beds, pressure drops across the beds, an imperfect heat sink, etc. Therefore, the quasi-static analysis is a best-case consideration. It is applicable in a general sense to understand the physics behind a sorption cooler and as is shown, can usefully identify several important design issues and opportunities. In most published work where the thermodynamic (quasi-static) efficiency of a sorption cooler is optimized, the Coefficience of Performance (COP) of a complete sorption cooler is calculated as a function of the relevant parameter settings4,5. This is a quantitatively sound method, but it does not give a good qualitative insight what exactly influences the COP of the system if the parameters are varied. The system consists of a compressor with aftercooler and a cold stage, and variation of most of the system parameters has different effects on these system components. To be able to study the behaviour of the compressor and the cold stage separately as a function of the system parameters, the exergy potential is introduced in this paper. This
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thermodynamic potential makes it possible to express the available useful energy at the outlet of the compressor (before and after the aftercooler) and at the inlet of the cold stage, so that the COP of the compressor and the cold stage can be calculated separately. This greatly clarifies the influence of compressor temperatures and pressures on the performance of the system components separately, as well as the influence of compressor container materials and dead volumes. In the paper we first discuss the quasi-static analysis of the compressor. By means of the exergy approach the COP of the compressor is determined. The compressor model is described and results of a parameter study are presented. Next, the cold stage is analyzed and it is shown that a straightforward combination of a compressor and a cold stage can lead to poor cooler performance. Finally, two solutions to this problem are considered in the form of a novel two stage compressor arrangement and a precooling configuration. We emphasize that the analysis and the parameter studies in this paper are based on our specific case of a microcooler with a warm-end temperature of 300 K. Also, a specific combination of sorbent and gas was considered: activated carbon and xenon. Nevertheless, the results are applicable in a very general sense. Other gases, materials or temperatures yield different numbers but the trends in the thermodynamic behaviour and the physics behind it remain the same. In our study the combination of carbon and xenon appeared to be appropriate for our microcooler design to construct a first stage that cools from 300 K down to roughly 165 K. COMPRESSOR STATIC ANALYSIS Definition of Exergy
To study the COP of the sorption compressor, a thermodynamic potential must be defined to express the work that is available. The maximum amount of work that can theoretically be
performed with the pressure difference that is generated by the compressor is obtained if no thermodynamic losses are introduced in the work generating process, which means that no entropy is being generated in this process. Consider a compressor that generates a high pressure at a high temperature At the inlet of the compressor a low pressure is maintained at a low temperature Figure 3 shows a qualitative TS diagram of this situation. The maximum amount of work in a process from state Q to state T is obtained along QST: adiabatic expansion QS from to followed by isothermal expansion ST at low temperature towards the final pressure During this route no entropy is generated and This maximum work is called the exergy, and it can be shown that in specific terms it equals:
where is the enthalpy difference between the two states expressed in J/g, and is the entropy difference expressed in J/gK. The result reduces to the change in the Gibbs free energy if It can be shown that Eq. (1) holds also if
Figure 3. TS diagram to illustrate the definition of exergy (see text).
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Compressor modelling
By using the definition of exergy, the COP of the compressor can now be expressed as follows:
where is the exergy of the compressor and is the total heat that is put into the compressor. is found by multiplying the specific exergy of the gas, with the amount of gas that is coming out of the compressor during one cycle: In this expression is the mass of the sorber material and is the mass of the gas liberated from the sorption cell relative to corrected for the gas that remains in the dead volume of the
sorber material and that does not take part in the mass flow out of the cell,
Hence,
where is the fraction of the dead volume in the sorber material (interparticle voids and macro pores), is the density of the gas at the high pressure and temperature and is the density of the sorber material. The value of the specific exergy of the gas, in Eq. (3) is dependent on which state (temperature) is taken for the high pressure gas that is coming out of the compressor. The exergy can be calculated with the state of the high pressure gas before and after the aftercooler, corresponding to states Q and R in figure 3. The difference is the loss of exergy due
to the (after)cooling of the gas. In this paper the exergy of the gas is calculated with the state of the high pressure gas after the aftercooler. This exergy eRT is relevant for driving the cold stage. The total heat put into the compressor is the heat required to heat up the thermal mass of the sober material and the container, plus the heat required to heat up the adsorbed gas (approximated by ), plus the desorption energy of the gas that is liberated from the surface of the adsorption material4: where and stand for the specific heat of respectively the sorber material, the container material and the adsorbed gas, the mass of the container, the temperature difference of the cycle and the desorption energy of the gas/sorber combination expressed in J/g. The mass of the container is determined by the wall thickness that is required to withstand the high pressures. For a cylindrical configuration this thickness is given by:
In this expression is the high pressure, R is the radius of the cylinder and is the maximum allowed tensile stress in the container material. If the mass of the endcaps of the cylindrical container is neglected compared to the mass of the cylinder itself, then the ratio of the containersorber mass can be determined as follows:
As can be seen, this ratio is independent of the container mass or size itself. Combining
expressions 6 and 8, the input power follows as:
The COP in Eq. (2) can now be calculated by deviding equations 3 and 9. If the adsorber properties and adsorption isotherms for a desired gas are known, then the compressor COP can be calculated under various operating conditions.
THERMODYNAMICS OF MICROMINIATURE SORPTION COOLER
557
Parameter study
In this section a parameter study is described that shows the influence of gases, adsorber materials and container materials. As an example of the process, the compressor performance is calculated with xenon as the working gas and a typical high surface area active carbon, Anderson charcoal6. This type of charcoal is readily available, in contrast to Saran charcoal which has a much lower dead volume fraction. Calculations with both types of charcoals are compared. The sorption data of Xe on Anderson and Saran charcoal were measured at NIST7, and the gas properties were calculated with Cryodata's fluid property program GASPAK8. The following material properties were assumed in the analysis. Anderson charcoal: dead volume fraction Saran charcoal: heat of adsorption for xenon on both charcoals: stainless steel container: In general, the low and high temperatures and pressures of the sorption compressor (respectively denoted by ) are the main parameters that have to be chosen, and are, therefore, of major interest in a parameter study. directly determines the maximum amount of gas that can be adsorbed at a certain (see figure 2). In the present study was taken at ambient level (i.e. 300 K). only influences the compressor performance, and can be chosen freely within practical limits. In a JT expansion stage, determines the cold temperature of the cooler, but it can also strongly affect the COP of the compressor. This influence is evaluated in the compressor parameter study. The high pressure pH strongly influences both the performance of the compressor and that of the cold stage. In figure 4 two plots are shown of the as a function of for different values of in figure 4a bar and in figure 4b bar. Two competing effects influence the COP as increases. For low values of increasing increases whereas
in Eq. (3) hardly decreases, which causes a net rise of the COP. At a
certain starts to decrease significantly - partly because of an increase of with and partly because of an increased amount of gas that is stored in the dead volume, (see figure 2). When this decrease in can no longer be compensated by an increase of the COP starts to decrease. The COP drops to zero when all liberated gas from the adsorber surface is lost in the dead volume of the adsorber material, and no gas is liberated from the compressor anymore. Another effect which tends to decrease the COP at higher values of is the increased wall-thickness that is required at higher pressures, see equations 7 and 8. As a consequence, relatively more heat is lost in the walls of the compressor. In general, an increase in increases the net amount of gas that is liberated from the
compressor, because is increased and the density of the gas that is lost in the dead volume is lowered. However, a higher also increases the required input power. At very low values of an increase of does not liberate a significantly higher amount of gas out of the compressor and it, therefore, leads to a reduction of the COP. At these values of most of
Figure 4. The compressor COP as a function of the high pressure for
bar (a) and
The calculated points for a two stage compressor in figure (a) are discussed later in the text.
bar (b).
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SORPTION CRYOCOOLER DEVELOPMENTS
Figure 5. (a) The maximum compressor performance as a function of the low pressure, for optimized and (b) Influence of the dead volume fraction and container material on the compresor COP.
the gas is already liberated from the adsorber surface at lower values of Above a certain value of the increased amount of liberated gas at higher values of is of significant benefit. An increase of has two effects on in Eq (3). Firstly, is decreased and, secondly, the amount of gas adsorbed at and is increased. At small adsorption rates the adsorption rate increases approximately linearly with
(see figure 2), whereas
decreases less than
linearly with increasing (at 300 K 20% in the case that increases from 1 to 2 bar for Xe). The net result is an increase of the COP with increasing At higher adsorption rates, does not increase linearly with anymore because the isotherms enter the saturation region. In that case starts to decrease with increasing because of the decreasing see figure 5a. Saturation of the charcoal is, therefore, a limit for compressor performance. There are several other parameters that can influence the general described behaviour, these
are separately discussed below. Isotherms. It is clear that the compressor performance improves for adsorption materials with higher amounts of gas adsorbed at the same temperatures and pressures. These adsorption rates are strongly related to the internal surface area that is available for adsorption. A microporous carbon like Saran or Anderson is very suitable in this respect3.
Dead volume fraction. The large influence of the dead volume fraction that was mentioned
before, has been highlighted in several publications9. The influence on the analysed Xe system is illustrated in figure 5b, where the COP is compared with for Anderson and Saran charcoal It appears that the reduced dead volume fraction slightly increases the maximum COP at low pressures, but that it increases the COP especially at higher pressures. A high dead volume fraction and associated low sorbent density also reduces the COP at higher pressures indirectly, because relatively more heat is lost in the container wall, see also the discussion about material choices below.
Material choices. The ratio of the heat capacities of the container and the sorber materials is a measure for the relative amount of heat that is lost in the compressor container. By using Eq. (8), this ratio can be written as:
This expression can be used as a handsome tool to evaluate the container heat capacity losses as a function of the material properties and the high pressure In table 1 a comparison is made between different possible container materials. The parameter can be used to compare materials with respect to each other. As a typical example, is calculated in the last column for the density of Anderson charcoal and a high pressure bar. It follows that it can be advantageous to use high strength alloys like Titanium or Inconel if higher pressures are
THERMODYNAMICS OF MICROMINIATURE SORPTION COOLER
559
required, and also ceramic containers can be attractive with respect to minimisation of heat capacity losses. The use of high strength alloys, however, may result in practical problems in the
case of small compressors because very thin wall thicknesses are required in that case, which may be difficult to realize. For example, in the case of a 1 cm diameter Titanium compressor suited for a high pressure bar, a desired wall thickness of about is obtained. Note that for a certain the container heat capacity losses are reduced if a sorbent is used with a high density and low dead volume fraction, like Saran, instead of a low density charcoal. Figure 5b illustrates the influence of the container material choice on the COP as a function of the high pressure, as well as the dead volume influences. From this figure it appears that the influence of the container material is small for pressures of interest, both for Anderson and Saran
charcoal. Compressor conclusions Some concluding remarks can be made with respect to the compressor modelling. For the special case of a Xenon - Anderson charcoal compressor operating at and bar, constructed of straightforward stainless steel 316 container material, a maximum COP of 3.5% at bar can be obtained. At higher pressures the COP decreases rapidly to zero at about 35 bar. The maximum COP can slightly be increased (to about 4%) by using high strength materials, and somewhat higher pressures can be obtained by increasing to 700 K. Much higher pressures can be obtained by using an adsorber material with a low dead volume fraction, e.g. with Saran at 40 bar a COP of 2.5 % can be realized.*
COLD STAGE STATIC ANALYSIS
In order to model the Linde Hampson cold stage, it is assumed that this stage including the counterflow heat exchanger is without losses. In that case the cooling power equals the enthalpy difference that is created between the low and high pressure sides of the warm inlet of the CFHX. Now the COP of the LH cold stage can be defined as:
In this expression is the exergy or Gibbs energy at the inlet of the CFHX as given by Eq. (1), or the minimum work of compression that is required. The subscript w refers to the warm * From thermodynamic point of view, the COP’s of a sorption compressor should not directly be compared with the large COP’s of a mechanical compressor, but with the COP’s of an engine. Like in an engine, in a sorption
compressor thermal energy is converted into mechanical work. Typical engine performances are less than 40%. A COP of 4 - 5% for a sorption 'engine' without moving parts that is, therefore, easily scalable to small sizes is, from a fundamental point of view, a promising result. Moreover, if a somewhat larger system is allowed including heat regenerating facilities that can recover more than 75% of the heat10, then COP's close to 20% should be obtainable.
This number is not very far away from the engine Carnot efficiency for the temperatures that are used! Of course the comparison with the thermodynamic engine efficiencies is less relevant when electrical heaters are used to drive the compressor cells, since high grade energy is in that case first degraded to thermal energy.
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SORPTION CRYOCOOLER DEVELOPMENTS
Figure 6. COP of the Xe cold stage as a function of the high pressure for different warm temperatures.
end of the CFHX. For a certain working gas the parameters that can be varied are
and is fixed by Figure 6 gives for Xenon a plot of as a function of for different values of for the case bar. The COP is normalized on the Carnot efficiency to make a better comparison possible between the curves for varying For low values of the performance is rather poor, since only small enthalpy differences can be created. The steep increase of the COP at higher pressures is because the fluid liquifies during compression. This transition moves to higher pressures if is increased, at the highest values of the increase is not so steep because the temperatures are above the critical temperature of xenon, but still reasonable performances can be obtained because of the strong non-ideal behaviour of the fluid in that regime. Very similar plots can be obtained for larger values of COMBINATION OF SORPTION COMPRESSOR AND COLD STAGE
If the heat sink temperature of the sorption compressor is also taken as the precooling temperature of the gas that enters the cold stage, then the COP of the Xenon sorption cooler can be obtained by multiplying and If the curve for of figure 7 is multiplied with one of the curves in figure 4a, then it appears that a very poor overall performance is obtained. The reason is that the sorption compressor only performs well at low pressures, whereas the cold stage requires high pressures to obtain a reasonable performance. This statement does not hold for compressors that are based on chemical sorption (e.g. hydrogen/metalhydride). These operate well at high pressures3. We see three possible solutions to overcome the bad matching of the (physical adsorption) compressor and the cold stage: • The dead volume fraction of the adsorption material can be reduced by using a high density charcoal like Saran. The compressor high pressures are in that case still limited to about 60 bar, but the overall performance does improve. • A two stage sorption compressor can be applied to enable generation of much higher pressures.
• The gas at the inlet of the cold stage can be precooled to lower temperatures with another cooler in order to improve the cold stage performance at low pressures. The last two options will be considered in detail below. Two stage sorption compressor In a two stage sorption compressor, that was earlier proposed in a different composition by S. Bard9, the gas is compressed from a low pressure to some intermediate pressure in a first stage, then flowed into a second stage where the gas is compressed from intermediate to high pressure. The cycle is illustrated in figure 7a. Each compressor stage operates in a similar way as the single stage that was described before. Because both compressor stages have a limited
THERMODYNAMICS OF MICROMINIATURE SORPTION COOLER
561
Figure 7. (a) Cycle of a two stage compressor; (b) Integrated two stage compressor (see text).
pressure ratio it is possible to operate them close to their thermodynamic optima (i.e. the peaks in figure 4) so that high overall COPs can be obtained at very high pressures. Moreover, the influence of the dead volume fraction is greatly reduced, and under some conditions completely irrelevant which makes adsorber selection much easier. Figure 7b shows the novel interconnection scheme of the 4 low pressure and 4 high pressure cells that we propose. It was recognized that one specific low pressure cell always blows the gas into one specific high pressure cell, so only one check valve is required to interface these two cells. During the cyclic operation the low and high pressure cells are in a constant phase difference with respect to each other and can, therefore, easily be combined in one sorption unit consisting of two compartments. This enables the use of a single heater and heat-switch for a unit of a combined low and high pressure cell. In this way the combined low and high pressure cells are always in a similar cooling or heating phase of the compressor cycle, but at different pressure levels. For instance, the low pressure cell of unit I is in phase B and blows the compressed gas in the high pressure cell of unit III, and the high pressure cell of unit I is in phase F and blows the compressed gas into the cold stage. A two stage compressor that compresses gas from to is only feasible if, for a certain intermediate pressure both stages deliver the same amount of gas. If this condition is satisfied, the net amount of gas that is freed from the low pressure stage at has to be adsorbed at the high pressure stage at This condition can be satisfied by chosing the proper sorber mass fraction for the two stages so that where stands for the net amount of gas that is liberated from stage i (1 or 2):
Now the condition
gives the required sorber mass fraction to obtain
The COP can now easily be evaluated by calculating the exergy that comes available and the total heat that has to be put into the two stage compressor. Doing this for for different intermediate pressures an optimum in is found. The calculated optimum COP for different high pressures is added to figure 4a, for the case bar, Anderson charcoal and a stainless steel container. The intermediate pressure ranges from 6 to 18 bar for increasing It can be concluded that a two stage compressor facilitates the generation of much higher pressures compared with a single stage compressor, even with very straightforward sorber and container materials. Now a reasonable cooler performance can be obtained by combining a two stage compressor with a cold stage that is precooled at 300 K. For instance, a total
is obtained for
bar.
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SORPTION CRYOCOOLER DEVELOPMENTS
Figure 8. Schematic diagram of LH cold stage with precooling arrangement.
Precooling of the cold stage The performance of the LH cold stage at lower pressures can be improved by lowering the temperature of the high pressure gas before it enters the CFHX, as was indicated in figure 6. A
similar effect can be obtained by actively precooling of the high pressure gas in the CFHX, see figure 811. For the total of this system can be written:
where and are the heats required to drive the compressor and the precooler, is the coefficience of performance of the precooler and is the ratio of the power that has to be cooled away by the precooler relative to the cooling power. Precooling is useful if the can be increased by increasing This is only possible if the reduction of is not taken away by a large increase of In the case of a sorption compressor with a relatively low precooling can be very attractive because can be reduced significantly with only little required If the CFHX behaves ideally, the heat taken away by the precooler can be calculated using an enthalpy balance:
If the performance of a certain precooler is known, this expression can be used to evaluate the proper precooling temperature. It is a logical choice to use another sorption cooler as a precooler by choosing a different gas that operates at a higher temperature. This concept was used in several coolers at JPL12. Also thermoelectric precooling has been used. The significant improvement of the performance of the xenon cooler is illustrated in figure 9, in which the total performance is plotted as a function of the high pressure, for different temperatures of the TE cooler. The dramatic improvement results from the preliquefaction of the refrigerant by the TE cooler and the subsequent near ideal thermodynamic performance of the JT expansion process. The numbers are calculated by application of equations 14 and 15, and COP values for the TE cooler that were obtained from
Figure 9. Performance of a Xe sorption cooler with TE precooling for different values of
THERMODYNAMICS OF MICROMINIATURE SORPTION COOLER
563
Melcor for multi-stage coolers13. By precooling the high pressure gas to 230 K for instance, a reasonable can be obtained for bar and Anderson charcoal in a stainless steel compressor. This COP can slightly be improved by using Saran instead of Anderson charcoal. CONCLUSIONS
The thermodynamic behaviour of sorption coolers is explained by a systematic analysis in which the compressor and cold stage are treated separately, both for quasi-static conditions. The parameter studies in this analysis are based on our specific case of a microcooler with a warmend temperature of 300 K. Nevertheless, the results are applicable in a much broader sense.
In general, the considered compressor performs well but only at relatively low pressures, whereas the Joule-Thomson cold stage requires high presures for proper performance. Two solutions were discussed to overcome this conflict: a novell two stage compressor and (TE) precooling of the gas in the cold stage. In this way, a COP of about 3 % can be obtained for a carbon/xenon cooler operating between 300 K and 165 K. This can be used as a first stage in our microcooler.
ACKNOWLEDGEMENTS
This research is supported by the Dutch Technology Foundation (STW). REFERENCES
1. 2. 3. 4. 5. 6. 7.
Nisenoff, M., Cryocoolers and high temperature superconductors: advancing toward commercial applications, Cryocoolers 8, Plenum Press, New York (1995), pp. 913-917. Burger, J.F., ter Brake, H.J.M., Elwenspoek, M., Rogalla, H., Microcooling: Study on the application of micromechanical techniques, Cryocoolers 9, Plenum Press, New York (1997), pp. 687-696. Wade, L.A., An overview of the development of sorption refrigeration, Adv. in Cryogenic Eng. 37 (1992), pp. 1095-1106. Bard, S., Development of an 80-120 K charcoal-nitrogen adsorption cryocooler, Proc. 4th Int. Cryocooler Conf. (1986), pp. 43-56. Chan, C.K., Optimal design of gas adsorption refrigerators for cryogenic cooling, Proc. 2nd Biennial Conf. on refrigeration for cryogenic sensors and electronic systems (1982), pp. 323-341. Commercially available from The Kansai Coke and Chemicals Company, Ltd., Japan. R.Radebaugh, National Institute of Standards and Technology, Boulder (1992).
8. Cryodata Inc., Niwot, Colorado. 9. Bard, S., Improving adsorption Cryocoolers by multi-stage compression and reducing void volume, Cryogenics, vol. 26 (1986), p. 450-458. 10. Alvarez, J.A., Krylo, R.J., Snapp, R.D., Weston, C., Sywulka, P., Abell, G.C., Development of an advanced sorption compressor and its application in a 125 K cryocooler, Cryocoolers 8, Plenum Press, New York (1995), pp. 569-579. 11. Lester, J., Closed cycle hybrid cryocooler combining the Joule-Thomson cycle with thermoelectric coolers, Adv. In Cryogenic Eng., vol. 35 (1990), pp. 1335-1340. 12. Bard, S., Jones, J.A., Schember, H.R., A two-stage 80 K – 140 K sorption cryocooler, Proc. ICEC 12, Butterworths, Guildford, UK (1988) 13. Melcor thermoelectronics, Trenton, NJ.
Fast Gas-Gap Heat Switch for a Microcooler J.F. Burger, H.J. Holland, H. van Egmond, M. Elwenspoek, H. J.M. ter Brake. and H. Rogalla
University of Twente, Faculty of Applied Physics P.O. Box 217, 7500 AE Enschede, The Netherlands
ABSTRACT
A sorption compressor requires heat switches to thermally isolate the cells during heating, and to connect them to a heat sink during cooling. The requirements for these heat switches are discussed and related to important compressor parameters. It is shown that under certain conditions a sorption compressor can be operated without heat switches at all. Furthermore, the static heat transfer behaviour of a gas gap is modelled in detail and compared with experiments on a 300 µm gas gap. Finally, the dynamics that limit the switching speed are discussed. INTRODUCTION
A microminiature sorption cooler is under development at the University of Twente1. The compressor of such a cooler contains four sorption cells that generate a pressure difference by the cyclic ad- and desorption of a working gas on a sorption material, which is accomplished by cooling and heating of the sorption material. In order to thermally isolate the sorption material during heating and thermally connect it to a heat sink during cooling, heat switches are required between the sorption cells and the environment that acts as a heat sink. The properties of these heat switches strongly influence the compressor performance and must, therefore, carefully be designed to fulfill the requirements. This paper presents a design analysis of a fast miniature gasgap heat switch that is used in the microcooler development. In the paper first the heat-switch requirements for a sorption compressor are extensively discussed. Next, it is argumented why a gas-gap heat switch is one of the attractive concepts for a small sorption compressor. The thermal conductivity of a gas gap is subsequently modelled to show the influence of different parameters, and this is experimentally verified. Finally, the dynamic behaviour during on and off switching is treated. This behaviour limits the switching speed, which is an important parameter for small compressor cells. HEAT-SWITCH REQUIREMENTS
The heat-switch requirements are related to the dynamic temperature cycling of the sorption cells. Fig. 1 illustrates the temperature of one compressor cell during one complete cycle of heating and cooling. To pressurize the cell and to generate a high-pressure gas flow out of the cell, it is heated uniformly. To depressurize the cell and to generate a low-pressure gas flow into
Cryocoolers 10, edited by R G. Ross, Jr. Kluwer Academic/Plenum Publishers, 1999
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SORPTION CRYOCOOLER DEVELOPMENTS
Figure 1. Illustration of the temperature cycle of one compressor cell.
it, it is cooled passively to the environmental temperature heat switches can now be deduced as follows.
The different requirements for the
1. Thermal resistance in the ON-state. The required thermal resistance can be related to the compressor input power via the following argumentation. The heat switch thermal resistance is an important part of the thermal system that determines the cooling behaviour of a compressor cell. If a sorption cylinder with a large aspect ratio is assumed (which is attractive for several reasons), then only the radial thermal behaviour is of real importance. Fig. 2a gives a typical radial thermal model of one sorption cell. It consists, respectively, of the sorption cylinder containing sorber material that can be modelled as a distributed heat capacity and thermal resistance; the heat switch thermal resistance; another distributed heat capacity and
thermal resistance representing the thermal link between the heat switch and the heat-sink device; and the thermal transition resistance to the heat-sink temperature. Under the highly desirable condition that the heat switch thermal resistance controls the cooling behaviour of the cell, the model reduces to the lumped model of Fig. 2b. This is possible if the system is designed such that only limited temperature gradients are present in the sorption cylinder and in the thermal link to the heat sink. In this case the temperature difference during cooling falls across the heat-switch resistance, and it can be used as a tuning variable during design and operation. During operation of a sorption compressor with four cells, the input power, is put in two of the four cells, so that each cell is heated with during half of the total cycle period, This means that the amount of heat put in one cell during a full cycle equals:
and this heat causes a temperature increase of the cell, which is determined by the heat capacity of one cell (for simplicity it is assumed that the heat of desorption is much smaller than the heat put in the heat capacity2, a typical value for the fraction of these two for charcoal compressors is 10–20 %): If the heat-switch thermal resistance has a constant value during cooling down of the compressor cell, the temperature of the cell follows an exponential decrease towards the heat-sink temperature where
is the RC-product of heat-switch resistance and sorption cell heat capacity.
Figure 2. (a) Typical radial thermal model of a sorption cell. (b) Simplified lumped model (see text).
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From our sorption thermodynamic modelling it follows that the compressor performance rapidly
reduces for increased minimum compressor temperatures2. As a consequence, in Eq. (3) should approach as close as possible at the end of the cooling cycle, after a period where n determines how close
within 1% of
is approached. For example, for
approaches
which is a reasonable value. Under the assumption that not the rate of
heating but instead the cooling rate is the limiting factor for the cycle period, the cooling period must be synchronized with the heating period of so that: Combination of equations 1, 2 and 5 yields:
This is a handsome expression that can be used in the heat-switch design phase to relate the heatswitch ON-resistance to the desired temperature difference and the input power of the sorption compressor. Note that this expression is independent of the compresor cell dimensions or the cycle timing. However, the expression is not useful anymore if the assumption is violated that no temperature gradients are present in the compressor cell. And this assumption is, of course,
strongly dependent on the chosen dimensions for a certain compressor input power! Some heat-switch concepts can also be used to regulate the heat-switch thermal resistance during cool down of the cell, for instance to modify the exponential cool down behaviour to control the low pressure in the cells. However, to be able to reach the same cold temperature after a much lower value of should be obtainable in the last part of the cool down period. 2. Thermal resistance in the OFF state. The heat switch in the OFF-state should isolate the compressor cell during the heating part of the cycle. The heat conduction through the OFF resistance during this period is a loss term, and should be minimized. If a linear heating ramp between and is assumed during the heating part and is considered to be independent of the temperature, then the total average power loss for the four cells is given by:
Now the total input power equals compressor input power can be expressed as
The ratio of the loss term and the useful so that can be written as:
where should preferably be much smaller than unity. Apart from this ratio can be defined by dividing Eq. (8) and (6), leading to
also an ON-OFF
From this expression it can be seen that if, for example, the compressor cells should cool to within 1% of and at most 5% of heat may be lost during the heating cycle then the ON-OFF ratio should be at least 50. An interesting effect occurs when which means that no thermal switch is present at all, but instead a fixed thermal resistance. If is chosen, it follows that and This means that, in principle, a sorption compressor can be operated without heat switch at all with the penalty of a reduced efficiency, for this example with about a factor of 2. Such losses can be reduced by using more than four sorption cells, so that the cooling period of one cell can be extended over a longer time: where
remains the heating period and m is the number of sorption cells Using this expresion, Eq. (6) can be recalculated for m sorption cells and from that the following ON-
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SORPTION CRYOCOOLER DEVELOPMENTS
OFF ratio for m cells is obtained:
If no heat switches are present then and if, for instance, and is chosen, then is obtained. This means that with a few more than four sorption cells and a slightly higher input power, a sorption compressor can readily be operated without heat switches at all. For some applications this may be an attractive alternative, for instance in the case that none of the heat-switch alternatives that are mentioned later is feasible, e.g. due to small dimensions. 3. Heat-switch speed. The heat switch should be able to switch in a relatively short period
of time, after the start of the heating or cooling part of the compressor cycle, for instance within the first 10% of it. This would imply As a consequence, the required heatswitch speed is proportional to the total compressor cycle period, which follows by combining Eq. (l) and (2):
The thermal mass in this expression is proportional to the volume of one compressor cell, and the compressor input power can be considered proportional to the required gas mass flow and cooling power2. As a typical example, we are working on compressor cells which are 5 mm in diameter and 5 cm in length, that are operated with a total compressor input power of about 5 W and a temperature difference of 300 K. For this case which leads to typical cycle periods of 100 s and a required heat-switch speed of less than 5 s. 4. Life time. Sorption coolers have the potential to reach life times of ten years or more because of the absence of moving parts. If a small sorption compressor is operated with a cycle period of 100 seconds, this would require switch actions of the heat switch. Obviously, this is a hard requirement in the design of a heat switch. 5. Heat-switch temperatures. The heat switch of a sorption compressor cell is connected
between a heat-sink at constant temperature and the sorption cell that is thermally cycled between and The heat switch should be resistant against these temperature variations and gradients. Typical values for activated carbon as the sorption material are: 300 K and 6. Heat-switch thermal mass. The relatively large input power that is required for the cyclic heating of the thermal mass of the compressor cells is responsible for the relatively low Coefficient of Performance of sorption compressors and sorption coolers2. This thermal mass should, therefore, be kept as small as possible and as a result the heat switch should not significantly contribute to this thermal mass. Now two different heat-switch arrangements can be considered: an ‘external’ heat-switch device that is connected to the compressor cell via a thermal link, and an ‘internal’ heat switch in which the compressor wall is an intrinsic element of the switch. In general, the thermal conductivity of sorption materials is low and to prevent temperature gradients during heating and cooling, the thermal path of the heat to be conducted away should be kept as small as possible. One important way to reach this is to use the complete outer surface of the compressor cell to conduct away the heat*. An ‘external’ heat switch that is connected to the complete outer surface of the cell requires a thermal link that will add much thermal mass to the compressor cell, which deteriorates its performance. In contrast, an ‘internal’ heat switch that connects directly to the outer surface of the compressor cell, using this as the
* A further reduction of the required thermal path and associated temperature gradients is obtained by choosing the aspect ratio of the cell (with a given sorber volume) as large as possible.
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temperature varying part of the heat switch, does not add any thermal mass to the cell at all, but still can use the complete outer surface area to conduct away the heat. MODELLING OF GAS-GAP STATIC THERMAL BEHAVIOUR
A number of heat-switch alternatives have been investigated in the past among which mechanical3 and fluid-flow switches4. Another approach is a gas-gap switch. This type of heat switch consists of two parallel surfaces with a gap in between that can be filled with a gas. The thermal resistance can be regulated by variation of the gas pressure in the gap5. In principle, a pressure increase can be realized by supplying gas from a storage bottle, and a pressure decrease by pumping away the supplied gas. However, a widely used method is the use of a sorption pump that can reversibly vary the pressure in a closed system by ad- or desorbing the gas3,5,6. Dependent on the detailed requirements, different physical or chemical sorbers and gases can be used. Hydrogen gas is often used because of the high thermal conductivity. (Notice the difference between a sorption pump that is used to vary the pressure in the gas gap around a compressor cell, and sorption compressor cells that are used to drive a sorption cooler.) The gas-gap heat switch appears particularly suitable for our application. An ‘internal’ heat switch around a compressor cell can be scaled to very small sizes, with thermal resistances suitable for our application and without adding any thermal mass to the compressor cells. Moreover, sorption pumps can also be scaled to very small sizes and it was reported that high switching speeds can be obtained with it6. In the next sections these aspects will be discussed in more detail. Theory
The process of heat transfer by gases is in the viscous state different from that in the molecular state. In the viscous state the totality of molecules is responsible for the heat transfer, whereas in the molecular state the individual molecules carry the heat from wall to wall7. The transition between both regimes is determined by the Knudsen number, which is the ratio of the
mean free path L and the distance d between the heat exchanging surfaces: Generally, the gas is considered to be in the continuum if Kn < 0.01 and in the molecular regime if Kn > 1. The mean free path can be derived from kinetic theory of gases and equals7
where k is Boltzman’s constant, is the molecule diameter and p is the pressure. The thermal conductivity in the continuum regime can be related to the viscosity and the volumetric specific heat by7: where
is the ratio of the specific heats at constant pressure and volume. The viscosity and the specific heat for a di-atomic gas are given by7:
where is Avogadro’s number, M is the molecular mass and R is the universal gas constant. If the pressure is low enough to be in the molecular regime, the flux of heat between two surfaces equals the flux of molecules on the walls times the amount of energy they transfer from wall to wall per molecule. The heat transfer coefficient h expressed in equals7:
The so-called accomodation coefficient is used in this expression to account for the incomplete energy exchange between a wall and a molecule. An effective accomodation coefficient is used to account for both walls:
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SORPTION CRYOCOOLER DEVELOPMENTS
Figure 3. Theoretical and experimental heat transfer coefficients for hydrogen gas. The theoretical
values are calculated for three different gap widths; the corresponding transition regions are indicated below the curves. The experimental data is discussed later in the text.
where
and
are the accomodation coefficients for both walls, and
and
are the surface
areas. The first expression in Eq. (18) reduces to the second for parallel surfaces with a thin gap
in between and identical accomodation coefficients. Eq. (17) is independent of the wall thickness, since d does not influence the particle flux nor the energy transport per molecule, and proportional to the pressure, since the particle flux is proportional to p. The transition region between the molecular and continuum region covers two decades of the pressure range and may be an important region for the gasgap operation. An expression for the heat transfer coefficient in this region can be derived by using the concept of a temperature jump between the wall and the gas in the presence of a temperature gradient, which reduces the heat transfer by an effective increase of the wall separation8 . Using this method, an expression for the heat transfer coefficient that covers the three regions was derived:
It can be shown that for low pressures Eq. (19) reduces to Eq. (17) and for high pressures it reduces to Eq. (15) divided by the gap distance d. Eq. (19) acts as a useful tool to calculate the gas-gap behaviour under various conditions. Fig. 3 shows the heat transfer coefficient as a function for the hydrogen pressure, for different gap widths and calculated with Eq. (19). These plots are discussed later in the text together with the measurements that are included in the plot.
Limiting ON and OFF thermal resistances
In general, the thermal resistance of a gas-gap heat switch is put together by three thermal resistances in parallel that represent the heat transfer through the gas, radiation through the gap, and parasitic conduction through the construction that maintains the gap. Dependent on the operating conditions, one or more of these resistances dominate the heat-switch thermal resistance. At first, the limiting ON and OFF resistances will be estimated under the condition that the pressure can be adjusted without limitations. Secondly, it is shown what pressures can be obtained with a typical ZrNi hydrogen sorption actuator and what this means for our application. Thermal resistance in the ON-state. For a heat switch with a usefull ON-OFF ratio, the ON resistance will be dominated by the gas conduction and The minimal gas thermal resistance is limited by the continuum region:
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and it follows that this resistance is proportional to the gap width and inversely proportional to
the conductivity of the gas, which is dependent on the kind of gas and the temperature. Equations (13) and (14) can be used to estimate the required pressure to reach this ON resistance. For a sorption cell of 5 cm in length and 5 mm in diameter that we are working on, a gap separation of can easily be realized. If hydrogen gas is used, can be achieved when This value can now be compared with the required in Eq. (6), which equals 24 K/W for and Therefore, the resistance can easily be matched on the requirements and a gas-gap heat switch is not the limiting factor for this case.
Thermal resistance in the OFF-state. The OFF thermal resistance is limited by the parasitic losses: radiation through the gap and conduction through the gap separating
construction. The radiation thermal resistance
can be expressed as
where and are the temperatures of the cold and hot surfaces, is the effective emissivity and is Boltzman’s constant. It is clear that is a strong function of especially because of the non-linear temperature dependence of radiation. can be maximized by application of clean and polished surfaces in the gas gap; in this way can readily be obtained7. If this is required, a further reduction of radiation losses is possible by application of radiation shields in the gap, for instance in the case of very high temperatures; in the case of a miniature gas gap this solution is virtually impossible. The relative importance of the thermal resistance of the gap
separating construction,
depends of the detailed design of the heat switch. It can be
made large by choosing a proper construction and low conductance construction materials, i.e. glass etc. For purpose of this study it is assumed that so that Now the required pressure to make sufficiently low (for instance ) can be calculated by combination of equations (17) and (22). For a compressor cell of 5 mm in diameter and 5 cm in length, is taken for a worst case estimation of the radiation losses, resulting in The pressure should be reduced below Pa to make This thermal resistance can be compared with the required in Eq. (8), and it follows that the radiation causes a 5% loss term. ON-OFF ratio. The limiting on/off ratio for the above mentioned conditions can be calculated by dividing equations (22) and (21). The ratio is inversely proportional to the gap width, and equals about 650 for It can be concluded that especially for thin gaps very high ratios can be obtained, of course under the condition that the required high and low hydrogen pressures can be supplied and that the conduction losses of the gap maintaining
construction are reduced below the radiation losses. Pressure actuation. Many different hydrogen chemical absorbers exist with different absorption isotherms and other properties such as degradation in time9. An important parameter for selection is the so-called equilibrium plateau-pressure for the lowest temperature that is available to cool the hydrogen actuator. This pressure determines the minimal OFF heat transfer rate of the gas gap. From a comparison it follows that ZrNi has a minimal pressure of about 0.5 Pa at 300 K that causes a heat transfer that is just below the parasitic radiation losses, see Eq. (22). High pressures of Pa can easily be reached by heating the material to somewhat below 200 °C, which means that for most gap widths the continuum conduction region can be reached. However, this hydride material is very sensitive to contamination and the speed of switching to the OFF state is limited because the material has to cool to a temperature very close to the heat-sink temperature. Clearly, this is a field for further research.
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SORPTION CRYOCOOLER DEVELOPMENTS
CHARACTERISATION OF GAS-GAP STATIC THERMAL BEHAVIOUR
To validate the gas-gap model, the heat transfer coefficient h was measured in a test setup as a function of the pressure. If a temperature difference is maintained across a gap between two parallel surfaces and the heat that flows through the gap is also measured, then h can be
calculated from where A is the surface area of the gap. The experimental setup is given in Fig. 4a and a detailed cross section of the experimental gas-gap construction is given in Fig. 4b. It consists of two cylindrical copper parts that fit into each other with a gap of in between. The inner cylinder measures 1 cm in diameter and 2 cm in length and is suspended via two thin glass tubes (outer diameter: ) to obtain a high thermal resistance for the gap separating link. The glass rods are mounted in a spoke construction that is attached to the outer cylinder; the rods are free to slide in the inner cylinder to account for thermal expansion effects. The spoke constructions on both sides are also used to adjust the gap width around the cylinder. A small ceramic heating resistance is mounted inside the inner cylinder and two thermocouples are mounted directly below both surfaces of the gap. The pressure in the gap can be regulated via the combination of a gas supply and a two stage vacuum pump, with two adjustable valves incorporated in the lines. The pressure was measured with a two stage membrane pressure transducer, fabricated by MKS Instruments10. If a measured input power P is supplied to the heater and thermal equilibrium is established (i.e. the temperatures are stabilized), then an ‘effective’ heat transfer coefficient h can be calculated if the temperature difference
is also measured. These measurements can be done for different pressures yielding h(p), which is an effective value because the loss terms are included. Measurements were done for hydrogen and nitrogen gas. The results for hydrogen are included in Fig. 3. The results for nitrogen showed a similar behaviour, except for the maximum heat transfer at high pressures, which was about a factor 7 lower in comparison to hydrogen. This corresponds to literature values. It can clearly be observed from the measurements that the minimum effective h at low pressures is limited by the losses, and that the maximum h at high pressures is limited by the gas conduction in the continuum region. From the measured data also
a transfer coefficient was calculated which only accounts for the heat transfer through the gas. This was achieved by separately measuring the heat flow losses by pumping the gap to a high vacuum; the losses were measured as a function of to account for the non-linear temperature dependence of the radiation. The corrected transfer coefficient is obtained by subtracting the losses from the measured heat flows, and is included in Fig. 3. The theoretical h in the molecular regime (see Eq. (19)) was fitted to these corrected data by adjusting the accomodation coefficient; was found for hydrogen and for nitrogen. Values reported in the literature are 0.3 - 0.7 for hydrogen and 0.6 – 0.9 for nitrogen, both around 300 K, and the lower values for clean metallic surfaces. Dirty unpolished surfaces with possible adsorbed layers of
Figure 4. (a) Experimental set-up to characterize gas-gap behavior. (b) Detailed cross-section.
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other gases in our experimental setup may explain the high values for that were found. In the transition region the measured data for hydrogen deviates from the modelled curve, to a maximum of about 35%; for nitrogen the measured data fitted the model much better. In both cases the deviation was within the range of the measuring accuracy. This accuracy is relatively low at the higher pressures because of the low temperature differences. Furthermore, from the measurements it can be concluded that an ON-OFF ratio of about 170 was obtained for this simple gas-gap configuration. Reduction of the severe parasitic losses would increase this ratio dramatically. From the curves it can also be concluded that it is attractive to use a smaller gap than strictly required for the desired ON-resistance. In this way much lower pressures are needed to obtain the required ON-resistance. GAS-GAP DYNAMIC BEHAVIOUR
If the gas-gap heat switch is actuated with a sorption actuator, there will be a finite delay between the moment the sorption actuator is switched ON or OFF and the actual switching of the gas gap between both states. Several factors can be distinguished that may influence this dynamic behaviour and these are briefly discussed below. 1. At first there is the thermal system as drawn in Fig. 5 that determines the heating and cooling behaviour of the sorption actuator. In this system represents the heat capacity of the sorption material and its ‘holder’ including a heating facility; a fixed thermal resistance between the sorber unit and the heat sink; and the input power required to maintain the high temperature of the sorption unit relative to the heat sink. The thermal resistance must fulfill two conflicting requirements. Firstly it determines together with the thermal mass the time constant that characterizes the cooling (switching) speed after switching OFF the heater of the sorber unit. Here,
is proportional to the amount of hydride that is required to
actuate the gas gap, which amount is proportional to the volume that has to be filled with gas times the maximum pressure that is required in the gap. Secondly, the resistance should limit the heater input power to a reasonably low value while in the ON state. In our case of a cell with and these conflicting requirements can be matched, but it leads to a
very small device with a very large required thermal resistance to the heat sink. Micromechanical techniques are perfectly suited to create a device of such dimensions1. 2. ZrNi is one of the typical absorber materials that can be used for the actuator. When this material is activated it consists of small grains that should be kept together in a small container or so. During hydriding and dehydriding the grains expand and contract several percents, which makes fixing of the grains to the container impossible. Moreover, contamination issues make fixing even more difficult11. As a consequence, a thermal resistance is present between the container and the grains, which is dependent on the gas pressure around the grains. Especially at low gas pressures when the switch is going to the OFF state, this thermal resistance can become very high – thus severely limiting the cooling down speed. To solve this dynamic problem for small devices we propose to use a thin film hydride actuator. 3. In general it is assumed that hydrogen absorption in hydrides consists of three important steps: adsorption of H2 at the surface and formation of H-atoms, diffusion of H-atoms into the bulk of the hydride, and formation of a ‘stable’ bond between the host metal and the H-atoms11. All these processes will require some time, but not much information is available on this topic. What can be said is that by increasing the active surface of the hydride relative to the volume, the time delays will reduce. This can be achieved very well by application of thin film techniques.
Figure 5. Thermal system that determines the heating and cooling behavior of the sorption actuator.
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SORPTION CRYOCOOLER DEVELOPMENTS
4. In a practical implementation of a sorption compressor cell with integrated gas gap, the gas volume that has to be pumped out consists of the gas-gap volume plus the two large void volumes at both ends of the sorption cylinder that contain the compressor support structures. If the gas-gap actuator is located in one of these volumes, the pumping speed of the gas gap is limited by the large amount of gas that is located in the volume at the other side of the compressor cell and that has to be pumped through the gap to or from the actuator. This will especially be a limiting factor when the actuator is switched OFF and the gas has to be pumped out of the volume when the flow is in the molecular flow regime. For our specific situation, a pumping speed of about 2 s was found which approaches the critical limit that was set. However, this pumping speed can easily be increased by including a ‘bypass’ pumping channel with large diameter between both volumes at the ends of the compressor cell in parallel to the gas gap. CONCLUSIONS
From the discussion of the heat-switch requirements it follows that the required ON and OFF resistances can be related to the important compressor parameters by simple expressions. It can also be concluded that a sorption compressor can be operated without heat switches at all, with the penalty of a power loss. These extra losses can be reduced by operating a sorption compressor with more than four compressor cells. Furthermore, the heat transfer behaviour of a gas-gap heat switch in the molecular, transition and continuum regime can be described with one closed expression, and fair agreement was obtained with experimental results. It can be
concluded that a gas-gap heat switch fits the heat-switch requirements for a miniature sorption cooler very well.
ACKNOWLEDGEMENTS
This research is supported by the Dutch Technology Foundation (STW). The authors acknowledge the contributions of JPL, especially Bob Bowman and Larry Wade. REFERENCES
1.
Burger, J.F., ter Brake, H.J.M., Elwenspoek, M., Rogalla, H., Microcooling: Study on the application of micromechanical techniques, Cryocoolers 9, Plenum Press, New York (1997), pp. 687-696. 2. Burger, J.F., Holland, H.J., Wade, L.A., ter Brake, H.J.M., Rogalla, H., Thermodynamic considerations on a microminiature sorption cooler, Cryocoolers 10, Plenum Press, New York (1998). 3. Chan, C.K., Self-actuated heat switches for redundant cryocoolers, Proc. 2nd Interagency Meeting on Cryocoolers, Easton MD (1986). 4. Bard, S., Jones, J.A., Regenerative sorption compressors fro cryogenic refrigeration, Advances in cryogenic engineering, Vol. 35, Plenum Press, New York (1990). 5. Johnson, D.L., Wu, J.J., Feasibility demonstration of a thermal switch for dual temperature IR focal plane cooling, Cryocoolers 9, Plenum Press, New York (1997). 6. Kashani, A., Helvensteijn, B.P.M., McCormack, F.J., Spivak, A.L., Helium liquid- and gas-gap heat switches, Proc. 7th Cryocooler Conf. (1992), pp. 355-370. 7. Roth, A., Vacuum technology, Elsevier, Amsterdam (1990). 8. Corruccini, R.J., Gaseous heat conduction at low pressures and temperatures, Vacuum (1959). 9. Wade, L.A., Performance, reliability, and life of hydride compressor components for 10 to 30 K. sorption Cryocoolers, Adv. in Cryogenic Eng., vol. 39, Plenum Press, (1994), pp. 1483-1490. 10. MKS Instruments, Inc., Six Shattuck Road, Andover, MA 01810. 11. Bowman, R.C., Personal communication (1998).
Development of a High Efficiency 0.5 W Class 4K GM Cryocooler T. Satoh1, R. Li1, H. Asami1, Y. Kanazawa1, and A. Onishi2 1
Research & Development Center, Sumitomo Heavy Industries, Ltd. Kanagawa, 254-0806, JAPAN 2 Precision Products Division, Sumitomo Heavy Industries, Ltd. Tokyo, 188-0001, JAPAN
ABSTRACT The effect of displacer stroke on the performance of a 4K Gifford-McMahon (G-M) cryocooler has been investigated. Displacer strokes of 30mm, 25mm, 20mm and 15mm are examined. The largest cooling capacity at 4.2K and also at 7K is nearly equal for all the strokes, but the first stage cooling capacity at 50K is much improved by optimization of the stroke. The optimum stroke is 20mm from the experiment. The cycle speed at which the maximum 4.2K cooling capacity is obtained is different for each stroke, and is smaller when the stroke is longer. The reciprocal of the optimum cycle speed is in proportion to the stroke. The input power dependence of the 4.2K cooling capacity and the COP at 4.2K has also been investigated for strokes of 30mm, 25mm and 20mm to improve the efficiency of the cryocooler. The experiment is carried out at the optimum cycle speed for each stroke. The 4.2K cooling capacity and the COP at 4.2K has basically the same input power dependence for all strokes. The largest COP at 4.2K is obtained at an input power of about 2 kW. INTRODUCTION The performance of a 4K G-M cryocooler has been much improved by developing magnetic regenerator materials with larger heat capacity1,2 and optimizing the intake/exhaust valve3, etc. A 4K G-M cryocooler with a cooling capacity of 1.5W at 4.2K has been developed4 and now a cryocooler with 1W at 4.2K is commercially available. A 4K G-M cryocooler has a large potential application market. In recent years, 4K G-M cryocoolers have been widely applied to many uses, such as helium recondensation in MRIs5, cooling of cryogen-free superconducting magnets6, SIS mixer cooling7 for radio astronomy, and so on. Not only the cooling capacity at 4.2K, but also the efficiency of a 4K G-M cryocooler is a very important factor for practical use, and a cryocooler with much higher efficiency is desired. The authors have investigated the effect of displacer stroke on the performance of a 4K G-M cryocooler. Expansion volume and displacer cycle speed are dependent on the stroke. The expansion volume is related to the PV work at the cooling space, and the displacer cycle speed is related to heat exchange between regenerator materials and helium. Thus, the displacer stroke is expected to have a strong effect on the 4.2K cooling capacity.
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GM REFRIGERATORS AND LOW-TEMPERATURE REGENERATORS
The relationship between compressor input power and the 4.2K cooling capacity has also been investigated. The 4.2K cooling capacity is thought to become larger when the compressor input power is increased. On the other hand, the efficiency of the cryocooler is expected to become largest at a certain input power level. EXPERIMENTAL APPARATUS
A two-stage type cold head was used in this experiment. The second regenerator is a hybrid type and is composed of lead spheres and spheres. Lead (150g) was stuffed into the higher temperature region, and into the lower temperature region of the second displacer. The first regenerator is composed of copper screens in the higher temperature region and lead spheres in the lower temperature region. The first and the second cylinder diameters are 52mm and 25mm, respectively. The authors obtained a cooling capacity of 1W at 4.2K by using a second cylinder with an ID of 35mm. Then the second cylinder inner diameter for the experiment was chosen so that its cross sectional area is one half of that for the 1W cooler — because the authors’ target cooling capacity is 0.5W. The first cylinder size was selected so that the inner diameter ratio of the first and second cylinder is the same as that for the cylinder of the cold head with cooling capacity 1.5W at A set of optimized intake/exhaust rotary valves that was obtained from a different experiment was installed. A germanium resistance thermometer was installed to measure the second stage temperature, as was a platinum-cobalt alloy resistance sensor to measure the first stage temperature. Electric heaters were installed to measure the cooling capacity of each cooling stage. The displacers were driven by an AC synchronous motor, and the displacer speed was varied by changing the supply frequency. RESULTS AND DISCUSSIONS
Effect of displacer stroke The effect of displacer stroke on cryocooler performance was investigated using strokes of 30mm, 25mm, 20mm and 15mm; the respective expansion volumes were and This experiment was carried out with a helium compressor from Sumitomo Heavy Industries, Ltd., mold CKW21. The input power was about 2.6 kW, and the charge pressure was the same for all strokes. Figure 1 shows the cycle speed dependence on the no-load first and second stage temperatures. The minimum first stage temperature is dependent on the stroke. The lowest minimum first stage temperature is 34.7K at 72 and 78 rpm at a stroke of 25mm. For all the strokes tested, the second stage temperature decreased when the cycle speed decreased. The second stage temperature was minimum at 42 rpm for a stroke of 30mm, but the minimum is not shown for other strokes in this cycle speed range. The lowest temperature is about 2.3K, independent of the stroke. The shorter the stroke is, the wider the cycle speed range is for second stage temperatures below 3K. As is shown in Figure 1, the second stage temperature is lower than 2.6K for cycle speeds between 42 and 144 rpm when the stroke is 15mm, i.e., the cycle speed dependence of the second stage temperature is very small. The cooling capacity at 4.2K is shown in Figure 2. The first stage temperature is fixed at 50K by a heater in this experiment. The cycle speed dependence curve of the cooling capacity at 4.2K becomes steeper as the stroke is increased. This tendency is related to the cycle speed dependence of the no-load second stage temperature. The cycle speed at which the cooling capacity is maximum becomes higher when the stroke is shorter. The maximum cooling capacity is nearly the same for all strokes8. The cooling capacity at 7K was also measured as shown in Figure 3. In this test a stroke of 15mm was not examined. Also at 7K, the maximum cooling capacity is the same for all strokes. Figure 4 shows the first stage cooling capacity at 50K when the second stage temperature is
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Figure 1. Cycle speed dependence of no-load first and second stage temperature.
Figure 2. Cycle speed dependence of 4.2K cooling capacity.
4.2K. The first stage cooling capacity has a maximum at some cycle speed as does the second stage. But the difference in the maximum values is much larger than that of the 4.2K cooling capacity. The maximum cooling capacities are summarized in Figure 5. The cycle speed at which the cooling capacity is maximum is noted near each point. The maximum first stage cooling capacity for a stroke of 25mm is more than twice as large as that for a stroke of 15mm. On the other hand, the 4.2K cooling capacity is improved only about 5%.
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GM REFRIGERATORS AND LOW-TEMPERATURE REGENERATORS
Figure 3. Cycle speed dependence of 7K cooling capacity.
Figure 4. Cycle speed dependence of first stage cooling capacity at 50K.
Figure 5. Stroke dependence of maximum cooling capacity at 4.2K and 50K.
Figure 6. Stroke dependence of a reciprocal of cycle speed at which a maximum 4.2K cooling capacity is obtained.
The reciprocal of the cycle speed at which the 4.2K cooling capacity is maximum is shown in Figure 6, and has a linear dependence on the stroke. As shown in the same figure, the same dependence is also shown for the 1.5W class cryocooler This figure shows that when the 4.2K cooling capacity is maximum, the product of cycle speed and stroke has the same value for every stroke. The PV work per unit time and the displacer speed are proportional to the product of cycle speed and stroke. The cooling capacity results from the PV work. Thus, the linear dependence of Figure 6 seems to mean that the PV work at the maximum 4.2K cooling capacity condition is independent of stroke. On the other hand, the displacer speed is closely related to heat exchange between the flowing helium and the regenerator material in the second displacer. Therefore, the linear dependence of Figure 6 seems to be also related with the heat exchange phenomena. The largest 4.2K cooling capacity, 0.76W, is obtained when the stroke is 20mm, the cycle speed 84 rpm, and the first stage temperature 50K. The measured input power is 2.61 kW.
DEVELOPMENT OF 0.5W CLASS 4K GM CRYOCOOLER
Figure 7. Input power dependence of COP at 4.2K and 4.2K cooling capacity.
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Figure 8. Pressure ratio dependence of COP at 4.2K and 4.2K cooling capacity.
Effect of Input Power on the COP
The input power dependence of the 4.2K cooling capacity and the COP was also investigated to improve the efficiency. In this experiment, three different compressors including CKW21 were used. The compressor capacities other than CKW21 are 2 kW and 5 kW, respectively. The input power is regulated by changing the charge pressure and/or opening the bypass valve between the supply and return helium lines connecting the compressor to the cold head. The helium flow rate and the room temperature pressure of the first cylinder were also measured in this
experiment. The experiment was carried out at strokes of 30mm, 25mm and 20mm. The cycle speed was fixed at the value at which the maximum 4.2K cooling capacity was obtained, i.e., 54 rpm for a 30mm stroke, 66 rpm for a 25mm stroke, and 84 rpm for a 20mm stroke. The 4.2K cooling capacity and the COP at 4.2K is shown in Figure 7. The input power in this figure is calculated for the isothermal condition from the measured flow rate and the pressure ratio of the highest and the lowest measured pressures. The compressor efficiency is 0.49. Figure 7 includes the experimental data for the systems with the three compressors. The experimental data obtained with different compressors are smoothly connected. Figure 7 shows that all the data plot on nearly the same line independent of stroke. This means that the 4.2K cooling
capacity and the COP at 4.2K are not affected by stroke and that the maximum 4.2K cooling capacity is nearly the same for every stroke. From Figure 7, the COP is largest at an input power of about 2 kW for the system. The effect of the stoke on the 4.2K cooling capacity and the COP at 4.2K is more clearly shown in Figure 8. This figure plots the COP value and the 4.2K cooling capacity directly against the pressure ratio. Though the 4.2K cooling capacity increases monotonically as the input power increases, the COP at 4.2K has a peak at a certain pressure ratio. The pressure ratio of the peak is different for each stroke and is larger when the stroke is larger. The largest COP at 4.2K is 3.65x10-4 and is obtained at a stroke of 25mm, a cycle speed of 66 rpm, a pressure ratio of 3.00, a cooling capacity of 0.70W, and an input power of 1.92 kW. The largest cooling capacity at 4.2K is 1.01W at a stroke of 20mm and an input power of 4.66 kW. CONCLUSIONS
The effect of the displacer stroke on the performance of a 4K G-M cryocooler has been investigated. The first stage cooling capacity is much improved by optimization of the stroke, though the 4.2K cooling capacity is only slightly improved. The optimum stroke is 20mm and the
optimum cycle speed is 84 rpm in this study.
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The reciprocal of the cycle speed at which the 4.2K cooling capacity is maximum is proportional to the stroke. This phenomena seems to be explained by the PV work and heat exchange in the second regenerator. The effect of the compressor input power on the 4.2K cooling capacity and the COP at 4.2K has been investigated to improve the cryocooler efficiency. The COP at 4.2K has the same dependence on the input power for all the examined strokes, and is maximum when the compressor input power is about 2 kW independent of the stroke. REFERENCES 1. G. Ke, H. Makuuchi, T. Hashimoto, A. Onishi, R. Li, T. Satoh and Y. Kanazawa, “Improvement of two-stage GM refrigerator performance using a hybrid regenerator,” Advances in Cryogenic Engineering, vol. 40, Plenum Press, New York (1994), pp. 639-647. 2. A. Onishi, R. Li, T. Satoh, Y. Kanazawa, H. Makuuchi, S. Aikawa and T. Hashimoto”, A 4K-GM Cryocooler with Hybrid Regenerator of Magnetic Materials,” Proceedings of Fourth Joint SinoJapanese Seminar on Small Refrigerator, (1993), pp. 44-48. 3. R. Li, A. Onishi, T. Satoh and Y. Kanazawa, “Influence of valve open timing and interval on performance of 4K Gifford-McMahon cryocooler,” Advances in Cryogenic Engineering, vol. 41, Plenum Press, New York (1996), pp. 1601-1607. 4. T. Satoh, A. Onishi, R. Li, H. Asami and Y. Kanazawa, “Development of 1.5W 5K G-M cryocooler with magnetic regenerator material,” Advances in Cryogenic Engineering, vol. 41, Plenum Press, New York (1996), pp. 1631-1637. 5.
M. Nagao, T. Inaguchi, H. Yoshimura, S. Nakamura, T. Yamada, T. Matsumoto, S. Nakagawa,
K. Moritsu and T. Watanabe, “4K three-stage Gifford-McMahon cycle refrigerator for MRI magnet,” Advances in Cryogenic Engineering, vol. 39, Plenum Pressure, New York (1994), pp. 13271334.
6.
K. Watanabe, S. Awaji, T. Fukase, Y. Yamada, J. Sakuraba, F. Hata, C.K. Chong, T. Hasebe and M. Ishihara. “Liquid helium-free superconducting magnet and their applications,” Cryogenics, vol.
7.
M. Takahashi, H. Hatakeyama, T. Kuriyama, H. Nakagome, R. Kawabe, H. Iwashita, G. McCulloch, K. Shibata and S. Ukita, “A compact 150 GHz SIS receiver cooled by a 4K GM refrigerator,” Proceedings of the 7th International Cryocooler Conference, Air Force Phillips Laboratory Report
8.
T. Kuriyama, Y. Ohtani, M. Takahashi, H. Nakagome, H. Nitta, T. Tsukagoshi, A. Yoshida and T. Hashimoto, “Optimization of operational parameters for a 4K-GM refrigerator,” Advances in Cryogenic Engineering, vol. 41, Plenum Press, New York (1996), pp. 1615-1622.
34, 15th ICEC Supplement, 1994, pp. 639-642.
PL-CP-93-1001, Kirtland AFB, NM (1993), pp. 495-507.
Development of a High Efficiency 4K GM Refrigerator Y. Ohtani, H. Hatakeyama, H. Nakagome Toshiba Corporation Kawasaki, Japan, 210-0083 T. Usami, T. Okamura, and S. Kabashima Tokyo Institute of Technology Yokohama, Japan, 226-0000
ABSTRACT
This paper describes performance measurements made on a 4 K Gifford-McMahon (GM) refrigerator that incorporates a magnetic regenerator material. The method used to achieve high efficiency was optimization of the pressure ratio. Three types of cold heads were tested: Type #1 with a 32-mm inner diameter of the second stage cylinder, Type #2 with an inner diameter of
40 mm, and Type #3 with an inner diameter of 56 mm. The second stage regenerator materials were and lead. The pressure ratio was changed by using several types of compressors and regulating the opening of a bypass valve that was connected between the intake and exhaust gas lines from the compressor. The compressor work was estimated by measuring the intake and exhaust pressure and mass flow rate. The highest measured Coefficient of Performance (COP) at for Type #1, for Type #2, and for Type #3; these were obtained at pressure ratios of 2.21, 2.38 and 2.23, respectively. At these pressure ratios, the cooling power at 4.2 K was 0.335 W, 1.01 W, and 1.99 W, and the compressor work was 0.78 kW, 2.20 kW, and 3.34 kW, respectively. A maximum cooling power of 3.68 W was obtained at 4.2 K using the Type #3 cold head at a high pressure ratio of 3.25, and a power of 7.99 kW.
INTRODUCTION
A 4K GM refrigerator that uses magnetic regenerator materials,1,2 has been an essential item for many superconducting magnet systems. For example, in conduction cooled superconducting magnet systems, ease of operation of high magnetic field has been realized by using a compact 4K GM refrigerator to cool superconducting coils of NbTi or Nb3Sn. A 1 W class 4K-GM refrigerator, which has a 32-mm inner diameter second cylinder, has been used to conduction cool superconducting magnet systems as reported previously.3,4 The 10 T class superconducting magnet had a bore diameter of 100 mm. In the future, demand will be there to cool larger magnet systems or to increase the ramp rate for generating the magnetic field. For future systems, the heat load for the 4 K refrigeration stage will similarly increase and the capacity of the 4K-GM refrigerator will have to increase to over 1 watt. 4K-GM refrigerators of this size and larger have been investigated by several researchers.5,6,7 There are several methods to increase refrigeration capacity. One way is to increase the Cryocoolers 10, edited by R. G. Ross, Jr. Kluwer Academic/Plenum Publishers, 1999
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operating pressure difference. Second way is to extend the expansion volume. Yet another way is to increase the refrigeration efficiency. GM refrigerator has a theoretical loss compared with Carnot cycle cooler. This loss is dependent on pressure ratio, and the higher the pressure ratio, the larger this loss. So we decreased the operational pressure ratio compared with the previous operations and optimized the pressure ratio for maximum COP. We also extended the second stage cylinder diameter to
increase cooling power, and also adopted the low pressure ratio operation to obtain high efficient and large cooling power at 4.2 K.
EXPERIMENTAL APPARATUS Cold Head Three types of cold heads of GM refrigerator were tested in this study. Table 1 shows the parameters of the cold heads. Type #1 is the smallest cold head. The inner diameter of the second cylinder is 32 mm. Stroke of displacer is 20 mm. We used magnetic regenerator
materials of
and
in this refrigerator as the second stage regenerator, and already
obtained more than 1 W cooling power at 4.2 K. But in this study to increase COP we re-tested this cold head. The inner diameter of the cylinder of Type #2 and Type #3 are 40 mm and 56
mm, respectively and displacer stroke of these two cold heads are both 32 mm. Figure 1 shows the picture of Type #3 cold head.
Figure 1. Cold head of 4K GM refrigerator (Type #3).
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Figure 2. Configuration of 4K GM refrigerator.
Compressor Figure 2 shows the experimental configuration of 4K GM refrigerator. The pressure ratio was changed by using several types of compressors and regulating an opening of a bypass valve, which was connected between intake and exhaust gas line from the compressor. The compressor
work was estimated by measuring the intake and exhaust pressures and mass flow rate in this experiment. We calculated iso-thermal compressor work and assumed that the compression efficiency was constant value at 0.48. A compressor input power in this work was defined as measured iso-thermal compressor work divided by constant compression efficiency. So the
results of this work did not depend on the difference of efficiency of each compressor. PERFORMANCE OF COLD HEAD TYPE #1
To investigate the performance of 4K GM refrigerator depended on the operating pressure ratio, Type #1 cold head, which has a 32 mm inner diameter of second cylinder, was tested. Figure 3 shows the experimental results of the cooling power at 4.2 K depended on the pressure ratio.
Figure 3. Pressure ratio dependence of cooling power.
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GM REFRIGERATORS AND LOW-TEMPERATURE REGENERATORS
Figure 4. Pressure ratio dependence of COP.
Two operating frequency of 26 rpm and 50 rpm were tested in this experiment. The cooling powers were increasing by increasing the pressure ratio in each frequency. About two times large cooling power was obtained at 50 rpm operation compared with 26 rpm at same pressure
condition. But COP, which is the cooling power divided by the compressor input power, was large at 26 rpm compared with 50 rpm shown in Figure 4. The COP value at 26 rpm operation was increased by decreasing the pressure ratio and maximum COP of was obtained. At each operation frequency, the cooling powers were 0 W at pressure ratio of about 1.7 and COP were 0 at this condition. Figure 5 and Figure 6 show the cooling power and COP at 4.2K versus the compressor input power, respectively. At the same pressure ratio condition, the compressor input power at 26 rpm was smaller than that at 50 rpm because of its small mass flow rate. The cooling power at 26 rpm was larger in the region below about 1.5 kW input power. The COP at 26 rpm was also higher than that at 50 rpm in the same input power region. Therefore Type #1 cold head is
suitable for small input power operation by low operation frequency. The minimum compressor input power to reach 4.2 K were 0.41 kW at 26 rpm and 0.93 kW at 50 rpm.
Figure 5. Compressor input power dependence of cooling power.
Figure 6. Compressor input power dependence of COP.
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Figure 7. Pressure ratio dependence of cooling power of Type #2 and #3 cold head.
PERFORMANCE OF COLD HEAD TYPE #2 AND TYPE #3
Cold heads Type #2 (40 mm cylinder diameter) and Type #3 (56 mm) were also tested to obtain larger cooling power than Type #1. In this experiment the operation frequency was constant of 35 rpm and the performances of these two cold heads were compared. Figure 7 shows the cooling powers of Type #2 and Type #3 cold head as a function of the pressure ratio. At the same pressure condition, more than two times large cooling power was obtained by using Type #3 because of about two time large expansion volume (see table 1). Figure 8 shows the COP related to the pressure ratio. In all pressure ratio regions, the COP of Type #3 was higher than that of Type #2. Type #3 cold head has small ratio of first / second expansion volume, so the efficiency of the second stage at 4.2 K was higher than that of Type #2. However, the efficiency of the first stage was smaller. The cooling power and COP dependence of compressor input power were shown in Figure 9 and Figure 10. Type #3 cold head is suitable for large input power operation above 2 kW compared with Type #2 (and also Type #1). Maximum cooling power of 3.52 W at 4.2 K was obtained at the 3.7 pressure ratio and 7.8 kW compressor input power. We also optimized the operation frequency for larger cooling power and 3.68 W at 4.2 K was obtained at 42.8 rpm using 7.99 kW compressor input power.
Figure 8. Pressure ratio dependence of COP of Type #2 and #3 cold head.
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GM REFRIGERATORS AND LOW-TEMPERATURE REGENERATORS
Figure 9. Compressor input power dependence of cooling power of Type #2 and #3.
Figure 10. compressor input power dependence of COP of Type #2 and #3.
CONCLUSION
To increase COP at 4.2 K of a GM refrigerator, we optimized operational pressure ratio of a cold head of the GM refrigerator and obtained about two or three times larger COP than previous operation. Three types of the cold heads were tested. Inner diameters of the second cylinders are 32 mm (Type #1), 40mm (Type #2) and 56 mm (Type #3). Cooling powers and compressor input powers of the refrigerators at their maximum COP operation were 0.335 W at 0.78 kW using Type #1 cold head, 1.01 W at 2.20 kW using Type #2 and 1.99 W at 3.34 kW using Type #3. The optimized pressure ratio of each cold head was about 2.2 to 2.4. Further, maximum cooling power at 4.2 K was 3.68 W using Type #3 cold head at an operation frequency of 42.8 rpm and a pressure ratio of 3.25 and a compressor input power of 7.99 kW. REFERENCES 1. T. Kuriyama et al., Adv. Cryog. Eng., Vol.41:1615(1996). 2. T. Tukagoshi et al., Adv. Cryog. Eng., Vol.41:1615(1996). 3. Y. Ohtani et al., Proc. of the 16th Intl. Cryog. Eng. Conf., 1113(1995). 4. S. Mine et al., presented at Cryog. Eng. Conf., (1997). 5. A. Ohnishi et al., Proc. of the 16th Intl. Cryog. Eng. Conf., 351(1995).
6. T. Inaguti et al., Proc. of the 16th Intl. Cryog. Eng. Conf., 335(1995). 7. J. N. Chafe et al., Proc. of the 9th Intl. Cryocooler Conf. (1995).
Analysis of a High Efficiency 4K GM Refrigerator Operating at a Lower Pressure Ratio T. Usami, T. Okamura, S. Kabashima Y. Ohtani*, H. Hatakeyama* and H. Nakagome* Tokyo Institute of Technology, Yokohama 226, Japan *Toshiba Corporation, Kawasaki 210, Japan
ABSTRACT
Recently, regenerative refrigerators such as Gifford-McMahon (GM) refrigerators have achieved liquid helium temperature levels using magnetic regenerator materials that have a much larger specific heat capacity below 10K than conventional second regenerator materials of lead. In this investigation a high efficiency 4K GM refrigerator using magnetic regenerator materials was developed and investigated. A refrigeration capacity of 2.04W at 4.2K was obtained for 3.4 kW of compressor input power by optimization of the operating pressure ratio. A maximum
coefficient of performance (COP) of
was achieved at 4.2K at a pressure ratio of 2.43 and an operating frequency of 30 rpm. In the investigation the refrigerator was operated at
various pressure ratios, and at each pressure ratio measurements were made of compressor work, of pressure-volume (P-V) characteristics of the second expansion volume, and of refrigeration capacity at 4.2K. Refrigeration losses were estimated at 4.2K using the experimental results and are discussed. Operating frequency was also optimized to maximize second stage refrigeration performance. INTRODUCTION
Recently, a 4K GM refrigerator using magnetic regenerator materials was put to practical use cooling a superconducting magnet.1,2 The efficiency of the 4K GM refrigerator, however, was lower than that of a commercial 20K GM refrigerator because helium behaves as a non-ideal gas at 4K. Therefore, the running cost of the 4K GM refrigerator was too high, and an improvement in the refrigeration efficiency is needed. In this paper, the effect of operating pressure ratio and frequency on the performance of a 4K GM refrigerator is experimentally investigated. Two kinds of efficiencies are discussed in each of the experimental conditions: 1) the P-V work divided by compressor input power, and 2) the cooling power divided by the P-V work. Additionally, a refrigeration loss is estimated to evaluate the performance of the refrigerator. EXPERIMENTAL APPARATUS
Figure 1 shows a schematic diagram of the experimental apparatus of the two stage 4K GM refrigerator. Inner diameters of first and second cylinder were 90 mm and 56 mm, respectively. Cryocoolers 10, edited by R. G. Ross, Jr. Kluwer Academic/Plenum Publishers, 1999
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GM REFRIGERATORS AND LOW-TEMPERATURE REGENERATORS
Figure 1. Schematic diagram of experimental apparatus.
In the high temperature region, the first regenerator used Cu mesh (#180); Pb shot (0.2-0.3 mm)
was used in the low temperature region. The second regenerator used Pb shot in the middle temperature region, and the magnetic regenerator materials, and in the low temperature region. Stroke of the displacer was 32 mm and periodic motion of the displacer was measured by a displacement gauge. The displacement and pressure were both measured at room temperature and displayed on an oscilloscope to obtain P-V work in the expansion volume. Four types of compressors listed in Table 1 were used for the series of experiments. These compressors are referred to as A, B, C and D in a descending order of input power. A real compressor, however, has its own efficiency of compression that is affected by its operating conditions; as a result, electrical input power to the compressor also changes. Not to be affected by the differences in the efficiency among the compressors, the compressor work was estimated from the isothermal compression work using the measured values of the operation pressure ratio and mass flow rate of the working fluid. The compressor efficiency was assumed to be constant at 0.48. Operating frequencies were 30 rpm, 35 rpm and 40 rpm. In this study, the first refrigeration stage was kept at the no-heat-load condition. REFRIGERATION EFFICIENCY
The theoretical coefficient of performance (COP) of a GM refrigerator is expressed as follows:
where
The COP of a real refrigerator is the ratio of the refrigeration capacity Q to the compressor input power i.e.
ANALYSIS OF 4K REFRIGERATOR AT LOWER PRESSURE RATIO
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or
Q,
and later section.
were obtained by experiment. COP is separated into two kinds of efficiencies which were experimentally obtained. These two efficiencies are discussed in a
EXPERIMENTAL RESULTS
Figure 2 shows the experimental results of refrigeration capacity versus operating pressure
ratio. Operating frequencies of 30 rpm, 35 rpm and 40 rpm were tested. To change the operating pressure ratio, four types of compressors, listed in Table 1, were used. In Figure 2, a set of data points that are circled used the compressor. When the operating frequency with a given compressor increases, mass flow rate generally increases and the pressure ratio decreases. For each frequency, the cooling capacity increases as the pressure ratio increases. The cooling power at
high pressure ratios was the highest for high frequency operation, when compressor A with the largest input power was used. Using compressor A, the maximum cooling capacity at 4.2K was 3.64W at 40 rpm. On the other hand, at low pressure ratios, the cooling power was the highest at low frequency when compressor D with the smallest input power was used. The maximum cooling capacity using Compressor D was 1.35W at 30 rpm. Next, Figure 3 shows the pressure ratio dependence of COP. In each operating frequency, the COP values were increasing as pressure ratio decreases, and maximum COP values were obtained at pressure ratios between about 2 and 2.5. The maximum values of the COP were
Figure 2. Pressure ratio dependence of refrigeration capacity.
Figure 3. Pressure ratio dependence of Coef. of Performance (COP).
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GM REFRIGERATORS AND LOW-TEMPERATURE REGENERATORS
Figure 4. Pressure ratio dependence of and
Figure 5. Pressure ratio dependence of
at 30 rpm, at 35 rpm, and at 40 rpm. Below these pressure ratios, the COP steeply decreases. The COP values at different frequencies were almost the same using compressor A, but when small compressors were used, the COP at low-frequency operation was higher than that at high frequency. DISCUSSION
In this section the cooling efficiency is discussed with respect to the experimental results and calculated losses. In Figure 4, the theoretical of the GM refrigerator and experimental measurements of were compared as a function of pressure ratio. increases as the pressure ratio decreases and is equal to the Carnot efficiency at the pressure ratio of 1. On the other hand, also increased as the pressure ratio decreased. as a function of pressure ratio was found to be independent of the operating frequency. The difference between and was the influence of the non-ideal nature of helium gas at 4.2K and compressor efficiency. Figure 5 shows the pressure ratio dependence of the In the case of operating frequency, above a pressure ratio of 2.5, there was only a small change in values, but it steeply decreased below a pressure ratio of 2.5. Over the complete range of pressure ratios, was higher at lower frequency. Figure 6 shows the pressure ratio dependence on first stage temperature of the refrigerator with no heat load applied to the first stage. Above 2.5 pressure ratio, the temperatures of the first stage stayed at about 35K for each frequency. But below the pressure ratio of 2.5, temperature
Figure 6. Pressure ratio dependence of 1st stage temperature.
ANALYSIS OF 4K REFRIGERATOR AT LOWER PRESSURE RATIO
Figure 7. Pressure ratio dependence of total loss shuttle and conduction.
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Figure 8. Pressure ratio dependence of regenerative loss.
steeply increased as pressure ratio decreased. And the increase of the temperature at high frequencies was higher than that at low frequency. Heat losses from the first stage to the second
stage at 4.2K increased at lower pressure ratios. The principal heat losses are the shuttle loss, conduction loss, and regenerative loss. Radiation loss can be neglected. The shuttle loss and conduction loss depend on the temperature of the first stage. These two losses were calculated at each experimental condition3 and the total of these losses is shown in Figure 7. This total loss depends on the first stage temperature and was independent of the operating frequency and pressure ratio. As a result, the loss increased below a pressure ratio of 2.5. Next the regenerative loss is estimated in Figure 8. The refrigeration loss was estimated as the total of the regenerative loss and above two losses, and is equal to the difference between and the experimentally determined cooling capacity. The regenerative loss generally depends on the mass flow rate, so the higher the operation frequency, the larger the loss. This loss also increases as pressure ratio increases because of the increase of mass flow rate. The regenerative loss was 3 or 5 times larger than the total of shuttle and conduction losses when pressure ratios were high. However, at low pressure ratios, the shuttle and conduction losses are almost equal to the regenerative loss. Both the net cooling capacity and
the regenerative loss decreased as pressure ratio decreased, but the shuttle and conduction losses did not decrease. Thus, the ratio of the shuttle and conduction losses to the total loss increased. This is why the COP value fell steeply at low pressure ratios. CONCLUSION
To obtain a higher efficiency 4K GM refrigerator, the effect of pressure ratio and frequency on the cooling capacity and efficiency has been experimentally investigated. The maximum
cooling capacity at 4.2K was 3.64W at a frequency of 40 rpm and a compressor input power of 8.0 kW. The maximum COP was when the frequency was 30 rpm and the pressure ratio was 2.43. In this optimized condition, the cooling capacity was 2.04 W at 4.2K and the compressor input power was 3.4 kW. One reason for the increased COP was that the
increased as the pressure ratio decreased in the same manner as the theoretical COP of a GM refrigerator. Another reason was that the did not appreciably change for pressure ratios higher than about 2.5. Refrigeration losses have been examined and the shuttle and conduction loss at low pressure ratios is considered to be the cause of the sharp decrease in COP and REFERENCES 1.
T. Inaguchi et al., “Effect of thermal conductance of cooling stage in 4K-GM cryocooler on cooling capacity”, Adv. in Cryogenic Engineering, Plenum Press, New York (1998).
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GM REFRIGERATORS AND LOW-TEMPERATURE REGENERATORS
2.
Y. Ohtani et al., “Development of a 11.5T liquid helium-free superconducting magnet system,” Adv. in Cryogenic Engineering, Plenum Press, New York (1996), pp. 1113-1116.
3.
M. Thirumaleshwar, et al., “Gifford-McMahon cycle - a theoretical analysis” Cryogenics (1986) Vol. 26, pp. 177-188.
Numerical Simulation of 4K GM Refrigerator T. Inaguchi, M. Nagao, K. Naka, and H. Yoshimura Mitsubishi Electric Corporation, Advanced Technology R&D Center 8-1-1, Tsukaguchi-Honmachi, Amagasaki, Hyogo, 661-8661 Japan
ABSTRACT
This paper describes a simulation model of a 4K-GM refrigerator. The simulation model is composed of a regenerator, an expansion space, and a cooling stage. The volumetric change of the expansion space and the actual thermophysical properties of the working fluid, helium are considered. The basic equations are made up of one-dimensional fluid equations and an energy equation of the regenerator material and the cooling stage. The fluid equations are expressed in the general coordinate system of which a coordinate axis moves with time to take account of the volumetric change of the expansion space. These basic equations are differentiated by using the TVD MacCormack method. In order to verify the simulation model, the calculation results were
compared with results of experiments and the theoretical validity of the calculation results was checked. As a result it was confirmed that the simulation model is appropriate. INTRODUCTION
Although GM refrigerators traditionally have only achieved no-load temperatures of not lower than approximately 10 K, the use of magnetic regenerator materials has allowed them to reach temperatures lower than 4.2 K, where helium can be liquefied1. The refrigeration capacity
at 4.2 K has improved continuously, and now surpasses The operating mechanism within the GM refrigerator, however, has not yet been completely
clarified. Researchers have principally depended on experiments to determine optimal values. To improve the efficiency of the GM refrigerator even further and optimize its design, it is necessary to make its mechanism clear and obtain optimal values using an analytical method. In previous analyses of the GM refrigerator, Daney et al.3 and Matsubara et al.4 considered the dead volume of the regenerator to evaluate its effect on regenerator efficiency. Seshake et al.5 focused on the temperature dependence of the regenerator material to investigate the effect of the material's specific heat on regenerator efficiency. Tominaga6 studied the heat flow in the regenerator from a thermoacoustic viewpoint and discussed the efficiency of the regenerator. Kurihara et al.7 made a model of the whole GM refrigerator system, including the compressor, to determine its refrigeration capacity. These investigations primarily concerned the analysis of the regenerator's behavior. They either ignored the expansion space or treated it as a space with uniform physical characteristics. If the expansion space is taken into account, however, because of its volumetric change, the fluid equations are changed. The authors of this paper express the fluid equations in a general coordinate system whose axes change with time. This allows us to use the same equations to formulate both the dead volume of the regenerator and the expansion space. It enables us to solve a simulation model consisting of a regenerator (including dead volume), an expansion space and a cooling stage by using the finite difference method (TVD MacCormack method).
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The new simulation model has two striking features: because it takes the volumetric change of the expansion space into account, it makes it possible to accurately calculate the fluid flows in the regenerator and expansion space, heat transfer, and the refrigeration capacity. In addition, the use of the actual thermophysical properties of the working fluid (helium) allows the calculation of the refrigeration capacity of a 4K-GM refrigerator. This paper will present this simulation model and its formulation as well as a discussion of the calculated results.
Main Symbols : heat transfer area at the bottom of the cooling stage : flow passage area
P : pressure Pr : Prandtl number
Q : heat transfer from the wall to the fluid
: heat transfer area of the gap between the
: heat transfer in the hot part of the regenerator (per unit length of ) : heat transfer in the cold part of the regenerator (per unit length of ) : heat transfer in the central part of the
displacer bottom and cooling stage : sectional area
: heat transfer area on the side wall of the cooling stage : heat transfer area between the wall and fluid : sectional area of the wall
regenerator (per unit length of Re : Reynolds' number : hydraulic radius
: specific heat of the wall per unit mass
T : fluid temperature
: diameter of the displacer bottom : inner diameter of expansion space
: wall temperature t : time
: total energy per unit mass e : internal energy per unit mass f : friction coefficient H : total energy flux per unit time
h : heat transfer coefficient : heat transfer coefficient of expansion space
: heat transfer coefficient of the gap between the displacer bottom and cooling stage : flow passage length
: wall length : mass flow in the hot part of the regenerator : mass flow in the cold part of the
regenerator : mass flow in the central part of the regenerator
)
u: velocity : velocity of the end of expansion space x : x -coordinate : porosity : viscosity coefficient : heat load on the cooling stage
: density : wall density
:
-coordinate
Suffices hend : hot end of regenerator lend : cold end of regenerator 0 : value at hot end of regenerator
CALCULATION METHOD Simulation Model
Figure 1 shows a schematic representation of the two-stage GM refrigerator. In the present analysis, a model of the 2nd regenerator, 2nd expansion space, and 2nd cooling stage was created to clarify the operating mechanism of the expander which operates at about 4 K. Figure 2 shows the simulation model. It consists of a regenerator, an expansion space and a cooling stage. The regenerator comprises the dead volume, which is used as a fluid passage, and the regenerator material. The expansion space changes in volume periodically. As the boundary conditions, the pressure and temperature are given at the high-temperature end of the regenerator, and the condition that no fluid passes through the end of the expansion space in either direction are given. The temperatures, pressures and velocities of the fluid in the dead volume of the regenerator and the expansion space are calculated, as well as the temperatures of the regenerator material and cooling stage. The fluid in the dead volume of the regenerator and the expansion space exchange heat with the regenerator material and the cooling stage, respectively. The cooling stage is subjected to an external heat load, which is transferred via the cooling stage to the fluid in the expansion stage.
NUMERICAL SIMULATION OF 4K GM REFRIGERATOR
Fig. 1 Schematic diagram of GM refrigerator.
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Fig.2 Simulation model.
If we use the term "wall" for the regenerator material and the cooling stage, the simulation model analyzes the heat exchanges between the fluid and the wall. It should be noted that this fluid is subjected to periodic changes in pressure and flow velocity. These periodic variations are caused by the boundary conditions set at the high-temperature end of the regenerator and volumetric changes in the expansion space. The following are assumed for simulation: (1) The fluid and the wall are both one-dimensional,
(2) The heat transfer in the fluid and the wall (regenerator material and cooling stage) in the flow direction is negligible.
(3) The regenerator is thermally insulated, it is subjected to no external heat load. Assumptions (2) and (3) are appropriate, considering the structure of the actual refrigerator. Assumption (1) means that the sectional areas of the regenerator's dead volume and the expansion space must be the same. If they are made the same, however, the volume of the expansion space will deviate from the actual value. Thus the volume of the expansion space was made equal to its actual volume by changing the stroke. Because the simulation model doesn’t move the regenerator and let the volume of the expansion space change, it doesn't include the shuttle loss caused in the gap between the displacer and the cylinder. The shuttle loss was calculated separately together with heat conduction losses of the cylinder and the displacer, and was added as heat loss in the expansion space.
Basic Equations The basic equations consist of a fluid equation and a wall energy equation. The fluid equation is a one-dimensional Euler's equation with additional friction and heat transfer between the fluid and wall. In the Cartesian coordinate system, it is expressed by:
where U, E, S are the vectors shown below:
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GM REFRIGERATORS AND LOW-TEMPERATURE REGENERATORS
To reflect the expansion space's change with time, the coordinates (t, x ) of the fluid equation (1) are converted into Using these relations,
Eq. (1) becomes Eq. (2) in a conservation form:
The elements of equation (2) are as follows:
Temperature T and pressure P can be obtained from the state equation of helium, Eq (4). In order to precisely consider the thermophysical properties of helium, a helium thermophysical properties program9 was actually used for this purpose. The energy equation of the wall is given by:
For the specific heat of the wall,
, the actual specific heat of the regenerator material and
cooling stage were used. Equations (3) to (5) are the basic equations. Figure 3 shows the computational mesh of the fluid. Figure 3(a) represents the computational mesh of the fluid in the (t, x) coordinate system. The expansion space changes its volume with time. Figure 3(b) represents the computational mesh in the coordinate system. can be taken arbitrarily as long as (t, x) corresponds to one-to-one. Here the system is taken so that the computational mesh width is 1 over the entire fluid region. Therefore the volume
of the expansion space does not change in the
coordinate system.
Equations (3) and (5) are finite-differentiated by the TVD MacCormack method10 and solved under specified boundary conditions. The first term, of the left hand side of Eq. (3) was calculated to satisfy the geometric conservation law11.
Heat Transfer from the Wall to the Fluid and Friction Coefficient Heat transfer from the regenerator material to the fluid and from the cooling stage to the fluid were calculated as follows:
Fig. 3 Computational mesh of fluid.
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(a) Heat transfer from the regenerator material to the fluid
Heat transfer Q from the regenerator material to the fluid is expressed by equation (6): Equation (7) 12 was used to express the heat transfer coefficient of the regenerator:
(b) Heat transfer from the cooling stage to the fluid An enlarged view of the cooling stage is shown in Fig. 1. After passing through the regenerator, the fluid passes through the gap (Region D in Fig. 1) between the displacer bottom and cooling
stage, reaching the expansion space. Region D and the expansion space exhibit different heat transfer coefficients because they have different hydraulic radii. The hydraulic radius of the expansion space was expressed as and the hydraulic radius of region D was expressed as The cooling stage is made of oxygen-free copper and oxygen-free copper has a sufficiently good thermal conductivity at about 4 K. Therefore the temperature across the thickness of the cooling stage is assumed to be uniform. The heat transfer Q from the cooling stage to the fluid is expressed by equation (8).
where is the heat transfer coefficient at the expansion space and is that of region D. Both are expressed by Eq. (9). When Reynold's number is greater than 3000, the equation employs the Petukhov-Gnielinski equation13, which expresses the heat transfer coefficient of turbulent flow in a cylinder. When Reynold's number is not greater than 3000, it uses the equation14 of the heat transfer coefficient of laminar flow in a cylinder.
The friction coefficient in the regenertor is estimated by data obtained by Kays and London15. The friction coefficient between the cooling stage and fluid is expressed by the Blasius equation13 when Reynold's number is greater than 3000. Otherwise it is expressed by the
equation16 of friction coefficient of laminar flow in a cylinder: Boundary Conditions
The boundary conditions at the high-temperature of the regenerator and the end of the expansion space are given as follows: (a) Boundary conditions at the high-temperature end of the regenerator
(b) Boundary conditions at the end of the expansion space
In these equations,
is an arbitrary value,
is an arbitrary value given as a function of time.
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SIMULATION RESULTS
Parameters of Simulation Model Table 1 lists the main parameters of the simulation model. The parameter values correspond to the GM refrigerator which has a refrigeration capacity of 3 W at 4.2 K2. The high-temperature side of the regenerator is filled with globes at a mass ratio of 74 %, while the lowtemperature side is filled with at a mass ratio of 26 %. The specific heats of these regenerator materials are shown in Fig. 4. The flow passage area of the regenerator is and the sectional area of the expansion space is 28.3 cm2. A stroke value of 9.8 cm is used to equalize the expansion space's volume in the actual apparatus and model. The dead space in the expansion space is 10 % of the expansion space. On the basis of experimental results, the boundary value of the temperature at the high-temperature end is set to 45 K. The boundary value of the pressure at the high-temperature end is given by data measured in the room-temperature space. Figure 5 shows the pressure boundary condition at the high-temperature end and the stroke of the expansion space. The shuttle loss was calculated using equation17 by Zimmerman et al. The shuttle loss and heat conduction loss in the cylinder and displacer total 0.5 W. These losses are considered as heat losses in the expansion space. The regenerator and the expansion space include 18 and 5 computational meshes, respectively. The time step is set so that Courant number is 0.9. The convergence condition is that the relative error between the cooling stage temperature in the current and previous cycle is in the order of This is an appropriate condition. For example, if the cycle frequency is 41 rpm and the operating temperature is 4 K, the condition means that further operation of 1 minute leads to a temperature change of less than 0.0016 K.
Fig. 4 Specific heat of
and
Fig. 5 Boundary condition of pressure and stroke of expansion space.
NUMERICAL SIMULATION OF 4K GM REFRIGERATOR
Fig. 6 Comparison between calculations and experiments about refrigeration capacity.
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Fig. 7 Comparison between calculations and experiments about temperature of fluid at the low end of regenerator.
Comparison of Simulated and Experimental Results Figure 6 compares the simulated and experimental values of the refrigeration capacity. The noload temperature is 2.79 K in the simulation and 2.72 K in the experiment, the difference being only 0.07 K. Temperature under a 3 W heat-load is 4.40 K in the simulation and 4.23 K in the experiment, the difference being 0.17 K. The maximum difference between the simulation and experiment under heat loads of 0 to 5 W is 0.17 K. This reveals that the simulated results agree
well with the experimental results.
Figure 7 compares the simulated and experimental values of the fluid temperature at the lowtemperature end of the regenerator. The fluid temperature was measured by inserting a CGR temperature sensor into the low-temperature end. The temperature sensor was sandwiched between felt matting to avoid direct contact with the regenerator material. The simulated results agree well with the experimental results, although the experimental values are slightly higher because the installation of the CGR sensor led to an increase in the heat load. Discussion of Simulated Results
To verify the theoretical validity of the simulated results, we integrated some calculated values over 1 cycle and examined them. In Fig. 8 and onwards, the simulated cooling stage is subject to a heat load of 3 W.
Fig. 8 Mass flux.
Fig. 9 Heat transfer from wall to fluid.
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Figure 8 shows the mass flux. The horizontal axis represents the coordinate. The ranges from 3 to 20 and from 21 to 25 correspond to the regenerator and expansion space, respectively. The broken line indicates an integration over 1 cycle. It is zero because mass flux is periodically stationary. The solid line indicates the integration of mass flux of inflow alone. The mass flux of inflow alone varies according to the position; the figure shows that it is greatest at the lowtemperature end of the regenerator. This is the effect of the dead volume of the regenerator. In the expansion space, the mass flux of inflow alone drops as the position comes closer to the end and vanishes there. This agrees with the assumption of the simulation model because it indicates that no fluid passes through the end of the expansion space in either direction. Figure 9 shows the heat transfer from the wall to the fluid per unit length of .. The solid line is the integration of the positive values alone, representing the quantity of heat exchange between the fluid and regenerator material. Although most of the heat is exchanged at the hightemperature side, a small peak can be seen around Around this point the regenerator material changes from to The rise in heat exchange can be explained by the fact that the specific heat of has a peak around this temperature. This accounts for the experimental results indicating that using a regenerator material with a greater specific heat at the low-temperature side increases the refrigeration capacity18,19. The broken line in Fig. 9 shows the integration of the heat transfer from the wall to the fluid over 1 cycle. The region corresponding to the regenerator, which is assumed to be insulated, exhibits no heat transfer. In the expansion space, it takes positive values because that region is under a heat load. The results agree with the assumptions of the simulation model. Figure 10 shows the integration of total energy flux over 1 cycle. To obtain the total energy flux, in energy equation of Eq. (3) the quantity in parentheses in the second term on the left side is multiplied by the flow passage area. The region for the regenerator, which is assumed to be insulated, shows a constant total energy flux of 92.8 J/cycle. In the region for the expansion space, which is under a heat load of 5.2 J/cycle, the total energy flux increases, reaching 98.0 J/cycle at the end of the expansion space. Figure 11 shows a schematic representation of the results shown in Fig. 10. Let us consider the meaning of total energy flux at the end of the expansion space, . From the boundary condition (11), Therefore the total energy flux at the end is: The integration of
over 1 cycle equals the indicated work. Actually, indicated work is
98.0 J/cycle, which is equal to the integration of over 1 cycle. Therefore, it can be said that the refrigeration capacity is equal to the difference obtained by subtracting from the indicated work the total energy flux flowing into the expansion space. Figures 10 and 11 reveal that the law of energy conservation is met over the entire range, and that verifies the simulation model.
Fig. 10 Total energy flux.
Fig. 11 Energy balance in regenerator and
expansion space.
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Fig. 12 Temperature distribution of fluid and wall.
The discussions in the previous section and this section show that the simulated results agree very well with the experimental results and that the simulated results are theoretically valid. It is therefore concluded that the simulation model is appropriate. The following section presents time variations in temperature, flux, heat transfer and so on. Time Variations in Other Quantities
Figure 12 shows the temperature distributions in the fluid and the wall. Curves
to
represent the results obtained at five points equally dividing 1 cycle. The divisions are shown in Fig. 5, Curves to are in the intake process, while curves to in the exhaust
process. Around the center of the regenerator, both processes show great temperature fluctuations of up to approximately 30 K. The low-temperature end of the regenerator and the expansion space exhibit small temperature fluctuations. In addition, the temperature differences between the fluid and the wall are small.
Fig. 13 Change of heat transfer and mass flux.
Fig. 14 Particle paths of fluid.
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Figure 13 shows time variations in heat transfer and mass flux from the wall to the fluid. The graphs represent, from the top, the high-temperature part, the central part and low-temperature part of the regenerator. The mass flux takes positive values when the fluid flows into the regenerator. The point of time at which heat transfer reaches a maximum varies depending on the position of the regenerator. In the process of outflow from a regenerator, the heat transfer at the lowtemperature part of the regenerator reaches a maximum the moment the fluid flows out, while heat transfer at the central part reaches a maximum slightly later. At the high-temperature part of the regenerator, the heat transfer reaches a maximum upon completion of the process of outflow from the regenerator. In the process of inflow into the regenerator, conversely, the heat transfer at the high-temperature part of the regenerator reaches a maximum the moment the fluid flows in, while heat transfer at the low-temperature part reaches a maximum upon completion of the process of inflow into the regenerator. These differences in the time of maximum heat transfer are caused by the heat capacity of the regenerator material. Figure 14 shows the particle paths of fluid. The horizontal axis represents the x coordinate. The region from 0 to 0.15 meters corresponds to the regenerator, and the upper region corresponds to the expansion space. The thick line shows the end of the expansion space. The fluid particles are divided into four groups according to their behavior: (1) Particles in group A flow out of the regenerator immediately they flow into it; (2) particles in group B oscillate in the regenerator after they flow into it; (3) particles in groups C and D oscillate in the regenerator and expansion space. In addition, there are particles which flow out of the regenerator after oscillating in it (not shown). Fluid particles travel over great distances in the regenerator. Particles in group B travel over 83 % of the regenerator length and those in group C over 69 %. CONCLUSION
A simulation model of a 4K-GM refrigerator has been developed to clarify the operating mechanism and accurately calculate the refrigeration capacity. The model is composed of a regenerator, expansion space and a cooling stage. In formulating the model, the fluid equations were described using a generic coordinate system whose axes vary with time. It was found that simulated results agreed very well with the experimental results. An examination of the simulated results established the theoretical validity and effectiveness of the simulation model. REFERENCES 1. H. Yoshimura, M. Nagao, T. Inaguchi, T. Yamada, M. Iwamoto, Rev. Sci. Instrum., 60 (1989), 3533-3536. 2. T. Inaguchi, M. Nagao, K. Naka, H. Yoshimura, submitted to Adv. Cryog. Eng.(1997). 3. D. E. Daney and R. Radebough, Cryogenics, 24 (1984), 499-501. 4. Y. Matsubara and Y. Hiresaki, Cryocooler 6, (1990), 173-182. 5. H. Seshake, T. Eda, K. Matsumoto, and T. Hashimoto, Adv. Cryog. Eng., 37(1992), 995-1001. 6. A. Tominaga, Cryogenic Engineering, 26 (1991), 30-36. 7. T. Kurihara and S. Fujimoto, Cryogenic Engineering, 31(1996), 197-202. 8. D. A. Anderson, J. C. Tannehill, and R. H. Pletcher, Computational Fluid Mechanics and Heat Transfer, (1984), 422, Hemisphere. 9. CRYODATA INC., HEPAK Ver.3.30, (1994). 10. H. C. Yee, NASA TM 89464, (1987). 11. P. D. Thomas and C. K. Lombard, AIAA Journal, 17 (1979) 1030-1037. 12. W. M. Kays and A. L. London, Compact Heat Exchanger 3rd ed., (1984), 150, McGraw-Hill. 13. JSME Heat Transfer Handbook, (1993), 74, Maruzen. 14. JSME Heat Transfer Handbook, (1993), 68, Maruzen. 15. W. M. Kays and A. L. London, Compact Heat Exchanger 3rd ed., (1984), 213, McGraw-Hill. 16. JSME Data Book: Hydraulic Losses in Pipes and Ducts, (1993), 22, Maruzen. 17. F. J. Zimmerman and R. C. Longsworth, Adv. Cryog. Eng., 16 (1970), 342-351. 18. A. Onishi, R. Li, H. Asami, T. Satho and Y. Kanazawa, Cryogenic Engineering, 31(1996), 162-167. 19. T. Inaguchi, M. Nagao, K. Naka, and H. Yoshimura, Cryocooler 9, (1997), 617-626.
Numerical Fluid Analysis of Pumping Loss K. Naka, T. Inaguchi, M. Nagao and H. Yoshimura Mitsubishi Electric Corporation, Advanced Technology R&D Center 8-1-1, Tsukaguchi-Honmachi, Amagasaki, Hyogo 661-8661, Japan
ABSTRACT This paper describes a numerical fluid analysis of pumping loss. As a result of the calculations, it was found that the pumping loss becomes small as the low end temperature of the clearance increases, and the pumping effect begins to contribute to the cooling capacity when the low end temperature increases to more than 215 K. In the clearance, the working fluid discharges heat in the high temperature part and absorbs heat in the low temperature part. It becomes a loss when the heat absorbed in the low temperature part is larger than the heat discharged in the high temperature part, while it contributes to the cooling capacity when the heat discharged is larger than the heat absorbed. We also examined the effects of various parameters on the pumping loss. As a result, we found that the pumping loss increases to a maximum value as the clearance
increases, and that it also increases as the frequency increases.
INTRODUCTION
Pumping loss is considered to be one of the major losses of regenerative cryocoolers with displacers such as GM crypcoolers, Stirling cryocoolers and Vuilleumier cryocoolers. However there has been little analysis of the pumping loss up to now and the mechanism of the pumping loss has not been clarified quantiatively and qualitatively. Pumping loss is generally thought to occur as follows. The working fluid in the low temperature part in the clearance between the cylinder and displacer moves to the high temperature part of the clearance during the compression process, and its temperature rises. The same working fluid then returns to the low temperature part again during the expansion process. However with regard to the regenerator and the clearance in the cryocooler, the clearance could be considered to be a kind of pulse tube and its structure to be similar to a basic pulse tube cryocooler. Thus, in a regenerative cryocooler, the working fluid might discharge heat in the high temperature part of the clearance, and might absorb heat in the low temperature part. According to this way of thinking, a pumping loss might not always occur and there might be a contribution to the cooling capacity. However, the truth of this hypothesis has not been proven. In order to clarify the pumping effect, we constructed a calculation model and calculated the pumping loss using computational fluid dynamics. The calculation model consists of only the
clearance. We supposed the heat capacity of the cylinder and displacer to be infinity and
supposed the wall of the clearance to be an isothermal wall. We input the pressure at the low temperature end of the working fluid in the clearance, and made the boundary condition at the high temperature end of the working fluid a fixed wall. Under these conditions, we calculated the heat loss transferred to the low temperature end, and clarified the mechanism of the pumping effect. We also examined the effect of various parameters on the pumping effect.
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CALCULATION MODEL AND CALCULATION CONDITIONS
A schematic diagram of the regenerative cryocooler and the analysed region are shown in Figure 1. A mesh division of the calculation model is shown in Figure 2. The calculation model consists of only the clearance and is divided into a pie slice with a symmetrical boundary. We supposed the heat capacity of the cylinder and displacer to be infinity and supposed the wall of the clearance to be an isothermal wall. This is an effective supposition in the region where the heat capacity of the working fluid is sufficiently small compared to the heat capacity of the cylinder and displacer. The main parameters of the calculation are shown in Table 1. As the boundary conditions, the low temperature end of the clearance is the pressure boundary, and the cylinder and displacer walls are isothermal walls. The high end temperature of the isothermal wall is 300 K, the low end temperature is 200 K and the cylinder and displacer walls have a linear temperature distribution. In this model, the temperature and velocity distributions in the radial and circumferential directions are neglected. The pressure condition at the pressure boundary is shown in Table 2 and in Figure 3. The pressure changes from 7×105 Pa to 19×105 Pa at 1 Hz. The temperature of the fluid from the pressure boundary is fixed at 200 K. For the initial condition, the high end temperature is 300 K, the low end temperature is 200 K and the fluid has a linear temperature distribution. 1The initial pressure is 13×105 Pa. We used the fluid analysis software Star-cd. We calculated the temperature, velocity, pressure and density distributions and calculated the heat transfer from the wall. The working fluid was helium and the properties of the fluid retained their average temperature values. The heat transfer from the wall to the fluid was calculated by the temperature difference and the distance between the wall and the mesh.
Figure 1. Schematic diagram of regenerative cryocooler and the analysed region.
NUMERICAL FLUID ANALYSIS OF PUMPING LOSS
Figure 2. Mesh division.
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Figure 3. Pressure fluctuation at pressure boundary.
ANALYSIS RESULTS
The convergence of the calculation is shown in Equation 1. This equation shows the ratio of the temperature change for every cycle. When the value of this equation becomes smaller than we judged that the calculation had converged.
The temperature change during the cycle in the axial direction is shown in Figures 4 and 5. This figure shows the relative temperature between the temperature of the fluid and the wall. It
has positive value when the temperature of the fluid is higher than the temperature of the wall. The temperature change in the high temperature part is larger than the change in the low temperature part. The change in the temperature and the heat transfer in the middle section of the mesh for one cycle are shown in Figure 6. This figure also shows the pressure fluctuations. The temperature of this figure is the actual temperature and it is positive when the heat is transferred from the wall to the fluid. From 0 to 0.25 sec., the pressure of the fluid increases, but its
temperature decreases. This is because the heat transfer is positive, in other words, the heat transfers from the fluid to the wall, this decrease in temperature is larger than the increase in
Figure 4. Temperature change during 1 cycle. Figure 5. Temperature change during 1 cycle.
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temperature due to the pressure increase. From 0.25 to 0.5 sec., the temperature of the fluid decreases as the pressure decreases. However heat is transferred from the wall to the fluid. In this way, the change of the temperature and the heat transfer follow behind the pressure change. In this model, the temperature change is behind the pressure change by about 3/4 cycle, and the transferred heat is behind by about 1/4 cycle. The same thing can be said about the mass flux. The change of the mass flux in the central part of the mesh is shown in Figure 7. It is positive when the mass flowing to the high temperature part. From this figure, the change in the mass flux is behind the pressure change by about 3/4 cycle, the same as the temperature change. When comparing this figure with Figure 6, we can understand that when the change in the pressure is the greatest, the mass flux reaches its maximum or minimum. And when the mass flux is at a maximum or minimum, the change in the heat transfer is a minimum or maximum and thus the temperature also reaches a minimum or maximum. The total change in the heat transfer from the wall to the fluid is shown in Figure 8. The
heat transfer follows behind the pressure change. We calculated the total heat transfer for each cycle by integrating the change in heat transfer. As a result, the total heat transfer for one cycle is 0.085 W. On the other hand, the pumping loss is the integrated energy for one cycle at the low temperature end and is usually calculated by Equation 2.
A schematic diagram of the energy balance is shown in Figure 9. The energy at the low temperature end is equal in the whole heat transfer, and the pumping loss is 0.085 W. In order to understand the pumping loss in detail, we calculated the effects of the low end temperature on the pumping loss. In the previous calculation the low end temperature was 200K. In these calculations, only the low end temperature was changed, to 50 K, 100 K, 150 K, 250 K and 300 K. The parameters of these calculations are the same as Table 1 other than the low end temperature. However by changing the low end temperature, we also changed the initial
Figure 6. Change in the temperature and
transferred heat. (Central part of the clearance).
Figure 7. Change of the mass flux .
(Central part of the clearance).
NUMERICAL FLUID ANALYSIS OF PUMPING LOSS
Figure 8. Total heat transfer from wall to fluid.
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Figure 9. Schematic diagram of the energy balance.
temperature, the temperature of the isothermal wall and the temperature of the fluid at the pressure boundary. The relationship between the low end temperature and the pumping loss is shown in Figure 10. For comparison, the pumping loss calculated from an expression of Leo2 is
shown in this figure. The pumping loss decreases as the low end temperature increases. Also the decrement decreases as the low end temperature increases. The tendency of Leo’s equation and the results of the present analysis are similar. However the decrement of the present analysis is larger than of Leo. It should be noted that the pumping effect takes a negative value when the low end temperature increases to more than 215 K. This shows that the pumping effect
contributes to the cooling capacity and is a pumping gain. To understand this phenomenon, the distribution of the heat transfer in the axial direction integrated over one cycle is shown in Figure 11. This figure shows the change of the heat transfer when die low end temperatures are 200 K and 250 K. It is positive when the heat is transferred from the fluid to the wall and it is negative when the heat is transferred from the wall to the fluid From this figure we can see that the working fluid discharges heat in the high temperature part of the clearance and absorbs heat in the low temperature part. This mechanism is similar to the
Figure 10. Relationship between low end temperature and the pumping loss .
Figure 11. Distribution of the heat transferred in the axial direction.
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Figure 12. Relationship between inner diameter of the cylinder and pumping loss.
Figure 13. Relationship between clearance and pumping loss.
Figure 14. Relationship between length in the axial direction and pumping loss.
Figure 15. Relationship between mean pressure and pumping loss.
Figure 16. Relationship between pressure fluctuation and pumping loss .
Figure 17. Relationship between frequency and pumping loss.
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behavior of the pulse tube of basic pulse tube cryocoolers. In the case of 200 K, the integrated
value of this transferred heat takes a positive value and there is a pumping loss, and in the case of
250 K, it takes a negative value and the pumping effect begins to contribute to the cooling capacity. In this way, the mechanism of the pumping effect is that the working fluid discharges heat in the high temperature part and absorbs heat in the low temperature part. This mechanism is
similar to the behavior of the pulse tube of pulse tube cryocoolers, and when the absorbed heat is larger than the discharged heat, there is a pumping loss, however when the discharged heat is
larger than the absorbed heat, there is a pumping gain and the pumping effect begins to contribute to the cooling capacity. EFFECT ON PUMPING LOSS OF VARIOUS PARAMETERS
The eight parameters used by this calculation are shown in Table 1. We examined effects on the pumping loss of all the parameters other than the high and low end temperatures. The following describes each one.
Effect of inner diameter of the cylinder In these calculations, only the inner diameter of the cylinder was changed, to 10.7 mm, 20.7 mm, 40.7 mm and 50.7 mm. The other parameters of the calculation were the same as Table 1. The relationship between inner diameter of the cylinder and the pumping loss is shown in Figure 12. We can see from this figure that when the inner diameter of the cylinder increases, the pumping loss increases in proportion. Effect of clearance In these calculations, only the clearance changed, to 0.3 mm, 0.5 mm and 1.0 mm. The relationship between the clearance and pumping loss is shown in Figure 13. From this figure we can see that the pumping loss becomes larger when clearance increases up to 0.7 mm, but it becomes smaller when clearance increases to more than 0.7 mm. Effect of length of axial direction In this calculation the length in the axial direction were changed to 0.2 m, 0.5 m and 1.0 m. The relationship between the length in the axial direction and the pumping loss is shown in Figure 14. The pumping loss becomes larger when the length in the axial direction increases. Effect of mean pressure In this calculation the mean pressures were changed to 10, 15, 18, 20 and The relationship between the mean pressure and pumping loss is shown in Figure 15. The pumping
loss has no effect on the mean pressure change. Effect of fluctuating pressure
In this calculation, the fluctuating pressures were changed to 2, 4 and
The
relationship between the fluctuating pressure and pumping loss is shown in Figure 16. The pumping loss becomes larger when the fluctuating pressure increases. Effect of frequency In this calculation the frequencies were changed to 1/6 Hz, 1/3 Hz, 1/2 Hz, 2/3 Hz and 4/3 Hz. The relationship between the frequency and pumping loss is shown in Figure 17. The pumping loss becomes larger when the frequency increases.
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CONCLUSION
In order to clarify the mechanism of the pumping loss, we made a calculation model and
calculated the heat loss transferred to the low temperature part. As a result of the calculation, the following points have been clarified. 1. The pumping loss becomes small as the low end temperature increases, and the pumping effect begins to contribute to the cooling capacity when the low end temperature increases to more than 215K. 2. In the clearance, the working fluid discharges heat in the high temperature part of the clearance and absorbs heat in the low temperature part. 3. The pumping loss increases linearly as the inner diameter of the cylinder and the length in the axial direction increases. The pumping loss increases as the frequency increases, but its increments become larger as the frequency increases. 4. The pumping loss increases as the clearance increases, however the pumping loss decreases
when the clearance increases beyond a certain point.
5. The pumping loss hardly changes when the mean pressure increases, but the pumping loss
becomes large as the fluctuation pressure increases. REFERENCES
1. 2.
Computational Dynamics, “Star-cd Ver3.0 Manual” Leo, B., “Vuilleumier Cycle Cryogenic Refrigeration System Technology Report”, AFFDL-
TR-71-85, WPAFB, Dayton, Ohio
Multilayer Magnetic Regenerators with an Optimum Structure around 4.2K H. Nakane1, T. Hashimoto1, M. Okamura2 H. Nakagome3, and Y. Miyata1 1
Kogakuin University Nishi-shinjuku, Shinjuku-ku, Tokyo, 163-8677, Japan 2,3 Toshiba Corporation 2 Sugita-cho, Isogo-ku, Yokohama-shi, Kanagawa, 235, Japan 3 Ukishima-cho, Kawasaki-ku, Kawasaki-shi, Kanagawa, 210, Japan
ABSTRACT In order to obtain high regenerative effectiveness, the heat capacity of the regenerator materials must be larger than that of helium as the working gas. For a magnetic regenerator material to be effective in the low temperature range, its transition temperature must be within the range where the helium regenerative operation is performed. Conventional experimental results indicate that anti-ferromagnetic or ferromagnetic materials with large spin values are suitable. As regenerator materials, sufficient regenerative effectiveness cannot be obtained when only one of these materials or their compounds is packed into the regenerator. In this study, materials with different heat capacity are arranged in multi-layers in the regenerator to obtain higher regenerative efficiency. The most effective multi-layer regenerator for liquefied helium is analyzed by computer simulation and experimental results. New materials, including heavy rare-earth and antimony (Sb) compounds, are used as the magnetic regenerator materials. Antimony compounds have especially sharp single-phase transition in the low temperature range. The relation between refrigeration capacity and regenerative effectiveness was discussed when these multi-layer regenerative materials were packed into the second stage of a GM refrigerator. INTRODUCTION Application of cryocooler-technic using the magneto-thermal effect has mainly been studied as the subject of magnetic refrigerator and magnetic regenerative materials with especial emphasis on regenerative materials to obtain better refrigeration capacity at very low temperature. In a regenerative refrigerator, a high heat-exchange effectiveness (regenerative effectiveness) in the regenerator is an important factor for improving the refrigeration capacity. In 4.2K GM refrigerators, only and gases are used as the working gas. gas is
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generally used since gas is costly. The specific heat peak of pressurized He gas is at temperatures below 10K. The heat capacity of magnetic materials in the regenerator must be larger than that of He gas. The specific heat of Pb, which is conventionally used as the
regenerative material, is near zero below 5K. Pb has no practical value around 4.2K. For a magnetic regenerative material to be effective in the low temperature range, its transition temperature must be within the range where the He regenerative operation is performed. When is packed into the second stage of a GM refrigerator, a refrigeration rate of 0.5~0.8W at 4.2K has been attained. The peak in the specific heat curve of a single magnetic material such as is sharp compared to that of He gas. Since the heat-exchange region of He gas is wide, a single magnetic material can not cover the specific heat peak of He gas. However, various single rare-earth compounds, which have a specific
heat comparable to that of He gas, can be used at the finite temperature range. We propose that magnetic materials with different peak heat capacity be arranged in multi-layers in the regenerator to obtain a heat capacity (heat capacity) comparable to that of He gas over the whole temperature range.1 A computer simulation at 4.2K clearly indicated that the refrigeration capacity of a four-layer regenerator into which ErNi and were inserted in sequence from cold to hot, is expected to be twice that of a single-layer To improve the refrigeration capacity of a multi-layer regenerator, it is necessary to find a magnetic material whose specific heat peak is very large at different temperatures and whose synthetic specific heat is larger than that of He gas. As yet, however, not much research into optimum multi-layer structured regenerators packed with magnetic materials has been carried out. To obtain higher regenerative effectiveness, we became aware of the sharp specific heat peak around (magnetic phase transition temperature) of antimony (Sb) compounds and magnetic material. Sb and rare earth compounds (Ho, Dy and Gd with large magnetic moment) were made and their specific heat was measured. Furthermore, in order to improve refrigeration capacity, the best heat distribution capacity for the regenerative operation in a
regenerator was investigated.
The regenerative effectiveness of magnetic materials in the
multi-layer regenerator was analyzed by computer simulation.
SELECTION OF REGENERATIVE MATERIALS AND MEASUREMENT OF SPECIFIC HEAT
Magnetic regenerative materials having various properties have already been developed. The magnetic specific heat of the materials is given by the entropy change due to the order-disorder magnetic phase transition. The relation between entropy S and specific heat C(T) is obtained by (1), and the integration over the whole temperature range is finite.
It can be seen from the above that the half-width of the specific heat peak is narrow when a very sharp specific heat peak exists at low temperatures. The absolute value of the peak
becomes small when enlargement of the half-width is tried. Fig. 1 clearly shows that the half-width of specific heat of is wide and the absolute peak value is small. The half-width of is narrow but the absolute value is close to that of He gas.
Therefore, neither single magnetic regenerative material can cover the specific heat of pressurized and depressurized He gas over the whole temperature range. A multi-layer
regenerator is essential to compensate for this deficiency in the magnetic materials.3 Fig. 2 is the schematic of a multi-layer regenerator. Magnetic materials having different phase transition temperatures are packed from low to high temperatures. By doing this, it is expected that a heat capacity comparable to that of He gas can be obtained over the whole temperature range.
MULTILAYER MAGNETIC REGENERATORS AROUND 4.2 K
Figure 1. Specific heat of Pb,
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and helium(8 and 20 atm).
Based on the above assumption, let us analyze the appropriate regenerative materials to be packed into a multi-layer regenerator. According to thermodynamics, the specific heat C(T) of the material is given as:
The temperature dependence of C(T) for magnetic materials is non-linear since C(T) has a very sharp peak near Therefore, the best kind of magnetic materials can not be selected
solely on the basis of temperature dependence. On the other hand, according to statistical mechanics, the entropy change due to the phase change from perfect order phase to the disorder phase through Tc is given by:
where k is the Boltzmann constant. From (3), as the condition to obtain large S, the number
Figure 2. Schematic figure of the multi-layer regenerator.
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of spins N and the spin value J must be large. From this, rare-earth compounds such as Er, Ho, Dy and Gd are expected to be effective as multi-layer regenerative materials. Since He gas has a large specific heat below 10K and a gentle decrease above 10K, materials having a specific heat larger than that of He gas below 10K and a wide peak above 10K must be found. In this paper, a very sharp specific heat peak around Tc of Sb compounds and magnetic materials was noted, and Sb and Ho, Dy and Gd, whose magnetic moments are large, were finally selected and their compounds made. The specific heat was measured with a MagLab Heat Capacity instrument made by Oxford Instruments Ltd. based on thermal relaxation method, whose temperature range is 0.5~200K and the applied magnetic field 12T. Since the specific heat was measured by (J/K·g) with this instrument, the value of was obtained by measuring the density. The density measured by the Archimedean principle and obtained from literature4 are shown in Table 1. The results of specific heat, HoSb, DySb and GdSb measured are shown in Fig. 3. Sb and Ho, Sb and Dy compounds, whose local and orbital magnetic moments are large, have a very sharp peak around below 10K. As for the comparison of absolute peak values, the peak value of HoSb at is 6 times that of at and 3 times that of at Sb and Gd compounds, whose local magnetic moment and non-Anisotropy (spin angular moment is maximum and orbital angular moment is zero) are large, manifested a comparably sharp peak at The property of specific heat in Fig. 3 was substituted into (1), and entropy S was obtained.
The temperature dependence of S is
shown in Fig. 4. From the results, these compounds can be expected to be effective as
regenerative materials. Experiments indicated that HoSb must be packed into the cold end, DySb into the intermediate and GdSb into the hot end of the regenerator.
Figure 3. Specific heat of RSb(R = Dy, Gd, Ho) and ErNi.
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Figure 4. Temperature dependence of entropy. COMPUTER SIMULATION
In this computer simulation, the heat-exchange rate of materials in the regenerator is assumed to be constant. A model, whose heat exchange effectiveness is proportional to only the specific heat exclusively, was considered. This computer simulation was done to obtain
information on the most suitable structure of a multi-layer regenerator. Both the properties of specific heat of the regenerative materials and He as the working gas must be considered. The regenerative effectiveness, which means the effectiveness of heat-exchange, is defined as the effectiveness of the enthalpy and it is as follows:5
where and are the actual and ideal enthalpy changes of the gas, respectively. In an ideal regenerator, is 100%. When is close to 100%, the specific heat distribution for the regenerative operation is considered best. As regards the multi-layer structured regenerator, the values of calculated with a computer are compared with the experimental values of the regeneration capacity. Table 2 shows the comparison of the values of obtained by computer simulation and by conventional experiment for the refrigeration capacity at 4.2K6 when the materials in Fig. 5 are packed into the second stage regenerator of a GM refrigerator. It clearly shows that the refrigeration capacity increases according as the value of increases. The refrigeration capacity is considered to be the highest when is the largest. The parameters used in the computer simulation are shown in Table 3. The temperature dependence of the specific heat in Fig. 6 is used. Properties (1)~(3) in Fig. 6 are used for the materials on the low temperature
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Figure 5. Combination of regenerator materials.
Figure 6. Specific heat of magnetic regenerator materials and helium(8atm).
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side and properties (4) and (5) for the materials on the high temperature side of the multi-layer regenerator. Property (6) is for often used in a single-layer regenerator. It is considered that maximum regenerative effectiveness is obtained when the synthetic specific heat of the multi-layer regenerative materials can cover the specific heats of both pressurized and depressurized gases.
SIMULATION RESULT In order to ascertain the effectiveness of a three-layer regenerator packed with the new materials (HoSb, DySb and GdSb) to improve the refrigeration capacity, a comparison of these new materials with conventional materials was carried out by computer simulation as shown in Fig. 6. We investigated the possibility of composing a regenerator which can cover the specific heat range of He gas. A simulation of a two-layer regenerator was carried out with a simple model. For this model, Er compounds with below 10K were used as materials for the low temperature side of the regenerator, and was packed into the high temperature side. Fig. 7 shows the calculated results of maximum regenerative effectiveness obtained while changing the value x, where the rate of materials packed into the low temperature side of the regenerator is x, and the rate at the high temperature side is
Figure 7. Calculated result between the regenerator effectiveness and heat capacity distribution of regenerator type II.
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In this figure, the longitudinal axis shows the regenerative effectiveness and the horizontal axis shows the rate of materials packed into the low temperature side (x). In the case of only the high temperature side was packed, i.e., with is the maximum value of As can be seen from the simulation result in Fig. 7, the effectiveness is higher when a better combination of materials which can cover the specific heat of He gas is achieved. The synthetic specific heat of the two-layer materials in Fig. 7 is higher than that of a one-layer For a three- and a four-layer regenerator, the simulation results between regenerative effectiveness and x is shown in Fig. 8. For the three-layer regenerator, the material for the low temperature side was packed at the ratio of x, (0.3 – x), 0.7, and for the four-layer regenerator at x, (0.25 – x), 0.35, 0.40. The results of a two-layer (Type E in Fig. 7) and a one-layer regenerators are also shown. The simulation clearly indicated that the regenerative effectiveness increases as the number of layers is increased. However, the rate of increase in the regenerative effectiveness decreases as the number of layers is increased. Our conclusion is that perhaps a two- or a three- layer regenerator is the most practical and experiments to determine the best combination of regenerative materials be conducted. For the two-layer regenerator, HoSb and DySb for the low temperature side was packed at the ratio of x and (1.0 – x), respectively. The simulation result is also shown in Fig. 8. The value of can not be evaluated exactly because is close to 1.0. However, it can be deduced that the refrigeration capacity improves remarkably because the regenerative effectiveness of a
Figure 8. Calculated result between the regenerator effectiveness and heat capacity distribution of regenerator type.
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two-layer structure packed with these new materials is larger than that of a four-layer regenerator packed with conventional materials. For three-layer regenerator including additionally GdSb, the regenerative effectiveness will be remarkably improved. The relation between past experimental data on the refrigeration capacity at 4.2K and the regenerative effectiveness obtained by simulation was arranged by the least square method and the result is shown by solid line in Fig. 9. The broken lines show the maximum refrigeration capacity for each layer obtained by simulation. In the experiment6, the GM refrigerator used had a cylinder whose inner diameter at the single stage was 70mm, and at the second stage 32mm. Fig. 9 shows that the refrigeration capacity of a three- and a four-layer regenerator packed with conventional materials is twice as efficient as a one-layer Furthermore, the refrigeration capacity of a regenerator packed with the new materials (HoSb and DySb) is expected to be about two times larger than that of a one-layer regenerator. At present, refrigeration capacity close to 1W is obtained by using only as the regenerative material. A refrigeration capacity over 2W should be achievable if the new materials are used. CONCLUSION
An attempt to develop the most suitable magnetic regenerative materials and a computer simulation of the regenerative effectiveness were carried out for a 4K regenerator. (1) It was found out that Sb compounds and rare-earth materials are effective as suitable regenerative materials, having a very large heat capacity comparable to He gas below 10K. (2) The computer simulation clearly indicated that the regenerative effectiveness increases as the number of layers in the regenerator is increased. (3) The computer simulation to analyze the regenerative effectiveness of a multi-layer regenerator packed with Sb compounds and rare-earth materials showed the possibility of obtaining a very high refrigeration capacity, judging from previous data on experimental refrigeration capacity. There is a possibility of obtaining a refrigeration capacity remarkably higher than that of a one-layer regenerator using only if a multi-layer regenerator packed with Sb compounds and rare-earth materials is made and experiments are carried out. Preparations to start production and experiments are now underway.
Figure 9. Relation between regenerator effectiveness and the refrigeration capacity at 4.2K.
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ACKNOWLEDGMENT
The authors would like to thank Dr. T. Numazawa for his helpful discussion and assistance. REFERENCES
1.
T. Hashimoto et al., "Excellent Character of Multi-Layer Type Magnetic Regenerator near 4.2K", Cryocoolers 8, Plenum Press, New York (1995) pp. 677-683.
2.
Y. Miyata et al., "Optimum Multi-layer Structure of a Regenerator with Magnetic Materials", Proc. of JSJS-5 (1997) pp. 216-221.
3.
T. Hashimoto, "Recent Progress in Magnetic Regenerator Materials and Their Application", Trans. of JAR 10 (1993) p. 357.
4.
M. E. Mullen et al., "Magnetic-ion-lattice interaction: Rare-earth Antimonides", Phys. Rev. B10 (1974) pp. 186-199.
5.
H. Seshake et al., "Analysis of Rare Earth Compound Regenerators Operating at 4K", Adv. Cryog. Eng. 37B, Plenum Press, New York, pp. 995-1001.
6.
T. Hashimoto et al., "Effect of High Entropy Magnetic Regenerator Materials on Power of the GM refrigerator", Adv. Cryog. Eng. 40 (1994) pp. 655-661.
Advances in Neodymium Ribbon Regenerator Materials Thomas Felmley Concurrent Technologies Corporation Johnstown, PA 15904
ABSTRACT
Advances in regenerator reliability, as well as material cost reduction, can be obtained by optimizing the geometry of the regenerator material. This ongoing development effort for <10K Gifford-McMahon (G-M) refrigerators utilizes coiled ribbons of Neodymium (Nd) instead of the typical spheroidal powder. The ribbons, which have small ridges on one face, are coiled to form flat pancakes. The height of the ridge in relation to the thickness of the tape determines the void volume. These flat coils are then stacked to form a cylinder and packed into the second stage of a G-M refrigerator, in the place of the Nd spheres. This paper describes the fabrication process used to create ribbons from the Nd rare-earth material. Although easier to fabricate than intermetallic regenerator materials, Nd work-hardens and oxidizes rapidly. Modifications were made to conventional extrusion, drawing and rolling processes to overcome these difficulties. The production process developed has much greater yields (>80%) than current Nd powder atomization techniques, which yields only 15% to 30%. Test results show that the amount of void volume, and therefore the pressure drop, can easily be controlled and optimized by varying the ribbon geometry. Whereas the void volume of packed Nd powder is normally roughly 35%, ribbon geometries with void volumes of 15% and 20%, were manufactured and tested. Current efforts are described to further improve the geometry and increase cryogenic performance well beyond that of spherical powder materials. INTRODUCTION
Neodymium ribbons were described as a possible alternative to intermetallic spheres as early as 1987 by Green, et al1. However, it was not until recently that ribbons of Nd have been produced in sufficient lengths2 to allow optimization and cryogenic testing. These ribbons, used in the second stage of a Gifford-McMahon (G-M) refrigerator, show promising results, especially when potential cost savings and reliability improvements are considered. The following describes the background, ribbon design, the processing steps, and some pertinent test results, including product characterization and cryogenic performance.
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BACKGROUND
This work supports the efforts of the Advanced Lightweight Influence Sweep System (ALISS) project, managed by Naval Surface Warfare Center (NSWC). ALISS, a magneticinfluence sweep system for shallow water applications, utilizes a superconducting magnet to reduce the size and weight of the system. This will allow the system to be fielded on aircushioned landing craft. The magnet is cooled by two G-M refrigerators. The current system, utilizing two cryocoolers with neodymium spheres in the second stage, is currently undergoing sea-trials. One goal of this project is to increase the cooling power of each refrigerator by thirty percent, by using Nd ribbons. This would allow the sweep system to either reduce weight by using only one cryocooler, or increase ruggedness and redundancy by having a second unit as a back-up.
At the temperature range where the superconducting magnet system operates, below 10K, only a few materials have useful heat capacity. All of these materials, which are intermetallics, are normally only available for this application in the form of spheres, typically about 0.2-mm (0.008-in.) diameter. Packed, uniform spheres approach a theoretical porosity limit of just under thirty percent. In actual use, a well packed regenerator exhibits 32% to 37% porosity.3 A coiled ribbon or tape of a rare earth metal or intermetallic is not limited by the geometric packing efficiency of spheres. By designing the shape and carefully forming the ribbon, the porosity can be set to predetermined levels. By optimizing the geometry of the ribbon, it is theoretically possible to outperform spherical materials, generating the optimum heat transfer area to pressure-drop ratio.3 In addition to performance advantages, Nd in ribbon or tape form offers cost advantages by
increasing product yields. The yield of Nd spheres in the proper size range from the atomization process is only approximately 20 percent. In contrast, from a similar billet starting size, ribbon production yields of 70 to 80 percent are possible. RIBBON DESIGN
The design of the ribbon is governed by two driving concerns: optimize its performance as a regenerator material, and maximize its producibility. Throughout the design process, the product and the process development have been integrated to achieve these goals. This design integration starts with material selection, continues through development of achievable ribbon geometries and includes process development as well. Material Selection. Pure (99.9%) neodymium is used as the regenerator material in the ribbon. Neodymium is not very ductile when compared to many other more common metals, due to its hexagonal-close-packed crystalline structure. Yet, it does offer some ductility and workability, especially when compared to the intermetallic compounds of similar cryogenic behavior, such as which behave like brittle ceramics. There are a variety of challenges to cost-effective ribbon processing with Nd. Neodymium metal is difficult to extract and is costly. The current market price is approximately $250 per kilogram ($114 per pound). In addition, Nd oxidizes in air and rapidly oxidizes in moist environments; Nd is pyrophoric; and the metal work hardens quickly, requiring frequent annealing. Ribbon Geometry. The physical geometry of the ribbon is determined by the desired void volume fraction or porosity. This void volume fraction regulates the pressure drop of the cooling
gas as it flows through the regenerator column. Each ribbon is created to precise tolerances, with ridges on one side distributed at set intervals, perpendicular to the length of the ribbon, as shown in Figure 1. Tapes have been produced and evaluated with ridge heights of 0.025 mm (0.001 in.) and 0.051 mm (0.002 in.) respectively. The tolerances on the ridge height are plus or minus 0.005mm (0.0002 in.). Each finished ribbon is then coiled to form a flat, thin pancake. These pancakes are then stacked in the second stage of the cryocooler. The ridges keep the each individual wrap of the coil separated from its nearest neighbors. The void volume fraction is determined by dividing the height of the ridge by the sum of the ridge height and the thickness of
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Figure 1. Diagram of Ribbon Geometry.
the ribbon. Thus, a ribbon with a ridge height of 0.051 mm (0.002 in.) and a thickness between ridges of 0.254 mm (.010 in.) yields a coil with a theoretical void volume fraction of 16.67%. The geometric variables to be optimized include ribbon thickness, width, ridge-height, and the ridge shape. These factors, as well as tension and coiling method, control the void volume fraction in the coil and directly affect refrigeration capability. PROCESSING
Neodymium is not readily available in either long wire or strip form. Therefore, a processing program was established to produce the ribbons by a sequence of extrusion, drawing and rolling steps. The Nd ribbon is then coiled and inserted into the second-stage regenerator tube. Extrusion. In the current efforts, the starting billet is 2 inches in diameter and 6 inches in length. After initial inspection, machining, and processing, the billet is extruded and drawn by Supercon, Inc. (Shrewsbury, MA). The neodymium is hot extruded in an evacuated copper can, with a niobium layer between the two metals, to prevent inter-diffusion. The composite canister is extruded to 12.7-mm (0.5-in.) diameter. The can prevents oxidation of the Nd metal and provides a good lubricant for drawing. Wire Drawing. After extrusion, the clad Nd is drawn down to a size convenient for rolling to ridged tape. The Nd metal work hardens quickly, and must be annealed in an inert atmosphere several times during the drawing process.4 A wire cross-section of diameter 2-mm (0.08-in.) is shown in Figure 2. Initially, after the drawing process was completed, the copper was etched away in an acid bath. This exposed the Nb layer, which is allowed to remain to prevent oxidation of the Nd. However, continued development has shown that the rolling process benefits as well from copper cladding. Therefore, one aspect of current development work is characterizing the effect of the copper on the cryogenic performance of the ribbons.
Rolling. Rolling is conducted in our facility on a Fenn rolling mill, with digital tension control. The wires, ranging in size from approximately 2.0 mm (0.08 in.) down to 1.15mm (0.045 in.), are rolled to tape in a seven-pass process. The initial rolling passes are aggressive – greater than forty-percent reduction per pass – to take advantage of the relative softness of the annealed ribbons. Smaller reductions, with a finer control of tolerances, follow, with the final pass imparting the ridges to the tape. A special embossed roll was fabricated to create the ridges. Figure 3 shows a transverse cross section of a tape of 0.25-mm (0.010-in) thickness. These tapes have the copper layer intact. The outer Cu layer is clearly visible, though the thin, light-colored Nb layer is much harder to discern. The core, slightly darker, is the gray Nd metal. Figure 4, a longitudinal cross-section at a higher magnification, highlights a single ridge, 0.050 mm (0.002 in.) in height. Copper comprises most of the ridge. The niobium has been thinned in places by the Nd, which roughens as it is worked because of the limited number of slip systems available for deformation.
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Figure 2. Neodymium wire clad with Nb and Cu. 80X.
Figure 3. Transverse section of 0.245-mm (0.010-in.) Nd ribbon. 80X.
As with the wire drawing process, the rapid work-hardening requires frequent annealing to prevent edge cracking and ribbon breakage. Other important factors for tight tolerances and minimal breakage are consistent tension and careful handling. Table 1 shows the matrix of sizes to be produced and evaluated. The sizes in italics have been produced in lengths up to several hundred feet. Coiling and Insertion. After a final annealing step, the ribbons are delivered to NSWC for coiling and cryogenic testing. The NSWC has developed a coiling system that uses adjustable tension clutches to control the tension and enhance uniform flow distribution.5 It has been observed that test coils wound too tightly have areas where the void space has collapsed. Conversely, if the coil is wound too loosely, certain areas of the coil will have larger void volume, leading to preferential flow and reduced cooling effectiveness. Figure 5 depicts a coil wound from 0.254-mm (0.010-in.) ribbon with 0.05-mm (0.002-in.) ridges. The coils are packed into the lower, cold, end of the displacer tube in the second stage of a G-M cooler. The top half, which sees temperature above 10K, is packed with lead spheres. The regenerator used requires coils that are 25 mm (1 in.) in diameter. The coils, 3.2 mm (0.125in.) in width, are stacked in the tube to a height of approximately 50 mm (2 in.). TESTING
To evaluate the quality and performance of the ribbons and coils produced, several metallographic, mechanical testing and cryogenic performance measures are used. Test results from three of these measures, which offer the most critical information, are described here. Image analysis is used to determine the void volumes of the coil. However, the two main performance indices, the pressure drop and the cooling power, are measured after the regenerators are assembled.
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Figure 4. Longitudinal section of 0.254-mm (0.010-in.) ribbon, showing ridge detail. 500X.
Characterization. Some ribbons, once coiled, are returned our facility for evaluation of the void volume using an automated image analysis (IA) system. This system uses an optical
microscope with a computer-controlled stage and a video image capture system. With this system, the overall void volume of a coil can be measured directly. It can also map areas of variation in void volume across the face of the coil, as shown in Figure 6. This system, once refined, will be able to indicate the percent thickness of the copper and niobium layers. Cryogenic Testing. Performance testing of the coiled ribbons is conducted at the Naval Surface Warfare Center Cryogenics Division (NSWCCD) in Annapolis, MD. They have the ability to measure the pressure drop across the regenerator, as well as the cooling power of the unit. Figures 7 and 8 show the pressure drop and cooling power of regenerators constructed with 0.18-mm (0.007-in.) ribbon and 0.25-mm (0.010-in.) ribbon. Data for Figure 8 was compiled with no heat-load on the first stage of the G-M cryocooler.5 CONCLUSION
Neodymium ribbons are now being produced in sufficient lengths to allow thorough cryogenic evaluation. The work presented here is part of an ongoing effort to optimize the properties of the ribbons and their production process. This production process will be scaled up to produce the tapes with the best performance and the technology developed will be shared with industry.
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Figure 5. Coiled ribbon, made from clad, 0.254-mm (0.010 in.) Nd ribbon. 3X.
Figure 6. Image analysis of 0.254 mm (0.010 in.) thick ribbon coil. Grayscale indicates density.
NEODYMIUM RIBBON REGENERATOR MATERIALS
Figure 7. Pressure drop of two different coil thicknesses versus a standard regenerator.
Figure 8. Second-stage cooling power of various regenerator configurations.
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ACKNOWLEDGMENT This work was conducted by the National Center for Excellence in Metalworking Technology, operated by Concurrent Technologies Corporation under contract No. N00140-92C-BC49 to the U.S. Navy as part of the U.S. Navy Manufacturing Technology Program. We would also like to acknowledge the support of the U.S. Navy Program Executive Office for Mine Warfare, Surface Mine Counter Measure Systems. REFERENCES
1.
Green, G., Patton, W., Stevens, J., Low Temperature Ribbon Regenerator, in Proceedings of The Second Interagency Meeting on Cryocoolers, David Taylor Naval Research and Development Center, Bethesda, MD (1987)
2.
Kajuch, J. and Felmley T., Rapid Response for Fabrication Issues of Neodymium Ribbon Regenerators for use in Cryogenic Refrigerators, Final Report, NCEMT TR No. 97-110, Concurrent Technologies Corporation, Johnstown, PA, July 1997
3.
Ackerman, R. A., Cryogenic Regenerative Heat Exchangers, Plenum Press, New York, (1997), pp. 53-58.
4.
Wong, T., Rudziak, M., “Niobium Clad Neodymium in Wire and Strip Form for Use as a Regenerator”, accepted for publication in Advances in Cryogenic Engineering, vol. 43.
5.
Chafe, J., Green, G., “Neodymium Ribbon Regenerator Cooling Performance in a Two-Stage Gifford-McMahon Refrigerator”, accepted for publication in Advances in Cryogenic Engineering, vol. 43.
Gd-Zn Alloys as Active Magnetic Regenerator Materials for Magnetic Refrigeration V.K. Pecharsky and K.A. Gschneidner, Jr. Ames Laboratory, US DOE, and Department of Materials Science and Engineering, Iowa State University, Ames, IA, U.S.A 50011-3020
ABSTRACT
Stoichiometric GdZn crystallizes in the cubic CsCl-type crystal structure and has been
reported to order ferromagnetically at 270 K. Experimental measurements of the magnetization and heat capacity of GdZn as a function of temperature and magnetic field confirm the ordering temperature and the type of magnetic order. The calculated magnetocaloric effect (in terms of the adiabatic temperature rise, and the isothermal magnetic entropy change, peaks at 270 K and reaches values of 6.5 K (-7 J/kg K) and 10.5 K (-11 J/kg K) for magnetic field changes of 0 to 5 and 0 to 10 T, respectively. The maximum magnetocaloric effect in GdZn is approximately 30% smaller than that observed in pure Gd, which is consistent with the amount of non-magnetic Zn in the intermetallic compound. Modeling of the magnetocaloric effect of different two-phase alloy compositions, including the eutectic composition alloy (50 mol.% GdZn + 50 mol.% Gd), indicates that Gd-Zn alloys with less than 50 at.% (~30 wt.%) Zn can be used as high performance active magnetic regenerator materials. Both Gd and GdZn are magnetically soft showing negligible magnetic hysteresis. The behavior of and can be adjusted between two boundary conditions: (1) and decreasing almost linearly from ~300 and ~270 K, and (2) and remaining practically constant over the range ~300 to ~270 K. This ability to adjust the magnetocaloric effect properties allows one a flexibility in designing different refrigeration cycles and highly effective regenerator materials.
A possible application for these alloys is for climate control magnetic refrigeration devices and refrigerators/freezers. Alloying Gd with Zn significantly reduces melting temperature of the alloys (the eutectic alloy melts at ~860°C) compared to that of pure Gd (1313°C) and also improves the ductility over the GdZn intermetallide. This should simplify their fabrication into useful shapes (spheres, thin sheets, wires, etc.) for magnetic regenerator beds. INTRODUCTION
The binary compound GdZn is one of the several intermetallics, which order magnetically in the vicinity of room temperature and, therefore, a study of its magnetocaloric properties presents a significant interest with regard to its use as a magnetic refrigerant material. According to the Gd-Zn phase diagram, GdZn melts congruently and it is the Gd-richest intermetallic phase in the system.1 It orders ferromagnetically at 270 K.2 Alloys containing more than 50 at.% Gd are equilibrium two-phase alloys containing both GdZn and Gd. Since the Curie temperature of Gd is it offers a possibility to develop two-phase alloys containing two materials with Cryocoolers 10, edited by R. G. Ross, Jr. Kluwer Academic/Plenum Publishers, 1999
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different magnetic ordering temperatures and, therefore, by changing the ratio of the two components one can design magnetic refrigerant materials with different behaviors of the magnetocaloric effect. The magnetocaloric effect in GdZn is expected to be lower than that in pure Gd because of the dilution of the magnetic Gd lattice with non-magnetic Zn. However, the formation of an intermetallic compound typically has a positive effect particularly when commercial (and significantly less expensive) metals are used to form an alloy because many of the impurities soluble in pure metals tend to segregate on grain boundaries of the intermetallic compounds and thus effectively “purify” the intermetallic phase. In this paper we present the experimental results on measurements of the magnetocaloric effect in GdZn using different experimental techniques. The magnetic entropy change, was calculated from heat capacity and magnetization. The adiabatic temperature rise, was calculated from heat capacity. The heat capacity and magnetocaloric effect of several two-phase alloys was modeled using the experimental heat capacities of GdZn and pure Gd. EXPERIMENTAL PROCEDURES
The GdZn alloy was prepared by melting a stoichiometric mixture of pure Gd and Zn in a sealed Ta crucible filled with pure He using a vacuum induction furnace. The gadolinium was prepared by the Materials Preparation Center, Ames Laboratory and was 99.9 at.% (99.99 wt.%) pure. The major impurities were as follows: O-0.044 (0.004), C-0.020 (0.002), H-0.016 (0.000), N-0.009 (0.001), Fe-0.004 (0.001), and F-0.003 (0.000), where the number after the chemical symbol represents impurity content in at.%, and the number in parenthesis is the impurity content in wt.%. The zinc, which was purchased commercially, was 99.99+ wt.% pure. The alloy (total weight approximately 20 g) was re-melted 3 times with the crucible being turned over after each melting to ensure its homogeneity. The Ta crucible was then machined off. The x-ray powder
diffraction study of the as-cast alloy, performed using an automated Scintag powder diffractometer and radiation, showed that the alloy was essentially single-phase material. The GdZn phase has a cubic structure, which is isotypical with CsCl. The x-ray data yielded a lattice parameter of which is in good agreement with the literature.1 The heat capacity was measured from ~3.5 to 350 K in magnetic fields of 0, 2, 5, 7.5 and 10 T using an adiabatic heat-pulse calorimeter.4 The accuracy of the heat capacity measurements was ~0.5% over the whole temperature region. The ac susceptibility and dc magnetization was measured using a LakeShore susceptometer/magnetometer model No.7225 from ~4 to 325 K in applied fields up to 5.6 T. RESULTS AND DISCUSSION
Heat capacity and magnetic properties
The heat capacity of GdZn from ~3.5 to 350 K in magnetic fields 0, 2 and 7.5 T is shown in Fig.l (the heat capacity was also measured at 5 and 10 T, but for clarity the data are not shown in Fig.l). At zero magnetic field a well-defined -type anomaly is observed at and corresponds to a second order magnetic phase transition from a paramagnetic to a ferromagnetic
structure. The Curie temperature determined from the zero magnetic field heat capacity inflection point is and this value agrees well with that reported earlier.2 Magnetic field broadens the -type anomaly shifting it towards higher temperature, which is consistent with ferromagnetic ordering. A least squares fit of the low temperature heat capacity (T < 7 K) yields the electronic specific heat constant,
This value is typical of many other lanthanide intermetallic compounds. The Debye temperature of GdZn is which is somewhat higher than the Debye temperature of pure Gd (169 K), and which is consistent with the lower atomic weight of the second component (Zn) in the intermetallic compound. The relatively large uncertainties in both electronic heat capacity and Debye temperature are due to the lack of
Gd-Zn ALLOYS AS MAGNETIC REGENERATOR MATERIALS
Figure 1. The heat capacity of GdZn from 4 to 350 K in magnetic fields 0, 2, and 7.5 T.
Figure 2. The inverse magnetic susceptibility of GdZn from ~5 to 325 K (symbols). The
solid line shows the linear least squares fit of the data above ~295 K to the Curie-Weiss law.
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Figure 3. The magnetic entropy change,
of GdZn for a magnetic field change
from 0 to 2, 0 to 5,0 to 7.5, and 0 to 10 T calculated from magnetization (filled symbols) and heat capacity measurements (open symbols). The of commercial Gd for
magnetic field change from 0 to 7.5 T (thick solid line) is shown as a reference.
experimental heat capacity data below 4 K because of the low temperature limitations of our apparatus.4 Magnetization isotherms measured between ~5 and 325 K in magnetic fields from 0 to 5.6 T confirm that GdZn is ferromagnetic below ~270 K. The magnetic susceptibility (Fig.2) obeys the Curie-Weiss law above ~295 K with an effective magnetic moment of peff=7.82(6) µB and a Weiss constant The value of the measured effective magnetic moment agrees well with the theoretically expected value for a free ion (7.94 ). The large and positive Weiss constant also supports the ferromagnetic ground state of GdZn.
Magnetocaloric effect The magnetic entropy change, which is shown in Fig.3, was calculated from the magnetization using the Maxwell relation:
and from magnetic field and temperature dependent heat capacity as described elsewhere.5 The results obtained from the two different experimental techniques are in excellent agreement with one another. The peak in is observed at T=269±1 K, i.e. at the temperature where the
spontaneous ferromagnetic ordering occurs and remains practically independent of the magnetic field. The difference of the peak values of of GdZn and of Gd is consistent with a 30 wt.% dilution of magnetic Gd by non-magnetic Zn The magnetocaloric effect, calculated from magnetic field and temperature dependent heat capacity 5 for magnetic field changes from 0 to 2, 0 to 5, 0 to 7.5, and 0 to 10 T is shown in
Fig.4 together with that of a commercial Gd sample. Note that the low temperature anomaly in
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Figure 4. The magnetocaloric effect, of GdZn for a magnetic field change from 0 to 2, 0 to 5, 0 to 7.5, and 0 to 10 T calculated from magnetic field and temperature dependent heat capacity. The of commercial Gd for magnetic field change from 0 to 7.5 T shown as a thick solid line for reference.
the magnetocaloric effect of commercial Gd is due to the presence of a substantial amount of gadolinium oxide, in the Gd sample. The peak value of the magnetocaloric effect in GdZn reaches approximately 65% of the peak value of the magnetocaloric effect observed in pure Gd for the same magnetic field change. Both the magnetic entropy change and the adiabatic temperature rise show a simple caret-like behavior peaking at the Curie temperature,6 which is consistent with a simple ferromagnetic ordering of the Gd atoms in GdZn.
Magnetocaloric effect of the two-phase Gd-Zn alloys The heat capacity is an additive property, that is if the system contains two or more individual phases having different heat capacities, the heat capacity of the mixture is the prorated sum of the heat capacities of all of the components. This allows an easy and accurate way of modeling the heat capacity, the total entropy and the magnetocaloric effect of multi-phase mixtures. This may be particularly useful when modeling the magnetocaloric effect in binary
metallic systems such as Gd-Zn, where the two phases have different but close to one another magnetic ordering temperatures. The heat capacity of alloy, which is a eutectic composition containing 50 mol.% Gd and 50 mol.% GdZn, is shown in Fig.5 in two magnetic fields – 0 and 7.5 T. The heat capacity for another alloy with the chemical composition which is slightly richer in Zn and has 40 mol.% of Gd and 60 mol.% of GdZn is shown in Fig.6. The heat capacity of these two alloys was modeled by calculating the weighted average of the heat capacity of Gd and GdZn in respective magnetic fields. For example, the heat capacity of alloy was calculated as where and are the molar heat capacities of Gd and GdZn, respectively. The validity of this procedure was verified experimentally.7
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Figure 5. The heat capacity of the eutectic alloy containing 50 mol.% Gd and 50 mol.% GdZn in 0 and 7.5 T magnetic fields calculated from the individual heat capacities of GdZn and commercial Gd.
Figure 6. The heat capacity of the alloy containing 40 mol.% Gd and 60 mol.% GdZn in 0 and 7.5 T magnetic fields calculated from the individual heat capacities of GdZn and
commercial Gd.
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The zero field heat capacities of both two-phase alloys display two distinct -type anomalies at ~270 and ~292 K which are due to the ferromagnetic ordering in GdZn and Gd phases, respectively. The major difference between the zero field heat capacities of the two alloys is observed in the height of the -type anomalies and not in their positions. The calculated and values of the and alloys are shown in Figs. 7 and 8, respectively. The eutectic alloy shows a constant entropy change over the range of temperatures 270 to 293 K, while its adiabatic temperature rise increases nearly linearly over the same temperature interval. By adjusting the composition of this alloy to (i.e. by increasing the relative amount of GdZn phase) its entropy change decreases almost linearly (Fig.7) and the adiabatic temperature rise is constant between 270 and 293 K (Fig.8). Even though both and of the Gd-Zn alloys are the same below 270 K or somewhat lower above 270 K than that of pure Gd, the flexibility in modifying the temperature dependencies of and offered by adjusting the Gd to Zn ratio may allow the design of more effective magnetic refrigerant, and/or different magnetic refrigeration cycles. For example, for an AMR (active magnetic regenerator) cycle, which ideally requires a linear adiabatic temperature rise with increasing temperature, one would chose an alloy near the composition. Likewise for an Ericsson cycle, which ideally requires a constant entropy change over the specified temperature range, one would also chose an alloy near the composition. In addition to being able to easily modify the magnetocaloric effect properties, the use of a two phase alloy, at least in the case of Gd-Zn alloys, provides a great advantage in fabricating the magnetic regenerator material. Pure Gd melts at 1313°C and GdZn melts congruently at The eutectic alloy has composition ~25 at.% Zn (i.e. approximately 50 mol.% Gd and 50 mol.% GdZn) and melts at ~860°C1. This melting temperature is significantly lower than those of the two components and this could simplify processing of the alloys into the spherical powders required for magnetocaloric beds. Furthermore, since the two phase mixture is reasonably ductile, it should be relatively easily fabricated into sheets or wires. The temperature range of the maximum magnetocaloric effect properties, 260 to 300 K (see Figs. 7 and 8) make these two phase alloys ideal candidate materials for air conditioning units (cold temperature of 288 K, 60°F) and refrigerators/freezers (265 K, 17°F).
CONCLUSIONS
Experimental measurements of the magnetization and heat capacity of GdZn as a function of temperature and magnetic field confirm that it orders ferromagnetically at 269 K in zero magnetic field. The maximum magnetocaloric effect in GdZn is approximately 30% smaller than that observed in pure Gd, which is consistent with the amount of non-magnetic Zn in the intermetallic compound. The magnetic entropy change calculated from two different experimental techniques is in excellent agreement with one another. Modeling of the magnetocaloric effect of different two-phase (Gd+GdZn) alloy compositions indicates that the Gd-Zn alloys with less than 50 at.% (~30 wt.%) Zn may be useful high performance active magnetic regenerator materials. The behavior of and in the Gd-Zn alloys can be easily adjusted to be constant or to change linearly with temperature between ~270 and ~293 K by changing the alloy composition. The ability to modify the magnetocaloric properties by
varying the Gd to Zn ratio allows a flexibility in the choice of thermodynamic cycle and the design of highly effective magnetic refrigerant regenerator materials. An additional benefit is the significant reduction of the melting temperature of near eutectic alloys compared to that of Gd and GdZn, which may simplify the processing of the two phase alloys into useful regenerator shapes (i.e. spheres, sheets or wires) for packing magnetocaloric beds.
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Figure 7. The magnetic entropy change, of the and alloys for a magnetic field change from 0 to 7.5 T calculated from the modeled heat capacity.
The magnetic entropy change of commercial Gd for the same field change is shown as a reference.
Figure 8. The magnetocaloric effect, of the and alloys for a magnetic field change from 0 to 7.5 T calculated from the modeled heat capacity. The magnetocaloric effect of commercial Gd for the same field change is shown as a reference.
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ACKNOWLEDGMENT
Ames Laboratory is operated by the U.S. Department of Energy (DOE) by Iowa State University under contract No. W-7405-ENG-82. This work was supported by the Office of Basic Energy Sciences, Materials Science Division. REFERENCES
1. Bruzzone, G., Fornasini, M.L., and Merlo, F., “Rare Earth Intermediate Phases with Zinc”, J. LessCommon Met., vol.22 (1970), pp.253-264. The Gd-Zn phase diagram found in several phase diagram handbooks have been constructed from the data published in this reference and the expected similarity between the known Pr-Zn, Nd-Zn and Y-Zn phase diagrams. 2. Eckrich, K., Dormann, E., Oppelt, A., and Buschow, K.H.J., “The Hyperfine Fields in Ferromagnetic rare-Earth Zinc Compounds with CsCl Structure”, Z. Phys. B, vol. 23 (1976), p. 157-171. 3. Dan’kov, S.Yu., Tishin, A.M, Pecharsky, V.K., and Gschneidner, K.A., Jr., “Magnetic Phase
Transitions and the Magnetothermal Properties of Gadolinium”, Phys. Rev. B, vol. 57, (1998), pp. 3470-3490. 4. Pecharsky, V. K., Moorman, J. O., and Gschneidner, K. A., Jr., “A 3-350 K fast Automatic Small Sample Calorimeter”, Rev. Sci. Instrum., vol. 68 (1997), pp. 4196-4207. 5. Pecharsky, V.K., and Gschneidner, K.A., Jr., “Comparison of the Magnetocaloric Effect Derived
from Heat Capacity, Direct, and Magnetization Measurements”, Adv. Cryogen. Eng., vol. 42A (1996), pp. 423-430.
6. Tishin, A.M., Gschneidner, K.A., Jr., and Pecharsky, V.K., , “The Magnetocaloric Effect and Heat Capacity in the Phase Transition Region”, to be published. 7. Gschneidner, K.A., Jr., Pecharsky, V.K., and Malik, S.K., “The
Alloys as Active Magnetic Regenerators for Magnetic Refrigeration”, Adv. Cryogen. Eng., vol. 42A (1996), pp. 475482.
Magnetocaloric Properties of V.K. Pecharsky and K.A. Gschneidner, Jr.
Ames Laboratory, US DOE, and Department of Materials Science and Engineering, Iowa State University, Ames, IA, U.S.A 50011-3020 S. Yu. Dan’kov and A.M. Tishin
Physics Department, M.V. Lomonosov Moscow State University, Moscow Russia 119899
ABSTRACT
The magnetic and thermal properties of have been studied using four different experimental techniques. The magnetocaloric effect was measured directly in quasi-static fields up to 2 T and in pulsed fields up to 8 T, and calculated from the temperature and magnetic field dependencies of the heat capacity, and the magnetization.
The compound orders
ferrimagnetically at 281 K in zero magnetic field. In magnetic fields larger than ~3 T a metamagnetic transition occurs between 40 and 50 K. The complex magnetic properties of also influence the behavior of the magnetocaloric effect. For a magnetic field change from 0 to 2 T the adiabatic temperature rise which is a single maximum, is ~2 K at the Curie temperature As the magnetic field increases from 2 to 5 T and greater, additional low temperature anomalies (two maxima and two minima) develop in the temperature range 15 to 150 K. Magnetocaloric effect maxima occur at 50 K J/kg K) and at 281 K (∆ Τad = 7.7 K, ∆S mag = -7.2 J/kg K) for a magnetic field change from 0 to 10 T. Experimental data obtained from the different measurement techniques are in excellent agreement with each other. INTRODUCTION
Recently there has been an increase in systematic studies of the magnetocaloric properties of both crystalline and amorphous intermetallic alloys containing lanthanide metals.1,2 These studies are quite useful because they contribute to the basic science by providing a rich spectrum of information about the nature of magnetic phase transitions. Simultaneously, they are of importance with respect to applied science and engineering because many of the lanthanide compounds display a large magnetocaloric effect (MCE) near the magnetic phase transition temperature. The availability of a large magnetic entropy, a wide variety of magnetic ordering temperatures, a small or negligible magnetic hysteresis, and sufficient thermal conductivity warrants the possibility of using some of these alloys as effective magnetic refrigerants.1-3 Since the efficiency of a magnetic refrigerator is proportional to the magnetic field induced magnetic entropy change of the magnetic refrigerant, it is, therefore, necessary to select magnetic solids displaying the largest For the best refrigerant performance, the large Cryocoolers 10, edited by R. G. Ross, Jr.
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must exist over a wide range of temperatures. It is known that the maximum
for a
ferromagnet is observed in the vicinity of its Curie temperature. In this paper we present the experimental results on the measurements of the magnetocaloric effect in the binary intermetallic compound using different experimental techniques. The magnetic entropy change was calculated from the heat capacity and magnetization, and the adiabatic temperature rise was calculated from the heat capacity and measured directly in quasi-static and pulsed magnetic fields. EXPERIMENTAL PROCEDURES
The
alloy was prepared by arc-melting a stoichiometric mixture of pure Gd and Al on
a water-cooled copper hearth in an argon atmosphere under ambient pressure. The Gd was prepared by the Materials Preparation Center of the Ames Laboratory and was 99.9 at.% (99.99 wt.%) pure. The major impurities were as follows: O-0.044 (0.004), C-0.020 (0.002), H-0.016 (0.000), N-0.009 (0.001), Fe-0.004 (0.001), and F-0.003 (0.000), where the number after the chemical symbol represents impurity content in at.%, and the number in parenthesis is the impurity content in wt.%. The aluminum, which was purchased commercially, was 99.99+ wt.% pure. The alloy (total weight approximately 10 g) was arc-melted 7 times with the button being turned over after each melting to ensure its homogeneity. The weight losses after the arc melting were negligible (<0.25 wt.%) and, therefore, the alloy composition was accepted as nominally
prepared. melts peritectically at 980°C, and the liquidus temperature for this composition is 1060°C. Thus when arc-melted alloy is cooled form the liquid state some additional phases may form even though the cooling rate is quite large for an arc-melted sample sitting on a watercooled hearth. For an alloy of the composition one might expect small amounts of GdAl and in the sample. The x-ray powder diffraction study of the as-cast alloy using an automated Scintag powder diffractometer with radiation, showed that the alloy was essentially single-phase material (i.e. the amount of second phase was < 5%). Since both GdAl and order magnetically at 42 and 43 K, respectively, heat capacity measurements may enable one to detect the presence of these compounds. A close examination of the zero field heat capacity indicates two small anomalies (bumps) at 40 and 47 K. The magnitude of the anomalies indicates that less than 2% of each may be present in the sample, which is consistent with the xray diffraction data. The heat capacity was measured from ~3.5 to 350 K in magnetic fields of 0, 2.0, 5.0, 7.5 and
10.0 T using an adiabatic heat-pulse calorimeter.4 The accuracy of the heat capacity measurements was ~0.5% over the whole temperature region. The dc magnetization was measured using a LakeShore susceptometer/magnetometer model No.7225 from ~4 to 325 K in applied fields up to 5.0 T. The magnetocaloric effect was measured directly from ~78 K to 350 K in quasi-static fields up to 2 T and in pulse fields up to 7 RESULTS AND DISCUSSION
Heat capacity and magnetic properties
The heat capacity of from ~3.5 to 350 K in magnetic fields 0, 5 and 10 T is shown in Fig.l. The heat capacity was also measured at 2 and 7.5 T, but for clarity the data are not included in Fig.l. A -type anomaly, observed at K, corresponds to a second order magnetic phase transition from a paramagnetic to a ferromagnetic or ferrimagnetic state. A magnetic field broadens 281 K peak shifting it towards higher temperatures, which is consistent with ferro- or ferri-magnetic ordering. When the magnetic field is 5 T or larger a heat capacity anomaly (a broad hump) is observed at ~`40 K, see inset of Fig.l. This is due to a metamagnetic (field induced) transition, which is confirmed by magnetization measurements, see next paragraph. Magnetization isotherms measured between ~5 and 320 K in magnetic fields from 0 to 5.6 T are shown in Fig.2. They confirm that is ferromagnetic below ~50 K at fields > 4T with
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Figure 1. The heat capacity of from ~5 to 350 K in magnetic fields 0, 5 and 10 T. The insert clarifies the behavior of low temperature heat capacity anomaly in magnetic fields 0, 5, and 10 T in a form of C/T vs. T plots.
an ordered magnetic moment of ~21 /formula unit. However in magnetic fields below ~2.5 T the ordered magnetic moment (~14 /formula unit) corresponds to 2/3-rds of the Gd atoms ordering ferromagnetically (the theoretical ordered magnetic moment per Gd atom is 7 ). This
Figure 2. Magnetization of
as a function of magnetic field from ~5 to 320 K.
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Figure 3. The inverse magnetic susceptibility of dc magnetic field of 0.2 T.
from ~5 to 325 K measured in a
indicates that the magnetic structure of is probably ferrimagnetic below ~280 K in the absence of magnetic field. A metamagnetic transition is observed below ~60 K in magnetic fields larger than 2.5 T, which brings about the fully aligned ferromagnetic structure in The crystal structure of is rather complex. There are three independent 4-fold Gd sites in this non-centrosymmetric tetragonal lattice. The magnetic structure of the ferrimagnetic phase, in which 2/3-rds of the Gd spins are aligned parallel, is probably even more complex, since a simple antiparallel alignment of two of the three independent Gd sites would lead to an effective magnetic moment of ~7 per formula unit. Two simple spin arrangements could account for the observed magnetic moment: (1) the spins in one set of Gd atoms are aligned antiparallel to each other, while the spins of those Gd atoms in the remaining two sets are in a parallel arrangement; and (2) the spins of half of the Gd atoms in two of the antiparallel sets of Gd sites are aligned parallel to the remaining Gd spins.
The magnetic susceptibility measured in 0.2 T dc magnetic field (Fig.3) obeys the CurieWeiss law above ~300 K with an effective magnetic moment of 7.88(8) and a Weiss constant The value of the effective magnetic moment agrees with the theoretical value expected for a free ion (7.94 ) within experimental error. The large and positive Weiss constant supports the ferrimagnetic ground state of Magnetocaloric effect
The magnetic entropy change, which is shown in Fig.4, was calculated from magnetization (Fig.2) using the Maxwell relation:
and from the magnetic field and temperature dependent heat capacity (Fig.1) as described elsewhere.7 The results obtained from the two techniques are in excellent agreement with one
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Figure 4. Magnetic entropy change in from ~5 to ~350 K in magnetic fields up to 10 T as determined from magnetization (filled symbols) and heat capacity (open symbols) measurements.
another (Fig.4). The upper temperature peak in is observed at K, i.e. at the temperature where the spontaneous ferrimagnetic ordering occurs, and remains practically independent of the magnetic field (Fig.4). This maximum is the only one observed for a low magnetic field change from 0 to 2 T. Consistent with the magnetic field induced transition, the second anomaly in develops when the magnetic field is increased beyond 2 T. Unlike the upper peak, the position of the low temperature peak in is strongly dependent on the magnetic field, which indicates that a metamagnetic transition is both temperature and field dependent. The magnetocaloric effect, was measured directly in a quasi-static magnetic field changing from 0 to 2 T, and in a pulsed magnetic field from 0 to 2 and 0 to 5 T. It was also calculated from the magnetic field and temperature dependent heat capacity for magnetic field changes from 0 to 2, 0 to 5, 0 to 7.5, and 0 to 10 T. The results, which are shown in Fig.5, are in excellent agreement with one another. The upper magnetocaloric effect peak occurs at K. The two lower temperature peaks are observed below 60 K when magnetic field increases to 5 T and above. The peak at is consistent with the low temperature shoulder visible in the curves. Its position remains independent of the magnetic field. The K (0 to 5 T) and the K (0 to 10 T) peaks are consistent with the corresponding peaks in magnetic entropy change. This behavior also indicates that the magnetic structure of the in magnetic fields in excess of 2 T is more complicated than a simple ferromagnet. The adiabatic temperature rise of was previously measured between 170 and 280 K for a 0 to 5.5 T field change.8 The results are in fair agreement with our data over this temperature range. The magnetocaloric effect peak occurs at about 10 K lower than ours. This difference is probably due to the impure Gd used to prepare the by Nikitin et al.8 The value of the magnetocaloric effect peak in near room temperature is approximately 40% of the peak value of the magnetocaloric effect observed in pure Gd for the same magnetic field change. The low temperature magnetocaloric effect in is also significantly smaller when compared to other prototype materials, such as alloys.1
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Figure 5. Magnetocaloric effect, in from ~5 to ~350 K in magnetic fields up to 10 T measured directly in pulsed and quasi-static field (filled symbols), calculated from heat capacity (open symbols).
The reduction of both can be understood based on the facts that (1) the two magnetic phase transitions in are more than 200 K apart on the temperature scale and that (2) the total
magnetic entropy available for utilization in the magnetocaloric effect is limited to per mole of Gd atoms (where R is the universal gas constant, and J is the total angular momentum [7/2 for Gd]) and is divided between the two transformations. Nevertheless, the presence of multiple magnetic ordering phenomena in brings about a significant magnetocaloric effect in the temperature range from ~10 to ~ 310 K. The cooling capacity of in the temperature range from 10 to 310 K, which can be defined as the amount of energy transferred from a cold reservoir to a hot reservoir in one ideal refrigeration cycle, and can be calculated as
is 610 J/kg for a magnetic field varying from 0 to 5 T, and it increases more that two-fold to
1400 J/kg for magnetic field change from 0 to 10 T. The best performance of the magnetic refrigerant is expected between 10 and 80, and 250 and 310 K in magnetic fields exceeding 5 T. CONCLUSIONS
The study of the magnetic, thermodynamic and magnetocaloric properties of using four different experimental techniques (magnetization, heat capacity, and magnetocaloric effect in quasi-static and pulsed field as a function of temperature and magnetic field) shows excellent agreement between the results obtained from different experimental measurements. The compound orders ferrimagnetically at 281 K in zero magnetic field. In magnetic fields larger than ~3 T a metamagnetic transition from ferrimagnetism to ferromagnetism occurs between 40 and 50 K. Additional changes in the magnetic structure of at ~ 13 K and magnetic fields
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in excess of 5 T are evident from the magnetocaloric effect studies. The complex magnetic properties of is evident from the behavior of the magnetocaloric effect, which displays several peaks at different temperatures between ~13 and ~281 K. The magnetocaloric effect in is considerably smaller compared to the best prototype materials in the respective temperature ranges. The best performance of the magnetic refrigerant is expected in magnetic field stronger than 5 T and in the temperature ranges from 10 to 80 K and from 250 to 310 K. ACKNOWLEDGMENT
Ames Laboratory is operated by the U.S. Department of Energy (DOE) by Iowa State University under contract No. W-7405-ENG-82. This work was supported by the Office of Basic Energy Sciences, Materials Science Division (KAG and VKP) and by a NATO Linkage
Grant No. 950700 (all authors). REFERENCES
1. Gschneidner, K. A., Jr., Pecharsky, V. K., and Malik, S. K., “The Alloys as Active Magnetic Regenerators for Magnetic Refrigeration”, Adv. Cryogen. Eng., vol. 42A (1996), pp. 475482.
2. Liu, X. Y., Barclay, J. A., Földeáki, M., Gopal, B. R., Chahine, R., and Bose, T. K., “Magnetic Properties of Amorphous and Alloys”, Adv. Cryogen. Eng., vol. 42A (1996), pp. 431-438. 3. Gschneidner, K. A., Jr., Takeya, H., Moorman, J. O., and Pecharsky, V. K., " (Dy0.5Er0.5)Al2: A Large Magnetocaloric Effect Material for Low-Temperature Magnetic Refrigeration", Appl. Phys. Lett., vol. 64 (1994), pp. 253-255. 4. Pecharsky, V. K., Moorman, J. O., and Gschneidner, K. A., Jr., “A 3-350 K Fast Automatic Small Sample Calorimeter”, Rev. Sci. Instrum., vol. 68 (1997), pp. 4196-4207. 5. Dan’kov, S. Yu., Tishin, A. M., Pecharsky, V. K., and Gschneidner, K. A., Jr., “Experimental Device for Studying the Magnetocaloric Effect in Pulse Magnetic Fields”, Rev. Sci. Inst., vol. 68 (1997), pp.
2432-2437. 6. Baenziger, N.C., and Hegenbarth, J.J., “Gadolinium and Dysprosium Intermetallic Phases. III. The Structures of
and
”, Acta Crystallogr., vol. 17 (1964), pp.
620-621. 7. Pecharsky, V.K., and Gschneidner, K.A., Jr., “Comparison of the Magnetocaloric Effect Derived from Heat Capacity, Direct, and Magnetization Measurements”, Adv. Cryogen. Eng., vol. 42A (1996), pp. 423-430.
8. Nikitin, S.A, Spichkin, Yu.I., and Tishin, A.M., “Magnetocaloric Effect and Bulk Magnetostriction of ”, Fiz. Tverd. Tela, vol. 31 (1989), pp. 250-253. Engl. Transl: Sov. Phys. Solid State, vol. 31
(1989), pp. 1241-1242.
Development of a Dilution Refrigerator for Low-Temperature Microgravity Experiments Pat R. Roach NASA Ames Research Center Moffett Field, CA 94035 and Ben P. M. Helvensteijn Sterling Software Redwood Shores, CA 94065
ABSTRACT A dilution refrigerator (DR) is the most common precooling stage for sub-millikelvin demagnetization experiments. The usefulness of the DR comes from its ability to provide cooling at 0.02-0.04 K for long periods of time while the heat of magnetization is being rejected by the demagnetization stage. In order to make these advantages of the DR available to researchers who need the microgravity of space for their experiments, we are developing a continuously-operating DR that will function in microgravity. We have previously demonstrated that the liquid helium of the DR can be controlled by the use of capillary forces in sintered metal sponges. We have found, however, that the small pores needed to control large heights of liquid on the ground are too small to allow sufficient liquid flow for effective cooling. We have built a shallow single-cycle version of the refrigerator that does not require large heights of liquid to be supported by capillary forces. The liquid chambers are next to each other and are filled with sinter with relatively open pores; these pores will allow much freer flow of the helium. The gravity independence of this design will be tested by tilting the system so that one chamber is slightly above or below the other and by inverting both chambers. The operation of the refrigerator should be unaffected by tilts of 5-10 degrees or by the inversion of the chambers. A design for a continuously-operating dilution refrigerator is presented. It includes all the advantages of the single-cycle refrigerator while adding the large advantage of continuous cooling for long periods of time. INTRODUCTION Research at low temperatures is an extremely fruitful field because of the many phenomena that occur only there. Unusual phases of matter such as superconductors and superfluids occur at low temperatures and many subtle behaviors that are obscured by thermal motion at higher temperature can be studied in great detail at low temperatures. Cryocoolers 10, edited by R. G. Ross, Jr. Kluwer Academic/Plenum Publishers, 1999
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Type of refrigerator needed To carry out research at low temperatures it is necessary to have a refrigerator that 1) cools to the required temperature, 2) is reliable and, 3) if possible, operates continuously for the duration of the experiment, whether that is hours or days. On the ground the need for temperatures below 0.3 K is almost universally met by the He-3-He-4 dilution refrigerator. Its usefulness arises from the fact that it operates continuously, it can provide a substantial cooling power at temperatures from around 1.0 K down to 0.010 K and below and it can run uninterrupted for as long as several months. There are many very interesting physics experiments that need the unique microgravity environment of space but which also need lower temperatures than are currently available. In order to investigate phenomena that occur at very low temperatures, particularly in superfluid He-3, the capability for extending research to temperatures of 0.001 K in space needs to be developed. On the ground, temperatures to 0.001 K and below are reached with adiabatic demagnetization systems that are precooled with helium dilution refrigerators. Similar temperatures can be achieved in space if the dilution refrigerator can be adapted to work in microgravity.
Microgravity Research
An important example of microgravity research is the study of liquid He-4 and He-3, both normal and superfluid phases; this has been a very productive field for many years because of the unique nature of these two very different liquids. The availability of the low gravity of space is a boon to this research because gravity has a major effect upon the behavior of the liquid. At phase transitions of the liquid, for example, the effect of gravity is to spread out the region over which the transition occurs. This can seriously mask important details of the transition. Other phenomena, such as spin-spin relaxation in He-3, are strongly influenced by the surface of a container, and the ability to form freely floating drops in microgravity would allow the influence of the container to be eliminated. Adapting a Dilution Refrigerator for Space
The helium dilution refrigerator relies on the unique properties of liquid He-3 and He-4. Cooling to 0.010 K and below is produced when He-3 atoms cross the phase boundary that exists between liquid He-3 and liquid He-4 at low temperatures. (Essentially, He-3 ‘evaporates’ into the liquid He-4.) We have been studying the capabilities of a special dilution refrigerator;1 this refrigerator is unusually compact and reliable, making it especially suitable for space applications. On the ground, gravity provides the force that keeps the two liquids in their required places so that the cooling can happen when and where it is needed. In space this force can be replaced with capillary forces that arise when the liquids are confined in porous sponges. We have shown2,3 that it should be possible to develop a helium dilution refrigerator that will confine the liquids with capillary forces and still provide the cooling that makes the dilution refrigerator so valuable. This approach should work even better in microgravity.
We are sure that it is possible to adapt the dilution refrigerator to operate in microgravity. If this can be achieved, the same features that make the dilution refrigerator so attractive for laboratory
research would become available to researchers in space. PREVIOUS DEVELOPMENT
Principle of Single-Cycle Dilution Refrigerator
Figure 1 shows how such a refrigerator operates. The lowest temperatures occur in the mixing chamber where there is a phase boundary between liquid He-3 and liquid He-4. Cooling is produced when He-3 crosses this boundary into the He-4. From the mixing chamber this dilute He-3
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Figure 1. Operation of compact single-cycle dilution refrigerator using a charcoal pump.
flows through the He-4 to a higher temperature chamber where it is fractionally distilled from the He-4. The resulting He-3 gas is collected by the charcoal pump. The cooling cycle ends when all the He-3 is in the charcoal pump. Because the refrigerator uses adsorption onto charcoal for its pumping, all operations can be controlled by heaters and, as a consequence, there are no moving parts in the refrigerator. Modification for Microgravity
On the ground, the operation of a dilution refrigerator depends on gravity to keep the liquid He3 and He-4 in their correct chambers. (The charcoal pump contains no liquid and is gravity independent.) Within the dilution refrigerator there are two liquid-vapor interfaces and one liquidliquid interface. All of these interfaces must be stably located in the absence of gravitational forces in a way that allows the free flow of the evaporated gasses and of the He-3 within the liquid phases of the refrigerator. Previous experiments4 have shown that capillary forces in a matrix of fine pores can successfully contain liquid helium in microgravity. We have extended this approach to He-3He-4 mixtures in a ground-based demonstration. The modifications we have made involve filling the liquid chambers of the dilution refrigerator with a sintered, porous metal matrix that confines the liquids to their correct positions by capillary forces. A critical aspect of this is the need to prevent the phase boundary between the liquid He-4 and liquid He-3 from leaving the mixing
chamber. This is greatly complicated by the fact that the interfacial tension between these two phases is exceedingly small5, only compared with for the He-3 liquidvapor surface tension and for the He-4 liquid-vapor surface tension. However, if the pores outside the region of the He-3 are small enough, and if the osmotic pressure trying to push the He-3 into these pores is not too large, the liquid He-3 will be prevented from entering the small
pores containing the He-4 by the interfacial tension; the He-3 will stay in the mixing chamber where it is needed.
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An obvious demonstration of the gravity independence of the dilution refrigerator would require a system many centimeters in diameter to operate properly in all orientations; this, however, is unrealistic. To use capillary forces to control the position of the various liquid-liquid and liquidvapor boundaries under an adverse gravitational head of many centimeters of liquid on the ground, it is necessary to use very fine pores ( diam.) to contain the liquid. However, we have found that such small pores seriously impede the flow of dilute He-3 from the mixing chamber to the still; this limits the cooling power achieved and prevents the attainment of the temperature goal desired. This need not be a problem for a space-based system since the dilution refrigerator can actually use quite large pores to overcome the very small accelerations likely to be encountered in orbit. The problem is that it is necessary to test the concept of capillary confinement in a dilution refrigerator on the ground (to at least a limited extent) before committing to a flight test to provide final confirmation of the approach.
Shallow Version. We have built a shallow version of the dilution refrigerator that has a mixing chamber and still that are only 0.5 cm high; this allows us to use sinter with rather large pores (40 to diam. in different locations, see Fig. 2), which we expect will permit excellent operation of the system. This design can verify a limited amount of gravity independence of the operation on the ground. In its normal position with the shallow still next to the shallow mixing chamber, and the pumping line coming out of the top of the still, it will operate even without sinter in the chambers. It would not continue to operate with the system tilted slightly so that one cham-
ber was above the other. With the coarse sinter in the chambers, however, we expect to be able to
tilt the system 5-10 degrees in either direction with little change in operation (see fig. 3). If the system is tilted more than this, either one or the other of the sinters where the connecting line attaches becomes empty, stopping the dilute He-3 circulation, or else the He-3 in the mixing chamber escapes into the surrounding small pores, allowing it to leave the mixing chamber. In a more convincing demonstration of gravity independence, the chambers can be inverted so that the still pump line is on the bottom (see fig. 4) and the system should still operate normally. Clearly, no operation in this orientation would be possible without the sinter; the liquid in the still would simply run into the pumping line. Even in this orientation the system can be tilted 5-10 degrees in either direction before the capillary forces are overcome by gravitational forces and the
Figure 2. Arrangement of different size sinters for confining liquid helium in still and mixing chamber.
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Figure 3. Distribution of liquid He as the shallow dilution refrigerator is tilted (on the ground).
Figure 4. Shallow dilution refrigerator can be tilted even while its chambers are inverted to convincingly demonstrate gravity independence.
Figure 5. Details of the low-temperature chambers of the continuously-operating dilution refrigerator.
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Figure 6. Components of a continuously-operating dilution refrigerator for microgravity use.
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liquid runs from one chamber into the other. These limited confirmations of capillary confinement will be a good indication that the system will work well in space. CONTINUOUSLY-OPERATING DILUTION REFRIGERATOR
We are developing a continuously-operating dilution refrigerator that will directly cool to 0.040 K in a microgravity environment. Such a refrigerator could also precool an adiabatic demagnetization stage for reaching temperatures of 0.001 K and below. Figure 5 shows the configuration we propose to test on the ground. The still and mixing chamber of this design are similar to those of the single-cycle refrigerator we have built. The dilute He-3 flows out of the mixing chamber into the still as before. But the He-3 gas, instead of being pumped from the still into a charcoal pump, now goes to a new chamber, the condenser, at 0.4 K, where it condenses back to a liquid and pure He-3 returns to the mixing chamber. Thus this He-3 never leaves the low-temperature region. As long as the still is heated to maintain its temperature at 0.6 K and the condenser is cooled to maintain its temperature at 0.4 K, He-3 will be continuously pumped from the still into the condenser and forced back into the mixing chamber. This continuous circulation of He-3 will produce continuous cooling in the mixing chamber where He-3 crosses the phase boundary from pure He-3 into the He-4. The notable feature of this design is the method by which the condenser is continuously cooled (while maintaining the advantages of compactness, reliability and the complete absence of moving parts). The condenser is cooled by a pair of independent, single-cycle He-3 refrigerators (see Fig. 6), each with its own charcoal pump, and each thermally linked to the condenser by a gas-gap heat switch. While one He-3 pot is cold and coupled to the condenser, the other He-3 pot is isolated from the condenser while it is being refilled at high temperature. Then, before the first He-3 pot runs empty, the second He-3 pot would be cooled down and coupled to the condenser by its heat switch. The first He-3 pot could then be decoupled and refilled and there would have been no interruption of cooling to the condenser. CONCLUSIONS
We have built a 'shallow' single-cycle dilution refrigerator to demonstrate the principle of capillary confinement in a refrigerator that can reach 0.1 K or below in microgravity. We have designed a continuously-cooling version of a dilution refrigerator that builds on the design of the single-cycle refrigerator while maintaining its advantage of no moving parts. The continuouslycooling version will be very useful by itself for microgravity experiments that require cooling to as low as 0.04 K; it will also be invaluable for experiments that require temperatures as low as 0.001 K because it can be used to precool adiabatic demagnetization systems that can reach those temperatures. REFERENCES
1. Roach, P. R. and Gray, K. E., "Low-Cost, Compact Dilution Refrigerator: Operation from 200 to 20 mK", Advances in Cryogenic Engineering, vol. 33, Plenum Press, New York (1988), pp. 707-712. 2. Roach, Pat R., "Thermal Efficiency of a Zero-G Dilution Refrigerator", Proc. Fifth Intern. Cryocoolers
Conf., Monterey, CA (August 1988) pp.195-204. 3. Roach, Pat R , and Helvensteijn, Ben, "Development of a Compact Dilution Refrigerator for Zero
Gravity Operation", Advances in Cryogenic Engineering, vol. 35, Plenum Press, New York (1990) pp. 1045-1053. 4. Duband, L., Alsop, D., Lange, A.E., and Kittel, P., "He-3 Refrigerators for Space", Adv. Cryo. Engin., vol. 35B (1990) p. 1447.
5. Guo, H. M., Edwards, D. O., Sarwinski, R. E., and Tough, J. T., "Fermi-Liquid Behavior of He-3 Adsorbed on Liquid Helium", Phys. Rev. Lett., vol. 27, no. 19 (1971) pp. 1259-1263.
Preliminary Experimental Results Using a Two Stage Superfluid Stirling Refrigerator A. B. Patel and J. G. Brisson Massachusetts Institute of Technology Cambridge, MA 02139
ABSTRACT
This paper describes the first operational two stage SSR and reports on its preliminary experimental performance. Previous SSR's were single stage machines consisting of a hot
(compressor) platform connected to a single cold (expander) platform by a heat exchanger. In order to achieve lower ultimate temperatures and higher cooling powers at low temperatures, we have constructed a two stage SSR that has one compressor platform and two expander platforms, which are connected to each other by two heat exchangers. Our two stage SSR has a total internal volume of and uses and Kapton heat exchangers. Operating from a high temperature of 1.0 K and with a 3.0% mixture, this SSR achieves a low temperature of 282 mK and delivers net cooling powers of 1 mW at 617 mK, at 485 mK and at 361 mK. This performance is equivalent to the best low temperature results of previous single stage SSR's. Based upon the reults of this SSR, we expect that the use of larger Kapton recuperators will significantly improve the performance of future two stage SSR's. INTRODUCTION The superfluid Stirling refrigerator (SSR) is a Stirling cycle refrigerator which uses a mixture as the working fluid to provide cooling below 2 K. At these temperatures, the mixture behaves to first approximation as an ideal gas of in an inert background of superfluid The basic components of a single stage SSR are a hot (compressor) piston and a single cold (expander) piston connected by a regenerator. Superleaks bypass each piston and allow the superfluid to flow freely between the pistons. These superleaks enable the thermodynamically active to be expanded and compressed through the Stirling cycle which results in cooling. A more detailed explanation of the cycle is provided by Watanabe, Swift, and Brisson.1 A single stage SSR was first demonstrated by Kotsubo and Swift in 1990.2 In 1992, Brisson and Swift3-6 improved the design by operating two refrigerators 180 degrees out of phase and using a counterflow recuperator made of CuNi tubes for the regenerator. They achieved a low temperature of 296 mK operating from a high temperature of 1.05 K. Using the same machine, Watanabe, Swift, and Brisson7 later reached 168 mK operating from a high temperature of 383 mK. Recently, Patel and Brisson8, using high efficiency plastic recuperators, achieved lower ultimate temperatures and higher cooling powers with a larger SSR of the Brisson and Swift
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design configuration. To date, the best performance of a single stage SSR has been a Patel and Brisson SSR that used a Kapton heat exchanger having 120 cm3 devoted to recuperative heat transfer. Operating from a high temperature of 1.0 K, this machine achieved an ultimate low temperature of 291 mK and delivered net cooling powers of at 418 mK and at
Unfortunately, the low temperature performance of a single stage SSR can not be significantly improved without great difficulty and expense. Improving the performance of a single stage SSR requires higher efficiency recuperators. However, the large Kapton recuperators used in the Patel and Brisson SSR's are already quite efficient (~90%), and improvements which further increase efficiency will be both difficult to design and expensive to construct for three reasons: 1. Higher efficiency low temperature recuperators will require very large surface areas because the Kapitza resistance to heat transfer between helium and most heat
exchanger materials increases by at least three orders of magnitude between 1.0 K and 100 mK.
2. High efficiency recuperators require uniform flow distribution among recuperator passages and between the recuperator sides. In order to maintain the magnitude of the pressure oscillation within the SSR (and thereby the cooling of the SSR), the recuperator must have a relatively small volume. This constraint requires a large surface area to volume ratio in the recuperator, which results in recuperators with large numbers of extremely narrow passages. However, as the passage size decreases and the number of passages increases, geometric variations among the passages and headering issues make uniform flow distribution much harder to achieve. 3. Due to the oscillating pressure and oscillating flow within a reciprocating machine such as a Stirling refrigerator, very high efficiency recuperators are difficult to model and design. The standard steady flow, constant pressure heat transfer correlations are not applicable, and second order effects such as heat transfer hysteresis loss and heat transfer phase shift can no longer be neglected.10,11 Another method of improving the low temperature performance of the SSR, which does not require more efficient recuperators, is to construct SSR's that have multiple expander platforms (stages). In 1996, we proposed building a two stage SSR.12 As shown in Fig. 1, a two stage SSR has a single compressor platform and two expander platforms connected by two heat exchangers. Our theoretical analysis shows that the second expansion stage should enable the SSR to reach lower ultimate temperatures and deliver higher cooling power at low temperatures than a single stage SSR.13 However, the price of this improved low temperature performance is reduced high temperature cooling powers. This paper describes the first two stage SSR and reports its preliminary experimental performance. This two stage SSR has a total internal volume of 120 cm3 and uses 3.0 cm3 and 4.8 cm3 Kapton heat exchangers. Operating from a high temperature of 1.0 K and with a 3.0% mixture, this SSR achieves a low temperature of 282 mK and delivers net cooling powers of 1 mW at 617 mK, at 485 mK and at 361 mK. DESCRIPTION OF THE TWO STAGE SSR
Figure 1 shows a schematic of our SSR. This refrigerator uses the Brisson and Swift design configuration, so it has two SSR's operating 180 degrees out of phase with each other and counterflow recuperators as the regenerators. The SSR consists of a hot (compressor) platform connected by a Kapton recuperator to an intermediate (expander) platform, which in turn is connected to a cold (expander) platform by another Kapton recuperator. The hot, intermediate, and cold platforms of this SSR are made of solid blocks of OFHC copper on which the pistons are mounted. The pistons are made with edge welded stainless steel
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bellows14 which have convolutions that nest into one another to minimize void volume. The effective areas of the pistons, based on the manufacturer's specifications, are for the hot platform pistons, 13.16 cm3 for the intermediate platform pistons, and for the cold platform pistons. The hot platform pistons are rigidly connected together and driven sinusoidally
using a push rod from a room temperature drive. The intermediate and cold platform pistons are similarly connected and driven together using a common push rod. The hot platform temperature is pinned at approximately 1.0 K by a evaporation refrigerator. Within each piston platform, there are superleaks made from porous Vycor glass which allow the superfluid to flow freely between the halves of the SSR during operation. In the hot platform, the superleaks are three Vycor cylinders 6.03 cm in length with diameters of 1.39 cm, 1.35 cm, and 0.72 cm. In the intermediate platform, the superleaks are three Vycor cylinders 10.63 cm in length with diameters of 0.74 cm. In the cold platform, the superleaks are three Vycor cylinders 15.16 cm in length with diameters of 0.74 cm. The large number of Vycor
Figure 1. A cross sectional view of the two stage SSR. The top pistons of each platform form one SSR half which operates 180 degrees out of phase with the SSR half formed by the bottom pistons. The counterflow recuperators between SSR halves act as the regenerators.
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cylinders in the platforms provide a total superleak cross sectional area of
and allow
the SSR to run at higher speeds, at higher temperatures, and with larger volume displacements than previous SSR's, without exceeding critical velocities. The total volume of the Vycor glass in the SSR is
Since 28% of the Vycor glass15 is void space, the glass contributes
15 cm3 to the total mixture volume of the SSR. The that diffuses into this volume, though, does not participate in the operation of the SSR because its diffusion times are substantially longer than the SSR cycle period. This is effectively trapped and does not undergo either compression or expansion during the SSR cycle. Within each piston platform, there are also isothermal heat exchangers. These heat exchangers are necessary for the SSR to deliver high cooling powers because most of the
expansion (compression) of within the pistons of a large SSR occurs adiabatically and heat must be added (removed) from the working fluid before it enters the recuperator. To accomplish this, large surface areas are needed to overcome the Kapitza boundary resistance between the copper and the mixture. However, this surface area must come with minimal void volume in order to maintain the magnitude of the pressure oscillation within the SSR.
Our SSR's isothermal heat exchangers are made from nested OFHC copper cylinders press fit into the piston platforms. A gap exists between the inner wall of an outer cylinder and the outer wall of an inner cylinder. At the top and bottom of each cylinder, there is a flow distributor 0.64 mm deep and 0.32 cm wide around the cylinder circumference. Each half of the hot piston platform contains one cylinder, 2.14 cm in length with a diameter of 3.80 cm, which provides a total heat transfer area of Each half of the intermediate piston platform contains two cylinders that provide a total heat transfer area of The first cylinder is 3.97 cm in length with a 4.11 cm diameter while the second cylinder is 4.88 cm in length with a 3.52 cm diameter. Each half of the cold piston platform contains four cylinders that provide a total heat transfer area of The cylinders are 6.71 cm in length and have diameters of 4.44 cm, 3.92 cm, 3.41 cm, and 2.89 cm. The total volume devoted to isothermal heat exchangers in the SSR is ( per SSR half). To ensure that fluid can not flow directly from the hot platform to the cold platform, the gaps created by the two nested copper cylinders of the intermediate platform heat exchanger form two distinct passages with heat transfer areas of and respectively. In the half of the SSR corresponding to the top pistons of each platform, fluid coming from the hot platform flows through the first heat exchanger passage which corresponds to the outer gap and into the intermediate platform piston. After mixing in the intermediate piston, fluid then passes through the second heat exchanger passage which is formed by the inner gap and into the bottom recuperator. In the bottom SSR half, the flows are reversed; fluid from the hot platform flows first through the second heat exchanger passage before entering the intermediate platform piston. The recuperator used in this SSR is a new design made of plastic. As shown in Fig. 2, the recuperative portion of the heat exchanger consists of alternating layers of Kapton17 film and Kapton film glued together using Stycast 1266.18 Each layer has five passages 2.38 mm in width and 20 cm in length. In the upper recuperator, ten layers, separated by nine layers, form 50 flow passages (25 per SSR half) which are arranged in a counterflow heat exchanger pattern. The bottom recuperator consists of sixteen layers separated by fifteen layers and has 80 flow passages (40 per SSR half). The total volume of the upper recuperator is of which only ( per SSR half) is devoted to recuperative heat transfer. The total volume of the lower recuperator is of which only ( per SSR half) is devoted to recuperative heat transfer. A complete description of the construction and design of this type of heat exchanger will be given by Patel and Brisson.16 The total volume of the SSR including the mixture trapped in the Vycor superleaks is Excluding the volume trapped in the Vycor, the SSR's volume is 120 cm3 ( per SSR half). The two fill lines into each of the SSR halves are sealed at low temperature by valves mounted on the hot platform. These valves are actuated manually from room temperature and
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Figure 2. The arrangement of alternate layers of Kapton film within the recuperator to form a counterflow heat exchanger.
are needed to prevent the mixture from moving up and down the fill capillaries during the operation of the SSR. Such a movement would put a significant heat load on the evaporation refrigerator and would also decrease cooling by reducing the magnitude of the pressure oscillation within the SSR. Calibrated ruthenium oxide19 and germanium20 thermometers mounted on the outside of the piston platforms are used to monitor the temperature. The precision of our temperature measurements is ± 0.67 mK at 1.0 K and ± 1.02 mK at 300 mK. Cooling powers are measured by monitoring the voltages across and currents through heaters made of wound manganin wire mounted on the cold and intermediate piston platforms. The precision of our cooling power measurements is
PROCEDURE AND RESULTS
The SSR was prepared for operation by first cooling the refrigerator to 1.0 K, then centering the pistons on each platform to ensure equal volumes of working fluid in each SSR half, and
finally filling the refrigerator with a 3.0% mixture. The fill lines to the SSR were then closed and the SSR was operated at various speeds using a hot piston stroke of 1.00 cm (17.7 volume displacement) and a intermediate/cold piston stroke of 0.98 cm ( and 7.53 volume displacements) to find a minimum temperature of 282 mK at a speed of one cycle every 25 seconds. We then measured cooling power as a function of operating speed and temperature. This was done by measuring the average cold and hot piston temperatures during the cycle at each operating speed while supplying a constant heat load to the cold piston
platform. The average piston temperature was determined by averaging the maximum and minimum temperature of the platform during a cycle. For all of the platforms, the peak to peak temperature difference during a cycle did not exceed 13 mK under any of the operating conditions. Typical
values were 6 mK, 7 mK, and 8 mK respectively for the cold, intermediate, and hot piston temperatures.
Figure 3 provides a map of the performance of the SSR. As in previous SSR's, Fig. 3a shows that there is an optimal operating speed for a given cooling power to minimize the cold piston temperature. Operating from a high temperature of 1.0 K, the SSR achieves a low
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temperature of 282 mK and delivers net cooling powers of 1mW at 617 mK, at 485 mK, and at 361 mK. Figure 3b provides the corresponding intermediate piston and hot piston temperatures for data shown in Fig. 3a. The figures show that the hot piston temperature is kept almost constant for a given operating speed by the evaporation refrigerator. The
Figure 3. (a) Cold piston temperature versus cycle period for constant cold piston cooling powers. (b) Intermediate piston temperature and hot piston temperature versus cold piston temperature for data given in Figure 3a. The dotted lines represent the intermediate piston temperature while the solid lines represent the hot piston temperature for cycle periods of 10 seconds
15 seconds
25 seconds
and 40 seconds
Figure 4. (a) Intermediate piston temperature versus cold piston temperature for various constant cooling power combinations of the cold and intermediate pistons using a 15 second cycle period. From bottom to top, the solid lines represent heat loads on the intermediate piston of 0 mW, 0.5 mW, and 1
mW. From left to right, the dotted lines represent heat loads on the cold piston of 300 1 mW, 1.5 mW, and 2 mW. (b) Hot piston temperature versus cold piston temperature corresponding to the data given in Figure 4a.
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intermediate piston temperature varies from 550 mK to 890 mK with cycle frequency, but it is almost constant with cold piston temperature. Figures 4a and 4b provide a performance map of the SSR operating at a constant 15 second cycle period but with different combinations of heat loads on the intermediate and cold pistons. Figure 4a shows that the intermediate piston temperature does not significantly vary with
different cold piston heat loads and a constant intermediate piston heat load. Together with the high intermediate piston temperature, this result indicates that the upper recuperator of our SSR is undersized. Basically, the heat load due to imperfect recuperation of the upper recuperator is so large that the increased heat load on the intermediate platform from the heating of the cold
piston platform does not significantly affect the intermediate piston temperature. With a properly sized upper recuperator, we expect that the intermediate piston temperature will rise at a slightly slower rate than the cold piston temperature as the cold piston heat load is increased. Figure 4a also shows that the cold piston temperature for a given cold piston cooling power changes only slightly with the different intermediate piston temperatures and heat loads. This result indicates that the lower recuperator is adequately sized. The temperature ratios between the hot piston and intermediate piston and between the intermediate piston and the cold piston provide another indicator that the upper recuperator is undersized and the lower recuperator is adequately sized. Our experience with single stage SSR's suggests that with properly sized recuperators and no heat loads on the SSR, the temperature ratios between hot and intermediate pistons and between intermediate and cold pistons should be between 2.0 and 3.5, depending on the operating speed and the mixture concentration. From the temperatures given in Fig. 3 and Fig. 4, the temperature ratio between hot and intermediate pistons at the faster operating speeds is approximately 1.3 while the temperature ratio between the intermediate and cold pistons is approximately 2.2. The upper recuperator of our two stage SSR is clearly undersized while the lower recuperator is adequately sized. Despite the inadequate upper recuperator, the experimental performance of our two stage SSR is encouraging. The low temperature experimental performance of this two stage SSR is equivalent to the best results of previous single stage SSR's. Compared to the single stage SSR, the two stage SSR provides slightly smaller cooling powers at low temperatures ( at 361 mK for the two stage SSR versus at 342 mK for the single stage SSR) but achieves a lower ultimate temperature (282 mK versus 291 mK). Based on the experimental performance of this SSR, we believe that replacing the upper recuperator of this SSR with a large Kapton recuperator (having devoted to recuperative heat transfer) will significantly increase the low temperature cooling powers of the two stage SSR and allow an ultimate low temperature of 230 mK. CONCLUSIONS
This paper describes the first operational two stage SSR and reports its preliminary experimental performance. This two stage SSR has a total internal volume of and uses and Kapton heat exchangers. Operating from a high temperature of 1.0 K and with a 3.0% mixture, this SSR achieves a low temperature of 282 mK and delivers net cooling powers of 1 mW at 617 mK, at 485 mK and at 361 mK. The experimental performance of this two stage SSR is equivalent to the best low temperature results of previous single stage SSR's. We expect to significantly improve upon the performance of this two stage SSR by using a larger Kapton upper recuperator. ACKNOWLEDGMENTS
This work has been supported by National Science Foundation grant CTS-9416689.
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REFERENCES
1.
Watanabe, A., Swift, G. W. and Brisson, J. G., "Uniform temperature cooling power of the superfluid Stirling refrigerator", Journal of Low Temperature Physics, vol. 103, (1996), pp. 273293.
2.
Kotsubo, V. and Swift, G. W., "Superfluid Stirling-cycle refrigeration below 1 Kelvin", Journal of Low Temperature Physics, vol. 83, (1991), pp. 217-224.
3.
Brisson, J. G. and Swift, G. W., "Superfluid Stirling refrigerator with a counterflow regenerator", Proceeding of the Seventh International Cryocoolers Conference, Santa Fe, NM, (1992), pp. 460465.
4.
Brisson, J. G. and Swift, G. W., "A recuperative superfluid Stirling refrigerator", Advances in Cryogenic Engineering, vol. 39B, (1994), p.1393.
5.
Brisson, J. G. and Swift, G. W., "High Temperature Cooling Power of the Superfluid Stirling Refrigerator", Journal of Low Temperature Physics, vol. 98, (1995), p. 41.
6.
Brisson, J. G. and Swift, G. W., "Measurements and modeling of a recuperator for the superfluid Stirling refrigerator", Cryogenics, vol. 34, (1994), p. 971.
7.
Watanabe, A., Swift, G. W. and Brisson, J. G., "Measurements with a recuperative superfluid
Stirling refrigerator", Advances in Cryogenic Engineering, vol. 41, (1996), p. 1527. 8.
Patel, A. B. and Brisson, J. G., "Experimental performance of a single stage superfluid Stirling refrigerator using a small plastic recuperator", Journal of Low Temperature Physics, vol. 111, (1998), pp. 201-212.
9.
Patel, A. B. and Brisson, J. G., to be published.
10. Kornhauser, A. A., "Gas-wall heat transfer during compression and expansion", Ph.D. Thesis, Massachusetts Institute of Technology, (1989), pp. 15-17. 11. Sun, Z. F. and Carrington, C. G., "Oscillating flow modeling of Stirling cycle cryocooler", Advances in Cryogenic Engineering, vol. 41, (1996), pp. 1543-1550. 12. Patel, A. B. and Brisson, J. G., "High cooling power superfluid Stirling refrigerator", Proceedings of the 1996 NASA/JPL Low Temperature Microgravity Physics Workshop, Edited by D. Strayer and U. Israelson, NASA Document D-13845, (1996), pp. 46-50.
13. Patel, A. B. and Brisson, J. G., "Theoretical performance of single stage and two stage superfluid Stirling refrigerators using Kapton recuperators", Advances in Cryogenic Engineering, vol. 43, (1998). 14. Types 60055-1, 60050-1, and 60035-2 from Metal Bellows Division, Senior Flexonics Inc., Sharon MA. 15. Document PI-VPG-91, Corning Glass Works, Coming, NY. 16. Patel, A. B. and Brisson, J. G., "Design and construction of a plastic heat exchanger for sub-Kelvin use", to be published. 17. Dupont High Performance Films, Circleville OH.
18. Grace Specialty Polymers, Emerson and Cumming Inc., Lexington MA. 19. Oxford Instruments Ltd., Oxford, England. 20. Lake Shore Cryotronics, Inc., Westerville, OH.
Investigation of Microscale Cryocoolers J.M. Shire, A. Mujezinovic, and P.E. Phelan
Arizona State University Department of Mechanical and Aerospace Engineering Tempe, AZ 85287-6106
ABSTRACT
With the advent of micro electrical mechanical system (MEMS) technologies and other advanced manufacturing techniques comes the opportunity to fabricate devices on a smaller
scale. One area of current interest to the microscale community is thermal systems. Many components of thermal systems have been constructed including motors, valves, pumps, nozzles, combustors, and heat exchangers. Although the pieces are in place, they have not been combined to form complete thermal systems. This paper compares the results of four cryocooler models reduced to small length scales and aspects of microcooler manufacturing. A brief review is given of candidate systems for miniaturization. Three pulse tube models and one Stirling cycle model are discussed. The manufacturing aspects of the project are deliberated and the conclusion is that stainless steel is a potential construction material. Results detail the capacities of the coolers and the minimum temperatures they can achieve as their length scale decreases. Results demonstrate that a net refrigeration of 300 mW may be possible at a length of approximately 5 mm, and at a temperature of 71 K. INTRODUCTION
The idea of small cooling devices is not new. Little discusses small Joule-Thomson devices that operate at temperatures of 65 K.1 These devices run as small as 1.0 cm diameter and 0.1 cm thick. The disadvantage of such a system is that it requires a supply of very high pressure (10-20 MPa) compressed gas to operate. Once the gas supply is exhausted, the device is done. The equipment required to produce the gas has a length scale significantly larger than the small Another form of small scale coolers is a thermoelectric device using the Seebeck effect to produce cooling. These devices can be very small, on the order of integrated circuits, and have
no moving parts. Unfortunately, low capacities and high minimum temperatures reduce their potential.2 The focus of this work is scaling down existing cryocoolers. The traditional vapor compression cycle has been studied in many works, and a vast volume of information is available on them. Vapor compression cycles have the disadvantages of requiring a high pressure
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ratio to operate and not reaching the cryogenic temperature range. Our attention is therefore turned to two less common refrigeration systems, the pulse tube refrigerator and the Stirling cycle.
Some work exists on small Stirling heat engines which are essentially Stirling coolers operating in reverse. Nakajima et al.3 constructed a Stirling heat engine designed to act as a small actuator with a swept piston volume of only It produced 10 mW of power at an operating frequency of 10 Hz. Experiments have been carried out on small pulse tubes with external compressors. Xu et al. constructed a 50-mm-long coaxial pulse tube which achieved a low temperature of 159.4 K.4 In general, coolers involve a compressor somewhere in their workings. An important step in the development of small scale coolers will be creating a small compressor incorporated with the system, instead of relying on an outside source of compression. As the devices become smaller, the pressure ratio that the compressor is capable of generating is likely to decrease. This will be compensated for by operating the smaller devices at higher frequencies to maintain capacity. A common misconception about pulse tubes is that a compressor must generate a continuous supply of compressed gas, the stream exergy of this gas being expended as the device operates. All that is actually needed is a way to generate pressure and mass oscillations within the tube. This could be accomplished, for example, by an oscillating membrane, which is considerably simpler than a compressor. Much work on oscillating membranes has been accomplished in the
field of silicon micropumps.5 This knowledge base should easily be applied to the current application. MODELING
The results of four different models are compared in this paper; the nomenclature used is defined in Table 1. The first model is de Boer’s analysis of the basic pulse tube refrigerator (BPTR) illustrated in figure 1a.6 The basic pulse tube refrigerator (BPTR) was invented in the 1960’s by Gifford and Longsworth.7 It is a relatively simple device compared to other cryocoolers. An oscillating piston at one end creates mass and pressure oscillations in the tube. Heat is absorbed and released by the tube walls, creating a surface heat pumping effect from the cold end to the hot end. The cold-end heat exchanger removes heat from the load, and the hotend heat exchanger rejects this heat to a sink. The major advantages of the pulse tube are that it has only one moving part, thereby increasing reliability. There are no moving parts at low temperatures, avoiding the common problems with seals that plague other cryocoolers. No exotic refrigerants are required to operate the cycle; air is a
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Figure 1. System schematics. (a) Basic Pulse Tube Refrigerator (BPTR) (b) Basic Pulse Tube Refrigerator with Regenerator (c) Orifice Pulse Tube Refrigertor (OPTR) (d) Stirling Cycle
Refrigerator.
suitable working fluid, although helium is preferred because of its higher specific heat ratio and other thermophysical properties. The pulse tube refrigerator proves difficult to model for several reasons. The working fluid
does not undergo a single thermodynamic cycle. Rather, each element of gas undergoes a cycle slightly different than its neighbors. The phenomenon of surface heat pumping causes the refrigeration effect. Each gas molecule completes a cycle of being compressed and heated, giving heat to the wall, expanding and cooling, and absorbing heat from the wall. This generates a net flow of heat from the cold-end heat exchanger to the hot-end heat exchanger. In the model, a finite number of “shuttles” is assumed. A shuttle is a packet of gas undergoing the cycle. By choosing important parameters such as the non-dimensionalized length of the heat exchangers, pressure ratio, and operating frequency, de Boer’s model is able to calculate the COP, the work per cycle, and the refrigeration capacity. There is no simple formula to calculate the quantities,
rather a mathematical process is required. The next major improvement to the pulse tube was the implementation of a regenerator allowing for the storage and release of heat during alternate phases of operation. de Boer added another facet to his model when he accounted for the existence of a regenerator in the pulse tube (figure 1b).8 This is a realistic addition because practical tubes generally are equipped with regenerators, de Boer demonstrated that the regenerator improves the refrigeration capacity of the refrigerator by a factor of:
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Mikulin introduced the orifice pulse tube refrigerator (OPTR) in 1984.9 The OPTR uses an orifice and reservoir to shift the phase of the pressure, mass, and temperature oscillations,
drastically increasing the refrigerating capacity of the device. The third model examined is Storch and Radebaugh’s10 OPTR model (figure 1c). Storch and Radebaugh attack the problem by combining basic thermodynamic analysis with phasors to track the time-varying properties of mass, pressure, and temperature throughout the tube. After making assumptions such as small pressure ratios and fixed temperature heat exchangers, the following relationship is obtained for the gross refrigeration power of the device:10
The Stirling cycle is the most common form of regenerative cycle. Although more exotic arrangements are possible, the typical Stirling device consists of two heat exchangers, a compressor, a displacer, and a regenerator (figure 1d). The cycle involves isothermal
compression and expansion, and constant volume cooling and heating. The fourth model
considered is Peterson and Al-Hazmy’s Stirling cycle analysis.11 It includes the basic thermodynamic analysis of the Stirling cycle. Their major improvement on existing models is that conductive heat transfer through the regenerator is considered. As the scales of devices decrease, phenomena that were previously negligible become dominant, such as this parasitic
conductive heat loss. These phenomena eventually grow to the point where they impose a lower limit on the size of the device. The minimum length scale for a Stirling cooler is shown to be:11
Peterson and Al-Hazmy demonstrated that a Stirling cooler with a length scale less than one
centimeter may operate with a positive net refrigerative effect. Similar to the Peterson and Al-Hazmy11 analysis, here we incorporate conductive heat losses along the tube walls with the above pulse tube models, and observe this effect on their refrigeration capacity at small length scales. A linear heat transfer model given by:
is used.
The ratio of wall heat transfer area to tube area is assumed to be 1:10. An absolute pressure of 3 bar is fixed for each model and the devices are all assumed to operate at 500 Hz. Helium is the working fluid of choice and a representative specific heat ratio of 1.66 is selected. The results of these models are shown and discussed in the next section. MANUFACTURING ASPECTS Material selection is a critical part of the design phase. Selecting the material with which to
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Figure 2. Material Thermal Conductivity Comparison
build the pulse tube limits the manufacturing methods to be used. Some important considerations are thermal conductivity, specific heat, and machineability. Thermal conductivity is a very important factor when designing these devices on small scales as parasitic heat losses eventually become a dominating force. Low thermal conductivities are desirable because they minimize these losses. Figure 2 illustrates the thermal conductivity of several candidate materials, silicon, gallium arsenide, stainless steel 304, polyvinyl chloride (PVC), and silicon dioxide, over a range of temperatures. Because the goal is to operate a device down to the cryogenic range, it is important to consider the thermophysical properties at low temperatures.
Silicon is a good material because of all the technology associated with its manufacturing on a small scale. Unfortunately it has a relatively high thermal conductivity, increasing even further at low temperatures. PVC has a very low conductivity, but little information is available about its manufacturing and processing on small scales. Stainless steel appears to be a good
compromise. Its conductivity continues to decrease with temperature and its manufacturing is well understood. For this reason stainless steel is chosen as the manufacturing material of the
devices and its properties are used for the remainder of this paper. RESULTS
Figure 3 details results with Storch and Radebaugh’s orifice pulse tube refrigerator (OPTR) model.10 At short lengths, approximately 3 mm, conduction through the tube walls is the
dominant force. It is an order of magnitude greater than the refrigeration, counteracting all useful effects of the device. As the tube lengthens, the temperature gradient from the hot end to the cold end is decreased, making the conduction weaker. Eventually the refrigeration far
surpasses the conduction and a reasonable net refrigeration is obtained. At a tube length of 9 mm, the conductive losses are approximately equal to the refrigeration. At a length of 2.4 cm, their magnitude is only 10% of the refrigeration. This particular simulation has the tube operating between a hot end temperature of 293 K and a cold end temperature of 71 K. Figure 4 again uses Storch and Radebaugh’s OPTR model. An important parameter for cryocoolers is the lowest temperature they can achieve. This limits their applications. A major factor in performance is often the ratio of the maximum pressure to the minimum pressure the device receives. Figure 4 plots the temperature ratio versus pulse tube length for three different pressure ratios. The curves represent a COP of zero. The refrigeration from the device is just matching the conductive losses through the tube walls, giving a net refrigeration of zero. Low pressure ratios are used because the compressor design is possibly the most challenging aspect of
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Figure 3. Orifice Pulse Tube Refrigerator Analysis
Figure 4. Orifice Pulse Tube Cycle Limits .
these small-scale coolers. They are not expected to achieve high pressure ratios. To go from a room temperature of 20 °C to a low temperature of –196 °C (77 K) is equivalent to a temperature ratio of 0.263. As expected, with an increasing pressure ratio, lower temperatures may be
reached at smaller tube lengths. The second curve, for a pressure ratio of 1.5, indicates that a temperature ratio of 0.2 is attainable at approximately 10 mm. Figure 5 compares four different analytical cryocooler models: de Boers basic pulse tube,6 de Boers pulse tube with regenerator,8 Storch and Radebaugh’s orifice pulse tube,10 and Peterson and Al-Hazmy’s miniature Stirling cycle.11 The curves represent the net refrigeration after the conductive losses have been subtracted. All four devices operate between 293 K and 71 K. Judging from the results of these four analytical models , Storch and Radebaugh’s OPTR is clearly the best system for the miniature design. The Stirling cycle refrigerator and the BPTR with regenerator are fairly close to one another, although the simplicity of the BPTR with regenerator makes it a better choice. The BPTR without regenerator is barely visible on the
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Figure 5. Cryocooler Cycle Comparison.
graph because of its low refrigeration power. One consideration that must be made before whole-heartedly endorsing the OPTR is that its design on small scales will not be trivial. The orifice typically contains a needle valve that must be adjusted to optimize system operation. CONCLUSIONS
This work demonstrates the feasibility of small scale cryocoolers. The orifice pulse tube form is the most viable option because of its superior performance and relative simplicity. The next step in the development process should be experimental work, particularly involving the small-scale compressors. ACKNOWLEDGEMENTS
One of the authors (P.E.P.) gratefully acknowledges the support of the National Science Foundation through an NSF CAREER Award (Grant No. CTS-9696003). The authors wish to thank Dr. de Boer for his assistance. REFERENCES 1
Little, W.A., “Advances in Joule-Thomson Cooling,” Advances in Cryogenic Engineering, vol. 35 (1990), pp. 1305-1314. 2 Mino, C.C., Cochrane, J.W., Volckmann, E.H., and Russell, G.J., “Cryogenic Thermoelectric Cooler with a Passive Branch,” Journal of Electronic Materials, vol 26, no. 8 (1997), pp. 915-921. 3 Nakajima, N., Ogawa, K., and Fujimasa, W., “Study on Microengines: Miniaturizing Stirling Engines for Actuators,” Sensors and Actuators, vol 20 (1989), pp.75-82. 4
Xu, M., He, Y., Wu, P., and Chen, Z., “Experimental Research of a Miniature Pulse Tube Refrigerator Using Nylon Tube,” Cryogenics, vol. 36, no.2 (1996), pp. 131-133. 5
Zengerle, R., Richter, M., Brosinger, F., Richter, A., and Sandmaier, H., “Performance Simulation of Microminiaturized Membrane Pumps,” Proc. Transducer ’93, Yokohama, Japan (1993), pp. 106-109. 6
de Boer, P.C.T, “Thermodynamic Analysis of the Basic Pulse-Tube Refrigerator,” Cryogenics, vol. 34, no.9 (1994), pp. 699-711.
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7 Gifford, W.E., and Longsworth, R.C., “Pulse-Tube Refrigeration,” Trans ASME:J Eng Ind Series B (1964) 86, pp. 264-268. 8
de Boer, P.C.T., “Analysis of Basic Pulse-Tube Refrigerator with Regenerator,” Cryogenics, vol. 35 (1995), pp. 547-553.
9
Mikulin, E.I., Tarasov, A.A., and Shkrebyonock, M.P., “Low Temperature Expansion Pulse Tubes,” Advances in Cryogenic Engineering, vol. 29 (1984), pp. 629-637. 10
Storch, P.J., and Radebaugh, R., “Analytical Model for the Refrigeration Power of the Orifice Pulse Tube Refrigerator,” NIST Tech. Note 1343, National Institute of Standards and Technology, Boulder, Colorado, (1990). 11
Peterson, R.B., and Al-Hazmy, M., “Size Limits for Stirling Cycle Refrigerators and Cryocoolers,” 1997 IECEC.
Development of Advanced Cryogenic Integration Solutions D. Bugby and C. Stouffer Swales Aerospace Beltsville, Maryland 20705 T. Davis, Lt. B. J. Tomlinson,
and Lt.M. Rich
Air Force Research Laboratory Kirtland Air Force Base, New Mexico 87117 J. Ku and T. Swanson NASA Goddard Space Flight Center Greenbelt, Maryland 20771 D. Glaister The Aerospace Corporation Kirtland Air Force Base, New Mexico 87117
ABSTRACT This paper describes the development of three advanced cryogenic integration devices for future space-based cryogenic systems. The first device is the cryogenic capillary pumped loop (CCPL), a miniaturized two-phase fluid circulator for thermally linking cryogenic cooling sources to remote cryogenic components. Two other devices, a nitrogen triple-point cryogenic thermal storage unit (CTSU) and a hydrogen gas-gap cryogenic thermal switch (CTSW), are also described. Each of the three components will fly on STS-95 in October 1998 as part of the CRYOTSU Flight Experiment. This paper describes the design, operation, flight configuration, and ground test results of each device and provides discussions of several new concepts for cryogenic integration. INTRODUCTION Cryogenic integration is a largely ad-hoc process that depends almost exclusively on the use of flexible conductive links (FCLs) to couple cryocoolers to components. This approach, while effective for short transport distances and low heat loads, will simply not work for future systems that may involve, for example, the transport of cooling across a two-axis gimbal. In order to meet such requirements, an overall initiative to develop new and enabling cryogenic integration technologies is needed. This paper describes such an effort, dubbed the Integrated Cryogenic Bus (ICB), and focuses on the design, operation, and ground testing of three specific ICB devices. Cryocoolers 10, edited by R. G. Ross, Jr. Kluwer Academic/Plenum Publishers, 1999
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Definition of an Integrated Cryogenic Bus (ICB)
An ICB is an optimized thermal control system comprised of one or more cryogenic heat transport devices and various auxiliary components. Heat transport device options include the
CCPL, the cryogenic loop heat pipe (CLHP), the cryogenic flexible diode heat pipe (CFDHP), the cryogenic pumped gas loop (CPGL), and the traditional FCL. Auxiliary components include the CTSU for load-leveling, the CTSW for diode action, multi-layer insulation (MLI) for low radiation parasitics, and kevlar cable systems for low conductance structural supports. Fluid circulation-based heat transport devices offer high flexibility (using bellows lines), small size, low weight, and diode action. Two-phase systems with small heat exchangers (such as the CCPL and CLHP) offer the most potential for miniaturization and, therefore, optimum performance. SBIRS-Low Application for an ICB
From an applications standpoint, SBIRS-Low (the successor to DSP) may be the first system to benefit from ICB technology. For this system, an ICB featuring a CCPL (or CLHP) and several CTSUs is being designed to transport 15 W of cooling across a two-axis gimbal. Figure 1 illustrates the concept. However, before this system can be implemented, CCPL and CTSU functionality in a zero-g environment must be demonstrated. This flight demonstration, the first ever for a CCPL, will take place on STS-95 in October 1998 as part of the CRYOTSU Flight Experiment. Flexible Conductive Links Cannot Meet Future Requirements
To demonstrate why an FCL alone will not work for future applications, Figure 2 shows the parasitics and mass of a cylindrical conduction bar (CB) as a function of transport length and bar material for an operating temperature of 100 K, a heat load of 10 W, and a of 5 K. As indicated in the figure, as transport lengths increase beyond about 0.50 m, the mass and parasitics of a CB become excessive. An FCL of equal conductance would be even heavier and have even higher parasitics, thus it is not viable for long transport length cryogenic applications. Organization of Paper
The remainder of this paper is organized as follows. First, the scope of the problem is defined and important background information is provided. In this section, descriptions of the applications, requirements, and cryogenic integration issues are provided. Next, the various ICB
solution options are described. Also included in this section is a description of existing ICBs used on DSP, AIRS, NCS, and others.1 Next, the three specific ICB devices that are central to this paper -- the CCPL, CTSU, and CTSW -- are described in detail. Lastly, the CRYOTSU flight experiment and the cryogenic integration approaches utilized therein, in particular the kevlar cable (low-conduction) suspension systems, are outlined and described.
Figure 1. SBIRS-Low integrated cryogenic bus (ICB) system.
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Figure 2. Parasitics and mass of a cylindrical conduction bar.
BACKGROUND
This section provides a discussion of the pertinent background information associated with cryogenic integration technology, including the important applications, the principal issues, and
the various cryogenic devices that are available.
Space Applications of Cryogenic Integration Technology The need for and application of cryogenic integration technology varies widely depending on the specific space system. The purpose of this subsection is to briefly discuss the integration technologies and emphasize where and for what application they are best suited and beneficial. Flexible Conductive Links (FCLs). The simplest cryogenic integration configuration is also the most common. This configuration involves the attachment of the cryocooler cold tip to
the cooled instrument interface via an FCL. This configuration is adequate when the cooler can be mounted near the instrument, there are no significant heat load variations, and parasitics from a redundant "off" cooler are not significant. The FCL is necessary to avoid side loads (due to launch vibration or differential material contraction) on the sensitive cryocooler cold-tip. The FCL also helps in the instrument-to-cooler alignment process. The most important integration technology for this configuration is the FCL. The current state-of-the-art is silver or aluminum
foil FCLs with a thermal resistance of about 2 K/W. Future improvements are possible by using graphite epoxy or carbon-carbon fibers that might reduce the thermal resistance by a factor of 2. Cryocooler Redundancy. For the near future, all low risk (non-experiment) cryocooler space applications will probably require redundancy due to the lack of statistical life test data on the mechanical, long-life cryocoolers. Cryocooler redundancy can significantly increase the cooling loads by adding parasitics from the non-operating or off cryocooler (which is attached at one end to ambient and at the other end to the cooled instrument). For typical Stirling or pulse tube cryocoolers, this additional heat load is about 0.3 to 0.5 W at 60 K. If the cryocooler has a capacity much larger than this load (with adequate margin), the simplest approach is to just absorb this extra heat load into the cryocooler operating budget. However, if margins are tight and/or the cryocooler has a small capacity relative to this off cryocooler load, the aforementioned ICB fluid circulator cryogenic devices can be very beneficial. Each of these devices has the capability to thermally isolate the off cooler from the instrument and reduce the off cooler parasitic heat loads by at least 3 to 5 times. It should also be mentioned that as the instrument temperature decreases, the cryocooler power efficiency becomes increasingly poorer and it becomes even more important to minimize heat loads. Thus, at 10 K, where a typical cryocooler capacity may only be 0.1 W, a thermal switching device may be essential for system feasibility.
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Variable Heating Loads. Another system characteristic that may require advanced
cryogenic integration technology is the variable heat load. The cryogenic heat load from an infrared instrument can come from several sources. For the focal plane, the heat load essentially consists of the electrical dissipation from the detector amplifiers (or readouts) and the thermal parasitics from the environment. Sources for the latter include conduction through wire leads or
radiation from the surrounding cavity. While the thermal parasitic load on a focal plane remains fairly constant, the electrical dissipation varies depending on whether the detector is powered on. For most systems with low-to-medium sized detector arrays, the electrical dissipation is usually relatively small and doesn’t have a large contribution to the total cryocooler heat load. Often for these small detector systems, the detectors are left "on" continuously. However, for larger detector arrays (or very cold ones such as those near 10 K), the electrical power can be very significant and sometimes can completely drive the total heat load. For these systems, there can be very significant system benefits in taking advantage of the fact that the detector is not continuously operating. By averaging the heat load, the cryocooler system can be sized to the average, instead of the maximum, instrument heat dissipation. The benefits of averaging the heat load also exist for applications where the cryogenically cooled optics are gimbaled. For these systems, the telescope or fore optics are usually gimbaled towards the target during operation, but are gimbaled away from the Earth and towards deep space during non-operating periods. This operating procedure results in lower, but variable, heat loads on the optics. Thus, for either large detectors or gimbaled optics, there are significant system benefits to averaging the heat loads. Thermal Storage. The aforementioned averaging can be accomplished using two principle cryogenic thermal storage options -- sensible heat storage or latent (phase change) heat storage.
The simplest approach is to absorb the peak heat loads using a sensible heat device such as a block of metal or a bottle of gas. While simple in design, a sensible heat device is not as weight efficient or as thermally stable as a phase change material (PCM) device. Depending on the temperature stability requirements, the operating temperature, and the materials used, a PCM device can be 10 to 50 times lighter than the sensible heat device. Using either thermal storage unit (TSU) option on a system with significant load variations can result in an effective system savings of over 100 times the mass of the TSU by reducing the cooling system size. Transport Length. The use of an FCL (alone) to interface one or more cryocoolers to an instrument is acceptable only when spacing permits the close placement of the coolers (cold tip
less than 0.15-0.25 m from instrument interface). Often the cryocooler must be a significant distance away from the instrument interface. This particular situation could be due to heat rejection requirements, a desire to minimize or isolate the instrument from the cryocooler motor
vibrations, or the need to cross a gimbaled joint. As indicated earlier, over any significant distance, the temperature drop associated with an FCL or even a solid CB quickly becomes excessive. This temperature drop drives down the cryocooler cold tip temperature and quickly increases the system power consumption. A CCPL, CLHP, CFDHP, or CPGL has the capability to transport heat from the instrument to the cryocooler over a significant distance (up to 1-2 meters) with a minimal temperature drop of up to 10 times less than a solid CB of the same mass. Gimbaled Optics. For the gimbaled optics case, there are often significant system savings to remotely mounting the optics cryocoolers. Any mass added to the gimbal (either from the cryocooler or its heat rejection radiator) is amplified by the additional mass necessary in the joint structure and motors. An FCL across the gimbal joint would produce both unacceptably large temperature drops as well as add significant torque resistance to the gimbal motors. A flexible heat pipe would also result in a significant torque loading on the motors. But, because there is no wick enclosed in the liquid and vapor lines, the CCPL, CLHP, or CPGL has the capability to use extremely thin (1.5 – 3 mm diameter) lines across the gimbal joint. Thus, these technologies could potentially be configured in a very flexible manner to produce minimal impact to the motors and enable the remote mounting of the cryocoolers off the gimbaled telescope.
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SOLUTION OPTIONS
As indicated earlier, an ICB provides an optimized thermal link between cryogenic cooling sources and cryogenic components. Although it is comprised of both heat transport devices and various auxiliary components, the "engine" of an ICB is the heat transport device. In general, a heat transport device carries heat via conduction or convection (fluid circulation). This section will focus on fluid circulation-based ICB systems. There are four principal types of fluid circulators to thermally link cooling sources to cooled components. These are: (1) cryogenic capillary pumped loop (CCPL); (2) cryogenic loop heat pipe (CLHP); (3) cryogenic flexible diode heat pipe (CFDHP); and (4) cryogenic pumped gas loop (CPGL). The first three devices are two-phase systems while the last is a single-phase system. To better understand the operation of the two-phase systems, a discussion of ambient temperature CPL technology is provided below followed discussions of each of the aforementioned devices. Ambient CPL Technology
Based on heat pipe principles, CPLs are passive thermal control devices capable of transporting heat over substantial distances with minimal temperature drop. A CPL is comprised of an evaporator, condenser, reservoir, transport lines and working fluid. Figure 3 shows a CPL with a 3-port evaporator, an approach that is typically employed in single evaporator systems. In this system, the reservoir and liquid lines connect directly into the evaporator core. Fluid circulation ensues when heat is applied to the evaporator, which contains a porous saturated wick. The liquid within the saturated wick evaporates, but it is immediately replenished via capillary action. Since only that which evaporates is replenished, the system self-regulates. The vapor then flows to the condenser before returning to the evaporator as liquid. To prevent evaporator bubble formation, the condenser must provide a few degrees of subcooling. In all CPLs, there is a thermostatically controlled, cold-biased reservoir that controls the saturation temperature. Variations in load (i.e., evaporator power) and condenser environment are handled passively. The fluid level in the reservoir rises or falls as needed to balance the saturation pressure as the vapor front in the condenser moves in or out depending on how the power and/or condenser environment are changing.
Cryogenic Capillary Pumped Loop (CCPL)
Although CCPLs and ambient CPLs are similar, the supercritical environment surrounding a CCPL necessitates two unique design features. The first is a liquid-cooled shield (LCS). The second is a "hot reservoir". A CCPL also has a normal CPL reservoir, referred to as the "cold reservoir". Figure 4 illustrates a typical CCPL plumbing arrangement. The LCS surrounds the liquid and cold reservoir lines of a CCPL to reduce parasitics and maintain subcooling. The source of LCS cooling is a section of condenser tubing that has been diverted from the condenser to the evaporator and back. The LCS is directly attached to this diverted section of tubing.
Figure 3. Ambient CPL with 3-port evaporator
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Figure 4. CCPL flow diagram.
The hot reservoir (which is actually an ambient temperature reservoir) serves two functions. The first is to lower the fill pressure of the system. A very high fill pressure would be required to provide the necessary mass of fluid for cryogenic temperature operation. The second is to aid in CCPL start-up. By appropriate plumbing and application of cold reservoir heater power, the working fluid can be shuttled between the reservoirs to cool the evaporator during start-up. As indicated in Figure 4, a CCPL needs five transport lines -- liquid, vapor, cold reservoir,
LCS out, and LCS return. For minimum ICB volume and maximum flexibility, the transport lines can be small diameter bellows lines (< 3 mm ID). One important advantage the CCPL has over cryogenic heat pipes is the higher pumping capability of the CCPL evaporator wick due to its small pore size. This feature enhances ground testability. Cryogenic Loop Heat Pipe (CLHP)
The CLHP is very similar to the CCPL, but there are two main differences. First, the CLHP cold reservoir (also called the "hydro-accumulator") is thermally linked to the evaporator,
whereas the CCPL cold reservoir is thermally linked to the condenser. Second, there is a heat pipe ("wicked") link between the CLHP hydro-accumulator and the evaporator. These features
give the ambient LHP the ability to self-start, whereas the ambient CPL and CCPL require a start-up sequence. Unfortunately, these features also make it impossible to start-up an unmodified CLHP. That is, because most cryogenic working fluids have critical temperatures below ambient, the CLHP evaporator and hydro-accumulator can never get cold enough, on their
own, to start-up. Fortunately, the CLHP can be modified to become a viable solution. Dual-Loop CLHP. The addition of a parallel LHP with an ambient working fluid can transform the CLHP into a viable solution. The additional LHP provides the means to cool the
CLHP evaporator and hydro-accumulator below the critical temperature upon which the CLHP will then self-start. Figure 5 illustrates a dual-loop CLHP flow diagram. As indicated in Figure 5, a dual-loop CLHP needs six lines -- liquid, vapor, LCS out, LCS return, ambient LHP liquid, and ambient LHP vapor. Like the CCPL, the transport lines can be very small diameter bellows lines and the high pumping head of the small pore size evaporator wick gives the CLHP a ground testability advantage over cryogenic heat pipes. Condenser-Mounted CLHP. Another option that can make the CLHP a viable solution is to mount the evaporator and hydro-accumulator on the condenser and apply heater power to generate a flow rate of vapor. Heat transport to a remote component is achieved by rerouting the exiting vapor back to the condenser and the resulting liquid on to the cooled component before returning to the condenser and evaporator. A miniaturized heat exchanger located at the cooled component extracts the cooling. Figure 5 illustrates the concept. This system requires two lines between the component to be cooled and the condenser (the LCS is no longer needed). The main drawback of this system is that it is only 50% efficient at best. To generate 1 W of cooling at the cryogenic component to be cooled, 1 W of heater power must be applied to the CLHP evaporator. Parasitics will decrease this efficiency even further.
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The efficiency of the condenser-mounted CLHP can be improved, at the expense of additional lines, by increasing the number of condenser and cooled component passes before returning the flow to the evaporator. One additional pass increases the number of lines to four and the maximum efficiency to 67%. Two additional passes increases the number of lines to six and the maximum efficiency to 75%. Figure 5 illustrates these various options. Due to the close proximity of the condenser and evaporator (which provides the pumping), condenser-mounted CLHP systems are insensitive to cooled component orientation during ground testing.
Advanced CLHP. The most promising CLHP system is the advanced CLHP. This device features a novel plumbing arrangement that uses a CCPL-type evaporator and a CLHP
evaporator/hydro-accumulator to reduce the number of lines between cooler and the cooled component to three and, at the same time, eliminate the LCS. This system auto-starts like a condenser-mounted CLHP (50% efficiency) and then operates like a normal high-efficiency CCPL. A LCS is not needed because any vapor bubbles entering the evaporator are flushed in a parallel loop driven by a small, intentionally applied heat load on the LHP evaporator that matches the parasitic load on the liquid line. The advanced CLHP has the same ground testability characteristics as the CCPL or dual-loop CLHP. This system is also illustrated in Figure 5.
Figure 5. CLHP flow diagrams
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Cryogenic Flexible Diode Heat Pipe (CFDHP) The CFDHP, as its name suggests, is a flexible heat pipe that can also function in a diode mode. The basic design of a CFDHP is similar to an ambient flexible diode heat pipe. The design consists of a stainless steel, over-braided bellows line and an internal wire-mesh wick that extends the entire length from the evaporator to the condenser. Diode action is achieved with a cold-biased reservoir. Two CFDHP devices were successfully flown and tested in two previous STS missions. The working fluids for these two flight CFDHPs were methane and oxygen. These particular CFDHPs were pressurized to around 13.8 MPa. Figure 6 conceptually illustrates the operation of a CFDHP. A CFDHP-based ICB system would require just one line between the cooling source and component to be cooled. However, due to the low pumping capability of the wire mesh, ground testing of this system would probably require the heat pipes to be run in a reflux mode. Cryogenic Pumped Gas Loop (CPGL) The most straightforward ICB heat transport device is the CPGL. A CPGL is comprised of a mechanical pump, a working fluid such as helium or neon, two single-phase heat exchangers (one at the cooling source and the other at the component to be cooled), and the necessary tubing. Figure 7 illustrates the concept. A few existing systems already utilize a CPGL including DSP and NCS.1-2 The CPGL is driven by a mechanical pump that circulates the gas in a continuous loop. Heat is carried by the gas from the component to be cooled to the cryogenic cooling source. The advantages of this system are its simplicity and its orientation insensitivity. The disadvantages include the parasitic heat added by the mechanical pump, moving parts in the pump, and the larger heat exchangers (HX) due to the single-phase nature of the heat transfer. A CPGL requires two lines between the cooling source and component to be cooled -- the outgoing line and the return line. To optimize system performance, the lines must be large enough for low pressure drop at the required mass flow rate, and small enough to minimize system parasitics, including radiation, conduction, and power dissipation in the pump.
Figure 7. CPGL flow diagram
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Existing ICB Systems
The ICB approach has been used in the past to solve a limited number of cryogenic thermal control problems that could not be solved with a simple FCL. Table 1 lists the characteristics of various existing ICB systems.1-2
DESIGN AND OPERATION OF ICB COMPONENTS
In this section, the flight unit designs and ground test results of the three ICB cryogenic components -- the CCPL, CTSU, and CTSW -- are detailed. As indicated, each of these three components will fly on STS-95 in October 1998 as part of the CRYOTSU Flight Experiment. Cryogenic Capillary Pumped Loop Flight Unit
The CCPL-5 flight unit is pictured in Figure 8 below.3 This flight unit was designed to bolt directly to the cold head of a Hughes 7044H tactical cryocooler. Within a very small flight envelope, a transport length of 0.25 m was effected by coiling the stainless steel, 1.27 mm inner diameter lines as shown in the figure. To ensure the system would perform properly, a test
program was carried out. That test program is described in the remainder of this subsection.
Test Objectives and Set-Up. The objectives of the performance test were to verify CCPL-5 functionality, identify the optimum charge pressure for flight, and evaluate the acceptability of the flight instrumentation. The test set-up is illustrated in Figure 9. As indicated in the figure, the CCPL-5 condenser was bolted to an OFHC copper bracket that simulated the Hughes 7044H cryocooler mounting interface. This bracket was bolted to the 2nd stage cold head of a G-M cryocooler. Also attached to the 2nd stage cold head was an OFHC copper strap that was connected to an OFHC copper bar bolted to the G-M cooler 1st stage cold head. The strap and bar were used to augment the cooling power of the 2nd stage for shorter cooldowns and enhanced test capability (such as the maximum power tests). As indicated in Figure 9, the tests
were conducted with the unit in a vertical orientation. That is, the tests were conducted with the evaporator completely above the condenser (evaporator flow axis parallel to the gravity vector).
Figure 8. CCPL-5 flight unit
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Figure 9. CCPL-5 ground test set-up
Test Procedure. The operational tests that were carried out with CCPL-5 consisted of the following: (a) start-up; (b) power cycling; and (c) high power. Each of these tests was carried out for a given system charge pressure. The charge pressures that were investigated were 1.86, 1.76, 1.65, 1.55, 1.45, and 1.34 MPa. The target condenser set-point for all tests was 80 K. Test Results. The test results are shown in Tables 2-4. Listed are the steady-state
temperatures of the evaporator condenser and cold reservoir as well as the measured system pressure and the saturation temperature corresponding to A plot of the results for the 1.34 MPa charge is provided in Figure 10. Overall, CCPL-5 performed very well, exhibiting stable behavior over a wide range of charge pressures and applied heater powers. The maximum power carrying capability of CCPL-5 ranged from 9-13 W.
Tables 2-4 indicate some apparently large values for the system temperature difference Since two-phase systems like CCPLs can operate with very small values, an explanation is warranted. The quantity is the sum of two terms: (equal to ) and (equal to ). The first term is a measure of subcooling. From a practical standpoint, 3-5 K of subcooling is needed to overcome parasitics and provide system robustness. One can limit by setting such that is the desired amount above If exceeds this ideal fill level, will be higher than necessary. For CCPL-5, a charge in the range 1.34-1.45 MPa seems to be the ideal match for a condenser set-point of 80 K. The second term
is a measure of evaporator effectiveness. It is a minimum when the
applied load is small. Table 3 shows that is only about 0.5 K for all charge pressures. At higher heat loads, is higher. In general, the magnitude of is a function of evaporator design. An evaporator body made of Al or Cu, instead of stainless steel (the CCPL5 construction material), could reduce In sum, proper choice of and an improved evaporator can likely lower to less than 5 K at 10-15 W loads.
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Cryogenic Thermal Storage Unit Flight Unit The cryogenic TSU flight unit is illustrated in Figure 11. The theory of CTSUs has been described elsewhere.4-5 The unit is a dual-volume system with a 140 cc cylindrical beryllium heat exchanger and a 16,760 cc spherical stainless steel storage tank. The system was filled to 0.61 MPa (88.5 psi) with nitrogen, which represents a charge of 117.8 gms. Since the vapor pressure of nitrogen at its triple-point is 0.012 MPa (1.8 psi), only 115.4 gms is available for phase change. With a liquid density of 0.87 gm/cc and a solid density of 1.02 gm/cc at the triplepoint, the heat exchanger void volume will be 95% filled with nitrogen when liquid and 80% filled with nitrogen when solid. With a heat of melting of 25.7 J/gm, the theoretical energy storage capacity of the flight CTSU is 2967 J.
Heat Exchanger Design. Figure 12 illustrates the design of the heat exchanger. As indicated, the heat exchanger is constructed of two halves into which are drilled a large number of 2.8 mm ID holes. To join the two halves and obtain a hermetic seal, several different methods including autogenous welding and diffusion bonding were tried unsuccessfully. The method that was successful is a process known as Hot Isostatic Pressure (HIP) bonding. In this process, the
two halves are pressed against each other at high pressure and temperature. After a prescribed amount of time, the beryllium surfaces bond to each other without any bond-line or seam. Figure 13 illustrates the seamless result. The figure also illustrates the brazing of a bimetallic (Ti-SS) fill tube into the side of the beryllium heat exchanger using an aluminum-based braze material.
Figure 10. CCPL-5 ground test results for 1.34 MPa charge pressure.
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Figure 11. Dual-volume cryogenic TSU flight unit.
Figure 12. Cryogenic TSU heat exchanger design.
Figure 13. Cryogenic TSU heat exchanger seamless bond and fill tube braze.
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Test Program. To verify the performance of the flight CTSU, a test program was carried out. Because of time constraints (this flight unit was completed a short time before it was integrated into the flight experiment), a very limited number of tests were performed. However, because of its similarity to previous single and dual-volume test units, there was a high degree of
confidence that it would perform successfully, which it indeed did. Test Set-Up. The test set-up is illustrated in Figure 14. In this test, the heat exchanger was cooled by a two-stage G-M cryocooler. A specially designed aluminum adapter bracket was used to mount the heat exchanger to the second stage cold head. The cooling power of the second stage was augmented by a thermal connection to the first stage as shown in the figure. Test Procedure. Although not shown in Figure 14, the tanks and CTSU heat exchanger were first evacuated with a turbo vacuum pump. The 2250 cc metering tank was then detached and filled with 100 gms of nitrogen. After reattaching the metering tank, the valves were opened and the system pressure rose to about 0.45 MPa (65 psi). Thus, 97.2 gms or 2498 J was available for energy storage due to the residual nitrogen vapor pressure in the tank of 0.012 MPa (1.8 psi) at 63.15 K. The cryocooler was then turned on to cool the system down. During cooldown, the
heat exchanger autonomously filled with nitrogen; no user interaction was required. When the heat exchanger temperature reached 60 K, it was fully charged with frozen nitrogen. Once charged, three full-discharge cycles (FDC) and two and one-half partial-discharge cycles (PDC)
were carried out with a heater power
of 5 W. A definition of these cycles is provided below.
A FDC starts with the working fluid fully-frozen, and is a full melt followed by a full
refreeze. Thus, melting and freezing begin with the working fluid in a single-phase state during a FDC. A PDC starts with the working fluid slightly melted, and is simply a partial melt followed by a partial refreeze. The working fluid stays entirely in the two-phase region during a PDC. Test Results. The results of the test program are illustrated in Figure 15. As indicated from
the FDC results, the durations of melting
and freezing
are roughly 13 and 20 minutes,
respectively. Using this information, the measured storage capacity and the net cooling rate can be computed from Equations 1 and 2. The resulting values are 2364 J (about 5% below
the fill mass value) and 1.97 W. Equations 1 and 2 can be derived from energy balances during melting and freezing. The data also show that the maximum temperature variation during a PDC,
which is how an actual component would operate with an attached CTSU, is only about 0.25 K.
Figure 14. Cryogenic TSU ground test set-up.
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Figure 15. Cryogenic TSU flight unit ground test results.
Cryogenic Thermal Switch Flight Unit
The cryogenic thermal switch (CTSW) flight unit is shown diagrammatically in Figure 16. The CTSW is a hydrogen gas-gap cryogenic thermal switch that was developed based on a JPL design.6 This section will describe the CTSW design, operation, and ground test program. Design and Operation. The CTSW is constructed of beryllium, titanium and stainless steel. The critical manufacturing step is the setting of the gas-gap, an extremely narrow (2 mil) flat slot that separates two beryllium cylindrical parts. The beryllium parts are held in place by a thinwalled titanium tube. A stainless steel bellows seals the system. A small diameter stainless line connects the switch to a hydride pump. The hydride pump, which is able to desorb or adsorb hydrogen depending on whether a small heater on the hydride pump is powered on or off, is what actuates the CTSW either on or off. The mass of hydrogen required in the CTSW is about 1 mg. The hydride pump was designed by JPL and its design was not altered.
Test Program. To separate the performance of the hydride pump from the intrinsic thermal performance of the CTSW, "on" testing was carried out by filling the switch with helium gas and "off" testing was carried out by evacuating the switch with a turbo pump. The performance goals were an on conductance greater than 1 W/K and an off resistance greater than 1000 K/W. The target operating temperatures were 30-140 K for the cold side and 30-300 K for the warm side.
Figure 16. Cryogenic thermal switch (CTSW) flight unit.
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Test Set-Up. The switch was mounted on a copper plate attached to the second stage of a two-stage G-M cryocooler. Figure 17 illustrates the CTSW test set-up. Due to manufacturing problems relating to the hermetic sealing of the various bi-metallic joints in the unit, and time constraints associated with the flight experiment, CTSW results with the hydride pump were not available for this paper. For this paper, only one cold end temperature (~140 K) was tested. Test Procedure. The "on" test was very straightforward and was accomplished as follows.
The switch was filled with 0.069 MPa (10 psi) of helium, the cold end was cooled to 140 K, and the steady-state temperature difference with no heater power was measured where is the temperature of the warm end and is the temperature of the cold end). This step provided the datum from which the silicon diode (relative) measurement errors and the effects of external parasitics could be eliminated. Then, 4 W of heater power was applied to the
warm end and the new
was measured. The "on" conductance was the heater power divided
by the change in temperature difference. Equation 3 provides the analytical relationship. The "off" test was less straightforward, owing to the comparatively large magnitude of the
external parasitic heat flow. This test was accomplished with a procedure that was developed over a number of tests with a non-hermetically-sealed unit. The switch was evacuated, the cold end was cooled to about 140 K, and heater power was applied to the warm end until the parasitics were "zeroed out". The off resistance was the resulting divided by the heater power. The parasitics were said to be zeroed out when the temperature of the warm end of the switch was equal to the temperature of the source of the parastics, the chamber wall. Equation 4 illustrates the analytical relationship.
Test Results. Table 5 lists the CTSW test results. The results indicate an "off" resistance of 1030K/W and an "on" conductance of 1.37 W/K. The resulting "switching ratio" is 1411. The performance of the switch with the hydride pump has not yet been measured, but it will probably be somewhat poorer.
Figure 17. Cryogenic thermal switch test set-up.
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CRYOTSU FLIGHT EXPERIMENT The CRYOTSU Flight Experiment packages the CCPL, CTSU, CTSW and an ambient temperature thermal storage experiment, the Phase Change Upper End Plate (PCUEP), into a Hitchhiker (HH) Get Away Special (GAS) Canister.7 This canister is known as the Cryogenic
Test Bed (CTB) and it provides five Hughes 7044H cryocoolers and associated electronics for carrying out cryogenic experiments in space. Each CTB split-Stirling cryocooler provides 3.5 W
of cooling at 80 K. Previous flights of the CTB were CRYOHP, CRYOTP, and CRYOFD.8-10 Due to the 100 W that each cryocooler dissipates, the CTB cannot run indefinitely with five (or even four) coolers turned on. So, the PCUEP was added both as an experiment and as part of the thermal control system. This device, with 600 W-hrs of energy storage capacity at 318 K, will extend the operating time of the experiment. Two axially-grooved ammonia heat pipes transport a portion of the waste heat to the PCM device. A diagram is provided in Figure 18. Cryogenic Integration Techniques To reduce conduction parasitics, the CTSU is suspended by 16 kevlar cables within a cryogenically-cooled, MLI-blanketed radiation shield. The radiation shield is a 1 mm thick 6061 aluminum box that surrounds the CTSU and CCPL. Two CTB cryocoolers are used to cool the CTSU, two are used to cool the radiation shield., and the fifth cools the condenser of the CCPL.
To further minimize parasitics, the radiation shield is suspended by kevlar cables that span between G-10 support posts mounted to the canister lid and G-10 spools that bolt to the outside of the radiation shield. The CTSW cold end is mounted to the bottom of the radiation shield while the hydride pump is mounted to the inside surface of the PCUEP.
Figure 18. CRYOTSU flight experiment (HH-G canister shown upside-down).
Objectives and Planned Flight Testing The principal objective of the CRYOTSU mission is to demonstrate the functionality of each
thermal control device in micro-gravity. For the CTSU, attaining this objective will require one or more ambient-to-cryogenic temperature cooldowns, followed by several FDCs to demonstrate its energy storage capacity and PDCs to demonstrate its temperature stability. For the CCPL, attaining this objective will entail one or more ambient-to-cryogenic temperature cooldowns and start-ups, followed by a battery of power cycling tests to compare with ground test data. For the
CTSW, several "on" and "off" tests will be carried out and compared to ground test data. For the PCUEP, a few FDCs and PDCs will be carried out to measure the energy storage capacity and temperature stability of the ambient PCM device.
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SUMMARY
This paper has described the design, operation, flight configuration, and ground test results of three advanced cryogenic integration devices for future space-based cryogenic systems. The cryogenic capillary pumped loop (CCPL), cryogenic thermal storage unit (CTSU), and cryogenic thermal switch (CTSW) were each successfully ground tested in preparation for an upcoming flight experiment in October 1998. These three devices, and several other cryogenic integration concepts described in the paper, are important parts of an overall initiative to incorporate new and enabling cryogenic technologies into space systems. This initiative, dubbed the Integrated Cryogenic Bus (ICB), endeavors to combine a range of cryogenic integration solutions to meet future requirements. Overall, this paper has attempted to elucidate the features and benefits of the ICB and of the advanced concepts and specific components upon which it is based.
ACKNOWLEDGMENT The authors would like to acknowledge the Air Force Research Laboratory (AFRL), the Ballistic Missile Defense Organization (BMDO), and the NASA Goddard Space Flight Center (GSFC) for funding the work described herein. The authors would also like to acknowledge the valuable contributions of R. Hagood, B. Marland, and E. Kroliczek of Swales Aerospace.
REFERENCES
1. Bugby, D., P. Brennan, T. Davis, et al., "Development of an Integrated Cryogenic Bus for Spacecraft Applications," Space Technology and Applications International Forum (STAIF96), Albuquerque, NM (1996). 2. Nellis, G., F. Dolan, W. Swift, H. Sixsmith, "Reverse Brayton Cooler for NICMOS," Cryocoolers 10, Plenum Press, New York (1998). 3. Bugby, D., T. Nguyen, E. Kroliczek, et al., "Development and Testing of a Cryogenic Capillary Pumped Loop Flight Experiment," 33rd Intersociety Engineering Conference on Energy Conversion (IECEC-98), Colorado Springs, CO (1998). 4. Bugby, D., R. Bettini, C. Stouffer, et al., "Development of a 60 K Thermal Storage Unit," Cryocoolers 9, Plenum Press, New York (1996). 5. Bugby, D., C. Stouffer, and Lt. M. Rich, "Experimental Verification of a 60 K Thermal Storage Unit," 32nd Intersociety Engineering Conference on Energy Conversion (IECEC97), Honolulu, HI (1997). 6. Johnson, D. and J. Wu, "Feasibility Demonstration of a Thermal Switch for Dual Temperature IR Focal Plane Cooling," Cryocoolers 9, Plenum Press, New York (1996). 7. Stouffer, C. and D. Bugby, "Cryogenic Thermal Storage Unit (CRYOTSU) Flight
Experiment," 32nd Intersociety Engineering Conference on Energy Conversion (IECEC-97), Honolulu, HI (1997). 8. Beam, J., P. Brennan, and M. Bello, "Design and Performance of a Cryogenic Heat Pipe Experiment (CRYOHP)," AIAA 27th Thermophysics Conference (1992). 9. Swanson, T., M. Buchko, P. Brennan, M. Bello, M. Stoyanof, "Cryogenic Two-Phase Flight Experiment; Results Overview," 1995 Shuttle Small Payloads Symposium, Camden Yards, Baltimore MD, September 25-28 (1995). 10. Thienel, L., P. Brennan, M. Buchko, M. Stoyanof, and D. Glaister, et al., "Design and Performance of the Cryogenic Flexible Diode Heat Pipe (CRYOFD) Flight Experiment," Paper 981583, SAE Conference, Boston, MA (1998).
Cold Accumulators as a Way to Increase Lifetime and Cryosystem Temperature Range V.T. Arkhipov, V.F. Getmanets, A. Ya. Levin, R.S. Mikhalchenko Special R&D Bureau (SR&DB) in Cryogenic Technologies Kharkov, Ukraine H. Stears Orbita Ltd, Kensington, MD, USA
ABSTRACT
The SR&DB has adopted the use of cold accumulators in conjunction with a cryocooler as the key means of achieving multi-year life cryogenic cooling systems. The approach involves cyclical operation of the cooler with low duty cycle, while the continuous cryogenic load is supplied by the
thermoaccumulator which contains a melting-freezing, evaporation-condensation, or other type of thermal phase transition material. The cryocooler is typically either a Joule-Thomson (J-T) throttle cycle based on mixed gases, or a split-Stirling cryocooler. Such cyclical operation increases system lifetime by up to 20 times. By changing the J-T working fluid and the phase change material it is possible to address a broad range of temperatures (4 - 150 K). Often the system can be upgraded to different operating temperatures without any changes in the cryocooler or accumulator hardware; only a replacement of the operating medium is required. INTRODUCTION
A key goal for future space missions is to increase the lifetime of the space vehicle and its components, including cryocoolers, up to 10-20 years and more. For this purpose we use cyclically operating J-T systems based on gas mixtures, or cyclically operating split-Stirling cryocoolers, which are thermally connected to a melting-freezing cold accumulator or accumulator containing some other type of thermal phase transition material1. A second advantage of this cryogenic system approach is ensuring temperature stability over the service life of the application; this is critical for many applications. The third important advantage is the application of gas mixtures2. By switching the operating medium of a J-T throttle system or accumulator, one can achieve the above-mentioned advantages at many discrete temperatures from 70 to 150 K. An additional advantage is minimal vibration and electromagnet interference in the vicinity of the cold load due to the ability to locate the compressor at a distant location. Lastly, the system can be easily adapted to many applications with minimal change to the cryocooler or accumulator hardware; instead, only the operating medium need be changed. This paper describes the experience of the SR&DB in applying this cooling system concept to applications in both the 70-150 K, and the lower 4-65 K temperature range. Cryocoolers 10, edited by R. G. Ross, Jr.
Kluwer Academic/Plenum Publishers, 1999
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Figure 1. Schematic of cooling system with a J-T cryocooler and cold accumulator: 1- inlet window; 2- IR-detector; 3- cryostat; 4,8,14- cold conductors; 5- superinsulation; 6- cold accumulator with heat conductor ribs; 7- suspension system; 9- evaporation heat exchanger; 10- throttle; 11- counterflow heat exchanger; 12- inlet and exit tubes; 13- heat radiator; 15- compressor.
CRYOCOOLERS WITH COLD ACCUMULATORS
The problem of achieving a substantial increase in space cryocooler lifetime was first taken on by the SR&DB about 20 years ago. The first solution developed1 involved the concept of a cryocooler mated with a melting-freezing accumulator of solid argon (see Fig. 1 and Table 1). Although this scheme involved a throttle cryocooler, it could also have been implemented with a
split-Stirling cryocooler (Fig. 2). The J-T cryocooler was selected due to a number of advantages 1,3-5. For example, this cryocooler is connected to the cold accumulator only by two thinwalled stainless steel tubes that provide low parasitic heat inflows to the accumulator. Also, the compressor of the throttle cooler can be located at a large distance from the coldhead, and the throttle part has no moving mechanical elements. As a result, minimal vibration and electromagnetic disturbance is generated in the vicinity of the cold load.
With the invention of nitrogen-hydrocarbon gas mixtures6 (created in the Ukraine in the 1970s) J-T cryocooler efficiency has increased and closely approached that of the best Stirling cryocoolers (see Table 1). In addition, such mixtures reduce the operating pressure down to 45 bar (instead of 200 bar for pure nitrogen, or 70-100 bar for nitrogen-freon mixtures); this in turn has reduced the compressor complexity and weight. We have found the most suitable systems for our applications to involve a lubricant-free compressor with piston-cylinder clearance seals that will not block cryocooler heat exchangers with cryodeposits. Also for our applications, a twostage compressor was created with pistons that move on special linear ball bearings 3,4 (Fig. 3). Compressors designed for mobile applications have two motors that rotate in opposite directions and colinear first- and second-stage pistons that move in opposite directions to create momentum cancellation. As a result, vibration of these compressors is reduced to minimal levels.
Figure 2. Schematic of cryostat system with a Stirling cryocooler and cold accumulator: 1- inlet window; 2- IR-detector; 3- superinsulation; 4,10,13- cold conductors; 5- heat conductor ribs; 6- cold accumulator; 7- suspension system; 8- cryostat; 9- cold finger; 11- copper tube; 12- displacer; 14- compressor; 15- heat radiator.
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Figure 3. Lubrication-free piston compressors created at SR&DB.
Tests at the SR&DB have shown that compressor lifetime is about 10,000 hours in either continuous operation or as the sum of repeatedly actuated on-off cycles. These results were the basis for the development of long-life space cryosystems which have also required the solution of
other complex problems including: 1) choice of optimal gas mixtures that address the particular compressor characteristics2; 2) design of a system of thermal fins to improve conduction within the melting-freezing accumulator7-10; 3) development of design and manufacturing techniques for highly effective cryocooler heat exchangers with minimum external and axial heat flows22; and 4) improvement of contamination removal technology for optical surfaces and achievement of minimum parasitic heat inflows through the accumulator and cryocooler superinsulation8.
The latter problem was previously solved at the SR&DB during the development of sublimation accumulators for cryogenic space application10-13. For these applications the laws of convective and contact heat transfer during sublimation and melting1,7,14,20 were investigated, the pro-
cesses of heat- and mass transfer in porous cryocrystals10,15 were investigated, and optimum designs of fixed and continuously rotated heat-transfer surfaces in sublimation cold accumulators11,12 were developed. This research, as a matter of fact, was also an inducement for the development of longer lifetime, more complex cryocoolers incorporating melting-freezing accumulators1. SUPERINSULATION AND CRYODEPOSITS During the last few years three main superinsulation problems have been solved, namely: 1) Achieving many years vacuum maintenance; achieved through the use of YCHT-10 (USNT10) spacers containing adsorbing fibers that cryopump the space between the MLI layers, and through the incorporation of new thermal-vacuum outgassing techniques 8,9,16,17 2) Elimination of water vapor cryodeposits on the superinsulation screens 3) Reduction of superinsulation blankets heat degradation associated with unequal temperature design optimization, and elimination of contact between layers. Besides cryodeposits internal to superinsulation, cryocooler surfaces, including heat exchangers and J-T throttle valves can also be degraded by condensable cryodeposits. The source of these deposits is vapor from oils diffusing from vapor-diffusion and ruffing pumps, and also low molecular weight constituents outgassing from cryocooler constructional materials. For monitoring the buildup of such contamination and cryodeposits, including identification of their thickness and composition we have developed and used three non-contact (optical) techniques. The method of an interference using a helium-neon laser is applicable for contamination layers with a thickness over 1 micrometer. For contaminant layers of a lessor thickness (down to
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Figure 4. A combination system for deep cooling.
several nanometers) inspection can be carried out using a laser ellipsometry method in the visible band. IR-spectrometry in the 2.5-25 micron range allows the identification of the source of contamination by specific absorption or reflection lines of films; calibration is achieved by view-
ing the condensed outgassing products of various known materials. The ellipsometry method can also be used for material identification by measurement of a film's refraction factor18,19. Accumulated data on refraction factors and optical spectra of the polymers most frequently used in cryostats helps in the identification of cryocontamination sources.
SOME SR&DB CRYOCOOLERS WITH COLD ACCUMULATORS
Table 1 presents SR&DB cryocoolers that have been used in space applications (Items 1-6) and in ground applications (Items 7,8). Most have been completed and have passed a full cycle of ground tests. The exception is cryocoolers N5 and N6, which are still in the development stage5,21. Except for the 65 K cryocooler, all of them are constructed on the basis of lubrication-free-piston compressors with slot-hole seals and ball piston suspension (Fig. 3); they have a lifetime of more than 1 year (up to 10,000 hrs). The hybrid system for deep cooling (N1) uses a solid-argon accumulator and fills a space vehicle bay (Fig. 4). It is designed for maintaining IR-instrumentation at nitrogen temperature levels. Using the cold accumulator to multiply the compressor's 1-year lifetime, the system can cool a cryogenic load with a heat release of 2-3 W for a mission life of 3-5 years. The N2 system uses the same cryocooler, but with the compressor at half the power level (with one drive)3. For system N3, with a temperature level of 35-40 K, a solid nitrogen accumulator is used based on the crystalline lattice structure phase transition that occurs in solid nitrogen at 35 K. The periodic cooling of the nitrogen accumulator is accomplished by a neon J-T system. For precooling of the neon down to 85-90 K, a mixed-gas throttle-cycle cryocooler is used with an argon-melting accumulator, similar to system N1. The N4 system with a neon-melting accumulator for operation at 25-30 K has three J-T cryocoolers. Cooling of the argon accumulator to 85-90 K is accomplished with a mixed-gas throttle cryocooler similar to the N1 system. For neon accumulator cooling, a two-cascade neonhydrogen throttle cryocooler with compressors as described above is periodically actuated. In the N5 system, a similar compressor is used in a closed-loop helium throttle-cycle 4-10 K
cooling system. The Precooling of the helium down to 19-20 K is fulfilled using a deuterium melting-solidification cold accumulator, which in turn is periodically cooled by a two-stage Stirling cryocooler. More detail on this cryosystem is presented elsewhere in this proceedings.4
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We have used the idea of a cold accumulator with cryogen evaporation-condensation in nonmagnetic cryocoolers N7 and N8 for superconducting magnetometers. Very low levels of magnetic noise have been achieved with such cryosystems by application of nonmagnetic materials and the use of a large spatial separation between the radiotransparent cryostat and the throttle-cycle compressor units. Vibration of the cooled object is also minimal due to the short-term cryocooler cycling. Nitrogen or argon vapor from fiberglass cryostats is removed into an external receiver or gas-holder. The throttle-cycle microcooler is actuated for the liquefaction of nitrogen (argon) only one or two times a week, for some hours. As a result, it successfully solves the problem of minimizing magnetic and electromagnetic disturbances. During evaporation, the temperature of the liquid and cooled object can be held to within 0.01 K, and even tighter if an intermediate pressure regulator is used together with a "soft" gas-holder for collecting of gas. As explained above, the use of accumulators has been found to have three key advantages: • Increased cryogenic system lifetime (up to 20 times) • Maintenance of a high level of temperature stability • Low levels of vibration and magnetic interference at the position of the cooled object In addition, it is possible to realize universal cryosystems (using the same hardware) for a wide variety of temperatures from 4 K up to 150 K (or above). This problem is considered in more detail below. UNIVERSAL CRYOCOOLERS WITH ACCUMULATORS FOR USE IN THE 4 – 150 K TEMPERATURE RANGE
The data of Table 1 show that by using the same type of compressor (or a modification) the SR&DB has managed to construct cryocoolers with accumulators for temperatures from 4 up to 90 K. This range is easy to expand for higher temperatures, as cooling capacity of all coolers increases with increasing temperature. Analysis and test results show that it is possible to extend the use of the existing cryocooler and accumulator hardware to any discrete temperatures in the range 4 – 150 K. In this case, it is only necessary to substitute higher temperature phase change materials for use in the accumulator. These can take advantage of any type of phase transitions including melting-solidification, evaporation-liquefaction, II-nd order phase transitions, magnetic phase transitions, superconducting transitions, etc. Qualified phase change heat sink materials are listed in Table 2. The discrete set of possible operating temperatures for the accumulator are determined by the temperatures of the documented phase transitions. As indicated, it is not necessary to change the working fluid of the J-T or Stirling cryocoolers for these higher temperatures. The increased capacity of the coolers at higher temperatures with allow the relative duration of their operation to be reduced; that automatically increases the lifetime of the cryogenic system as a whole. For universal J-T cryocoolers with an accumulator, the effective temperature range is 70150 K and above. For single-stage split-Stirling cryocoolers this range can be even wider: 35 – 150 K and more. The combination of two-stage split-Stirling coolers and J-T cryocoolers allows the range to be extended down to 4 K (see item 5 in Table 1). For universal J-T cryocoolers with gas mixtures there is one more additional modification possible using the same refrigeration machinery; this is to change the gas mixture composition of the J-T cryocooler in accordance with the increased operating temperature of the accumulator. Excluding most low-temperature components (at first then and at last ) from the mixture, it is possible to greatly increase the J-T cryocooler cooling capacity. For a constant cryogenic load level, this capability also increases the cryogenic system lifetime. CONCLUSIONS The SR&DB has adopted the use of cold accumulators in conjunction with a cryocooler as the key means of achieving multi-year life cryogenic cooling systems that can be selected to provide cryogenic cooling at a diverse number of cold-end temperatures. As usual, it is generally necessary to pay something for enhanced capability, and in this case the penalty is a certain increase in
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overall size and mass of the cryocooler system; however, the increase in weight and size is much less than the resulting increase in system lifetime. Key accomplishments of the SR&DB include: 1. A number of cryocooler systems with cold accumulators have been designed, developed and tested, providing discrete cooling in the temperature range 4 – 90 K; the accumulators have been based on three types of phase changes: melting-solidification, boiling-evaporation, and phase transitions of the second kind. 2. An increase in cryocooler system lifetime by a factor of 20 or more has been achieved through the incorporation of cold accumulators. 3. For cryocoolers with cold accumulators a high level of temperature stability has been achieved; J-T based systems have additionally reduced vibration and electromagnetic interference to minimal levels. 4. Analytical tools have been developed for the rapid and successful design and implementation of cryocoolers with accumulators, including analyses in the areas of: - optimum solutions for thermal insulation design - design of heat-transport fins and counterflow heat exchangers - refrigeration cycle parameter calculations that allow the selection and implementation of optimum designs of cryocoolers with accumulators 5. Accelerated life-test methods have been developed for cryocoolers with accumulators. These accomplishments and the experience gained with cryocoolers incorporating cold accumulators has fostered the implementation of a family of universal (for 4 – 150 K) cryocoolers with cold accumulators. This family, based on a common set of cryocooler and accumulator hardware, allows a wide variety of discrete cooling points to be provided through the simple replacement of the accumulator phase change material and/or the J-T cryocooler working fluid: • From 70 to 150 K using throttle cryocoolers with mixed gases • From 35 to 150 K using single-stage split-Stirling cryocoolers • From 4 to 150 K using hybrid systems with a split-Stirling upper stage and with an additional low-temperature J-T cooling stage.
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REFERENCES 1.
2. 3. 4. 5. 6.
7.
8. 9. 10. 11. 12.
13. 14. 15. 16. 17.
18.
19. 20. 21.
22.
Verkin B.J., Kirichenko Yu.A., Mikhalchenko R.S., Getmanets V.F., “Onboard cryostatting systems,” Cryogenic Engineering, Kiev, Naukova Dumka, 1985, pp. 25-47. Arkhipov V.T., Yakuba V.V., Lobko M.P., Yevdokimova O.J., “Multicomponent Gas Mixtures for J-T Cryocoolers,” Cryocoolers 10, Plenum Press, New York, 1999. Arkhipov V.T., et al., “Long Life Cryocooler for 84-90 K, ” Cryocoolers 10, Plenum Press, New York, 1999. Arkhipov V.T., Getmanets V.F., Levin A.Ja., “Long Life 5-10 K Space Cryocooler System with Cold Accumulator,” Cryocoolers 10, Plenum Press, New York, 1999. Arkhipov V.T., Getmanets V.F., Levin A.Ja., “Experimental Complex of the Non-Magnetic RadioTransparent Cryostatting System,” Cryogenics, 1992, v.32, ICEC Supplement, pp. 203-206. Alfeev V.N., Nikolsky V.A., Yagodin V.M., “Throttle Cryogenic Systems on Multicomponent Gas Mixture,” Electronic engineering, Ser. 15. A Cryogenic Electronic Engineering, Release 1 (3), 1971, pp. 95-103. Mikhalchenko R.S., Arkhipov V.T., Ostrovskiy I.N., “A study of solidification of liquid cryogens in cryostats,” Proc. of Third Soviet-West Germany Symposium on Heat Exchange, In: Low Temp. Physics, n.9, 1990, pp. 476-480. Getmanets V.F., Zhun G.G., Mikhalchenko R.S., et al., “Cryogenic Superinsulation with Increased Efficiency,” Adv. in Cryogenic Engin., v.43B, Plenum Press, New York, 1998, pp. 1319-1325. Mikhalchenko R.S., Getmanets V.F., Pershin N.P., Annikova T.A., “New Efficient Composit Superinsulations,” Cryogenics, v.32, ICMC Supplement, 1992, p. 60. Verkin B.J., Getmanets V.F., Mikhalchenko R.S., “Low temperature sublimative cooling thermophysics,” Kiev, Naukova Dumka, 1980, 256 p. Verkin B.J., Mikhalchenko R.S., Arkhipov V.T., et al., “Development experience for onboard sublimating cold accumulators,” In: Low Temperature Engineering, Kiev, Naukova Dumka, 1973. Mikhalchenko R.S., Vakulenko V.D., Arkhipov V.T. et al., “Two components sublimating cold accumulator,” KT-12, Same, pp. 22-30. Mikhalchenko R.S., Arkhipov V.T., Getmanets V.F., “Development principles for controlling-filling equipment of sublimating cold accumulators,” Same, pp. 31-47. Verkin B.J., Mikhalchenko R.S., Getmanets V.F., Goncharenko L., “Contact Heat Transfer in Solid Cryogens,” Adv. Cryogenic Eng., v.25, Plenum Press, New York, 1981, pp. 431-437. Verkin B.J., Getmanets V.F., Mikhalchenko R.S., “Thermophysics of the phenomena of gradientless heat transfer in porous solid cryogens,” Cryogenics, 1979, 19, n.1, pp. 17-23. Getmanets V.F., Grigorenko B.V., Zhun G.G., et al., “Accelerated Cryocooler Life Tests for Cryodeposit Failures,” Cryocoolers 10, Plenum Press, New York, 1999. Getmanets V.F., Mikhalchenko R.S., Shapovalenko V.V., Yurchenko P.N., “Universal magnetoradio-transparent cryostats for helium and nitrogen,” In: High temperature superconductivity, n.1, 1990, pp. 104-110. Getmanets V.F., Grigorenko B.V., Kurskaya T. A., et al., “Cryogenic-vacuum contamination sources identification for reinforced cryostats heat protection,” Same, pp. 86-92. Grigorenko B.V., Getmanets V.F., Mikhalchenko R.S., et al., “Cryovacuum systems contamination sources identification using IR-spectroscopy,” Chem. Oil Apparat. Develop., n.12, 1989, pp. 18-19. Getmanets V.F, Levin A.Ja., Potemina L.G., “Heat transfer through parallel axis-symmetric fins system during coolant phase change, ” Sov. Eng.-Phys. J., Minsk, v.59, n.6, pp. 903-910. Arkhipov V.T., Lubchenko V.N., Povstyany L.V., et al., “Low Weight and Long Life 65 K Cooler,” Cryocoolers 10, Plenum Press, New York, 1999. Getmanets V.F, Gorpinko Yu.I., Levin A.Ja., “Heat optimization for counterflow heat exchanger of cryogenic systems,” Cryogenics’90, Koshice, 1990, pp. 236-237.
Test Results of a Nitrogen Triple-Point Thermal Storage Unit B.G. Williams and I.E. Spradley Lockheed Martin Missiles & Space Company Advanced Technology Center Palo Alto, CA 94304 USA
ABSTRACT Mechanical cryogenic refrigerators, used to cool space based instruments, are often sized to provide constant cooling at a specified temperature and available input power. Intermittent sensor heat loads greater than the designed value will result in unsteady sensor temperatures. A thermal storage unit (TSU) is a device which absorbs the spike heat loads while providing a stable temperature reservoir for the sensor. The authors have constructed and tested a single volume TSU which absorbs heat loads by melting nonexpendable, totally-retained nitrogen at its triple point in a constant pressure-temperature-volume process. This system is designed to provide approximately 5 W-hr of thermal storage capacity at 63.1 K. Presented in this paper are the results of the measured thermal capacity and stability, investigating supercooling and superheating phenomena, and instrument simulations in 0° and 180° orientations. INTRODUCTION During the last two decades, Lockheed Martin has built and delivered over 20 highquality, long-life cryogenic cooling systems for a variety of sensors on scientific and military missions. These include passive and active cryogenic systems of various designs. To expand this capability and integrate the active cooling provided by cryocoolers with the temperature stability provided by passive solid cryostat systems, a closed-system thermal storage unit has been developed. This system absorbs heat loads by melting a nonexpendable, totally-retained substance at its triple point in a constant pressure-temperature-volume process (see Figure 1). The TSU system contains nitrogen and is designed to operate at it triple point of 63.1 K with a thermal storage capacity of 5 W-hr (18,000 J). This configuration is ideal for a variable duty cycle sensor because the cryocooler needs to only be sized to remove the average heat load from the system; the TSU acts as a thermal capacitor, absorbing the peak heat loads and providing a constant temperature heat sink throughout sensor operation. The energy absorbed by
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Figure 1. Pressure-temperature phase diagram showing the operating point of a triple point TSU as well as an expendable solid cryogen cooler.
the TSU during peak heat loading is removed by the cryocooler once the sensor heat load is decreased. A TSU system combines the advantages of both stored cryogen and active cryocooler systems: it provides the temperature stability of stored cryogens, along with the potential longlife of closed-cycle cryocoolers. The combined cryocooler/TSU system is also ideal for cooling sensors operating intermittently on low duty-cycle missions. The cryocooler is turned off while the TSU absorbs the heat loads during sensor operation. This provides an essentially vibrationless environment for the sensor. Then, during instrument quiescent periods, the cryocooler is turned on and the TSU substance is refrozen. Only enough phase change substance needs to be stored to provide cooling for one sensor operation cycle, thus reducing the required mass of the system. This paper discusses a nitrogen triple point thermal storage unit and presents the results from thermal capacity, thermal stability, and instrument simulation testing.
DESCRIPTION OF TSU The nitrogen triple point TSU consists of an aluminum 6061-T6 spherical tank, with an outside diameter of approximately 9 inches, and integral instrument and cryocooler interfaces (see Figure 2). Two tanks were constructed, one filled with aluminum foam and the other without. The one without foam was structurally tested by thermally cycling it between 300 K
and 77 K two times, pressurizing it to 3000 psi fifty times, and then pressurizing until failure at 6800 psi. The other unit with the foam, was charged to 1900 psi (room temperature) and sealed off; this provided a total nitrogen mass of Two strip heaters were mounted to the tank in order to provide as uniform as possible heat source to the phase change system during thermal capacity measurement tests. In addition, two
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Figure 2. Photograph of the aluminum TSU tank.
platinum resistance temperature (PRT) sensors were attached to the top and bottom hemispheres of the tank. A single thin-walled fiberglass tube was used to support the TSU tank inside of a thermal shield. This shield encompassed the TSU tank, contained two integral cryocooler interfaces, incorporated a liquid cryogen cooling loop, and provided a secondary instrument interface. The shield had the capability of being cooled by either a cryocooler or a liquid cryogen. Mounted on this shield was a strip heater in conjunction with a silicon diode temperature sensor, so that the temperature of the shield could be monitored and controlled to a desired temperature. The shield is supported from a specially made vacuum shell by a folded fiberglass tube arrangement. All cold sections (the thermal shield and the TSU tank) were blanketed in multilayer insulation (MLI) to minimize parasitic heat leaks from their warmer surroundings. For the characterization testing presented in this paper, a liquid helium (LHe) cooling loop was used in place of a cryocooler to cool the TSU tank. A heat meter was designed which
simulated a cryocooler with 1.5 W of cooling at 65 K, while at the same time providing a mechanism to measure the cooling being applied to the tank. This heat meter was calibrated prior to installation in the system so that the heat flow through it would be known by monitoring silicon diode temperature sensors located on each side of the device. As part of the scenario, in which it was desirable to simulate the operation of a cryocooler, the heat meter was designed to interface with the tank at the same location as a cryocooler. Integral to the top side of both the tank and the shield were shrink-fit interfaces in which an instrument to be cooled could be thermally connected to the TSU system. The instrument interfaced with the tank is termed the primary instrument, whereas the one interfaced with the shield is designated the secondary instrument. Simulated instrument heat loads could be applied to the system through test caps interfaced at these locations; the test caps contained heaters and PRT sensors.
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TEST RESULTS
Figure 3 shows a plot of the initial cool down of the tank/phase change material from room temperature through the nitrogen triple point. This graph is a composite of data taken over a three day period. Initially, the system was cooled with liquid nitrogen until the temperature was about 150 K. From that point on, liquid helium was used. A change in the slope of the temperature versus time curve, at approximately 122 K, is evident in Figure 3. This corresponds to the onset of liquefaction and is the value predicted based on the mass of nitrogen contained within the tank. The onset of liquefaction refers to the temperature at which an initially allgaseous substance, undergoing a constant specific volume cooling process, becomes saturated; after this point, with continued cooling, the phase change substance will consist of a saturated liquid/vapor mixture. At the triple point, the physics of the system dictate that a constant temperature process must occur and was observed as a flat temperature plateau. After the nitrogen was frozen, the system was cooled to a temperature slightly lower than the triple point in preparation for the first thermal capacity measurement test. Thermal Capacitance Measurement
Thermal capacitance measurements were taken for heat rates of 0.5 W, 1 W, and 2 W, while the shield temperature was maintained at 63 K to minimize parasitic heating. Figure 4 shows the triple point melt curve for the 0.5 W case; note that only the values for the sensors mounted to the tank surface are shown. For this case, the thermal capacitance was determined to be 4.69 W-hrs compared to the estimated value of 4.7 W-hrs. The estimated value was based on the final charge mass of nitrogen. The temperature rise from start to end of melt was about 0.5 K. The temperature stability of the melt plateau will be discussed later in this paper.
Figure 3. Composite plot of the initial cool down and freezing of the nitrogen TSU.
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Figure 4. Plot of TSU tank temperatures for the 0.5 W melt case.
Since the design of the heat meter allowed for a fairly accurate knowledge of the heat transfer rate through the link during freezing, the freezing curves could also be used to calculate the capacitance. However, there is still a margin or error associated with the determination of the cooling provided by the LHe. This is a result of fluctuations in the cold side of the link’s temperature (up to 10 K) during maximum cooling; variations of 10 K could result in about 1/3 W differences in the cooling rate. Figure 5 shows a sample of a freezing curve which was performed at the maximum cooling rate of about 2.8 W. Table 1 summarizes the thermal capacitances as measured for both melting and freezing tests.
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Figure 5. Example of triple point freezing process.
Temperature Stability
It was observed from the melting curves that the triple point plateau, for this TSU system, consisted of two distinct sections (refer to Figure 4). The first section is very flat and has a duration of between 15 to 35 percent of the total melt time. The second section starts with a very sharp step-increase in slope and quickly tapers off to a constant slope; this slope is greater than the flat section’s but yet much less than the slope of either before or after the melt. The temperature rise of the TSU during the second section accounts for the majority of the temperature change during melting. There are several items of interest concerning this phenomena that are worth noting. First, this slope change is only observed in the melting cases. It is postulated that the initial flat plateau is the result of the nitrogen phase change occurring in the nitrogen closest to the wall of the tank. After the step rise, the nitrogen phase change becomes uncoupled from the sensible heat of the tank. Since the tank wall is a much better conductor than the foam/nitrogen combination, the heat from the heaters is distributed faster through the tank than into the phase change material. For example, aluminum foam, with a relative density of approximately 3%, was bonded to the internal surface of the tank hemispheres prior to the welding of the tank. At 63 K, this foam only has a thermal conductivity of 1.6 W/m-K; this is about an order of magnitude greater than that of the nitrogen but still about an order of magnitude less than that of solid 6061 aluminum (the tank material). This implies that heat is transferred much more readily through the tank walls than into the foam/nitrogen matrix. Second, it has been postulated that the step-increase in the melting curve is inherent in solid-to-liquid phase transformations with a very low percent volume of liquid. In support of this theory and as an after-thought to the testing procedure, the TSU was cooled down through its solid-to-solid phase (alpha/beta) transition point. This transition occurs at approximately 35 K with nitrogen. The system was then allowed to warm up through the solid transition with only parasitic heat loads; the shield slowly warmed up to about 80 K since the LHe had been depleted during the cool down. A step-increase in slope midway through the phase change plateau was not observed.
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Table 2 shows the temperature rise over each part of the melt plateau for the various heat loads. Before the step-change in slope, the variation in temperature, especially in the lower heat load cases, is small. After the step-change, the temperature rises more rapidly but yet could still be considered relatively small. For example, even with the 2 W heat load, the temperature variation over the entire melt plateau is only about ± 0.6 K. With the 5 W heat loading and greater, the termination of the melting plateau was difficult to determine. Investigation of Supercooling Effects Supercooling is defined as the cooling of a liquid below its solidification temperature as a result of the formation of crystalline structures in a thermodynamically metastable state. If the latent heat released upon freezing is sufficient enough, then the temperature of the substance will rise rapidly back to its transition temperature for the completion of the phase change. Supercooling should not be confused with subcooling of the solid, which is defined as cooling the solid to lower temperatures after the completion of the freeze period. For cryocooler applications of a TSU system, supercooling could cause concern for the control system which operates the cryocooler. For example, if a control system is set to turn off the cryocooler when the temperature drops one degree beneath the triple point, and if supercooling is present and is greater than one degree, then the cryocooler would be turned off before the TSU had an opportunity to freeze. The result is that a flat plateau during instrument operation would not be available. From the freezing curves (refer to Figure 5 as an example), at maximum cooling capacity, it was determined that supercooling of the nitrogen did not occur. Supercooling effects would have been evident at the beginning of the freeze plateau. It is postulated that nitrogen, with its simple single element diatomic molecule, requires little if any effort to align into the proper crystalline form in order for solidification to occur. More complex molecules, such as methanol or nitrogen trifluoride, require significant effort to align their molecules into the proper crystalline structure, hence supercooling would be observed with these substances.1,2
Instrument Simulation
Instrument simulation testing was performed in two orientations: 0° and 180°. The 0° orientation refers to the instrument link interface, which is connected to the primary test cap, being pointed upwards and the liquid nitrogen pooled at the bottom of the tank at a point furthest from this interface. The 180° orientation refers to the entire system being rotated 180° upside down; the liquid nitrogen would then pool at a location closest to the instrument link interface.
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Figure 6. Instrument simulation at 0° orientation and a 50% duty cycle.
This rotation was performed with the nitrogen at a temperature greater than the triple point so that no nitrogen would be trapped in the foam in its solid phase. Each simulation cycle consisted of a total of 100 minutes. For example, a 50% duty cycle would have an instrument “on” period of 50 minutes and an instrument “off” period of also 50 minutes. During the “on” period, the heater mounted on the primary test cap was set to 3.5 W. For the “off” period, this heater was set for 0.5 W. The secondary test cap heater, connected to the shield, was maintained at a constant 1.8 W for the entire 100 minutes. The shield temperature for all instrument simulation testing was maintained at 140 K. For all instrument simulation tests, the cooling through the heat meter was set to approximately 1.4 W. It was a concern during the shorter duty cycle testing that the solid would subcool away from the triple point plateau. In order to fully simulate the operation of a cryocooler, a temperature controller was connected to the heater on the tank side of the heat
meter. If the temperature at this point dropped below about 61 K, then the heater would apply heat to keep that temperature from decreasing any further; this simulated the “ramp down” of a cryocooler as it approaches its set point. Figure 6 shows a plot of the two tank temperature sensors, the primary test cap sensor, and the tank side of the heat meter sensor, for a 50% duty cycle test. The temperature of the primary test cap appeared to be asymptotically approaching a steady state value, but the primary heater was set back to 0.5 W before this steady state temperature was obtained. The maximum variation in this temperature was about 1 K. There was also about a 4.5 K difference between the top of the tank and the primary test cap. This is a result of the thermal link connecting the tank to the test cap not being optimized for a particular design. The design incorporated was a generic, easily assembled fixture to be used for characterization testing only The complete primary interface link’s resistance was 1.43 K/W for an estimated steady state temperature drop of 5 K. Also from Figure 6, it can be noted that the nitrogen experienced complete melting partway through cycle #4’s “on” period. The recovery after this “on” period was not sufficient enough to restart the freezing process. Therefore, cycle #5 was performed with the nitrogen completely melted and produced a temperature versus time curve of constant slope. This is
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Figure 7. Instrument simulation at 180° orientation and a 50% duty cycle.
representative of a system in which a latent heat of transformation is not present but only sensible heat (specific heat capacity of aluminum and nitrogen). For all duty cycles, there was a noticeable temperature difference between the bottom of the tank and the top of the tank. This was expected since the bottom of the tank is closest to the cooling source (heat meter) and to the PCM. The top of the tank is closest to the heat source (primary test cap). Figure 7 shows the 50% duty cycle for the upside down orientation (180°). It was observed that the temperature of both the primary test cap and the top of the tank form a much flatter plateau as compared to the 0° orientation (see Figure 6). Also, in addition to the spike seen at the beginning of cycle 1, a small bump is evident at the beginning of each of the other cycles. The rise in slope seen at the end of cycle 2 and about 1/3 of the way through cycle 3 can be attributed to the step increase observed during the capacitance measurement testing. Complete melting occurred at the end of cycle #4; the “off” period of cycle 4 provided enough cooling to start to refreeze the nitrogen. The nitrogen was then completely melted towards the beginning of cycle #5. Finally, the temperature difference between the top of the tank and the bottom of the tank was much less for the 180° orientation than compared to the 0°. Also, a sharp “spike” in the temperature in both the primary test cap and top of the tank
temperature sensors was observed at the beginning of the first “on” cycle This spike was present at the transition from a subcooled solid to a triple point substance. However, this spike was only observed in the 180° orientation; there was nothing similar to it in all of the testing performed at the 0° orientation. It has been postulated that this spike is the result of localized superheating or stratification of the nitrogen near the tank wall. In addition, it could be the result of the relative placement of the heat sources and cold sinks. CONCLUSIONS
The characterization testing successfully measured the triple point thermal capacity of a single volume nitrogen TSU system. This capacity was found to be 4.69 W-hrs,
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which was very close to the calculated 4.7 W-hrs based upon the final nitrogen charge mass. A relatively flat temperature plateau was produced over the useful cooling capacity of the system with only about a 1 K temperature rise (2 W heat load case). A small step-increase in the slope of the melt plateau was observed during melting and occurred at about 25% of the way through the melt; this step-increase was only seen during
melting and never during freezing. It is postulated that the phenomena is inherent in solid-toliquid phase transformations which have a very low percent volume of liquid. Also, it was postulated that this step-increase could be partially the effect of the latent heat transformation becoming uncoupled from the sensible heat rise in the aluminum tank. Since the sensors and heaters were mounted on the outside of the tank, and also since the tank was a much better thermal conductor than the internal structure (foam/nitrogen), slightly increasing temperatures with a constant slope were seen after the initial plateau. However, for all melting cases, the temperature rise at the various locations was relatively small (about 1 K maximum) over useful operational time periods. During all freezing processes, supercooling effects were never observed. However, in the inverted 180° orientation, small spikes in the temperature of the primary test cap and top of the tank were observed and could possibly be attributed to either local superheating or stratification effects. The TSU system also demonstrated simulated instrument operation with cyclic varying heat loads. For a 50% duty cycle scenario, a temperature rise in the primary test cap of about 1.3 K was seen across the 50 minute plateau of the 0° orientation whereas about a 0.35 K temperature rise was seen across the plateau of the upside down case (180°). The temperature difference between the tank and the primary test cap, as observed during the instrument simulation tests, was a result of the generic design used for the link. This link has a resistance of 1.43 K/W. This system can readily be modified to adapt to a new tank based on design and analysis which can be performed to maximize the performance of a TSU. The design and analysis of this tank could be directed towards specific requirements by investigating and optimizing the internal heat exchanger (foam or other options), the geometry of the tank (spherical, cylindrical, flat can shaped, or other), and instrument interface link. Since the existing TSU system was originally designed to integrate cryocoolers to cooling the tank and shield, and this system had to be modified to incorporate the LHe cooling system, it could easily be adapted back to using actual cryocoolers. Also, the existing system is ideal for an integrated cryocooler/TSU testbed; slight modifications would only need to be performed to allow different cryocoolers or TSU tankage systems to be used. It should also be noted that the existing system is an ideal candidate for the testing of a TSU system using the alpha/beta solid-to-solid phase transformation of nitrogen. No further modifications would need to be performed so that testing could be accomplished. REFERENCES
1. Bugby, David C., et al. “Development of a 60 K Thermal Storage Unit”, In Cryocoolers 9, ed. R. G. Ross, Jr., Plenum Press, New York (1997), pp. 747-764. 2. Williams, Brian G., “Integrating a Triple Point Phase Change System with Cryogenic Mechanical Refrigerators”, PhD Dissertation, Utah State University, Logan (1997).
Optimal Integration of Binary Current Lead and Cryocooler H.M. Chang1 and S.W. Van Sciver2 1
Hong Ik University, Dept. of Mechanical Engineering
Seoul, 121-791 Korea 2
National High Magnetic Field Lab Tallahassee, FL 32306
ABSTRACT
An optimal integration of a binary current lead and a two-stage cryocooler has been analytically sought to minimize the required refrigerator power. The binary current lead is a series combination of a normal metal conductor at the warmer part and a high Tc superconductor (HTS) at the colder part. The lead is cooled by direct contacts with the first stage of the two-stage cryocooler at the joint and with the second stage at the cold end. No helium boil-off gas is present. A new and simple analytical method is developed to calculate the cooling loads at the two locations in the binary lead. A mathematical expression for the required power input for the loads is derived by incorporating a model depicting the performance of actual cryocoolers. With a new graphical method, the optimal conditions are found for the cooling temperature at the joint and the dimensions of the two parts to minimize the refrigerator power per unit current. The results show that there exists an optimal relation between the length and the current density of the metal lead, which is independent of the HTS part or the cryocooler. It is also demonstrated that the current density of the HTS and the joint temperature have unique optimal values respectively to minimize the refrigerator power per unit current, when the length of the HTS part and its critical properties are given. The actual power input to the cryocooler in the
optimal conditions is compared with its minimum as a thermodynamic limit, which can be obtained with reversible refrigerators. In addition, a useful dimensionless number is introduced for the optimal cooling of the binary current leads. INTRODUCTION
A binary current lead is a series combination of normal metal conductor as a higher temperature part and high Tc superconductor (HTS) as a lower temperature part. Since the HTS
materials are perfect electrical conductors and have much lower thermal conductivity than the normal metals, the heat leak to the cryogenic temperatures through the binary leads could be
considerably smaller than that through the conventional metallic leads. A number of studies1-6 have been performed during the past several years, to apply the bulk HTS materials to the leads carrying a high current density and to develop the effective cooling technology. The cooling method for the binary leads could be quite different from the standard helium-vapor-cooling of
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the conventional metallic lead, depending on how the liquid cryogens and/or the cryocoolers are employed1. Recent progress in the development of 4 K Gifford-McMahon refrigerators has raised the possibility of the liquid-free or the cryocooler-cooled superconducting magnets7,8. Since there is no liquid cryogen in those superconducting systems, the current leads should be conductioncooled in vacuum by contact with cryocoolers. Two typical configurations of the superconducting systems cooled by two-stage cryocoolers are schematically shown in Figure 1. The second stage of the two-stage GM cryocooler absorbs heat from the magnet and the cold end of the HTS current lead, while the first stage cools the radiation shield and the joint between the two parts of the binary lead. The cooling at the joint is necessary to reduce the amount of heat leak to the lower temperatures and to maintain the HTS in a superconducting state. The present authors think that the feasibility of the conduction-cooled HTS current lead has been demonstrated by the recent construction and operation of several prototypes7,8 and the next crucial step towards the practical application could be the development of energy-efficient current leads. The cooling of the HTS leads without the boil-off helium gas has been partly considered in some of the previous publications1-6. Most of this research work is, however, related with the design or the analysis for the HTS leads whose ends were cooled by liquid nitrogen or liquid helium and still might not provide enough information on the optimal cooling scheme for the liquid-free HTS leads. For conventional metallic leads, the conduction-cooling method was examined and completely optimized by Hilal9. In the theoretical work, Hilal showed by the method of calculus of variations that the refrigerator power could reach an absolute minimum with optimally distributed Carnot refrigerators and optimally sized leads. A few years before Hilal’s work, Bejan and Smith10 derived an absolute minimum of the refrigerator power required to cool a given geometry of mechanical supports for cryogenic apparatus. From a thermodynamic point of view, the mechanical supports are quite similar to the HTS current leads that do not generate heat in a superconducting state. Recently, the present authors published a new optimization technique for the conductioncooling of the binary current leads11, by combining the two optimization methods mentioned above. They have revealed that the refrigerator power has an absolute minimum as a thermodynamic limit, when the lead is cooled by optimally distributed Carnot refrigerator along the length of the lead and the dimensions of the lead are optimized. They have also presented the optimal conditions for the binary current lead cooled by a two-stage Carnot refrigerator. These results are considered as thermodynamic limits, because the refrigerator has been assumed to be reversible and the HTS is marginally superconducting at the optimal conditions.
Figure 1. Typical configurations of cryocooled-cooled superconducting magnet systems.
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Figure 2. Binary current lead cooled at cold end and at joint by a two-stage cryocooler.
This paper aims at the optimal operating conditions for the binary current lead cooled by an actual two-stage cryocooler to minimize the required refrigerator power. In order to achieve a practical usefulness, the optimization should include the performance characteristics of the contemporary cryocoolers and the realistic stability margins. A new analytical and graphical method is presented first with specific examples for a quantitative discussion. Then, a more general optimization technique is discussed for its application to every kind of binary lead cooled by a two-stage refrigerator. COOLING LOADS
A binary current lead cooled by a two-stage cryocooler is schematically shown in Figure 2. The HTS part and the normal metal part of the binary lead are denoted by subscripts 1 and 2, respectively. The heat current through the lead in a direction from the warm to the cold end is defined as Q. The heat removed at the cold end by the second stage of the cryocooler is denoted by
and the heat removed at the joint by the first stage is denoted by
between the heat from the metal at the joint,
is the difference
and the heat to the HTS at the joint,
It is assumed that the HTS does not generate heat in a superconducting state. and are the temperatures at the cold end, the joint and the warm end, respectively. For the HTS part of the lead, the heat current is constant along the axis and is identical to the cooling load at the cold end, since no heat is generated.
where A and L are the length and the cross-sectional area of the lead, respectively and k is the thermal conductivity.
For an infinitesimal length of the metal lead shown in Figure 2, the energy balance equation can be written as and the heat generation rate, is expressed by combining the one-dimensional equations for the Fourier heat conduction and the Ohmic heat generation.
where is the electrical resistivity and I is the current that the lead is carrying. After Equation (2) is multiplied by Q and integrated over the metal length, it can be rearranged for the heat current from the metal lead to the joint,
It is immediately observed that
has its minimum when the heat current at the warm end,
is zero. If has a positive value, is larger than the minimum because of the excessive heat conduction through the metal lead. If has a negative value on the contrary, is also
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larger because of the excessive heat generation. The axial temperature gradient should be zero at the warm end when the heat conduction and the heat generation are optimally balanced. This condition for the minimum is identical to the case of the conventional vapor-cooled metal lead12. The minimum heat current to the joint is now found in a closed form by letting
Thus the minimum cooling load at joint is obtained from Equations (1) and (4).
At any arbitrary axial location of the metal lead, the optimal heat current is found as a function of temperature by integrating Equation (2) from the axial location to the warm end and letting
from which the optimal dimensions of the metal lead are found.
It should be noted that the optimal conditions for the metal lead are related to the joint temperature of the two parts but are independent of the dimensions the HTS lead.
Equations (4) and (7) are very simple and useful expressions for the conduction-cooled metal lead, which have been reported recently by the present authors11. Similar optimal conditions were reported in previous publications with simple assumptions for the material properties, which could be derived as special cases of these general expressions. If the electrical resistivity and the thermal conductivity are assumed to be constant, Equation (4) is reduced to which was described by Seol and Hull5. For materials that obey the Wiedemann-Franz law, Equation (4) is directly integrated to
as Yang and Pfotenhauer6 mentioned. REFRIGERATOR POWER
The total power input to a cryocooler for the two cooling loads can be generally expressed as,
where FOM is the figure of merit, defined as the ratio of the actual coefficient of performance (COP) to Carnot’s coefficient of performance. In Equation (10), it has been assumed that the warm end temperature, is identical to room temperature at which the cryocooler rejects heat.
The FOM of a cryocooler depends on the cold head temperature, the type of refrigeration cycle, the refrigeration capacity, the performance of its components, and so on.
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Figure 3.
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COP (coefficient of performance) and FOM (figure of merit) of some commercial cryocoolers and the refrigeration models in current study.
If the FOM’s are known for a specific cryocooler, the refrigerator power can be calculated from Equation (10). On the other hand, for the purpose of a more quantitative demonstration of the optimization, it has been suggested in this study to construct a simple yet reasonable model to the FOM’s depicting the performance of actual cryocoolers. An extensive survey has been executed on the performance of commercial cryocoolers that can be applied to the cooling of the binary current leads, and the COP and the FOM are plotted as functions of the refrigeration temperature in Figure 3. It is noticed that the COP and the FOM of the second stage of two-stage GM cryocoolers or the single-stage coolers have greater values than those of the first stage of two-stage GM cryocoolers. In most practical cases, the cold end of the binary lead is cooled by the second stage of a two-stage cooler and the joint of the two parts are cooled by the first stage, so the two FOM’s are expressed by simple functions,
as indicated by two dashed curves in Figure 3. In this model, the has a value between 0.02 and 0.05, and has a value between 0.01 and 0.035 in the valid temperature ranges. OPTIMIZATION The integration of the binary current lead and the two-stage cryocooler can be optimized such that the total refrigerator power per unit current1-3,6,9 has a minimum. The mathematical expression for the refrigerator power per unit current is derived by substituting Equations (1) and (4) into Equation (10) and dividing it by I.
where is the current density at the HTS. It is worthwhile to notice in Equation (12) that the power per unit current is a function of and only, when the FOM’s, the material properties and the end temperatures are given. The readers may be reminded that and are not included in Equation (12) because they have been already optimized as in Equation (7). The refrigerator power per unit current has been calculated with Equation (12) for various values of the current density and the joint temperature in a copper+Bi2223 binary lead. Figure 4 shows contours of the refrigerator power per unit current on a current density vs. joint
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Figure 4.
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Contours of refrigerator power per unit current and critical current density of HTS on current density and joint temperature diagram of Cu+Bi2223 when
temperature diagram for and FOM’s given by Equation (11). The thermal conductivity of Bi2223 is taken from Hermann et al.3 and the properties of copper are taken from Maehata et al.13 for RRR=60. The cold and warm end temperatures of the lead are fixed at 4 K and 300 K, respectively, throughout this paper. Generally speaking, the refrigerator power per unit current decreases as the current density of the HTS or the joint temperature increases. However, when the joint temperature is low or the current density is high, the power per unit current is almost independent of the current density since the cooling load for the HTS or the first term in Equation (12) is relatively small. On the contrary, when the joint temperature is high or the current density is low, the first term is dominant and the current density is relatively more significant in the total power per unit current than the joint temperature. Since Equation (12) has been derived with the assumption that the HTS does not generate heat, the superconductivity should be confirmed by incorporating the critical properties of the HTS, which will establish the final process of the optimization. The current density of the HTS should not exceed the critical current density, which can be represented reasonably well by a linear function of temperature for Bi2223 1,4.
where is the critical current density at 0 K and varies over a wide range, depending upon the size, the shape and the fabrication method as well as the applied magnetic field. For the purpose of more quantitative discussions in this paper, it is assumed that JC0 = 10,000 A/cm2 and 104 K for Bi2223 at zero magnetic field4 and Equation (13) is plotted on the
diagram of
Figure 4. Clearly, there exist unique optimal values for the current density and the joint temperature to minimize the refrigerator power per unit current while the HTS is superconducting,
as indicated by the square dot. At higher joint temperatures and smaller current densities than the optima, more refrigerator power per unit current is required because of a greater refrigerator power to cool the HTS. At lower joint temperatures and larger current densities, more power per unit current is also required because of a greater power to cool the metal part of the lead. The concave shape of the curves at low region indicates that for a constant current density, the optimum joint temperature to minimize the power per unit current should be significantly lower than the critical temperature, as discussed by Yang and Pfotenhauer6. The theoretical minimum of the power per unit current is about 2.66 W/A with the present cryocooler model, which is about 29 times greater than the minimum for a two-stage Carnot refrigerator11 and about 47 times the absolute minimum as a thermodynamic limit 11 that can be
INTEGRATION OF BINARY CURRENT LEAD AND CRYOCOOLER
obtained with distributed Carnot refrigerator. The corresponding optimal values of
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and
are
and 97.7 K, respectively. Because of the nature of the superconductivity, the theoretical optimum to minimize the refrigerator power per unit current is always determined at a marginally superconducting state as shown in Figure 4. In practice, the HTS current leads should be designed such that the current density and the joint temperature are lower than their theoretical optima in order to be stable from a certain level of thermal disturbance. Two different design schemes are suggested for the determination of the actual operating conditions with some stability margins. The first one is to find the operating condition among the points where the operating current density of the HTS has a certain fraction of its critical value at any arbitrary temperature. The dotted curve on Figure 4(a) shows the points where is 50 % of and the refrigerator power per unit current on the curve has a unique minimum, 2.93 W/A as indicated by a circle dot. This design is based on a safety factor of 2 in the operating current density. The corresponding optima for and are 415 A/cm2 and 95.4 K, respectively. The second design scheme is to find the operating condition among the points where the joint temperature is lower by a certain amount than its critical temperature at any arbitrary current density. The temperature difference is related closely with the magnitude of the disturbance energy to initiate a quench in the HTS lead or to propagate the normal zone. The dotted curve on Figure 4(b) shows the points where TJ is lower by 5 K than the critical curve. The refrigerator power per unit current on the curve has a unique minimum, 2.94 W/A as indicated by a triangle dot, when and are and 93.3 K, respectively. In this specific example, the optimally designed refrigerator power has about 10% more than the theoretical minimum in the both cases. The above procedures have been repeated for various values of the HTS length, and the results have been plotted in Figures 5 through 7. In these figures, the theoretical optima are marked by the squares and the two suggested designs are marked by the circles and the triangles, respectively, as in Figure 4. It is observed in Figure 5 that as increases, the refrigerator power per unit current decreases for every design. However, the power per unit current does not vary
significantly if is greater than about 10 cm, which means that the length of the HTS does not need to be very long as far as it is optimally cooled. As decreases to zero, the power per unit current approaches the values required by single-stage cooling of an optimized metallic lead. If is infinitely large, all contours of in Figure 4 are vertical and the theoretical power per unit current will be reduced to 1.99 W/A, which can be directly calculated by the asymptotic behavior of Equation (12).
In this simple case, the cooling load at the cold end is negligible and the joint temperature is the critical temperature of the HTS.
Figure 5. Optimized refrigerator power per unit current as a function of HTS Length for
Figure 6. Optimal joint temperature as a function of HTS Length for
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Figure 7. Optimal current density of HTS as a function of HTS Length for
Figure 8. Optimal current density and length of metal (copper) as a function of joint temperature for various RRR’s.
In Figure 6, the optimal increases and approaches to the critical temperature of the HTS, as increases. It is noted that if is 10 cm or 30 cm as is typical of commercial HTS lead, the practical design values of are found in a range between 90 K and 95 K, while the theoretical optima are slightly higher. It is observed in Figure 7 that as
increases, the optimal values of
decreases monotonically and the optima including the stability margins have smaller values than the theoretical optima. Once the joint temperature is optimally determined, the design for the metal lead should be finally designed by Equation (7). Figure 8 shows the optimal product of the current density and the length of the metal lead as a function of the joint temperature for various RRR values of
copper. The properties of copper have been taken from Maeheta et al.13, again. It should be kept in mind that if the material and the end temperatures of the metal lead are given, there are an infinite number of combinations for and to minimize the cooling load, but the minimum load is the same for every optimized combination, as given by Equation (4). The theoretical optimization and the suggested designs for leads are summarized and compared with the corresponding thermodynamic limits11 in Table 1. The optimized results for are the cases of the metallic lead, in which the optimal single-stage cooling at the cold end can be directly derived by letting in Equations (4) and (7). When the cooling load at the cold end is negligible and the optimal joint temperature is the critical temperature of the HTS as described by Equation (14). It is noted that the optimal values of for the metal lead are not strongly dependent on the cooling method or the design scheme, because the optimal joint temperature does not vary significantly in this specific example. A NEW DIMENSIONLESS NUMBER
In the previous section, the optimization for the cryocooler-cooled binary lead has been demonstrated with the variables having real dimension. It is worth performing a simple dimensional analysis, because the number of independent variables may be reduced in the optimization process.
The minimum of the refrigerator power per unit current, Equation (12), with a constraint, Equation (13), can be expressed in general as a function of the two end temperatures, the critical properties and the length of the HTS lead, the thermal conductivity and the electrical resistivity for two materials as functions of temperature, and the performance of the cryocoolers for any optimization scheme.
Equation (15) can be written in a dimensionless form as
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where a significant dimensionless variable is defined as
and will be called the CV number after the present authors.
The CV number is composed of the properties of the HTS and the metal and the length of the HTS, but has nothing to do with the cooling method. The number is physically interpreted as the
relative magnitude of the cooling load of the optimized metal lead with respect to the HTS lead. It is obvious that the cooling load of the HTS lead can be reduced by increasing the current density or the length as they are in the numerator of Equation (17), or by decreasing the thermal conductivity as it is in the denominator. It is also noted that the square root term in the CV number is a constant for metals that obey the Wiedemann-Franz law. The value of the CV number is approximately 1,550 for the lead when and
An evident usefulness of the CV number can be illustrated with Figures 5 through 7. The graphs for the optimal conditions have been generated as functions of when As mentioned earlier, depends upon the size, the shape, the fabrication method and the applied magnetic field. Even if of the available HTS lead may not be Figures 5 through 7 could be used to determine the optimal values for
and
with an equivalent
that yields the same CV number.
SUMMARY
A complete design method is developed for the optimal cooling of the binary current lead
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with a two-stage cryocooler. The optimization aims at the minimum of the required power input per unit current, provided that the critical properties and the length of the HTS lead, the refrigeration performance of the cryocooler and the two end temperatures of the binary lead are given. The analytical or graphical design procedure can be summarized as the following steps. (1) The required refrigerator power input is calculated by Equation (12) with an assumption that the dimensions of the metal lead are optimal. (2) The optimal values for the joint temperature and the current density of the HTS should be determined “simultaneously” by minimizing Equation (12) with a constraint, Equation (13) as in Figure 4. A proper stability margin may be included in this step. (3) With the optimal joint temperature, the product of the length and the current density of the metal lead is optimized by Equation (7) as in Figure 8. Any combination of the length and the current density will result in the same cooling load for the metal.
ACKNOWLEDGMENT This paper has been accomplished under the Financial Support Program for Faculty Research Abroad provided by Korea Research Foundation. The National High Magnetic Field Laboratory is jointly supported by the State of Florida and the National Science Foundation under grant DHR-9527035.
REFERENCES 1. Wesche, R. and Fuchs, A.M., “Design of superconducting current leads,” Cryogenics, vol.34, no.2 (1994), pp.l45-154.
2.
Herrmann, P.F. et al., “Cryogenic load calculation of high Tc current lead,” Cryogenics, vol.33, no.5 (1993), pp.555-562.
3.
Herrmann, P.F. et al., “European Project for the Development of High Tc Current Leads,” IEEE Trans Applied Superconductivity, vol.3, no.1 (1993), pp.876-880.
4.
Balachandran, U. et al., “Application of Sinter-Forged Bi-2223 Bars to 1500-A A.C. Power Utility
Services as High-Frequency Current Leads in a 77-4 K Temperature Gradient,” Applied Superconductivity, vol.3, no.6 (1995), pp.313-320
5. Seol, S.Y. and Hull, J.R., “Transient analysis and burnout of high temperature superconducting current leads,” Cryogenics, vol.33, no.10 (1993), pp.966-975. 6.
Yang, S. and Pfotenhauer, J.M., “Optimization of the Intercept Temperature for High Temperature Superconducting Current Lead,” Advances in Cryogenic Engineering, vol.41, Plenum Press, New York (1996), pp.567-572.
7. Hasebe, T. et al., “Cryocooler Cooled Superconducting Magnets and Their Applications,” presented at Cryogenic Engineering Conference and International Cryogenic Materials Conference, Portland, Oregon, (1997). 8. Watanabe, K. et al., “11 T liquid helium-free superconducting magnet,” Cryogenics, vol.36, no. 12 (1996), pp.l019-1025. 9. Hilal, M.A., “Optimization of Current Leads for Superconducting Systems,” IEEE Trans Magnetics, vol.MAG-13, no.1 (1977), pp.690-693. 10. Bejan, A and Smith, J.L., “Thermodynamic optimization of mechanical supports for cryogenic apparatus,” Cryogenics, vol.14, no.3 (1974), pp.158-163. 11. Chang, H.-M. and Van Sciver, S.W., “Thermodynamic optimization of conduction-cooled HTS current leads,” to be published in Cryogenics (1998). 12. Wilson, M.N. Superconducting Magnets, Oxford University Press, Oxford (1993), p.261.
13. Maehata, K., Ishibashi, K. and Wakuta, Y., “Design chart of gas-cooled current leads made of copper of different RRR values,” Cryogenics, vol.34, no.11 (1994), pp.935-940.
Cryogenic Systems Integration Model (CSIM) S. D. Miller and M. Donabedian
The Aerospace Corporation El Segundo, CA, USA 90245
D. S. Glaister
The Aerospace Corporation Albuquerque, NM, USA 87119
ABSTRACT
The Cryogenic Systems Integration Model (CSIM) is a Microsoft Windows® 95 based software tool for the simulation and analysis of spacecraft cryogenic mechanical refrigeration thermal control systems. Previous experience has shown that cryogenic systems exhibit large
analytical uncertainties. Historically, cryogenic thermal integration has been critical and often
inadequately considered during the preliminary design phase resulting in significant revisions and system penalties. CSIM development was initiated in response to the need for an efficient method for the preliminary design and parametric analysis of spacecraft cryogenic mechanical refrigeration systems. CSIM provides the capability to model and simulate the thermal performance of a cryogenic mechanical refrigeration system consisting of instruments, cryocoolers, thermal straps, thermal
storage units, thermal switches, heat pipes, intermediate shields, radiators, and brackets. The simulation outputs provide a complete breakdown of temperatures, heat flows, dimensions, weight, power, and total system penalties. Version 2.1 was presented at the 9th International Cryocooler Conference (ICC9). An updated and improved version (3.0-030) is presented in this paper. Enhancements include: a) the addition of a context sensitive help system, b) a Carnot correction and stroke interpolation routine which allows the user to interpolate or extrapolate the cooler performance to a heat rejection temperature and stroke (or power level) different from that provided in the database, c) revisions and improvements to the network display and dialog boxes to make the program easier to use, and d) expansion of the cooler database which includes incorporation of several coolers characterized at the Air Force Research Laboratory (AFRL), Albuquerque, New Mexico, during the last few years. The program has been licensed to a number of aerospace industry contractors and government agencies and is available to qualified users.
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INTRODUCTION
Integration, analysis, and evaluation of sensor cooling systems at cryogenic temperatures using mechanical refrigerators (cryocoolers) has become increasingly complex, difficult, and
time consuming. The large number of variables and options available to the designer, together with an increasing number of cryocoolers available or that are being developed, begs for a more efficient method of integrating and analyzing these systems. Past experience has shown that systems integration is not adequately considered during the early design stages leading to significant system performance degradation or drastic design changes. In response to these needs, a Microsoft Windows® 95 based software tool was developed at The Aerospace Corporation. It is intended that this program will be used as a preliminary design tool to aid in parametric and trade-off analysis and provide a means to simulate the performance of various system options. The beta version of CSIM was presented at 8th International Cryocooler Conference while the first operational model was presented at An updated and improved version 3.0-030 is presented in this paper. Enhancements include: a) the addition of a context sensitive help system, b) a Carnot correction and stroke interpolation routine which allows the user to interpolate or extrapolate the cooler performance to a heat rejection temperature and stroke (or power level) different from that provided in the database, c) revisions and improvements to the network display and dialog boxes to make the program easier to use, and d) expansion of the
cooler database which includes incorporation of several coolers characterized at the AFRL
during the last few years. Description of Improvements and Features Added
A number of improvements and many new features have been added to this current version (3.0-030) of CSIM since the beta version was presented at the
in 1994 and version 2.1 was presented at the in 1996. The enhancements were made to reduce the effort required to add new cryocoolers to the database, and to reduce the time required for new users to learn to use CSIM. These enhancements are described in detail in the remainder of this section.
Context Sensitive Help. In prior versions of CSIM, an extensive user’s manual was written which described the CSIM’s underlying model and assumptions, functions and procedures for adding and changing the database, and running the simulation. This version of CSIM has a context sensitive help feature. This feature provides help information available for the function currently being displayed. The new help feature is based upon the Windows® Help System. When a help button is pressed, the help information is displayed for the corresponding dialog box. All of the features of Windows® Help System are available for the CSIM Help text. The information can be reviewed in a standalone mode, and the information can be searched by specific words.
Error Messages. In prior versions of CSIM, a Help menu was included as part of the user’s guide to describe the user’s actions required for warning and error messages. Since each element has a large number of input characteristics, it is easy for the user to create an integrated system of elements which are not necessarily compatible with each other. As a result, there are many
opportunities to create situations where the program will not converge on a solution. Approximately 40 warnings or error messages can be generated. With the revised version of CSIM, when a specific message is displayed as result of a problem, the corresponding reason and
all of the options available to fix the problem are displayed when the help button is selected. An example of the help information is seen in Fig. 1. A complete list is provided in the user’s guide which is included with the software.
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Figure 1. Context sensitive help messages
Cryocooler Database Changes. The cryocooler database has been expanded significantly since the original database was presented at the A number of coolers characterized at the AFRL over the last few years have been added including the TRW model 3503, 3585, and 6020 pulse tubes; the Creare 65 K standard spacecraft cooler (SSC) diaphragm Stirling cooler and the Texas Instruments (Raytheon) 1.0 W Stirling tactical cooler3 configured for the Space Technology Research Vehicle-2 (STRV-2) space flight. Also, included is the Creare 65 K single-stage reverse Brayton (SSRB) engineering development model which has been under life test at the AFRL and has accumulated in excess of 20,000 hours.
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Other new data includes the Matra-Marconi Space (United Kingdom) 50-80 K Stirling unit using an intercept strap at 150 K and 190 K to enhance cooling capacity. Previous data was for the standard unit without an intercept strap. New data from Hughes Aircraft (Raytheon) has been included for the improved standard spacecraft cooler #4 unit with intercept strap cooling at 140 K and 170 K. This unit with the 170 K intercept strap is currently one of the baseline coolers for the United States Air Force Space Based Infrared Systems-Low (USAF SBIRS-L) focal plane cooling system. Because of the increasing use of several tactical coolers for use in short-term space flight experiments, data was also added from the study of Sparr and Banks4 which recorded performance degradation for up to 3000 hours for three different tactical coolers which were then defined as: 1) Hughes 7050H, 2) Magnavox MX 7049 and, 3) the Texas Instruments 1.0 W cooler. These coolers are now of course all part of the Raytheon systems product line. Cryocooler Database Interpolation and Extrapolation Capabilities. For the beta version of CSIM presented at and version 2.1 presented at there were three different sections of the cryocooler database; one for actual test data, a second where a manual Carnot correction had been made for different heat rejection (compressor case) temperatures, and a third for miscellaneous reference data. In the current version, only the first set has been retained. The second set was deleted because there is now an automatic Carnot correction available (defined below), and the third set has been deleted because it was no longer considered useful. The beta version of CSIM included a compressor case temperature interpolation if there were data for multiple case temperatures but it did not allow stroke interpolation. The current version now allows the user to select any stroke and the program will conduct a linear interpolation of the performance. Also, if there are test data for only one compressor case temperature, the new Carnot correction routine will allow the user to run cases based on estimated performance for other temperatures using a theoretical Carnot correction. Carnot Correction. As indicated above, the cryocooler element now includes an automatic Carnot correction capability which is useful in estimating cooler performance at heat rejection temperatures (compressor case temperatures) that are not available as part of the existing cooler database. This feature activates automatically when there is performance data for only one compressor case temperature and it does not match the selected value for the simulation. In addition, the user has an option to further modify this correction by inserting additional empirical correction factor. This was put in the program because evaluation of typical Stirling and pulse tube coolers showed that they do not follow the theoretical Carnot correction very well at higher temperature due to rapid decrease in compressor motor efficiency at elevated temperatures. The Carnot correction routine assumes the cooler load line is the same as the original data and merely changes the input power values in the original matrix. Thus, the user must be aware that this extrapolation may not be valid if the cooler data is already based on a near maximum stroke or power level and thus this correction is intended primarily for cases where a nominal stroke or power level has been utilized in the database. The new power input value calculated is based on the following equation which essentially derives a new power input data points using the existing cold tip temperatures and power input data points.
where CF
= = = = =
The original total power input The new computed total power input The correction factor manually supplied for the new case temperature Compressor case temperature Cold tip temperature
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Figure 2. Revised network display.
This is derived from the theoretical Carnot efficiency which can be expressed by
Note. Since the new power being computed in Eq. (1) is being computed for a common cold tip temperature, the in the numerator of the theoretical Carnot efficiency Eq. (2) cancels out in the equation used in calculation of the new power. Also, if the new calculated power exceeds
the maximum power shown in the database for that cooler, CSIM will halt the simulation and display an error message “that the cooler maximum power has been exceeded.” The user will be instructed to either select a lower power setting, a lower compressor case temperature, or modify a number of other parameters to reduce the load and required power input for the cooler. Details are covered in the user’s guide and dialog boxes are provided in the program.
Updated User’s Guide. An updated and expanded user’s guide has been prepared and is provided with the program to qualified organizations who request a copy of the program and complete the licensing agreement.
Network Display. More descriptive figures used in the network display were added to
CSIM as an enhancement. New figures were created for the following components: cryocoolers, thermal straps, thermal switches, conduction bars, heat pipes, radiators, focal plane arrays, thermal storage units, isolators, brackets, and vacuum enclosures. These figures are displayed for their respective components. An example of the revised display is seen in Fig. 2.
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SUMMARY AND CONCLUSION
CSIM is a Microsoft Windows® 95 based software tool for analysis and simulation of integrated cryogenic cooling systems using mechanical refrigerators (cryocoolers); it has been developed with extensive modifications from the earlier versions presented at ICC8 and ICC9. The program includes logic, algorithms and database necessary to carry out a simulation on an entire system and or conduct parametric analyses. The enhancements have been made to the
program to make it a more powerful tool for analysis of integrated cryogenic cooling systems using cryocoolers. The improvements to the user’s manual and the addition of the context
sensitive help feature make CSIM easier to use for new users. The simplifications to the cryocooler database, and the Carnot correction reduce the effort required to add cryocooler
performance data as the new data becomes available. The program has been licensed to several organizations and the request forms and licensing agreements are available by contacting the authors. ACKNOWLEDGMENT The support for this activity was supplied jointly from The Aerospace Corporation’s engineering methods funds and the United States Air Force Research Laboratory in Albuquerque, New Mexico. REFERENCES
1. Donabedian, M., S. D. Miller and D. S Glaister, “Cryogenic Systems Integration Model (CSIM), version 2.1,” Cryocoolers 9, Plenum Press, New York (1997), pp. 861-872. 2. Donabedian, M., D. S. Glaister and D. Bernstein, “Cryogenic Systems Integration Model, (CSIM),
beta version,” Cryocoolers 8, Plenum Press, New York (1995), pp. 695-707. 3. Private communication, 1st Lt. B. J. Tomlinson, AFRL Kirtland Air Force Base, Albuquerque, New Mexico, Data for the STRV-1D/2 cryocooler (25 November 1997). 4. Sparr, L. and S. Banks, “Adaptation of Tactical Coolers for Short Duration Space Flight Missions,” Cryocoolers 8, Plenum Press, New York, (1995), pp. 695-707.
Heat Rejection Effects on Cryocooler Performance Prediction Lt. B. J. Tomlinson Air Force Research Laboratory Kirtland AFB, NM USA 87117 A. Gilbert and J. Bruning Nichols Research Corporation Albuquerque, NM USA 87106
ABSTRACT Satellite developers have the complex task of designing spacecraft thermal management systems that must remain stable despite extremely dynamic operating environments. Developing spacecraft thermal management systems that depend on predictable cryocooler performance is
one of the most challenging aspects of space thermal system design. Developers must often rely on very sparse cryocooler performance data when conducting their system-level tradeoffs. While significant data is becoming available on newly emerging coolers, most experimentation is performed at or near a cooler’s design point operation based on somewhat generic or dynamic spacecraft requirements. This makes it extremely difficult to examine all possible spacecraft operating scenarios and their effects on cryocooler performance. Off-design point performance mapping is also often limited by time constraints, either imposed by system or follow-on cooler development schedules. However, in characterizing cryocoolers at the Air Force Research Laboratory (AFRL), considerable off-nominal performance mapping has allowed empirical modeling of quantifiable parametric relationships. These models tie together the interrelationships between key cryocooler parameters, both operational and environmental, forming precise performance prediction methods useful over most potential operating ranges. This paper presents an approach to developing performance prediction models for the coolers undergoing characterization and evaluation at AFRL’s Cryocooler Characterization Laboratory (CCL). Cooler concepts under investigation include various Stirling cycle, pulse tube, and reverse Brayton cycle machines. Particular emphasis is placed on better quantifying the relationship of heat rejection temperature to both independent and dependent cryocooler control parameters. These models may become extremely useful design tools that could assist system developers in selecting the appropriate cooler for their operational and mission needs. It will also aid spacecraft developers in adequately compensating for cooler parametric sensitivities and time dependent performance drift in their satellite thermal management designs.
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INTRODUCTION
Several cryocooler concepts capable of handling different cooling loads spanning the temperature range between 8.5-150 Kelvin were developed to meet known or projected military space cryogenic cooling requirements. Stringent cold end refrigeration performance must be met despite a wide range of heat rejection temperatures associated with the thermal management capabilities of varying spacecraft system designs and orbital environments. A composite study of multiple military mission analyses predict spacecraft thermal bus temperatures that span the range of 225 K to over 343 K under certain operating scenarios. These rejection temperatures form the high and low extreme thermal boundary conditions for the thermodynamic operation of the coolers. Performance of the various space cryocooler concepts now undergoing evaluation show different levels of sensitivity to heat rejection temperature and whether it is applied in a steady-state or transient fashion.
Cryocooler operation can be thought of as a continuous process yielding finite samples of measured performance data. There are limits to the effectiveness of the information gained from finite steady state samples of data viewed as discrete event-oriented trends meant to represent a continuous process. Simulation techniques are better suited for modeling machine performance as it relates to a continuous process. The results of cryocooler performance characterization over
time can best be analyzed by developing mathematical models that lead to high fidelity simulation of the machine as it operates during its design life while integrated into a spacecraft.
Simulation then allows cooler users to forecast their system response to a series of complex and interrelated cooler transient and steady-state parameters, spacecraft operating conditions, and space environmental factors. The development of a cryocooler performance simulation tool is itself a four-step process as shown in Figure 1. Each of the four steps shown are interrelated activities that begin with known information about the cryocooler to be simulated. The end results is the simulation itself, which is capable of both real time and accelerated performance prediction. This paper focuses on activities required to accomplish the first task of the simulation development process, while incorporating elements of the fourth task in order to verify accuracies of derived numerical equations. The end result is the empirical mathematical models that accurately predict the performance of the cryocoolers under evaluation at the CCL. These models, in turn, will form the basis for future development of complete simulation tools that depend on time history data and spacecraft operational and mission parameters not yet available at this stage of characterization of these long-life machines. Obviously, the development of the mathematical models described here are not meant to be an in-depth treatise on the theoreticallybased thermodynamic performance prediction normally used during the design/development phase of each cooler. Rather, the methods employed are meant to precisely describe the actual
Figure 1. Cryocooler Performance Simulation Development Process.
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operation of each specific cooler experiment article, resulting in a user’s tool for better understanding existing hardware.
METHODOLOGY FOR MODELING CRYOCOOLER PERFORMANCE Objectives There are two objectives for developing the mathematical models for performance prediction. The first identifies the most significant parametric variables that affect cooler performance. The variables identified are then analyzed to quantify the strength of the association among the independent variables and the outcome or dependent variable. The resulting performance prediction equation based on the extent, magnitude, and strength of the relationships among the variables is then used to describe the operation of the machine. In particular, the effects of heat rejection temperature on cooler performance and input power requirements will be quantified. Approach
Experience with analyzing the performance of the coolers under evaluation at the CCL shows that their operation can be described as functions of two or more independent variables. A linear least squares multivariable regression analysis proved to be the best approach available for analyzing how the cooler operates based on multiple variable interactions. This technique involves analyzing how two or more independent variables affect a single dependent variable. The single dependent variable should be the one controllable by the spacecraft based on desired operation of its cooled subsystem. Results of multivariable regression analyses are used to characterize the relationship between the dependent and independent cooler operating parametric variables by determining the extent, direction, and strength of association among the interrelated factors. Assumptions
There are important assumptions applicable to all of the coolers being modeled that are key elements affecting the results of their multivariable regression analyses. These assumptions are: 1. The actual independent variables affecting the cooling load could be readily recognized
through a qualitative review of the data from experiments on each of the units. 2. The performance trends associated with the independent variables for each of the experiment articles were smooth and could be approximated by linear or second order polynomial expressions for the independent variables and their coefficients. 3. The range of interest with respect to independent and dependent variable interactions is within the design envelope of the coolers or its expected space application. 4. Strong relationships among variables (dependent and/or independent) do not prove cause and effect, only correlation of trends as observed in discrete data. 5. Measurement error is taken into account when evaluating the accuracy of regression analysis results. 6. Turbo Brayton evaluation and regression analysis assumes no temperature difference between the compressor inlet and the cooler heat rejection temperature. 7. Certain controllable parameters that are not normally adjusted during cooler operation are not included in the variable assessments (e.g., charge pressure, phase angle, operating frequency).
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Construct Initial Defining Linear or Nonlinear Curve Fitting Equation The initial defining equation is determined based on the number of independent variables identified or selected to define the cooler’s performance. In most cases, there were three independent variables and one dependent variable. The basic form of the equation for determining the least squares curve fit of the data using multivariable regression is
where and are constants representing coefficients of the polynomial expression determined through regression analysis and and represent the independent variables while 7 represents the dependent variable. There are variations of this defining equation, where n is the number of independent variables. Each variation of the equation contains different combinations of the independent variables as they could relate to the dependent variable. Regression of all possible equation constructs as well as stepwise regression for each of the coolers showed that Eq. 1 represented the best form of the equation for determining performance prediction models, except in the case of the reverse Brayton cycle cooler as describes later. Independent and Dependent Variable Selection and Ranking There are a number of parameters associated with cooler operation that could serve as independent variables by which the dependent variable could be defined. These include heat rejection temperature cooler input power cold end temperature heat load operating frequency (f), charge pressure (p), and phase angle. In searching for the best possible variable candidates for the defining equation, regression analyses were performed and the resulting coefficients were tested for stability as each separate term was eliminated. Through this method and by examining the resulting regression analyses for correlation coefficient strengths of variable interaction, the following variables were identified as the most significant (in rank order of significance): (1) (2) (3) When was determined to be the dependent variable. Therefore,
Final Performance Prediction Equation Development and Validation
Based on the variable selection and use of the defining equations for the type of regression analysis conducted, the following equation serves as the final performance prediction equation form used to conduct the least squares multivariable regression analysis on the coolers under evaluation (again, the reverse Brayton cooler was the exception as described later in the paper): where E is the residual or error associated with this approximation technique. Using the computer applications program Microsoft Excel with its regression analysis capabilities, the coefficients of the expression shown in Eq. 3 were determined for each of the coolers. The accuracy of the resultant expression was evaluated based on the value of the squared multiple regression correlation coefficient or value. The closeness of fit of the resulting curve defined by the performance prediction equation is represented by how close the value approaches unity. values less than 0.91 contained unacceptable levels of variation and were not used. Other tests such as normality of residuals, estimation of error based on near neighbors and variable stability during elimination tests were also conducted to determine adequacy of the final performance prediction equations.
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Figure 2. CSSC Composite Performance Summary.
For Figure 2;
where
and
The final equation is as follows:
CRYOCOOLER PERFORMANCE PREDICTION MODEL SUMMARIES Diaphragm Flexure Stirling Cooler (CSSC)
The steady state performance envelope for the CSSC is shown in Figure 2. There is very little evidence of interaction of the variables within the range of interest with respect to their upper and lower boundaries as indicated by the overlaying of the performance trend lines in the composite plots. Tests for parallelism show a high degree of nesting of the data with respect to each of the possible variations in parameter correlations. There is a slight downward concavity present in the load line composite plot, but linearity of the coefficients of the regression model using the basic form of the multivariable equation yields good results. The performance prediction equation for the CSSC unit under evaluation is shown in Eq. 4. Validation testing
using actual experiment data confirms the accuracy of the prediction model, with a standard error of less than 0.27 W predicted versus actual values over 16 runs made between the 270 K and 310 K range. The high value indicates very good correlation of the data with the multivariable regression model results.
35 K/0.3 Watt Pulse Tube Cooler (3503) The steady state performance envelope for the 3503 pulse tube cooler is shown in Figure 3. Just as with the CSSC unit, there is very little evidence of interaction of the variables within the range of interest with respect to their upper and lower boundaries as indicated by the overlaying of the performance trend lines in the composite plots.
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CRYOCOOLER INTEGRATION AND TEST TECHNOLOGIES
Figure 3. 3503 Composite Performance Summary.
For Figure 3;
where and
The final equation is as follows:
Tests for parallelism show a high degree of nesting of the data with respect to each of the possible variations in parameter correlations. Performance trends with respect to each of the variables show a high degree of linearity, validating the use of the basic linear least squares multivariable regression equation. The resulting performance prediction equation for the 3503 pulse tube unit under test is shown in Eq. 5. Validation runs using actual experiment data confirms the accuracy of the prediction model, with a standard error of less than 0.72 W predicted versus actual values over 20 runs made between the 280 K and 310 K range. The high value indicates very good correlation of the data with the multivariable regression model results.
60 K/2.0 Watt Pulse Tube Cooler (6020) The steady state performance envelope for the 6020 pulse tube cooler is shown in Figure 4. Just as with the CSSC and 3503 units, there is very little evidence of interaction of the variables within the range of interest with respect to their upper and lower boundaries as indicated by the overlaying of the performance trend lines in the composite plots. Tests for parallelism show a high degree of nesting of the data with respect to each of the possible variations in parameter correlations. Performance trends with respect to each of the variables show a high degree of linearity, validating the use of the basic linear least squares multivariable regression equation. The resulting performance prediction equation for the 6020 pulse tube unit under evaluation is
shown in Eq. 6. Validation runs using actual experiment data confirms the accuracy of the prediction model, with a standard error of less than 0.862 W predicted versus actual values measured over 16 runs made between the 280 K and 310 K range. The high value indicates very good correlation of the data with the multivariable regression model results.
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Figure 4. 6020 Composite Performance Summary.
For Figure 4;
where
and
The final equation is as follows:
35 K/1 Watt Pulse Tube Cooler (3585) The steady state performance envelope for the 3585 pulse tube cooler is shown in Figure 5.
Although most combinations of independent variable correlations with
do not indicate a high
degree of interaction of these factors, there is evidence that a small degree of interaction exists among variables. Interaction is illustrated in the load line convergence about the 63 K cold end temperature as shown in Figure 5. Tests for parallelism show a high degree of nesting of the data with respect to and However, parallelism fails when correlation factor tests include rejection temperature. There is a slight change in the slope of the load lines as a function of rejection temperature for this cooler. Performance trends with respect to each of the variables show a high degree of linearity, validating the use of the basic linear least squares multivariable regression equation. However, the interaction that is evident shows that the basic form of the regression equation should include a or expression for better accuracy. The resulting performance prediction equation for the 3585 pulse tube unit under evaluation is shown
in Eq. 7. Validation runs using actual experimental data confirms the accuracy of the prediction model, with a standard error of less than 3.85 W predicted versus actual values measured over 35 runs made between the 280 K and 310 K range. This equates to less than three percent error under normal operating conditions. The somewhat lower R2 value caused by the omission of the interaction term in the expression is still high enough to indicate very good correlation of the data with the multivariable regression model results.
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CRYOCOOLER INTEGRATION AND TEST TECHNOLOGIES
Figure 5. 3585 Composite Performance Summary.
For Figure 5;
where and
The final equation is as follows:
Single Stage Turbo Brayton Cryocooler (SSTC) In evaluating the significance of variables for the SSTC engineering model cooler, significant differences were evident in the causality of the relationships between independent
variables and the input power. Figure 6 shows the steady state load line performance of the SSTC. Just as with the CSSC, 3503, and 6020 units, there is very little evidence of interaction of the variables within the range of interest with respect to their upper and lower boundaries as indicated by the overlaying of the performance trend lines in the composite load line plots. Tests for parallelism show a high degree of nesting of the data with respect to each of the possible variations in parameter correlations. However, performance trends with respect to each of the variables show a much lower degree of linearity than with the other coolers. The nominal set of load lines at the setting shows obvious downward concavity, leading to the need for a negative second order term in the final mathematical performance prediction equation. In addition to this, a thorough review of the cooler’s inherent sensitivities to lower heat load applications indicate the need to include a term representing the sensitivity of the response to the cycle efficiency. The resulting performance prediction equation for the SSTC is shown in Eq. 8 in terms. Note that the variable rankings for this cooler indicated higher strengths of association of all variables to the heat load. The resulting equation is then solved for in terms of the same independent variables evaluated for the other coolers and including the significant impacts of the nonlinear trend lines and cycle efficiency effects. Validation runs using actual experimental data confirms the accuracy of the prediction model, with a standard error of less than 0.862 W predicted versus actual values measured over 16 runs made between the 280 K and 310K range. The very high value indicates exceptionally good correlation of the data with the multivariable regression model results. Again, it should be noted that this multivariable regression analysis assumes no significant temperature difference exists between the heat rejection interface and the compressor inlet gas temperature.
HEAT REJECTION EFFECTS ON CRYOCOOLER PERFORMANCE
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Figure 6. SSTC Composite Performance Summary.
where Solving Eq. 8 in terms of
and and substituting coefficients from regression analysis yields
HEAT REJECTION EFFECTS ON COOLER PERFORMANCE For most of the coolers undergoing evaluation at the CCL, it can be stated that the rate of change of the input power required to obtain a target operating condition for the cooler is independent of the heat rejection temperature. This does not mean that and are not related, but the relationship of and is independent of the relationship between and Therefore, does not interact with nor does interact with This is why there is no or interaction term in the polynomial expression of the linear multivariable regression equation. The exception to this observation is the slight interaction observed on the 3585 cooler where a and/or a -by-TR expression is needed for higher accuracy. The result of this relationship is that the effects on caused by and can all be
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CRYOCOOLER INTEGRATION AND TEST TECHNOLOGIES
independently evaluated. From all observations except the slight interaction case of the 3585 unit, merely shifts the linear trend lines relating and to The response curves for each cooler representing and affects on Pin all have the same general shape (linear or slightly curved) and differ from one another by the amount of shifting caused by an additive
constant related only to SUMMARY AND CONCLUSIONS The general solution equation for linear least squares multivariable regression analysis yields very good performance prediction equations for each of the coolers evaluated at the CCL. The addition of contributing expressions relating thermodynamic performance to cooler input power, as well as observed trend lines not accurately represented by linearization improve accuracies of the resulting performance prediction equations. Future inclusion of transient effects, timehistory performance trends, extended operating range data, and investigations of improved mathematical terms in the expressions describing nesting and interaction of independent variables will also help improve cryocooler prediction models. When linked with spacecraft operating and environmental data, these performance prediction models can be used to develop complete cryogenic system simulation tools to help with system level thermal management design activities.
REFERENCES
1.
Roberts, T. and Bruning, J., “Hughes Aircraft Company SSC I & II Performance Mapping Results”, Cryocoolers 9, Plenum Press, New York (1997), pp. 38-42.
2.
Kleinbaum, D. and Kupper, L., “Applied Regression Analysis and Other Multivariable Methods”, Duxbury Press, North Scituate, MA, 1978.
3.
Roberts, T., “Cryocooler Transient Performance Modeling”, Cryocoolers 9, Plenum Press, New York (1997), pp. 163-171.
4.
Roberts, T. and Bruning, J., “Creare SSTC Acceptance Test and Performance Mapping Results”, Phillips Laboratory (AFRL) Interim Report, Kirtland AFB, NM (1996).
Cryocooler Working Medium Influence on Outgassing Rate V.F. Getmanets Special Research & Development Bureau in Cryogenic Technologies Institute for Low Temperature Physics & Engineering National Academy of Sciences of Ukraine Kharkov, 310164, Ukraine G.G. Zhun' Kharkov Polytechnic University Kharkov, 310002, Ukraine
ABSTRACT The influence of four cryocooler working fluids or purge fluids on the outgassing rate of internal materials is studied. At gas exchange rates higher than 0.5 1/min it is found that the outgassing products, cryodeposit growth, and the transportation rate of outgassing products to heat exchangers depends only on the rate at which outgassing products diffuse to the surface of internal materials. Gases with a high affinity for adsorbing onto surfaces slightly decrease the outgassing rate in the cryocooler by blocking the transport of products to/ from material surfaces. INTRODUCTION In modern cryocoolers, numerous polymers and composite materials are used. They are used as piston liners, displacer elements, electronic circuit boards and components, wire potting and insulation, and structural elements made of carbon fiber and fiberglass. In a vacuum, and especially at elevated temperatures, they heavily outgas various products that easily condense at low temperatures. These outgassing products are gases dissolved inside the constructional materials. Such contaminates often shorten the lifetime of cryogenic systems by plugging heat exchangers, coating optical windows and mirrors1, or saturating vacuum vessel adsorbers2. Condensed contaminants also increase the emissivity of thermal shields3, and worsen the thermal characteristics of superinsulation3,4 (see Fig. 1). Contamination of vacuum-system applications has been studied for a long time, and progress toward the solution of these problems has been achieved. The same gas contamination processes take place in cryocoolers. However, due to the difficulty of such investigations, little data exist on the contamination processes interior to a cryocooler. In the few data that do exist, some authors have speculated that outgassing ceases in an atmospheric gas medium. They thus propose filling vacuum volumes with carbon dioxide5,6.
Cryocoolers 10, edited by R. G. Ross, Jr. Kluwer Academic/Plenum Publishers, 1999
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Figure 1. Variation of emissivity
over time due to condensation and cryopumping of cryodeposits
throughout the thickness of multilayer superinsulation (MLI) surrounding a liquid nitrogen dewar; the
noted temperature reflects a location within the MLI ranging from room temperature on the outside to 77 K at the inner layer.
This article describes an experimental study of the dependence of outgassing rate on the flow rate of four cryocooler working fluids or purge gases and The experimental approach and apparatus are schematically illustrated in Fig. 2. The researched material (fiber-
Figure 2. Sketch of the experimental apparatus used for the material outgassing investigation: 1chamber containing the tested material; 2- electric heater for temperature control; 3- tank of purge gas; 4known volume for gas flowrate measurement; 5- throttle valve; 6- pipe for filling chamber with purge gas;
7, 8, 15, 18- vacuum lines; 9- membrane with calibrated orifice; 10, 11, 12, 13- pressure transducers for pressure measurement before and after the diaphragm; 14- cutoff valve; 16- adsorption pump for the collection of outgassing products; 17- temperature sensor; 19- liquid nitrogen vessel for cooling the adsorption pump; 20, 21- evacuation couplings.
WORKING MEDIUM INFLUENCE ON OUTGASSING RATE
735
glass or aluminum alloy AMC with bands of width 25-30 mm, length 80-100 mm, and thickness of 0.8-1.2 mm) was placed in a 35 liter AMC alloy chamber. The relative evacuation rate through one pipeline was equal to (where total material surface, and
the effective evacuation area of the pipeline). The material evacuation and outgassing investigations were conducted at room temperature (294 K), and also at higher temperatures, attained by electric heater -2. Chamber temperature was fixed by several temperature gauges -17. The materials outgassing was measured by the flow method7 using calibrated orifice diaphragms -9, and pressure gauges -10, -11, -12, -13 located before the and after the diaphragm, giving pressure drop
The analyzed outgassing products (including
were evacuated and adsorbed by adsorption
pump -16. Sampling of outgassing products for analysis was conducted using adsorption pump
-16 at liquid nitrogen temperature. Following Getmanets's theory7-9 of evacuation and outgassing, the process of gas removal follows in four sequential phases: 1) The gas is evacuated from the open volume being evacuated; 2) The outer layers of multi-layer adsorbed gases come off from interior material surfaces; 3) The final monolayer of adsorbed gases leaves ulterior material surfaces; 4) Gases diffuse to the surface from the interior of solids within the evacuated volume; Only in the fourth (diffusion phase) is the outgassing rate an unambiguous function of the material. The beginning of the diffusion phase is determined by the criterion7,8. where the characteristic lifetime of gases dissolved in a solid body, and in the vacuum volume (with allowance for repeated readsorption in the monolayer) are determined by the ratios:
where is the diffusivity of the gas dissolved in the solid, is the thickness of the solid (halfthickness for the case of evacuation from two sides), and is the time of adsorption.
The most important characteristics of the four listed phases are the pressure of the molecules in the monolayer, and the outgassing rate associated with the beginning of the diffusion phase. However theory7,8 does not give analytical methods of definition for the above parameters of evacuation and outgassing phases. They must be determined experimentally. Typical dependencies of specific outgassing rate on time for fiberglass are shown in Fig. 3; included are two temperatures and several relative evacuation rates for the desorption and diffusion phases. Analysis of these data shows that in the first three phases of outgassing, the outgassing rates are increasing due to the increase of relative evacuation rate S. At the function has a shape similar to that of the letter "L" at 294 K and 390 K. The knee of the "L" is associated with the completion of the monolayer desorption, and the beginning of the diffusion phase from the interior of solids. The location in Fig. 3 where the knee is the most prominent is at this location has been designated by a large star. Let us consider that the diffusion outgassing phase8,9 begins at this point. The transition from desorption-controlled outgassing to diffusion-controlled is indicated by the change of evacuation rate9. In the desorption phase, increasing increases outgassing rate, and in the diffusion phase, the outgassing rate remains constant (see Fig. 3). The outgassing rate in both the desorption, and diffusion phases is increased by increasing temperature. For example, for a temperature increase from 294 K to 390 K the outgassing rate increased by a factor of 3.6-3.8 (see matching curves: 1 and 4, 2 and 5, and 3 and 6 in Fig. 3). Experimentally, it has been observed that for the same material samples, outgassing at various temperatures (for example, 294, 350, 390 and 420 K) before the beginning of the diffusion
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CRYOCOOLER INTEGRATION AND TEST TECHNOLOGIES
Figure 3. Dependence of specific outgassing rate from fiberglass on time of evacuation Temperature of evacuation (degassing) K: 1-3, l'-3', l"-3"; K: 4-6, 4'-6', 4"-6"; Relative evacuation rate l'-3', 4'-6'-outgassing in volumetric-desorption phase with relative speed twice the value for processes described accordingly 1-3, 4-6; l"-3", 4"-6"= outgassing in the diffusion phase with relative speed twice the value for processes described by dependencies 1-3 and 4-6, accordingly; 7, 8: volumetric-desorption and diffusion phases respectively; 9- changes of outgassing of a material in the volumetric desorption phase after doubling of the evacuation rate; 10- Point designating the beginning of the diffusion phase of material outgassing.
phase, that after consequent cooling down to room temperature, the outgassing rates become identical and equal to the value at 294 K. Water Molecule Interactions
Next, we studied the effect of a water-molecule monolayer on the diffusion rates of gasses from solid materials at various vacuum levels. In these tests, fiberglass samples were evacuated
to different pressures, all being lower than the monolayer filling pressure Next, the evacuation was stopped and the pressure was fixed at various levels. The results are shown in Figs. 4a and 4b. As we see, after monolayer filling by water molecules, the pressure is essentially constant. This confirms that a water-molecule monolayer on the surface prevents other gas species from departing into the vacuum volume outside of the solid material. Only a small pressure increase above is observed due to solid body outgassing of molecules of nitrogen, methane (curves 16), and hydrogen (curve 1') which are passing from the material body to the vacuum volume through the water monolayer. and molecules also are dissolved in the material body, but to a much lessor extent than the water molecules. Purge Gas Interactions Next, the outgassing of fiberglass plastics and AMC alloy was investigated with various purge gasses, but only for the diffusion evacuation phase. The approach of these experiments was as follows: first, the outgassing of fiberglass and AMC alloy was measured at temperatures of 294 and 390 K over 250 hours. After that, the investigated material was extracted from
chamber 1 in Fig. 2. Next, new samples of a similar material were placed in the chamber. For
WORKING MEDIUM INFLUENCE ON OUTGASSING RATE
737
Figure 4. Influence of water vapor pressure, and thus filling of a surface monolayer of water, on outgassing from bulk fiberglass (a) and insulation spacer fibers (b): 1- pressure rise after evacuation cutoff at hydrogen getter presence; 1'- pressure rise at the absence of getter; 4,5- superposition of the curves 2 and 3 on curve 1 after being shifted by the time 6- defined in Figs. 4a and 4b, respectively.
them, the evacuation was conducted until the diffusion outgassing phase was reached. At this time a flow of a purge gas was initiated through chamber 1 (with the help of valve -5) with a
pressure of After a certain time interval, flowing ceased and evacuation of the purge gas was initiated; measurement of outgassing from the material followed. Then the flow of the purge gas was again continued through the chamber for a certain time, etc. Figures 5-8 show a comparison of the outgassing rates obtained in the diffusion phase for fiberglass plastics and an AMC alloy for purges using gases of and Analysis of Figs. 5-8 shows that outgassing occurs into the purge gas medium. For He and Ar, which are characterized by a small energy of adsorption (see Figures 5 and 6), outgassing values similar to those for outgassing into a vacuum were obtained at high purge rates. For example, for fiberglass plastics (with temperature 390 K), with a helium purge with volumetric flowrate outgassing values were obtained which are the same as those for vacuum outgassing (see dependencies 5 and 6 in a Fig. 5). For lower helium purge rates V, the rate of outgassing is higher than for the diffusion phase (for comparison see dependencies 3 and 5, and also 4 and 5 in Fig. 5). To reach outgassing equal to that in vacuum, a purge at 294 K should be executed with a velocity of 0.46 1/min. This velocity is smaller by 13% than at 390 K; this is seen from curves 6 and 8 in Fig. 5. For the AMC alloy, similar peculiarities are observed during the use of a Helium purge. However, in this case to attain outgassing equal to diffusion into vacuum, the purge rate at 294 K and 390 K needed to be 3.8 and 4 times smaller, respectively. Purging of the fiberglass plastics and AMC alloy with gaseous Ar also allowed one to obtain outgassing equal to diffusion into a vacuum. However, this required gas flows 11-17 % higher for the same temperature than for He (see Figs. 5 and 6). For purging with gaseous the outgassing rate did not achieve that equal to diffusion into a vacuum. It was 30-50% above diffusion, as is seen from curves 6 and 7 in Fig. 7. This is due to molecules adsorbing onto the fiberglass surface during purging, thus blocking material outgassing. Therefore, after purging of the fiberglass for a certain time, the outgassing rate into a vacuum was measured. The value of W appeared higher than that for a diffusion phase and the divergence between curves 6 and 7 increased with time. Outgassing into was the poorest of the four investigated gases owing to the significant adsorptance of the onto both the fiberglass plastics, and onto the AMC alloy at 294 and 390 K. Therefore, using this gas for purging at any purge rates could not achieve outgassing values equal to diffusion. This is seen from a comparison of curves 5 and 6, and 4 and 7 for fiberglass plastics, and also 9 and 10, and 8 and 11 for the AMC alloy in Fig. 8. From the above results it is clear that substantial outgassing occurs into a purge gas. The outgassing rate is determined by the temperature, by the gas used, and by the flow rate of the
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Figure 5. Variation of outgassing rate (W) in the diffusive regime with evacuation or purge time with helium; in curves 1-4 the materials were evacuated, while in curves 5-12 the materials were purged with dry helium; the purge rate with dry He in (1/min) for each curve is, respectively: 5- 0.21; 6- 0.35, 7- 0.52; 8- 0.46; 9,10- 0.06, 11- 0.14; 12- 0.12.
Figure 6. Variation of outgassing rate (W) in the diffusive regime with evacuation or purge time with argon; in curve 1 the material was evacuated, in curves 2-11 the material was purged with dry argon; the purge rate with dry Ar in (1/min) for each curve is, respectively: 5- 0.36; 6- 0.50, 7- 0.59; 8,9- 0.10; 100.17, 11-0.14.
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Figure 7. Variation of outgassing rate (W) in the diffusive regime with evacuation or purge time with dry nitrogen; in curve 1 the material was evacuated, in curves 2-11 the material was purged with dry nitrogen; the purge rate with dry in (1/min) for each curve is, respectively: 5- 0.35; 6- 0.46, 7- 0.52; 80.65; 9-0.06, 10-0.10, 11-0.16; 12-0.19.
Figure 8. Variation of outgassing rate (W) in the diffusive regime with evacuation or purge time with dry carbon dioxide; in curve 1 the material was evacuated, in curves 2-11 the material was purged with dry carbon dioxide; the purge rate with dry in (1/min) for each curve is, respectively: 5- 0.51; 6- 0.71, 7- 0.96; 8- 0.988; 9- 1.20, 10- 0.30, 11- 0.48; 12- 0.26; 13- 0.41.
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Figure 9. Time variation of outgassing rate and its composition for the EVTI-28 superinsulation at 323 K: 1- water; 2- hydrogen; 3- nitrogen; 4- methane; (– – –) after cooling to room temperature.
purge gas blowing by the material. Use of a purge gas with a high adsorptance onto material surfaces (such as and to some extent can substantially decrease the outgassing rate. On the other hand, it is seen that at sufficiently high purge rates (0.5 liters/min for He), outgassing into the purge gas increases and approaches that into a vacuum. At the same time, if the purge is absent, outgassing reduces to near zero. It is seen that results of these experiments are absolutely identical to the outgassing phenomena in vacuum for two regimes: 1) the mode of a high evacuation rate and diffusion-controlled outgassing; and 2) the mode of switched-off evacuation (see Fig. 4). In the switched-off case, the pressure quickly rises up to the level, the monolayer filling pressure of the adsorbed gas molecules. Again, after monolayer filling by water molecules, pressure almost does not change. Thus, outgassing into the operation media has been stopped. So if the purging or vacuum pumping is stopped, despite their difference, the partial pressure of water molecules grows again up to and then stops growing. Outgassing into the working medium stops also. With the initiation of purging, water molecules begin to be removed from the volume. This results in the decrease of the monolayer filling, and therefore to increased gas diffusion from the solid body. When the purge rate becomes large enough to remove all of the outgassing products, we again transition into the diffusion outgassing phase, but into the purge gas volume. This phenomenon will also be observed in the cryocooler working volume filled with gaseous helium as the working medium. The matter is that inside the cryocooler periodical gas exchange from the compressor and back occurs due to the compressor cyclic operation. Being dissolved in the gaseous helium, water vapor molecules in a cyclic manner enter into the displacer unit and condense on the regenerator surfaces and other cold surfaces. The then dried-out helium returns to the compressor and captures a new water molecule for transport to the displacer. As helium flow rates in a cryocooler are close to the critical value of 0.5 1/min, the outgassing from the interior materials surfaces is close to its value in a vacuum. CONCLUSIONS 1. Firstly, a method was developed and experimental studies were conducted on material outgassing rates into four cryocooler working fluids or purge gases. 2. It was established that there exists some critical flowrate of the working gas, above which outgassing into this working medium becomes close to or equal to the outgassing rate into a vacuum.
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3. It is shown that at flowrates above the critical one, poorly adsorbed gases (helium, argon) do not influence the outgassing rate into the working medium. Gases which are adsorbed well (nitrogen, carbon dioxide) decrease to some extent the outgassing rate into the working medium due to partial filling by them of the adsorbed monolayer on the material's surface. 4. It was revealed that for cyclically operating cryocoolers the growth rate of the cryo deposit layer is determined by the outgassing rate and volumetric helium flow rate inside the cryocooler body. At helium flow rates higher than 0.5 1/min, the cryodeposit growth and its transportation rate to the heat-exchangers depends only on the materials outgassing in the diffusive phase. REFERENCES 1.
Getmanets, V.F., Grigorenko, B.V., Kurskaya, T.A. et al., “Identification of the cryogenic-vacuum
impurities of the fiberglass cryostats thermal protecting systems”, Kharkov, FTINT ANUkr, Preprint N 44-89, 1988, 11 P. 2.
Zhun', G.G., Podolsky, A.G., Getmanets, V.F., et al., “Complex of highly-effective equipment for
surfaces preparation for metallization,” Proceedings of the Ukrainian Vacuum Society, Kharkov, 3.
1996, V. 2, pp. 250-259. Zhun', G. G., Getmanets, V. F. et al., “ Investigation and elimination of the factors worsening cryovessel efficiency”, Inzh. Phys. Zhurn, 1989, V. 56, N. 2, pp. 271-276.
4.
Zhun', G.G. Podolsky, A.G., et al., “About the influence of the thermal insulation evacuation conditions on the cryovessel lifetime,” Investigation of the thermophysical properties of the working media and processes of the cryogenic engineering and air-conditioning, Leningrad, LTI, 1987, pp.
5.
Perkins, P., Dengler, R., Self-Evacuated Multilayer Insulation of Light Weight Prefabricated Panels for Cryogenic Storage Tanks, NASA TM-405, March 23, 1958.
6.
Kaganer, M.C. Heat and mass exchange in low temperature constructions, Moscow, Energy, 1979, 256 P. Getmanets, V.F., Large-scale onboard cryostatting systems (design fundamentals, heat and mass transfer, development of the effective thermal insulation, Diss. Dr. Sci., Kharkov, ILTPh&E of
39-45.
7.
Ukraine, 1985, 364 P.
8.
9.
Getmanets, V.F., Mickhalchenko, R.S., “Evacuation and outgassing of the vacuum systems being heated,” Kiev. Naukova Dumka, In: Investigation of the processes in cryogenic and vacuum systems, 1982, pp. 36-86.
Verkin, B.I., Mikhalchenko, R.S., Getmanets, V.F., Mikheev, V.A., “Application of Multilayer insulation in Cryogenic Engineering and Improvement of its Efficiency,” Proc. of ICEC-10, 1984, pp. 529-538. 10. Zhun', G.C., Schalaev, V.J., In: Studies of thermophysical properties of operating medium, cryogenic processes and air conditioning, Leningrad, Leningrad Technol. Inst., 1986, pp. 9-13.
Accelerated Cryocooler Life Tests for Cryodeposit Failures V.F. Getmanets, G.G. Zhun Special Research & Development Bureau in Cryogenic Technologies Institute for Low Temperature Physics & Engineering National Academy of Sciences of Ukraine Kharkov, 310164, Ukraine
H. Stears Orbita, Ltd., Kensington, MD
ABSTRACT In other work by these authors1 it was experimentally found that with a helium circulation rate of over 0.5 1/min internal to a cooler, the outgassing process inside the cryocooler is as intensive as it is under a high vacuum. During 20 years of theoretical and experimental outgassing studies2,3,7 we have also found that the outgassing rate in the diffusion mode is quite repro-
ducible. Using these results, an accelerated cryocooler test method for cryodeposit failures has been developed and experimentally validated. The proposed approach allows the required test time to be reduced by a factor of 20 to 50. INTRODUCTION
Cryodeposits in cryocoolers are solids that form on cryogenic surfaces such as regenerator screens and heat exchanger surfaces due to the condensation of contaminant gasses. The contaminant gasses are products that have outgassed into the working medium, usually helium, from various porous constructional materials internal to the cryocooler, particularly from polymers and composite materials. It has been experimentally found1 that for helium circulation rates over
0.5 1/min, the outgassing process inside the cryocooler is as intensive as it is into a high vacuum. Outgassing is a process that occurs continuously over time; there are no observable ways to significantly speed the process up for purposes of accelerated testing. Even a method of preheating the entire cooler up to 100 to 120°C (which is critical for standard cryocoolers) does not seem to give a large effect. Thus, a more analytical approach to life prediction has been developed, based on extrapolating two key experimental measurements made on the subject cooler: 1) Measurement of the volume of condensed gas that is needed to cause an unacceptable cryodeposit level associated with reduced cooler efficiency or increased coldend temperature 2) Measurement of the long-term average outgassing rate of contaminant gasses internal to the cooler from the start of operation up to the point of unacceptable cryodeposit level Cryocoolers 10, edited by R. G. Ross, Jr.
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Figure 1. Comparative outgassing at room temperature of a 35-liter cryostat as a function of evacuation time in the diffusion mode: 1- vessel with superinsulation having a weight of 1.4 kg without adsorber; 2- the same as #1, but with a carbon adsorber having a mass of 0.6 kg.
The test period needed for these measurements lasts no more that 0.1 to 0.01 of the desired cryocooler lifetime; thus one achieves a very accelerated life prediction. Based on the above, a cryocooler's predicted lifetime as limited by Cryodeposits is estimated as follows:
One means of determining the value of is to introduce a selected contaminant gas mixture into the operating cryocooler at a controlled rate. The moment of cryodeposit failure is determined by monitoring the power/efficiency performance of the cryocooler over time. The main difficulty is selecting the mixture composition of the introduced gas. The overriding issue is selecting an approach that will provide a useful forecast of the likely outgassing process over the cooler life (say 5 to 20 years) based on data collected in a one- or twomonth period of accelerated testing. A successful approach contains four elements: • • • •
Available outgassing experimental results (e.g. Figs. 1 and 2) as a function of time A good understanding of outgassing process physics Highly reproducible experimental results Specialized test facilities and equipment
Figure 2. Comparison of the outgassing volumes from superinsulation on a 35-liter cryostat from duration of maintenance data: 1 - Vessel with superinsulation of weight 1.4 kg without adsorber; 2 Vessel as in item 1 with carbon adsorber having weight 0.6 kg; 1,2 - are obtained by an integration of curves 1,2 on - volumes obtained from the ratio (1) on the outgassing accordingly at the end of 1st, 6th, 12th and 24th months of evacuation (according to curve 1 in Fig. 1).
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Figure 3. Time variation of the outgassing of glass papers according to the data of different au-
thors: 1, 3, 7-10- flow method; 2- accumulation method; 4-6- weight method; (SBShT) glass paper, Getmanets et al.; 3- Tissuglass, Glassford A.P.M.; 4- Dexiglass, Glassford A.P.M., Lin C.K.; 5- Tissuglass, Glassford A.P.M., Lin C.K.; 6- Tissuglass, Keller; (SBSh-T), Kupriyanov V.I. et al.; (SBR-M), Fetisov Yu.M., Kaganer M.G.; 9- Dexter-1246, Kurzner K., Wietzke I.; (EVTI-7), Kaganer M.G., Fetisov Yu.M.
Figure 4. Investigation of outgassing of a SBSh-T glass paper by the flow method and accu-
mulation method: 1 - Flow method; 2 - method of accumulation; 3 - modified method of accumulation with a correction factor
OUTGASSING LAWS AND PROCESS REPRODUCIBILITY
Until recent times, the above interpretation of the outgassing phenomena was not understood. The general belief was that the outgassing process, even for a single material, was not reproducible due to crucial unit-to-unit differences in the manufacture and storage of individual material specimens. Actually, such a viewpoint sounds reasonable, since outgassing measurements made by various researchers on the same material (like fiberglass, for instance) are often dramatically different (Fig. 3). Also, measured outgassing rates for the same material may dramatically vary depending on the measurement method used, such as the flow or accumulation method (Fig. 4), or even depending on the gas-sample evacuation rate used (Fig. 5).
Figure 5. Effect of the evacuation rate on the measured outgassing rate in a SI blanket with a thickness of 50 mm, and comprising of
100 layers of perforated diffraction screens with a at evacuation rates and 8.8 1/s.
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Figure 6. Time variation of outgassing from glass papers according to data from different authors: 1,3 - flow method; 2- accumulation method; 4- weight method; (SBSh-T) glass paper, Getmanets et al.; 3Tissuglass, Glassford A.P.M.; 4-Dexiglass, Glassford A.P.M., Lin C.K.
To resolve this history of uninterpretable results, we have drawn upon 20 years of personal research experience in experimental and theoretical fields2, and have tried to find a reproducible way to quantify outgassing. As a measure of our success, Fig. 6 shows a time-related plot of the outgassing characteristics of four different fiberglass papers as obtained by four author groups3-6 using different measurement methods and test equipment. Note that all four characteristic curves virtually coincide. The four data groups in Fig. 6 are united with a single, but very important normalization, based on the physics of outgassing that we develop below.
Physics of Outgassing
For the purpose of understanding the underlying fundamentals, we appeal to Fig. 7, which uses Getmanets' model2 to break the process of outgassing into four distinct regimes, each described by its own physics: 1) Evacuation of the open interior volume of the cryocooler 2) Desorption of the outer layers of multi-layers of adsorbed gasses on interior surfaces 3) Desorption of the final monolayer of an adsorbed gas on interior surfaces 4) Diffusion of gasses from the bulk of internal construction materials
Bulk Diffusion (Region 4) — The ultimate outgassing rate is controlled by the fourth "diffusion" regime, which is characterized by the fact that the outgassing rate is dominated by the output of gases dissolved in the interior of contructional materials. We have both theoretically and experimentally determined 2,3,7 that it is only in this diffusion regime that the outgassing rate does not depend on the vacuum pumping conditions, but is determined as a typical time-related function (Figs. 1,2,6). This statement follows from the following experimentally obtained relationship2 for the outgassing rate in the diffusion regime:
Figure 7. Pressure variation during material evacuation according to the Getmanets' model: 1- gas removal from the bulk volume; 2- desorption from multiple adsorbed surface layers; 3- desorption from the final surface monolayer; 4- molecular diffusion from the interior of solid materials
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for for
where: – average thermal velocity of molecule migration – quantity of molecules per unit area in saturated, surface-adsorbed monolayer – dissolved-gas-molecule diffusion time through bulk material – length parameter over which diffusion occurs – Diffusion coefficient for gasses dissolved in material
a – interatomic space of outgassing material R – universal gas constant T – absolute temperature – energetic barrier for molecule transition from solid-state body to its surface – constant value is described by the equation:
where: m – molecular weight of outgassed products – Avogadro number – molecule adhesion coefficient (for molecule-to-surface collision)
Polylayer Desorption (Region 2) — In comparison to the diffusion regime (see equations (2 and 3), outgassing in the desorption regime2 conceptually depends on: – total surface area of all materials in the evacuated volume (cryocooler interior) – evacuation hole surface area (pump-out port) – theoretical probability of a molecule escaping through the evacuation hole, summed for each of the following three factors: a) the evacuation-hole post-counterpressure, b) the probability of a molecule flying up to the evacuation-hole, and c) the quantity of molecule-to-surface collisions prior to the escape of the molecule.
Monolayer Desorption (Region 3) —According to Getmanets' theory2 the outgassing rate in the third regime, that of desorption from the final monolayer, is depicted by the following equations:
where: – pressure of adsorbed monolayer filling and saturation – distance and time of molecule flight between two surface collisions – molecule lifetime within gas volume until escape into evacuation system – time of adsorption
For water molecules at room temperature some approximate values are: and thus the outgassing parameter is largely defined by physical adsorption processes associated with interior surfaces. As a result, the outgassing rate in the desorption regime will abruptly decrease with a decrease in the specific evacuation rate and the corresponding decrease in the probability of a molecule escaping from the system (which depends on the
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Figure 8. Time variation of outgassing rate and its composition for EVTI-28 superinsulation at 323 K: 1- water; 2- hydrogen; 3- nitrogen; 4 - methane; ( – – ) after cooling to room temperature.
length of the evacuation system structure). Also, the outgassing rate will continue to slowly decrease over time due to the observable exponential dependency of on the molecular lifetime parameter within the pumped-out volume (see (6)). A sharp increase in is associated mainly with an increase in the number of repeated wall re-adsorption cycles associated with the decreased evacuation rate. These fundamental dependencies governing the outgassing process have generally not been taken into consideration by previous researchers. For example, it is clear from the physics that water vapor1 (see Fig. 8), has a very long lifetime (up to for its surface adsorbed molecular layers2,3,7. Up until the moment of escape into the vacuum system, every molecule makes a huge number of repeated collisions and surface adsorptions (averaged as where is the ratio of the total surface area of all materials in the vacuum space (cryocooler interior) to the evacuation-hole area. Hence, the total time of residence of every molecule in the adsorbed status can amount to several seconds or even minutes. As a result, with piston-bore clearance structures at low evacuation rates (as within a cryocooler compressor), the time required for the removal of surface-adsorbed poly- and monolayers of water may last (at room temperatures) for many days or even weeks. It is worth saying that in regimes of adsorption from poly- and monolayers, the outgassing rate is strongly temperature dependent and speeds up considerably at temperature as high as 80 to 100°C. Because the durations of experimental research studies on outgassing often do not exceed more than a few days, many tests only involve the vacuum desorption regimes and never reach the diffusion regime. Unfortunately, the desorption regimes are the regimes that exhibit a strong dependency on the vacuum pump-out conditions as defined by the surface-area ratio This very fact can explain the strong outgassing dependency on vacuum conditions as seen by various authors (see Figs. 3 to 5). In particular, the effect of multiple repeated adsorption cycles, which is associated with a high value of fully explains the difference between outgassing values obtained using the flow method versus the accumulation method, Fig. 4. With the accumulation method, for instance, the pressure continuously increases after evacuation is terminated1. As a result, with the pressure rise, the number of molecules in the adsorbed monolayer grows; wherein remaining here is the prevailing mass of molecules diffusing from the material volume. Therefore, the accumulation method may underestimate (by one or two orders of magnitude) the resulting decreased outgassing effect, which can be estimated by the following ratio:
Here and are the distance and duration, respectively, of a molecule free flight in the volume until a repeated collision with the material surface and a consequent adsorption; parameter is the average thermal velocity of molecule migration. Once experimental data from the accumulation method have been multiplied by the K-parameter from equation (7), the resulting data will correlate remarkably well with outgassing data obtained by the flow method (see Fig. 4).
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The above discussion brings us to an important practical conclusion. Experimental outgassing data are quite reproducible functions of material and system factors, provided the two following conceptual conditions are observed: 1) During the outgassing process, it is important to complete the desorption regime and proceed into the material-controlled diffusion regime 2) Adequate methods and technical means need to be used to monitor the outgassing rate parameters With respect to the methods for monitoring outgassing rate, we employ a flow method using calibrated orifice diaphragms with associated pressure-drop gauges. In the flow-method, the outgassing factor is defined by the evacuation hole area parameter vs. pressure-drop thereon. The usage of such a diaphragm enables us to easily control the test conditions of the diffusionregime setup. When the diffusion regime is achieved, the monitored outgassing rate does not depend on larger or smaller diaphragm hole sizes any longer; in contrast, during the desorption regimens, it does. It is worth noting that the reproducibility of the outgassing effect (being a characteristic function of the material) is likely to be dependent on the specific chemical composition and manufacturing process of the material under study. Thus, if, for example, prior to the experi-
ments on outgassing phenomenon, one of two similar-identity tested samples is subjected to longterm evacuation and air-free storage under inert gas or vacuum conditions, its outgassing curve characteristics can expect to be significantly different. This change will actually be manifested as a time-shift; still, in a broader study of the aspect, some more complex unpredicted effects may also occur. ACCELERATED TEST METHOD
The above procedure has proven its worth as an accelerated test method for predicting the expected life of a cryocooler with respect to failure caused by plugging of its cryogenic heat exchanger by cryocondensate. Specific steps of the procedure include: 1) The cryocooler unit is first evacuated (it is possible at high temperature) until commencement of the diffusion regime. In the diffusion regime at room temperature, the vacuuming process lasts for about a month, during which the following data are monitored: i) the outgassing rate W0, and ii) the composition of the outgassed products (after cooling the unit down to room temperature).
2) The cryocooler is then filled with helium, is cooled down, and contaminant gasses matching the outgassing products are added at a controlled rate. The gas volume is then monitored to determine its level when excessive losses in efficiency or increased coldend temperature are achieved. 3) The value of the cryocooler life at the time of predicted cryodeposit-caused failure is estimated as:
Comparing this equation with Eq. 1, it can be seen that the average outgassing rate over the expected operational lifetime, is assumed to be equal to half of the measured magnitude of the outgassing rate Such an assumption (with plus-minus allowance) is based on the fact that in a 5 to 10-year period, the rate of outgassing diminishes to zero (or reduces by 10 to 12-times as compared to the beginning of diffusion regime, Fig. 1). Therefore, the real outgassing curve is replaced with a straight-line (with some overestimation). It is obvious from Fig. 2 that there is also a variation in the quality of the estimated outgas-
sing parameters depending on the test length over which the outgassing data are taken. Practically, it is quite enough to monitor the above data obtained within 1 or 2 test months, wherein the total volume value may be overestimated by a factor of two, but with a sufficient statistical reserve to compensate for any possible measurement errors plus any unexpected test factors. It is worth saying that the above disclosed methods are also quite appropriate for evaluation
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of a cryocooler cryostat lifetime via vacuum-maintenance capability parameter. It is just necessary to find, via an accelerated approach, a maximum permissible outgassing volume parameter that is responsible for unacceptable increased parasitic heat loads. EXPERIMENTAL VERIFICATION OF THE METHOD
The trustworthiness of the described accelerated outgassing method and its applicability to studying critical materials for long-life cryosystems has been proven by many years of the author’s experience in the field of expertise. Fig. 41 illustrates an example of the comparative pressure change in 5 years (in a real time scale) for two liquid-nitrogen filled vessels. One vessel was
manufactured using traditional technology, and the other was modified to comply with results of accelerated testing. The availability of accelerated test methods has led to the improvement in the thermal insulation design, the adsorption pump structure, and the technology of thermal-vacuum outgassing. As a result of these efforts, a 5-year lifetime for the vessel has been achieved without any measurable addition to the thermal parasitic influx rate. CONCLUSIONS 1. Updated accelerated life-test methods have be proposed and validated for testing susceptible cryogenic elements such as heat exchangers concerning cryodeposit failures. The methods are based on a long history of experimental and theoretical research conducted by the authors2 in the field of outgassing phenomenon for materials under vacuum. 2. The developed methods are based upon pure physical simulation of accelerated tests, and are based on two general and experimentally proven facts1,2: • The outgassing process from a material into a gaseous helium medium (wherein helium is circulation-pumped at a considerable rate) occurs with the same rate as in a vacuum (other gases, such as or for example, may somewhat decrease the outgassing rate).
• After termination of the surface desorption regimes from polylayers and monolayer, and after transition into the diffusion regime whereby dissolved gases are withdrawn from the material interior, the outgassing factor for materials or devices of standard-manufacture technology (other than vacuum technology) is a highly reproducible individual time function. 3. The principle of acceleration, or speeding-up, used in the accelerated life tests described here involves the separate measurement of two parameters: • The equilibrium rate and composition of the outgassing by-products in the diffusion regime • The volume of outgassing by-products required to reach a state of operational failure (i.e. end-of-life) These tests can be run consecutively on a single cryocooler unit. 4. The proposed life-test approach ensures a 20 to 50-times acceleration of the test program with preservation of an appropriate lifetime-forecast accuracy.
5. Such an approach has been successfully used by the SR&DB under real-world conditions for prediction of the lifetime of: • Vacuum-maintaining cryostats, along with their superinsulation • Cryogenic optics, in terms of contamination thickness buildup on cold surfaces ACKNOWLEDGMENT The authors express their gratitude and acknowledgment to the U.S. Air Force Research Laboratory, Kirtland AFB, New Mexico, and particularly Lt B.J. Tomlinson for the support of this work.
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REFERENCES
1.
Getmanets, V.F. and Zhun', G.G., “Cryocooler Working Medium Influence on Outgassing Rate,” Cryocoolers 10, Plenum Press, New York, 1999.
2.
Getmanets, V.F., Large-scale On-board Cryostatting Systems — Design Fundamentals, Heat and Mass Transfer, Development of the Effective Thermal Protection, Dissertation for Dr. Sci., Kharkov, ILTPh&E of Ukraine, 1985, 364 p.
3.
Getmanets, V.F., Mikhalchenko, R.S., Vacuuming and Outgassing in Unheated Vacuum Systems, Naukova Dumka, Kiev, (1982), pp. 36-86. Kutzner, K., Wielzke, I., “Mesverfahren zur Bestimeing von Ausgasungeraten Unteren Suchungen von Superisolations Materials,” Vacuum Technik., v.21, n.2-3, (1972), pp. 54-56. Gassford, A.R.M. “Outgassing Behavior of Multilayer Insulation Materials,” J. Spac. Rockets, v.7,
4.
5.
n.12, (1978), pp. 1464-1468.
6.
7.
8.
Kupriyanov, V.I., Chubarov, V.V., Tarasov, N.N., Dryamov, V.A. “Study of Material Properties in Vacuumed Multi-Layer Insulation”, Processes, Technology and Control in Cryogenic Machinebuilding, NPO Kryogenmash, Balashikha, (1976), pp. 141-148. Verkin, B.I., Mikhalchenko, R.S., Getmanets, V.F., Mikheev, V.V., “Application of Multilayer Insulation in Cryogenic Engineering and Improvement of its Efficiency,” Proc. of ICEC-10, 1984, pp. 529-538.
Getmanets, V.F., Zhun', G.G., Mikhalchenko, R.S., et al., “Cryogenic Superinsulations with Increased Efficiency,” Advances in Cryogenic Engineering, Vol. 43B, Plenum Press, New York, 1998, pp. 1319-1325.
Thermal Resistance Across the Interstitial Material Kapton MT at Cryogenic Temperatures L. Zhao and P.E. Phelan Department of Mechanical & Aerospace Engineering Arizona State University, Tempe, AZ 85287-6106
ABSTRACT
The present work concentrates on the measurement of the thermal resistance of a copper/Kapton MT/copper junction, in a flat-plate geometry, at cryogenic temperatures. Kapton
MT is a polyimide film with alumina filler particles and has a relatively low thermal resistance, but yet a high voltage standoff capability. The thermal resistance consists of two components: thermal contact resistance at the copper/Kapton MT interfaces, and the thermal conduction resistance across the Kapton MT film. The measured thermal resistance indicates that increasing the contact pressure reduces the thermal resistance, to a limit determined by the film conduction resistance. Increasing the contact pressure and the average interface temperature, and decreasing
the thickness of the interstitial layer, tends to decrease the thermal resistance. INTRODUCTION
Thermal contact resistance arises in the region of contact where two solid specimens are pressed together. It has long been realized that surfaces are rough on a microscopic scale, which causes the real contact area to be significantly smaller compared to the nominal contact area.1 Interstitial materials may be deliberately introduced in order to control the thermal contact
conductance (hc), hc = q/∆T, where q and are the heat flux and the temperature drop across the interface respectively. To enhance hard surfaces may be coated with soft metals of high
thermal conductivity by electroplating or vacuum deposition.2 The insert’s thermal conductivity and hardness, relative to the values of the corresponding properties of the base materials forming the contact, together with the insert’s thickness, dictate the resulting change in A particular category of thermal enhancing inserts under investigation is electrically insulating interstitial materials. One application is the superconductor current lead heat intercept connection, which consists of a structure where two concentric copper cylinders are
separated by a composite film. As the purpose of the heat intercept connection is to efficiently transfer the heat generated in the copper portion of the current lead, plus that conducted in from the ambient, away from the low-temperature superconductor magnet, it is imperative that the
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Figure 1. Schematic diagram of the experimental apparatus.
total thermal resistance of the heat intercept connection in the radial direction be kept as small as possible. At the same time, electrical isolation must be maintained between the inner and outer copper cylinders, necessitating the use of an electrically insulating interstitial material. It was determined previously that at the interface between the two copper cylinders, separated by a thin-walled G-10 tube, provided the largest component of the total thermal resistance.3 Therefore, in the most recent design, the G-10 tube is replaced with a thin Kapton film which, while still providing electrical isolation, is softer than G-10 and has a higher thermal conductivity, offering the potential for a reduced and thus a lower total thermal resistance.4 The present work concentrates on the measurement of the thermal resistance of a copper/Kapton MT/copper junction, in a flat-plate geometry, at cryogenic temperatures. The effects of the contact pressure, the average interface temperature and the thickness of the interstitial layer are evaluated. EXPERIMENTAL APPARATUS
The test system illustrated in Figure 1 is utilized to investigate a copper/Kapton MT/copper junction. It consists of two samples, upper and lower, and two calibrated heat flux meters which are pressed together and aligned to the center line of the apparatus. The alignment can be adjusted by the point contacts of the ball bearings. The test specimens are made of bare ETP copper. The material chosen for the insulating film is Kapton MT, a polyimide film with alumina particles as a filler, giving it a relatively low thermal resistance, but yet a high voltage standoff capability. The thermal conductivity of this material at cryogenic temperatures is sufficient to result in a small temperature drop across the sheet at the desired heat flux input, while the high
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dielectric strength of Kapton
allows only a thin layer to be used. Kapton MT sheets of various thicknesses are investigated in the experiments described later. The thermal conductivity of the heat flux meters was previously calibrated against a reference 304SS bar obtained from the National Institute of Standards and Technology. The temperature drops along the heat flux meters and across the sample interface are measured by
differential Type E thermocouples. The differential thermocouples are calibrated at a single point by immersing one junction into liquid nitrogen and the other into an ice bath, and applying the resulting percent deviation from the standard table to all measurements. The silicon diodes mounted on the copper heat mounts are utilized as the reference temperatures for the differential thermocouples. The thermocouples are inserted into diameter holes drilled to the center lines of the samples and heat flux meters, and which are partially filled with fine copper powder in order to provide reliable thermal contact. Indium foil is inserted at all junctions besides the sample interface in order to reduce extraneous contact resistance. Two thermal shields are placed above and below the heat flux meters in order to reduce the radiation heat transfer. The entire test column is surrounded by a thermal shield to minimize heat gain from the ambient. Before each experiment, a small amount of initial pre-stress is applied to the samples via a loading screw mechanism to guarantee a good contact. The pre-pressure, which has a magnitude of 0.809 MPa around room temperature, reduces to 0.007 MPa at 123 K and vanishes below 100 K. In order to apply and control the contact pressure at cryogenic temperatures, a bellows located below the lower ball bearing is pressurized with helium gas to obtain an evenly distributed pressure, which is measured by a load cell. The outputs of the thermocouples, silicon diodes, and the load cell amplifier are monitored and recorded by a Macintosh PowerPC computer, then simultaneously converted and analyzed by a labVIEW control program. If the temperature drops across the sample interface as well as those along the upper and lower heat flux meters are all within a set criterion for the last 50 data points, the system is considered to be in steady state. Starting with the reference silicon diodes located on the copper heat mounts, each temperature point is calculated sequentially from the differential thermocouple readings. The temperature drop at the sample interface is overdetermined by the calculation downwards and upwards, which are averaged to determine the final The heat flux at the sample interface is considered as the average of the upper and lower heat flux meters. All measurements are conducted under a vacuum condition of torr on average, which is measured by a cold cathode vacuum gauge. EXPERIMENTAL RESULTS AND DISCUSSION
Separate experiments are conducted for three different thicknesses of the interstitial Kapton MT sheets, (Kapton 150MT), (Kapton 300MT) and (Kapton 600MT). Before each experiment, a small amount of initial pre-stress is applied to the sample to guarantee a good contact. This pre-pressure decreases with reducing temperature, and eventually vanishes at cryogenic temperatures. The contact pressure is set in turn to approximately 0 MPa, 6.55 MPa, 13.10 MPa and 19.65 MPa by a pressurized bellows, while the average interface temperature remains fixed and is allowed to approach steady state. Measured values of the thermal contact conductance as a function of the average temperature across the sample interface for Kapton MT are presented in Figure 2, which displays the experimental data over an average temperature ranging from 35 to 110 K. Observe that increases with the contact pressure, P, because when the two surfaces are pressed tighter together, the soft Kapton film is squeezed into the voids between the actual contact spots,
thus increasing the effective contact area and causing
to diminish.
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Figure 2. Effect of contact pressure on thermal contact conductance for Kapton MT. The thermal conductance across the Kapton sheet, resistance, is defined based on its thermal conductivity, K, as:
where µm or
due to its thermal conduction
is the thickness of the interstitial Kapton layer under consideration, either 76 and which is assumed to remain constant, that is, to be independent of contact
pressure. The thermal conductivity, K, is taken from NIST.5
A well-known correlation for interstitial material) is given by6:
between two contacting randomly rough surfaces (no
where is the surface roughness, k the harmonic mean thermal conductivity, H the microhardness of the softer material in contact, and c and n are empirical coefficients found by fitting to experimental data. From this, is expected to increase with P. Moreover, the thickness of the Kapton layer is expected to reduce when it is pressed, therefore decreasing its contribution to the thermal resistance and resulting in an increased Presumably, increasing P further would yield a further increase in up to the limit posed by the resistance of the Kapton layer itself,
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Figure 3. Thermal contact conductance vs. average interface temperature at contact pressure of 6.55 MPa for three thicknesses of the Kapton MT layer.
given by Eq. (1). However, represents the maximum allowable pressure in our experimental apparatus. Figure 3 graphs versus the average temperature at the interface, for different thicknesses of the interstitial layer, at fixed contact pressures of 6.55 MPa and 19.65 MPa, respectively. The graph shows that increases as the thickness of the Kapton sheet is reduced. While the interstitial material gets thinner, its resistance decreases accordingly, which contributes to the enhancement of the total thermal contact conductance. Even so, Kapton 150MT provides substantial electrical insulation (dielectric strength = 4323 V). Figure 4 shows the effect of the contact pressure on for three thicknesses of the Kapton MT sheet, at fixed average interface temperatures of 111 K. The graph is consistent with our earlier observations, namely that increases with increasing contact pressure and decreasing thickness. CONCLUSIONS
The thermal contact conductance at a copper/Kapton MT/copper interface is investigated at cryogenic temperatures. The results indicate that increasing the contact pressure increases the contact conductance, in accord with expectations, for the greater the compression, the greater the effective contact area as a result of the soft interstitial material squeezing into the voids between the actual contact spots. The value of also varies inversely with the thickness of the interstitial layer. The thinnest layer provides the lowest thermal contact resistance, but yet sufficient electrical insulation.
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Figure 4. Effect of contact pressure on thermal contact conductance for three thicknesses of Kapton MT at average temperature of 111 K. ACKNOWLEDGMENTS
All the authors wish to thank Lisa De Bellis and Shijin Mei for their assistance in the experiment. P.E.P. gratefully acknowledges the support of the National Science Foundation through a CAREER Award (Grant No. CTS-9696003). REFERENCES
1. Fletcher, L.S., “Recent Developments in Contact Conductance Heat Transfer,” J. Heat Transfer, Vol. 110, no. 1059(1988). 2. Lambert, M.A. and Fletcher, L.S., “A Review of Thermal Contact Conductance of Junctions with Metallic Coatings and Films,” AIAA Paper 92-0709, AIAA 30th Aerospace Meeting, Reno, NV (1992). 3. Phelan, P.E., Niemann, R.C., and Nicol, T.H., “Thermal Contact Resistance for a Cu/G-10 Interface in a Cylindrical Geometry,” A\SME Proceedings of the 31st National Heat Transfer Converence, Vol. 5, no. 185 (1996). 4. Zhao, L., Phelan, P.E., Niemann, R.C. and Weber, B.R., “Thermal Resistance across a Copper/Kapton/Copper Interface at Cryogenic Temperatures,” submitted to the 1997 Cryogenic Engineering Conference, Portland, Oregon
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5. Rule, D.L., Smith, D.R. and Sparks, L.L., “Thermal Conductivity of a Polyimide Film between 4.2 and 300K, with and without Alumina Particles as Filler,” NISTIR 3948, NIST (1995), p. 40. 6. Ochterbeck, J.M., Peterson, G.P. and Fletcher, L.S., “Thermal Contact Conductance of Metallic Coated BiCaSrCuO Superconductor/Copper Interfaces at Cryogenic Temperature,” J. Heat Transfer, Vol. 114, no.21 (1992).
Cryocooler Subsystem Integration for the High Resolution Dynamics Limb Sounder (HIRDLS) Instrument D.J. Berry, D. Gutow, J. Richards, and R. Stack Ball Aerospace & Technologies Corp. Boulder, CO, USA 80306
ABSTRACT The High Resolution Dynamics Limb Sounder (HIRDLS) Cooler Subsystem provides 62 K cooling to the detector subsystem using a cryocooler symmetrically supported in a heat-rejecting radiator. The electronics, with software, is mounted on an independent panel of the instrument structure. Detail design of the cooler subsystem is underway, following the instrument system configuration trade studies and the instrument preliminary design review. The technical challenges associated with vibration cancellation are compounded by the necessity to simultaneously reject heat from the system. We discuss our system layout and how we provide active vibration cancellation while providing thermal system compatibility and isolation. We discuss the flexible vacuum enclosure and orthogonal bellows, which provide omnidirectional and low-interface forces during instrument integration and alignment. The enclosure also provides a vacuum environment for ground testing. We discuss the cooler-to-detector interface with the Ball S-link, which is highly conductive and compliant.1 We discuss the electronics reliability and strategic redundancy implemented in its circuitry to assure the highest reliability versus cost and weight. We also discuss the software implementation. INTRODUCTION Ball is integrating its flightworthy Stirling-cycle mechanical cryocooler into HIRDLS, an instrument slated to fly on the Earth Observing System Chemistry Platform. The HIRDLS cooler subsystem, developed under subcontract for the NASA Goddard Space Flight Center Earth Observing System Program, consists of a sophisticated and highly reliable single-stage, fixed regenerator;2 verifiable, noncontacting, Stirling cryocooler3,4 and its drive electronics; interfacing hardware; cryocooler support bracketry; and the cryocooler radiator. The HIRDLS cooler is leveraged from the NASA 30 K Phase IV cooler,3 currently at Goddard Space Flight Center under life test. HIRDLS COOLER SUBSYSTEM DESCRIPTION The cooler lifts 770 mW from the detector, plus 120 mW from parasitics, onto the cold components of the displacer and cold link assembly for a total of 890 mW at 62 K. Total input power,
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Figure 1. Major components of the HIRDLS Cooler Subsystem (to the same scale).
Figure 2. Performance curves for the single-stage cryocooler breadboard test.
including the electronics, is 62 W, and the mass is 15 kg. The entire cooler subsystem mass, including the radiator and cold tip interfacing hardware, is 26.3 kg. The major components are shown in Figure 1. The single-stage breadboard performance curves for the HIRDLS cooler are shown in Figure 2.
Symmetric Mount
Active vibration cancellation1,2,4 provides attenuation of the axial modes for the compressor and displacer.1 To optimize the control algorithms’ performance, the compressor and displacer axial centerlines coincide with the radiator centerline plane. The mounting brackets create a symmetric structural load path from the compressor and displacer, as well as rejecting cooler waste heat.
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Radiator Instrument Attachment The radiator-mounted cooler interfaces to the instrument structure with five titanium mounts as shown in Figure 3. The mounts are four blade-style flexures oriented toward one tube-style mount. This orientation mitigates in-plane loads resulting from thermal expansion differences between the instrument structure and the weight-relieved aluminum radiator. The geometry of the blades controls the instrument-to-radiator heat leak path, the fundamental dynamic mode of the
flexure-mounted system on the radiator, and the loads induced to the structure. The current radiator design shown in Figure 4 is the result of an extensive trade study of the radiator-mounting configuration, which trades off cost, mass, and vibration levels transmitted from the cooler into the instrument.
INDUCED VIBRATION CONSIDERATIONS Two sources of vibration impact the system design and must be considered. The first source is the uncontrolled lateral vibration of the cooler being transmitted to the structure. The second source is the displacer cold finger pulsations pushing on the detector. The pulses are the result of the Stirling-cycle pressure wave. Using a NASTRAN model of the HIRDLS instrument structure from Matra Marconi Space in Portsmouth, England, we constructed a cooler vibration system impact model for evaluation. Cooler vibration data was collected on a 6-axis dynamometer.2,5 The spectral response of the cooler was added to the NASTRAN model and the force response data was determined. Based on the results, we determined that when the radiator is hard mounted on the flexures, the force response at the instrument structural interface is greater than when the radiator is soft mounted on elastomeric isolators. The criticality of the vibration is not at the instrument interface, but rather
at the jitter sensitive detector. Figure 5 illustrates relative forces at the detector and at the instrument mount.
Figure 3. Four blade-style flexures oriented to one rigid tubular mount minimize thermal and
structural loads into the instrument structure.
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Figure 4. A simple, weight-relieved panel provides margin on the cooler thermal interface and provides the necessary structural mount. Instrument side shown; space-view side opposite.
Figure 5. Hard mounting the radiator minimizes vibration forces at the detector vacuum flange.
When soft mounting was evaluated, the vibration that is not transmitted across the radiator structural mounts shows up at the detector interface. Therefore, the detector subsystem interface experiences higher vibration levels when the radiator is soft mounted. To minimize the vibration at the detector subsystem, the program decided to use the hard mount. The residual Stirling-cycle pressure wave vibration is easily characterized and mitigated with the Ball S-link,1 which interfaces the cold finger to the detector subsystem. Figure 6 shows the anticipated force levels acting on the detector subsystem cryogenic interface. The S-link interfaces directly to the detector subsystem with the S-link opposite end attached to the cold rod. The other end of the cold rod connects directly to the displacer cold tip.
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Figure 6. Established models verified by test predict milli-newton forces at the detector cryogenic interface.
LAUNCH LOAD CONSIDERATIONS
Launch loads contributing to the displacer cold finger torsion and bending stresses are mitigated by noncontacting vespel snubbers as shown in Figure 7. The cold finger bending loads snubber is held in place by a titanium extension tube mounted to the cryo-vac housing. Titanium matches the thermal expansion of the cold finger and provides the highest stiffness per unit mass. The snubber also provides a support for small radiation shields around the cold finger. Parasitic heat loads from the 300 K end of the cold finger to the 60 K cold tip are minimized by several
radiation shields. The cold finger requires torsional snubbing of the cold rod to prevent excessive torsional loads at the cold finger base. The cold rod snubber is a polyimide wagon wheel device that provides a small area and a long heat leak path. It also provides the structural restraint between the cold rod and the bellows extension tube. We chose the polyimide because it eliminates the conductive heat path through the snubber as the vespel shrinks away from the bellows extension tube during cool down, creating a thermal disconnect. CRYOGENIC INTERFACES The flexible vacuum enclosure provides a vacuum shroud to the cryogenic interface for ground test as shown in Figure 7 while allowing compliance during the cooler subsystem-to-instrument integration. The flexible vacuum enclosure is made of a stainless steel two-piece junction box weldment connecting two orthogonal bellows. The two-piece junction box allows the cooler subsystem radiator-cooler assembly to be integrated, with the final interface of the cold rod to the displacer cold finger exposed, until the junction box is mated. The two halves join along an O-ring flange while creating an interlocking hinge opposite of two clamping fasteners. The arrangement holds the O-ring in contact with the flanges as the vacuum is applied inside the enclosure and through the pump-out port. The O-ring is then compressed and seals the vacuum. A similar seal concept is used at the flexible vacuum enclosure-to-detector subsystem vacuum interface, except two Orings are used, with one on each side of the bellows extension tube flange. This floats the flange
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Figure 7. A lightly restrained orthogonal bellows provides a low-spring rate, omni-directional,
vacuum shroud during telescope focus adjustment.
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between the O-rings when vacuum is not applied, creating a damped vibration load path to the detector vacuum flange on orbit. When the vacuum is applied, during ground test, the elastomer is compressed, creating a vacuum-tight seal. The pump-out port is open during launch and vents out on assent. The flexible vacuum enclosure is required only for ground operations, but it cannot be removed before flight due to access limitations. Launch vibration testing is performed prior to delivery. During test, the flexible vacuum enclosure is restrained to the radiator. When a vacuum is applied during cooler operation ground tests, the undesirable resultant forces are canceled from the sensitive detector by a counterforce device shown in Figure 8. The counterforce device is required to reduce one axis of the orthogonal bellows pulling force on the detector when a vacuum is applied. The other axis is loosely fixed to the radiator. The resultant force is canceled at the robust bellows extension tube counterforce reaction flange. All the bellows are the same, allowing some flexibility for detector displacement during instrument focus adjustment. ELECTRONICS
The cooler control unit (CCU) shown in Figure 1 houses the cooler operating drive and control electronics, data collection, telemetry, and firmware. Reliability of electronics is always a
major concern on spaceflight instruments. The CCU circuitry derives its heritage from the NASA 30 K Phase IV cooler.3 An extensive quality program has been invoked on the HIRDLS cooler control unit to meet flight standards. Flight status has been achieved by enforcing a rigid parts program including both Ball and NASA parts control board. Further quality assurance has been achieved by performing single-event upset and radiation susceptibility analyses. An affordable level of redundancy has been incorporated into the design, bringing the cooler subsystem reliability to 0.92. The CCU has a radiation hardness of 30 krad and houses redundant motor drive converters, compressor pulse-width modulation motor drives,3 and displacer linear motor drive electronics. Redundant launch lock relays are employed to short the compressor and displacer motor windings, creating a magnetic brake on the moving masses during launch. The launch locks have been tested during 14.1 g root mean squared random vibration and 15-g sine wave with no contact of the mechanical stops.3,4
Figure 8. Removed for flight, the counterforce device nulls out bellows-induced forces prior to reaching the detector vacuum interface.
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The aluminum chassis rejects heat from the circuitry by thermally conducting to the walls and then radiating to space. The thermal analysis shows a large junction temperature margin with worst-case power and a chassis temperature of 35 °C. Orientation of the chassis on the HIRDLS instrument is not as critical as the location. The chassis is a four-segment assembly. Each segment contains a specific circuit, and all are interconnected with a jumper cable housed under an electromagnetic interference tight cover. The flanged base contains the power section and in-rush current-limiting circuitry. The other segments contain the analog, digital, and pulse-width modulated boards. The cooler control unit connects to the cooler radiator assembly with two bulkhead connectors on the top of the radiator and two mating connectors on the cooler control unit. The cooler control unit interfaces to the instrument power supply to receive spacecraft noisy bus power. Engineering data is transferred through the instrument processor unit on an RS-422 serial protocol interface.3 SOFTWARE The flight software calculates and outputs the proper signals to drive the compressor and displacer motors. These drive signals control the amplitude, frequency, and phase of the oscillation of each motor. The oscillation frequency of the four motors can be varied, but is the same for all. The phase between the compressor motors and the displacer motors is user selectable. The amplitude of the oscillation is also controlled by the software. Different modes of operation cause the amplitude to be controlled in different ways.
There are three operational substates of the cooler subsystem: manual, thermostatic and warm-up. In the manual substate, the motor strokes are controlled to a commanded amplitude. In the thermostatic substate, the stroke is varied to maintain a constant temperature at the cold tip. The warm-up substate is similar to the manual substate, except that the phase angle between the compressor and displacer is set to approximately the conjugate from the normal (cool down) phase angle. This reverses the Stirling cycle, and the cold tip warms up. The software also performs vibration cancellation. This may be done concurrently with either the manual or thermostatic substates. When performing vibration cancellation, one side of the compressor and the counterbalance motor drive waveforms are modified such that the vibration caused by each motors’ reciprocating motion is canceled. An electronic circuit detects if any of the motors go beyond their normal range (100% amplitude). If this occurs, the electronic motor drives are automatically shut down and a status bit is set. Included in the electronics is a watchdog circuit, which times out after approximately 50 seconds. The software periodically resets the watchdog circuit to prevent it from timing out. Should a timeout occur, the watchdog circuit shuts down the motor drives and causes a reset of the microprocessor. The cooler subsystem is under control of a host computer; this computer commands which state and substate of operation the cooler should be in, as well as the various operating parameters. This communication takes place through the command and telemetry port. Commands come to the cooler from the host computer through this interface, and telemetry messages are sent out. The HIRDLS cooler subsystem can also be operated independently from the instrument host
computer using ground support equipment. The HIRDLS firmware and ground support equipment software operates with Lab View from the stand-alone cooler test rack. CONCLUSION The cooler subsystem is symmetrically mounted within the radiator to maximize the vibration control algorithm performance. Vibrations induced to the detector have been mitigated by hard mounting the radiator, rubber isolating the detector vacuum interface flange, and using a Ball S-link at the detector cryogenic interface. Launch load stresses are eliminated by means of snubbers protecting the displacer cold finger from bending and torsion. The flight electronics and software designs are documented and in test.
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The cooler subsystem will be subjected to flight qualification testing including electromagnetic interference and electromagnetic compatibility, random and sine vibration, and thermal vacuum thermal balance. ACKNOWLEDGMENTS
This work has been performed under subcontract to Lockheed Martin Missiles & Space in Palo Alto, California. We wish to thank Willy Gully and Tom Yarnell at Ball Aerospace for their outstanding technical contributions to the HIRDLS cooler and system integration design advances.
REFERENCES 1. Arentz, R.F. et al., “A Verified, Broadly Applicable Design for Interfacing Cryoccolers to Sensors 2. 3. 4. 5.
or Devices,” Advances in Cryogenic Engineering, Vol. 41, Plenum Press, New York (1995), pp. 1113-1120. Berry, D., H. Carrington, and W.J. Gully, “Two-Stage Cryogenic Refrigerator for High Reliability Applications,” in Advances in Cryogenic Engineering, Vol. 41, Plenum Press, New York (1995), pp. 1585-1592. Berry, D., et al., “System Test Performance for the Ball Two-Stage Stirling-Cycle Cryocooler,” Proceedings of the 9th International Cryocooler Conference, Plenum Press, New York (1996), pp. 6977. Carrington, H., W.J. Gully, M. Hubbard, and C. Vainer, “Multi-stage Cryocooler for Space Applications,” Cryocoolers 8, Plenum Press, New York (1995), pp. 93-102. Carrington, H., et al., “Functional and Life Test Data for a Two-Stage Stirling Cycle Mechanical
Cryocooler for Space Applications,” in Proceedings of STAIF-98, Am. Inst. Phys., CP420 (1998), pp. 199-204.
EMI Performance of the AIRS Cooler and Electronics D.L. Johnson, S.A. Collins, and R.G. Ross, Jr.
Jet Propulsion Laboratory California Institute of Technology Pasadena, California 91109
ABSTRACT
The TRW pulse tube cryocooler for JPL's Atmospheric Infrared Sounder (AIRS) instrument is required to meet stringent requirements for radiated electric and magnetic fields, conducted emissions on the input power bus, and electromagnetic susceptibility. To meet the radiated magnetic field requirements, special mu-metal shields were designed, fabricated, and fined to the cooler following an extensive period of magnetic testing with mock-up cooler hardware. Excessive magnetic fields is a generic issue with linear-motor cryocoolers, as is excessive levels of input ripple current. Solving the ripple current issue required the addition of a dedicated ripple filter as part of the spacecraft power system. As one of the first cryocoolers with flight electronics available for testing, the AIRS cooler offered an important opportunity to measure and understand these important issues. This paper describes the development of the magnetic shields to bring the AC magnetic fields of the AIRS cooler within the requirements of MIL-STD-461C, includes before and after data on the achieved field levels, and presents extensive data on the suite of EMI characteristics of the AIRS cooler with its flight electronics. Because of its importance, levels of input ripple current are presented as a function of compressor input power to allow future applications to judge the ripple-current compatibility with specific power system capabilities and to serve as data for scoping the design of ripple suppression hardware. INTRODUCTION
Instrument Overview The objective of the Atmospheric Infrared Sounder (AIRS) instrument is to make precision measurements of atmospheric air temperature over the surface of the Earth as a function of elevation. It is scheduled to be flown on NASA's Earth Observing System PM platform in the year 2000. The technical foundation of the AIRS instrument is a cryogenically cooled infrared spectrometer that uses a pair of TRW 55K pulse tube cryocoolers1-4 to cool the HgCdTe focal plane to 58 K. The spectrometer operates over a wavelength range from visible through and places demanding requirements on the EMI performance of the cryocooler and its electronics. Cryocoolers 10, edited by R. G. Ross, Jr. Kluwer Academic/Plenum Publishers, 1999
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Figure 1. Overall AIRS instrument showing proximity of instrument electronics to the cryocoolers.
Figure 1 illustrates the overall instrument and highlights the key assemblies and the close proximity between the cryocoolers and the sensitive instrument electronics. Physically, the instrument is approximately in size, with a mass of 150 kg, and an input power of 220 watts. Early in the design of the AIRS instrument, two key decisions of design philosophy were established that served as fundamental ground rules for the approach to meeting the cryocooler EMI design requirements. These included: 1) cooler drive fixed at 44.625 Hz, synchronized to the instrument electronics—to minimize pickup of asynchronous EMI noise (or vibration) from the cryocooler, and 2) cooler drive electronics fully isolated (dc-dc) from input power bus Because the required cryocooler EMI performance was more stringent than the measured EMI levels any existing cryocooler at the beginning of the AIRS development effort, the AIRS Project established a collaborative in-house/contractor teaming approach to achieve the necessary cryocooler EMI performance. This integrated-product-team approach involved the cryocooler developer (TRW Space & Technology Division of Redondo Beach, CA), who took primary responsibility for the cryocooler's EMI design, the AIRS instrument developer (Lockheed Martin IR Imaging Systems of Lexington, MA), who took primary responsibility for the instrument compatibility, and JPL, who took primary responsibility for the AC magnetic shield and supplemental EMI filter development, EMI acceptance testing, and spacecraft interface negotiations. The remainder of the paper details the EMI requirements on the AIRS cryocoolers, summarizes the overall EMI design approach undertaken including development tests, and describes the final results of the extensive series of Qualification Acceptance tests conducted on the flight cryocooler (PFM) units. AIRS CRYOCOOLER RADIATED AND CONDUCTED EMISSIONS
The Electromagnetic Interference (EMI) design requirements on the AIRS cryocoolers are fundamentally interface requirements associated with assuring electrical operational compatibil-
ity between the AIRS cryocoolers and their electrical neighbors; these neighbors include the very sensitive electronics in close proximity to the coolers within the AIRS instrument itself (particularly the electronics associated with the focal plane readout), the electronics associated with neighboring instruments and the spacecraft itself, and the spacecraft power system that provides the 28 Vdc bus power to the cryocoolers. The EMI requirements associated with these interfaces fall into three areas, which, in the case of AIRS, are tailored from the popular MIL-STD-461C baseline:5 • Radiated magnetic field emissions (30 Hz to 50 kHz) • Radiated electric field emissions (14 kHz to 18 GHz) • Conducted AC currents on the 28 Vdc power bus (30 Hz to 50 MHz)
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Figure 2. Cryocooler system showing internal compressor operational elements.
Radiated Magnetic Field Emissions In interpreting the radiated magnetic field requirement with respect to the AIRS compressor it is useful to first understand the fundamental electromagnetic structure of the cryocooler. The AIRS compressor, shown schematically in Fig. 2, is a mechanically resonant system with two pistons acting into a common compression space. Each flexure-suspended piston assembly operates much like a loudspeaker, whereby it is driven via a moving coil in a permanent magnetic field. Piston motion and gas compression is generated by simultaneously applying an alternating current through the coils of the two piston assemblies at the drive frequency of 44.625 Hz. This frequency was chosen to optimize the overall performance of the AIRS instrument and cooler, and the compressor was then tuned to be near mechanical resonance at this frequency to maximize the drive motor efficiency. Compressor AC Magnetic Emissions. Two sets of AC magnetic field measurements are typically made to quantify cryocooler AC magnetic field emissions: 1) at a 7-cm distance, corresponding to the MIL-STD-461C RE01 test specification5, and 2) at a 1-m distance, corresponding to a MIL-STD-462 RE04 test method. The measurements are made using a standardized 37-turn loop antenna positioned the required distance from the outer surface of the compressor or electronics box. Figure 3 summarizes the RE01 performance of a number of representative Oxford-type space cryocoolers, not including AIRS, which use the same fundamental piston-drive approach as
Figure 3. AC magnetic field emissions measured for a variety of Oxfordtype space cryocoolers (versus MIL-STD-461C RE01 requirements).
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Figure 4. Magnetic shielding studies used this "magnetic mock-up" of the AIRS compressor with various
shield materials, thicknesses and configurations.
the AIRS cooler.6 The data are plotted in decibels above 1 pT; the breaks in the measured data are due to changes in the amplifier gain and spectrum analyzer bandwidth settings. Note that the radiated magnetic field emission levels for the fundamental drive frequency are typically above the MIL-STD-461C specification. After the first three or four harmonics, the levels rapidly drop and reach background ambient levels above 1 kHz. Radiated magnetic emissions observed above 10 kHz are typically emissions at the harmonics of the switching power supply drive frequency. Given the inability of most previous space coolers to meet the MIL-STD-461C RE01 requirement, the AIRS cooler was designed from the beginning to incorporate magnetic shielding. Originally, mu-metal shields were to be incorporated around each compressor motor internal to the compressor pressure housing. However, early measurements of the shields' effectiveness showed that the shields were saturating from being too close to the magnetic source and were providing shielding levels below 5 dB. A number of magnetic shielding studies were next run using various configurations of CO-NETIC AA7 and Moly Permalloy8 shields on a magnetic mockup of the AIRS cooler as shown in Fig. 4. These tests showed that field reductions of the order of 20 dB could be achieved with external shields, and still meet the tight volume and mass restrictions imposed by integration constraints within the AIRS instrument. The final shield design, shown in Fig. 5, utilizes 0.5-mm (0.020") thick shields hydroformed from Moly Permalloy
Figure 5. Flight magnetic shield and one of two mounting rings used to support the shield from the compressor pressure housing (left), and final flight cooler with four shields installed (right).
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Figure 6. AC magnetic field emissions measured for the AIRS mechanical cooler with and without the addition of the flight mu-metal shields (versus MIL-STD 461C RE01 requirements).
high-permeability magnetic sheet material and stood off from each compressor end cap using structural aluminum rings as shown in the left-hand figure. The right-hand photograph in Fig. 5 shows the flight cooler system with the four shields installed. Figure 6 highlights the measured AC magnetic field emissions from the completed AIRS cooler in contrast to measurements made before the shields were installed. With the shields, the cooler meets the MIL-STD-461C RE01 requirement on radiated magnetic field emissions. Compressor DC Magnetic Field Emissions. In addition to generating AC magnetic fields
associated with the AC coil currents, the compressor generates DC magnetic fields associated with the permanent magnets and iron pole pieces used to provide the magnetic circuit for the drive motors. The resultant DC magnetic dipole field falls off proportional to with increasing distance away from the cooler body. Figure 7 describes the DC magnetic field profile as measured for the AIRS compressor along the compressor centerline as a function of distance away from the compressor endcap (left figure), and along the length of the compressor at a 16-cm radial distance from the compressor centerline (right figure). Measurements were made using a Hall generator that was zeroed with the Earth’s magnetic field so that the Earth’s field contribution is not included in the measurements. Note that the left-hand graph of Fig. 7 presents data for the compressor both with and without the mu-metal shields shown in Fig. 5; it also shows the classic dependence of the magnetic field as a function of distance from the end of the compressor.
Figure 7. DC magnetic field emissions measured for the AIRS mechanical cooler.
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Figure 8. AC magnetic field emissions of the AIRS cryocooler electronics.
Electronics AC Magnetic Field Emissions. Although not a principal source of magnetic fields, the AIRS cooler electronics were also measured with respect to MIL-STD-462 RE01 and RE04 test methods. Figure 8 summarizes the RE01 performance of the electronics and highlights two magnetic field peaks associated with the 45 kHz PWM switching frequency. Radiated Electric Field Emissions
As noted above, the AIRS cooler drive electronics utilize pulse-width-modulated power converters (PWMs) to synthesize the compressor drive waveform with maximum efficiency and low harmonic distortion. Early in the program it was recognized that the high internal currents associated with the 45 kHz PWM switching frequency are a major source of electric field emissions and must be carefully contained in an EMI-proof enclosure. To this end, the AIRS cryocooler electronics were packaged in an all-aluminum enclosure, shown in Fig. 9, with extensive internal compartmentalization and filtering of inter-compartment penetrations. The radiated electric field emissions of the AIRS electronics were measured using MILSTD-462 RE02 narrowband and broadband electric field emission specifications. Measurements were conducted at a distance of 1 meter from the geometric center of the electronics. Several antennas were used to measure the emissions up to a frequency of 10 GHz.
Figure 9. AIRS electronics in heavily EMI-shielded enclosure.
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Figure 10. Electric field emissions of AIRS electronics.
Figure 10 shows the low radiated electric field emissions achieved with the AIRS electronics. Discontinuities in the data are changes in the antennas, amplifiers, and bandwidths used to cover the different frequency bands. The localized peak above the RE02 limit at 20 MHz is associated with the computer clock frequency and is due to inadequate shielding of a non-flight cable used for the testing. Conducted Emissions on the 28 V Power Bus
In addition to generating large magnetic fields as noted above, the sinusoidal currents drawn by the linear motors at their 44.625 Hz drive frequency result in a large input ripple current at twice the drive frequency; this corresponds to full wave rectification of the drive current. The magnitude of the ripple current is inversely related to the operating DC voltage, and proportional to the operating power. It is difficult to significantly filter this primary ripple current because of its large magnitude and low frequency. To help accommodate this large ripple current, the AIRS cryocoolers are supplied by a set of dedicated 28 Vdc "dirty bus" power circuits that allow an input ripple current as high as 200% p-p/average. In addition, the cooler power passes through the AIRS instrument where additional powerline filtering and inrush current suppression is accomplished. The 28 Volt power lines of the AIRS cooler were tested for ripple current emissions in both the narrowband and broadband frequency spectrums. Measurements were conducted on both the high-side (positive) and return (negative) lines using a current probe. A line impedance simulation network was inserted in the 28 Volt line to closely simulate the spacecraft bus power impedance; the impedance, which is a function of frequency, is 0.25 ohms for frequencies below 1 kHz. Figure 11 is the narrowband conducted emissions profile on the 28 Volt positive lead. The specification line is that of MIL-STD-461C CE01/03. The harmonics of the 44.625-Hz drive frequency are clearly observable, as are extensive harmonics of the 45 kHz pulse-width-modulated power converters. The return (negative) lead current emission profiles had nearly identical emission levels. The extensive over-spec peaks in the region from 90 kHz to 20 MHz are filtered out by the AIRS instrument power line filter, and are not allowed onto the spacecraft power bus. Input Ripple Current Test Results (Time Domain). The time-domain waveform of the input ripple current was also measured for a range of electronics input power levels. Figure 12 presents example input-current waveforms, while Fig. 13 plots the measured relationship between ripple-current level and cooler input power. Note that the input ripple exceeds 200% at the
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Figure 11. Conducted ripple-current emissions of AIRS electronics on 28 V power bus.
higher cooler power levels; this will be partially attenuated by the filtering accomplished at the instrument level. The high ripple current levels reflected onto the spacecraft bus by the AIRS cryocooler strongly support the decision to provide it with a separate “dirty” 28 V bus.
Figure 12. Comparison of the current waveforms on the 28 V power bus for the S/N 301 and 302 coolers, each with a system input power of 110 W.
Figure 13. Relationship between ripple-current level and cryocooler input power level.
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Power-On Inrush Current/Transient Voltage (Time Domain). The AIRS cryocooler flight electronics was tested for inrush current as well as transient voltage when the unit was switched from OFF to ON. A peak inrush current of approximately 9 amps was measured associated with charging of the internal circuitry; a voltage transient of -2 volts was recorded at the same time. It should be noted that the cooler has a slow start circuit upon powering of the compressor, so no further reflected voltage or current is observed at this point beyond the peak ripple current noted in Fig. 13. AIRS CRYOCOOLER RADIATED AND CONDUCTED SUSCEPTIBILITY
The AIRS cryocooler and its electronics must not only produce low levels of EMI to be compatible with its surroundings, but must also withstand similar levels of EMI from external sources such as other spacecraft instruments. EMI susceptibility is also important because, in addition to voltage and current ripple on the power bus, the cooler power-ups can produce inrush current spikes that can draw down the voltage available to the operating cooler over short periods of time. The cooler and electronics must be able to maintain normal operation without malfunctioning under allowable levels of input voltage ripple and turn-on voltage transients. The Electromagnetic Compatibility (EMC) tests were run on the AIRS cooler in three areas: Susceptibility to Radiated Electric Field Emissions (14 kHz to 18 GHz). The AIRS cryocooler was subjected to AC electric fields according to MIL-STD-461C RS03. During operation, both the compressor and electronics were subjected to electric field strengths of 2 V/m from 14 kHz to 2 GHz, and then to a field strength of 10 V/m from 2 GHz to 18 GHz. No anomalies were observed. Susceptibility to Radiated Magnetic Field Emissions (30 Hz to 200 kHz). The AC magnetic field susceptibility test involved exposure of the operating cooler to fields over a range of 30 Hz up to 200 kHz. The field strength level was 120 dBpT at the surface of the cryocooler. While being subjected to the AC magnetic fields, the cooler and electronics were constantly monitored for any anomalies. None were found. Susceptibility to Conducted Emissions on the 28 Vdc Power Bus (30 Hz to 400 M Hz). AC Voltage Ripple. The MIL-STD-461C CS01 requirement is volts peak-to-peak from 30 Hz to 2 kHz. Before this level could be applied to the power bus, the peak current had reached 5 amps. While the cooler operation was being monitored, no change in the performance occurred. At 2 kHz, the injected voltage ripple was reduced (per specification) as frequency increased to 50 kHz, where it was then constant at 3 V p-p up to 400 MHz. The performance of the cooler was not affected. Injected Voltage Transient. The requirement of method CS06 is that a transient voltage be injected on to the +28 Vdc line relative to chassis, on to the return line relative to chassis, and on to the +28 Vdc line relative to the return line. The transient is to have a 40 V peak amplitude, duration, and be repeated at 60 pulses per second for 5 minutes. The test was conducted and no anomalies were observed in the cryocooler operation or performance. SUMMARY AND CONCLUSIONS
The AIRS cryocooler EMI development and testing activity has been a highly collaborative effort involving cooler development at TRW, integration studies at LMIRIS, and magnetic shielding and acceptance testing activities at JPL. As one of the first cryocoolers to be required to meet stringent EMI constraints, the AIRS cooler offered an important opportunity to measure and understand the important EMI/EMC issues and parameters. Extensive development and test results have been presented including development of the magnetic shields to bring the AC magnetic fields of the AIRS cooler within the requirements of MIL-STD-461C, and data on the entire suite of EMI characteristics of the AIRS cooler with its flight electronics. Particular attention has been given to characterizing the levels of input ripple current to allow future applications to judge
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the ripple-current compatibility with specific power system capabilities and to serve as data for scoping the design of future ripple suppression hardware. ACKNOWLEDGMENT The work described in this paper was carried out by the Jet Propulsion Laboratory, California Institute of Technology, Lockheed Martin IR Imaging Systems, and TRW, Inc; it was sponsored by the NASA EOS AIRS Project through an agreement with the National Aeronautics and Space Administration. Special credit is due P. Narvaez, M. Gross, and B. Ruff of the JPL EMI/ EMC Laboratory, who performed the EMI/EMC measurements presented here. The cryocooler team at TRW designed and fabricated the AIRS cooler and electronics, except for the compressor magnetic shields, which was the work of S. Leland of JPL. R. Schindler of the AIRS Project Office was responsible for overseeing the spacecraft and instrument EMI/EMC requirements and the design of the cryocooler ripple/inrush EMI filter installed in the AIRS instrument. REFERENCES
1. Ross, R.G., Jr., Johnson, D.L., Collins, S.A., Green K. and Wickman, H. "AIRS PFM Pulse Tube Cooler System-level Performance," Cryocoolers 10, Plenum Publishing Corp., New York, 1999. 2. Ross, R.G., Jr. and Green K., "AIRS Cryocooler System Design and Development," Cryocoolers 9,
Plenum Publishing Corp., New York, 1997, pp. 885-894. 3. Chan, C.K., et al., "Performance of the AIRS Pulse Tube Engineering Model Cryocooler," Cryocoolers 9, Plenum Publishing Corp., New York, 1997 pp. 195-202. 4. Chan, C.K., et al., "AIRS Pulse Tube Cryocooler System," Cryocoolers 9, Plenum Publishing
Corp., New York, 1997, pp. 895-903. 5. Electromagnetic Emission and Susceptibility Requirements for the Control of Electromagnetic Inter-
ference, MIL-STD-461C, Department of Defense, Washington, DC (1986). 6. Johnson, D.L., et al., "Cryocooler Electromagnetic Compatibility," Cryocoolers 8, Plenum Publishing Corp., New York, 1995, pp. 209-220. 7. CONETIC AA alloy is manufactured by the Magnetic Shield Division, Perfection Mica Company,
Bensenville, IL. 8. Moly Permalloy is manufactured by Allegheny Ludlum Steel Corp., Pittsburgh, PA.
The Application and Integration of Mechanical Coolers R.M. Wilkinson1, S.R. Scull1, A.H. Orlowska2, T.W. Bradshaw2, and C.I. Jewell3 1
Matra Marconi Space, Filton, Bristol, England Rutherford Appleton Laboratory, Chilton, Oxfordshire, England 3 ESA ESTEC, Noordwijk, The Netherlands 2
ABSTRACT
Space qualified, single stage, Stirling cycle coolers are available as standard products in batch manufacture and their application to instrument cooling has become an accepted solution to cryogenic requirements. MMS, as a major space cryo-cooler supplier, has become involved in the issues associated
with the practical application and integration of these coolers on a number of programmes. Critical issues include the connection of the cooler cold finger to the instrument cold bus bar, thermal shielding, thermometry installation, heat rejection, structural mounting to survive launch and ensure good alignment to minimise out of balance forces and, in the case of use with a low vibration Cooler Drive Electronics, force transducers and thermal straps. Hardware developed for these applications include a cold finger flexible thermal link, developed by RAL, and laminated thermal straps, which conduct away waste heat from the cooler whilst by-passing the force transducers used in micro-vibration cancellation systems. This paper describes the cooler integration issues addressed and solved on several major European space programmes due for launch over the next 5 years. INTRODUCTION
The Matra Marconi Space (MMS) 50-80K cooler is a single stage, split Stirling cycle cooler.1 Its design is derived from the MMS 80K cooler, and the original Oxford University cooler. On completing the current batch of 10 coolers a total of 54 flight standard 50-80K and 80K coolers will have been manufactured. This paper reviews some of the integration issues of the 50-80K cooler in three spacecraft cooler sub-systems, which will be referred to by their respective spacecraft or instrument names; MIPAS, Helios II, and INTEGRAL. • MIPAS (Michelson Interferometer for Passive Atmospheric Sounding)2 is an instrument
which measures key trace gases in the atmosphere and will fly on the ENVISAT spacecraft in 2000. The cooler sub-system uses a pair of coolers which were supplied and integrated to the cooler structure by MMS. It has completed qualification testing and the flight model is being integrated into the instrument. Cryocoolers 10, edited by R. G. Ross, Jr. Kluwer Academic/Plenum Publishers, 1999
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• The Helios II cooler system uses a pair of coolers and an active vibration cancellation system using force transducers for feed back. Qualification of the coolers to meet the launch and space environments is complete. • INTEGRAL is an ESA space telescope which will be launched in 2001. The detector of the gamma-ray spectrometer has a cooler system consisting of two pairs of coolers. Vibration and thermal vacuum qualification testing has been completed. There are many issues to consider when integrating mechanical coolers for use on spacecraft. Of these, three issues have solutions which are fairly unique to mechanical coolers and will be looked at in more detail; vibration cancellation, cold finger integration and design of the cooler system structure. VIBRATION CANCELLATION
A single MMS 50-80 K cooler operating at full stroke generates large vibration forces at the compressor and displacer. These forces are generated along the axis of the cooler due to the sinusoidal displacement of masses inside the compressor and displacer. Because of this, the coolers are operated in pairs mounted 'back to back' (Fig. 1) and in phase so that the majority of
the forces are cancelled out. There are two types of vibration cancellation system using either conventional drive electronics or an active Low Vibration Drive Electronics (LVDE), both of which use the ‘back to back’ mechanism configuration.
Conventional drive electronics system
The conventional drive electronics system works by matching the vibration forces of a pair of coolers during ground testing. Whilst measuring the resultant vibration forces using a force plate, the stroke of one of the coolers is adjusted via the Cooler Drive Electronics (CDE). Once the resultant forces have been minimised, the CDE maintains this balance. The disadvantage of this system is that adjustments cannot be made in orbit and it can only be used in systems where a specification of 2 Nrms or greater, out of balance force, can be tolerated. The INTEGRAL
cooler sub-system is an example of a conventional drive system.
Figure 1. Two 50-80 K coolers in a 'back to back' configuration.
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Active system An active system measures the resultant out of balance force generated by the pair of mechanisms, so that it can be minimised whilst in orbit. One method of doing this is to mount the compressors on a common structure which is mounted in turn onto the spacecraft structure via force transducers which measure the forces being transmitted to the spacecraft structure. A LVDE system adjusts the drive current to the mechanism in anti-phase to the measured force in order to minimise the exported vibration. Force transducers, however, form a poor thermal path for the necessary conduction of the heat dissipated by the compressors to the spacecraft structure. To overcome this, thermal straps are fitted between the compressor structure and the spacecraft structure. The straps must have a high conductivity but a low stiffness to ensure that most of the vibration forces are transmitted through the force transducers and not the thermal straps. A laminated thermal strap has been developed consisting of thin sheets of copper with a 90° bend (Fig. 2). The laminates are brazed together at the two ends but left separate at the bend to increase the flexibility. Straps are mounted onto the structure in close proximity to each of the compressor flanges to shorten the conductive path and to spread the thermal heat flow through the support structure. If optimum performance is to be achieved, great care must be taken to minimise the stiffness and to match the ratio of the stiffness of the straps to that of the primary structure. Failure to do this will result in unbalanced forces, not measured by the force transducers, being transmitted to the structure. The Helios II cooler system uses this method and has achieved system level axial
vibration in the order of 0.2 N peak in the harmonics. Figure 3 shows a schematic of this system. Accelerometers may be used in place of the force transducers to detect the micro-vibrations and such a system has been developed by Rutherford Appleton Laboratory (RAL) for use on MIPAS. A disadvantage of such a system is that structures designed to survive launch tend to be
rather stiff and therefore deflections are very small and difficult to measure. The structure therefore needs to be optimised in order to ensure that these levels are detected and in general it is our experience that such systems are less sensitive than those using force transducers. An advantage of this method is that the coolers structure can be hard mounted to the instrument structure without the need for thermal straps. COLD FINGER INTEGRATION
A conductive link is attached to the cold finger to provide a thermal path between the cooler and the instrument detector. It is important that the temperature gradient between the cold finger and detector is kept to a minimum so that the cooler can operate at the highest temperature
possible. However, the cold finger must be protected from mechanical loads applied by the conductive link which could cause damage and limit the life of the cooler.
Figure 2. Thermal strap.
Figure 3. Force transducer vibration cancellation.
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The cold finger is a thin-walled titanium tube which encloses the regenerator. During operation of the cooler the regenerator oscillates along the axis of the finger with a nominal
clearance of a few microns. In order to prevent the tube deflecting sufficiently to cause a rubbing contact with the regenerator during operation, the lateral mechanical load applied to the 50-80K cooler cold finger must not exceed 0.4 N. Loads that can be applied to the cold finger during launch are much higher and are limited
by the strength of the finger. A launch support tube is fitted to all 50-80K cooler cold fingers to limit deflections and therefore the loads that are imparted but since the launch support tube does not contact the cold finger when operating it does not form a parasitic conduction path. Of the launch and operation requirements the most difficult to achieve is the 0.4 N lateral force limit during operation and by achieving this the launch requirement is automatically achieved within normal launch vibration levels. In operation, the mechanical loads applied to a cold finger are the combined effect of the conductive link mass supported by the cold finger (when operated in a 1g environment) and the elastic forces due to deflection of the conductive link. Deflection is caused by: misalignment during integration of the conductive link onto the cold finger; thermal contraction of the conductive link, the detector support structure, and the cold finger; deflection of the detector support structure relative to the cooler support structure due to gravitational loads. As a result of these deflections the conductive link must be both flexible and have a high conductivity. In addition, the mass of the conductive link which is supported by the cold finger must be kept to a minimum in order not to impact ground testing. A conductive link which meets these requirements, and has been qualified on the MIPAS and INTEGRAL systems, is the Thermal Link Assembly (TLA) developed by RAL. The TLA consists of a coiled copper braid brazed onto two copper blocks which form the interface with the cooler and the detector bus bar. A process of brazing the TLAs has been developed which prevents the braze wicking along the braid so that maximum flexibility is maintained. Figure 4 shows TLAs installed on two of the four cold fingers of the INTEGRAL system.
Figure 4. TLAs shown with thermal shield removed.
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The TLAs in both the MIPAS and INTEGRAL systems are enclosed by a thermal shield
which is thermally anchored to the detector bus-bar. This reduces radiation parasitics which would otherwise be caused by the high emissivity of the TLA braid material. A summary of the TLA conductivity and stiffness properties is given in Table 1. The stiffness in the X axis is the greatest and for this reason the TLA is orientated so that the X axis is parallel to the axis of the cold finger, which can support high loads without causing lateral deflections of the finger. The total allowable deflection of the TLA is calculated by subtracting the weight of the TLA, supported by the cold finger, from the allowable load and dividing this by the TLA stiffness. COOLER STRUCTURE DESIGN
The cooler structure performs several functions. It supports the cooler enabling survival of the launch environment including achieving a sufficiently high structural frequency (approximately 120 Hz). It also aligns the coolers for vibration cancellation, provides a conduction path to dissipate heat from the coolers to the spacecraft radiator and assists in conducting heat across the face of the radiator. The maximum power dissipation of a MMS 50-80K cooler is 50 W, of which approximately 40 W is dissipated at the compressor and the remainder at the displacer heat rejection interfaces. Thermal gaskets are fitted between the structure and cooler mounting flanges to improve joint conductivity. A thermal housing is fitted to the compressor cylinder head to provide a second conduction path in addition to the compressor mounting flange. Where possible the structure can include a vacuum enclosure around the cold fingers for use during ground testing of the coolers. This avoids the need to perform all tests on the cooler subsystem and instrument inside a vacuum chamber.
An example which includes all these features is the INTEGRAL structure (Figs. 5 and 6). To maximise stiffness and conductivity the structure was machined as a single component from a billet of aluminium alloy.
Figure 5. INTEGRAL structure.
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Figure 6. INTEGRAL cooler sub-system.
SUMMARY
Three design issues which are unique to the integration of mechanical coolers, have been described along with solutions which have been developed on three spacecraft systems. The same integration issues are also relevant to use of pulse tube coolers but with a relaxation of the cold finger lateral load limit. REFERENCES
1. Jones, B.G., et al. “The Batch Manufacture of Stirling-Cycle Coolers for Space Applications Including Test, Qualification, and Integration Issues”, Cryocoolers 9, Plenum Press, New York, 1997
2. Endeman, M. et al. “Michelson Interferometer for Passive Atmospheric Sounding - MIPAS” SPIE Proceedings Vol.1934, 13-15 April 1993, Orlando
Cooling System for Space Application GuoLin Ji and Yinong Wu Shanghai Institute of Technical Physics Chinese Academy of Sciences Shanghai 200083, P.R. China ABSTRACT
In this paper a miniature single-stage Stirling cryocooler for cooling a MCT infrared detector is introduced. The concept of the cryocooler is based on the Oxford-type Stirling refrigerator featuring special flexure suspension springs, linear drive motors, and clearance seals without wear for long-term operation. Cooling tests have demonstrated a no-load temperature of 43 K and a cooldown time of 20 minutes for a compressor power consumption of about 30 W. To reduce mechanical vibration disturbances, a head-to-head arrangement of the two refrigerators was adopted in the system configuration; the effect of electromagnetic interference (EMI) of the refrigerators on the detector was also considered. The cooling power is transmitted from the cold heads to the cold platform, on which the detector is seated, through a flexure cold link. The overall system can provide 300 mW of cooling power at the detector at 80 K. Power consumption is less than 80 W (including controller), the induced vibration force is less than 0.2 N, and the EMI noise is greatly reduced.
INTRODUCTION This paper describes a spaceborne Stirling cycle cryocooler. It is an Oxford-style Stirling cooler developed to cool a MCT infrared detector. A cooling capacity of 500 mW at 80 K is required to cover the detector heat dissipation and passive heat load from the surroundings. To minimize the self-induced vibration of the Stirling cooler, the cryocooling system consists of twin opposed Stirling coolers, a cryogenic bus bar joining the MCT detector to each cold tip of the coolers, and an electronic controller. The temperature difference between the MCT detector and the cold finger is designed to be as small as possible to decrease loss of cooling capacity. Because the EMI generated by the cryocoolers is also reduced, the system can meet the operational requirements of the MCT detector.
Description of the cooler The cryocooler for space application is a single stage split Stirling cycle cooler. The cooler consists of a compressor, a displacer, and a connecting tube. As shown in Figure 1, the special flexure diaphragm suspension spring, linear driving motor, LVDT positioning sensor, and clearance seal are adopted in both the compressor and displacer in order to achieve long life and high reliability for space application. The shafts of the compressor and displacer are supported by two sets of flexure diaphragm springs, respectively. This kind of spring is rigid radially, allowing Cryocoolers 10, edited by R. G. Ross, Jr. Kluwer Academic/Plenum Publishers, 1999
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Figure 1. Stirling cryocooler system.
only translational movement along its axis, and is used to prevent contact across the clearance seal.
Annealed Be-Cu alloy was selected for the diaphragm suspension springs. The diaphragm springs for the compressor and displacer are designed to have an appropriate natural frequency and small displacement in the lateral direction during operation. The piston is driven sinusoidally with a linear motor while maintaining the close-tolerance clearance seal. The electronic controller
can maintain highly reliable close-loop control of the cooler operation with measured and generated position feedback from the LVDT position sensors. The aluminum body is shaped for excellent heat dissipation. The materials in the cryocooler are selected for minimizing working gas contamination. The swept volume of the compressor is The stroke of the compressor and displacer is 8 mm and 2~4 mm, respectively. The total mass of the Stirling cooler is 4.2 kg: 3.1 kg for the compressor and 1.1 kg for the displacer. The compressor has a 120 mm diameter and 180 mm length; the displacer has a 90 mm diameter and 240 mm length (including the vacuum chamber); and the linked pipe has a 3 mm diameter and 300 mm length. Cryocooling system
1. System description. The cryocooling system consists of twin opposed Stirling coolers, a cooling box, and an electronic controller. As shown in Figure 1, the cooling box consists of a cryocooling bus bar, vacuum chamber and supporting element, etc. The twin compressors and displacers are mounted opposite each other for vibration control. 2. Vibration balance. The major concerns of the cryocooler application to space infrared remote sensing are the levels of vibration, heat transfer, EMI, etc. The inertial force generated by the linear motor and piston of the Stirling cooler is 10 N ~ 100 N. In order to minimize vibration of the MCT detector, two compressors and displacers, controlled by closed-loop electronics, are arranged in a head-to-head system configuration. According to the test results, a majority of the forces have been balanced, and because of the flexible cryogenic bus bar between the cryocooler and detector and the rigid support between the detector and cooling box, the vibration of the detector has been significantly diminished.
3. EMI. The piston assemblies of both the compressor and displacer are driven by moving coils in a permanent magnetic field. Mechanical motion is generated by applying an alternating current through the coils at the driven frequency, about 40 Hz, which was selected to optimize the thermodynamic performance of the Stirling cooler. However, the sinusoidal current through the coils generates the EMI. As test results show, the EMI is so large that the noise volt level of the system
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is about 1 V. In our system the EMI has been significantly reduced by using proper filters and shields, and the noise volt level has been reduced to ~20 mV, which has satisfied the requirements of the detector. 4. Heat transfer. The waste heat is discharged via conduction in the cryocooling system. The cooling power between the detector and coldhead is transferred through the cryocooling bus bar, which is fabricated using a copper pigtail. However, the temperature difference along the cryocooling bus bar can reduce the thermal efficiency of the cryocooling system. On the other hand, there is transfer efficiency in the cooling capacity from the cryocooler to the detector through the copper pigtail. While the cooling capacity is directly proportional to the cryocooling temperature of the cold finger, the transfer efficiency of the cryocooling bus bar can be defined as:
where:
= = = =
transfer efficiency of cooling capacity detector temperature cooling finger temperature minimal temperature of the cooler
With a temperature difference across the cryocooling bus bar of 2.5 K, the transfer efficiency equals 89%
Performance Test Results Typical performance of the Stirling cooler and cryocooling system is as follows:
1. Stirling cooler Cooling capacity (normal): Temperature on the mounting surface: Cooldown time (to 80 K): Power consumption (compressor): (displacer): Operation frequency: Filling pressure (He): Weight: 2. Cryocooling system Cooldown time (80 K): Temperature difference of cryocooling bus bar: Noise volt of MCT detector: Residual vibration:
Weight:
500 mW at 80 K 316 K 15 min 22 W 0.6 W 40 Hz 1.3 MPa 4.2 kg
~1.5 hrs 2.5 K 20 mV <0.2 N 14 kg
Figure 2 shows the effect of drive frequency on the cooling performance. Figure 3 represents the relationship between the frequency and the power consumption per cycle. These two figures clearly show that although goes down as the frequency goes up, there exists an optimum frequency. When the cooler operates at this frequency, which is very close to the resonance frequency of the mechanism, the minimum power consumption is attained for every cycle. The effect of phase angle on was measured for different compressor strokes, and the result obtained is shown in Figure 4. Evidently, the optimum phase angle is located between 70 degrees and 80 degrees. Figure 5 is the cooldown line of the cryocooler system (40 Hz, Compressor Stroke: 8 mm).
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Figure 2. Minimum cold tip temperature versus frequency.
Figure 4. Minimum cold tip temperature versus phase angle.
Figure 3. Power consumption per cycle versus frequency.
Figure 5. Cooldown characteristic of the cryocooling system.
CONCLUSIONS
1. The balanced arrangement and the flexible cryocooling bus bar significantly reduce the vibration. 2. By using proper filters and shields, the EMI has been reduced to levels that satisfy the requirements of the system. 3. The prototype cryocooling system has been successfully developed for space application.
REFERENCES
1.
Y. Wu and G. Du, “Development of a Miniature Cryocooler for Detector Cooling”, Proc. of 5th JSJS, Osaka, Japan (1997).
2.
G. Davey, “Review of Oxford Cryocooler”, Adv. Cryo. Eng., Vol. 35 (1990), pp. 1423-1430.
Drive and Control System for a Stirling Cryocooler Wang Biao, Guolin Ji, and Yinong Wu Shanghai Institute of Technical Physics Chinese Academy of Sciences Shanghai 200083, P.R. China
ABSTRACT A drive and control system for a Stirling cryocooler has been developed. By accurately controlling the stroke, frequency, DC bias, and phase between compressor and displacer, the cooler can reach its lowest temperature of 43 K with no heat load. In order to minimize the induced vibration, two “Oxford type” single stage split Stirling cryocoolers are positioned faceto-face and driven 180 degrees out of phase. An induced vibration level of less then 0.2 N is achieved. INTRODUCTION Because of the fast development of space infrared remote sensing technology, it is getting increasingly urgent to manufacture miniaturized, long-life, highly reliable mechanical coolers for space use. The “Oxford type” split Stirling cycle cooler ensures the long-life operation with stable performance by means of its unique structural characteristics. It is different from other refrigerators; the “Oxford type” split Stirling cooler includes a complex drive and control electronic system. In order to cool the MCT detectors at very low induced vibration level, two coolers are positioned face-to-face and driven 180 degrees out of phase. SYSTEM DESCRIPTION The reliable, lightweight, compact, high efficiency electronic controller is installed in a small box. It contains a moving mass position signal processor for the compressors and displacers, power supplies, telemetry interface, vibration cancellation circuitry, and logic. The servo control system, which uses a single-chip microcontroller (8031) as its main controller, is shown in Figure 1. The control system obtains the four LVDT sensor signals using A/D converters to determine the operating parameters of the moving masses including dc bias, frequency, stroke, and phase between compressor and displacer. It then iterates the feed-forward outputs via D/A converters until the integral of the error signal is zero. By accurately adjusting these parameters, the cooler can reach its lowest temperature with minimum power consumption. As shown in Figure 2, two “Oxford type” single stage split Stirling coolers are arranged faceto-face and driven 180 degrees out of phase to minimize the induced vibration. The hardware of the electronic control system consists of four boards: sensor board, power supply board, digital board, and special board for space applications. The first three parts set up the Cryocoolers 10, edited by R. G. Ross, Jr. Kluwer Academic/Plenum Publishers, 1999
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Figure 1. The servo control system for balanced Stirling cooler.
Figure 2. Balanced Stirling cryocooler.
framework of the whole control system. The last one solves many problems for space applications, such as remote test, remote control, and isolation.
The core software system uses a structured programming method and is divided into several blocks: start-up block, main control block, gas pressure force process block, self-test block, telemetry block, etc. The main functions of these blocks are described below. 1. Start-up block: Primes the register; adjusts the bias and stroke of moving mass to the predisposed values step by step. 2. Main control block: Determines the operating parameters of the moving masses according to the feedback position signals from the LVDT sensors; it then changes the feedforward outputs until the integral of the error signals is zero. 3. Pressure force process block: With the clues of states of pneumatic force, the program has better servo control over the Stirling cooler. 4. Self-test block: If one of the two coolers is inoperable it will be turned off automatically, while the other is driven continuously. If both coolers are inoperable, they will be cut off together. 5. Telemetry block: Sends the telemetry signals to the monitor.
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PERFORMANCE TESTING AND RESULTS
A prototype Stirling cooler driven by the electronic controller has been developed. When the operating parameters, which include stroke, phase, frequency, and so on were adjusted experimentally, it was shown that these parameters do have remarkable effects on the cooler. Although the lowest temperature goes down as the frequency goes up, there exists an optimum frequency. When the cooler operates at this frequency, which is very close to the resonant frequency of the mechanism, it takes the minimum power consumption for every cycle. Increasing the piston stroke of the compressor generates more cooling capacity, and the optimum stroke of the displacer is related to the stroke of the compressor. For this cooler, the optimum stroke of the displacer is about 4 mm when the stroke of the compressor is 8 mm. The optimum phase angle is between 70 and 80 degrees. The top level performance of the cooler is shown below. ITEM
BASELINE PERFORMANCE
Cooling power Power consumption Lowest temperature with no load System mass: Cooler Electronic controller Induced vibration Control precision for Stroke Control precision for phase
1.0 W at 80 K 44 W 43 K
Control precision for frequency
< 1 Hz
8.4 kg 6.2 kg (aluminum) < 0.2 N < 0.2 mm < 2 deg
CONCLUSION
A prototype electronic controller has been successfully developed for a balanced Stirling cryocooler. By precise control of the moving masses, the cooler can reach its lowest temperature with no heat load and offers 1000 mW of maximum cooling power at 80 K. The induced vibration is less then 0.2 N. When one of the two coolers is inoperable, the system can be turned to the single working cooler automatically. REFERENCES 1. CA. Lew, "Long-life Stirling Cycle Coolers for Application in 60~110 K Range", Proc. of 3rd Symp. on Space Thermal Control and Life Support system, Noordwijk, The Netherlands, Oct. 1988. 2. B. Hocking, "The Development of a Range of Low Vibration Cryocoolers for Space Applications," Proc. 4th Euro. Symp. on Space Environmental and Control Systems, Florence, Italy, Oct. 1991.
Testing of Infrared Detectors Using a Zero Gravity Dilution Refrigerator R.S. Bhatia, J.J. Bock and P.V. Mason
Observational Cosmology Group California Institute of Technology Pasadena, CA 91125 A. Benoît Centre de Recherche des Trés Basses Température, CNRS 38042 Grenoble, France
M.J. Griffin Astrophysics Laboratory, Department of Physics
Queen Mary & Westfield College London, UK El 4NS
ABSTRACT We describe characterisation of the 100 mK dilution refrigerator, detectors and readout system for potential use on submillimetre spaceflight instruments. Relevant instruments include
the High Frequency Instrument (HFI) on the Planck Surveyor mission for measurement of the anisotropies in the cosmic microwave background. The HFI will use a dilution refrigerator precooled by a 4 K mechanical cryocooler. We have used a modified variation of this zero gravity flight dilution system, precooled by a superfluid cryostat rather than mechanical cryocoolers. The cooling power at 100 mK was 102 nW. The prototype detector was of the bolometric spider web type with a germanium thermistor, as presently baselined for the HFI. The bolometer differential readout system used a sinusoidal AC bias, and is being further developed for flight use. Closed loop temperature control about a setpoint of 100 mK was also used. Dark bolometer tests were performed including measurement of detector heat capacity, time constant, noise and microphonic susceptibility. The bolometer achieved the sensitivity required for all channels of the HFI, but was a factor of three times slower than the slowest required device. A mechanical shaker was used to simulate the microphonic effects of cryocooler operation. For a bolometer acceleration of at 100 mK with 73 nW of residual cooling power, the system exhibited no increase in low frequency noise above those levels measured for the response without microphonic excitation. This acceleration level was ~ 70 times greater than the level expected from the flight cryocoolers at the 4 K enclosure for the HFI. This gives us confidence in the proposed use of sensitive bolometers with cryocoolers for space missions.
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INTRODUCTION
Instruments operating at far-infrared and submillimetre wavelengths have important applications in astronomy and Earth atmospheric sciences. Although ground-based instruments can be used in limited window regions of high atmospheric transmission, their sensitivities are compromised by the high photon flux from the dense lower layers of the atmosphere. Cryogenic cooling is essential for most instruments operating at these wavelengths, in order to reduce background radiative emission and to achieve the required levels of detector sensitivity. Spaceborne cryogenic instrumentation alone can therefore provide both the low emissivity environment, and for Earth observations, the global spatial coverage which are required. Along with the developments in detectors, there have been major advances over the past decade in the technology of spacecraft cryogenic systems. Open cycle cryostats generally do not create significant disturbances within the instrument, but have the disadvantages of high launch mass and limited lifetime. Closed cycle mechanical coolers have recently been developed and have low mass and a lifetime of several years, but also generate vibrations and electromagnetic interference (EMI) which can potentially degrade the performance of the instrument to unacceptable levels. There were two main aims to this work: the first was to build up a 100 mK testbed using a design of dilution refrigerator suitable for use in zero-gravity. The second aim was to use this testbed to evaluate the performance characteristics of infrared detectors suitable for space
missions. This included determination for the first time of the susceptibility of infrared detectors to cooler vibrations and EMI, with the tests serving as a benchmark against which the cryocooler
flight specifications can be realistically defined. For astronomy missions, the arbitrary ESA flight specification on the 4 K mechanical coolers operating with low vibration drive electronics is interpreted as allowing a maximum permissible acceleration of at the interface between the 4 K stage of the cooler and the instrument of 13 kg mass. A typical project to which this work is relevant is the High Frequency Instrument (HFI) on the Planck Surveyor spacecraft, which has baselined the use of cryocoolers (rather than a cryostat) to precool a dilution stage which will in turn cool an array of bolometric detectors to 100 mK. The cooling power available at this 100 mK stage is 100nW.
DETECTOR
Bolometers are the preferred type of detector at submillimetre wavelengths because of their high response at these wavelengths and also their high absorptive efficiency. The most important type of thermal detector is the bolometer, in which a temperature change of an absorber causes a change in the resistance of the thermometric material (usually a semiconductor) before the absorbed radiation is conducted away to a heat sink through a weak thermal link. These detectors are inherently subject to three types of noise. Photon noise arises from the random arrival rate of the incident photons at the detector. Johnson noise arises from random motions of charge carriers in the thermistor. Finally, phonon noise arises from the quantised nature of heat transport along the wires leading from the bolometer absorbing element to the heat sink. These latter two mechanisms are inherent to a bolometer and are both significantly reduced by reducing the detector operating temperature. A common figure of merit for a direct detector is the Noise Equivalent Power (NEP), defined as the incident signal power required to obtain an rms signal
level equal to the noise in a 1 Hz post-detection bandwidth. The optical NEP specification for the Planck HFI bolometers has been taken to be with a speed of response of 20 Hz for the slowest channel. The bolometer used in these tests was a second generation device of the spider web type12, as shown in Figure 1. The mesh absorber is made by patterning a thin film of low stress that is deposited in the shape of a spider’s web onto a silicon wafer. After the nitride has been etched, the wafer is diced and placed in a silicon etch, leaving only the mesh of nitride connected by a small number of legs to a solid silicon frame. Chromium then gold are then evaporated onto the centre region of the absorber through a shadow mask, producing a network
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of wires. The major advantage is that the true absorber area is now much less than for a standard composite bolometer. This leads to a significant reduction in heat capacity and hence an increase in sensitivity. Also, the suspended mass is now much smaller than for a solid absorber, and so the eigenfrequency for this structure is correspondingly much higher.
DILUTION REFRIGERATOR Problems arise with adaptation of the classical dilution system for use in space because gravity is used to maintain the phase separation in the still and also because of the complex pumping installation needed to cycle the The phase boundary in the still is of particular concern. Here the dilute solution is a superfluid, so the walls are coated with a superfluid film which can creep into the pumping lines and contribute a substantial circulation to the cryocooler. This will reduce the cooling power. In zero gravity, the film creep can reach through the cold section of the pump lines and the level will be much higher than in one g operation. Therefore, the phase boundary in the still must be controlled for successful zero g
operation. A dilution refrigerator for possible use in a zero g environment has been developed at NASA Ames3. This uses gas desorption from charcoal pumps for all pumping operations. The system has a porous metal sinter to confine and control the liquid phases at low temperatures. The sinter pore size is coarser than that usually used to confine because it was found that a very fine pore size impeded the flow of dilute from the mixing chamber to the still4. This basic system is most suitable for single-cycle operation, but can be modified relatively easily for continuous operation. A circulation refrigerator for possible zero g use has also been investigated by Hendricks5. It is similar to the so-called Leiden dilution refrigerator6 in that it circulates but there is only one phase boundary, which is in the mixing chamber. The rest of the dilute solution has sub-critical concentration. The demixing chamber can then operate at temperatures above the phase separation point and it is feasible that a vapour precooler may not be needed. Tests of a porous metal phase separator for this system have shown that trapping of a mixture against gravity can be achieved. A method using electrostiction to separate and hold the two phases in the mixing chamber has been investigated by Jackson7, but this requires high electric field intensities and is yet to be demonstrated. Instead of using osmotic pressure to extract from the dilute phase in the mixing chamber, it is possible to use the mutual friction between dissolved and to drive out of the mixture8, provided that the velocity of the is high enough. This is the principle of operation of
Figure 1 . Spider Web Bolometer.
Figure 2 . Schematic of HFI Thermal Design.
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the Benoît dilution refrigerator. The two pure liquid isotopes are precooled in a heat exchanger, then mixed together to produce the cooling due to the enthalpy difference between the pure
isotopes and the dilute mixture9. In order to keep viscous heating to a minimum, the inlet gases flow through cylindrical tubes of very small diameter. The gases are at a pressure higher than the critical pressure and there is only a small pressure drop due to viscous losses. The mix return line is of slightly larger diameter and the three tubes are soldered together forming a recuperative heat exchanger. In thermodynamic equilibrium, the concentration decreases with increasing temperature. To extract the at the warmer end of the heat exchanger, it is necessary to avoid the counterflow diffusion of in the tube by maintaining the injection rate at a level higher than the solubility limit. In these narrow tubes, the surface tension maintains a sequence of concentrated and dilute drops to avoid diffusion in the dilute phase10. As the temperature increases, the bubbles dissolve and in the warmest part of the heat exchanger it is the mutual friction between and which prevents counterflow11. To run such a system, it is not necessary to use a mechanical pump or circulator, but only to store the pure isotopes at high pressure (and ambient temperature) and allow the gases to flow at the correct rates and in the correct ratios into the cryostat using massflow meters. This system precooled using the 4 K Stirling/J-T closed cycle cooler has been adapted for potential use on astronomical missions12,13. A schematic is shown in Figure 2. The Joule-Thomson cooling effect of the returning mixture is essential for thermal shielding of the 0.1 K stage. The cooling power of the system is 100 nW at 100 mK.
TESTBED
The cryostat (referred to as Yogi) was an HDL-10 cryostat14 with a 254 mm diameter work surface. The cryogenic stages of this system in comparison to the expected flight system are shown in Figure 3, and the testbed itself is shown in Figure 4. The gas handling system allowed a designated mass flow rate of and/or to flow through the capillaries and collected the returning fluid. Following capillary evacuation, the system was purged to remove paniculate impurities from the system by flowing through both the and the capillaries. Each of the input lines passed through a liquid nitrogen trap to condense gaseous impurities. The required and mass flow rates were set using two electronic controllers15. In this laboratory system, the return mix was stored to enable the to be recovered.
Figure 3 . Yogi and HFI Schematic.
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The dilution stage was precooled to 300 mK using a sorption refrigerator. The refrigerator consisted of a charcoal-filled pump, an evaporator and a connecting tube. The system was filled to 60 bar absolute of At low temperatures the is adsorbed on the charcoal. This is driven off by heating the pump to 35 K. It condenses under higher pressure and drips
into the evaporator. The pump is allowed to cool and in doing so it pumps on the liquid, reducing the vapour pressure above the liquid and so reducing the temperature of the evaporator to ~ 300 mK. This was a single stage refrigerator, and so in order for the to condense below its critical temperature of 3.3 K, it is first necessary to pump on the main bath. To increase the condensation efficiency of the the bath was pumped down to 1.6 K. This temperature reduction also reduced the radiative heatload on the 300 mK and 100 mK stages. A heater was located on the sorption pump to cycle this refrigerator. The heatswitch between the 300 mK stage and the 100 mK stage was of the gas gap type space qualified on the Far-Infrared Photometer on the IRTS mission16. This was operated in passive mode, i.e. current was not
needed to maintain the switch in the on state. The wiring from 300 K to 1.6 K comprised two harnesses of twisted manganin pairs. The first harness carried the bias and FET drain lines and the second harness the thermometer, accelerometer and heater lines. The thermistor in Channel 1 was used in a standard bridge circuit for temperature control using an AC bias with lock-in amplification17, as shown in the upper schematic of Figure metal film resistor heater was located on the bolometer plate for heatload determination and for temperature control using a Proportional Integral Derivative (PID)
Figure 4 . Yogi 100 mK Testbed.
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Figure 5 . Temperature Control and Readout Circuits.
controller18. In Channel 2, a resistor was used as a dummy bolometer, in order to discriminate between thermal effects exhibited by the bolometer and purely electrical ones. The thermistor in Channel 3 was used for readout of the stage temperature, and the spider web bolometer was wired into Channel 4. The thermistor, bolometer and load resistor were read out as shown in the lower panel of Figure 5. The noise for each JFET19 preamplifier pair was All lines going into and out of the cryostat were enclosed in RF tight metal boxes. The thermometer, heater and accelerometer lines were screened using filters20. Additional thermometry included germanium resistance thermometers21 located on the 300 mK and 100 mK stages read out using a bridge circuit22, and carbon resistors on the sorption pump and at the condensation point which was thermally anchored to the 1.6 K base plate. A mechanical shaker23 and amplifier24 were used in these tests to simulate the microphonic input of the closed cycle coolers. The shaker was mounted on top of the cryostat, with the armature coincident with the central axis of the cryostat and the 100 mK support structure. A uniaxial cryogenic accelerometer25 was located at the 100 mK detector blanking plate to measure the cold stage acceleration in the vertical axis at a point as close to the bolometer as possible. A second such accelerometer was located at the 1.6 K plate at which the JFETs are located, again measuring acceleration in the vertical axis. The accelerometers were read out using a standard room temperature charge amplifier26.
OPERATING PROCEDURE The cryostat was precooled in the standard manner using liquid nitrogen and liquid helium to 4.2 K. During this time, was circulated through both and capillaries to remove any residual paniculate contamination. The sorption refrigerator was cycled several times, thereby functioning as a gas gap heat switch to cool down the 300 mK stage to 4.2 K more quickly. The bath was pumped on, to obtain a base temperature of 1.6 K. The sorption refrigerator was cycled by heating the pump to 35 K for 10 minutes, then allowing the pump to cool. The cycle took approximately one hour in total, after which the intermediate stage was at a temperature of
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0.3 K. A mixture of and with flow rates set to 100 % and 100 % ( and respectively) was allowed to flow through the dilution capillaries, starting from a dilution stage temperature of about 1.7 K. This initially put a thermal pulse on the sorption pump which warmed
up the sorption stage, but the sorption system recovered from this within ~ 30 minutes. The silicon JFETs were heated to 120 K, above the charge carrier freeze-out temperature and to reduce their noise. Cooldown from start of pumping on the bath to achieving 100 mK was achieved within three hours. For these tests the flow rates were maintained at 100 %.
TEST RESULTS These tests were all performed with the system optically blanked off. The bolometer time constant was determined to be 25 ms by measuring its decay response to 60 keV rays from an source. From measurement of the nominal resistance of the bolometer during cooldown of the cold stage and using the bolometer model of Griffin27, the static thermal conductance to the heatsink at 100 mK was determined to be Bolometer noise spectra were taken under PID temperature control about a nominal temperature of 100 mK. The bolometer impedance was at this operating point. First, quiescent noise spectra were taken. The output signals from the differential amplifier were further amplified by a factor of 10 using a low noise amplifier28, with an AC coupling cut-on frequency of 15 mHz and a cut-off frequency of 10 Hz. This signal was digitised using an analogue-to-digital converter29 with 20 bits resolution and read out using a PC interface board30. The bolometer noise
power spectra were then taken under sinusoidal microphonic excitation at 32 Hz. The excitation levels were initially at the 100 mK blanking plate, and at the cold plate. The levels were increased to at the 100 mK blanking plate, and at the cold plate. FFTs were performed on the data using a Hanning window function. These measurements were repeated for the load resistor and the thermistor. The microphonic susceptibility of the spider web bolometer to 32 Hz microphonic excitation is shown in Figure 6. The quiescent noise spectrum is for the device under temperature control but without microphonic excitation. The noise floor above 0.15 Hz is Between 0.01 Hz and 0.15 Hz, there is a peak in the noise level at 0.08 Hz of Below 0.01 Hz, there is a significant increase in noise. There is a peak at 7 Hz, associated with the temperature control loop. There is some rolloff above 10 Hz, due to the filtering in the postamplifier. Plotted over this is the equivalent response for an acceleration of at the
Figure 6 . Bolometer Microphonic Susceptibility.
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bolometer stage and at the 1.6 K cold plate. The amplitude and frequency response for this are comparable to that obtained in the quiescent power spectrum. Over this is plotted the response to an acceleration level of at the bolometer stage and at the 1.6 K cold plate. This shows a small level of increase in the noise over 0.15 Hz to 7 Hz, compared to the quiescent response. The susceptibility of the load resistor to 32 Hz microphonic excitation is shown in Figure 7. The measured noise floor for the quiescent noise spectrum above 0.01 Hz is The theoretically expected value arising from Johnson noise is The quadrature sum of this with the JFET noise is Below 0.01 Hz, the noise level increases substantially. Again, there is a peak at 7 Hz, associated with the temperature control loop. There is some rolloff above 10 Hz, due to the filtering in the post-amplifier. Plotted over this is the equivalent response for an acceleration of at the bolometer stage and at the 1.6 K cold plate. This is comparable to the quiescent response. The same is true for the response to an acceleration level of at the bolometer stage and at the 1.6 K plate. The current through the heater to maintain temperature control at 100 mK was measured using the temperature controller. Although the cooling power decreased under microphonic excitation, this was probably due to the frictional heating effects of the components under vibration rather than reflecting an inherent problem with the operating mechanism of the refrigerator under microphonic excitation. This implies that care must be taken in the design of the Planck Surveyor HFI to ensure that the cooling power available at 100 mK is not compromised.
The flight instrument will have an array of bolometers and as such, there will be a much larger number of wires routed from the JFETs at ~ 120 K to the 100 mK stage devices. The available cooling power will still only be 100 nW at 100 mK, and heatsinking of wires along the length of the dilution capillaries will be necessary. The challenge remains of ensuring that the parasitic heatload on the 100 mK stage from the wiring and the wiring support is still comfortably within the cooling power of the dilution refrigerator, whilst ensuring that this support structure prevents any motion of the wiring. When the acceleration is increased to at the 0.1 K stage, there is an appreciable increase in the noise floor for the bolometer, but not for the load resistor. This would indicate that the increase has a thermal origin, perhaps arising from the increasing level of temperature fluctuations due to the reduced cooling power of the dilution stage at the higher level of microphonic excitation. Problems were experienced with the temperature stability which limited the lowest
Figure 7 . Load Resistor Microphonic Susceptibility.
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frequencies which could be successfully read out. The increase in noise at ~ 0.08 Hz was present in the bolometer and thermistor spectra but not the load resistor spectra, indicating that this effect had a thermal origin. It was probably due to the limitations of the temperature control. With a current of 0.28 nA through the bolometer, the responsivity of the device was
For a measured noise level of this translates into an electrical (blanked) NEP of 2 The effective conductance of the bolometer thermal link was The time constant was measured to be 25 ms, giving a heat capacity of As such, the device achieved the required NEP for all the HFI channels, but was three times slower than the slowest required device. CONCLUSIONS (i) The dilution refrigerator has been commissioned. The cooling power at 100 mK was 102 nW. (ii) Basic characterisation of this spider web bolometer has been performed. At 100 mK, the thermal conductance was the time constant was 25 ms and the heat capacity was The electrical responsivity of the device was The measured noise level for the bolometer was The electrical NEP was The bolometer achieved the NEP required for all channels of the Planck Surveyor HFI, but was three times slower than the slowest required device. (iii) The Caltech design of AC biased bolometer readout electronics has been successfully used for these measurements. (iv) For a bolometer acceleration of at 100 mK, under temperature control with 73 nW of cooling power and an AC readout system with differential wiring, the system exhibited no increase in low frequency noise above those levels measured for the quiescent response without microphonic excitation. This acceleration level is a factor of ~ 70 times greater than the level expected from the flight cryocoolers at the 4 K enclosure for the Planck Surveyor HFI. There was a slight level of increase in the noise floor for a bolometer acceleration of with 40 nW of cooling power. These results have demonstrated that with the wiring configuration and readout electronics used here, it is possible to achieve a high degree of immunity to the effects of microphonics. ACKNOWLEDGEMENTS
We wish to thank Jeff Beeman at LBNL for supply of the NTD germanium. The bolometers were fabricated at the JPL Center for Space Microelectronics Technology. This work was partly supported by NASA Innovative Research Grant NAG5-3465 and NASA NAG5-6573 for US involvement in Planck Surveyor. R. Bhatia acknowledges a graduate studentship funded by QMW College (University of London) and QMC Instruments Ltd.
REFERENCES
1. Mauskopf, P. D., J. J. Bock, H. Del Castillo, W. L. Holzapfel and A. E. Lange, “Composite Infrared Bolometers with
Micromesh Adsorbers”, Applied Optics, Vol. 36, No. 4, 1 Feb 1997, pp. 765-
771
2. Bock, J. J., H. M. Del Castillo, A. D. Turner, J. W. Beeman, A. E. Lange and P. D. Mauskopf, “Infrared Bolometers with Silicon Nitride Micromesh Absorbers”, Proceedings of the 30th ESLAB Symposium,
‘Submillimetre and Far-infrared Space Instrumentation’, 24-26 September 1996,
ESTEC, Noordwijk, The Netherlands, ESA SP 388, December 1996, pp.119-122
3. Roach, P. R. and B. Helvensteijn, “Development of a Dilution Refrigerator for Low Temperature 4. 5.
Microgravity Experiments”, Proceedings of the 10th International Cryocooler Conference, Monterrey, California, 26-28 May 1998, Ed. R. G. Ross, Jr.,, Plenum Press, New York Roach, P. R., NASA Ames Research Center, CA 94035-1000, Personal Communication, June 1997 Hendricks, J. B., “ Dilution Cryocooler for Space”, Final Report, NASA Contract No. NAS8-37437, 11 March 1991
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6. Pennings, N. H., K. W. Taconis and R. De Bruyn Outober, “The Leiden Dilution Refrigerator II”, Physica, Vol. 84B, 1976, pp. 102-109 Dilution Refrigerators Operate Aboard Spacecraft?”, Cryogenics, Vol. 7. Jackson, H. W., “Can 22, February 1982, pp. 59-62 8. Kuerten, J. G. M., C. A. M. Castelijns, A. T. A. M. de Waele and H. M. Gijsman, “Comprehensive Theory of Flow Properties of Moving Through Superfluid in Capillaries”, Physics Review Letters, Vol. 56, 1986, pp. 2288-2290 9. Benoît, A. and S. Pujol, “A Dilution Refrigerator Insensitive to Gravity”, Physica B, Vol. 169, 1991, pp.457-458 10. de Bruyn Ouboter, R., B. van den Brandt and J. W. Tierolf, “Visual Observations of the Counterflow of the Two Liquid Phases in the Dilution Refrigerator”, Physica B+C, Vol. 107,
1981, pp.557-558 11. Paragina, A., “Etude de la Réfrigération par Dilution d’Hélium en Cycle Ouvert. Application à un Satellite d’Observation Submillimétrique”, PhD Thesis, CRTBT, CNRS, Grenoble, France (1997) 12. Benoît, A., T. Bradshaw, C. Jewell, T. Maciaszek, A. Orlowska and S. Pujol, “A Future Potential Long Life Refrigerator for 0.1 K Cooling in Space”, Proceedings of the 24th International Conference on Environmental Systems and 5th European Symposium on Space Environmental Control Systems, 20-23 June 1994, Friedrichshafen, Germany, SAE Technical Paper Series No. 941276 (1994) 13. Benoît, A. and S. Pujol, “Dilution Refrigerator for Space Applications with a Cryocooler”, Cryogenics, Vol. 34, No. 5, 1994, pp. 421-423 14. Model HDL-10 Cryostat, Infrared Laboratories, Inc., Tucson, Arizona 85719, USA 15. Model 5850i Mass Flow Controller, Brooks Instrument, Hatfield, Pennsylvania, USA 16. Duband, L., D. Alsop and A. Lange, “A Rocket-borne Refrigerator”, Advances in Cryogenic Engineering Vol. 35, Ed. R. W. Fast, Plenum Press, New York, USA, 1990, pp. 1447-1456 17. Bock, J., P. Mauskopf, W. Holtzapfel, V. Hristov and A. Lange, “AC Bridge Readouts”, Proceedings of a Workshop on Bolometers for Millimetre and Submillimetre Space Projects, 15-16th June 1995, Orsay, France, Ed. J. M. Lamarre, J. P. Torre and V. Demuyt, Institute Astrophysique Spatiale Report
RS 95-02, June 1995, pp. 181-189 18. TS-530 Temperature Controller, RV-Elektroniikka Oy Picowatt, Vantaa, Finland. Distributed by Oxford Instruments Ltd., Witney, Oxfordshire, UK 19. JFET, custom-made by InterFET Corporation, 1000 N. Shiloh Rd., Garland, Texas 75042, USA 20. RF Filter, Part 1212-0502, Spectrum Controls, Inc., Fairview, Pennsylvania, USA. 21. Model GR-200A-50 GRT Sensor, Lake Shore Cryotronics, Inc., Westerville, Ohio 43081-2399, USA 22. AVS-47 AC Resistance Bridge, RV-Elektroniikka Oy Picowatt, Vantaa, Finland. Distributed by Oxford Instruments Ltd., Witney, Oxfordshire, UK 23. V400 Shaker, Ling Dynamic Series, Royston, Hertfordshire, UK 24. PA 100 Power Amplifier, Ling Dynamic Series, Royston, Hertfordshire, UK 25. 7422 Piezoelectric Accelerometer, Endevco Corporation, San Juan Capistrano, CA 92675, USA
26. Model 104 Charge Amplifier, Endevco Corporation, San Juan Capistrano, California, USA 27. Griffin, M. J., “Theoretical Modelling of Bolometer Performance”, Proceedings of a Workshop on Bolometers for Millimetre and Submillimetre Space Projects, 15-16th June 1995, Orsay, Paris, France, Ed. J. M. Lamarre, J. P. Torre and V. Demuyt, Institute Astrophysique Spatiale Report Number RS 95-02, June 1995, pp. 27-36
28. Model SR 560 Lock-in Amplifier, Stanford Instruments, Stanford, CA, USA 29. Model DDC101 20-bit ADC, Burr-Brown Corporation, Tucson, AZ 85734-1400, USA 30. Model DEM-DDC101P-C PC Interface Board, Burr-Brown Corporation, Tucson, AZ 85734-1400, USA
Design of a 90 K Cryogenic Passive Cooler for the IASI Instrument D.J. Doornink
Fokker Space B.V. Leiden, The Netherlands
ABSTRACT
IASI (Infrared Atmospheric Sounding Interferometer) is part of the core payload of the EUMETSAT Polar System (EPS) METOP-1 and will contribute to the primary mission objective of EPS which is the assessment of meteorological parameters. IASI on METOP will operate from a low altitude, sun synchronous polar orbit. The core of the IASI instrument is a cryogenic detection system with an operational temperature of 90 K which is to be achieved with passive cooling. Fokker Space has performed a design study for this passive cooler. The performance analyses indicate a net cooling power of 108 mW at 90 K. This paper will describe on one hand the main requirements imposed and other hand the design concept chosen to ensure adequate performance.
INTRODUCTION Compared to active cooling a passive cooler has advantages in terms of power consumption, reliability, mass and exported vibrations. However the performance of a passive cooler decreases rapidly when the required temperature levels become lower. Challenging in this respect is the low temperature to be achieved on the coldest stage of the IASI main cooler. The heat rejection capability of a passive radiator drops fast around a temperature level of 90 K because of the law of heat transfer by radiation. On the other hand a multi-stage concept can reduce radiative parasitic heat flows on the coldest stage by virtue of the same law. Theoretically speaking this principle of increasing the number of stages can be pursued until the heat balance finally becomes governed by the parasitic heat flows through the structural supports of the coldest stage and through the electrical wiring that connects the cryogenic detector in the cold box unit to the outside world. In practice the structural stiffness and strength of the assembly and its increasing complexity tend to limit the number of stages to two or three. This delicate balance between thermal performance and mechanical integrity needs to be supported by adequate verification. For this reason in an early stage of development the design is
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Figure 1. Envelope available for the IASI main cooler.
supported by tests on a breadboard model as well as the usual analyses. In this paper the scope and results of performance analyses and breadboard tests will be presented.
MAIN REQUIREMENTS Following is a selection of the main design driving requirements imposed on the radiator assembly.
Operational Cold Box Temperature. The temperature of the cold box shall be kept below 100 K, given a cold box dissipation of 68 mW and a harness heat leak of 40 mW. In a conceptual definition like described in this paper a margin of 10 K on the temperature to be achieved seems
appropriate, so the aim is to have the cooling performance at 90 K. Non Operational Cold Box Temperature. The maximum temperature of the cold box interface shall be below 60°C. This is a design constraint which becomes active during S/C SAFE mode in which solar flux impinges on the radiators. Position Stability of the Cold Box. Between begin-of -life and end of life the actual field stop centre shall be located at the position of the desired field stop centre with a position tolerance of
Minimum Fundamental Frequency. The minimum fundamental frequency for each axis direction shall be greater than 110 Hz. Mass. The radiator assembly mass shall be less than 18 kg. IASI Main Cooler Envelope. The dimensions of the envelope for the IASI main cooler are indicated in Figure 1.
The IASI Environment on the METOP-1 Satellite IASI (Infrared Atmospheric Sounding Interferometer) is part of the core payload of the EUMETSAT Polar System (EPS) METOP-1 and will contribute to the primary mission objective of EPS which is the assessment of meteorological parameters.
The METOP-1 orbit is sun-synchronous with 10:00 descending node local time and a mean altitude of 800 km. The -Z axis of the satellite is pointing in NADIR. This means that the
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Figure 2. The IASI instrument on the METOP-1 satellite.
METOP-1 top panel is always shaded from solar flux. Instruments like IASI on this panel can have surfaces without solar illumination viewing space suitable as a radiator with a stable flux environment and thus a good temperature stability. On the IASI instrument bench several units are mounted, among which the main cryogenic radiator is the largest . The IASI bench is oriented such that the earth horizon remains below its plane with a margin of 1°, as can be seen
in Figure 2. This means that in the nominal attitude of the satellite the top face of the bench will not receive earth fluxes. This is an important aspect for the design of the cooler. SELECTED COOLER CONCEPT
The cooler concept has been the subject of an extensive trade-off process. In this trade-off experience from previous passive cooler projects was used: The MeteoSat Second Generation (MSG) passive cooler is a two-stage design with a net heat rejection capability of 50 mW at 90 K. This cooler accommodates the focal plane system of SEVIRI (Spinning Enhanced Visible and InfraRed Instrument), which is the main instrument of this meteorological satellite. Compared with IASI the envelope volume is much larger however and MSG is not bothered by earth fluxes, since it operates in a geo-stationary orbit. The SCIAMACHY Radiant Cooler (SRC) is a two stage 130 K concept to be flown on the SCIAMACHY instrument on ENVISAT-1. The orbital flux environment on ENVISAT-1 is
similar to the IASI situation, but the envelope volume available to the cooler is larger. SRC also includes a thermal bus by which the remote detectors of the instrument are cooled. The details of this trade-off are beyond the scope of this paper. Simplified model calculations and the experience from other passive cooler projects have shown that the available envelope for the IASI cooler makes it very difficult to reach the 90 K with a two-stage concept. Two stage concepts are preferable in view of stiffness, strength and cold stage position stability performances, but with special attention for these aspects a three stage concept seems to be feasible.
The main purpose of the IASI main cooler radiator assembly is to provide a cold thermal environment both conductively and radiatively for the third stage radiator on which the cold box
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Figure 3. IASI Main Cooler Concept.
unit with the detector is mounted. This environment consists of the second stage, the reflector of the sun shield and deep space. Heat leaks towards the third stage through supports and harness must be reduced as far as possible. Detector dissipation and irradiation of the detector and its surroundings shall be minimised as well. Cooler Orientation. The radiator surfaces of the three stages are all in one plane, parallel to
the IASI bench, see Figure 2. This way in the nominal S/C attitude the radiator stages will not receive earth fluxes. Sun Shield. The sun shield prevents direct or indirect illumination of the first, second and third stage radiator surfaces by solar flux. The sun shield edges act as additional radiator surfaces to reduce the temperature of the sun shield structure. These sun shield radiator patches are covered with SSM at the space-side and with MLI at the back. At the back the sun shield is covered with MLI blankets to reduce heat input from the IASI internal environment. The sun shield is integrated on the first stage.
Reflector. Integrated with the sun shield, a high performance reflector prevents earth fluxes from reaching the radiator surfaces but at the same time maintains their view to space. The reflector is cylindrical with such a radius that parallel radiation from the earth in between the horizon and the region below that is reflected back over the front edge of the first stage radiator. The reflector has a low-emissivity, high specularity gold coating to minimise self-emission and scattering of earth fluxes onto the first and second stage radiators. The temperature of the reflector is of major importance to the cold stage heat balance and must be kept as low as
possible. Therefore the reflector is covered with MLI blankets at the back side. These reduce the heat input form the IASI environment towards the reflector. The sun shield radiator patches reject parasitic heat flows. First Stage. The first stage consists of the radiator stage with the reflector and the sun shield integrated on it. The first stage is constructed from aluminium to minimise thermal gradients The first stage radiator is coated with black paint to have maximum emissivity in the infrared. At the instrument side the inward facing surfaces of the first stage are goldised to minimise radiative heat leaks. The first stage has a thermal shield that prevents direct radiative heat exchange between the support frame MLI and the instrument facing side of the second stage. The optical baffle penetrates the stage structure and the thermal shield.
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Second Stage. The main thermal function of the second stage is to reduce the internal parasitic heat loads on the third stage. The second stage is constructed from aluminium The second stage radiator is coated with black paint to have maximum emissivity in the infrared. At the instrument side the inward facing surfaces of the second stage are goldised to minimise radiative heat exchange with the first stage interior. The second stage has a thermal shield that
prevents direct radiative heat exchange between the first stage interior and the instrument facing side of the third stage. The optical baffle penetrates the second stage and its thermal shield. Third Stage. The third stage is constructed from aluminium to minimise thermal gradients. The third stage radiator is coated with white paint to have sufficient emissivity in the infrared while avoiding excessive heating during spacecraft SAFE mode. At the instrument side the
inward facing surfaces of the third stage are goldised to minimise radiative heat exchange with the second stage interior. Optical Baffle. The signal radiation reaches the cold box unit via a three stage baffle. One baffle part is integrated on the cold box unit at the third stage and two more baffle parts are integrated in the second and first stage, respectively. The optical baffle interior is black. This triple configuration helps to reduce the radiative heat load on the optical entrance of the Cold box Unit. The optical baffle is goldised at the outside like the surfaces of the surrounding stages. Cooler Support Frame. This frame forms the interface between the IASI bench and the first stage structure of the cooler. The inner surfaces of the support frame facing the first stage
Figure 4. Exploded view of the cooler assembly (cold harness and optical baffle not shown).
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box are covered with MLI to minimise radiative heat exchange with the first stage and other
interior parts. Cover Door. In stowed configuration the door is closed, i.e. protecting the sensitive reflector surfaces and the radiator areas for contamination from the external environment of the S/C. In orbit the door is released and latches in the open position. The release mechanism is
derived from the well-proven thermal knife. In this configuration the door shields part of the earth fluxes from the reflector. The upper side of the door is exposed to solar flux. To reduce the thermal emission and reflection of solar flux this side of the door has a low-emissivity coating. In open configuration the cover door has Kapton VDA at the space side to reflect solar flux. At the earth side the door is covered with SSM to keep the door temperature both in open and in closed position as low as possible. Special Design Features
Inter-Stage Support Elements. The three stages are supported by glass fibre reinforced plastic (GFRP) elements to provide sufficient thermal insulation. These GFRP elements have been tested on strength/stiffness and fatigue behaviour in a representative breadboard model of a similar cooler. Radiative insulation from the highly reflecting inter-stage cavities is to be achieved with a low-emissivity coating on the GFRP. The design of the blades includes restriction hinges in the end-fittings. This avoids over-determination in the support of the stages
and thus minimises internal stresses during the temperatures excursions. The lay-up of fibres in the composite is adapted to the lines along which the internal stresses in the elements occur. Third Stage Position Stability. The detector assembly is mounted on the third stage. The position stability and the required stiffness are major design drivers for the structural design of the radiator assembly. The alignment of the IASI front end optics with a reference point in the cold box unit must be preserved under all operational conditions. The main disturbing influences are the vibration loads occurring during launch and the temperature excursions when the cooler is cycled in between operational temperatures and room temperature.
During the mission the cooler is decontaminated on a regular basis to remove volatile contaminants. The phenomena causing the displacements are 1) settling of material interfaces due to large internal stresses and 2) contraction or expansion due to temperature variations. Internal stresses are reduced as far as possible by having quasi-isostatic suspension of the stages.
Figure 5. Harness clamping on first and second stage.
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During cool down from room temperature to the operational temperature levels the internal stresses due to different rates of contraction are suppressed this way. Also small mismatches due to production tolerances will not lead to internal stresses and pre-tension. The alignment of the IASI front end optics to the optical centre of the cold box unit shall be preserved within certain margins. The thermal displacement of the cold box unit can be split into
two parts:
a) the motion of the third stage in the plane of the radiators. This motion is compensated to first order by an axial-symmetric positioning of the GFRP elements around the optical centre of the cold box unit. b) the motion of the third stage in a direction perpendicular to the radiator plane. In the mechanical design a provision for compensation of shrinking effects for this direction is foreseen. When the radiator assembly cools down from ambient to operational temperature levels the various construction elements will shrink according to their various thermal expansion coefficients. The configuration of the supporting elements is chosen such that to first order the
displacement in out of plane direction is zero. S/C SAFE Mode Provisions. The first and second stage radiators are coated with a highemissivity black paint to have maximum emissivity in the infrared. The third stage radiator
however is coated with white paint to avoid excessive heating during spacecraft SAFE mode. In the spacecraft SAFE mode the satellite is permanently oriented towards the Sun, with the spacecraft +Z axis sun-pointing. In this situation the radiator surfaces on the stages receive solar flux. For the third stage with the sensitive detector system the temperature would rise too high if black paint were applied, so white paint is selected here as a compromise. Decontamination Heating. On all stages and on the sun shield / reflector decontamination
heater circuits are installed in order to heat up the cooler assembly up to 30°C in operating
conditions in orbit. The total installed decontamination heater power is 140 W. Harness Clamping. To minimise the parasitic heat flow through the harness towards the cold box unit, the harness is clamped to the first and second stage, see Figure 5. These stages provide cold anchors that will bring down the local harness temperature. Smooth internal surfaces. The radiative heat flow between the stages is minimised by having a low-emissivity coating on the surfaces. Apart from the stages main structure surfaces, which are quite smooth, there are other small elements present in the cavities in between the stages. For instance the GFRP blade elements, the harness and other irregularities. In such a lowemissivity cavity the parts with higher emissivity tend to collect the multiple-reflected radiation on them, resulting in an increased heat transfer. Therefore the GFRP elements have a lowemissivity coating to minimise this effect. Also bolt heads and cavities are hidden behind lowemissivity shields.
PERFORMANCE ANALYSIS RESULTS
Temperature Distribution The thermal performance was analysed with a 37-node thermal model. In Figure 6 the temperature distribution of the cooler assembly is given. The calculations assume operating conditions under a worst case hot flux environment, end-of-life optical properties and a maximum interface temperature on the IASI bench.
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Figure 6. Temperature distribution of the cooler in operating conditions.
Third Stage Heat Balance
From the calculations a heat balance for the first, second and third stage was derived. The heat flows are based on the worst case hot circumstances. In the third stage heat balance the largest parasitic heat load (59 mW) comes from the reflector that irradiates the radiator surface directly. The dissipation of the detector (68 mW) is a given quantity. The third contribution (48 mW) comes from the GFRP elements that support the
third stage on the second stage.
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The main parasitic heat loads on the third stage are reasonably in balance. Critical areas for the third stage heat balance are firstly the heat load from the reflector. This heat flow is reduced by having a low-emissivity and highly specular coating on the reflector and its surface and by reduction of the reflector temperature. Next contribution is the heat leaks through the GFRP elements supporting the third stage on the second stage. The optimisation of the cold harness for low thermal conductance is another point of attention in the detailed design. This heat balance clearly indicates the importance of the reflector coating. In fact this is the main area in the design where there is direct heat transfer between third stage and first stage (reflector) which is against the principle of a multi-stage concept. We have considered the option of having an additional reflective shield on the second stage that blocks the view from the third stage to the first stage reflector. This option turned out to be ineffective due to flux trapping in the area between the shields. On the right hand side of the balance the heat rejection to space could be improved by applying a radiator coating with higher emissivity than white paint. Due to the requirements with respect to the SAFE mode in combination with the maximum allowable cold box unit temperature this is not allowed and the white painted third stage radiator remains the baseline. In SAFE mode the third stage currently reaches a temperature level of 40°C, which leaves some margin with the 60°C requirement.
Mechanical Analyses Derived from design drawings of the IASI main cooler as generated by Fokker Space a MSC-NASTRAN model of this cooler is created. In the FEM-model the parts as indicated in Figure 4 are represented. With this model several analysis runs were performed. The lowest two natural frequencies are 110.2 and 128.7 Hz. These modes consist of movement of the stages in the plane of the radiators. During the further development the margin on the 110 Hz requirement shall be increased by design measures like mass reductions and local stiffeners. Breadboard Test Program
Parallel to the analysis and design work a slightly simplified breadboard model of the cooler assembly was designed and manufactured. This breadboard model was based on a two-stage cooler with similar geometry as the IASI main cooler. Major objectives of this model were the qualification of the GFRP support elements and the verification of position stability of the cold stage with respect to the environment. The breadboard model has been subjected to various tests:
Stiffness, Strength And Fatigue Aspects. The breadboard model is subjected to sinusoidal and random vibration loads. The vibration modes and in particular the lowest eigenfrequency is determined this way. Random vibration with extended duration are used to simulate the launch circumstances to verify whether the assembly can withstand the launch loads and to verify the fatigue behaviour. Alignment Verification After Vibration Loads. The preservation of alignment after imposing the launch vibration loads is verified in tests on the breadboard model. These measurements were combined with the regular vibration tests mentioned above. The measurement of displacements of the cold stage in vacuum conditions was implemented with electromagnetic sensors. Alignment Verification After Thermal Cycling. The preservation of alignment after
thermal cycling from ambient to operational temperature is verified in tests on the breadboard model. These tests were carried out separately in a thermal vacuum test facility. Here also the electromagnetic sensors were used. Breadboard Testing Results
In general the breadboard testing was quite successful. In the design of the breadboard the
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sizing of the GFRP support elements some margin was allocated, which is common practice when dealing with composites. The mechanical properties of the GFRP was better than expected, leading to a comfortable margin on the stiffness requirements. Some optimisation of the GFRP cross section resulting in improved thermal insulation is allowed here. The vibration tests showed some initial settling of attachments after which the behaviour became stable with no further settling. The tolerance on the on the cold stage position was less than per stage in any direction. The fatigue behaviour was verified with a random vibration profile representative of the launch loads. No damage was observed until the load duration was exceeded significantly. With respect to the alignment after thermal cycling the measured displacements were below per stage in directions parallel to the radiator surface and reproducible. In out of plane direction the displacement was much larger, since the breadboard model did not have the compensation configuration as described above. CONCLUSION This design study has shown that for the requirements and constraints imposed by the IASI instrument a passive cooler with sufficient performance is feasible. The performance has been verified by analyses. Verification of structrural integrity has been the initial driver for the definition of a testing program. Critical technologies like the GFRP insulating support elements have been qualified by tests carried out on a representative breadboard model of the cooler assembly. Altogether this work is a good starting point for further development.
ACKNOWLEDGEMENTS
This design study and the breadboard verification testing program for the main passive cooler of the IASI instrument was supported by the Netherlands Agency for Aerospace Programs NIVR. S. Butterman, M. Lemmen, F. Maitimo, A. Morgownik and G. Velsink have contributed to the design and analysis work presented here. J. Webber has conducted the breadboard testing program.
NOMENCLATURE-ABBREVIATIONS
(F)SSM GFRP METOP MSG EPS IASI SCIAMACHY MLI
(Flexible) Second Surface Mirror Glass Fibre Reinforced Plastic Metereological Polar Mission Meteosat Second Generation EUMETSAT Polar System Infrared Atmospheric Sounding Interferometer Scanning Imaging Spectrometer for Atmospheric Cartography Multi-Layer Insulation
REFERENCES 1.
Wigbers, I. et al., “SCIAMACHY Radiant Cooler”, Paper of the International Cryocooler Conference 9, Cryocoolers 9, Plenum Press, New York (1997), pp. 917-923.
2.
H. Petersen et al. “MSG Performance Verification”, paper to be presented at the 26th International Conference on Environmental Systems.
Cryocoolers for Human and Robotic Missions to Mars P. Kittel 1 , L.J. Salerno1, and D.W. Plachta2 1
NASA Ames Research Center
Moffett Field, CA 94035-1000, USA 2
NASA Lewis Research Center Cleveland, OH 44135-3192, USA
ABSTRACT Future missions to Mars will make use of a number of different cryocoolers.
These
cryocoolers will be used to liquefy and preserve propellants and for transporting liquid hydrogen
feed stock. The earliest mission requiring a cryocooler is the 2003 robotic mission, which will demonstrate technology for In-Situ-Consumable-Production (ISCP) as part of the long-term goal of developing In-Situ-Resource-Utilization (ISRU). Later missions may require cryocoolers to preserve propellants on Trans-Mars Injection stages, Martian landers, and Mars return vehicles. In addition, ISCP may require liquid hydrogen feed stock as part of the chemical process to
produce propellants from the Martian atmosphere. Cryocoolers will be required to minimize or eliminate the boiloff of this hydrogen in transit to Mars. Propellants produced on Mars will need liquefiers and cryocoolers for storage on the Martian surface until the propellants are used in the ascent vehicle. This paper will present a description of the planned missions with an emphasis on the various cryogenic options being considered and on current estimates of cooling requirements. These estimates are continuing to evolve. They are based on ongoing design and trade-off studies. INTRODUCTION
NASA is planning an extensive set of exploration missions. Prominent in this set of missions is a series of human missions to Mars. These missions plan to make extensive use of cryogenic propellants, some of which will be manufactured on Mars. This In-Situ-Consumable-Production (ISCP) is part of the scheme to reduce the mass launched from Earth through In-Situ-ResourceUtilization (ISRU). Making propellants on Mars requires liquefaction, storage, and transfer of cryogens on Mars. Some of the proposed chemical processes for ISCP need hydrogen as feed stock. The hydrogen would most likely be transported from Earth as a liquid Other phases of the mission require long-term storage of cryogenic propellants. Current estimates of these requirements are summarized in Table 1.1 A more complete description of the mission scenario is given elsewhere.2 Prior to the human missions, there will be several opportunities to test ISRU components on robotic missions. Currently, the missions planned for the 2001, 2003, 2005, and 2007 launch windows are candidates for testing parts of ISCP. Three of these, the 2003, 2005 and 2007 missions may involve cryogens. The latter two missions are sample return missions. The 2003 mission includes a partial ISRU demonstration that may liquefy oxygen. One option being
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considered for the 2005 mission is for the return vehicle to use liquid oxygen and methane for propellants. These propellants would be brought from earth. The 2007 mission may make the propellants on Mars, an end-to-end ISCP demonstration. Estimates of these propellant requirements are shown in Table 2. Table 3 gives the liquefier requirements for both the human and robotic missions. Several options are being considered for the human missions. The baseline approach is
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reflected in Table 1 with 2 descent options. One option uses the same tanks for ascent and descent at Mars. The other option uses separate descent tanks. The human and robotic missions will require a variety of cryocoolers to liquefy and/or store propellants. This paper reports on a preliminary analysis of the cooling requirements for these missions. Two options were considered for storage tanks: passive insulation systems and hybrid systems incorporating passive insulation and cryocoolers. The results clearly show the advantage of using cryocoolers. In the study, none of the systems makes use of cooled shields whether vapor cooled or actively cooled. The addition of cooled shields will be considered in the next phase of the study. The mission costs depend principally on the total mass that is launched from Earth into low Earth orbit (IMLEO). The missions will not be feasible if IMLEO becomes excessive. Thus, the criterion for evaluating different mission options is the minimization of total system mass. This mass includes the tank, insulation, propellant (including propellant lost to boiloff), cryocooler, power supply (except in mission segments with excess power, such as the TMI stage), and thermal radiators. THE MODEL
An analytic model has been developed to estimate the mass of passively insulated storage tanks and hybrid systems. The hybrid systems include cryocoolers to reduce or eliminate the
boiloff. The system mass is where:
for argon and xenon, 15.1 for hydrogen, and 5.35 for oxygen and methane tanks;1 and V includes 3% ullage and 2% residuals. The right hand side of Eq. 2 contains terms for MLI, 1.27 cm of closed cell foam,3 and a purge bag. Foam insulation is only used on hydrogen tanks with less than 75 layers of MLI to reduce the effects of moisture and air condensation. Purge bags are used for the same reasons on hydrogen tanks with more than 75 layers of MLI. The thermal conductivities of the foam and purge bag are ignored, as their conductivities in vacuum are considerably greater than the conductivity of the MLI.3 Eq. 3 is is
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based on an informal survey of estimates of what flight cryocoolers typically weigh at present and assuming that a 10% reduction in mass is achievable. This is a conservative estimate as some small flight cryocoolers weigh 1/2 of this estimate. Neither does the estimate take into account the mass reductions allowed by the lower ambient temperature of Mars. In Eq. 4, Earth refers to systems that only operate near Earth, while Mars refers to systems that must also operate on or near Mars.4 The masses include the increased system mass and 25% contingency. The near Earth system is assumed to be 40% of a near Mars system. Eq. 6 is the mass of the radiator.5 Eq. 9 is based on a published correlation for calorimeter data.6 The degradation factor, f, takes into
account the reduced performance that MLI exhibits when applied to tanks with penetrations. It varied from 1.7 for 20 layers of MLI to 3.3 for 240 layers. Eq. 10 is based on the best test data available.3 The efficiency used in Eq. 12 (10%) is also conservative as some cryocoolers are
predicted to approach efficiencies of 20% in the next few years. The cryocoolers used in the liquefiers are sized by simply considering the enthalpy change of the condensing gas plus the parasitic loads on the tanks:
For the 2003 and 2007 missions the liquefiers will only operate when there is sufficient sunlight to power the solar arrays. This is about 8 hours per day. The parasitic load, in Eq. 13 represents the total daily heating being removed during the limited operating period. Sizing the liquefier involves an energy balance between the incoming gas flow and the
cooling provided through the heat exchanger. The former is a function of the mass flow rate, latent heat, and enthalpy of the gas, while the latter depends on the surface area of the heat exchanger and the temperature difference (subcooling) between the exchanger and the liquid. For
the case of a simple liquefier, laminar film condensation on a vertical plate applies. The applicable correlation is:7
where: and g is
on Mars.
Figure 1. Liquefier heat exchanger surface area for methane and oxygen, from Eq. 15.
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An evaluation of Eq. 15 is shown in Figure 1. There the required heat exchanger surface area is shown for both methane and oxygen with 0.5 K and 1 K subcooling. The curves were calculated assuming that the incoming gas is at 240 K and that the height of the heat exchanger (fin length) is 5 cm. The area is not a strong function of the fin length. Halving the fin length only reduces the required area by 16%.
COOLER REQUIREMENTS A number of different mission segments were analyzed. Each propellant tank was sized for two different scenarios. One used only passive insulation. The other used a cryocooler sized to eliminate the boiloff. The tank volume was larger in the first case to compensate for the boil off.
These tanks were sized to have the same propellant mass at the mission end as the zero-boiloff tanks have for the whole mission. The propellant masses at mission end are given in Table 1 and the system masses in Table 2. In each case, the minimum mass system was found with the constraint that None of the systems considered here used a meteroid shield. The results of these analyses are presented in Table 4. In all cases considered here the hybrid approach results in a significantly lower system mass. This is illustrated for a typical case in Figure 2. The hybrid system is very insensitive to the number of MLI layers. As the number of layers increases, the increase in MLI mass is accompanied by a reduction in the parasitic heat load and a reduction in the cooler mass. The passive system has a broad minimum. The principal reason the passive systems have more mass is that they are launched with more fluid and, thus, have larger, heavier tanks.
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Figure 2. Comparison of the hybrid and passive system masses for the TEI oxygen tanks.
While the descent/ascent tanks are used for 1605 days, they only store a full load of fuel for
less than 700 days at a time. Thus, the tanks are sized for the 700-day hold between filling on Mars and ascent. The passive version of the ascent tanks requires larger liquefiers or additional coolers to re-liquefy the boiloff. If the tanks are cooled to eliminate the heat leak during liquefaction then 4.7 W and 1.8 W of cooling are needed for the and tanks respectively. If the boiloff is re-liquefied, then the increased liquefier cooling powers are 7.7 W and 2.8 W for the and tanks respectively. The corresponding masses are not included in Table 4. However, the additional masses are not significant. CONCLUSION
There are a number of missions to Mars in the early planning phases that could benefit from the use of cryogenic propellants. Some scenarios make use of cryocoolers to preserve propellants, while others also liquefy propellants produced on Mars. Preliminary analyses of the cooling requirements for these missions show the benefit of using cryocoolers to preserve the propellants. In all cases considered here, using a cryocooler to eliminate boiloff reduced the system mass. In the study, none of the systems makes use of cooled shields whether vapor cooled or actively cooled. The addition of cooled shields will be considered in the next phase of the study. If all passive storage is used with in situ liquefied propellants, larger liquefiers or additional coolers will be needed to re-liquefy the propellants lost to boiloff. The greatest benefit from using cryocoolers comes during the mission phases that require the longest storage times. These mission segments also put the most stringent requirements on cryocooler lifetime, efficiency, and mass. REFERENCES 1. Kos, L. and Alexander, R.A. MSFC, private communication. 2. Salerno, L.J., and Kittel, P.; “Cryogenics and the Human Exploration of Mars”, Cryogenics 38
(1998) in press.
3. Leon, L.J. and Martin, J.J.; “Experimental Testing of a Foam/Multilayer Insulation (FMLI)
Thermal Control System (TCS) for use on a Cryogenic Upper Stage”, Space Technology and Applications International Forum, Albuquerque, NM (1998), American Institute of Physics Conference Proceedings 420, (1998), pp. 331-347. 4. Withrow, C.A. and Morales, N.; “Solar-Electrochemical Power System for a Mars Mission”, NASA Technical Memorandum TM 106606, December 1994.
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5. “Investigation of External Refrigeration Systems for Long Term Cryogenic Storage”, Lockheed Missiles & Space Report A981632 (1971) 6. Nast, T.; “Multilayer Insulation Systems”, Handbook of Cryogenic Engineering, G Weisend II, Ed., Taylor and Francis, (1998) chapter 3.4.1 7. Kreith, F.; Principles of Heat Transfer, 3rd Ed., Intext Press, NY (1973), p. 527
NOMENCLATURE A f g k L
surface area MLI degradation factor gravitational acceleration latent heat of condensation enthalpy of gas thermal conductivity of liquid fin length of liquefier
mass flow purge bag mass cryocooler mass
foam mass insulation mass meteorite shield mass MLI mass propellant mass radiator mass
solar cell mass system mass tank mass
N P
layers of MLI density of MLI (layers/cm) power input to cryocooler heat flow through foam heat flow through MLI heat flow through penetrations
heat load at Tc liquefier load
cold temperature hot temperature volume dynamic viscosity subcooling of heat exchanger tank factor
emissivity of MLI efficiency density of gas density of propellant
Design Considerations for Industrial Cryocoolers C.M. Martin and J.L. Martin Mesoscopic Devices, LLC, Golden, Colorado, USA 80401
ABSTRACT In sharp contrast to cryocoolers designed for use in space or laboratory applications, cryocoolers intended for installation in industrial or outdoor environments must operate over extremes of environmental temperatures, and must survive harsh environments including rain, salt fog and structural vibration. As cryocoolers move from protected laboratory environments to widespread industrial applications, new design rules will be required. Using the design of cryocoolers intended for cooling towertop telecommunications equipment for the wireless industry as an example, this paper describes the unique design considerations required to insure reliable operation in a wide range of outdoor installations. Since most outdoor applications will require air cooling, the design of the heat rejection system becomes a key element in the system optimization. We show that there is a maximum
value of the thermal resistance of the heat rejection system (°C/W) for which stable operation at design conditions is possible, and present a methodology for determining this critical value. INTRODUCTION One of the areas of potential future growth in cryogenics is industrial applications, driven in part by high temperature superconductors (HTS). Until now, small cryocoolers have been designed for three well-defined applications: laboratory use, space refrigeration and military tactical coolers. As the markets for industrial applications mature, it will be necessary to reexamine the design requirements of the coolers to better match these new applications. Industrial applications demand operation under a wide range of conditions which, for simplicity, can be combined into three groups, environmental requirements, installation issues and heat rejection requirements. This paper will consider the range of requirements posed by one such application - coolers for high temperature superconducting filters used in wireless telecommunications. One condition in particular will prove a significant challenge - extreme temperatures. Many industrial applications will not have cooling water available, forcing the coolers to reject their excess heat to air. In this case, the heat rejection system becomes critical to the operation of the cooler. We will explore the effects of heat rejection in greater detail below, and will show that there is a maximum allowable value of the thermal resistance that still enables stable cryocooler operation.
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ENVIRONMENTAL REQUIREMENTS
The required operating range will depend on the specific application being considered. For example, indoor and outdoor industrial applications will have significantly different conditions for operating temperature and precipitation. One application where small cryocoolers are already
being used is in the telecommunications area, where the coolers are used to cool high temperature superconducting filters for wireless telephone base stations. The filters, and hence, the coolers, are being installed both inside the environmentally controlled base station, and at the antennas outside. It is this second application, tower top mounted cryocoolers, that we will use as an example case. For the telecommunications industry, there are many sources of environmental standards. Two of the most common standards organizations are Bellcore in the United States and ETSI, the European Telecommunications Standards Institute, in Europe. The standards cover all issues associated with the operating environment. The following paragraphs provide a partial list of conditions relevant to towertop telecommunications applications. Temperature. A typical operating temperature range specified by Bellcore is -40 to +60°C
(-40 to +140°F). While cryocooler designers usually consider differential thermal expansion at the cold end, over this range of operating temperatures, differential thermal expansion throughout the cryocooler system becomes significant. Many materials commonly used to improve thermal
interfaces, such as thermal greases at the heat rejection system, can not be used at the high temperature extremes. Oil separation in oil lubricated compressors becomes very difficult at the
high temperatures. Oil properties change greatly over this temperature range, making start-up a concern at low temperatures, where the oil viscosity is very high. As an example of some of the problems which may be encountered, Superconducting Core Technologies found that differential thermal expansion between the warm end seals and displacer piston significantly reduced the cooling capacity of a pneumatically driven G-M cryocooler at both low and high ambient temperature extremes. Humidity. 0 to 100% relative humidity with condensation can be expected in some applications. The possibility of condensation requires that electronics must be in sealed enclosures, or must otherwise be protected against the humidity (i.e. conformal coating of printed circuit boards). Corrosion of connector contacts will lead to the use of specialized connectors. Appropriate solutions are available, but cryocooler designers will be forced to look to other areas, such as the automotive industry. Standard fans will not be acceptable in condensing environments. Fans with sealed electronics will be required.
Thermal cycling, thermal shock. In unprotected environments, the temperature will vary with the weather and the time of day. Bellcore specifies test to mimic these thermal cycles. Precipitation. Cryocoolers designed for outdoor applications will be exposed to all forms of precipitation, including rain, snow and freezing rain. In addition, they may also experience any form of precipitation in the presence of wind. As is true for humidity, this will place
significant demands on the protection of electronics and electrical connectors. In addition, for systems with forced air convection cooling, air intakes and exhausts must be designed to prevent blockage. Sand, blowing sand. Bellcore publication GR-13 specifies the requirements for operating in the presence of sand and blowing sand. Sand will affect all exterior sealing surfaces. Delicate electronics can be damaged by abrasive sand if not properly protected. Heat rejection systems must be able to deal with sand and dust. For natural convection, fouling of the surfaces must be
assumed so that adequate margin is available. For forced convection, robust fans and carefully planned air flow passages may be required. While air filters are an obvious solution, they may not be appropriate for low maintenance applications, where regular filter changes may prove unacceptably expensive.
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Salt fog. Salt fog is covered by Bellcore publication 2836, as well as by several ASTM standards. As with several other environmental conditions, electronics are one of the most
sensitive components in cryocoolers. In addition, installation in a salt fog environment will reduce the number of materials available for use. Many commonly used metals have unacceptable corrosion rates in salt fog. The presence of salt fog and humidity requires careful
consideration of electrochemical reactions. Not only must the different parts of the cryocooler be considered, but the designer must also consider what the cooler will be mounted to and the mounting hardware itself. Many standard material choices for cryocooler external components,
including many grades of aluminum, copper and even stainless steels, will prove unacceptable in environments where salt fog is present. As an example, G-M cryocooler helium flex lines are commonly supplied in grade 303 or 321 SS. Both of these alloys are subject to pitting corrosion in salt fog environments and will fail in an accelerated salt fog life test within a matter of hours. The requirement for careful materials selection forces the manufacturer to maintain close control over not only their manufacturing process, but also the raw materials used to fabricate cryocooler components.
Lightning. For the towertop telecommunications applications, lightning protection is very important. Lightning protection devices are commercially available, but to use them correctly, a
system level grounding solution must be planned and implemented. A key uncertainty in lightning protection is the large range of potential lightning strikes. The peak current and voltage can vary by orders of magnitude. When choosing among lightning protection devices, ability to arrest a power surge must be balanced against non-reusable parts, which can mandate service after every lightning strike. Wildlife, insects. Outdoor applications will expose cryocooled systems to a range of wildlife and insects. Both performance and maintenance can be affected. For ground mounted
systems, the heat produced by the cryocooler makes the system, or the space beneath it, an attractive place for small animals to spend the winter. These animals can damage electrical wiring and other components. Keeping insects out, particularly ones such as wasps and hornets, is important for maintenance. To keep these insects out, holes must be no larger than 1/8” (3.2 mm). A final concern is the possibility of birds nesting on the top of the system. For a natural convection heat rejection system, bird nests can significantly reduce the air flow and increase the thermal resistance.
INSTALLATION ISSUES
Unlike in laboratory environments, installation requirements for industrial applications can have a significant impact on the design. For the example case above, with HTS filters and cryocoolers mounted on a wireless telephone tower, the problems are especially severe. These towers are from 5 to 100 m tall (16 to 327 ft), and are often located in remote locations.
Vibration. Unlike several current cryocooler applications where designers try to minimize the vibration transferred from the cooler to the load, in this application the vibration transferred from the tower to the cooler is dominant. Since every tower is constructed differently, and the wind loads vary from site to site and from time to time, very little quantitative data is available on the vibration of the tower. A first order estimate of the accelerations indicates that the tower
vibrations applied to the cryocooler can be on the order of 0.5 g’s at 1 Hz and 1-10 g’s at 60 Hz. Vibrations are broadband in nature and significant accelerations are found up to a kHz or more. Vibration applied to an operating cryocooler is typically not considered in current designs, but it must be considered when determining the life of the cooler. Electronics and close-tolerance seals are just two areas that could be severely affected by tower vibrations. Mounting. Towertop mounting of a cryocooler forces the designer to consider several additional issues not typically encountered. First, the cooler may not be truly vertical when installed. For systems which were designed assuming gravity was aligned with the axis of a
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machine, this can lead to additional, unexpected wear. Maintenance in these installations is extremely difficult and expensive. This leads to requirements that there be no regularly scheduled maintenance for long periods, and that the lifetime be long. Cooler weight will clearly be important in towertop applications. Maximum receiver system weights of less than 40 kg can pose severe constraints on the design, particularly for passively-cooled systems which can have 20 kg or more in the heat sinks.
Noise. Maximum allowable noise levels may be well-quantified (as for standards from Bellcore and ETSI), or it may be highly subjective. Typical noise limits are on the order of 70 dB
for outdoor applications. A typical 100 CFM fan will have a noise level of 60 to 70 dB. High velocities at intake or exhaust, or blockage close to the fan inlet can increase this. Designers
must consider noise level in the selection of components, the design of air flow passages, mounting details and possible noise absorbing materials. HEAT REJECTION
On water cooled cryocoolers, heat rejection systems are a minor part of the design. However, for industrial applications where the heat is rejected to ambient air, the heat rejection system can be a driving force in the cryocooler design. The design of the heat rejection system typically involves a tradeoff between performance and cost, size and weight, with the designer attempting to find the least expensive and lowest weight system that will meet the required
performance goals. A number of potential limits exist which must be considered in the design. These limits include:
• cryocooler thermal stability, • compressor temperature limit, • input power limits, and • temperature limits on associated components (e.g. electronics). Each of these mechanisms will be discussed below. In the design of a cryocooler for an industrial application, each of these issues must be considered, and the most restrictive becomes the design constraint. Thermal Stability
Ross and Johnson1 showed that any cryocooler with a finite thermal resistance between the cryocooler heat sink and the ultimate heat sink has the potential for unstable operation. Unstable operation occurs when an attempt to service a higher heat load at a fixed temperature leads to such an increase in the heat rejection temperature that further increases in input power are required, causing the input power to increase without limit. Ross and Johnson described this effect in the context of a fixed thermal resistance between the heat sink and the ambient, and investigated the limits on the cryocooler operating region. Here, we consider the inverse problem: given a desired operating point and cryocooler performance characteristics, what heat rejection system performance is required to ensure stable operation? The reject, or heat sink, temperature can be determined by the equation (1) where is the reject temperature at the warm end of the cooler, is the ambient temperature, P is the input power that is being removed at the heat rejection system and R is the thermal resistance of the heat rejection system. This simple linear model is applicable for both natural and forced convection cooling over small ranges around the design point, where the heat sink temperature is linearly dependent on the power dissipation and the heat sink thermal resistance. Ross and Johnson presented this algorithm for calculating the input power required as a function of heat sink temperature
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where P is again the input power level, evaluated at a reject temperature of for a load temperature of is the baseline input power level evaluated at the baseline reject temperature, and is the proportionality constant between the heat sink temperature and the coldend temperature. Given the input power required as a function of load temperature for a baseline heat load,
and baseline reject temperature, the effect of thermal resistance on the input power can be determined. As an example, consider the input power function
Substitution of Eq. (1) into Eq. (2), and evaluation in Eq. (3) yields an equation for the input power as a function of the heat rejection system’s thermal resistance
where For data presented in Ross and Johnson for the 80K STC Stirling cooler, approximate values of the constants in Eq. (3) are: and For this cooler, the value of is 3.5°C/K. Figure 1 shows the input power required as a function of ambient temperature for this case. Figure 1 also contains the calculated input power for two additional cases – the STC
80K cooler in an ambient temperature of 60°C, and a TRW 3505 pulse tube cooler with The input power at first rises slowly with thermal resistance. Then, it begins to rise very quickly, and at some point, the system becomes unstable, and adding additional input power does not maintain the required cooling because the heat sink temperature is rising too quickly. For the STC 80 K cooler in a 60°C ambient, (filled squares in Figure 1), this limit occurs at approximately 0.6 °C/W thermal resistance. Stable operation is not possible for any heat rejection system with a higher thermal resistance. For higher ambient temperatures, this region of instability occurs at lower thermal resistances. Instability occurs at higher thermal resistances for coolers whose cold end temperature is less sensitive to the reject temperature.
Figure 1. Calculated input power as a function of heat sink thermal resistance.
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Compressor Temperature Limit For most cryocoolers, the reject temperature is essentially the same as the aftercooler temperature. For any design with motor magnets immersed in the working fluid, the magnets will be at the same temperature as the aftercooler, if not higher. The magnets’ Curie temperature sets an upper bound for the operating temperature of the motor, and hence the aftercooler. For systems which provide cooling for the motor separately from, or in addition to the cooling for the aftercooler, the limits for the two heat rejection systems must be considered independently. Input Power Limits
As Figure 1 showed, increasing the reject temperature increases the amount of input power required to achieve the desired cold end load and temperature. If the input power is limited, an analysis such as Figure 1 can be used to determine the maximum thermal resistance that will still allow the cooler to achieve its design cooling load. The input power can be limited either by cooler design limitations, such as stroke or electronics, or it can be limited by the cryocooler user. For example, customers for towertop telecommunications cryocoolers need to minimize the total power draw to maximize back-up battery life. Other Components In many industrial applications, the cryocooler will be in a common housing with other
components. These other components may or may not have heat to be dissipated in addition to the heat being dissipated by the cryocooler. Regardless, because they are in the same enclosure, some of the heat from the cryocooler will heat the other components. If the other components are electronics, there are two issues that must be addressed. The first is that the rated temperature of the electronics should not be exceeded. The second issue is that at elevated temperatures, the lifetime of electronics drops rapidly. A system level trade must be conducted to determine the appropriate resolution between the cost of high temperature electronics and the cost (and size, weight, etc.) of a lower thermal resistance heat rejection system. SUMMARY
For industrial applications, cryocooler designs will be driven be new constraints related to the installation site. These constraints will include environmental effects such as temperature extremes, humidity, precipitation and lightning. In addition, industrial environments can not be counted on to provide the clean locations typical of most current cryocooler installations.
Cryocooler designs will need to be modified to address issues such as airborne particulates and bugs that can block air inlets and rodents that can damage electrical wiring. One of the major differences between most current cryocooler installations and future industrial installations is the requirement for heat rejection directly to ambient air which may be at an elevated temperature. The potential for elevated reject temperatures will drive many design choices, and must be included in the design effort from the earliest stages. ACKNOWLEDGEMENTS
This work was funded in part by Superconducting Core Technology, Inc., Golden, Colorado. REFERENCES
1.
Ross, R.G., Jr. and Johnson, D.L., “Effect of Heat Rejection Conditions on Cryocooler Operational Stability”; 1997 Cryogenic Engineering Conference, Portland, OR, July 28- August 1, 1997.
Survey of Cryocoolers for Electronic Applications (C-SEA) Jill L. Bruning and Roberta Torrison
Nichols Research Corporation Albuquerque, New Mexico USA Ray Radebaugh
National Institute of Standards and Technology Boulder, CO USA Marty Nisenoff
Naval Research Laboratory Washington, D.C. USA ABSTRACT
Significant advances in cryocoolers over the last decade have allowed cooling technology to become capable of supporting longer lifetime applications, with the potential of very low cost. A survey of cryocooler technology has been funded by the Defense Research Projects Agency (DARPA) through the Naval Research Laboratory (NRL), to update, expand, and disseminate the current capability, availability, and cost of cryogenic coolers. This paper presents an initial summary of the data collected, and reviews the trends and analysis of prior surveys performed in the 1960’s, 1970’s and 1980’s1-8. This cryocooler survey included contacts in the US, Europe, China, and Japan. The scope of the survey was coolers for sensor, superconducting, and cold CMOSS applications of interest to both military and commercial markets. Life test data, production capabilities, and cost summaries are presented in addition to the traditional performance parameters of temperature, capacity, efficiency, power, weight, and others. Initial results show improved trends in efficiency. The data and analysis have been compiled on electronic media and the survey with associated browsing and searches will initially be available to parties who participated. INTRODUCTION
The uses for cryogenic cooling have increased significantly over the last decade. The advent of high temperature superconducting (HTS) microwave and wireless applications, cold electronics and processing, commercial and medical thermal imagery, and remote surveillance have been added to the traditional markets of cryopumping and military coolers for infrared sensors and low temperature superconducting electronics. Expeditious insertion of new cold electronics technologies into cost effective devices hinges on the ability to provide reliable, low cost cryogenic cooling.
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The cryocooler industry has responded to these expanding requirements with a variety of products, research and development, and low cost manufacturing techniques. Over the last 15 years, significant advances in cryocooler technology have been made with the development of flexure Stirlings, pulse tubes, linear drive motors, and novel refrigerators. DARPA/NRL has furthered technology development through programs focused on low cost (less than $1000/unit for production quantities of 10,000) and long life (3 years of continuous operation)6. Potential commercial suppliers have recognized cryocoolers as a critical enabling technology. In order to successfully “match” the most promising cryocooler to an HTS filter, a chilled CMOSS computer, or other application, the most up-to-date, comprehensive information on refrigeration technology must be available to potential users. The C-SEA project is a national and international cryocooler survey of vendors and developers, government research facilities, and universities. The goal is to provide a collective database of information on available cryocooler technology. APPROACH The approach to the cryocooler survey was to first collect and review previous survey data1-8. From this review a cooler data collection format was developed and reviewed with the NRL customer. The survey is not only an update of previous survey information, but also more expansive than previous collections. The parametric bounds for the survey are refrigeration temperatures from 4 to 240 degrees Kelvin and cooling capacities from 0 to 500 watts. The
lower temperatures were included to allow comparison between low temperature superconducting (LTS) and HTS devices in terms of cooling requirements and availability. The higher temperatures and capacities will provide data for cooled CMOSS device options. The technical parameters surveyed were: organization, point of contact, model, type/cycle, design point capacity and temperature, input power, volume, weight, cooldown time, application, production capacity, lifetime, and cost. Nichols Research generated a list of cryocooler vendors for the European and United States collections. Dr. Radebaugh of the NIST generated the contact list for the Japan and China contacts. The team then began collecting data through a series of telephone calls, faxes, emails, meetings, conferences, and internet information resources. Dr. Radebaugh visited both Japan and China to collect data from vendors and researchers in those countries. An electronic database was generated in Microsoft Access, which was selected based on its compatibility with Microsoft Office and the ability to save the database files in html format so that the database could be easily used on a web site. The collection format report is shown in Figure 1.
Figure 1. The above data points were requested from each cryocooler manufactured for each
vendor.
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RESULTS As the cryocooler specifications were received from the vendors, they were input into two database tables: one for point of contact information (company name, address, phone number,
etc.), and one for specific cryocooler information, as listed in figure 1. The data were normalized into common units (e.g., kg, W, K) to facilitate comparisons. The tables were linked on model number/name, as that was the one field that was distinct for each cryocooler. In addition to the two tables, a primary data screen was designed to lead the user into the data for the coolers. From this screen the user can view either the cryocooler specifications or the overall graphs that have been generated (shown further in this paper). After selecting the cryocooler specifications form, the user is given the option of viewing either corporation or cryocooler information for any model of cryocooler for which data exists in the database. There are a limited number of predefined queries that support the forms in the database. Microsoft Access users can generate their own queries to view the data. Future versions of the database will allow easy formation of ad hoc queries so users can select additional data of interest. Four graphs were plotted using the data collected from the survey, and are shown in the following section. Refrigeration vs. Temperature Figure 2 shows data points for all the coolers of all vendors surveyed. Clearly the greatest number of available coolers is in the 80 K temperature regime and at cooling capacities less than 100 W. This data represents 5 times more data points than most recent Robinson survey of 19847. It appears from the graph, one can obtain any capacity between .1 W and 300 W at 80 K.
Figure 2. Capacity versus Temperature plot shows the majority of the available coolers are in the 80 K.
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Capacity vs. Weight For each cryocooler, the design point capacities were plotted against the cooler’s total weight (in kilograms) for coolers in the 30 to 90 K temperature range, as shown in figure 3. This temperature range was selected in order to compare our results with those from the Strobridge ’69 paper1. For additionally robust comparison, we included the data from the Robinson paper, which was from 1977-19807. When looking at coolers with small capacities, the weight has increased from the 1960s through the 1970s, to the 1990s. This is probably due to the use of linear drive motors, which are generally heavier than rotary drive motors and have been incorporated over the last 15 years in order to increase cooler lifetime. This hypothesis cannot be verified with the data currently in the database, as we obtained very little data about cooler lifetime. The survey form requested respondents to include the weights of the compressor, expander, and electronics. No respondents itemized electronics weight. Where weights were given for both the compressor and the expander, the total of these weights was used. Where only one of the weights was given, that weight was used as the total weight. Cryocoolers for which no weights were given are not represented in the graph. The graph does not take into account differences among cryocoolers (e.g., space applications vs. vacuum systems) nor where only either the compressor or the expander is included in the weight total. In order to access the weight trends, further clarification from the vendors is needed.
Figure 3. This graph shows the changes over time in the capacity vs weight trends of cryocoolers.
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Input Power vs. Cost The input power at design point capacity was plotted against the cost of each of the cryocoolers, as shown in Figure 4. There are few data points on this graph, as very few of the respondents provided cost information, which was primarily from vendors that have large volume manufacturing. This graph can be considered for high volume production cryocoolers only. As can be seen from comparing this data with the Strobridge1 line, costs have come down considerably in the lower capacity (< 10 kW) coolers. As more cost data comes in for the higher capacity coolers, it is anticipated that there will be a change in the slope of the Bruning line, with a resultant crossover with the Strobridge data. Capacity vs. % Carnot
Figure 5 shows the design point capacity plotted against the percent Carnot efficiency for the data collected during our survey, and contrasts this data against the data presented in the
1968 Strobridge paper3. The data from our survey is not separated by temperature and thus
shows up fairly scattered. The following equations show how the % Carnot was calculated
= Input Power (watts) Q = Design Point Capacity (watts) = Heat Rejection Temperature (Kelvin) T = Design Point Temperature (Kelvin) FURTHER WORK
More data needs to be collected. Some vendors provided only partial data, others didn’t respond in time for this paper. The alpha version of the database will be distributed to all participating vendors and comments will be collected and, where appropriated, changes will be made to the database data fields, report and form formats, and additional queries generated. The final database will be distributed on a CD ROM and made web enabled so the data can be accessed via the world wide web. Ideally, the database will be updated annually, with changes submitted by vendors. CONCLUSIONS The cryocooler electronic data base provides a valuable resource for searching a specific cooling requirement, comparing parameters, or finding contact sources and information. Initial data analysis suggests additional data gathering is required for cost and at high cooling capacities. Clarification of weight data (i.e., what is included or not included) in the reported weights will further define the weight comparisons. Additional lifetime data will further enhance cooler mass data analysis.
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Figure 4. The above graph shows that costs have decreased since the 60s.
Figure 5. The 90s data are added to the data from the 60s and 70s for comparison.
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There are considerably greater numbers of smaller capacity (less than 1 watt) coolers than any previous survey and these are more efficient. The Bruning trend line does not include a lot of data at very high capacities and that is probably why there is a cross-over with the Strobridge efficiency trend. Further effort will focus on more analysis of the data and increased collections where there are limited or questionable results. ACKNOWLEDGMENT
This work was funded by the Research Projects Agency (ARPA) through the Naval Research Laboratory (NRL). Dr. Marty Nisenoff’s direction and contribution have been invaluable. The authors also wish to acknowledge and gratefully thank the vast number of industry participants who have not only generously provided data, but continually gave valuable suggestions and encouragement for this project. REFERENCES
1. Strobridge, T.R., “Refrigeration for Superconducting and Cryogenic Systems,” IEEE Transaction, Nuclear Science NS-16 #3 (1969), pp. 1104-1108. 2. Bridwell, M.C. and J.G. Rodes, “History of the Modern Cryopump,” Journal of Vacuum Science Technology, A3(3), May/June 1985, pp. 472-475. 3. Strobridge, T.R. and D.B. Chelton, “Size and Power Requirements of 4.2K Refrigerators,” Advances in Cryogenic Engineering, Volume 12, (1966), pp. 576 - 584. 4. Byrns R.A. and M.A. Green, “An Update on Estimating the Cost of Cryogenic
Refrigeration,” LBNL-40188, presented at CEC, (1997). 5. Strobridge, T.R. and R.O. Voth, “Refrigeration Technology for Superconductors,” IEEE Transactions on Nuclear Science, Volumes NS-24, NO. 3, June 1977, pp. 1222-1226. 6. Electronic Industries Association, Proceedings of M-CALC II, San Diego, CA, 15-16 January 1998. 7. Robinson, George, J. L. Smith, Jr., and Y. Iwasa, “MIT survey 1984.” 8. Jensen, H.L., T.C. Nast, A.P.M. Glassford, R.M. Vernon, and W.F. Ekhern, “Handbook of External Refrigeration Systems for Long-Term Cryogenic Storage,” prepared for Thomas L. Davies, Manned Spacecraft Center, National Aeronautics and Space Administration, No LMSC-A984158, 22 February 1971.
Construction and Tests of a High-Tc SQUID-Based Heart Scanner Cooled by Small Stirling Cryocoolers C.J.H.A. Blom, H.J.M. ter Brake, H.J. Holland, A.P. Rijpma, and H. Rogalla University of Twente, Faculty of Applied Physics P.O. Box 217, 7500 AE Enschede, The Netherlands
ABSTRACT
A heart scanner that can be equipped with up to 25
SQUID magnetometers was
designed at the University of Twente. In this design the mechanical cooler interference is reduced by operating two coolers in counterphase. The magnetic cooler interference diminished by positioning the coolers and the SQUIDs in a coplanar arrangement, and by separating the SQUIDs from the cold tips of the cryocoolers with a solid conducting thermal interface. In the present paper the construction of this cryogen-free SQUID system is described and test results are presented. A SQUID plate temperature of 60 K was realized in about 2 hours. The corresponding heat load to the coolers is roughly 0.9 W. With a preliminary magnetic shield around the compressors a noise level of 0.6 was measured in the frequency band 10 to 100 Hz and magnetocardiograms were recorded inside a magnetically
shielded room. A further reduction of the noise level is expected after optimizing the mumetal shielding of the compressors. INTRODUCTION
Superconducting QUantum Interference Devices (SQUIDs) are the most sensitive magnetic
flux-to-voltage converting sensors. So far, their main application has been in biomagnetic research where multichannel dc-SQUID based magnetometer systems are used. These systems are usually cooled by liquid helium and operated in magnetically shielded rooms to obtain an extremely low-noise measuring environment. Therefore, they are expensive, require helium refills, and cannot be transported in a simple manner. Because of their higher operating temperature, a much more flexible magnetometer system can be realized with SQUIDs. The ceramic copper oxides, as with a critical temperature of 92 K, allowed for SQUID operation in liquid nitrogen, boiling at 77 K. For cooling the SQUIDs also small-scale turnkey cryocoolers, originally developed for cooling infrared sensors, are available and very reliable.1 The clear advantages of cryocoolers compared to liquid-nitrogen cryostats are the lower operating temperature (which gives a potentially better SQUID performance2) and the turnkey operation (no refills required). In a recent paper we have presented the design of a heart scanner that could be equipped with up to 25 SQUIDs, which was cooled by two small Stirling-type cryocoolers.3 In the present paper the actual construction and first test experiments are presented. In the next section
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Figure 1. Heart scanner design 3 (schematic), depicted with nine SQUIDs: side view and bottom view.
the design of the heart scanner is reviewed briefly. Then, the actual construction is considered followed by a section describing the test experiments. These concern cryogenic tests (cool-down
temperature registration, heat load and operating temperature) and magnetometer tests (system
noise level, magnetocardiography). The paper concludes with a discussion on the established results. DESIGN OF THE HEART SCANNER
Cryocoolers used in the Heart Scanner
Two UP 7058 Stirling-type coolers are used, manufactured by Signaal USFA (Eindhoven, the Netherlands).4 In each cooler two compressor pistons are driven by linear motors in a dualopposed arrangement. A detailed description of the cooler operation is presented elsewhere.3, 5 The cooling power available at the cold tip depends on the cooler input power, the environmental temperature and the temperature of the tip. At 55 W input power and 20 °C the cooling power depends in a linear way on the tip temperature:
Eq. (1) which is based on measurement data supplied by Signaal USFA gives 12.5 W at 300 K and 1.6 W at 80 K. Proper heat sinking of the cooler is in this respect very important. An increase in the warm-end temperature of the cold head directly translates into a lower cooling power. At 53 W input power and the cold head at 323 K we measured
3 For an input power of 55 W this means in fair approximation in the case of a 323 K warm-end temperature: (resulting in a cooling power of 10.7 W at
300 K and 1.5 W at 80 K). Cryogenic Design
The cryogenic design of the heart scanner, with thermodynamic modelling and optimization, is described in detail in our previous paper.3 In this design, schematically depicted in Fig. 1, 25 SQUID magnetometers can be mounted on a silicon plate, which is cooled by the two Signaal USFA coolers. The cooler compressors are placed 0.5 m from the SQUID magnetometers in order to limit the magnetic interference. This interference is further minimized by positioning the compressors coplanar with the SQUID magnetometers. The vibrations of the coolers are reduced by applying two coolers, mounting
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Figure 2. Photograph of the vacuum box with from left to right: the silicon SQUID plate supported by three glass spacers, the thermal link held inside a G10 tube, and the cold fingers coaxially mounted on the box and heat sinked on the warm ends. Note that in comparison with the side view of Fig. 1 the arrangement of this figure is ‘upside down’.
them in a back-to-back configuration, and operating them in counterphase. In our design, the cold tips of the coolers are thermally linked to the SQUID plate by two flexible copper leads (each with an effective cross-sectional area of and a length of 30 cm). Together with the thermal leads the SQUID plate and the cold fingers should be wrapped in about 15 layers multilayer insulation (MLI) and placed in a rectangular non-metallic vacuum box. The cold heads are coaxially mounted on the side walls of the vacuum box. A thermodynamic simulation, using twice the cooling power for a single cooler as given by Eq. (1), yielded a SQUID plate temperature below 55 K in about 50 minutes.3 CONSTRUCTION OF THE HEART SCANNER
Vacuum Box
The vacuum box was manufactured from G10 glass-fibre reinforced epoxy plate material with a thickness of 17 mm. In order to minimize the distance of the SQUIDs to the outside, the
lid that closes the box on the side of the SQUIDs (see top of Fig. 1) has a thickness of only 10 mm. Furthermore, a section of of this lid, that directly faces the silicon SQUID plate was reduced to a 5 mm thickness. To limit the bending of the lid, two extra plates of G10 were glued in the box next to the thermal link. These supporting plates are not included in Fig. 1 but can be seen in Fig. 2. Under vacuum the bending at the centre of the 5 mm thick area was 0.5 mm. Two flanges were glued into the box: one for pumping, the other for electrical feedthroughs. In order to maintain the good vacuum conditions some activated carbon acting as getter material was placed on the cold tips and on the thermal link. Cooler Heat Sinking
As mentioned before, proper heat sinking of the coolers is very important. In that respect convective cooling by means of a forced water or air flow would be very attractive. However, because of the limitations regarding transportability and noise contributions we chose to design a thermal link with passive convective cooling only. Therefore, the coolers were mounted on an aluminium plate to which also cooling fins were attached. In the design special attention was paid to the thermal Johnson noise contribution from the material of the heat sink. This should be significantly less than the SQUID noise level. Anticipating on future developments this limit was set to
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Figure 3. Schematic of the heart scanner assembly. 1 : aluminium mounting plate; 2: aluminium compressor blocks; 3 : aluminium cold-head blocks; 4: “flexible” heat link 5 : glass-epoxy vacuum box; 6: cooling fins; 7: flanges for pump and feedthroughs; 8: wooden base plate; 9: wooden support.
Figure 4. Photograph of the cooler heat sinking (see text).
The passive heat sink is depicted in Figs. 3 and 4. The compressors are clamped in aluminium blocks that are rigidly mounted on an aluminium plate of about in size and 8 mm thick. To the other side of this plate a type S245 high-power extrusion heat sink of Thermalloy6 is attached. The mounting plate is fixed to the glass-epoxy vacuum box, and this as a whole is supported by a wooden base plate (see Fig. 3). Heat sinking of the cold heads warm ends appeared to be problematic. Firstly, the cold heads cannot be rigidly mounted to the aluminium plate because they are already fixed to the vacuum box (O-ring seals). Instead a more or less flexible link had to be applied with a higher thermal resistance then a direct rigid contact. Secondly, the contact area that can be used for heat sinking is for the cold head about one fifth of that of a compressor, whereas the heat dissipated in the cold head is roughly one third of the total cooler input power (i.e. 15 to 20 W). A substantial part of this heat is caused by a surface heat pumping mechanism in the split pipe. 7 Furthermore, the heat that is absorbed at the cold tips because of the cooler operation also has to be rejected to the environment. At start up, 300 K, this is an extra heat of about 12 W per cold head. First, we applied a thermal link of copper leads but later we replaced these by aluminium strips. With the aluminium strips a better thermal contact to the aluminium plate could be realized. Per cold head 12 strips were used with a width of 24 mm and a thickness of 0.5 mm The contact area on the cold-head side was about that on the plate The length in between the contacts was 11 mm. At all interfaces alumina-filled grease was applied to improve the thermal contact. A rough thermodynamic model can be made assuming firstly, the mounting plate to be at uniform temperature and secondly, all heat rejection to the environment to take place at the cooling fins. Then, a thermal node model can be formed with the compressors, the cold-heads warm ends, the mounting plate and the environment as nodes, connected via thermal resistances. These thermal resistances, however, are quite poorly defined. Especially the contact resistances are difficult to evaluate and also the resistance between the cooling fins and the environment had to be estimated via extrapolation from the manufacturer’s data.6 The following temperatures resulted from this analysis: plate 318 K, compressors 319 K, cold heads 326 K.
Also a more sophisticated model was developed in which the direct heat exchange of the compressors and the cold heads with the environment was taken into account. Furthermore, a one-dimensional profile was considered along the aluminium plate by dividing the plate in small strips. Because the direct exchange was included the resulting temperatures were somewhat lower: compressors 314 K, cold heads 323 K. Along the plate atemperature profile resulted from 314 K at the cold-head end to 311 K at the other end. During start up with an extra 12 W per cold head these temperatures can in the worst case rise to 318 K for the compressors and 330 K for the cold heads.
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The thermal noise contribution of the heat sink material can be evaluated by considering small volume elements dV. The thermal noise current density in this element is uniformly distributed in all directions and its RMS magnitude is given by 8:
Here, T is the temperature of the material, its electrical conductivity (for aluminium at room temperature: ), the frequency bandwidth. Consider the SQUID to be placed in the x,y-plane with its normal in the z-direction (i.e. the SQUID senses the z-component of the magnetic field). Further, consider a metallic plate in the x,y-plane which extends from to Let be the area of its cross section perpendicular to the x-axis. If is small compared to then the thermal noise contribution of the plate can be evaluated by means of Eq. (2) and the Biot-Savart law as:
Based on this equation we evaluated the worst-case noise contribution of the heat-sink material
as
This is sufficiently less than the anticipated SQUID noise level of
Thermal Link The thermal link was made of two Litze cables consisting of separate twisted and noninsulated copper wires, each cable with a cross-sectional area of and a length of 30 cm.
On one end the cables were soldered each to a conically shaped copper cap that could be attached to the cold tip of a cold head. The other ends of the cables were soldered to copper plates that could be clamped to the silicon SQUID plate. At both ends copper-filled greas was applied to improve the thermal contact.
The thermal noise contribution of the copper leads can simply be evaluated by Eq. (3). With and the result is that the SQUIDs have to be at least at a distance of 5 cm from the cable ends. In our construction the distance is 5.5 cm, resulting in a noise contribution of This is sufficiently less than the noise level of the current SQUIDs but once a resolution of is available, the link has to be improved (other material e.g. sapphire). SQUIDs + SQUID plate Two SQUIDs were available from the earlier development of a 7-channel liquid-nitrogen cooled heart magnetometer.9 These SQUIDs had an average noise level of
and were packaged in alumina chip carriers with outer dimensions of Because only two SQUIDs were used in the test experiments the silicon plate could be limited in size to
(thickness 0.7 mm). The two SQUID packages were glued onto the plate by
means of GE varnish 10. The modulation coils were mounted on the reverse side. The centre of the first SQUID was on 5.5 cm from the copper thermal link, the second SQUID centre at 11.9 cm (i.e. in gradiometer operation the baseline is 6.4 cm). Suspension of Thermal Link and SQUID Plate The copper leads of the thermal link were led through a G10 tube that was fixed between
the two supporting plates of the vacuum box, as can be seen in Figs. 2 and 4. At both ends of the tube nylon wires were used for this purpose.
The position of the SQUID plate inside the vacuum box was fixed by means of glass spacers. These were machined in a tetrahedral shape, three short ones with a height of 10 mm to be used at the SQUID side of the silicon plate, and three long ones of 48 mm for the other side. All were machined with a contact area of about at the top. At the base the short ones
have acontact area of
the long ones
In order to prevent the silicon plate from
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breaking because of the spacer forces, small G10 strips were glued onto the plate for distributing
the forces over it. Furthermore, small disks of silicon rubber with a thickness of 1.2 mm were placed between the spacers and the G10 box to allow for some spacer flexibility. Multilayer Insulation
As described in Ref. 3 the radiative heat load to the SQUID plate, the thermal link and the cold fingers is reduced by 15 layers of aluminized mylar foil, MLI type NRC-2.10 Because it was not possible to wrap the complete cryogenic part with a single sheet, we used 15 separate pieces that were folded one over the other. First, the SQUID plate and the thermal link (including the G10 support tube) were removed from the system. Then, the 15 sheets were placed in the box such that the glass spacers for the silicon plate pricked through. Next, the G10 support tube was placed in the box and fixed with the nylon wires that were pierced through the MLI blanket. Then, the copper leads of the thermal link were led through the tube and and attached to the cold fingers. The SQUID plate was fixed to the other end of the thermal link and was placed on the three glass spacers. After that, the MLI sheets were folded to close the MLI blanket. Finally, the lid was put on the vacuum box, at the same time piercing the three glass spacers in the lid through the MLI blanket, thus fixing the SQUID plate’s position within the box. Thermometry
The temperatures were recorded with Lake Shore diodes, specified accuracy ± 1.5 K or 1.5 %.11 Five diodes were used outside the vacuum box to monitor the temperatures of the two compressors, the two cold heads, and the plate with the cooling fins. For the cryogenic tests without the SQUIDs, six diodes were used inside the box: one on each cold tip, two on the SQUID plate (on the respective SQUID positions), and two on the copper thermal link (one close to the silicon plate, the other at about 5 cm from the cold fingers). The diodes were glued with GE varnish.10 After the cryogenic tests the SQUIDs were placed on the plate and only three diodes remained in the box: two on the cold tips and the third on the silicon plate, centred between the SQUIDs. Wiring
Each SQUID needs 8 connecting lines, two of which for the signal output. Except for these signal lines they can be high-ohmic. We, therefore, used twisted pairs of manganin wire. The signal lines were made with twisted pairs of 0.15 mm diameter copper wire. To reduce the conductive heat load through the wires, 30 cm of length was used between the feedthrough flange on the vacuum box and the cold tips where the wires were thermally anchored. This extra length of wire was wound around a kapton cylinder in the vacuum space. From the cold tips to the SQUIDs the lines each have a length of about 50 cm. The temperature diodes are connected in series and a current of is supplied via manganin wire. The voltages over the diodes are also measured via twisted pairs of manganin wire. Heat Loads
We consider two heat load terms. Firstly, the heat load via the vacuum space and the MLI blanket and secondly, the conductive heat flow through the spacers and the wires. The first is the
dominant term but also the most difficult to evaluate. In our model considerations we have used
an expression for an optimum layer spacing (i.e. 30 layers per cm) and no openings.3, 12 The resulting ideal-case heat flow through the MLI blanket in our construction would be 155 mW. However, the MLI sheets are not equally spaced. Especially near the spacers the sheets will be more or less pressed together. Furthermore, we have used 15 separate sheets that were folded to
close the blanket. That means a lot of openings resulting in a higher heat flow. Therefore, also a worst-case limit is considered. Because the emissivity of the MLI blanket is small
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compared to that of the G10 box, the heat flow from the box to the MLI can be expressed as
where is the outside surface of the MLI blanket. At the outside this blanket is at a temperature is the temperature of the G10 box. In the worst-case the MLI layers have so much thermal contact that and Eq. (4) reduces to
Because of the contacts near the spacers this worst-case approximation seems more realistic than the ideal case. In our construction With and the resulting worst-case heat flow is 0.79 W (a factor 5 more than in the ideal case). The heat flow through the spacers and the wires is less important. Neglecting the thermal resistances at the contact interfaces, we estimated this flow to be 49 mW (with the SQUIDs installed 68 mW). These flows add up to a heat load of 0.22 W in the ideal MLI case and of 0.86 W in the worst MLI case. These heat loads, combined with the cooling power that corresponds to a coldhead warm end temperature of 323 K, give the following tip temperatures: ideal case 48.6 K, worst case 68.7 K. The temperature increase along the thermal link from the cold tips to the SQUIDs can be evaluated by assuming the total heat load to be equally distributed over the length (L) of the link. In that case the temperature difference is given by
Here A is the cross-sectional area of the link and its thermal conductivity. For the first SQUID position on the silicon plate at 5.5 cm from the copper leads is evaluated as 11 K/W. Eq. (6) then gives the following SQUID temperatures: ideal case 49.8 K, worst case 68.7 K. TEST EXPERIMENTS
Vacuum Test An important parameter in the heat flow through an MLI blanket is the residual gas pressure. Thermal conduction via the gas becomes a relevant heat flow contribution when the pressure in the vacuum space rises above about 10 mPa 13. In our tests we pumped the vacuum box to 3 mPa. After the pump was switched off the pressure quite rapidly increased because of outgassing: a typical rate was an increase to 1 Pa in 45 minutes. We, therefore, glued some activated carbon on the cold fingers and on the thermal link. The carbon acts as a getter and the low pressure could now be maintained. After switching off the pump, the pressure rose to 40 mPa. At that time the coolers were started and the pressure dropped to 1.8 mPa. Cryogenic Test A registration of the temperatures during a cool-down after the carbon getter had been installed is depicted in Fig. 5. Unfortunately, after the thermal cycling in the experiments without carbon getter, a few temperature diodes got loose and gave unreliable temperature indications. In Fig. 5 the remaining temperatures are plotted. At min the coolers were started and their inputs were tuned up to 55 W in the following 3 minutes. At min the SQUID plate had cooled to 63.9 K with tip 2 at 59.1 K. After about 90 minutes of cooling the temperature of the SQUID plate decreased less than 0.1 K per minute. During this experiment the temperatures of the cold-heads warm ends rose to 330 K and 334 K (respectively cold head 1 and 2). At min a fan was switched on to actively cool the warm end of cold head 2. As a result the warm end temperature decreased by 9 K to 325 K. The SQUID plate cooled down further to 59.4 K at min, with tip 2 at 55.9 K. Taking into account the temperature measuring accuracy and the cooling power for a warm-end temperature of 323 K , the tip
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Figure 5. Cool-down temperature registration. Upper curves from top to bottom: solid: cold head 2, warm end; dashed: cold head 1, warm end; solid: compressors (1 and 2); dashed: heat-sink mounting plate; Lower curves from top to bottom: solid: SQUID plate at 11.9 cm from link; dashed: SQUID plate at 5.5 cm from link; solid: thermal link; dashed: cold head 2, cold tip.
temperature of 55.9 K corresponds to a heat load of Following Eq. (6) this results in a SQUID temperature of which corresponds to the measurement results. A comparison of these data with the values given in the section on heat loads indicates that we are about ‘half-way’ the worst MLI case and the ideal case. The longer cool-down time than expected is not primarily caused by the larger heat load,
because this mainly affects the end temperature. We expect the longer cool-down to be caused by extra heat capacities present in the vacuum space that also have to be cooled over some limited thermal resistance. For instance, the G10 guiding tube around the copper leads has to cool down to 60 K and via its thermal contact to these leads an extra thermal time constant is introduced in the thermodynamic behaviour. SQUID Magnetometer Experiments
After the SQUIDs were mounted, the system was installed in the magnetically shielded room of the Biomagnetic Centre Twente. Unfortunately, one of the SQUIDs that we used exhibited a continuous loss of performance in time. Quite steadily the critical current had decreased and so had the required operating temperature. During our experiments the operating temperature of this SQUID was so low (< 68 K) that it could not be operated simultaneously
with the other SQUID that functioned properly at higher temperatures. Therefore, the test experiments were performed with the latter SQUID only. First, the cooler interference at the operating frequency of 50 Hz was measured to be This is significantly higher than the design limit of Therefore, further suppression of the 50 Hz peak is required . Second, the noise spectrum was determined. Here, we had to apply an analog 50 Hz notch filter because we were limited by the dynamic range of the 12-bits A/D converter. The resulting spectrum is depicted in Fig. 6a. Except for the remaining noise peaks resulting from the 50 Hz interference, the noise level from 10 to 100 Hz is As the coolers are switched off, the noise level dropped to about (Fig. 6b). Because of the temperature drift in this case, the noise level is higher at lower frequencies. At 10 Hz the noise level is in this experiment. The white noise level of this specific SQUID in liquid nitrogen was measured in the 7-channel heart scanner 9 as
HIGH-Tc SQUID-BASED HEART SCANNER
Figure 6. Field noise spectra measured by SQUID: a: no magnetic shielding; b: magnetic shielding around compressors; c: coolers off.
845
Figure 7. Magnetocardiogram recorded with ‘cryogen-free’ SQUID; 0.1-100 Hz 50 Hz notch; inset: schematic MCG peaks.
The higher level in our experiment can be explained by a lower operating
temperature and, therefore, a lower effective measuring area of the SQUID.14
The higher white noise level with the coolers running is for the largest part due to noise in the compressor currents. These currents are generated by computer and set to adequate powers by means of audio amplifiers.15 In these circuits apparently a small thermal noise contribution is added to the 50 Hz power signal. Whereas the 50 Hz currents for the four different compressor coils are correlated and tuned in such a way that they counteract,3 this is not the case for the thermal noise. The thermal noise contributions are not correlated and, therefore, the noise powers add up. By means of an open box of mumetal foil covering the compressors as a preliminary magnetic shield, we were able to reduce the noise by a factor of 7 to a level of 0.6 see Fig. 6c. Because of the magnetic shielding, the 50 Hz noise peak also was reduced by one order of magnitude. In this set-up we could record a magnetocardiogram of an adult as depicted in Fig. 7. For this purpose we used a 0.1 Hz high-pass filter and a 100-Hz low-pass. The 50 Hz interference was further reduced by digital filtering (49.4-50.6 Hz). In the MCG recording
all relevant peaks can be distinguished: the P-wave and the QRS-complex respectively due to the depolarization of the atria and that of the ventricles, and the T-wave resulting from the repolarization of the ventricles.16 Despite the successful recording of the MCG we aim at a further suppression of the white noise. This can be established by means of cylindrical mumetal shields that completely cover the compressors. In previous experiments we have measured a shielding performance of such shields of a factor 40.17 However, for this purpose the aluminium compressor blocks of the heat sink will have to be adapted. CONCLUSION
A cryogen-free SQUID-based heart magnetometer was built and successfully tested. A SQUID plate temperature of 60 K was realized in about 2 hours. With a SQUID a noise level of was measured in the frequency band 10 to 100 Hz and preliminary magnetocardiograms were recorded inside a magnetically shielded room. In the near future we will apply optimized mumetal shielding of the compressors in order to reduce the system noise level to that of the SQUIDs (roughly Furthermore, the SQUID with the lower operating temperature will be replaced by a new SQUID. Then, a first-order gradiometer can be
formed electronically, and the system can be tested outside the magnetically shielded room.
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ACKNOWLEDGMENTS
We are indebted to P.J. van den Bosch, H.A. de Boer, and W.A.M. Aarnink for their contributions in the cryogenic design of the system. Also, J.F.C. Verberne is acknowledged for his contribution on the cooler noise suppression, and H. W.M. Veldhuis for the realization of the glass-epoxy vacuum box. Further, we thank Signaal USFA for the support and the fruitful contributions in the discussions: a special word of thanks in this respect to A.A.J. Benschop and P.C. Bruins. The research was also financially supported by the Dutch Research Program on Superconductors REFERENCES 1.
Walker, G., Miniature Refrigerators for Cryogenic Sensors and Cold Electronics: Monographs on Cryogenics, vol. 6, ed. R.G. Scurlock, Clarendon Press, Oxford, UK (1989).
2.
Rogalla, H., “Superconducting electronics”, Cryogenics, vol. 34 (ICEC suppl.) (1994), pp. 25-30.
3.
van den Bosch, P.J., ter Brake, H.J.M., Holland, H.J., de Boer, H.A., Verberne, J.F.C., and Rogalla, H., “Cryogenic design of a SQUID-based heart scanner cooled by small Stirling cryocoolers”, Cryogenics, vol. 37 (1997), pp. 139-151.
4.
Signaal USFA, P.O.Box 6034, 5600 HA, Eindhoven, The Netherlands (Tel: 31-40-2503603)
5.
Verbeek, D., Helmonds, H., and Roos, P., “Performance of the Signaal USFA Stirling cooling engines”, Proc. 7th Int. Cryocooler Conf., Santa Fe, 1992, (1993), pp. 728-737.
6.
Thermalloy, Inc. P.O. Box 810839, Dallas TX 75381-0839 USA.
7.
Benschop, A.A.J., v. Wordragen, F.C. and Bruins, P.C., “Reduction of surface heat pumping effect in split-Stirling cryocoolers”, Cryocoolers 9, Plenum Press, New York (1997), pp. 139-146.
8.
Bakker, C.J. and Heller, G., “On the Brownian motion in electric resistances”, Physica, vol. 6 (1939), pp. 262-274.
9.
ter Brake H.J.M., Karunanithi, R., Holland, H.J., Flokstra, J., Veldhuis, D., Vargas, L., Hilgenkamp, J.W.M., Jaszczuk, W., Janssen, N., Roesthuis, F.J.G., and Rogalla, H., “A seven-channel SQUID-based heart scanner”, Meas. Sci. Technol., vol. 8 (1997), pp. 927-931.
10. GE low-temperature varnish and superinsulation NRC-2 were supplied by Oxford Instruments Ltd.,
Eynsham, Oxford OX8 1TL, England 11.
Silicon diode temperature sensors (types DT-470/471-LR) and manganin high-resistance wire (type MW-36) were supplied by Lake Shore Cryotronics, Inc., Westerville, OH, USA.
12. ter Brake, H.J.M. and Flokstra, J., “Computer Aided Cryostat Design in Twente: recent
developments”, Proc. 12th Int. Cryogenic Eng. Conf., eds. R.G. Scurlock and C.A. Bailey, Butterworth, Guildford, UK (1988), pp. 88-92.
13. Scurlock, R.G. and Saull, B. “Development of multilayer insulations with thermal conductivities below ” Cryogenics, vol. 16 (1976), pp. 303-311. 14. ter Brake, H.J.M., Aarnink, W.A.M., van den Bosch, P.J., Hilgenkamp, J.W.M., Flokstra, J. and
Rogalla, H., “Temperature dependence of the effective sensing area of Supercond. Sci. Technol., vol. 10 (1997), pp. 512-515.
dc SQUIDs”
15. Rijpma, A.P., Verberne, J.F.C., Witbreuk, E.H.R., and ter Brake, H.J.M., “Vibration reduction in a set-up of two split type Stirling cryocoolers”, Cryocoolers 9, Plenum Press, New York (1997) 727-
736.
16. Erné, S.N. and Lehmann, J., “Magnetocardiography, an introduction”, in SQUID Sensors: Fundamentals, Fabrication and Applications, Kluwer Acad. Publ. (1996) pp.395-412. 17. ter Brake, H.J.M., van den Bosch, P.J. and Holland, H.J. “Magnetic noise of small Stirling coolers”, Adv. In Cryog. Eng., vol. 39 (1994), pp. 1287-1295.
Cryocooler Applications for High-Temperature Superconductor Magnetic Bearings R. C. Niemann and J. R. Hull Energy Technology Division Argonne National Laboratory Argonne, Illinois USA 60439
ABSTRACT
The efficiency and stability of rotational magnetic suspension systems are enhanced by the use of high-temperature superconductor (HTS) magnetic bearings. Fundamental aspects of the HTS magnetic bearings and rotational magnetic suspension are presented. HTS bearings can be cooled by liquid cryogen bath immersion or by direct conduction, and thus there are various applications and integration issues for cryocoolers. Among the numerous cryocooler aspects to be considered are installation, operating temperature, losses, and vacuum pumping. HTS MAGNETIC BEARINGS
Stable Levitation
In its simplest form, a superconducting bearing comprises a permanent magnet levitated in a stable position over a superconductor.1 A levitation force develops because of the superconductor's tendency to exclude magnetic flux (the Meissner effect), making the superconductors behave like a strong diamagnet. Accordingly, a superconductor with a permanent magnet positioned close above it develops a shielding current, which excludes the flux in such a way that the actual magnet "sees" its mirror image as shown in Fig. 1.
Figure 1. Levitation by the Meissner effect. Cryocoolers 10, edited by R. G. Ross, Jr.
Kluwer Academic/Plenum Publishers, 1999
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More specifically, if a permanent magnet is vertically magnetized with its north pole down, the image will also be vertically magnetized but with the north pole up, exerting a repulsive force on the real magnet. The closer the magnet is positioned to the superconductor, the stronger the repulsive force; the farther away the magnet moves, the weaker the force. This arrangement can yield a levitation that is vertically stable. Horizontal stability is attained by flux pinning. A flux-pinning center is an inclusion, crack, or other crystalline defect in the superconductor. Because the superconducting region surrounding the nonsuperconducting center is strongly inclined to exclude magnetic flux, a flux line through the center tends to become trapped there as shown in Fig. 2. When enough flux lines are trapped in this way, a permanent magnet will remain levitated in position, even over a flat surface. The flux lines between the permanent magnet and the superconductor surface behave rather like mechanical springs with attachments to magnet and superconductor. If the magnet is moved up and down or sideways, it will tend to be pulled by the "springs" back to its equilibrium position.
Bearing Details In superconducting wire applications, the current must pass from grain to grain over a relatively long distance, but in superconducting bearings the current needs to circulate only within individual grains. The present material of choice for superconducting bearings that
operate at liquid nitrogen temperatures is yttrium-barium-copper-oxide (YBCO). Grains of this material can grow to diameters of several centimeters when made by a process called melttexturing. In the present state of the art, the upper size limit to melt-texturing appears to be about a 10cm diameter. The levitation force that the superconductor can provide is proportional to its average magnetization, which is proportional to the product of its grain diameter times its current density. The critical current density in these HTS samples is which together with a diameter of several centimeters, allows levitation pressures between the superconductor and a neodymium-iron-boron permanent magnet to be as high as In practical bearings, the low levitation pressure available in the interaction between permanent magnet and superconductor is often augmented by various hybrid schemes in which interactions between pairs of permanent magnets provide the bulk of the levitation force. These interactions are unstable, but the inclusion of a properly designed HTS component is sufficient to stabilize the complete bearing.
Figure 2. Horizontal stability by flux pinning.
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THE BEARING SYSTEM
The bearing system considered is shown schematically in Fig. 3. The system consists of the (1) suspended mass to which is affixed (2) a permanent magnet that is levitated by (3) an HTS magnet that is connected by (4) supports to the (5) vacuum vessel. Details of the suspended mass vary with the specific applications and operational access may be required along the bearing axis. Correspondingly, the support and vacuum vessel details vary with each application. The HTS bearings are cooled to their operating temperature by an external refrigeration source. For purposes of the quantitative discussions (which follow) of cryocooler integration issues, the following representative HTS bearing parameters are used: inside diameter = 10.2 cm (4 in.), outside diameter = 22.9 cm (9 in.), thickness = 1.27 cm (0.5 in.), and mass = 1450 g. The HTS material is melt-textured YBCO. LIQUID BATH COOLING
The HTS bearings can be cooled by immersion in a cryogenic liquid bath. Advantages
(1) (2) (3) (4) (5)
The advantages of liquid bath cooling include the following: Effective heat transfer by direct contact with the HTS material. Thermal reserve to accommodate upset conditions. Operation with cryocooler recondenser and/or with external bulk liquid cryogen source. Potentially low vibration input. Optional nonmetallic bath containment.
(6) Optional common refrigeration source for both bearings. Disadvantages
(1) Need for liquid inventory that must be initially developed and then maintained. Associated details include (a) level monitoring/control, (b) upset response, and (c) pressure vessel safety. (2) Need for vacuum-leak-tight cryogen bath container and associated connections.
Figure 3. Schematic diagram of bearing system. (1) Suspended mass, (2) permanent magnet,
(3) HTS magnet, (4) support, and (5) vacuum enclosure.
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Representative System
A representative liquid bath cooling system that incorporates a cryocooler is shown schematically in Fig. 4. The system consists of the (1) HTS bearing magnet and (2) a liquid bath reservoir fed by a (3) liquid supply reservoir. All reservoirs and their interconnecting piping are located in the (4) system vacuum enclosure. More than one liquid bath reservoir can be fed by a common liquid supply reservoir. By locating the supply reservoir above the bearing reservoir, and by piping the bearing reservoir vent gas stream into the supply reservoir gas space, a quasi-thermal siphon is possible. For circulation to exist,
where supply column fluid density, return column fluid density, g = acceleration due to gravity, fluid head length, and pressure drop around loop. The supply system can be operated as an essentially closed system by incorporating a cryocooler-driven condenser. The condensing surface can be located in the gas space of the
supply reservoir. Depending on the capacity of the cryocooler and the system losses, the system could operate in the subcooled mode, which would allow operation at temperatures below the normal boiling point of the cryogen employed. CONDUCTION COOLING The HTS bearings can be cooled by solid conduction to a refrigeration source.
Figure 4. Schematic diagram of liquid bath cooling system. (1) HTS magnet, (2) bearing liquid bath reservoir, (3) common liquid supply reservoir, and (4) vacuum enclosure.
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Advantages
(1) No need for HTS bearing cooling liquid inventory and associated operational details. (2) Operation with a cold plate that is cooled by a cryocooler cold head. Cryocooler operation results in stand-alone, plug-in operation. (3) Gravity is not a factor in cooling. Disadvantages
(1) (2) (3) (4)
Potentially poor heat transfer with the HTS material. Limited thermal reserve to accommodate upset conditions. Potential vibration input. Potential metallic components in the bearing region, which could contribute operating losses.
Representative System A representative solid conduction cooling system that incorporates a cryocooler is shown schematically in Fig. 5. The system consists of the (1) HTS bearing, which is cooled via (2) a heat transfer connection to the (3) cryocooler cold head. The cold head is attached to the
(4) vacuum enclosure. CRYOCOOLER SYSTEM REQUIREMENTS
Operational Availability The HTS bearings will compete with conventional bearings, either active or passive, that are
generally very reliable and require only infrequent maintenance at manufacturer-recommended intervals. For the conventional bearings, unscheduled shut-down for maintenance is rare. The
HTS bearings, on the other hand, require cryogenic cooling. The maintenance schedule of the cryocooler should match the parent system's maintenance schedule.
Cooling Capacity The required cooling capacity of the cryocooler depends on the design of the HTS bearing system and the cooling method employed.
Figure 5. Schematic diagram of solid conduction cooling. (l) HTS magnet, (2) heat transfer connection, (3) cryocooler cold head, and (4) vacuum enclosure.
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During cool-down from ambient temperature to a nominal operating temperature of 77 K, the total thermal energy to be removed from a single conduction-cooled HTS bearing is If the HTS bearing is liquid-bath cooled and is installed in a representative epoxy fiberglass (G-10) reservoir with a mass of 4.1 kg (9.2 1b), an additional of thermal energy must be removed during cooldown; thus, the total cooldown energy removal is per bearing assembly.
Time to cooldown should be compatible with the parent system's requirements. Transfer Losses
Transfer losses during the delivery of refrigeration from the cryocooler to the bearing substantially contribute to the overall cryogenic load for the cryocooler. These losses would be greatly reduced if an innovative integral design of a cryocooler into the bearing could be developed and implemented with part of the cryocooler system such as the cold head, or the entire unit, built into the bearing and its containment. CRYOCOOLER INTEGRATION ISSUES Installation
While the cryocooler provides the refrigeration needed to operate an HTS magnetic bearing, its installation must complement the functionality of the device that employs the bearings. Therefore, the cryocooler should be compact and lightweight. Its interfaces with the vacuum vessel and the cold head connection point should permit cryocooler maintenance access and provide for straightforward cryocooler replacement. The axial separation, i.e., the gap between the permanent magnet and the HTS magnet, should be minimized because the levitation capacity of the bearing depends strongly on the distance between the permanent magnet and the HTS magnet. Thus, the thickness of any intermediate members (i.e., liquid bath reservoir walls, etc.) associated with bearing's refrigeration should be minimized. Operating Temperature
The HTS levitation capacity is a function of the material's operating temperature. Levitation force increases as material temperature decreases. Thus, one should avoid temperature gradients
in the HTS bearing elements due to cryocooler refrigeration coupling that could lead to losses in levitation capacity. The temperature of the surface of the HTS material that faces the permanent magnet is most important to its levitation performance. This surface is also closest to sources of ambient heating, i.e., temperature gradients, due to incident thermal radiation. The levitation capacity of the superconductor can be significantly improved by lowering its
temperature. For YBCO, levitation capacity at 77 K is increased by 13% at 66 K. Such conditions can be achieved with liquid bath cooling with the liquid operating in the subcooled mode (i.e., liquid boiling at subatmospheric pressure) and in the conduction cooled mode by lowering the cold head operating temperatures. The conduction cooling mode appears to be the preferred method to achieve such operation. Thermal Mass
A reduction in cryocooler output could result in increased temperature of the HTS bearing material, which would reduce its levitation capability and could lead to bearing failure. To allow adequate time for a safe, controlled shutdown of the parent system, thermal mass can be added to
slow the temperature rise of the HTS material. Thermal mass can be in the form of a solid material in contact with the HTS material or reserve liquid cryogen inventory. The inclusion of solid material thermal mass will affect cooldown.
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Losses The direct losses of a superconducting bearing consist of losses in the rotating permanent magnet and losses that appear in the HTS. The latter are more important, because heat deposited in the HTS must be removed at cryogenic temperatures. Losses in the HTS part of the bearing include magnetic hysteresis loss in the HTS itself and thermal losses arising from thermal conduction and radiative heat flux to the cryochamber. Magnetic friction. Magnetic hysteresis is the dominant loss in the HTS. The HTS experiences a changing magnetic field as the bearing rotates. Part of this changing field is due to
inhomogeneities in the permanent magnet. That is, if one measures the magnetic field near the surface of a cylindrical magnet at a fixed radius as a function of circumferential angle, one will measure a high mean value with a small variation that depends on angle. It is this small variation that creates the hysteresis loss in the HTS. Additional magnetic field variation will be experienced by the HTS if the levitated permanent magnet undergoes vertical or radial oscillation or exhibits a whirl amplitude. The drag torque of hysteresis loss is independent of rotational speed, and the thermal power deposited in the HTS is directly proportional to speed. Eddy current. The same magnetic field variation that causes hysteresis loss in the HTS will cause eddy current losses in any electrically conducting component of the cryochamber. Heat power input in this case is proportional to the square of the rotational speed. The field variation from the rotating magnet may also cause eddy currents in electrically conducting
components of the system outside the cryochamber. If the HTS is composed of an array of components or otherwise has a significant inhomogeneity, the magnetized array will exhibit its own magnetic field variation over the volume of the rotating magnet. If the magnet is
electrically conducting, it will experience eddy currents. Vibration. Vibrations imposed on the HTS bearing system will contribute to the losses of the system. The tolerance to vibration depends on the nature of the specific bearing system. An acceptable level could be several micrometers, with being generally unacceptable.
Thermal. Sources of thermal losses are support and piping solid conduction, residual gas conduction, and thermal radiation. Solid conduction can be minimized by conventional cryogenic design measures, and
residual gas conduction is negligible due to the bearing operational vacuum requirement of Thermal radiation can be significant due to the high emissivity of the YBCO bearing material, which is dull black. The same applies to the case where the bearing is enclosed in an epoxy fiberglass reservoir (G-10 EFG), where Thermal radiation can be controlled by coating the high-emissivity surfaces with a lowemissivity material such as aluminum or stainless steel. The type and thickness of the material must be carefully considered because a metallic interface between the spinning permanent magnet and the HTS magnet can contribute to eddy current losses. Thermal radiation can also be controlled by the use of multilayered insulation (MLI). As in the case of coatings, MLI can contribute to the bearing's eddy current losses. As an example, the thermal radiation between 300 and 77 K has been estimated for the representative bearing system (see section "The Bearing System"). Cold (77 K) surface areas are for conduction cooled (HTS bearing only) systems and for liquid-bath-cooled (reservoir) systems. Cold-surface emissivities for coatings are assumed to be 0.05. Warm-surface emissivity is assumed to be 0.05 (stainless steel vacuum vessel). For MLI, three layers were assumed with a heat flux of The results of the estimate are given in Table 1.
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Cryocooler efficiency. Efficiency of the cryocooler contributes to the overall efficiency of the parent system. This is particularly important where the bearing is incorporated into a
flywheel energy storage system. Vacuum Pumping
The cold surfaces of the HTS bearing assembly can be utilized for the maintenance of the vacuum required for efficient (low-drag) bearing operation. Such a possible installation is as shown in Fig. 6. Emissivity control is utilized to reduce the cooling of the permanent magnet, which would reduce its levitation capacity. Graded MLI is employed to reduce thermal radiation and control eddy current losses. A getter is installed for vacuum maintenance. CONCLUSIONS
• • •
• • • •
HTS bearings can provide stable, low friction levitation. The operation of HTS magnetic bearings can be improved by the use of cryocoolers to provide the necessary refrigeration. Cooling of the HTS material can be by immersion in a liquid cryogen bath (which can be maintained by a cryocooler) or by conduction cooling through direct connection to the cryocooler cold head. Operation of the parent system determines the cryocooler system requirements, which include operational availability, cooling capacity, transfer losses, and installation requirements. The cryocooler installation must complement the functionality of the parent system. Operating temperature must be such that adequate and stable levitation is provided by the HTS bearing. Thermal mass can be included to allow for a moderation of HTS temperature variations and to allow adequate time for a safe shutdown in case of cryocooler failure.
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Figure 6. Thermal control and vacuum pumping.
• •
Losses contribute to the overall efficiency of the integrated system. Factors to be considered in the design process include magnetic friction, eddy currents, thermal heat loads, and cryocooler efficiency. The cold surfaces of the HTS bearing can assist in the maintenance of the system's operating vacuum.
ACKNOWLEDGMENTS This work has been supported by the U.S. Department of Energy, Energy Efficiency and Renewable Energy, as part of a program to develop electric power technology, under Contract W-31-109-Eng-38. The authors acknowledge the manuscript preparation skills of J. A. Stephens and the editorial contributions of C. A. Malefyt.
REFERENCES 1.
J. R. Hull, "Flywheels on a Roll," IEEE Spectrum, (July 1997), pp. 20-25.
2.
Q. S. Shu, R. W. Fast, and H. L. Hart, "Heat Flux from 277 to 77 K through a Few Layers of Multilayer Insulation," Cryogenics, vol. 26, (1986), pp. 671-677.
Advanced Cryocooler Cooling for MRI Systems Robert A. Ackermann and Kenneth G. Herd General Electric Corporate Research and Development 1 Research Circle Niskayuna, NY 12309
William E. Chen General Electric Medical Systems 3001 W. Radio Drive Florence, SC 29501
ABSTRACT
Advances in cryocooler technology during the past several years have enabled the design of new cooling methods for magnetic resonant imaging (MRI) systems. Open cycle operation of MRI systems using a Gifford-McMahon (GM) cycle cryocooler to cool two thermal shields, one at 20 K and the second at 80 K, has been the standard approach used to minimize helium usage in these systems. The concept has worked very well and enabled the development of an important medical imaging modality. However, the 12 K temperature limit of these cryocoolers has limited the design flexibility of the MRI magnet system by requiring a cylindrical design with two thermal shields and a large helium container to extend the operating time. The development of Gifford-McMahon cycle cryocoolers capable of cooling below liquid helium temperature, or providing larger cooling capacities between 4.2 and 10 K, has removed this design barrier and provided greater overall system design flexibility. The paper describes the impact that new GM cryocooler developments, based on rare earth intermetallic compounds in the second-stage regenerator, has had on MRI designs. By extending the cooling capacity of these units to below 4.2 K with rare earth materials, new MRI products have been developed that operate as closed cycle systems without the need for replenishing liquid helium to maintain the magnet at temperature for long periods of time. The paper describes the evolution of MRI systems at the General Electric Company from open cycle systems to two new developments using conduction cooling and helium recondensing to eliminate the need for refilling with helium. The paper reviews the design of a conductively cooled system developed for an open MRI magnet used for interventional therapy and a helium recondensing system that was incorporated into GE’s product line. In addition to a description of these systems, the operational reliability of cryocooled systems will also be reviewed.
Cryocoolers 10, edited by R. G. Ross, Jr. Kluwer Academic/Plenum Publishers, 1999
857
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COMMERCIAL CRYOCOOLER APPLICATIONS
1. INTRODUCTION There have been four cryocooler advancements (see Table 1) that have had a major impact on the development of MRI systems. The first was in 1960 when Arthur D. Little Inc. introduced the first two-stage Gifford-McMahon cycle cryocooler.1 This introduction was significant because it enabled refrigeration to be produced at temperatures below 20 K in a simple, and relatively inexpensive, closed cycle regenerative cooler. The second major advancement occurred during the 70’s when the maintenance interval for GM cryocoolers was extended to 12 months. This advancement was related to the development of cryogenically cooled vacuum pumps, where commercial success of the cryopump was dependent on the performance and reliability of the cryocooler. Prior to this time, the maintenance interval for cryocoolers was measured in weeks. The first major advancement in cryocooler performance came in 1990 when researchers at Tokyo University2 lowered the second-stage temperature from 12 K to 4.2 K by using a rare earth intermetallic compound in the second-stage regenerator. Further developments with multiple rare earth compounds have increased the capacity at 4.2 K and enabled the commercialization of a 1.0 W 4.2 K cryocooler in 1996. These four developments, as we shall see later, have had a major impact on the design of GE’s MRI products in the following ways: 1. Provided greater flexibility in magnet and cryostat designs enabling new products to be developed. 2. Reduced operating costs by reducing on-site cryogen storage and transfer requirements. 3. Reduced manufacturing costs by simplifying cryostat designs. The critical cryocooler characteristics that have evolved as the most important in enabling these MRI developments are given in Table 2. The cooling capacity and temperature of the cryocooler is important because of its impact on the “openness” of the cryostat design and reduction in helium
ADVANCED CRYOCOOLER COOLING FOR MRI SYSTEMS
859
consumption, and the maintainability is important because of the cost of servicing an MRI system in the field. A summary of the changes these cryocooler developments have had on MRI system design is presented in Table 3. The first occurred in 1985, which related to the 1975 cryocooler development, when a two-stage GM refrigerator was used to cool the radiation shields surrounding the magnet. This development occurred primarily because of the improved capacity and maintenance interval of the GM cryocooler, which made it possible to reduce the helium consumption from 1.0 l/h to less than 0.2 1/h and increase the scheduled magnet maintenance interval from once a month to once a year. This was a major product advancement that enabled MRI systems to be installed in hospitals, clinics, and mobile vans throughout the country, not just research hospitals in major cities. The second major MRI development occurred in 1993 when an “open” conductively cooled magnet was installed in Brigham and Women’s hospital.3 This development was based on the 1990 cryocooler development that increased the second-stage cooling capacity at 10 K, thus enabling the use of a directly cooled superconductor. The significance of this was that it eliminated the need for a large helium vessel and greatly reduced the size of the magnet cryostat. Prior to this development, the requirement for large quantities of LHe led to cylindrical systems in which the patient has to lie in the bore and at the mid-plane of the magnet with no access to the patient during the imaging process. The confinement in the small bores of these systems also led to, on many occasions, patient discomfort and anxiety during imaging. The development of the conductioncooled magnet enabled a more “open” design and the introduction of interventional therapy during the imaging process. The third, and most recent, GE MRI development is the zero helium boil-off cryostat which is based on the 1996 cryocooler development using multi-compound rare earth materials in the second-stage regenerator. This development has enabled GE to reduce the size and complexity of its
standard LHe systems by reducing the size of the LHe vessel and eliminating one thermal shield, thereby enabling a more “open” design than previously possible with conventional LHe systems. A more detailed discussion of these MRI systems is presented in the following section. The correlation these tables show between the advances made in cryocooler technology and MRI designs is that MRI designers are adept in using cryocooler technology to advance new MRI products and that the progression from “closed,” restrictive, systems to “open” interventional systems was made possible by the development of a GM cryocooler with a rare earth regenerator. Previous attempts to use a three-stage GM / Joule Thompson 4.2 K cryocooler for recondensing proved impractical because of the poor reliability and cost of these coolers. Therefore we can conclude that as cryocooler capacity at 4.2 K and the maintenance interval are increased new MRI products will be developed and markets expanded. 3. GE MRI SYSTEMS
As presented in Table 3, the first of the GE MRI cryocooled products was the Signa 1.5 Tesla system. The cryostat design used two cryocooled thermal shields to reduce the heat leak into the helium bath. A schematic layout of the Signa cryostat is shown in Figure 1, and a picture of a GE Signa 1.5 T MRI magnet is shown in Figure 2. This system was the top of the GE product line and was well received by radiologists because of the high quality images it produced. However, its cost and size limited its sales to larger medical institutions and in 1990 the marketplace started emphasizing smaller, less expensive systems. The next development within GE was conductively cooled magnets in which the cryogens were replaced with a two-stage cryocooler that cooled both the magnet and a single thermal shield. This concept was chosen because the elimination of one thermal shield and the liquid helium container meant that the cryostat could be made considerably smaller and less expensive. In 1993, a 0.5 T conductively cooled magnet of a novel open configuration was developed to allow direct, unimpeded access to the operational field at the most homogeneous part of the magnet. As shown in Figure 3, the magnet consists of a split pair of three coils each, which are optimized to attain a wide
860
COMMERCIAL CRYOCOOLER APPLICATIONS
Figure 1. Schematic of a Signa 1.5 T MRI magnet cryostat.
ADVANCED CRYOCOOLER COOLING FOR MRI SYSTEMS
861
Figure 2. Signa 1.5 T MRI magnet cryostat.
separation gap with a large region of field homogeneity. To achieve the large gap, the coils were wound using tape conductor that was conductively cooled to 10 K by a pair of GM cryocoolers. The elimination of the helium vessel and one thermal shield enabled each coil set to be enclosed in a tightly spaced vacuum enclosure that maximized the gap between the assemblies. The two split pairs of coils and vacuum enclosures were connected through posts located at the top and bottom of the coils that also supported the attractive forces between the coil pairs and provided the thermal
cooling path between the two halves. The cooling was provided by two Leybold UCH-130 GM coldheads mounted in separate evacuated sleeves located at the bottom of one cryostat. A design having two cryocoolers mounted in a separate vacuum sleeve was chosen to obtain a highly reliable and easily maintained system. By incorporating two cold heads with sufficient capacity in each
Figure 3. 0.5 T conductively cooled magnet with novel open configuration.
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COMMERCIAL CRYOCOOLER APPLICATIONS
cold head to maintain the conductor below the critical temperature for this application, 13.5 K, and separate vacuum enclosures for the cryocoolers, the goal for having a minimum maintenance interval of 1 year along with the ease of removal of the cryocoolers without disturbing the main vacuum
was achieved. The primary components of the dual cryocooler assembly are shown in Figure 4. The two cold heads are mounted horizontally opposed to one another in individual sleeves. Each sleeve has a 40 K station in contact with the first stage of the coldhead and a 10 K station in contact with the second stage. Heat is conducted from the sleeve to the cold heads through spring-loaded indium gaskets. Thermal busbars connect the thermal stations to the cryostat thermal shield and to the magnet. High-temperature superconducting current leads are used in the cold region and are heat stationed dialectically to the 40 K and 10 K thermal stations. The dual refrigerator assembly mounts directly to the floor to attenuate the transmission of vibrations from the cold heads to the coils. Table 4 summarizes the steady-state temperatures of the dual cryocooler assembly, the magnet, and the thermal shield. The temperatures are given for both cases with one and both cold heads operating.
Figure 4. Dual cryocooler assembly.
ADVANCED CRYOCOOLER COOLING FOR MRI SYSTEMS
863
The split pair magnet, MRT, was installed as the centerpiece of an interventional suite at Brigham and Women’s Hospital Boston, MA, in 1994 and high quality intraoperative images have been produced in interventional procedures on a continuing basis since. The technical success of this system has been verified by the installation of 13 additional systems since. A photo of the MRT is
shown in Figure 5. The most recent development at the General Electric Company is a zero boil-off K4 system in which the magnet is immersed in LHe and a two-stage cyrocooler is used to recondense the helium and cool a thermal shield. Recondensing the LHe eliminates the need to replenish the magnet with helium for as long as the cryocooler maintains its capacity at 4.2 K and provides the potential for operating the MRI system for its full life without the need to refill the cryostat and only routine maintenance on the cryocooler. The K4 uses a GM cryocooler that contains a multi-compound rare earth 2nd stage regenerator and provides 1.0 W at 4.2 K with 40 W at 40 K. From the product requirements, The factors driving the zero boil-off design were: 1. The system could be made smaller, more “open,” and lighter. 2. Less expensive to manufacture 3. More flexible and versatile for future applications 4. Easier to service and requires less frequent servicing 5. Employs internal magnetic shielding
6. Minimizes the LHe usage over the life of the system A schematic of the K4 zero-boil-off magnet and cryostat is shown in Figure 6, and a photograph of an early production unit is shown in Figure 7. The unit consists of the inner helium container which contains the main NbTi field coils and shielding coils along with the LHe bath. The helium vessel is supported from the outer vacuum vessel with low heat leak tension straps. A single thermal shield surrounds the inner and outer surfaces of the helium vessel and its attached to the first stage of the cryocooler through the “thermal link” and cooled to 40 K. The cryocooler is mounted at the top of the cryostat and interfaces the recondensing chamber and thermal shield through the cryocooler sleeve that isolates the vacuum in the cryocooler sleeve from the main cryostat vacuum. The recondensing surface is cooled to 4.2 K by the second stage of the cryocooler and evaporated helium flows into the recondensing surface and returns to the LHe bath by natural convection. This gravity feed design also represents a thermal valve that allows the cryocooler to be
Figure 5. The 0.5 T MRT conductively cooled imaging system installed at Brigham and Women’s Hospital.
864
COMMERCIAL CRYOCOOLER APPLICATIONS
Figure 6. Schematic of the K4 zero-boil-off magnet and cryostat.
Figure 7. Early production unit of the K4 zero-boil-off magnet and cryostat.
ADVANCED CRYOCOOLER COOLING FOR MRI SYSTEMS
865
removed from the system without increasing the heat load to the LHe bath. That is, as the second stage of the cryocooler warms up above 4.2 K helium flow into the helium vessels from the re
condenser stops, and the cryocooler becomes thermally disconnected from the LHe, reducing the heat leak to only conduction through the “Liquid/Gas Tubes.” This passive isolation along with the cryocooler sleeve enables the cryocooler to be removed from the system for servicing with no
helium loss. Over a longer term, the thermal shield will warm, producing some loss in helium. The steady state temperatures of operation without the cryocooler in the system are given in Table 5. This zero boil-off design has become very popular in many markets where LHe is expensive and trained maintenance personnel are not readily available.
4. COMPARISON OF GE-MRI SYSTEMS As we have seen, the evolution of the GE MRI systems described above was made possible by
advances in cryocooler technology. We see that the advancement from cylindrical cryostats with claustrophobic patient bores to open systems that enable interventional therapy has occurred because cryocooler performance and reliability has improved greatly during the past 20 years. Table 6 compares some of the advantages and disadvantages of these developments and explains what the driving factors were in this evolution. From the first Signa MRIs to the conduction-cooled units, the change was driven primarily by the requirement for “openness” for interventional therapy and greater flexibility in magnet design. It should be noted that for image quality the Signa 1.5 T created a standard of excellence that even today is used to measure the performance of other systems. Other key features that made the Signa system a commercial success were: 1. Using the cryocooler to cool only the shields, and not the LHe bath, made the magnet
insensitive to cryocooler failures and able to operate for long periods of time without the cryocooler. This provided a high degree of reliability for the system. 2 Its large weight made it less sensitive to cryocooler vibrations, which added to its image quality. 3. Its internal magnetic shielding eliminated the need for costly external magnetic shielding to minimize stray magnetic fields. This improved siting flexibility and reduced siting costs. The major success of the MRT magnet was that it achieved the “openness” required for interventional therapy. However, the use of two cryocoolers, no internal shielding, and superconductor made it a very expensive system to buy and site. In addition, servicing was difficult
because of the interaction of the cryocoolers with the magnetic field. This interaction produced a large retaining force on the cold head motor that makes it difficult to remove the cryocoolers from their vacuum sleeves with the magnet at field. Also, the temperature sensitivity of the field homoge-
neity limits the temperature rise of the magnet during servicing, making it difficult to change out
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COMMERCIAL CRYOCOOLER APPLICATIONS
the cryocoolers without disturbing the field homogeneity. This characteristic is due to the current distributions in the tape, which will shift as the coil warms above its previous ramping
temperature, thus requiring a reshimming of the magnet to restore the field homogeneity. With only one cryocooler running, it has proven difficult to prevent this temperature rise. During installation and startup of the magnet, cryocooler reliability was also a major service issue. Premature performance degradation, with the associated rise in magnet temperature, required the replacement of cryocoolers at several installations. Examination of the failed units indicated that the problem was
due to helium contamination in the cryocoolers. Typical contamination levels found in cryocoolers removed from several sites are given in Table 7. In comparison, the longer term operation of the MRT magnet at Brigham and Women’s hospi-
tal has proven very reliable. Servicing of the cryocoolers has occurred three times in the four years that the system has been operating at the site. The most recent service on the system was performed in November 1997, which resulted from a degradation in performance of the cryocoolers after 18 months of continuous operation. To assess the cause of the degradation a gas analysis and mechanical inspection were performed on the two cold heads removed from the magnet. In comparison to the cryocoolers from the other 13 MRT sites, the contamination found in the Brigham and Women’s cryocoolers after 18 months of operation was considerably less, as shown in Table 8. In addition,
ADVANCED CRYOCOOLER COOLING FOR MRI SYSTEMS
867
the Ar was used to define the amount of air that would be diffused into the system through the
aeroquip coupling rubber seals in 18 months. A mass balance was performed on the measured CO, and in the reactants of the combustion equation. This led to good agreement between the measured and calculated levels of Ar in the gas sample. For the “A” cryocooler this is 31 ppm vs 29.5 ppm and for the “B” cryocooler this is 52 ppm vs. 43.1 ppm. It is postulated that localized corona effects in the compressor provide the energy for the thermal reaction to occur.
The future of the MRT product is not clear at this time. The reliability issues of the cryocoolers, along with the high system costs associated with having two cyrocoolers, using superconductor and siting costs due to external magnetic shielding requirements, have affected the commercial success of the product and shade the remarkable medical success achieved through interventional surgical procedures. The zero boil-off system is the most recent of the GE developments. The lower life cycle cost, the ease of service, and the ease of manufacturing of the zero boil-off technology have gained wide acceptance for the K4 product line both within GE Medical Systems and with GE’s customers. Its major advantages listed in Table 6 are proving to be driving its acceptability.
5. CONCLUSIONS
The development of MRI systems at GE has shown a strong correlation to the developments in cryocooler technology. This was most pronounced in the 1980’s and 90’s when advancements in reliability and performance enabled the cryocoolers to become an integral part of the MRI system. The first of these advances was to reduce the LHe boil-off to very low levels, and the second was to reduce LHe consumption to zero by either conduction cooling of the magnet or recondensing helium boil-off from the magnet. The requirement for “openness” in magnet designs will continue to drive this correlation and provide a commercial platform for cryocooler manufacturers. The limitation on this progress, however, has been both the initial cost and inconsistency of performance reliability. These two characteristics have put the MRT system at risk and could negate the benefits derived from a zero boil-off magnet system. Those manufacturers that perform best in these two areas will enjoy success and help expand MRI into many new and important diagnostic and interventional areas. REFERENCES
1. Gifford, W.E., “The Gifford-McMahon Cycle,” Advances in Cryogenic Engineering, Vol. 11, 1966 (K.D. Timmerhaus, ed.). Plenum Press, New York, p. 152. 2. Kuriyama, T., Hakamada, R., Nakagome, H., Tokai, T., Sahashi, M., Li, R., Yoshida, O., Matsumoto, K., and Hasimoto, T., “High-efficiency two-stage GM refrigerator with magnetic material in the liquid helium temperature region,” Advances in Cryogenic Engineering, Vol. 35B, 1990 (R.W. Fast, ed.), Plenum Press, New York, p. 1261. 3. Laskaris, E.T., Ackermann, T., Dorri, B., Gross, D., Herd, K., and Minas, C., “A Cryogen-Free Superconducting Magnet for Interventional MRI Applications,” IEEE Transactions on Applied Superconductivity, Vol. 5, No. 2, June 1995.
Proceedings Index This book draws upon the work presented at the 10th International Cryocooler Conference, held in Monterey, California, in May 1998. Although this is the tenth meeting of the conference, which has met every two years since 1980, the authors’ works have only been made available to the public in hardcover book form since 1994; this book is thus the third hardcover volume. Prior to 1994, proceedings of the International Cryocooler Conference were published as informal reports by the particular government organization sponsoring the conference — typically a different organization for each conference. Most of the previous proceedings were printed in limited quantity and are out of print at this time. For those attempting to locate references to earlier conference proceedings, the following is a listing of the nine previous proceedings of the International Cryocooler Conference. 1) Refrigeration for Cryogenic Sensors and Electronic Systems, Proceedings of a Conference held at the National Bureau of Standards, Boulder, CO, October 6-7, 1980, NBS Special Publication 607, Ed. by J.E. Zimmerman, D.B. Sullivan, and S.E. McCarthy, National Bureau of Standards, Boulder, CO, 1981. 2) Refrigeration for Cryogenic Sensors, Proceedings of the Second Biennial Conference on Refrigeration for Cryogenic Sensors and Electronics Systems held at NASA Goddard Space Flight Center, Greenbelt, MD, December 7-8, 1982, NASA Conference Publication 2287, Ed. by M. Gasser, NASA Goddard Space Flight Center, Greenbelt, MD, 1983. 3) Proceedings of the Third Cryocooler Conference, National Bureau of Standards, Boulder, CO, September 17-18, 1984, NBS Special Publication 698, Ed. by R. Radebaugh, B. Louie, and S. McCarthy, National Bureau of Standards, Boulder, CO, 1985. 4) Proceedings of the Fourth International Cryocoolers Conference, Easton, MD, September 25-26, 1986, Ed. by G. Green, G. Patton, and M. Knox, David Taylor Naval Ship Research and Development Center, Annapolis, MD, 1987. 5) Proceedings of the International Cryocooler Conference, Monterey, CA, August 18-19, 1988, Conference Chaired by P. Lindquist, AFWAL/FDSG, Wright-Patterson AFB, OH. 6) Proceedings of the 6th International Cryocooler Conference, Vols. 1-2, Plymouth, MA, October 25-26, 1990, David Taylor Research Center Report DTRC-91/001-002, Ed. by G. Green and M. Knox, Bethesda, MD, 1991. 7) 7th International Cryocooler Conference Proceedings, Vols. 1-4, Santa Fe, NM, November 17-19, 1992, Air Force Phillips Laboratory Report PL-CP-93-1001, Kirtland Air Force Base, NM, 1993. 8) Cryocoolers 8, proceedings of the 8th ICC held in Vail, Colorado, June 28-30, 1994, Ed. by R.G. Ross, Jr., Plenum Press, New York, 1995. 9) Cryocoolers 9, proceedings of the 9th ICC held in Waterville Valley, New Hampshire, June 25-27, 1996, Ed. by R.G. Ross, Jr., Plenum Press, New York, 1997.
869
Author Index Ackermann, R.A., 857 Alexeev, A., 475 Allen, M.S., 315 Arkhipov, V.T., 87, 467, 487, 529, 689 Asami, H., 575 Bauwens, L., 179 Bennett, J.M., 275 Benoît, A., 795 Benschop, A.A.J., 359 Berry, D.J., 761 Bhatia, R.S., 795 Biao, W., 791
Davis, T.M., 1, 21, 59, 671
Ito, S., 395
Dolan, F., 431 Donabedian, M., 1, 717 Doornink, D.J., 805
James, S.C., 43 Jeong, S., 249
Duband, L., 281
Jewell, C.I., 281, 513, 521, 781 Ji, G., 787, 791
Egmond van, H., 565 Elwenspoek, M., 565 Fabris, D., 257 Felmley, T., 621
Fernandez, R., 449 Frederking, T.H.K., 275 Fujimoto, S., 213
Johnson, D.L., 95, 119, 139, 771 Jones, B.G., 513 Ju, Y.L., 197, 205
Bock, J.J., 795 Boiarski, M., 457
Getmanets, V.F., 55, 467, 529, 689, 733, 743
Kabashima, S., 581, 587 Kanazawa, Y., 575 Kang, Y.M., 149, 213 Kawecki, T.G., 43 Khatri, A., 457, 535 Kiehl, W., 59
Bondarenko, S.I., 55
Gilbert, A., 33, 129, 723
Kirkconnell, C.S., 239
Borisenko, A.V., 467
Glaister, D.S., 1, 671, 717 Gong, M.Q., 481
Kittel, P., 351, 815
Green, K., 119 Griffin, M.J., 795 Gschneidner, Jr., K.A., 629, 639 Gully, W.J., 59, 505 Gutow, D., 761
Kotsubo, V., 157, 163, 171, 299 Kovalenko, V., 457 Ku, J., 671 Kukharenko, V.N., 405, 413 Kuo, D.T., 105, 191
Blom,C.J.H.A., 837
Bradley, W., 439 Bradshaw, T.W., 67, 513, 521, 781 Brisson, J.G., 655 Bruning, J.L., 829
Bruning, J.F., 33, 723 Bugby, D., 671 Burger, J.F., 553, 565 Burt, W., 129
Konkel, C., 439
Haberstroh, Ch., 475
Lanes, C., 129
Hagiwara, Y., 395 Hall, J.L., 343
Le, A., 275 Lee, J.M., 351
Cai, J.H., 337 Carrington, H., 59 Champagne, P.J., 163, 171
Hanes, M., 111
Lester, J., 505
Hashimoto, T., 611
Chan, C.K., 104
Helvensteijn, B.P.M., 647 Herd, K.G., 857 Hill, D., 535
Levenduski, R., 449, 505 Levin, A.Y., 529, 689 Levy, A.R., 545 Li, R., 575 Liang, J.T., 197, 205, 233, 337, 481 Lobko, M.P., 487 Loc, A.S., 105, 191
Chandratilleke, G.R., 221, 227 Chang, H.M., 707
Charles, I., 281 Chase, D., 111 Chen, H.E., 275 Chen, W.E., 857 Clappier, B., 171 Colbert, R., 139
Hatakeyama, H., 581, 587
Hiratsuka, Y., 149 Hoden, B.P., 77 Hofmann, A., 369 Holland, H.J., 553, 565, 837
Longsworth, R.C., 535 Lubchenko, V.N., 87 Luo, E.C., 481
Collins, S.A., 119, 771
Hooijkaas, H.W.G., 359 Hozumi, Y., 321 Hsu, I.C., 497
Corey, J.A., 181
Huang, P., 299
Martin, C.M., 181, 823
Curran, D.G.T., 1
Hull, J.R., 847
Martin, J.L., 181, 823
Dan’kov, S.Y., 639
Iida, T., 221, 321
Davidson, D., 129
Inaguchi, T., 593, 603
Mason, P.V., 795 Matsubara, Y., 149, 213, 221, 227, 291
871
872
AUTHOR INDEX
McCormick, J.A., 421, 431 Mikhalchenko, R.S., 467, 689 Miller, S.D., 717 Mimura, K., 221 Miquet, L., 281
Pfotenhauer, J.M., 387 Phelan, P.E., 663, 753 Plachta, D.W., 815
ter Brake, H.J.M., 553, 565,
Povstiany, L.V., 87, 467
Mitchell, M.P., 257, 379
Quack, H., 265, 475
Tishin, A.M., 639 Tomlinson, B.J., 21, 33, 59, 67, 129, 257, 671, 723
Miyata, Y., 611
837
Timmerhaus, K.D., 351
Torrison, R., 829
Mujezinovic, A., 663 Murakami, M., 321, 329
Raab, J., 139 Radebaugh, R., 351, 829
Usami, T., 581, 587
Nagao, M., 593, 603
Ravex, A., 281 Ravikumar, K.V., 291
Van Sciver, S.W., 707
Naka, K., 593, 603 Nakagome, H., 221, 227, 581, 587, 611 Nakame, H., 611
Nakamura, N., 329 Nakano, A., 329
Nam, K., 249 Nara, K., 395
Reilly, J., 21, 421
Richards, L., 761 Rijpma, A.P., 837
Wade, L.A., 545, 553
Roach, P.R., 647 Rogalla, H., 553, 565, 837
Wickman, H., 119 Wild, S., 369
Ross Jr., R.G., 119, 139, 343, 771
Williams, B.G., 171, 697
Rühlich, I., 265
Wollan, J.J., 315
Nguyen, T., 139
Niemann, R.C., 847 Nisenoff, M., 829 O’Baid, A., 111 Ohtani, Y., 221, 227, 581, 587
Okamura, M., 611 Okamura, T., 581, 587 Okuda, H., 221, 227 Olson, J.R., 157, 163, 171, 307 Onishi, A., 575 Orlowska, A.H., 67, 513, 521, 781
Patel, A.B., 655 Pecharsky, V.K., 629, 639 Penswick, L.B., 77
Wilkinson, R.M., 781 Wu, Y., 787, 791
Nast, T.C., 157, 163, 171, 299
Nellis, G.F., 421, 431
Wang, J.J., 233
Salerno, L.J., 815 Satoh, T., 575 Scull, S., 67, 513, 781 Shinohara, S., 221, 227 Shiraishi, M., 329 Shire, J.M., 663
Sixsmith, H., 421, 431 Smith, R., 439 Spradley, I.E., 697 Stack, R., 761 Stears, H., 87, 467, 487, 529, 689, 743
Stouffer, C., 671 Swanson, T., 671 Swift, G.W., 307, 315 Swift, W.L., 421, 431 Takamatsu, K., 329
Yakuba, V.V., 487
Yang, J.H., 205
Yang, L.W., 233, 337
Yevdokimova, O.V., 487 Yoshida, S., 275
Yoshimura, H., 593, 603 Yoshimura, N., 221, 227 Yuan, J., 387 Yuan, S.W.K., 105, 191 Yuen, W.W., 497 Zhang, P.S., 337 Zhao, L., 753
Zhou, S.L., 227 Zhou, Y., 197, 205, 233, 337, 481 Zhu, W.X., 197, 205, 337 Zhun, G.G., 733, 743
Subject Index Adv. Mobile Telecom. Tech., Inc., 395
2-stage coaxial pulse tube, 233 compact coaxial pulse tube, 197
Aerospace Corporation:
miniature multi-bypass coaxial PT, 205
Accumulators (see Thermal storage)
CSIM system design model, 717 flexure bearing trial results, 163 integration hardware developments, 671
space cryocooler overview, 1 Air Force Research Lab (formerly Phillips Lab): cooler endurance testing, 33 heat rejection effects, 723 MTI cooler system performance, 129
program overview, 21 AIRS cryocooler development, 119, 771 Ames Research Center (NASA):
dilution refrigerator, 647 pulse tube with inertance tube, 291
pulse tube modeling, 351 Ames Laboratory (Iowa State Univ.), 629 APD Cryogenics:
10K BETSCE cryostat, 535 mixed gas refrigerant optimization, 457 Applications of cryocoolers (see Integration with
cryocoolers) Argonne National Lab, 847 Arizona State University, 663, 753 Ball Aerospace cryocoolers: 10K Hybrid J-T/Stirling, 1, 505 30K two-stage Stirling, 1
35K/60K three-stage Stirling, 1, 21, 59 1.5W-55K HIRDLS Stirling, 1, 761
mixed gas experiments for J-Ts, 481
pulse tube long-term stability, 337 Chiyoda Corp., 321 Compressors: linear motor for, 111 flexure bearings for, 163 Concurrent Technology Corp., 621
Conductance measurements of: Kapton MT, 753 Contamination: accelerated testing for, 743 effect on pulse tube cooler, 337, 343 internal outgassing rate, 733 COOLLAR, 1, 21, 449 Creare Cryocoolers: 35K reverse-Bnyton, 1, 421
5W-65K reverse-Brayton, 1 65K SSC Stirling cooler, 1 NICMOS rev-Brayton, 1, 21, 431
miniature reverse-Brayton, 21 Cryenco, 315 CRYOBUS, 21 Cryocooler applications (see Integration of cryocoolers) Cryocooler integration technologies (see Integration of cryocoolers) CRYOTSU, 21, 671 CSIM cryosystem design software, 21, 717
COOLLAR closed-cycle J-T, 1, 21, 449 Bearings:
Aerospace flexure, 163 BEI: 1W tactical pulse tube cooler, 191
0.3W 78K tactical Stirling, 105 BETSCE, 21, 1, 535 Brayton cycle cryocoolers (see Reverse-Brayton cryocoolers)
British Aerospace (BAe) cryocoolers (see Matra
Daikin Industries: 1-5W 80K pulse tube cooler, 149 5-20W 80K pulse tube cooler, 213 DC flow, effect on P-T efficiency, 163, 281, 299,
307, 315 Decade Optical Systems, 77 Diaphragm compressor: Creare Stirling, 33 Dilution refrigerator, 647, 795
Marconi cryocoolers)
Eindhoven Univ. of Technology, 359 California Institute of Technology, 795
CEA/DRFMC (France): inertance & DC flow in pulse tubes, 281
Chinese Academy of Sciences, Cryogenics Lab:
Electric field emissions (see EMI/EMC measurements)
Electromagnetic interference (see EMI/EMC measurements)
873
874
Electronics:
cooler drive and control, 139, 791, 771 EMI/EMC measurements:
AIRS pulse tube cryocooler, 771 IMAS pulse tube cryocooler, 139
SUBJECT INDEX
J-T cryocoolers:
APD 10K BETSCE cryostat, 535 APD mixed gas optimization, 457 Ball closed-cycle COOLLAR, 449 Ball hybrid 10K J-T/Stirling, 505
Erbium regenerator materials, 611
BETSCE flight experiment, 1, 21, 535
European Space (ESA-ESTEC) activities: 4K hybrid J-T/Stirling development, 513 4K hybrid J-T/Stirling life tests, 521 pulse tube dc flow and inertance, 281 space cooler integration experience, 781
Chinese mixed-refrigerant experiments, 481 Dresden mixed gas w/ precooling, 475 microchannel heat exchanger for J-T, 497 MMR Technologies, 43
Flexure bearings (see Bearings) Fokker Space radiant cooler, 805 Forschungszentrum Karlsruhe, 369
porous plug J-T valve, 545
RAL/MMS hybrid 4K J-T/Stirling, 513, 521 SR&DB 5-10K hybrid J-T/Stirling, 529 SR&DB 84-90K mixed gas cooler, 467 SR&DB mixed gas study, 487 Jet Propulsion Lab:
Gas-gap heat switches (see Heat switches) General Electric MRI applications, 857 Gifford-McMahon Cryocoolers:
4K with Erbium regenerator materials, 497 displacer stroke sensitivity at 4K, 575
integration with MRI systems, 857 multi-layer magnetic regenerators for, 581
AIRS cryocooler system performance, 119 AIRS cryocooler EMI, 771
contamination effects on pulse tubes, 343 IMAS pulse tube cryocooler, 139
porous plug J-T valve, 545 TI 1W-80K cooler characterization, 95
Joule-Thomson Cryocoolers (see J-T cryocoolers)
neodymium ribbon regenerator for, 621 numerical simulation of, 593
Goddard Space Flight Center (NASA): cooler integration technologies, 671 HTS applications: binary current leads, 707
magnetic bearings, 847 magnetocardiography using SQUIDS, 837 Heat exchangers: microchannel, 497
perforated plate, 275 Heat switches, 21, 671, 565 Heat pipes, cryogenic, 671 Heat conduction (see Conductance measurements) High temperature superconductor applications (see HTS applications) Hong Ik University, 707 Illinois Inst. of Technology, 257 IMAS pulse tube cooler, 1, 139 Integration of cryocoolers with:
bolometers, 795 cryogenic loop heat pipes and CPLs, 671 flight electronics, 771, 791 heart scanner, 837 HTS applications (see HTS applications) human and robotic missions to Mars, 815 industrial applications, 823 magnetic bearings, 847 MRI systems, 857 space experiments (see Space experiments) space instruments (see Space instruments) SQUIDS, 837 systems integration software (CSIM), 717 thermal storage, 87, 467, 671, 689, 697 thermal switch, 671
Kapton MT conductance tests, 753 Kharkov Polytechnic University: modeling wave coolers, 405 outgassing rate study, 733 teaching pulse tube design, 413 Korea Adv. Inst. of Sci. and Tech., 249 Life test results: Creare 65K reverse Brayton, 33 Creare 65K diaphragm Stirling, 33
NRL low cost coolers, 43 RAL 4K hybrid J-T/Stirling, 521 TRW 3585 pulse tube, 33 TRW 6020 pulse tube, 33 Linear motor, variable reluctance, 111 Lockheed Martin IR Imaging Systems: AIRS cryocooler system performance, 119 Lockheed Martin Advanced Tech. Center: 2W-60K pulse tube development, 157 35K LADS Stirling cooler, 1 35K/60K pulse tube cooler, 21, 163
N2 triple-point thermal storage, 697 pulse tube coolers for HTS space appl., 171 pulse tube DC flow investigation, 299 Los Alamos National Laboratory, 307
Magnetic field emissions, 139, 771 Magnetic refrigerators: regenerator materials for, 629 refrigerant materials for, 639
Massachusetts Inst. of Technology, 655 Matra Marconi cryocoolers: 4K hybrid J-T/Stirling cooler, 1, 513 10K 2-stage Stirling cooler, 1, 21, 67 50-80K Stirling cooler, 1 flight applications of MMS coolers, 781
SUBJECT INDEX
Mesoscopic Devices: industrial cooler design considerations, 823 pulse tube cooler for telecom., 181
Microscale cryocoolers: miniature pulse tube & Stirling, 663 miniature sorption, 553, 565
Mitchell Stirling, 257, 379 Mitsubishi Electric, 593, 603 Mixed refrigerants (see Refrigerants, mixed gases) MMR Technologies, 43 MRI cooler applications, 857 Nat'1 High Magnetic Field Lab, 707 Nat’1 Inst. of Standards and Tech. (see NIST) Naval Research Laboratory: cooler life test results, 43 cooler survey for electronic appl., 829 Neodymium ribbon regenerator mat'l, 621 Nichols Research Corp. (NRC):
cooler endurance testing, 33 cooler survey, 829 heat rejection effects, 723 NICMOS: cooler flight electronics, 439 reverse-Brayton cooler, 431 Nihon University: 1-5W 80K pulse tube cooler, 149 4K space pulse tube cooler, 221
5-20W 80K pulse tube cooler, 213
875
double vortex tube as HX, 257 effect of working pressure, 227
experimental verification of models, 387 flow calculation, 321 flow visualization within, 329 gas temperature measurement within, 395
heat transfer characteristics of, 249 inertance tube, effect of, 281, 291
microscale PT development, 663 modeling of wave cooler, 405 modeling with harmonic simulation, 359 modeling with MS*2 Stirling code, 379 modeling with thermoacoustic theory, 369 multi-bypass coaxial pulse tube, 205 pumping loss analysis, 603 Raytheon 3-parallel tube, 239 tapered pulse tubes, 307, 315 teaching design in Ukraine, 413
Queen Mary & Westfield College, 795 RAL (see Rutherford Appleton Laboratory) Rare earth compounds, 611, 621 Raytheon (formerly Hughes Aircraft): LCC low-cost pulse tube, 43 PSC 60K Stirling cooler, 1, 21 pulse tube research, 239 SMTS 35K Stirling cooler, 1
Refrigerants:
effect of inertance tube on PT, 291
contamination of, 343, 733, 743
effect of working presssure on PT, 227
for magnetic refrigerators, 639
NIST: cooler survey, 829 Orbital Sciences Corp.: NICMOS cooler electronics, 439
Phase change materials (see Thermal Storage) Pulse tube cryocoolers: 1-5W 80K Daikin, 149 1W BEI tactical, 191 5-20W 80K Daikin, 213 Cryenco 1100W 125K, 315 LM-ATC 2-stage 35/60K, 163 LM-ATC 2W-60K, 157 LM-ATC for HTS space applications, 171 TRW 3503, 33 TRW 3585, 33 TRW 6020, 33 TRW AIRS, 119, 771
TRW IMAS, 139 TRW miniature, 1, 21
TRW MTI, 129 Pulse tube theory and investigations: 2-stage coaxial developement, 233
mixed gases, 457, 475, 481, 487 superfluid helium, 655 Refrigeration performance overviews: for electronic applications, 829
space coolers, 1 Regenerator: etched foil performance trial, 163 GdZn magnetic alloys for, 629 modelling of element geometries, 265 multi-layer magnetic, 611
neodymium ribbon, 621 optimization for 10K Stirling, 67 optimization for 4K GM, 581, 611 Reverse-Brayton cryocoolers: 35K Creare, 1, 421 5W-65K Creare, 1,33 Creare NICMOS, 1, 21, 431 miniature Creare, 21 NICMOS cooler electronics, 439 Rutherford Appleton Laboratory: 4K hybrid J-T/Stirling development, 513 4K hybrid J-T/Stirling life tests, 521 10K 2-stage Stirling, 67 space cooler integration experience, 781
2D corrections to 1D models, 351 compact coaxial pulse tube, 197
Sandia National Laboratories, 129
contamination effects on, 337, 343
Shanghai Inst. of Tech. Physics, 787, 791
dc flow, effect of, 163, 281, 299, 307, 315 design of 4K space pulse tube cooler, 221
Signaal USFA: harmonic simulation of PT, 359
876
Software: Aerospace CSIM model, 717 MS*2 Stirling code, 379 Sorption cryocooler: APD 10K cryostat for, 535 BETSCE, 1, 21, 535
microminiature, 110, 553
SUBJECT INDEX
STRV, 95 Sumitomo Heavy Industries, 575 Superconductor applications (see HTS applications) Superconducting Core Tech. (STC), 181
Superconductor Technologies Inc. (STI), 111 Superfluid helium as working fluid, 655
Swales & Associates, 671
porous plug J-T valve for, 545
Space experiments: BETSCE, 535
CRYOTSU, 671 NICMOS, 431, 439 STRV, 95
Space instruments: AIRS, 119, 771 Chinese, 787, 791 cooler integration trades for, 1 HIRDLS, 761 INTEGRAL, 781
missions to Mars, 815 MTI, 129 Planck, 795 SR&DB of the Ukraine:
5-10K hybrid J-T/Stirling cooler, 529 65K Stirling cooler, 87 84-90K J-T refrigerator, 467
accel. tests for contamination, 743 cooler outgassing rate, 733 gas mixtures for J-T coolers, 487 program overview, 55 thermal storage design, 689 Stirling cryocoolers:
Texas Instruments tactical Stirling coolers: 1W-80K performance characterization, 95 Thermal conductivity (see Conductance measurements) Thermal storage, 87, 467, 529, 671,689, 697
Thermal switch (see Heat switch) Throttle-cycle (see J-T cryocoolers) Tokyo Institute of Tech., 581, 587 Toshiba Corporation:
4K GM refrig. optimization, 581 4K space pulse tube cooler, 221 multi-layer mag. regen. for 4K GM, 611 pressure ratio effect on 4K GM, 587 working pressure effect on PT, 227 TRW cryocooler activities: 3503 pulse tube cooler, 1, 21 3585 pulse tube cooler, 1, 21, 33 6020 pulser tube cooler, 1, 21, 33 AIRS pulse tube cooler, 1, 119, 771 IMAS pulse tube cooler, 1, 139 miniature pulse tube coolers, 1, 21
MTI pulse tube cooler, 1, 21, 129 Turbo Brayton coolers (see reverse Brayton coolers)
Ball 1.5W-55K HIRDLS, 761
BEI0.3W 78K tactical, 105
Univ. of Calgary, 379 Univ. of California, UCLA, 275
Creare 65K diaphragm, 33
Univ. of California, UCSB, 497, 545
drive and control electronics, 791
Univ. of Dresden, 265, 475
miniature micromachined, 663
Univ. of Karlsruhe, 369
Ball 35/60K three-stage, 59
MMS/RAL 10K 2-stage, 67 SR&DB 65K with thermal storage, 87 STC for laser cooling, 77 STI HTS, 111 Texas Instruments 1W-80K, 95 variable reluctance linear motor for, 111 with superfluid He working fluid, 655 Stirling Technology Co., 77
Univ. of Kogakuin, 611 Univ. of Tsukuba, 321, 329 Univ. of Twente, 553, 837 Univ. of Wisconsin, 387 Vortex tube as pulse tube heat exchanger, 257 Wave cooler modeling, 405
