CRYOCOOLERS 12
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A publication of the International Cryocooler Conference
CRYOCOOLERS 12
Edited by
Ronald 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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0-306-47919-2 0-306-47714-9
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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. One class of pulse tubes that has reached maturity is referred to as “Stirling type” because they are based on the linear Oxford Stirling-cooler type compressor; these generally provide cooling in the 30 to 100 K temperature range and operate at frequencies from 30 to 60 Hz. The other type of pulse tube cooler making great advances is the so-called “Gifford-McMahon type.” Pulse tube coolers of this type use a G-M type compressor and lower frequency operation to achieve temperatures in the 2 to 10 K temperature range. Nearly a third of this proceedings covers these new developments in the pulse tube arena. Complementing the work on low-temperature pulse tubes is substantial continued progress on rare earth regenerator materials and Gifford-McMahon coolers. These technologies continue to make great progress in opening up the 2 - 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 to achieve low-cost cryocooler systems in the 65 to 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; the significant development here is the cost-effective incorporation of Oxford-like flexure spring piston supports so as to achieve an extended-life, low-cost product. The objective of Cryocoolers 12 is to archive these latest developments and performance measurements by drawing upon the work of the leading international experts in the field of cryocoolers. In particular, this book is based on their contributions at the 12th International Cryocooler Conference, which was held in Cambridge, Massachusetts, on June 18-20, 2002. The program of this conference consisted of 120 papers; of these, 105 are published here. Although this is the twelfth 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 fifth volume in this new series of hardcover texts for users and developers of cryocoolers. 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 is
v
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PREFACE
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 12 is organized into 20 chapters, starting first with an introductory chapter providing summaries of major government cryocooler development and test programs. The next several chapters address cryocooler technologies organized by type of cooler, starting with regenerative coolers; these include Stirling cryocoolers, pulse tube cryocoolers, Gifford-McMahon cryocoolers, thermoacoustic refrigerators, and associated regenerator research. Next, recuperative cryocoolers including Brayton, Joule-Thomson, and sorption cryocoolers are covered. The technology-specific chapters end with a chapter on unique sub-Kelvin and optical refrigerators. The last three chapters of the book deal with cryocooler integration technologies and experience to date in a number of representative space and commercial 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 the Massachusetts Institute of Technology, Department of Mechanical Engineering, which hosted the 12th 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 of the 12th ICC included:
CONFERENCE CO-CHAIRS John Brisson, MIT Peter Kerney, Conductus
CONFERENCE ADMINISTRATOR Doris Elsemiller, MIT
PUBLICATIONS Ron Ross, Jet Propulsion Laboratory
TREASURER Ray Radebaugh, NIST
PROGRAM CHAIRMAN William Burt, TRW PROGRAM COMMITTEE Robert Boyle, NASA / GSFC Peter Kittel, NASA / ARC Ralph Longsworth, APD Cryogenics Rod Oonk, Ball Aerospace Jeff Raab, TRW Alain Ravex, Air Liquide Klaus Timmerhaus, Univ. of Colorado Mark Zagarola, Creare
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 Joseph L. Smith, Jr., Michelle Douglas, Catherine Kerney, and the following staff of MIT Conference Services: Eva Marie Cabone, Virgina Lauricella, Cathi D. Levine, and Marie Seamon.
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Contents 1
Government Cryocooler Development Programs NASA Space Cryocooler Programs—An Overview
1
R.G. Ross, Jr., Jet Propulsion Lab, Pasadena, CA; and R.F. Boyle, NASA GSFC, Greenbelt, MD
Air Force Research Laboratory Space Cryogenic Technology Research Initiatives
9
B.J. Tomlinson, B. Flake, and T. Roberts, Air Force Research Laboratory, Kirtland AFB, NM
Status Report on the Linear Drive Coolers for the Department of Defense Standard Advanced Dewar Assembly (SADA)
17
W.E. Salazar, US Army Night Vision, Fort Belvoir, VA
Space Stirling Cryocooler Developments Development of a Low-Power Stirling Cycle Cryocooler for Space Applications
27 27
J.S. Reed and G.D. Peskett, Univ. of Oxford, Oxford, UK
High Capacity Flexure Bearing Stirling Cryocooler On-Board the ISS
31
T. Trollier, A. Ravex, and P. Crespi, Air Liquide DTA, Sassenage, France; J. Mullié, P. Bruins, and T. Benschop, Thales Cryogenics, Eindhoven, The Netherlands
Space Flight Qualification Program for the AMS-02 Commercial Cryocoolers
37
K.A. Shirey, I.S. Banks, S.R. Breon, and R.F. Boyle, NASA/Goddard Space Flight Center, Greenbelt, MD
Thermodynamic Performance of the Ball Aerospace Multistage Stirling Cycle Mechanical Cooler
45
W.J. Gully, D. Glaister, E. Marquardt, R. Stack, and G.P. Wright, Ball Aerospace, Boulder, CO
Performance Characterization of the Ball Aerospace 35/60K Protoflight Spacecraft Cryocooler
51
C.H.Y. Bruninghaus and B.J. Tomlinson, AFRL, Kirtland AFB, NM; and N. Abhyankar, Dynacs Engineering, Albuquerque, NM
Continued Characterization Results for the Astrium 10K Developmental Cryocooler
59
S.A. Yarbrough, B.A. Flake, and B.J. Tomlinson, AFRL, Kirtland AFB, NM; and N. Abhyankar, Dynacs Engineering, Albuquerque, NM
ix
x
CONTENTS
Tactical and Commercial Stirling Cryocoolers Technology Development Related to Tactical Cryocoolers at Raytheon Infrared Operations
69 69
B.A. Ross, M.L. Brest, and F.I. Mirbod, Raytheon Infrared Operations, Goleta, CA
Performance and Reliability Data for a Production Free Piston Stirling Cryocooler
75
M. Hanes and A. O’baid, Superconductor Technologies, Inc., Santa Barbara, CA
CMCEC Life Test Results and Related Issues
79
S.W.K. Yuan, D.T. Kuo, and T.D. Lody, CMC Electronics, Sylmar, CA
MTTF Prediction in Design Phase on Thales Cryogenics Integral Coolers
87
J.M. Cauquil, J.Y. Martin, Thales Cryogenics, Blagnac, France; and P. Bruins, T. Benschop, Thales Cryogenics, Eindhoven, The Netherlands
An Experimental Study of the Phase Shift between Piston and Displacer in a Stirling Cryocooler
95
S.J. Park , Y.J. Hong, H.B. Kim, D.Y. Koh, Korea Inst. of Mach. & Mat'ls, Taejeon, Korea; B.K. Yu, Wooyoung, Seoul, Korea; and K.B. Lee, Pusan Nat’l Univ., Pusan, Korea
Dynamic Analysis of a Free Piston Stirling Refrigerator
103
Y-J Hong, S-J Park, H-B Kim, D-Y Koh, Korea Inst. of Mach. and Materials, Taejeon, Korea
Tactical and Commercial Pulse Tube Cryocoolers Low Vibration 80 K Pulse Tube Cooler with Flexure Bearing Compressor
109 109
P.C. Bruins, A. de Koning, and T. Hofman, Thales Cryogenics, Eindhoven, The Netherlands
Development of 40-80K Linear-Compressor Driven Pulse Tube Cryocoolers
115
J. Liang, J.H. Cai, Y. Zhou, W.X. Zhu, L.W. Yan, W. Jing, Y.L. Ju, Y.K. Hou, and K. Yuan, Chinese Academy of Sciences, Beijing, China
Performance and System Design of 60K Pulse Tube Coolers Driven by a Linear Compressor for HTS Filter Subsystems
123
Y.L. Ju, K. Yuan, Y.K. Hou, W. Jing, J.T. Liang, and Y. Zhou, Chinese Acad. of Sciences, Beijing, China
High Capacity Pulse Tube Cryocooler
131
I. Charles, J.M. Duval, and L. Duband, CEA/SBT, Grenoble, France; T. Trollier and A. Ravex, Air Liquide DTA, Sassenage, France; and J. Y. Martin, Thales Cryogenics, Blagnac, France
Development of Single and Two-Stage Pulse Tube Cryocoolers with Commercial Linear Compressors
139
K.B. Wilson , Sunpower, Inc., Athens, Ohio; and D.R. Gedeon, Gedeon Assoc., Athens, Ohio
Development of a 5 W at 80 K Stirling-Type Pulse Tube Cryocooler L.W. Yang, G. Thummes, Univ. of Giessen, Giessen, Germany; N. Rolff, H.U. Häfner, Leybold Vacuum, Cologne, Germany
149
CONTENTS
Design and Test of a 70 K Pulse Tube Cryocooler
xi
157
Y. Yasukawa, K. Ohshima, K. Toyama, T. Itoyama, Y. Tsukahara, R. Kikuchi, and N. Matsumoto, Fuji Electric Corp. R&D; and T. Kamoshita and T. Takeuchi, Fuji Electric Co., Tokyo, Japan
Space Pulse Tube Cryocooler Developments Miniature 50 to 80 K Pulse Tube Cooler for Space Applications
165 165
T. Trollier and A. Ravex, Air Liquide DTA, Sassenage, France; I. Charles and L. Duband, CEA/SBT, Grenoble, France; J. Mullié, P. Bruins, and T. Benschop, Thales Cryogenics, Eindhoven, The Netherlands; M. Linder, ESA/ESTEC, Noordwijk, The Netherlands
Design and Characterization of a Miniature Pulse Tube Cooler
173
A.S. Gibson, R. Hunt, Astrium UK, Stevenage, UK; I. Charles, L. Duband, CEA/SBT, Grenoble, France; M. Crook, A.H. Orlowska, T.W. Bradshaw, RAL, Chilton, UK; and M. Linder, ESA/ESTEC, Noordwijk, The Netherlands
Low Cost, Lightweight Space Cryocoolers
183
C.S. Kirkconnell, G.R. Pruitt, K.D. Price, Raytheon ES, El Segundo, CA; and B.A. Ross and W.R. Derossett, Raytheon IS, Santa Barbara, CA
JAMI Flight Pulse Tube Cooler System
191
J. Raab, R. Colbert, J. Godden, D. Harvey, R. Orsini, G. Toma; TRW, Redondo Beach, CA
Performance Testing of a Lightweight, High Efficiency Cooler
199
L. J. Salerno, P. Kittel, NASA/ARC, Moffett Field, CA; and B.P.M. Helvensteijn, A. Kashani, Atlas Scientific, San Jose, CA
Development of a Lightweight Pulse Tube Cryocooler for Space Applications
205
T. Nast, J. Olson, P. Champagne, B. Evtimov, Lockheed Martin ATC, Palo Alto, CA.; T. Renna, Lockheed Martin CSC, Newtown, PA; and G. Sarri and C. Hernandez, ESA/ESTEC, Noordwijk, The Netherlands
Development of a Two-Stage Pulse Tube Cryocooler for 35 K Cooling
213
T.C. Nast, J. Olson, B. Evtimov, and V. Kotsubo, Lockheed Martin ATC, Palo Alto, CA
High Capacity Two-Stage Pulse Tube Cooler
219
C.K. Chan, T.Nguyen, and C. Jaco, TRW, Redondo Beach, CA; B.J. Tomlinson and T. Davis, AFRL, Kirtland AFB, NM
Development of a High Capacity Two-Stage Pulse Tube Cryocooler
225
W.G. Foster, J. Olson, P. Champagne, B. Evtimov, E. Will, A. Collaco, and T. Nast, Lockheed Martin ATC, Palo Alto, CA; R. Clappier, Clappier Consulting, Discovery Bay, CA; A. Mitchell and D. Jungkman, Northrup Grumman, Baltimore, MD; R. Radebaugh, NIST, Boulder, CO; and D.G.T. Curran, Aerospace Corp., El Segundo, CA
Two Stage Hybrid Cryocooler Development
233
K.D. Price and C.S. Kirkconnell, Raytheon ES, El Segundo, CA
Development of a 10 K Pulse Tube Cryocooler for Space Applications
241
J. Olson, T.C. Nast, B. Evtimov, and E. Roth, Lockheed Martin ATC, Palo Alto, CA
Linear Compressor Development and Modeling Scaling of Cryocooler Compressors P.B. Bailey and M.W. Dadd, Oxford Univ., Oxford, UK; C.F. Cheuk and N.G. Hill, Hymatic Engineering, Redditch, UK; and J. Raab, TRW, Redondo Beach, CA
247 247
CONTENTS
xii
The Linearity of Clearance Seal Suspension Systems
255
M.W. Dadd, P.B. Bailey, and G. Davey, Oxford Univ., Oxford, UK; T. Davis, B.J. Thomlinson, AFRL, Albuquerque, NM
Piston Resonance in the Orifice Pulse Tube
265
P.C.T. de Boer, J.-M. Duval, I. Charles, and L. Duband, CEA-Grenoble, France
Producibility of Cryocooler Compressors
275
C.F. Cheuk, N.G. Hill, R Strauch, Hymatic Engineering, Redditch, UK; P.B. Bailey, Oxford Univ., Oxford, UK; J. Raab, TRW, Redondo Beach, CA
GM-Type Pulse Tube Coolers for Low Temperatures Helium-3 Pulse Tube Cryocooler
283 283
I.A. Tanaeva and A.T.A.M. De Waele, Eindhoven Univ. of Tech., The Netherlands
Two-Stage Pulse Tube Cryocoolers for 4 K and 10 K Operation
293
C. Wang and P.E. Gifford, Cryomech, Syracuse, NY
Development of a 4K Two-Stage Pulse Tube Cryocooler
301
M.Y. Xu, P.D. Yan, T. Koyama, T. Ogura, R. Li, Sumitomo Heavy Industries, Tokyo, Japan
Performance of a 4K Pulse Tube Refrigerator and Its Improvement
309
S.W. Zhu, M. Nogawa, S. Katsuragawa, M. Ichikawa, T. Inoue, Aisin Seiki Co., Aichi, Japan
Experimental Investigation of a G-M Type Coaxial Pulse Tube Cryocooler
317
K. Yuan, J.T. Liang, Y.L. Ju, Chinese Academy of Sciences, Beijing, P.R. China
Hybrid Cryocoolers Using Pulse Tubes Experimental Study on Two-Stage Pulse Tube Refrigeration with Mixtures of Helium and Hydrogen
325 325
N. Jiang, Z.H. Gan, G.B. Chen, L.M. Qiu, Y.L. Jiang, Y.L. He, and N. Li, Zhejiang University, Hangzhou, China
Experimental Investigation of 4K VM Type Pulse Tube Cooler
331
W. Dai, Y. Matsubara, and H. Kobayashi, Nihon Univ., Funabashi, Japan
Affecting the Gross Cooling Power of a Pulse Tube Cryocooler with Mass Flow Control
337
A. Waldauf, M. Thürk, and P. Seidel, FSU, Jena, Germany; and T. Schmauder, Leybold Optics, Hanau, Germany
Pressure Wave Generator for a Pulse Tube Cryocooler
343
Y. Matsubara, W. Dai, Nihon Univ., Funabashi, Japan; and H. Sugita and S. Tooyama, NASDA, Tsukuba, Japan
A First Order Model of a Hybrid Pulse Tube/Reverse-Brayton Cryocooler G.F. Nellis, J.H. Baik, and J.M. Pfotenhauer, Univ. of Wisconsin, Madison, WI; and J.R. Maddocks and A. Kashani, Atlas Scientific, San Jose, CA
349
CONTENTS
Pulse Tube Analyses and Experimental Measurements The Role of the Orifice and the Secondary Bypass in a Miniature Pulse Tube Cryocooler
xiii
361 361
Y.K. Hou, Y.L. Ju, W. Jing, and J.T. Liang, Chinese Acad. of Sciences, Beijing, China
Surface Heat Pumping Loss in a Pulse Tube Refrigerator
371
J. Jung and S. Jeong, Adv. Institute of Science and Tech., Taejon, Korea
Numerical Model for Pulse Tubes using Method of Lines
379
A. Schroth and M. Sahimi, USC, Los Angeles, CA; and C. Kirkconnell, Raytheon ES, El Segundo, CA
Pulse Tube Refrigerator Analysis, Including Inertance Tube and Friction in the Regenerator
389
L. Fang, X. Deng, and L. Bauwens, Univ. of Calgary, Calgary, Canada
GM Refrigerator Developments Cooling Performance of a Small GM Cryocooler with a New Ceramic Magnetic Regenerator Material
397 397
T.Satoh, Sumitomo Heavy Industries, Ltd., Yokosuka, Japan; and T. Numazawa, Tsukuba Magnet Lab., Nat’l Inst. for Materials Science, Tsukuba, Japan
Improvement of 4K GM Cooling Performance with a New Regenerator Material
403
Y. Ikeya and R. Li, Sumitomo Heavy Industries, Tokyo, Japan; and T.Numazawa, Tsukuba Magnet Lab/NIMS, Tsukuba, Japan
Thermal Hysteresis at 4 K with a GM Cryocooler
411
G.E. Bonney, Advanced Research Systems, Inc., Allentown, PA
Thermoacoustic Refrigerator Investigations Study on the Onset Temperature Gradient of Regenerators Used for Thermoacoustic Prime Movers
421 421
H. Ling, E. Luo, J. Wu, M. Yang, X. Li, Chinese Academy of Sciences, Beijing, China
Study on Thermoacoustic DC-Flow Model for a Cyclic Regenerator due to Nonlinear Effects
425
E. Luo, Chinese Academy of Sciences, Beijing, China
Thermodynamic Analysis of a Traveling Wave Thermoacoustic Device by Use of a Distributed-Parameter Method
431
M. Yang, X. Li, and G. Chen, Zhejiang Univ., Hangzhou, China; and E. Luo, H. Ling, and J. Wu, Chinese Academy of Sciences, Beijing, China
Investigation of a High Frequency Travelling Wave ThermoacousticDriven System
439
Q. Li and Z. Yu, CAS, Beijing, China; and J. Wu, F. Guo, and Q. Tu, Huazhong Univ. of Science and Tech., Wuhan, China
The Influence of Thermal Natural Convection on a Traveling-Wave Thermoacoustic Engine H. Liu, E. Luo, H. Ling, and J. Wu, Chinese Academy of Sciences, Beijing, China
447
CONTENTS
xiv
Experimental Investigation of Thermoacoustically Driven Pulse Tube Refrigerator Using Noble Gas Mixtures
451
G.B. Chen, K. Tang, T. Jin, Y Shen, and Y.H. Huang, Zhejiang Univ., Hangzhou, China
Regenerator Materials Development Low Temperature Cryocooler Regenerator Materials
457 457
K.A. Gschneidner, Jr., A.O. Pecharsky, and V.K. Pecharsky, Ames Lab, Iowa State Univ., Ames, Iowa
Specific Heat and Magnetic Properties of GdSb
467
H. Nakane, S. Yamazaki, T. Yamaguchi, S. Yoshizawa, and T. Numazawa, Kogakuin Univ. and Meisei Univ., Tokyo, Japan
A New Ceramic Magnetic Regenerator Material for 4 K Cryocoolers
473
T. Numazawa, Tsukuba Magnet Lab, Tsukuba, Japan; T. Yanagitani and H. Nozawa, Konoshima Chemical Co, Kagawa, Japan; and Y. Ikeya, R. Li, and T. Satoh, Sumitomo Heavy Industries, Tokyo, Japan
Predicted Performance of a Low-Temperature Perforated Plate Regenerator
483
J.B. Hendricks, Alabama Cryogenic Engineering, Huntsville, Alabama
LIGA-Fabricated High-Performance Micro-Channel Regenerators for Cryocoolers
489
K. Kelly, A. McCandless, and S. Motakef, Mezzo Systems, Baton Rouge, LA
Improved Flow Patterns in Etched Foil Regenerator
499
M.P. Mitchell, Mitchell/Stirling, Berkeley, CA; and D. Fabris, Santa Clara Univ., Santa Clara, CA
Compact High Effectiveness Parallel Plate Heat Exchangers
507
E.D. Marquardt, Ball Aerospace, Boulder, CO; and R. Radebaugh, NIST, Boulder, CO
Regenerator Performance Analyses and Tests Measurement of Heat Conduction through Bonded Regenerator Matrix Materials
517 517
M.A. Lewis and R. Radebaugh, National Institute of Standards and Technology, Boulder, CO
Regenerator Loss Measurements at Low Temperatures and High Frequencies
523
J.M. Pfotenhauer, Univ. of Wisconsin, Madison, WI; and P.E. Bradley, M.A. Lewis, R. Radebaugh, NIST, Boulder, CO
Regenerator Characterization under Oscillating Flow and Pulsating Pressure
531
S. Jeong, K. Nam, and J. Jung, Korea Adv. Inst. of Science and Tech., Daejon, Korea
Oscillating Flow Characteristics of a Regenerator under Low Temperature Conditions K. Yuan, L. Wang, Y.K. Hou, Y. Zhou, J.T. Liang, Y.L. Ju, Chinese Academy of Sciences, Beijing, P.R. China
539
xv
CONTENTS
A Comparative Evaluation of Numerical Models for Cryocooler Regenerators
547
J.P. Harvey and P.V. Desai, Georgia Inst. of Tech., Atlanta, GA; and C.S. Kirkconnell, Raytheon ES, El Segundo, CA
Periodic Porous Media Flows in Regenerators
555
T. Roberts, AFRL, Kirtland AFB, NM; and P. Desai, Georgia Inst. of Tech., Atlanta, GA
Turbo-Brayton Cryocooler Developments Initial Operation of the NICMOS Cryocooler on the Hubble Space Telescope
563 563
W.L. Swift, J.A. McCormick, J.J. Breedlove, F.X. Dolan, and H. Sixsmith, Creare, Hanover, NH
Development of a Turbo-Brayton Cooler for 6 K Space Applications
571
M.V. Zagarola, W.L. Swift, H. Sixsmith, J.A. McCormick, M.G. Izenson, Creare, Hanover, NH
J-T and Throttle-Cycle Cryocooler Developments A Hybrid, Multistage 10K Cryocooler for Space Applications
579 579
R. Levenduski, J. Lester, Redstone Engineering, Carbondale, Colorado; and E. Marquardt, Ball Aerospace, Boulder, CO
Development of a Medium-Scale Collins-Type 10 K Cryocooler
587
C.L. Hannon and J. Gerstmann, Advanced Mechanical Technology, Inc., Watertown, MA; and M. Traum, J.G. Brisson, and J.L. Smith Jr., MIT, Cambridge, MA
Efficiency of the ARC and Mixed Gas Joule Thomson Refrigerators
595
A. Alexeev, Messer Cryotherm, Kirchen, Germany; and D. Goloubev, E. Mantwill, TU Dresden, Dresden, Germany
Further Development of the Mixture Refrigeration Cycle with a Dephlegmation Separator
603
M.Q. Gong, J.F. Wu, E.C. Luo, Y.F. Qi, Q.G. Hu, and Y. Zhou; Chinese Academy of Sciences, Beijing, China
Research on Adiabatic Capillary Tube Expansion Devices in MixedRefrigerant J-T Cryocoolers
609
Y.F. Qi, M.Q. Gong, E.C. Luo, J.F. Wu, and Y. Zhou; Chinese Academy of Sciences, Beijing, China; and Y. Cao, Zhejiang Univ., Hangzhou, China
Study of a Vortex Tube by Analogy with a Heat Exchanger
615
Y. Cao, G.M. Chen, Zhejiang Univ., Hangzhou, China; Y.F. Qi, E.C. Luo, J.F. Wu, M.Q. Gong, Chinese Academy of Sciences, Beijing, China
Thermodynamic Prediction of the Vortex Tube Applied to a MixedRefrigerant Auto-Cascade J-T Cycle Y. Cao, G.M. Chen, Zhejiang Univ., Hangzhou, China; M.Q. Gong, Y.F. Qi, E.C. Luo, J.F. Wu, Chinese Academy of Sciences, Beijing, China
621
CONTENTS
xvi
Sorption Cryocooler Developments
627
Evaluation of Hydride Compressor Elements for the Planck Sorption Cryocooler
627
R.C. Bowman, Jr., M. Prina, D.S. Barber, P. Bhandari, A.S. Loc, and M.E. Schmelzel, JPL, Pasadena, CA; D. Crumb and J.W. Reiter, Swales Aerospace, Pasadena, CA; and G. Morgante, IASF/CNR, Bologna, Italy
Initial Test Performance of a Closed-Cycle Continuous Hydrogen Sorption Cooler, the Planck Sorption Breadboard Cooler
637
M. Prina, A. Loc, M.E. Schmelzel, D. Pearson, J. Borders, R.C. Bowman, A. Sirbi, P. Bhandari, L.A. Wade, A. Nash, JPL, Pasadena, CA; and G. Morgante, CNR-IASF, Bologna, Italy
Construction and Operation of a 165 K Microcooler with a Sorption Compressor and a Micromachined Cold Stage
643
J.F. Burger, H.J. Holland, H.J.M. ter Brake, M. Elwenspoek, and H. Rogalla, Univ. of Twente, The Netherlands
Sub-Kelvin and Optical Refrigerator Developments Automated Closed-Cycle Cooling to 250 mK for the Polatron
651 651
R.S. Bhatia, V.V. Hristov, B.G. Keating, A.E. Lange, P.V. Mason, B.J. Philhour, G. Sirbi, and K.W. Yoon, Caltech, Pasadena, CA; and S.T. Chase, Chase Research, Sheffield, UK
Progress in the Development of a Continuous Adiabatic Demagnetization Refrigerator
661
P.J. Shirron, E.R. Canavan, M.J. DiPirro, J. Francis, M. Jackson, T.T. King, and J.G. Tuttle, NASA/Goddard Space Flight Center, Greenbelt, MD
Sub-Kelvin Mechanical Coolers
669
A. Ravex and P. Hernandez, Air Liquide DTA, Sassenage, France; and L. Duband, CEA/SBT, Grenoble, France
Preliminary Performance of a Superfluid
Compressor
675
F.K. Miller and J.G. Brisson, Massachusetts Institute of Technology, Cambridge, MA
Preliminary Experimental Results Using a Three-Stage Superfluid Stirling Refrigerator
681
C. Phillips and J.G. Brisson, Massachusetts Institute of Technology, Cambridge, MA
Dielectric Mirror Leakage and Its Effects on Optical Cryocooling
687
G. Mills, J. Fleming, Z. Wei, Ball Aerospace, Boulder, CO; and J. Turner-Valle, Optical Engineering Assoc., Longmont, CO
Cryocooler Integration Technologies Advanced Components for Cryogenic Integration
693 693
D. Bugby, B. Marland, C. Stouffer, and E. Kroliczek, Swales Aerospace, Beltsville, MD
Cryogenic Loop Heat Pipes as Flexible Thermal Links for Cryocoolers D. Khrustalev, Thermacore International, Lancaster, PA
709
CONTENTS
A Thermally Conductive and Vibration Protective Interface for Linear Cryogenic Coolers in Applications for Gimbaled Infrared Devices
xvii
717
A.M. Veprik, V.I. Babitsky, Loughborough Univ., Loughborough, UK; and S.V. Riabzev, N. Pundak, Ricor, Ltd., Israel
Space Cryocooler Applications Cryocooler Load Increase due to External Contamination of Cryogenic Surfaces
727 727
R.G. Ross, Jr., Jet Propulsion Lab, Pasadena, CA
Performance Characteristics of the ASTER Cryocooler in Orbit
737
M. Kawada, NIAIST, Tsukuba, Japan; H. Akao, M. Kobayashi, S. Akagi, Mitsubishi Electric, Kamakura, Japan; T. Maekawa, O. Nishihara, Fujitsu, Ltd., Kawasaki, Japan; M. Kudoh, JROSO, Tokyo, Japan; and H. Fujisada, Sensor Information Lab, Tsukuba, Japan
AIRS Pulse Tube Cooler System-Level and In-Space Performance Comparison
747
R.G. Ross, Jr., Jet Propulsion Lab, Pasadena, CA
Final Qualification and Early On-Orbit Performance of the RHESSI Cryocooler
755
R. Boyle, S. Banks, and K. Shirey, NASA/Goddard Space Flight Center, Greenbelt, MD
Operation of a Sunpower M87 Cryocooler in a Magnetic Field
761
S.R. Breon, K.A. Shirey, I.S. Banks, B.A. Warner, R.F. Boyle and S. Mustafi, NASA/ Goddard Space Flight Center, Greenbelt, MD
Active Vibration Cancellation in Astrium Stirling Cycle and Pulse Tube Coolers
771
S. Akhtar and R. Hunt, Astrium, Stevenage, UK
On-Orbit Cooling Performance of a Miniature Pulse Tube Flight Cryocooler
777
D.R. Ladner, Lockheed Martin Astronautics Operations, Denver, CO; R. Radebaugh, P.E. Bradley, M. Lewis, NIST, Boulder, CO; P. Kittel, NASA/ARC, Moffett Field, CA; and J. H. Xiao, Ethicon, Somerville, NJ
Commercial Cryocooler Applications High-Tc Squid-Based Gradiometer Cooled by a Cryotiger Gas-Mixture Cooler
789 789
A.P. Rijpma, H.J.M. ter Brake, H.J. Holland, and H. Rogalla, Univ. of Twente, The Netherlands
On the Development of a Non-Metallic and Non-Magnetic Miniature Pulse Tube Cooler
799
H.Z. Dang, Y.L. Ju, J.T. Liang, and Y. Zhou, Chinese Academy of Sciences, Beijing, China
Cryogenic Refrigerator Evaluation for Medical and Rotating Machine Applications R.A. Ackermann, General Electric R&D, Niskayuna, NY; D.A. Grey, GE Medical Systems, Florence, SC; and S. Funayama and K. Ito, Sumitomo Heavy Industries, Tokyo, Japan
805
CONTENTS
xviii
Helium Free Magnets and Research Systems
813
J. Good, S. Hodgson, R. Mitchell, and R. Hall, Cryogenic, Ltd., London, UK
Indexes
817
Proceedings Index
817
Author Index
819
Subject Index
821
NASA Space Cryocooler Programs—An Overview R.G. Ross, Jr.† and R.F. Boyle* †
Jet Propulsion Laboratory California Institute of Technology Pasadena, CA 91109 * NASA Goddard Space Flight Center Greenbelt, MD 20771 USA
ABSTRACT Mechanical cryocoolers represent a significant enabling technology for NASA’s Earth and Space Science Enterprises. An overview is presented of ongoing efforts at the Jet Propulsion Laboratory (JPL) and Goddard Space Flight Center (GSFC) in support of current flight projects, near-term flight instruments, and long-term technology development. Highlights of the past year include the launch into space of three new cryocooler systems aboard NASA missions: 1) a Sunpower 80 K Stirling cooler on the RHESSI gamma-ray spectrometer spacecraft launched February 5, 2002, 2) an 80 K Creare turbo-Brayton cooler added to the NICMOS instrument during the Hubble Space Telescope servicing mission of March 1-12, 2002, and 3) a pair of TRW 55 K pulse tube coolers on the AIRS instrument aboard the EOS Aqua platform launched May 4, 2002. In addition, a major NASA cryocooler development initiative referred to as the Advanced Cryocooler Technology Development Program (ACTDP) was kicked off with four parallel industry contracts in April, 2002. The ACTDP concepts are required to provide dual cooling at 6K and 18 K and are focused at NASA low temperature applications.
INTRODUCTION NASA programs in Earth and space science observe a wide range of phenomena, from atmospheric physics and chemistry to stellar birth. Many of the instruments require low-temperature refrigeration to enable use of cryogenic detector technologies that increase sensitivity, improve dynamic range, or to extend wavelength coverage. Over the last two decades, NASA has funded cryocooler technology development in support of many projects, and has also taken advantage of coolers developed under Defense Department and commercial funding.1 The largest utilization of coolers is currently in Earth Science instruments operating at medium to high cryogenic temperatures (50 to 80 K), reflecting the relative maturity of the technology at these temperatures. However, in support of studies of the origin of the universe and the search for planets around distant stars, interest has peaked in systems of low temperature refrigerators providing cooling down to 100 mK. NASA's development of a 20 K cryocooler for the European Planck spacecraft and its new Advanced Cryocooler Technology Development Program (ACTDP) for 6-18 K coolers are examples of the thrust to provide low temperature cooling for this class of missions. Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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GOVERNMENT CRYOCOOLER DEVELOPMENT PROGRAMS
COOLERS ON NEAR-TERM EARTH AND SPACE SCIENCE MISSIONS During the first four months of 2002 we have seen three new cryocooler systems launched into space to support NASA missions. Two of the three are based at least partially on the Oxford cooler technology that first flew on the Improved Stratospheric and Mesospheric Sounder (ISAMS) instrument in 1991; this type of cooler has demonstrated the potential for multi-year lifetime, and has been adopted by many long-life instruments to enable new and improved science. The third cooler, the NICMOS cooler, is the first space application of a turbo-Brayton cooler. These recently launched coolers, which are reviewed below, build upon the coolers of earlier NASA missions, such as those on the ISAMS, MOPITT and Hyperion instruments that have been described previously.2 Additional coolers, such as the TRW pulse tube coolers on the TES instrument and the Ball Aerospace Stirling cooler on the HIRDLS instrument, are in the queue for launch aboard NASA missions in the next couple of years and are also described below.
RHESSI Gamma-Ray Spectrometer The Ramaty High-Energy Solar Spectroscopic Imager (RHESSI) uses an array of nine large germanium gamma-ray detectors to observe solar flares from 3 keV to 25 GeV. The detector array is cooled to 75 K by a Sunpower M77B Stirling cooler (Fig. 1) running at 65K.3,4 Additionally, the cooler uses a heat intercept strap clamped to the Stirling coldfinger to provide simultaneous cooling to the instrument's higher temperature radiation shields at 155 K. This technique thus provides the capability of a two-stage cooler with an off-the-shelf single-stage cooler. Launched in February 2002, the cooler is maintaining the gamma-ray detectors at their required 75 K temperature, with a goal of continuing to gather science for up to two years on orbit.5 This mission also represents the first attempt to use a low-cost commercial cooler to achieve multiyear operation in space. NICMOS Cooling System The Creare NICMOS Cooling System (NCS) was designed to cool the Near Infrared Camera and Multi-Object Spectrometer (NICMOS) instrument of the Hubble Space Telescope. It was successfully launched and integrated into the Hubble Space Telescope during the HST's fourth servicing mission (SM3B) in March 2002.6 This followed an earlier flight test of the entire cooling system aboard a week-long Shuttle mission in October 1998.7 The NCS, shown in Fig. 2, is a large turbo-Brayton cooler with a nominal cooling capacity of 7 W at 80 K with 400 W input power. The NICMOS instrument was originally launched in 1997 using a solid Nitrogen dewar to cool its sensitive infrared focal plane. However, after a dewar failure led to a shorter than expected on-orbit life, the NCS was identified as a way to extend the mission and recover the original science objectives. Thus, the NCS was developed to be retrofitted to the NICMOS instrument during a subsequent HST servicing mission in space.
NASA SPACE CRYOCOOLER PROGRAM OVERVIEW
3
The NCS is designed to maintain the instrument’s detectors in the range of 75-85K by circulating refrigerated neon gas through the NICMOS dewar’s existing liquid helium freeze lines. The very large (7 W at 80 K) heat load is associated with the inefficiencies of using existing inspace Bayonet couplings on the NICMOS dewar to connect with the gas lines. To date, the system has performed flawlessly and the NICMOS instrument has been returned to its job of gathering infrared images of the far reaches of space.
Atmospheric Infrared Sounder (AIRS) instrument Another recently-launched NASA instrument with cryocoolers is the Atmospheric Infrared Sounder (AIRS) instrument. This instrument measures atmospheric air temperature using a HgCdTe focal plane operating at 58 K and is cooled by a redundant pair of 55 K TRW pulse tube coolers.8,9 Launched in May 2002 on NASA’s Earth Observing System Aqua platform, the instrument was designed and built under JPL contract by Lockheed Martin Infrared Imaging Systems, Inc. (now BAE Systems IR Imaging Systems) of Lexington, MA. Initiated in 1994, the cryocooler development effort was the first space application to select a pulse tube cryocooler. The highly collaborative development effort, involving cryocooler development at TRW and extensive cryocooler testing at JPL and Lockheed Martin, has served as the pathfinder for much of the pulse tube development to date. The AIRS flight pulse tube coolers, shown in Fig. 3, were originally delivered to JPL for testing in October 1997, and to the instrument for integration in January 1998. Since being launched in May 2002 the coolers have been performing flawlessly.10
TES Cooler Development The EOS Tropospheric Emission Spectrometer (TES) instrument is the next large cryogenic instrument presently under development at JPL. TES is an infrared instrument designed to measure the state of the earth’s troposphere. It is scheduled for launch into polar orbit aboard NASA’s third earth observing systems spacecraft (EOS-Aura) in the 2004 timeframe. TES uses two 57 K coolers to cool two separate focal planes to 62 K. The two coolers are identical and are a variant of the TRW AIRS pulse tube cooler, but configured with the pulse tube hard mounted to the compressor.11,12 The coolers were fabricated by TRW under contract to JPL, and have recently completed integration into the overall TES instrument. The instrument is scheduled to be integrated onto the TRW spacecraft later this year.
HIRDLS Cooler Development On the same spacecraft as the TES instrument, the High Resolution Dynamics Limb Sounder (HIRDLS) instrument uses a single-stage Stirling cryocooler manufactured by Ball Aerospace under contract to Lockheed Martin. The HIRDLS cooler, which provides 720 mW at 55 K for an infrared array covering 21 bands between 6-17µm, uses technology developed under a number of
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NASA and DoD contracts.13 It incorporates radial position sensors for establishing and monitoring the clearance seals in the cooler, prior to closeout of the housing. It is similar in design to a two-stage 30 K cooler delivered to GSFC in 1997, and life tested to 13,000 hours.
AMS-2 Charged-Particle Spectrometer A set of four Sunpower M87 coolers has been baselined to fly on the Alpha Magnetic Spectrometer–2 (AMS-2) mission in October 2004. The instrument, mounted on the International Space Station, will use a large superconducting magnet assembly in a search for antimatter nuclei from cosmic sources. The coolers will be used to intercept heat at the outer thermal shield on a 2500 liter helium tank. With a mass of over 2000 kg for the superconducting magnets and helium tank, it is extremely challenging to provide enough thermal isolation to allow a 3-year lifetime, even with the coolers operating at nominal power. The four coolers, each capable of 6-7 W of heat lift at 77 K, will be run at reduced power to provide a total of 20-25 W of cooling on the shield at 77 K. The coolers, operating in the stray field of the magnet system, will be specially qualified for operation in a magnetic field of 750-1000 gauss.
Planck Cooler Development As a precursor to the US low-temperature cryocooler missions, JPL is presently working on the development of a 1W at 18-20 K hydrogen sorption cryocooler for the Planck mission of the European Space Agency.14 The objective of the Planck mission is to produce very high resolution mapping of temperature anisotropy in the cosmic microwave background (CMB) radiation. Planck's Low Frequency Instrument (LFI) will have an array of tuned radio receivers based on High Electron Mobility Transistors (HEMTs) to detect radiation in the range 30-100 GHz. These receivers will be operated at a temperature of about 20 K. The High Frequency Instrument (HFI) will use bolometers operated at 0.1 K for frequencies from 100 GHz to 900 GHz. The redundant hydrogen sorption cryocoolers are being designed to cool the LFI detectors to 18 - 20 K and to precool the Rutherford Appleton Lab (RAL) 4 K helium J-T that cools the 0.1 K dilution refrigerators in the HFI cooling system. A successful test of the breadboard Planck sorption cooler was conducted in February 2002,15 following significant development of the refrigerators compressor elements.16 The flight coolers are currently in fabrication, with the first qualification/flight unit scheduled for delivery and instrument integration in early 2004, followed by the second flight unit a year later.
Other Applications Another NASA application for space cryocoolers is in propulsion systems. NASA’s Glenn Research Center and Ames Research Center are studying the use of cryocoolers to enable zeroboiloff storage of cryogenic propellants in space flight systems.17,18 At the Johnson Space Center, the Variable Specific Impulse Magnetoplasma Rocket (VASIMR) project is designing a system that will use high-temperature superconducting coils for plasma containment and acceleration.19
CRYOCOOLER DEVELOPMENT FOR FUTURE NASA MISSIONS Over the years, NASA has collaborated with the US Air Force to develop new cryocooler technologies for future space missions. Recent achievements include the NCS, AIRS, TES and HIRDLS cryocoolers described previously, and new smaller pulse tube coolers at TRW20, 21 and Lockheed Martin.22,23 The largest technology push within NASA right now is in the temperature range of 4-10K. Missions such as the Next Generation Space Telescope and Terrestrial Planet Finder plan to use infrared detectors operating between 6-8 K, typically arsenic-doped silicon arrays, with telescopes of greater than 5m diameter. Other missions call for large aperture telescopes operating as low as 4 K. Constellation-X plans to use X-ray microcalorimeters operating at 50 mK. Both NGST and TPF plan to passively cool their optics to 35 K, incorporating sophisticated sunshades and thermal isolation structures to minimize heat input, and incorporating large
NASA SPACE CRYOCOOLER PROGRAM OVERVIEW
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radiators to maximize heat rejection. This option is made possible by the orbits selected for these missions, well away from the thermally-disruptive presence of Earth.
Advanced Cryocooler Technology Development Program To develop the needed cryocooler technology for this mission set NASA has initiated the Advanced Cryocooler Technology Development Program (ACTDP) under the leadership of JPL and in collaboration with the NASA Goddard Space Flight Center. Four contractors have initiated the first study phase of the ACTDP effort to develop cryocooler designs capable of realistically completing flight unit development and delivery in the 2006 to 2007 timeframe. Each of the coolers is designed to provide greater than 7.5 mW of cooling at 6 K together with an additional 250 mW at 18 K. The four contractors include: Ball Aerospace & Technologies Corp. of Boulder Colorado Creare, Inc. of Hanover, New Hampshire Lockheed Martin ATC of Palo Alto, California TRW Space and Electronics of Redondo Beach, California The four concepts being pursued by these four contractors are summarized below. These concepts represent the starting point for the contractors studies and can thus be expected to evolve and be refined as the study phase progresses. It is planned that at least two of these concepts will be selected for fabrication of engineering model hardware starting in the fall of 2002. Ball ACTDP Cryocooler Concept. As shown in Fig. 5, Ball Aerospace's ACTDP cryocooler concept utilizes a multistage Stirling refrigerator to precool a J-T loop powered by a linear-motion Oxford-style compressor. The J-T loop provides remote cooling of the 6 K and 18 K loads and isolates the loads from compressor-generated vibration and EMI. No intermediate radiative precooling is required, and the compressor elements are easily separated by over 3 meters from the cryogenic loads. The multistage refrigerator is based on leveraging existing Ball flight-quality Stirling compressors, J-T cold-end technology, and drive electronics; these technologies are configured and adapted to meet the specific needs of the ACTDP mission requirements. The baseline concept has a projected total system mass of 27 kg (including flight drive electronics) and has an estimated input power of approximately 150 watts into the drive electronics with no intermediate radiative precooling. Creare ACTDP Cryocooler Concept. Creare's ACTDP cryocooler concept utilizes a multistage turbo-Bray ton refrigerator with optional precooling by a cryoradiator. The turbo-Brayton loop, which has remotely located turbo-expanders operating at 6 K and 18 K, generates minimal vibration and allows the 6 K and 18 K loads to be widely separated from the loop's room-temperature compressor and electronics. The multistage refrigerator is based on leveraging existing Creare flight-quality turbo-Brayton compressors, expanders, and drive electronics as well as new developmental hardware aimed at low temperature operation.24,25 These hardware elements are configured and adapted to meet the specific needs of the ACTDP mission requirements. The baseline
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concept, shown in Fig. 6, has a projected total system mass of 27 kg (including flight drive electronics) and has an estimated input power of approximately 105 watts into the drive electronics, with approximately 1.3 W dissipated at a 75 K intermediate-temperature radiator. Lockheed Martin ACTDP Cryocooler Concept. Lockheed Martin's ACTDP cryocooler concept utilizes a multistage pulse tube refrigerator, with optional cryoradiator precooling, to directly cool the 6 K and 18 K loads. The single-unit multistage refrigerator leverages existing Lockheed flight-quality pulse-tube compressors, cold heads, and drive electronics, and laboratory pulse tube technology26 that has demonstrated direct cooling down to 4 K; these are being configured and adapted to meet the specific needs of the ACTDP mission requirements. The baseline concept, shown in Fig. 7, has a projected total system mass of approximately 26 kg (including flight drive electronics) and has an estimated input power of approximately 208 watts into the drive electronics when no intermediate radiative precooling is utilized. Use of a 120 K precooler dissipating 8 W is estimated to reduce the input power to on the order of 106 watts. TRW ACTDP Cryocooler Concept. TRW's ACTDP cryocooler concept, illustrated in Fig. 8, utilizes a multistage pulse tube refrigerator, with optional cryoradiator precooling, to precool a J-T loop powered by a linear-motion Oxford-style compressor. The J-T loop provides remote cooling of the 6 K and 18 K loads and isolates the loads from any compressor-generated vibration
NASA SPACE CRYOCOOLER PROGRAM OVERVIEW
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and EMI. The multistage refrigerator is based on leveraging existing TRW flight-quality pulse tube compressors and drive electronics, and developmental J-T cold-end technology; these are configured and adapted to meet the specific needs of the ACTDP mission requirements. The baseline concept shown in the accompanying illustration has a projected total system mass of approximately 17 kg (including flight drive electronics) and has an estimated input power of approximately 207 watts into the drive electronics, with 2 W dissipated at the 85 K intermediate temperature radiator.
SUMMARY Cryocoolers are increasingly being adopted for usage in NASA science instruments, with a total of 10 cryocoolers launched into orbit over the past 10 years, and several more scheduled for the next few years. With flight cryocoolers widely available for the 30 K to 150 K temperature range, NASA-funded technology development is now focusing primarily on coolers in the 4-20 K temperature range, and on coolers for special applications such as storage of cryogenic propellants in space.
ACKNOWLEDGMENT The work described in this paper was carried out by NASA Centers and contractors including the Jet Propulsion Laboratory, California Institute of Technology and Goddard Space Flight Center; it was sponsored by the National Aeronautics and Space Administration.
REFERENCES 1 . Ross, R.G., Jr., “JPL Cryocooler Development and Test Program: A 10-year Overview,” Proceedings of the 1999 IEEE Aerospace Conference, Snowmass, Colorado, Cat. No. 99TH8403C, ISBN 0-78035427-3, IEEE, New York, 1999, pp. 115-124. 2. Boyle, R. and Ross, R.G., Jr., “Overview of NASA Space Cryocooler Programs,” Adv. in Cryogenic Engin., Vol 47B, Amer. Inst. of Physics, New York, 2002, pp. 1037-1044. 3. Boyle, R., Banks, S., Cleveland, P. and Turin, P., “ Design and Performance of the HESSI Cryostat,” Cryogenics 39 (12), 1999, pp. 969-973. 4. Boyle, R. et al., Cryocoolers for the HESSI Spectrometer: Final Report of the Cryocooler Tiger Team, Internal Document, Goddard Space Flight Center, 2001. 5. Boyle, R., Banks, S. and Shirey, K., “Final Qualification and Early On-Orbit Performance of the RHESSI Cryocooler,” Cryocoolers 12, Kluwer Academic/Plenum Publishers, New York, 2003. 6. Swift, W.L., Dolan, F.X. and Breedlove, J.J., “Initial Operation of the NICMOS Cryocooler on the Hubble Space Telescope,” Cryocoolers 12, Kluwer Academic/Plenum Publishers, New York, 2003.
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7. Swift, W.L., et al., “Flight Test Results for the NICMOS Cryocooler,” Adv. in Cryogenic Engineering, Vol 45A, Kluwer Academic/Plenum Publishers, NY, 2000, pp. 481-488. 8. Ross, R.G., Jr. and Green K., “AIRS Cryocooler System Design and Development,” Cryocoolers 9, Plenum Publishing Corp., New York, 1997, pp. 885-894. 9. Ross, R.G., Jr., et al., “AIRS PFM Pulse Tube Cooler System-level Performance,” Cryocoolers 10, Plenum Publishing Corp., New York, 1999, pp. 119-128. 10. Ross, R.G., Jr., “AIRS Pulse Tube Cooler System-Level and In-Space Performance Comparison,” Cryocoolers 12, Kluwer Academic/Plenum Publishers, New York, 2003. 11. Raab, J., et al., “TES FPC Flight Pulse Tube Cooler System,” Cryocoolers 11, Kluwer Academic/ Plenum Publishers, New York, 2001, pp. 131-138. 12. Collins, S.A., Rodriguez, J.I. and Ross, R.G., Jr., “TES Cryocooler System Design and Development,” Adv. in Cryogenic Engin., Vol 47B, Amer. Inst. of Physics, New York, 2002, pp. 1053-1060. 13. Kiehl, W.C., et al., “HIRDLS Instrument Flight Cryocooler Subsystem Integration and Acceptance Testing,” Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York, 2001, pp. 769-774. 14. Cornut, M. and Gavila, E., “Planck Spacecraft Cryogenic Chain,” Cryocoolers 12, Kluwer Academic/Plenum Publishers, New York, 2003. 15. Pearson, D., et al., “Test Performance of a Closed Cycle Continuous Hydrogen Sorption Cryocooler, ” Cryocoolers 12, Kluwer Academic/Plenum Publishers, New York, 2003. 16. Bowman, R.C., Jr., et al., “Evaluation of Hydride Compressor Elements for the Planck Sorption Cryocooler,” Cryocoolers 12, Kluwer Academic/Plenum Publishers, New York, 2003. 17. Hastings, L. et al., “An Overview of NASA Efforts on Zero Boil-off Storage of Cryogenic Propellants,” to be presented at 2001 Space Cryogenics Workshop, Milwaukee, WI. 18. Hedayat, A., et al.,“ Large Scale Demonstration of Liquid Hydrogen Storage with Zero Boiloff,” Adv. in Cryogenic Engin., Vol. 47B, Amer. Institute of Physics, Melville, NY, 2002, pp. 1276-1283. 19. Chang Dvaz, F. R., “Research Status of The Variable Specific Impulse Magnetoplasma Rocket,” Fusion Technology 35, 1999, pp. 87-93. 20. Chan, C.K., Ross, R.G., Jr., et al., “IMAS Pulse Tube Cooler Development and Testing,” Cryocoolers 10, Plenum Publishing Corp., New York, 1999, pp. 139-147. 21. Ross, R.G., Jr., “IMAS Pulse Tube Cryocooler Development and Testing,” Integrated Multispectral Atmospheric Sounder (IMAS) Instrument Technology Development and Demonstration, Final Report, Internal Document, Jet Propulsion Laboratory, Pasadena, CA, 1998, pp. 3-1 to 3-16. 22. Ross, R.G., Jr., et al., “Gamma-Ray Pulse Tube Cooler Development and Testing,” Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York, 2001, pp. 155-162. 23. Nast, T.C., et al., “Miniature Pulse Tube Cryocooler for Space Applications,” Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York, 2001, pp. 145-154. 24. Swift, W., McCormick, J.A. and Zagarola, M.V., “A Low Temperature Turbo-Brayton Cryocooler for Space Applications,” Adv. in Cryogenic Engin., Vol. 47B, Amer. Inst. of Physics, Melville, NY, 2002, pp. 1061-1068. 25. Zagarola, M. V., et al., “Development of a Turbo-Brayton Cooler for 6 K Space Applications,” Cryocoolers 12, Kluwer Academic/Plenum Publishers, New York, 2003. 26. Olson, J., Nast, T. and Evtimov, B., “Development of a 10K Pulse Tube Cryocooler for Space Applications,” Cryocoolers 12, Kluwer Academic/Plenum Publishers, New York, 2003.
Air Force Research Laboratory Space Cryogenic Technology Research Initiatives B.J. Tomlinson, B. Flake and T. Roberts Space Vehicles Directorate Air Force Research Laboratory Kirtland AFB, NM, USA 87117-5776
ABSTRACT The Air Force Research Laboratory (AFRL) Space Vehicles Directorate is actively pursuing cryocooler and cryogenic integration research to support technology needs for the Air Force, the Missile Defense Agency (MDA, formerly the Ballistic Missile Defense Organization), and the Department of Defense (DoD). This paper presents an overview of the in-house and external applied research efforts on a number of fundamental issues toward the development of long-life strategic cryocoolers and cryogenic integration technology. External activities include Small Business Innovative Research (SBIR) development, AFRL funded activities at the National Institute for Standards and Technology (NIST), and funded research at Oxford University. In house research efforts include work on advanced regenerators for Stirling cycle cryocoolers, and continued in-house research on products from the SBIR program. Additionally, this paper examines future research trends for AF and DoD cryogenic technology.
INTRODUCTION The Air Force Research Laboratory, Space Cryogenic Technologies Group (AFRL/VSSS) is developing next generation space cryogenic cooling and cryogenic integration technologies for space based applications. The principal application is for infrared sensor cooling to provide the warfighter with surveillance and tracking capability from space borne platforms. Space cryocooler technology development at the Air Force Research Laboratory (AFRL) centers on the development of long life, high efficiency, and low-mass active refrigeration for space. Technologies addressed include Stirling cycle (includes pulse tubes), reverse Brayton cycle, and Joule-Thomson cycle coolers. The primary development focus has been on achieving technology capable of long lifetime operation in excess of 7 years and, more recently, 10+ years with a high degree of reliability and confidence within heritage designs to space flight hardware. The technical efforts at AFRL concentrate on exploratory and advanced development programs that focus on the development of technology from concept and breadboard engineering models to protoflight models that are geared to experimental characterization and technology transition for flight demonstrations and, potentially, operational programs. Given the incremental success of the AFRL program in long life design, development focus has renewed emphasis on issues such as efficiency, mass, and induced vibration to ensure the feasibility of this technology for space use (Davis, et al.1).
Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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Cryogenic integration technology has been often a neglected system design activity. With the marked rise in flight demonstrations and operational programs using cryocooler technology, more attention has been placed on cryogenic integration technology. If integration is neglected, cryocooler cooling capacity can be easily overwhelmed by the heat parasitics of a poorly designed or inadequate cryogenic interface system. AFRL development focus has centered on the development of component level technology such as cryogenic thermal interfaces, cryogenic thermal switches (for redundancy), cryogenic heat transport (for heat transport over long distances), and cryogenic thermal storage (temperature stability and duty cycle operation). Current programs are focusing on near protoflight designs for miniature thermal switches and miniature cryogenic loop heat pipes that could act as thermal switches and heat transport devices (i.e. like a conductive strap).2 The AFRL Cryogenic Cooling Research Facility provides a laboratory examination of first of-a-kind and / or one-of-a-kind cryocooler and cryogenic integration technologies. The laboratory characterizes the full thermodynamic performance envelope of the technology. The characterization examination of unique cryocoolers is a critical and necessary component for emerging cryocooler technologies. Although a function of advanced, or protoflight development, the laboratory provides the framework for the in-house research capability and as a practical test bed for technologies that are near transition to flight status. The endurance evaluations of coolers in the lab provides the Air Force, industry, and system designers with detailed data on the suitability and reliability of developmental cryogenic technology for long life space applications. The significant interest in the endurance evaluation of cryocoolers (and the large database of performance data), is stark evidence of the concern of the user and system designer community over the capabilities of cooler technology. AFRL is unique in that the characterization and endurance evaluation of advanced cooler technology is synergistically intertwined with its developing research program in looking at basic thermodynamic, fluid flow, and heat transfer issues in cryogenic cooling. The AFRL research and development program, broken up into “basic” (Air Force 6.1 research) and “applied” (Air Force 6.2 research), is essential to the continued development of advanced cryocooler and cryogenic integration technology. With constrained research investment budgets, planning and leveraging are the keys to successful development and transition of applied and basic research. Balancing complementary in-house and external research efforts is integral in insuring efficiency of the overall program and is necessary to both produce results in the quickest possible time and eliminate nonproductive investments (see Suddarth3 for a discussion of this on a service wide scale).
ADVANCED DEVELOPMENT AND TRANSITION The primary emphasis of the current advanced cryocooler and cryogenic development efforts is to support the requirements for advanced DoD satellite programs that will fly within the next several years. The advanced development programs focus on the development of cryocoolers and integration technology that is primarily centered on multiple cooling loads at roughly 35 and 100 Kelvin. These requirements are not the focus of this paper, but serve as the baseline for the research and development planning and execution that AFRL is conducting to provide the basic evolutionary and revolutionary capabilities in future systems. Often times, the ideal development program of basic to applied to advanced to protoflight technology cannot be achieved. Advanced development programs, meant to be far more mature than “unproven” basic or applied research technology, serve as the test beds for the products of basic and applied research. Although this increases technology development risk, the payoff to risk ratio is large and could have significant impact on the capabilities of the warfighter. With respect to the developments for current and near term needs, capabilities that must be met justify program risk to insert R&D products into the advanced research programs. Although the 6.3, advanced programs focus on meeting the “now” requirements of the user, it is important to note that cryocooler technology is considered both necessary to complete the mission and immature. Consequently, the user community continues to provide support to cryocooler development programs and supports the characterization and long life endurance of these products to provide some degree of reliability confidence. This is a reaffirmation of the importance of our cryocooler characterization laboratory to look at performance issues that the con-
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tractor does not and to maintain the long life endurance evaluation that would be cost prohibitive for a contractor. Continued discussion of the transition of research and development products to advanced and protoflight development will continue below in sections concerning basic R&D, the role of small business, and synthesis and balance.
BASIC AND APPLIED RESEARCH AND DEVELOPMENT INITIATIVES Air Force Science & Technology basic and applied research has focused on the thermodynamics, fluid flow, and heat transfer characteristics of regenerators for Stirling cycle cryogenic cooling and fundamental issues with induced vibration in linear motor compressors. Future research initiatives will continue the work on these issues and initiate or expand research on neglected areas and new opportunities. Objectives of in-house research include advancing the knowledge of the fundamental heat, mass and momentum transport processes of cryogenic refrigeration through experimentation, modeling and simulation. With experimental data and results from highly detailed computational flow models, simplified equations and other models will be developed to predict cooler performance. The principle basis in practice for cryocooler design is the experience of the designers. As a relatively new technology, the level and breadth of experience is relatively small compared to, say, commercial refrigeration systems. Consequently, innovation has been relatively slow with substantial leaps in performance made almost solely by trial and error—an expensive and time consuming process. Although the principle industry developers have access to some commercial models and their own internal models, there are significant shortcomings in their modeling process that hinder full understanding of the internal processes and allow true optimal designs. To accelerate and innovate cryocooler designs, designers need tools to predict performance over a full range of geometries and operating conditions. A substantial body of knowledge exists at AFRL on the characterization and testing of protoflight cryocoolers. While some operating parameters can be varied to find optimum operation (e.g. frequency, stroke-length, phase-shift), actual coolers’ physical geometry and materials are fixed. As cryocooler program development costs run in the millions of dollars, it is a prohibitively expensive proposition to optimize cryocooler designs solely through experimental prototype variations. Simplified models that accurately predict thermal and flow performance will allow the investigation of optimal cryocooler designs that maximize cooling performance while minimizing system mass and power requirements. Achieving low temperatures with light weight, energy efficient cryocoolers is critical to fielding space based long wave infrared sensors. Preliminary work has begun on computational flow dynamics (CFD) simulation of the flows internal to regenerators and pulse tubes. Data from these studies will be used along with experimental data to develop predictive models for designers. As such small scale calculations cannot be empirically validated, current experimental work being conducted is seeking to directly measure the macro scale internal flows within active Stirling cycle components, and then to correlate this measured transient performance states with both CFD and macro scale models of cooler component performance (see Roberts and Desai4 in this proceedings for a discussion of this effort’s preliminary results). Vibration mitigation work, conducted with Oxford University, focuses on understanding the basic phenomena that produce nonlinear, off axis vibration in dual opposed linear motor compressors. By understanding the quasi-static and dynamic forces of motion in these gas filled compressors, new designs can be created to mitigate the vibration produced and allow for significant reductions in complexity for the control electronics. This research has great potential for influencing the basic design of Oxford class linear compressors currently in use by a number of industry developers for Stirling cryocoolers. Transition success is achieved when AFRL sponsored research is incorporated into next generation designs by industry and the benefits envisioned by the research are achieved. AFRL is working to ensure this transition through a number of channels, but principally through the dissemination of the research findings through conference papers and journal articles. Recent work with the National Institute of Standards and Technology (NIST) has explored
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potential designs for highly compact and effective recuperative heat exchangers for applications such as Joule-Thomson or reverse Brayton cycle coolers, where the heat exchangers often either determine or greatly contribute to the overall cooler efficiency and mass. Products of this research have been transitioned to other cryocooler programs and will be demonstrated by early 2003. Also, solid state laser cryogenic cooling is being investigated via SBIR programs and collaboration with Los Alamos National Labs and the University of New Mexico (e.g. Edwards, B.C. et al.5). Program leveraging with other organizations such as Sandia National Labs, the National Institute of Standards and Technology (NIST), and various academic institutions are being explored for future cooperation and development. Future research initiatives center around the basic thermodynamic processes within the various cryogenic cooling cycles as they are supportive of both special program office (SPO) mission requirements and general DoD and USAF research goals. The goals cited for DoD research sponsors are oriented along both basic research and applications research lines. For example, the current work in measuring and modeling regenerator flow and transport mechanisms is being re-scoped for AFOSR to show how this flow regime is an example of soliton propagation through a moderately viscous flow field. Growth areas for applications include investigation of Microelectromechanical System (MEMS) fabrication techniques for cryogenic cooling components such as regenerators, heat exchangers, expanders, and compressors. The sum of the research plan for space cryogenic technology is still being formulated but is grounded by the tenets of continued in-house regenerator research and leveraging of the Small Business Innovative Research program. The current strawman plan (seen graphically in Figure 1) for space cryogenic technology continues the building of the basic and applied research and development program in conjunction with the SBIR program to transition products to the advanced hardware programs. These research products will translate to performance enhancing, evolutionary, or revolutionary leaps in system capability. A very important component of this process is the feedback loop from the advanced programs and missions to continually update the aims of the research program to ensure AFRL is addressing stated and potential future mission needs. Although flexibility is a major strength, care must be taken to ensure a stable research program over the long run. Such a program must be able to endure the sometimes short sighted directional changes in the research program that prevent long term goals from coming to fruition.
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Careful coordination between the Air Force research effort (longer term view) and the warfighter (shorter term view) is necessary to ensure a responsive and successful research program. AFRL will continue to publish technical papers relating to research aims and research progress in order to provide the development community background for its own IR&D technology development programs. Most importantly, AFRL will strive to continue to transition research products from first-of-a-kind developments to protoflight level in order to widen and strengthen the technology base for systems designers and the warfighter. AFRL will also maintain the flexibility needed to address evolving mission requirements and opportunities that arise as a result of rapidly evolving technology in other arenas.
THE ROLE OF SMALL BUSINESS INNOVATIVE RESEARCH PROGRAMS The Small Business Innovative Research (SBIR) program has served many roles in the development of cryocooler and cryogenic integration technology. The program is basically a way for small businesses to develop and transition innovative ideas to commercially viable products with the government’s help. Our use of the SBIR program has taken two routes. One, the program serves as a path for “risky” ideas and research, and second, as a tool to develop components or concepts that have a clear path for insertion into advanced development programs for current and near term demonstration. The SBIR funding, from our point of view, is outside of our budget and is “free” money. This gives us a great deal of leveraging for manpower versus SBIR investment and is very attractive. SBIR developments for “risky” ideas provide us a tool that can investigate potential revolutionary technology with relatively low risk to our disposable development budget. Although highly risky, this is the type of R&D that is needed to make the critical leaps and break out of current technology paradigms and limitations. For example, we manage MDA sponsored SBIR programs investigating solid state cooling with lasers. Sustained and prudent selection of SBIRs over time will allow AFRL engineers to explore a large cross section of possible technology with acceptable risk to the government and an acceptable potential rate of success. Anti-Stokes laser cooling is now being considered for submission as an AFOSR Multidisciplinary Research Initiative. Many “new” opportunities in research have been addressed with the SBIR program by creating SBIR topics that focus on technology development for low temperature cooling, manufacturability, reliability, and Microelectromechanical System (MEMS) fabrication techniques for cryogenic cooling. By creating the topics for cutting edge research themes, the SBIR program takes on the guise of a speculative research initiative that is designed to augment and complement Air Force funded in-house and external research. SBIR developments for near term component development and transition come from the technology developments that have “plug and play” characteristics, enhance performance of current technology, or provide a niche solution to a cryogenic issue. These developments are only as good as the efforts from the SBIR contractor and AFRL to transition the technology to the more mature cryocooler and cryogenic integration technology programs, industry base, and the flight programs where they are needed. With respect to the in-house and externally funded basic and applied research and development programs, the SBIR program can be utilized as a stand-alone tool or as a complementary tool. Prudence would dictate that selection of SBIR programs would consider the transition of development products to in-house programs for continued exploration and characterization as either part of existing research or as a new branch of research. This has often been overlooked in the past, but is now a prime consideration in the structure of the SBIR programs that are funded and managed by AFRL.
SYNTHESIS, BALANCE, AND PLANNING In order to have a successful R&D program, the work to achieve stated future performance goals for cryogenic cooling needs to be coordinated, balanced within the development program, and has the benefit of detailed investment and transition planning. All of these characteristics are
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less scientific and more technical program management in nature, but understanding of the technology and technology direction is needed to make prudent decisions. Technology forecasting can only leverage from what is currently understood about existing technology. If you could predict the wildcards, you would not be doing R&D. A more conservative approach would be to look at evolving system requirements and incorporate that into a capability based technology goal for cryogenic systems. Improved discrimination utilizing high performance multicolor and multispectral focal planes will provide significant improvements in operational capability for surveillance and missile tracking and detection. Multiple spacecraft applications could require near 10 K operation with the use of doped Silicon infrared sensors for missions such as mid-course missile detection, satellite defensive systems, and spectroscopy surveillance where silicon is preferred for wavelength and/or uniformity. Traditionally, stored cryogens have been used where low temperature operation is required, but incur large system penalties and prohibitive mass penalties for most missions on the drawing board. This is due to the fact that dewars are mostly applicable for short duration (< 1 year) experiments or very small cooling loads. The end result is that efficient, low mass, active cryocoolers are needed to support the low temperature cooling requirements for Very Long Wave Infrared (VLWIR) focal planes. For the future, AFRL is examining technologies that will be capable of supporting out year system concepts and requirements. As shown in Figure 2, AFRL is pursuing technologies in 10 Kelvin cooling, high capacity 35 Kelvin cooling, and advanced cryogenic integration. Additionally, multistage cooling is an area of system level interest to maximize system cooler system efficiency for power consumption and mass. These trends are based on near term cooling requirements for available doped Silicon focal plane arrays. In all likelihood, future systems will require multiple or larger arrays with a resulting increase in heat loads. In addition to cooling the tracking sensor, VLWIR surveillance systems will cool off the aft optics at temperatures projected from 40 to 60K. Future systems will require cryocoolers with much larger cooling capacity to meet the expected increase in focal plane and optical system size. Projected cooling loads for a conceptual system are as high as 5 watts at 35 Kelvin for the sensor and >20 watts at 100 Kelvin for the optics. Large capacity, multi-load cryocoolers capable of simultaneously cooling sensors, optics, and optical benches will greatly simplify the cooling approach for next generation payloads. In addition, cryocoolers are potentially an enabling technology for future space systems that will have significant cooling requirements for the large cryogenic fuel or propellant tanks on board the spacecraft. A range of issues are currently being addressed including the use of cryogenic gas or liquid storage, the large volume requirements for mission life, and the cost for on-orbit replenish-
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ment. Cryocooler capacity and thermal / structural integration for large tanks will be significant issues with cooling loads like 50-100 watts at 100 Kelvin or 5 watts at 20 Kelvin. High capacity cryocoolers and long term (>20 years) on-orbit propellant storage are potentially enabling technology for many future space systems including orbital transfer vehicles and on-orbit propellant depots. Given the range of potential system designs, cryogenic cooling and integration are important issues, and planning is required to ensure positive progress toward developing “generic” technologies that support the envelope of performance requirements. Cooling requirements for each of these varied applications are compounded by the need to minimize mass, volume, input power, and provide for redundancy. Cryocoolers capable of multistage cooling (multiple loads at different temperatures) are much more system effective than having multiple cryocoolers for each specific cooling load. Multistage coolers increase system reliability by minimizing the number of redundant coolers needed and, consequently, minimize the number of drive electronics for those coolers. One of the largest hurdles to effectively designing multistage coolers is the lack of understanding of the thermodynamic processes within the cooler. In any event, the benefits of having one cooler to satisfy all optics and sensor cooling for a given application is far more attractive to systems designers than multiple coolers, each with their own peculiar control requirements and failure or degradation modes. It is important to note that with constrained budgets not all possible areas of research can be adequately explored. Given this, the Space Cryogenic Technology research program must weigh potential payoffs, insertion potential, and maturity for existing and new research efforts with appropriate funding levels. This is not an easy task and it will have significant impact on the out-year development of cryocooler technology. Given the varied requirements for cryogenic cooling systems, it can be seen that an integrated research and development plan is needed to efficiently invest technology funds for maximum return. The AFRL Space Cryogenic Technology group has initiated an internal effort to map out research investment funds, research initiatives, and transition opportunities based on current technology forecasts. The principle end goals of the research cover three separate areas: low temperature cooling, high capacity cooling, and multistage cooling. Successful research within these areas would provide the technology needed to meet the aforementioned out-year system requirements. However, the difficult work is not just in the research itself, but the careful construction of the overall research program over many years. Fiscal years 2003 and 2004 will be critical in the establishment of the AFRL research program and will provide the foundation of Air Force cryogenic technology development for the next ten years.
CONCLUSIONS Overall, the need for a balanced and integrated plan for technology development is evident to ensure the efficient use of resources to accomplish development goals. AFRL is designing a research and development program to take advantage of in-house expertise, external development expertise, the SBIR program, current advanced development programs, and planned out-year budgetary resources to come up with an integrated roadmap for research and transition. Current user and system designer emphasis on low temperature cooling, high capacity cooling, and multistage cooling have led to a number of basic and applied research initiatives and SBIR topics. The research portfolio is not complete. Additional inputs and analysis are needed from the warfighter, systems designers, cryocooler developers, academia, and government resources to round out the strategy of research and development for space cryogenic technology. Critical planning and initiation of research programs will conclude in the 2003 and 2004 timefrarne and will lay out a roadmap for technology development and insertion opportunities for the next 10 years. The AFRL’s leveraged approach to cryocooler development is providing technology to support near term requirements for DoD programs. Pursuit of advanced concepts can significantly reduce weight and improve thermodynamic performance to result in enhanced cryogenic system performance, integrated sensor/cooler packaging, improved reliability, and mission enabling capability for future DoD space systems.
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ACKNOWLEDGMENTS The AFRL developmental programs and research and development work described in this paper are a combined effort of Air Force civilian and military personnel, personnel from the Department of Defense, Dynacs Engineering support contractors, Aerospace Corporation engineers, and external organizations such as NIST, Oxford University, and a host of hard working small business contractors.
REFERENCES 1. Davis, T. M., Reilly, J., and Tomlinson, B. J., “Air Force Research Laboratory Cryocooler Technology Development,” Cryocoolers 10, R. G. Ross, Jr., Ed., Plenum Press, New York (1999), pp. 21-32. 2. Bugby, D., Stouffer, C., Davis, T., Ton-dinson, B. J., Rich, M., Ku, J., Swanson, T., and Glaister, D., “Development of Advanced Cryogenic Integration Solutions,” Cryocoolers 10, R.G. Ross, Jr., Ed., Plenum Press, New York (1999), pp. 671-687. 3. Suddarth, S., “Solving the Great Air Force Systems Irony,” Aerospace Power Journal, XVI/l, Eric A. Ash ed., US Air Force, Maxwell AFB (Spring 2002); also available at http://www.airpower.maxwell. af.mil/airchronicles/apj/apj02/sprO2/suddarth.html 4. Roberts, T. and Desai, P, “Periodic porous media flows in regenerators,” Cryocoolers 12, Kluwer Academic/Plenum Publishers, New York (2003). 5. Edwards, B.C. et al., “Demonstration of a Solid State Optical Cooler: An Approach to Cryogenic Refrigeration,” Journal of Applied Physics, 86 (1999), pp. 6489-6493.
Status Report on the Linear Drive Coolers for the Department of Defense Standard Advanced Dewar Assembly (SADA) W.E. Salazar U. S. Army Communications and Electronics Command, Research, Development, and Engineering Center Night Vision and Electronic Sensors Directorate Fort Belvoir, VA 22060-5806
ABSTRACT The Standard Advanced Dewar Assembly (SADA) is the critical module in the Department of Defense (DoD) standardization effort of scanning second-generation thermal imaging systems. DOD has established a family of SADA’s to address high performance (SADA I), mid-to-high performance (SADA II), and compact class (SADA III) systems. SADA’s consist of the Infrared Focal Plane Array (IRFPA), Dewar, Command & Control Electronics (C&CE), and the cryogenic cooler. SADA’s are used in weapons systems such as Comanche and Apache helicopters, the M1 Abrams Tank, the M2 Bradley Fighting Vehicle, the Line of Sight Antitank (LOSAT) system, the Improved Target Acquisition System (ITAS), and Javelin’s Command Launch Unit (CLU). DOD has defined a family of tactical linear drive coolers in support of the family of SADA’s. The Stirling linear drive cryo-coolers are utilized to cool the Infrared Focal Plane Arrays (IRFPAs) in the SADA’s. These coolers are required to have low input power, a quick cooldown time, low vibration output, low audible noise, and higher reliability than currently fielded rotary coolers. This paper: (1) outlines the characteristics of each cooler, (2) presents the status and results of qualification tests, (3) discusses issues that have impacted cooler reliability, and (4) presents the status and test results of efforts to increase linear drive cooler reliability.
INTRODUCTION The US Department of Defense (DoD) has chartered a strategy to standardize Second Generation Infrared (IR) components throughout the Services. The SADA’s are at the heart of this standardization effort. A SADA consists of the Infrared Focal Plane Array, Dewar, Command & Control Electronics (C&CE), and an integrated linear drive cryo-cooler that may be procured separately from the SADA. The US Army CECOM Night Vision and Electronics Sensors Directorate (NVESD) has developed a family of Stirling cycle linear drive coolers to support the SADA’s and the standardization effort.
Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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The SADA’s are defined by Government controlled performance specifications and interface control drawings. These SADA specifications and drawings in turn define the performance and interface requirements for the family of cryocoolers. The SADA specifications and drawings ensure the form, fit, and performance of SADA/coolers built by different manufacturers from the United States and abroad. The SADA coolers are dual opposed piston, Stirling cycle, linear drive coolers that require either external or internal cooler control electronics. These coolers were developed to address the shortcomings of rotary coolers such as low reliability, poor shelf life, multi-axes vibration and torque, excessive acoustic noise, and poor temperature stability. The integrated SADA/cooler assembly improved the thermal interface between the cooler and the IR focal plane array (IRFPA) by directly bonding the IRFPA to the cold-finger. The common module Dewar/Cooler requires a “fuzz-button” or bellows to make contact between the cooler coldfinger and the Detector. Table 1 highlights the key parameters of the family of coolers:
QUALIFICATION REQUIREMENTS FOR SADA CRYO-COOLERS SADA’s and Linear Drive Coolers are infrared imaging critical components that require qualification prior to first production delivery. These components are qualified once they pass a series of tests designed to reflect their use in the required military environment. The environmental requirements can be generally divided into airborne, ground, and man-portable type systems. The government or the manufacturer may perform the qualification tests. The government approves all test procedures, equipment, and test facilities prior to testing. This is required in order to ensure standardization of test methods, and to allow for the true comparison of test results from different manufacturers from around the world. Some of the weapon systems supported by this qualification effort include the Army’s Second Generation Forward Looking Infrared (FLIR) Horizontal Technology Integration (HTI), Comanche, Apache, Javelin, and ITAS. The tests listed in Table 2 are required for the cooler qualification effort. The family of Army cryo-coolers is shown in Figure 1.
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0.15-WATT LINEAR DRIVE COOLERS The 0.15-watt linear drive cooler was developed for use in the Javelin Command Launch Unit (CLU). Javelin is an anti-tank missile system. This 0.15-watt cooler was originally qualified in 1997, and it was re-qualified in 1999 following an Army funded Manufacturing Technology (Mantech) program with DRS (formerly Texas Instruments). This Army program covered both the 0.15-watt and 1.0-watt coolers. It successfully applied manufacturing process improvements to the compressor clearance seals, gas decontamination process, motor manufacturing, and cooler final assembly. The Army Mantech program also developed manufacturing processes to replace the compressor helical spring suspension system with a flat plate, flexure spring, suspension system. The goal of changing to a flexure springs system was to simplify motor assembly and double the life of the coolers from 4,000 to 8,000 hours MTTF. Three Javelin flexure springs coolers are still under reliability testing. They have accumulated an average of over 14,000 hours. Hundreds of 0.15-watt Javelin flexure springs coolers have been delivered to the US Army since 1999. This cooler continues to be mass-produced to meet demand. Before the successful mass-production of the Javelin flexure springs cooler, flexure springs cryo-coolers had only been associated with aerospace applications, very low manufacturing rates, and very high prices.
1.0-WATT LINEAR DRIVE COOLERS The 1.0-watt cooler was the focus of significant efforts and investments to qualify multiple sources, reduce manufacturing costs, and increase their reliability. These coolers are used with SADA II, and are critical components of many DoD programs to include the Army’s 2nd Generation FLIR Horizontal Technology Integration (2nd Gen. FLIR HTI) program and ITAS. Three sources are currently qualified. The DRS Infrared Technologies 1.0-watt cooler design was first qualified in 1997. DRS is one of the main suppliers of 1.0-watt coolers to the Army. AEG Infrared Modules (AIM) of Germany was the second qualified supplier. AIM was qualified in 1998 through a Foreign Comparative Testing (FCT) program with NVESD and the Army’s Program Manager for FLIRs (PM FLIR). The FCT program provided funds to purchase and test several AIM coolers. Northrop Grumman Life Support (formerly Litton) of Davenport Iowa, was recently qualified, and provides an additional source of supply to the expanding market for tactical cryo-coolers. The 1.0-watt cooler was also the beneficiary of an Army Mantech effort that resulted in a decrease in cooler manufacturing costs. As mentioned before, this Mantech program was performed on both the 0.15-watt and 1.0-watt coolers with funding from the US Army Mantech program, the Program Manager for Night Vision Reconnaissance, Surveillance and Target Acquisition (PM-NV/RSTA), and the Program Manager for Javelin. This Mantech program focused on manufacturing process improvements to the compressor clearance seals, gas decontamination process, regenerator/expander design, motor manufacturing, and cooler final assembly. The Mantech effort was completed in 1998 with the completion of environmental and reliability tests. This program established a lower cooler price threshold that is impacting the competitive procurement of current and future procurements. In order to maximize competition, the Mantech program included a technology transfer effort that provided DRS reports and briefings to approved cooler manufacturers. Figure 2 shows the cooling capacity of several qualified coolers. The figure clearly groups cooler performance at the three required ambient temperatures of –54°C, +23°C and +71°C. 1.75-WATT AND 1.50-WATT LINEAR DRIVE COOLERS The SADA I, designed to address the requirements for high-performance systems, is entering the competitive stage of development and procurements. The 1.75-watt and 1.50-watt coolers were designed to address the needs of SADA I applications. The Apache and Comanche helicopters are the main users of the SADA I/cooler assemblies.
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The 1.75-watt underwent evaluation and formal qualification testing as part of a Foreign Comparative Test program with NVESD and the Program Manager for Comanche (PM Comanche). This cooler demonstrated acceptable performance throughout the testing program. Several 1.75watt coolers were successfully integrated into Comanche, Apache, and several other high performance FLIR systems. The formal qualification of the 1.75-watt cooler was not completed as the qualification effort for a SADA I cooler was redirected towards the newly designed 1.50-watt cooler. AIM introduced the 1.50-watt cooler for use in the Apache program. Due to its lighter weight, the Comanche program adopted this cooler for its FLIR systems. The AIM 1.50-watt cooler design was derived from a modified 1.0-watt cooler compressor. It is under formal environmental testing in Germany. The reliability test will take place in the United States in a US Army facility. Initial testing has demonstrated that this new cooler meets basic performance requirements at the required ambient temperatures. The graph in Figure 3 shows the critical cool-down time performance measurements for both the 1.75-watt and 1.50-watt coolers. It also shows performance in the required hot, room, and cold ambient temperatures.
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ISSUES THAT HAVE IMPACTED RELIABILITY OF SADA II COOLERS Since 1995, the US Army has sponsored three sets of reliability tests involving the SADA II 1.0-watt linear drive cooler. The reliability test cycle in Figure 4 was standard in all the reliability tests. Several failures not related to wear-out have occurred during these tests. Several of these nonwear failures were found during periodic performance testing. The reliability test is stopped every 400 to 500 hours to allow for periodic performance testing. The following Army failure criteria are applied during the reliability test and during each periodic performance test: Inability to cool the required heat-load to 79.5K or below, when measured during the 72 hour reliability cycle (-32C to +52C) Failure to cool-down to 79.5K in 17 minutes or less Failure to meet the input power requirement of 60 watts when measured at any point during the 72 hour reliability cycle Failure to meet the leak rate requirement Failure to meet the vibration output requirements If the cooler under test fails to meet any of the above criterions, the cooler is taken out of the reliability test. The following is a discussion of the lessons learned and the paths taken to address issues that impacted cooler reliability.
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Cooler Electronic Issues The cooler control electronics make a significant impact on cooler performance and its expected life. The SADA II 1.0-watt cryo-cooler is designed to work with external cooler control electronics. In a 1996 test, the cooler control electronic board was replaced with a more efficient board. The change gave the cooler under test more input power margin that resulted in additional useful life for the cooler. Figure 5 shows how the input power levels decreased after the board was changed.
Dewar/Cooler Integration Issues Dewar and cooler manufacturers sometimes overlook the importance of this critical manufacturing process. Several coolers have failed due to loss of gas. This problem has occurred during controlled reliability tests or during field tests. All cooler failures occurred after the Dewar/cooler assemblies had passed all acceptance and environmental stress screening tests. The loss of gas was manifested as “Knocking” or noisy coolers, or the coolers suffered a sharp decrease in cooling capacity. The reasons for the loss of gas were the following: Loose Dewar/Cooler interface screws Dirty or scratched Dewar/Cooler mating surfaces The Dewar/cooler integrating facilities took successful steps to remedy these issues. These steps involved a top to bottom review of their integration operations. This resulted in the revision of integration and handling procedures, a change in inspection requirements, moving to new facilities, and the tightening of fabrication requirements for this critical interface.
RELIABILITY IMPROVEMENTS The low reliability of tactical cryo-coolers is one of the top lifecycle cost drivers in military infrared imaging systems. In the mid-1980’s, the Mean Time to Failure (MTTF) of military rotary cryo-coolers ranged from 1,000 to 2,500 hours. In the 1990’s, the introduction of linear drive coolers with control electronics doubled the MTTF requirements to 4,000 hours. The reliability of these coolers has been demonstrated with the completion of several reliability tests for both the 0.15-watt and 1.0-watt tactical coolers. Most of the reliability failures during the first tests were attributed to wear-out of the compressor pistons. The Army Mantech program demonstrated that the use of harder materials in the piston and displacer clearance seals results in increased reliability. The program developed and tested several cost effective manufacturing processes to fabricate different clearance seal materials. Since then, several manufacturers have introduced new wear-resistant materials in their linear drive com-
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pressors and expanders. In addition, more effective decontamination processes were developed and tested, and several companies are now using these processes in full-scale production. These steps have resulted in production coolers with predicted reliability in the 5,000 to 6,000 hour MTTF range. The most dramatic improvement in Army cryo-cooler reliability since the mid-1990’s has come with the introduction of flexure springs. The radial stiffness provided by the flexure springs significantly lowers the wear-out of the piston clearance seals. Flexure springs had for many years been used in long life cryo-coolers for aerospace applications, but this technology was not used in low-cost tactical coolers due to the high price associated with the piece parts and assembly processes. The introduction of this technology in tactical cryo-coolers was made possible by the development low-cost packaging and assembly processes. Figure 6 provides a snapshot of the official reliability test results for all SADA II 1.0-watt coolers tested since 1996. This graph clearly depicts the significant difference in reliability between linear drive coolers with helical springs and with flexure springs. The resulting mean MTTF of coolers with helical springs is at 5,792 hours. For an 80% confidence level the lower confidence limit is at 5,374 MTTF hours. Some of these failures were not related to compressor wear-out, so if all other failures are excluded from the calculations the MTTF is slightly higher at 6,243 hours. The mean MTTF for the flexure springs coolers is now at 12,888 hours. For an 80% confidence level the lower confidence limit is at 11,783 hours. Two of the four flexure spring coolers under test have failed due to wear in the expander. Therefore, further increases in reliability must come from improvements in the expander area.
SUMMARY The family of Standard Advanced Dewar Assemblies (SADA’s) is available to support a DoD standardization strategy. A family of linear drive coolers is established in support of the SADA assemblies and the standardization strategy. Weapon systems such as the Army’s Second Generation Forward Looking Infrared (FLIR) Horizontal Technology Integration (HTI), Comanche, Apache, Javelin, and ITAS have required or are currently requiring successful qualification tests as a prerequisite to production programs. The US Army CECOM Night Vision Directorate and US Army Program Managers were successful in efforts aimed at qualifying coolers and increasing cooler reliability. Qualification test efforts for the DRS 0.15-watt flexure spring and for three sources of 1.0-watt coolers were successfully completed in the last five years. Reliability tests show that flexure spring coolers provide a significant increase in cryo-cooler reliability. Several thousand 0.15-watt and 1.0-watt coolers have
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been produced and delivered to Army programs. The 1.50-watt cooler has replaced the 1.75-watt cooler in helicopter applications. The qualification effort for AIM’s 1.50-watt cooler is ongoing, and efforts to qualify the 1.0-watt flexure spring coolers will continue.
REFERENCES 1.
J. Shaffer and H. Dunmire. The DOD Family of Linear Drive Coolers for Weapons Systems,” Cryocoolers 9, Plenum Publishing Corp., New York, 1997, pp. 17-24.
2.
February 1999 Industry Review, Contract DAAB07-95-C-J513, Linear Drive Cooler Mantech Program, DRS Infrared Technologies.
3.
December 1997 Industry Review, Contract DAAB07-95-C-J513, Linear Drive Cooler Mantech Program, Raytheon Texas Instruments Systems.
4.
R.M. Rawlings and S. Miskimins, “Flexure Springs Applied to Low-Cost Linear Drive Cryocoolers,” Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York, 2001, pp. 103-110.
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Development of a Low-Power Stirling Cycle Cryocooler for Space Applications J.S. Reed and G.D. Peskett University of Oxford, Clarendon Laboratory Oxford, OX13PU, UK
ABSTRACT Stirling cycle cryocoolers have successfully been used to provide thermal control for many space applications, particularly Earth observation instruments employing infrared telescopes. These coolers typically provide cooling powers of 0.5-LOW at 80K. Recent developments have reduced heat lift requirements with the result that these coolers may be larger and more massive than required for some applications. Derating can also lead to a drop in efficiency. A small Stirling cycle cryocooler optimized for a l00mW heat lift is under development at Oxford University. An integral geometry has been chosen, and the cooler is to be mounted on a radiator. A prototype has been manufactured and assembled. Performance will be investigated over a range of operating conditions, and results will be integrated into a simple computer model of the cooler. Some preliminary results are presented.
MINIATURE STIRLING PROTOTYPE The prototype is an integral Stirling machine based around a linear fixed coil, moving magnet compressor motor (Figures 1 and 2). Helium is used as the working gas; other key parameters are shown in Table 1. The displacer/regenerator is driven by the pressure wave and so the system must be carefully set-up to provide the correct phase shifts. Both the displacer/regenerator and the compressor piston are supported by spiral type flexure bearings — the displacer/regenerator at the rear of the cooler (single-stack), and the compressor at the front and rear (double-stack). The compressor incorporates two clearance seals: one at the baseplate interface and one allowing the displacer/ regenerator shaft to slide through. The regenerator is constructed from a Vespel cylinder stacked with wire mesh punched to a circular cross section. The prototype is fitted with a LVDT sensor for the displacer/regenerator position and a linear capacitive sensor for the compressor piston position. The pressure and temperature are monitored at key positions. High-resolution real-time data are acquired during operation using a PC and a Computer-Boards DAS1000 multifunction PCI card. Custom software provides immediate calibrated views of system parameters. Software is also available to calculate phase shift information when required. The system (including software) has been validated during calibration.
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TEST FACILITY The test facility consists of a vacuum chamber, internal and external electronics, gas handling, data logging, and an pot (Figure 3). The link and heater system allow temperature control of the warm heat exchanger to simulate the radiator. A helium supply and filling system allow the cooler internal pressure to be varied. Additional support electronics and data acquisition are available outside the chamber. A drive signal is externally generated with variable frequency, amplitude, and offset. During testing the cooler is driven in open-loop mode with no position feedback.
MEASUREMENTS TO DATE The position, pressure, and temperature sensors have been calibrated. Particular care has been taken to measure the phase shifts within the position sensors. Initial measurements with an evacuated cooler show a compressor resonance at ~17.8Hz and a Q-factor of 45 indicating satisfactory alignment. The low frequency is due to the large moving mass of the magnet. The resonant frequency increases with filling pressure (Figure 4). Measurements of coil heating during operation show that up to 40W r.m.s of input power may be safely applied; this is beyond the expected operating point.
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Currently the prototype is ready to be run over a variety of operational parameters. Pressure driving of the displacer/regenerator has been observed with favourable phase shifts.
MODELING A simplistic 1D model has been developed based upon the work of Organ [1]. The model is written in C using the NAG Scientific Libraries. A dynamical core has been written and tested. Heat transfer is also included, but is yet to be tested The model treats the system as a set of coupled non-linear differential equations and solves these by a simple stepping algorithm that continues until the conditions at the end of a cycle match those at the beginning to a user defined tolerance. The pressure drop through the regenerator is incorporated using the static friction factor correlation from Kays and London [2]. This pressure drop, an external force, or a combination of the two can drive the displacer according to the user’s requirements. Heat transfer may be incorporated for the regenerator using NTUs calculated according
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to the Kays and London correlation. For the working spaces Annand’s correlations [3] are used, although the user may specify another if required. The model allows simple selection and optimisation of a range of operating conditions and parameters. Initial tests of the model have shown the expected behaviour. It remains to validate this model with the prototype cooler. Future improvements will include piston leakage, ‘shuttle’ heat transfer, and regenerator leakage.
REFERENCES 1.
Organ, A. J., “The miniature, reversed Stirling cycle cryo-cooler: integrated simulation of performance,” Cryogenics, no. 39 (1999), pp. 253-266.
2.
Kays, W. M., London, A. L., Compact Heat Exchangers, 3 ed. reprint, Krieger Publishing Company, Florida (1998).
3.
Annand, W.J.D., “Heat transfer in the cylinders of reciprocating internal combustion engines,” Proc. Inst. Mech. Engrs., vol. 177, no. 36 (1963), pp. 973-990.
rd
High Capacity Flexure Bearing Stirling Cryocooler On-Board the ISS T. Trollier, A. Ravex and P. Crespi(1) J. Mullié, P. Bruins and T. Benschop (2) (1) Air Liquide Advanced Technology Division, AL/DTA Sassenage, France (2) THALES Cryogenics B.V. Eindhoven, The Netherlands
ABSTRACT A high capacity Stirling cryocooler has been demonstrated at Development Model level during the year 2001 under AL/DTA and THALES Cryogenics co-funding. This development is based on a commercially-off-the-shelf LSF9320 type cryocooler from THALES Cryogenics (flexure bearing compressor and a standard wearing Stirling cold finger). It is now features a dual opposed piston compressor modified in order to pneumatically drive a Stirling cold finger, which is also implemented using flexure bearing technology. The pneumatically driven cold finger does not use any motor to achieve the movement and correct phase shift between the Stirling displacer and the pressure wave. The absence of this motor enhances the reliability of the system and simplifies the electronic control required to drive the system. This reliable and powerful cooler concept has been selected as the cooling system for the ESA/ CRYOSYSTEM vial freezers to be delivered by AL/DTA to ASTRIUM for use on board the International Space Station in 2006. The CRYOSYSTEM is a set of facilities for ultra-rapid cooling, preservation and storage of biological samples and protein crystals at -180°C. The actual performance is presented for various water-cooled heatsink locations that take into account the benefit of the Medium Temperature Loop (MTL) available on-board the ISS. Future performance improvements are also discussed.
INTRODUCTION AL/DTA was selected by ASTRIUM in February 2002 for the delivery of the vial freezers for the ESA/CRYOSYSTEM project. The CRYOSYSTEM is a set of facilities for ultra-rapid cooling, preservation, and storage of biological samples and protein crystals at -180°C on board the ISS. Cryocoolers are required to cool down the dewar magazine of the vial freezers as shown in Figure 1. Six Flight Models (FMs) are needed to support the CRYOSYSTEM vial freezers flow between the ISS, Kennedy Space Center and the Prime facilities. The delivery of the vial freezers is planned to start in 2006. Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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Figure 2 shows the present day flexure bearing Stirling cryocooler. This stainless steel version weighs 7.5 kg. The outer diameter of the compressor halves is 90 mm and the total length is approximately 200 mm. The compressor design is built around a moving-magnet linear motor that drives the pistons in a dual opposed configuration into a common compression chamber1. The moving magnet linear motor offers a big advantage over the conventional moving-coil design. This innovative concept allows the coils, which are the main source of gas contamination, to be placed outside the working gas. Additional advantages are the absence of flying leads and glass feed-throughs to supply current to the coils. Thus, moving magnet technology is applied in our compressor design to improve the reliability of the complete system. The main disadvantages of this configuration are the losses and the EMI, which are higher than in a conventional moving-coil design. High-performance, axiallymagnetized NdFeB magnets are used in the motors. Flexure-bearings are used in order to have a radial clearance between the piston and the cylinder. These flexure-bearings are round discs made of spring steel, with three arms. With this kind of flexure bearing, a very high radial stiffness can be reached. By changing the shape, the length and the thickness of the arm, the ratio between the axial and the radial stiffness can be changed without increasing the maximum stresses in the flexures. The fatigue limit of the spring steel is To have enough safety margin, the design-limit for the Von Misses stresses is set to as presented in Figure 3.
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CRYOCOOLER PERFORMANCE Water Cooled Heat Sinks As shown in Figure 4, a series of experiments was conducted using water heat sinking of the cooler at various dissipation locations. First, aluminium brackets with internal water flow channels were implemented alternatively around the compressor centre part, or around each coil and simultaneously around the centre part and the coils. The cold finger was also equipped with a water cooling bracket at the heat rejection path of the warm end as depicted schematically in Figure 4. As shown in the experimental results in Figure 5, the heat sinking of the centre part of the compressor alone is much more efficient than the heat sinking of the coils, so heat sinking of the
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entire compressor casing (centre part + coils) is not necessary to provide good performance. This will provide space for implementing mechanical support brackets at the coil locations and for integrating the water circuit internally to the stainless steel centre part of the compressor. Figure 6 displays measured load curves for various electrical input powers. These data are required to support the cool-down simulations of the vial freezers. At an input power of 150 W, a slope of about 180 mW/K is reached, which provides 7.3 W of cooling at 75 K or 10 W at 90 K. With half of the maximal input power during start up, the cooler is expected to provide more than 30 W of cooling to the cold magazine.
Compressor Skin and Internal Coil Temperatures During the cooling-power measurements presented above, the skin temperatures were measured at different places on the compressor case with thermocouples. Thermal insulation was put around the compressor coils to avoid additional natural convection, and the centre part was surrounded by the water-cooled heat sinks. Although the surface temperatures already give a good indication that the central heat sinking is very efficient, it is still interesting to complete the experiment with the measurement of the actual coil temperature in the same conditions. To get an indication of the temperature of a coil during operation, the resistance of the coil was measured immediately after switch off. Knowing this resistance at room temperature (Ta), we find the temperature with the following equation:
where is the temperature coefficient at the temperature To (for copper ). At room temperature (for a single coil). The calculated coil temperatures for various input powers and for 75 K cold tip temperature operation are also presented in Table 1. As shown, the temperature of the coils is very close to the temperature of the casing. This demonstrates again that water cooling of the centre part is very efficient and sufficient.
FUTURE PERFORMANCE IMPROVEMENTS Upcoming work will be dedicated to reducing the off-state parasitic heat losses; these can have a big impact when the cryocooler is turned off. In the present design, the cold finger is made entirely of stainless steel in order to ease the manufacturing of prototypes. As reported in Table 2, the total parasitic heat loss of the cold finger is 1400 mW for a 75 K to 300 K differential temperature. This is about one third of the total heat load into the cold magazine of the freezer.
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Some simulations are presented herein which consist of replacing stainless steel by high performance titanium alloy and also reducing the thickness of the cold finger and the regenerator tubes. The impact of the length of the regenerator and cold finger tubes is also depicted. Thus, with a length of 80 mm and the use of thin Ti6A14V tubes, the parasitic heat losses are expected to be reduced about 70%. This should provide 1.4 W of additional cooling at 75 K with 80 W of mechanical input power (about 125 W electrical input). The balance between thermal performance and mechanical resistance is also under analysis at the dewar level; the focus here is on the induced loads into the cold finger during launch and landing phases. In this framework, flexible thermal link solutions are presently under design.
CONCLUSIONS A compact, reliable and efficient flexure bearing Stirling cooler is under optimization for use on board the ISS within the CRYOSYSTEM vial freezers program. The displacer is supported by flexure bearings and is pneumatically driven; this is expected to enhance the reliability of the system and simplifies the electronic control required to drive the overall system. With traditional materials, a cooling capacity of 7.3 W at 75 K (or 10 W at 90 K) has been reached with 150 W input power and with water-cooled heat sinks. The future implementation of high performance titanium alloys is expected to provide a significant reduction of the parasitic heat losses of the cold finger, together with an important increase in the cooling capacity.
REFERENCES 1. M. Meijers, A.A. J. Benschop and J.C. Mullié, “High Reliability Coolers under Development at SignaalUSFA”, Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2001), pp. 111-118.
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Space Flight Qualification Program for the AMS-02 Commercial Cryocoolers K. A. Shirey, I. S. Banks, S. R. Breon, and R. F. Boyle NASA/Goddard Space Flight Center Greenbelt, MD 20771
ABSTRACT The Alpha Magnetic Spectrometer-02 (AMS-02) experiment is a state-of-the-art particle physics detector containing a large superfluid helium-cooled superconducting magnet. Highly sensitive detector plates inside the magnet measure a particle’s speed, momentum, charge, and path. The AMS-02 experiment will study the properties and origin of cosmic particles and nuclei including antimatter and dark matter. AMS-02 will be installed on the International Space Station on Utilization Flight-4. The experiment will be run for at least three years. To extend the life of the stored cryogen and minimize temperature gradients around the magnet, four Stirling-cycle Sunpower M87N cryocoolers will be integrated with AMS-02. The cryocooler cold tip will be connected via a flexible strap to the outer vapor cooled shield of the dewar. Initial thermal analysis shows the lifetime of the experiment is increased by a factor of 2.8 with the use of the cryocooler. The AMS-02 project selected the Sunpower M87 cryocoolers and has asked NASA Goddard to qualify the cryocoolers for space flight use. This paper describes the interfaces with the cryocoolers and presents data collected during testing of the two engineering model cryocoolers. Tests include thermal performance characterization, electrical characterization, launch vibration testing, and magnetic compatibility testing.
INTRODUCTION Since November 2001 two Sunpower M87 engineering model cryocoolers have been under test at NASA Goddard for the Alpha Magnetic Spectrometer-02 (AMS-02) International Space Station (ISS) Project. They are designated Engineering Model #1 (EM#1) and Engineering Model #2 (EM#2). NASA Goddard will be qualifying four Sunpower M87N cryocoolers for space flight use and two cryocoolers for flight spares on AMS-02. Delivery of the flight cryocoolers to Goddard is expected at the end of 2002. The engineering model cryocoolers have undergone extensive thermal characterizations and magnetic compatibility testing. EM#2 has undergone electrical characterizations and vibration qualification. The engineering model cryocoolers have each accumulated over 2,000 hours of run time with no change in thermal performance. A systematic trend analysis has been implemented to track the performance of the engineering model cryocoolers throughout the course of the project.
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CRYOCOOLER INTERFACES AMS-02 is scheduled to be launched in January 2006 and will be installed on the ISS for a minimum three year mission. The experiment consists of a large superconducting magnet and a number of highly sensitive detectors that will measure a particle’s speed, momentum, charge, and path in an effort to search space for the presence of dark matter, strange matter and antimatter. The AMS-02 superconducting magnet will be cooled by 2,600 liters of superfluid helium in a large annular tank. The magnet, superfluid helium tank, layers of super-insulation and 4 vaporcooled shields are suspended within a toroidal vacuum case. The vacuum case is machined out of aluminum with two large support rings on the top and bottom of the outer cylinder. The cold mass is approximately 2,090 kg. Space Cryomagnetics, Ltd. of Abingdon, England is developing the magnet, helium tank, and vapor-cooled shields. The vacuum case is being developed by Lockheed Martin in Houston, TX. In an effort to extend the life of the stored cryogen, four Sunpower M87N cryocoolers will be used to cool the outer vapor cooled shield. The baseline performance requirement is a total of 16 watts of heat lift at 80 K with 400 watts of input power. Initial thermal analysis shows the lifetime of the experiment is increased by a factor of 2.8 with the use of the cryocoolers. To minimize thermal gradients on the vapor cooled shield two cryocoolers will be mounted to ports on the upper vacuum case support ring and the remaining two cryocoolers will be mounted to the lower vacuum case support ring. Figure 1 shows a view of two of the cryocooler port locations on the AMS-02 vacuum case. The remaining two cryocooler port locations are 180 degrees from its pair on either support ring. The cryocooler mounting brackets must provide a hermetic seal to the vacuum case and must thermally and mechanically decouple the cryocooler from the vacuum case. A compliant (flexible, soft) mount is required in order to allow force attenuation using a passive balancer system. The cryocooler mounting bracket is being designed and analyzed by Swales Aerospace in Beltsville, MD. The cold tip of the cryocoolers will be connected via a flexible strap to the outer vapor cooled shield of the dewar. Integration of the strap will be through a 101.6 mm diameter access port adjacent to the cryocooler mounting port. The strap will span a distance of approximately 100 mm and allow for relative motions no more than 12 mm between the cold tip and the vaporcooled shield. Motion of the strap is expected during launch, vacuum pump down, magnet cool down, magnet charging and discharging and in the case of a quench. The thermal strap is being developed by Space Cryomagnetics, Ltd in Abingdon, England.
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Each cryocooler will reject heat to a capillary pump loop sunk to a direct condensing zenith octagonal radiator. One quadrant of the radiator will be dedicated to each cryocooler. The thermal control system will provide a nominal cryocooler operating temperature between 0°C and +10°C. Survival heaters will be implemented to maintain a minimum non-operating temperature of -40°C and to assure a minimum turn on temperature of -10°C. The maximum allowable operating temperature is 40°C. The external thermal control system is being developed by Carlo Gavazzi Space of Milano, Italy and OHB Systems of Bremen, Germany. The cryocoolers will be powered from either or both of the ISS 124 V DC buses (main and auxiliary). The electronics must provide the capability of being powered from either bus and must maintain galvanic isolation between the two buses. The power drive configuration has not yet been selected. Studies are being conducted on which waveform drive would be the most efficient. A sinusoidal or square waveform drive would use a pulse-width modulated (PWM) switching amplifier at 200 KHz. These methods will have switching losses as well as high EMI. A pulse duration square wave drive is being considered which would eliminate the switching losses and result in lower EMI and higher efficiency. The cryocooler electronics are being developed by ETH-Zurich in Switzerland. In order to eliminate beat frequencies, it is preferred that all cryocooler drive signals be synchronized. We are considering running each cryocooler 45° out of phase relative to each other such that the peak current draw will not occur simultaneously. The cryocooler system has been allocated a maximum of 400 W, which does not include power to the electronics. Nominal operation will provide 100 W to the 4 cryocoolers. The maximum operational scenario will provide 150 W to 3 cryocoolers. The cryocooler controller will monitor and maintain the following housekeeping parameters: cryocooler cold tip temperature, cryocooler case temperature, motor voltage, motor current, phase angle between the motor current and voltage, and cryocooler acceleration. The controller will provide automated shutdown in the event of an over-limit condition occurring with the heat reject temperature, motor current, and motor voltage. Ground station uplink commands will be available to enable/disable the launch lock mode and adjust the input voltage amplitude. AMS-02 will be the first space flight mission that will have Stirling-cycle cryocoolers operating within a substantial steady-state magnetic field. The cryocoolers will be mounted in locations with a magnetic gradient over the entire length of the cryocoolers and fields as high as 925 Gauss perpendicular to the cryocooler axis and 400 Gauss along the cryocooler axis. Tests are continuing to determine the effects of an external magnetic field on a cryocooler motor.
SUNPOWER M87 CRYOCOOLER The Sunpower M87 cryocooler is a commercial Stirling-cycle cryocooler designed for high volume manufacturing.1 The compressor piston is driven by a moving-magnet linear motor. The amplitude of the input voltage to the linear motor controls the stroke of the compressor piston. A pressure wave generated by the compressor piston drives the displacer. The displacer piston shuttles gas back and forth from the cold end to the warm end through a random fiber regenerator. The gas is expanded in the cold end to absorb heat from the thermal load and compressed at the warm end to reject heat to the environment. A gas bearing system is utilized to center the compressor and displacer pistons and to prevent touch contact between the moving parts. Vibration suppression from the linear motor is implemented by the use of a passive (tuned spring-mass) balancer system. The compressor has a broad resonance around 60 hertz, while the passive balancer has a narrow resonance at 60 hertz. The Sunpower M87 was designed to provide 7.5 watts of cooling at 77 K with 150 watts input power while operating at a 35°C heat reject temperature. This cryocooler has a designed lifetime of 40,000 hours. The M87 was not designed with space applications in mind. Cooler orientation during operation is restricted to the vertical orientation with the cold end facing down. Modifications were necessary to make the unit acceptable for space flight use. Sunpower
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made changes to the M87 to allow operation in all orientations. Sunpower refers to this modified cryocooler as the M87N. In order to accommodate the generation of a sinusoidal waveform drive from the ISS 124 V DC bus, we asked Sunpower to reduce the motor impedance of the M87N in order to achieve a maximum input voltage to the cryocooler of 75 V RMS. A flat mounting plate was added to the rear of the cryocooler to allow additional mounting locations.
ENGINEERING MODEL PERFORMANCE Test Setup Description The AMS-02 project purchased both a standard Sunpower M87 and the modified M87N for the two engineering models. Before shipping the units to Goddard, each cryocooler was run by Sunpower for a 250-hour qualification period. Upon arrival at Goddard, the cryocoolers were inspected and prepared for thermal performance characterization. The coldfingers on EM#1 and EM#2 were instrumented with a heater, to simulate a thermal load, and a Lakeshore silicon diode, then wrapped with 5 layers of multi-layer insulation. Resistance temperature detectors (RTDs) were mounted to the cryocoolers’ heat reject collars and cases to monitor the environmental temperatures. The cryocoolers’ heat reject temperatures were maintained by laboratory recirculating chillers. EM#1, tested in a vacuum chamber, was mounted in the vertical orientation, cold finger down. EM#2, tested on an optical bench, was mounted in the vertical orientation, cold finger up. The coldfinger on EM#2 was enclosed in a vacuum bonnet connected to a vacuum header. Both cryocooler mounts were designed to be compliant to allow force attenuation by the passive balancer. Each engineering model cryocooler is protected with Goddard-developed laboratory cryocooler shutdown electronics that protect against cold tip temperature overheat, cryocooler body temperature overheat, and loss of vacuum. A display showing the total number of hours accumulated on the cryocooler is implemented on the front panel of the electronics. The electronics can be switched between a sinusoidal waveform drive and the Sunpower control electronics drive. The electronics also allow for external input where an arbitrary waveform function generator and power amplifier combination could be used to produce a non-standard waveform. A data acquisition program written in Lab VIEW data logs the motor voltage, current, power, power factor, cryocooler body temperature and cryocooler cold tip temperature every minute.
Thermal Characterizations Extensive thermal characterizations were conducted on both engineering model cryocoolers to measure the cold tip temperature as a function of input power to the compressor, heat reject temperature, and heat lift. Compressor power was measured at the input to the compressor; therefore power losses in the drive electronics are not reflected in the results. EM#1 was tested with heat reject temperatures of –10°C, 0°C, 20°C, and 40°C. EM#2 was tested with heat reject temperatures of 20°C and 40°C. Both cryocoolers were tested with compressor power levels ranging from 50 W to 150 W and thermal loads ranging from 0 W to 10 W. Figures 2 and 3 show the thermal performance for EM#1 and EM#2, respectively. The thermal performance characterizations have verified that the four M87 cryocoolers will meet the AMS-02 cooling requirement of 16 W at 80 K with 400 W of input power. The engineering model testing has shown that at a reject temperature of 20°C, an input power of 100 W, and a cold tip temperature of 80K, EM#1 can lift 5.3 W and EM#2 can lift 5.7 W. A subset of the thermal performance characterization is repeated at least every three months and compared against the baseline performance curve established during acceptance testing of the cryocooler to verify there has been no thermal performance degradation. We have not seen any thermal performance degradation in either of the engineering model cryocoolers.
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Electrical Characterization A comparison study between different waveform drives and cryocooler efficiency was conducted on EM#2. To date testing was done using a sinusoidal waveform and a square waveform. A pulse duration square wave will be studied later this summer. While driving EM#2 with a sine wave, the amount of heat lift was measured while maintaining a constant input power and cold tip temperature (CTT). This test was repeated using a square wave drive. Table 1 shows the calculated percent Carnot and the percent heat lift for both sinusoidal and square wave drives at nominal input power to the cryocooler. The percent
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heat lift is a ratio of the heat lift for the square wave to the heat lift for the sine wave. The percent heat lift is normalized to the sine wave drive and therefore we show a percent heat lift of 100% for the sine wave drive. A slight drop in efficiency was found when driving the cryocooler with a square wave, compared to that when using a sine wave.
Random Vibration To qualify the M87N design, launch vibration tests have been conducted on EM#2. The cryocooler was rigidly mounted, with the motor coils shorted, and was subjected to a minimum workmanship level of 6.8 Grms in all three axes. The AMS-02 project expects that the vibration transmitted through the primary structure to the experiment components will be much smaller than minimum workmanship levels. A Kistler tri-axis accelerometer mounted to the cryocooler case collected the response data. Figure 4 shows EM#2 mounted on the vibration shaker. Thermal performance testing pre- and post-vibration indicates there was no thermodynamic degradation as a result of random vibration testing. Upon completion of the compliant mounting bracket design and machining, EM#2 will be put through random vibration a second time to verify the integrity of the compliant mounting bracket.
Magnetic Compatibility Magnetic compatibility testing2 was conducted on EM#1 and EM#2 at a cyclotron facility at the Massachusetts Institute of Technology. Figure 5 shows the test setup with a cryocooler mounted to the cyclotron magnet in the background and the cryocooler drive electronics in the foreground, a safe distance away from the magnet. The cryocoolers successfully operated in magnetic fields and gradients with magnitudes that were greater than twice the maximum expected AMS-02 levels. At these levels, we conducted multiple on-off cycles on EM#2. The cryocoolers were operated in the magnetic fields at compressor power levels ranging from the minimum to maximum allowable levels. A small thermal performance degradation of 1-4% was observed at field levels larger than the maximum expected. We are continuing to examine the degradation in performance at very high magnetic fields, however, at the nominal AMS-02 cryocooler operating conditions, no performance degradation was apparent.
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Cryocooler Trend Analysis A systematic trend analysis has been implemented to track the performance of EM#1 and EM#2 throughout the course of the project. The performance of the flight cryocoolers will be tracked using the same method. Parameters being tracked are the hours of operation in a particular orientation, number of start/stop cycles, number and cause of an automatic shutdown for out-of-limit conditions (including facility-related shutdowns), helium leak rate, and thermal performance. Table 2 gives a summary of the trend analysis that has been recorded as of the end of May 2002. The total accumulated hours listed does not include the 250 hours accumulated on each unit before being shipped from Sunpower.
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SUMMARY During the past year, NASA Goddard has been testing two Sunpower M87 engineering model cryocoolers for the AMS-02 project. These include a standard off-the-shelf M87 as well as a modified M87 for space flight application. The engineering model cryocoolers have just over 2,000 hours of accumulated run time. Extensive thermal performance characterizations have verified that four M87 cryocoolers will meet the AMS-02 cooling requirement of 16 W of heat lift at 80 K with 400 W of input power. EM#2 was subjected to random vibration at minimum workmanship levels. No thermal performance degradation was detected after vibration. The engineering model cryocoolers are capable of operating in magnetic fields twice the maximum AMS-02 expected levels. Electrical characterizations using various waveform drives is continuing in order to determine the most efficient method for driving the cryocoolers.
ACKNOWLEDGMENTS The authors wish to acknowledge Ed Quinn of Orbital Sciences for his support in all of the engineering model testing and Rick Foster of the MIT Center for Space Research for his overall support throughout the course of the project. We also wish to thank Dr. Ulrich Becker of MIT for access to the cyclotron facility and Messrs. Stephen Harrison and Steve Milward of Space Cryomagnetics, Ltd. for calculations of the magnetic field parameters. Much credit is due to Dr. Henning Leidecker of NASA Goddard Space Flight Center in formulating approaches to the magnetic compatibility issues.
REFERENCES 1.
Unger, R.Z., "The Advent of Low Cost Cryocoolers," Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2001), pp. 79-86.
2.
Breon, S.R.et al., “Operation of A Sunpower M87 Cryocooler in a Magnetic Field”, to be presented at the 12th International Cryocooler Conference, Cambridge, MA, 2002.
Thermodynamic Performance of the Ball Aerospace Multistage Stirling Cycle Mechanical Cooler W.J. Gully, D. Glaister, E. Marquardt, R. Stack, and G.P. Wright Ball Aerospace & Technologies Corp. Boulder, CO, USA 80306
ABSTRACT Ball Aerospace has been working on highly reliable Stirling cycle multi-stage mechanical coolers for more than a decade. We have produced, characterized, and delivered one-, two-, and three-stage coolers on a number of programs. Currently we are developing a more producible two-stage cooler with twice the capacity that uses the same envelope. The tendency for larger coolers to be relatively efficient has been a benefit. We have incorporated changes to the motor to improve its electrical efficiency. This paper discusses a full range of qualification testing on the mechanical unit. We completed thermal characterization to the point of regenerator saturation. We measured vibration export on a 6-axis dynamometer, and carried out vibration testing to 17 Grms. As a result of the vibration tests, we have corrected a few mistakes we made in our zeal to be more producible.
MOTIVATION FOR THE SB235 CRYOCOOLER The mechanical cryocooler development work at Ball Aerospace reflects the interest in multistage, long-life mechanical cryocoolers by NASA and DOD customers. Our original coolers were ‘research’ articles meant to explore in detail the operation of these springsupported, Oxford-style, non-contacting mechanisms. After acceptance testing to GEVS level requirements[1], each was placed into extended life testing. Currently, the SB230 30 K cryocooler has accumulated >14,700 hours, and the AFRL SB335 35/60 K cryocooler has accumulated >12,000 hours. These tests help us understand the long-term behavior of our coolers. As illustrated in Figure 1, these coolers were designed to support ‘first principles’ measurements of cryocooler behavior and included many specialized testing features. With this initial testing effort completed, we wanted to streamline the cryocoolers and tailor them more for flight applications. We also wanted to redesign it to meet current perceived system needs. Mechanical cryocooler loads have grown in step with the increasing size and complexity of cryogenic systems. A single cooler now is called upon to cool different parts of a system (e.g., both optics and detectors) as well as a redundant, non-operating copy of itself. Consequently, we focused on increasing the capacity of our dual temperature 35/60 K[2] cooler. Although we would have benefited more in the scale efficiency had we upgraded our three-stage design, we eschewed the efficiency advantage of the third expansion stage in favor of the producibility Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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advantage of a two-stage design. Still, as shown in Table 1, we achieved a substantial increase in the specific power and mass efficiency with our new cooler. The new cooler is shown in Figure 2, and some of the simplifications made on the displacer interface are highlighted in Figure 3. In addition to the hardware changes, we have streamlined and simplified our processes and procedures, a lesson learned from building a 30 K style cryocooler, the SB 160, on the HIRDLS flight program.
SB235 CRYOCOOLER TEST PROGRAM We performed a number of tests on the SB235 mechanical unit, including thermal characterization, export vibration, and two launch vibration campaigns.
THERMAL CHARACTERIZATION A diagram of the cryocooler’s performance under nominal conditions is shown in Figure 4. The cooler is most efficient under these conditions, and these curves were used to select the data shown in Table I. Power to the cooler can be increased by increasing its operating frequency and charge pressure. This generally results in more refrigeration, but at some point the regenerator will saturate and the available cooling power will peak. Figure 5 shows that the midstage responds to this increase in power, but that the net cold stage cooling has decreased. This occurs because the cold stage regenerator becomes less efficient at the higher mass flows. This is clearly demonstrated in Figure 6. We increased the power by stepping up the pressure and frequency and
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observed the cooling of the middle and cold stages. The midstage kept responding, but the cold stage reached an asymptote load temperature of 36 K under these conditions.
USING SATURATION TO SHIFT LOADS IN A STIRLING COOLER The ability to shift refrigeration capacity from one stage to another is an advantage in space applications because of the difficulty in predicting the loads in advance. The saturation effect discussed above gives the ability to do this with only a small power penalty. We can effect this transfer by using the phase angle between the compressor and displacer, giving the Stirling cooler another variable that can be used to optimize its performance. A Stirling cooler works best by shifting the phase to 90 deg, which results in the largest pV diagram. But it also compounds the mass flow that increases the regenerator losses in both regenerators. This is most significant when the regenerator loss is already high as it is in the cold regenerator, and can lead to decreased net cooling at the cold stage. Furthermore, in the twostage cooler this extra regenerator loss appears as extra refrigeration at the midstage, amplifying the shift in cooling.
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As shown in Figure 6, the cold regenerator became saturated at the higher power levels, but the warm regenerator did not. Figure 7 shows that under these conditions the warm-stage net refrigeration increases but the cold-stage net refrigeration actually decreases. In effect, the cooling switches from the lower to the upper stage. The power penalty is slight because of the higher Carnot factor for producing the extra cooling at the lower temperature of the colder stage.
EXPORT VIBRATION Our study also characterized the vibration export from the cooler. In tests, the cooler ran under nominal conditions with loads of 0.5 W and 1 W on the two stages. Figure 8 shows the axial export data when the cooler is under electronic control along the same axis. With the axial loads actively balanced this way, the worst loads are found in the cross axis. Under these conditions most of the cross axis forces were under 0.2 Newtons, but typically there was an exceptional harmonic whose force peak amplitude approached 0.5 Newtons.
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LAUNCH VIBRATION We tested our previous coolers to the 14.2 Grms launch vibrations typical of GEVS. We tested this cooler at even higher 17 Grms levels and in general succeeded, but with some minor difficulties. The cooler proper passed, although the counterweight mechanism will need to be requalified separately. In retrospect, we were taking two types of risk at these high levels. The first is that the cooler was built on an IR&D program, where hardware is typically built without the expensive quality assurance associated with a flight program. Successful testing depends upon the kind of attention to detail that QA brings. The second is that we had just finished a producibility revision of our design. Even though each change improved the cooler, we still deviated from previously qualified designs. FUTURE WORK To re-test the SB235 we plan to re-assemble our cooler and resume testing. This will include a checkout, a hermetic seal, thermal vacuum testing, and extended life testing. We will also do a counterbalance re-test to 17 Grms. E300 Electronics. We are continuing work on the E300 electronics. This is a derivative of the E200 electronics that was built for our flight cooler. The most important change was the upgrade of the power section for a capability of 200 W. These electronics are ideally suited for driving the SB235 and similar coolers under higher power applications. SB235E. We are designing an even more powerful cooler, the Enhanced version of the SB235. With the experience we gained with the SB235, there are ways to modify the cooler to increase its mid-stage capacity to about 8 W at 116 K, and 1.5 W at 36 K. REFERENCES 1. 2.
General Environmental Verification Specification (GEVS) for STS and ELV Payloads, Subsystems, and Components, Revision A, Table 2.4-4, Goddard Space Flight Center. Gully, W.J., et al., “Qualification Test Results for a Dual-Temperature Stirling Cryocooler,” Cryocoolers 10, Klewer Academic/Plenum Publishers, New York, (1999), pp. 59-65.
Performance Characterization of the Ball Aerospace 35/60K Protoflight Spacecraft Cryocooler C.H.Y. Bruninghaus, B.J. Tomlinson and N. Abhyankar* Air Force Research Laboratory, Space Vehicles Directorate Kirtland AFB, NM 87117 *Dynacs Engineering Co. Albuquerque, NM 87106
ABSTRACT This paper outlines results from the preliminary characterization of the Ball Aerospace 35/60K Protoflight Spacecraft Cryocooler performed at the Air Force Research Laboratory (AFRL) Space Vehicles Directorate. The cooler was developed under joint sponsorship by NASA Goddard Space Flight Center and AFRL for space based infrared sensing applications. This cryocooler is a unique, three stage Stirling cycle protoflight cryocooler unit capable of simultaneously lifting 0.4 W at 35 K and 0.6 W at 60 K. The physical characteristics and component details of the cooler are included in this report. The thermodynamic performance of the cryocooler is presented in terms of the effect of variable heat loads on the cold end, mid-stage, and upper stage temperatures. Based on the acquired data, performance models are presented using two-variable regression methods.
INTRODUCTION The Ball Aerospace 35/60K cooler is a protoflight, three stage, split-Stirling, mechanical cryocooler designed to support space based infrared sensing applications. It was developed under joint sponsorship by NASA Goddard Space Flight Center and the Air Force Research Laboratory. The cryocooler provides simultaneous refrigeration at two different temperatures. It is designed to lift 0.4 W at 35 K at the cold end and 0.6 W at 60 K at the mid-stage with 80 W of total input power. This enables cooling of a focal plane or detector at one temperature while simultaneously cooling a radiator or heat shield at a higher temperature. Although the cooler has three stages of fixed regenerators, only two are capable of supporting heat loads. The third stage is used to intercept internal parasitic heat loads. Fixed regenerators are used because they improve reliability (by reducing the number of moving parts), allow for easy integration, and help mitigate vibration. Characterization experiments at the AFRL Cryogenic Cooling Research Facility (CCRF) are conducted to investigate a cooler’s performance over the complete range of thermodynamic and environmental parameters that encompass the normal space mission requirements. Sometimes, special requirements are established by other sponsoring organizations and are integrated into the experiment plan. Characterization typically involves verification of the cooler’s ability to meet Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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design requirements, cooldown time determination, verification of the cooler’s ability to maintain long-term temperature stability at design conditions, thermal vacuum and thermal cycling experiments at extreme heat rejection temperatures, and parasitic heat load determination. These experiments are interlaced with periodic baseline experiments that are done to track any changes in performance over time. The complete set of characterization experiments, which extends over a wide range of operating parameters and environmental conditions, provides system designers and spacecraft integrators with a performance map of the capability of this cryocooler. The cooler has dual opposed compressors and a displacer with a momentum counter-balancer to cancel vibration. All pistons are mounted on spiral flexures, which provide frictionless motion between the cylinders and piston liners. The flexures improve the reliability of the system, which is required for space missions. The cooler is controlled by flight-like control electronics. The command parameters are modified through a software program via an RS232 interface. In the laboratory, the unit is mounted in a 24” vacuum chamber with a heat rejection fluid system. The cooler is instrumented with silicon diodes to measure temperature on the cold end, mid-stage, and upper stage. Two heaters are mounted to the cold end and mid-stage to apply a heat load, simulating the heat produced by the sensor package and heat shields. The integrated test assembly is shown in Figure 1, which shows the balancer in front of compressor. The serpentine hoses circulate chiller fluid used to maintain the heat rejection temperature. Thermocouples (TCs) are attached to various locations on the cooler to obtain its temperature distribution. The chiller setpoint temperature is adjusted to provide a fixed heat rejection temperature for nominal operation, or to provide a cyclic variation between extreme heat rejection temperatures for the thermal-cycle and thermal-vacuum tests. The control electronics box is also integrated with the cooler in the vacuum chamber. The cooler is equipped with Ball Aerospace’s uniquely designed Differential Impedance Transducers (DITS) to monitor the internal clearance between the piston and cylinder sleeve inside the compressors and displacer. The radial position of the armature is tracked by proximity sensors.
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The coaxial feed-through cables for the sensors are shown in Figure 1 at the top end of the compressor and displacer casings. The DITS system is a useful tool, but it is difficult to implement because it requires precision alignment, careful bookkeeping of voltages from multiple proximity sensors (eight per cylinder), and requires an extremely sensitive electrical signal measurement system.
PERFORMANCE EVALUATION A typical cool down curve is shown in Figure 2, which shows cold end and mid-stage temperatures, stroke length as a percentage of maximum stroke length, and input power to the system. The cooler is run with fixed stroke lengths and no heat load applied to the cold end or mid-stage. During most of the cooldown, the mid-stage temperature is lower than the cold end temperature. This indicates that some internal heat shuttling may be occurring between the cold end and mid-stage until steady state is reached. After steady state is reached, 0.4 W is applied to the cold end, and 0.6 W is applied to the midstage. For these heat loads, with nominal compressor and displacer stroke lengths, the design requirements state that the cooler should maintain temperatures of 35 K at the cold end and 60 K at the mid-stage. The corresponding input power, excluding the power consumed by electronics, should be less than 80 W. The results from the design point verification are shown in Table 1. The corresponding long-term stability performance (over a month’s worth of data) is shown in Figure 3. To understand the behavior of the cooler for other heat load conditions, a matrix of seven cold end heat loads (0, 0.1, 0.3, 0.5, 0.7, and 0.9 W) and four mid-stage heat loads (0., 0.3, 0.7, 0.9 W) was evaluated. The cold end loads were gradually increased while the mid-stage heat load was held
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constant. A steady-state criterion is always defined, which serves as a condition to be met before the heat loads are changed to acquire the next data point. Normally, the steady-state criterion of ± 0.15 K for 15 minutes is used for characterization experiments. However, this criteria was tightened to ± 0.1 K for 50 minutes for the cold end, ± 0.6 K for 30 minutes for the mid-stage, and ± 0.15 K for 15 minutes to detect steady-state conditions accurately during automated experiments. Cold end, mid-stage, upper stage, and heat rejection temperatures, compressor and displacer stroke lengths, input power, and other parameters were recorded for each combination of cold end and mid-stage heat loads. These data are presented in Figures 4 through 6. The cold end and midstage temperatures are plotted in Figure 4, which shows that cold end temperature is not very sensitive to the mid-stage heat load. The mid-stage temperature lines are almost parallel to each other with respect to the cold end heat load. This trend indicates that both the temperatures vary
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approximately linearly with respect to the mid-stage load. The upper stage temperature (closest to the base of the regenerator) is plotted in Figure 5. It varies quadratically with respect to both the cold end and mid-stage heat loads and peaks near where the cold end heat load is 0.5 W. The input power, shown in Figure 6 is mostly linear with respect to both mid and cold end loads, but has an inflection near the no-load at tip. At the no-load condition, the parasitic load determines the power consumption. Based on the temperature behavior in Fig. 4, it can be observed that for certain combinations of mid and coldend loads, the cold end temperature is higher than the mid section at steady state. The steady state data acquired for the loadlines presented in Figures 4 through 6 provide valuable insight into the cooler’s behavior. The thermodynamic variables can be empirically related as a function of mid-stage and cold end heat load. With the cold end and mid-stage loads as independent variables, the following equations fit the data with reasonable accuracy (statistically ).
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In Figure 4 the narrow spread between cold end temperature lines indicates that constant cold end temperature lines are achievable for a short range of cold end heat loads. For example, a constant temperature of 35 K at the tip can be maintained with cold end heat loads ranging from 0.33 W to 0.44 W over the range of measured mid-stage heat load values. These values can be derived from Eq (l). The equations and figures presented above were generated for constant stroke length (90% on both compressors and the displacer) and a constant heat rejection temperature of 300 K. These two parameters have significant influence on the behavior of the cooler. Future characterization experiments, in addition to thermal vacuum, thermal cycling, parasitic load determination, will include variation of these parameters for regression analysis.
PERFORMANCE SHIFT Over the past year and a half, the cooler has been running in endurance at its design conditions, with occasional breaks during characterization experiments and equipment exchanges. During this time, the cooler’s performance appears to have shifted. After the cooler was integrated with the vacuum chamber in January 2001, it was run with the stroke length command set at 91%. The maximum allowable stroke length is set at 95% to mitigate the risk of damaging the cooler by making contact at the end stops. However, in May 2001, AFRL technicians noticed that the temperature started gradually increasing while the stroke length command remained constant. The decision was made to increase the stroke length command to 92% in order to maintain a cold end temperature of 35 K at the design conditions. Over time, the cold end temperature continued to rise with all other conditions remaining constant. Figure 7 shows the constant stroke length data from January 2001 through February 2002.
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During February 2002, Ball personnel came out to AFRL to help troubleshoot the cooler. The decision was made to switch from running the cooler with constant stroke length commands to running in constant temperature control mode. The cold end temperature was set at 37 K to avoid having the cooler running at 95% while trying to reach and maintain 35 K. Since switching to constant temperature control, the stroke length and input power have gradually continues to increase, while maintaining a constant cold end temperature of 37 K with all other conditions set at their nominal values. Figure 8 shows temperature control data from January through May 2002. Load lines were also run at various times during the past year and a half. The load lines also show a shift in performance. For constant stroke length set at 90%, the temperature for each heat load combination was consistently higher. AFRL is currently investigating the cause of the performance shift. The investigation is taking into consideration both the experiment stand and the cooler as the cause of the shift. The Ball cooler sits in a vacuum chamber that shares its vacuum pump with four other chambers, which are in parallel down stream of the main vacuum line. One of these chambers houses the Raytheon (formerly Hughes) Standard Spacecraft Cryocooler (SSC), which is known to have a leak of approximately 1 atm per year. One possibility is that contaminates from the SSC are being collected on the Ball cooler cold end (it runs at the lowest temperature of the other coolers), increasing the parasitic heat load and thus causing the cooler’s inability to reach design conditions. The cooler was shut down and allowed to warm up to approximately 100 K. As the cooler warmed up, a gradual increase in the vacuum level was noted, indicating that something was released into the vacuum. Figure 9 shows the decrease in vacuum level with increase in temperature. After the warm up, the cooler was started again with the temperature control set at 37 K. There was no change in stroke length or input power. However, the 100 K outgassing temperature was probably not hot enough to have released any water-ice film that may have collected on the cold end.1
CONCLUSIONS The Ball Aerospace 35K/60K is undergoing characterization at Air Force Research Laboratory Cryogenic Cooling Research Facility. Preliminary experiments have shown that it can simultaneously lift 0.4 W with a cold end temperature of 35 K and 0.6 W with a mid-stage temperature of 60 K. The data analysis, with heat loads at cold tip and mid-stage as independent variables, provides an empirical relationship between heat loads, temperatures, and input power. The cause of the performance shift is still being investigated by AFRL and Ball Aerospace personnel. Once the
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performance shift investigation is complete, the effects of heat rejection temperature and stroke length variation on the cooler’s performance will be investigated.
ACKNOWLEDGMENTS Authors would like to acknowledge advisory assistance from Maj. Barrett Flake and Mr. Tom Roberts of the Air Force Research Laboratory Space Vehicles Directorate.
REFERENCES 1. Ross, R.G., Jr., “Cryocooler Load Increase due to External Contamination of faces,” Cryocoolers 12, Kluwer Academic/Plenum Publishers, New York, 2003.
Cryogenic Sur-
Continued Characterization Results for the Astrium 10K Developmental Cryocooler S.A. Yarbrough1, B.A. Flake1, B.J. Tomlinson1, and N. Abhyankar2 Air Force Research Laboratory, Space Vehicles Directorate Kirtland AFB, NM 87117-5776 2 Dynacs Engineering Co., Albuquerque, NM 1
ABSTRACT Under a technology development program, Astrium in Stevenage, United Kingdom, developed a Stirling cycle cryocooler with four Oxford flexure compressors and a two-stage expansion cold end designed to lift 45 mW at 10.4 K. The cooler has completed a performance characterization evaluation at the Air Force Research Laboratory Space Vehicles Directorate, Kirtland AFB, NM. This report presents the characteristic load lines, results of optimization trials, where frequency, phase angle, and expander stroke lengths were varied, and lessons learned during the characterization process. A performance map of cooler performance at several rejection temperatures is also included. Recent anomalous behavior and performance degradation over time will be discussed. The cooler is currently undergoing long-term endurance evaluation and anomaly investigation.
INTRODUCTION In order to enhance the Air Force capability in space surveillance and missile detection with the silicon based very long wavelength infrared sensors, the Missile Defense Agency (MDA) and US Air Force sought to develop a cryocooler for cooling in the 10K temperature region. The 10K cooler operation can also be effective for cooling semiconductor electronics. The development of such coolers for space application is a technological challenge due to the high power requirement and specialized cold finger configuration required to achieve cooling at such low temperatures. Astrium developed a prototype cooler (Figure 1) based on their previous experience with twostage-20 K and single stage 50-80 K cooler programs. European Space Agency supported these programs. To enhance its capabilities to 10 K, various modifications were implemented, including increasing the number of compressors from two to four and a refinement of the cold finger design and the associated regenerator. The cooler was delivered to the Air Force Research Laboratory (AFRL) Cryocooler Cooling Research Facility for further detailed characterization. The following items were undertaken as part of the characterization process: 1) heat rejection temperature effects on cool-down, 2) design and load line characteristics, 3) optimization evaluation (frequency, phase, expander stroke length), 4) performance map characterization at different rejection temperatures, and 5) endurance evaluation. The cryocooler description, some lessons learned, and the preliminary characterization results were presented in a previous paper.1 Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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CHARACTERIZATION The characterization experiments done at AFRL provide an objective evaluation of the cryocooler’s mechanical and thermodynamic performance envelope as well as experimental data for the validation of empirical models developed at AFRL to predict cryocooler performance at conditions for which no experimental data exists. Table 1 shows the cooler’s normal operating conditions. The cooler’s nominal stroke length was originally 9 mm on all four compressors. However, one of the compressors was damaged during initial testing at Astrium, and as a result cannot be run beyond 8 mm without encountering
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noise problems. Due to this limitation, Astrium established the normal stroke lengths as 9.2mm on compressors 1, 2, and 4, and 8 mm on compressor 3. Figure 2 shows a cooldown curve for the Astrium 10 K at two rejection temperatures, 275 K and 290 K, with a fixed 6-mm compressor stroke length. The steady-state values for each rejection temperature are listed in Table 2. During the cool down, the cold tip temperatures (TC1) for both reject temperatures reach steady state relatively quickly. The midstage temperatures (TC3) require at least 5 hours after start-up before they reach steady state. The nominal stroke lengths for the cooler are 9.2, 9.2, 8.0, and 9.2 mm for the four compressors. To reach the nominal stroke lengths the cooler is brought to steady state at 6 mm first. Once steady state is reached then the stroke lengths are increased to 7 mm and the cooler is allowed to reach steady state again. The stroke lengths are then increased to 8 mm and finally to the nominal stroke lengths and steady state is achieved.
BASELINE PERFORMANCE COMPARISON The performance baseline of the cooler is comprised of data collected under nominal conditions with sequentially varied heat loads. Steady state is reached at each value of the load. The heat rejection temperature is held fixed. Curves of cold end temperature and input power versus heat load characterize the cooler. Cooler baseline performance is considered to be the December 2000 data reported previously.1 Four load lines showing cold end temperature versus load are depicted in Figure 3. This data was taken during Dec 2000, June 2001, and Jan 2002. The December and June test results are
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similar, however the values in Jan 2002 show a noticeable drift in performance. The effect is more apparent in Figure 4 for the input power versus heat load. For the same stroke conditions, the input power required to drive the cooler is lower than earlier tests, which could possibly indicate a decrease in charge/fill pressure.
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The midstage temperature is shown in Figure 5. The midstage temperature increased 2 K from Dec 2000 to Jan 2002. This is approximately the same temperature change as seen at the cold end. As pointed out in the discussion of Figure 2 the midstage temperature requires significantly more time to reach steady state. For all of the data shown, the steady state temperature criteria used is ±0.15 K for 15 minutes. After the 0.1 W data point there was an instability with the stroke control on 14 Jan 02. The instability caused a decrease in the total input power (Figure 4) and an increase in the cold end temperature (Figure 3). This decrease in power and increase in temperature caused the normally linear load lines to be offset for the data points between 0.15 and 0.35 W.
OPTIMIZATION Optimization trials investigate the combination of operating parameters that enable optimum cooler performance. The parameters include phase angle, drive frequency, compressor offset, and expander stroke length. In optimization experiments, all controllable operating parameters are kept constant except for one. The one parameter is varied to determine values corresponding to the most cooling at the lowest temperature and input power. An optimization trial was completed at Astrium before delivery to AFRL. AFRL optimization trials will be used to verify the optimum values or identify small changes in performance due to experiment stand differences or other factors.
Phase Angle Variation The sensitivity of the cold end temperature and total input power to phase angle (between the displacer and compressors) is determined by bringing the cooler to steady-state at fixed stroke lengths (9.2, 9.2, 8, 9.2, 3, 3) mm and an operating frequency of 30 Hz. The heat rejection temperature remained set at 290 K. Data points were obtained for discrete heat load points of 0,45,60, and 80 mW. The drive signals to the compressors and displacer were manually adjusted to keep the stroke lengths constant. The cold end temperature and input power values at a constant load of 45 mW are plotted against phase angle in Figure 6. In Figure 7, the percent Carnot COP is plotted against phase angle for the same data points in Figure 6. In order to obtain a minimum in cold end temperature as well as a maximum in %Carnot efficiency, repeated tests were performed to cover range of phase angles from 45 to 69 degrees. The
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data points are not repeatable, especially for the cold end temperature, and the cause of the difference is not known. The input power and specific power minimum values occur at a phase angle of 63 degrees. The results show that the maximum %Carnot efficiency corresponds to the cold end temperature at a phase angle of 55 degrees. The data confirms that the 55-degree phase angle design point will result in the lowest cold end temperature and highest efficiency.
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Frequency Variation To determine the effect of operating frequency on the cooler’s performance, data were collected at five operating frequencies (28-32 Hz). The tests were repeated twice to examine repeatability. Figures 8 and 9 show the effects of frequency on the coldend temperature, input power, %Carnot efficiency, and specific power. Although the data points were not repeatable, the trends in the plots clearly indicate an optimum frequency around 30 Hz from both an input power and cold end temperature standpoint.
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PERFORMANCE MAPPING Characterization load lines, where both the heat rejection temperatures and compressor and displacer stroke lengths are varied from their nominal values, were completed at AFRL. These load lines provide engineers with a map of the cryocooler’s performance capabilities during steady state operation. Figures 10 and 11 show the Astrium 10 K performance maps for reject temperatures of 290 K and 275 K, respectively.
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OBSERVATIONS Degradation in the cold end temperature has occurred since June 01 (Figure 3). The cause of this is under investigation at the current time and nothing concrete has presented itself. Some of the issues that we are investigating are cryopumping (contamination on the cold tip), stiction, and helium leaks. Cryopumping has been occurring in the vacuum chamber and is discussed at length in a previous paper.1 By monitoring the vacuum and cycling the cooler between 10 and 40 K prior to testing, the impact of the cryopumping was minimized in the tests. This process has been practiced on every test that is reported in the report, which may indicate that the gradual degradation is not due to cryopumping. To further investigate this we are planning to add a GM cold finger, operating at 6 K, to the vacuum chamber in order to attract the contaminants away from the Astrium cold head. Additional stiction tests have been conducted since the degradation was noticed. The data have not changed from the results that were reported in July 2001.1 Since the results were similar, contamination within the compressors and expander has been ruled out. Helium leak tests have been performed on the chamber to see if the cooler is losing pressure. The helium detector shows a very small steady leak rate. A single turbomolecular pump is able to keep the vacuum chamber pressure at With this low vacuum level, it is unlikely that the leak is large enough to cause the degradation. At present, the Astrium 10 K cryocooler is running continuously at the normal conditions listed in Table 1. The diagnostic tests, previously mentioned, are being performed to understand the cause of the degradation. Once the diagnostic tests are completed, the cooler will be moved onto tabletop for long-term endurance testing.
ACKNOWLEDGMENTS We thank T. Roberts, J. Kallman, M. Martin and G. Lybarger for their technical assistance.
REFERENCE 1.
Bruninghaus, C. H. Y, J. P. Kallman, B. J. Tomlinson, Jr., E. Myrick, “Performance Characterization of the Astrium 10 K Developmental Cryocooler,” Adv. in Cryogenic Engineering, Vol. 47B, Amer. Institute of Physics, Melville, NY (2002), pp. 1109-1116.
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Technology Development Related to Tactical Cryocoolers at Raytheon Infrared Operations B. A. Ross, M. L. Brest, and F. I. Mirbod Raytheon Infrared Operations Goleta, CA 93117-3090
ABSTRACT Two technology areas have been investigated recently at Raytheon Infrared Operations (RIO). First, we have developed a small cryocooler with good performance at high operating temperatures. The design approach for this cryocooler will be reviewed, and the measured performance data at the integrated dewar/cooler assembly level will be presented, with emphasis on performance at high operating temperatures. The first production run of over 100 cryocoolers was recently completed. Second, the design of a long-life compressor for tactical cryocoolers was completed. A flexure bearing suspension system is used in order to minimize contact between the pistons and their cylinders. The flexure design is challenged by severe constraints in terms of package size and overall compressor mass. The design approach and the resulting configuration is described.
INTRODUCTION Tactical rotary cryocoolers have been in production at Hughes/Raytheon since 1978. Production of linear cryocoolers began in 1994. More than 50,000 tactical coolers have been delivered, and the current production capacity is over 300 units per month At RIO the technology development activities are guided by several objectives. First, we are refining our products to incorporate a consistent design approach using standard production processes. This has driven down cost and improved hardware quality. Since major design changes on a single product are difficult due to cost and product qualification constraints, we have adopted a continuous improvement methodology. This is done by using funding on active products to make incremental design and process improvements. Then, the next active product incorporates and builds upon the previous improvements. This results in “leap frog” type improvements that are faster and more efficient than the usual evolution of a single product. Second, we are working to achieve good performance and reliability at high operating temperatures. As infrared and other cooled devices become more compact and find use in more extreme environments, the desire for good performance and reliability at high operating temperatures increases. Third, we are developing technology for low-cost tactical cryocoolers with operating lifetimes exceeding 20,000 hours. Applications are emerging that require the cooled sensor to
Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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operate for thousands of hours per year. Implementation of cryocoolers with extended lifetimes greatly reduces life-cycle costs and as well as the logistics efforts needed for sensor replacement. The fourth objective is closely related to the third objective. We need to continue to share technology and production methods within Raytheon, between RIO tactical cryocooler operations and El Segundo space cryocooler operations, in order to develop an effective long-life tactical cryocooler. In fact, we believe that there is potential for a long-life tactical cryocooler to find use as a low-cost space cryocooler with only minor modifications. In addition, we are investigating the use of pulse tube expanders in tactical cryocoolers. Staff from El Segundo have contributed significantly to the space cryocooler technology base, as described in recent ICC papers (Price1,3, Kirkconnell2,4). In the midst of these technology objectives, our prime concern is to meet the needs of our customers. We continue to find that the interfaces with the cryocooler are critical to cooling performance of the system (Ross5). Accordingly, we continue to pay strict attention to the power input characteristics, heat sinking of the compressor and expander, and thermal interface between the detector and cold finger.
RECENT PRODUCTION EXPERIENCE Results of a recent production run indicate some positive benefits of the evolution of cryocooler design at RIO. A low-rate initial production lot of 7052-196S cryocoolers with a quantity of 150 was produced between December 2001 and March 2002. The actual production cost was below the cost target, partly because of the 98.4% first-time yield that led to a low amount of rework. The 7052-196S cryocooler has a state-of-the-art production design that is intended to keep manufacturing costs low while meeting performance and operability requirements. The configuration is somewhat unique, as the compressor is large, the expander is small because of envelope constraints, and the transfer tube is long. The compressor pressure vessel is made from stainless-steel, allowing the closure weld to be made in-house. The assembly of the components is straight-forward, with few matched parts. A number of assembly processes were upgraded to reduce cycle time. Tooling was upgraded, particularly in the area of displacer assembly. In some cases the processes were improved so that subsequent traditional production steps could be eliminated. Implementation of a state-of-the-art automated test minimized the touch labor required for final acceptance test. A very important part of the production readiness process was to identify 13 critical suppliers, and work with them to reduce material defects and cost. We have eliminated features that needlessly added cost, and have identified critical features that directly affect process yields in the RIO factory. The 7050 compressor is similar to the 7052 compressor; it has a center-exit transfer tube while the 7052 compressor has an end-exit transfer tube. Based on the success of the 7052 compressor, the 7050 compressor was recently redesigned. By using lessons learned on the 7052 and 7062 compressors, major design improvements were made with little to no technical risk. The compressor housing was redesigned to utilize the same stainless-steel material. This allowed the weld to be performed in-house, reducing cycle time and eliminating a painting
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operation. The indium union seal at the compressor was replaced with the conical type seal used on the 7062 compressor. This seal greatly increases the reliability of the compressor. Heat sizing of pistons was implemented to minimize debris generation. Aside from improving reliability, this allowed us to eliminate a considerable amount of touch labor and cycle time. Through concurrent engineering with manufacturing, design improvements and lean manufacturing have led to a dramatic cost reduction for the 7050-260L product line. Currently, prototype compressors have been built and tested showing excellent performance. A production run of the upgraded 7050 compressor is scheduled for the second half of 2002.
7062-260S CRYOCOOLER DEVELOPMENT RIO developed a new IR sensor engine in 2001, with the objective of good performance at high operating temperatures in an extremely compact package. The needed performance was similar to that achieved by the existing 7060-260S and 7061-260S cryocoolers at RIO, but two avenues of change were followed. First, the compressor geometry had to be made more compact (1.32 inch outer diameter, 4.5 inch length). Second, the compressor design approach needed to more closely follow the approach used on the 7050 and 7052 compressors described above. The three small tactical cryocoolers are shown in Figure 2. The design was changed to reduce manufacturing costs. The magnetic configuration of the linear motor was changed, as was the core material. The parts count in the displacer was reduced while a bond joint was eliminated. Displacer alignment tooling was upgraded to improve the consistency of the final product. Changing the compressor pressure vessel to be like the 7052 pressure vessel allowed the seal-welds of the compressor to be done in-house and eliminated a painting process. The design was upgraded early in 2001, and prototypes were refined in the July to October 2001 time frame. Production of the first lot of over 100 units took place from December 2001 through May 2002. Cooling performance for 20 integrated dewar/cooler assemblies is shown in Figure 3. The performance parameters of interest are cooldown time at soak temperatures of 27 and 71°C, and steady-state input power when the compressor skin is at 23 and 95°C. The lines of Figure 3 represent the specified maximum values, and the data points show the means and standard deviations of the 20 assemblies. The performance margin is substantial.
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7080-514 CRYOCOOLER DESIGN The objective of this design activity is to develop a long-life (>20,000 hours) tactical cryocooler, with a focus on implementing a flexure bearing suspension system in the compressor. The standard RIO -514 expander was used for analysis purposes, but other expanders are expected to be used with the compressor, including the –388 expander and a pulse tube expander. The pumping capacity of the compressor was designed to be approximately 30% greater than the 7052 compressor described above. Tactical applications for larger capacity cryocoolers are the initial focus, and the needs of multiple programs expected to be met with a single configuration The design follows traditional tactical compressor designs where possible. The pressure vessel is essentially scaled up from the 7052 compressor. The processes associated with the piston/cylinder wear couple are the same, as is the linear motor configuration. The main difference between the 7080 compressor and other compressors produced by RIO is the implementation of a flexure bearing support system for the pistons. A maximum allowable stress state for the flexure material was established to ensure long life without fatigue failures. The flexure bearings are made as large in diameter as possible within the 2.375 inch compressor outer diameter constraint, in order to minimize the peak stress state. The flexures are designed so that the pistons do not contact their cylinders unless side loads are very high (which is not expected for much of the operating lifetime). By eliminating friction and debris generation, operating lifetime will be greatly extended compared to the non-supported piston assemblies currently in use. The characteristics of the flexures are favorable when the flexure material is thin, but there is a trade-off with a practical number of flexures. The axial spring rate of the flexure stacks, two per piston, must meet the dynamic tuning needs of the compressor, as defined by the SAGE analysis code when simulating the 7080-514 cryocooler. The flexure analysis was based on non-linear finite element methods. A simple beam flexure configuration that could be analyzed by more simple linear methods (Marquardt6) was considered, but the design requirements could not be met with that configuration. The flexure design parameters that were considered in the analysis included: Outer diameter (active and actual) Inner diameter (active and actual) Flexure thickness Flexure material properties Number of arms and their geometry
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Implementation of flexure bearings requires a precision alignment operation that is not required for our other tactical cryocoolers. RIO has relied on the Raytheon space cryocooler staff in El Segundo, California for guidance on achieving the alignment required for tactical applications. The linear motor is similar in configuration to the one used in the 7052 compressor, but the available geometric envelope was quite different. The compressor diameter was allowed to grow from 1.94 inches to 2.375 inches, allowing the linear motor to grow accordingly. The motor needed to be as short as possible, in order to keep the overall compressor length and mass down. In addition, the motor needed to provide a higher force over a shorter stroke in order to accommodate the flexure design. Finite element methods were used to evaluate the magnetic circuit and the forces produced by the motor as a function of position of the moving coil. The design parameters included outer and inner diameters, amplitude and frequency, magnetics configuration, core geometry, magnet volume and coil volume. Once all of the design trade-offs were considered, a design was developed that was viable in terms of expected performance, flexure stress state, and linear motor configuration. The overall compressor diameter is 2.375 inches, the length is less than 5 inches, and the estimated mass is 3.7 pounds. There is potential to reduce mass further, but this was beyond the scope of the initial design activity. The compressor is expected to be suitable for SADA-I, SADA-II, and other applications, and will be the largest-capacity cryocooler in the standard product line at RIO. The production cost of this long-life compressor is expected to be slightly higher than existing compressors, due to the cost of the flexure bearing components and the alignment required compared to the existing machined springs. However, we feel that this is an enabling technology that will lead to greatly-reduced life-cycle costs, and thus will prove to be attractive to our customers.
CONCLUDING REMARKS Tactical cryocooler technology and the production that results continues to evolve at RIO. We are still making rotary cryocoolers, but that is expected to continue for only the next few years. RIO is consolidating its product line with the intent of meeting the needs of many customers with a low number of cryocooler models. The RIO cryocooler factory is encouraged with the results of the design upgrade and process improvements associated with the 7052-196S cryocooler. Because of these changes, the cost
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target was met for the first full production run in early 2002, with a first-time yield of 98.4%. RIO has applied similar changes to the 7050 (transfer tube center exit) compressor. Prototype hardware was successfully operated and full production is scheduled for later in 2002. The 7062 compressor was developed based on the 7060 compressor, and it also incorporates the improvements first demonstrated by the 7052 compressor. The 7062-260S cryocooler has demonstrated good performance at high operating temperatures, and is significantly less expensive to build than the 7060-260S cryocooler. Moving the product line toward a consistent design and process approach continues to drive down costs. An advanced design for a long-life tactical compressor that will allow an operating lifetime exceeding 20,000 hours was developed in 2001. The compressor is expected to find use in a variety of tactical applications, with a variant targeted for a low-cost space cryocooler application. The working relationship within Raytheon between the RIO tactical cryocooler staff and the El Segundo space cryocooler staff is a critical component of the development of this long-life cryocooler.
ACKNOWLEDGMENT Mike Jacoby provided invaluable assistance with the non-linear flexure analysis of the flexure bearing designs. Dennis Ebejer assisted with the prototyping and testing of the 7062260S cryocooler. Bobby Thompson and his factory team are responsible for the production accomplishments described in this paper. Ray Laithrup of Servomagnetics Inc. provided the analysis of the linear motor design.
REFERENCES 1. Price, K., Reilly, J., Abhyankar, N., and Tomlinson, B., “Protoflight Spacecraft Cryocooler Performance Results” Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2001), pp. 35-43.
2.
Kirkconnell, C. S and Price, K. D., “Thermodynamic Optimization of Multi-Stage Cryocoolers” Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2001), pp. 69-78.
3.
Price, K., and Urbancek, V., “95 K High Efficiency Cryocooler Program” Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2001), pp. 183-188.
4.
Kirkconnell, C. S., Price, K. D., Barr, M. C., and Russo, J. T., “A Novel Multi-Stage Expander Concept” Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2001), pp. 259-263.
5.
Ross, B. A. and Black, S. H., “Advances in High-Performance Cryocoolers and Production Variants at Raytheon Infrared Operations” SPIE Infrared Technology and Applications XXVII, Vol. 4369 (2001), pp. 427-435.
6.
Marquardt, E. and Radebaugh, R., “Design Optimization of Linear-Arm Flexure Bearings” Cryocoolers 8, Kluwer Academic/Plenum Publishers, New York (1995), pp. 293-304.
Performance and Reliability Data for a Production Free Piston Stirling Cryocooler M. Hanes and A. O’baid Superconductor Technologies Incorporated Santa Barbara, CA 93111
ABSTRACT The use of a free piston, gas-bearing Stirling cycle cryocooler for commercial high temperature superconducting (HTS) RF filters for the wireless communication industry dictates that the cooler must not only be low cost, but also have long life and high reliability. Performance and reliability data are presented for the cryocooler that is currently in production. About 1260 of these cryocoolers have been deployed and, up to this point in time, have accumulated 11.2 million hours of run time. This cryocooler has very high efficiency (6W lift at 77K, 100W input power, 35°C heat reject temperature) and demonstrated long life. The goal for the life of this cooler was 40,000 hours mean time between failure (MTBF). Actual field reliability data is presented confirming that this goal has been met and is indeed surpassed. There is also data showing the performance characteristics of this cooler as well as data from ongoing, in house life tests. INTRODUCTION The Superconductor Technologies Inc. (STI) cryocooler is used for cooling HTS RF filters to 77K for use in the wireless communication industry. The performance and physical characteristics of this cooler are listed below: Performance: 6 watts at 77K with 100 watts input power and 35°C heat reject temperature Operating temperature range: -40°C to 60°C Input power: 60Hz, 140W max. Weight: 7.5 lbs Dimensions: 3.5 in. OD x 11.5 in. length Operating orientation: Any Vibration output with passive balancer: 2.2 N The STI cryocooler is used in conjunction with a dewar which is also produced at STI. The maximum heat load of the dewar and the performance of the cryocooler are shown in Figure 1. Most systems are operating at about 23°C ambient temperature and Figure 1 shows that the cryocooler has more than adequate margin of heat lift versus the dewar heat load. The heat lift of the cooler as a function of input power and at various heat reject temperatures is shown in Figure 2.
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RELIABILITY DATA FROM THE FIELD To date, STI has over 1260 systems operating in the field that utilize the present cooler; they have a combined run time of over 11.2 million hours. Of these, there are 100 coolers with more than 18,500 hours of run time. Figure 3 shows the accumulated run time and number of coolers in the field as a function of date.
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There have been 14 random cooler failures, which results in an MTBF of approximately 800,000 hours. This number is calculated from Eq. (1): MTBF = Cumulative run time hours / Number of failures
(1)
The use of Eq.l for MTBF calculation assumes that the failure modes of the cryocooler are occurring randomly, there is no infant mortality and no recurring wear out mechanism detected. Figure 4 shows the failures of cryocoolers in the field and the effect of those failures on the MTBF. As can be seen from the graph, the MTBF has been relatively constant for the past year, which is a good indication that this number is the actual value for this cooler.
IN HOUSE LIFE TEST DATA STI has an extensive in house life test program with forty five coolers running continuously. As shown in Fig. 5, twenty five of these coolers are running “on the bench” with the cold end insulated with a foam dewar and a constant input power. Periodically, these coolers are removed
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from life test and performance tested to determine if there has been any degradation in performance, or a failure. The other twenty coolers are running with a vacuum dewar and a heat load applied to the cold end. These coolers have the heat reject temperature, input power, heat load, and cold end temperature data taken once per hour on a continuous basis. This set up allows any variation in cooler performance to become immediately apparent. Some of these coolers are periodically turned off to simulate more closely the actual field conditions. The hours of run time on each cooler varies from 9,000 to over 23,000 and the combined run time on all 45 coolers is ~ 670,000 hours with one failure to date. Figure 5 shows the monitored life test set up and Figure 6 shows the adjusted cold end temperature data from 10 of the coolers on monitored life test. This data is adjusted to compensate for any fluctuations in input power (due to supply voltage variations) and heat reject temperature (due to room ambient variations).
SUMMARY STI had set a goal 5 years ago of producing a cooler which has a minimum 40,000 hour life. The data presented here demonstrates this goal has been met, and, surpassed. This conclusion is based on field data from 1260 coolers with a combined run time of 11.2 million hours and in house life test data from 45 coolers with a combined run time of 670,000 hours.
CMCEC Life Test Results and Related Issues S.W.K. Yuan, D.T. Kuo, and T.D. Lody Cryocooler Group CMC Electronics Sylmar, CA 91342
ABSTRACT CMC Electronics, Cincinnati (CMCEC) is conducting life tests on all models of our linear Stirling coolers and has demonstrated a life of ~10,000 hours on each model. In this paper, an update of this continued effort is documented. In particular, the life test of the half watt cooler (Model B512B) and the 5 watt cooler (B5000D) are discussed in detail. The B512B cooler has demonstrated a 20,000 hour life MTTF and 40,000 hour MTBF. Other life related issues like contamination and leak rate are also discussed.
INTRODUCTION With the advances in technology, the life of tactical cryocoolers has extended way beyond the initial goal of 4,000 hours. Recently, a significant collection of empirical data has provided credible support to life prediction models1,2 based on mechanisms of wear and contamination control.3,4 Table 1 shows the updated life test results of CMCEC’s linear coolers. In this paper, the life test results of the half watt model (B512B) and the 5 watt model (B5000D) are presented.
THE B512B COOLER The Life Test The B512B cooler is a half watt linear Stirling cooler.5, 6,7 The original life test was performed for our customer, under the Second Generation Dewar Assembly (SGDA) program. Four coolers were picked from the production line and put on life test. The coolers have been operating continuously on a bench at room temperature, with acceptance tests performed at scheduled intervals. After successfully meeting the specification of 4,000 hours, CMC decided to continue the life test
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under an IRAD program, to demonstrate life beyond the original requirement. The life test exit criteria are listed in Table 2. Figures 1 to 3 depict the degradation of the cooler performance as a function of operation time. One of the units (S/N 013) was found to exhibit a much faster rate of degradation compared to the rest of the units, and it failed the performance test at around 10,000 hours. A failure analysis performed on this unit found heavy rubbing along the length of the displacer and that the coldfinger
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cylinder was not straight. After correcting the problem, the performance was restored (see data points of S/N 013 at 10,000 hours), indicating that the failure was not related to wear of the compressor. Assuming a definition of MTTF8 that is based on compressor wearout failure, the performance of S/N 013 does not factor in the MTTF calculation. The three remaining coolers exceed 20,000 hours MTTF life. Contamination Issue. Unit S/N 014 failed at 20,500 hours due to build up of gaseous contaminants. A gas chromatography analysis showed that the methane level exceeded the allowable value (100 ppm) in the cooler, consistent with early findings.3,4 After purge and fill, the performance of the cooler returned to the value consistent with the wearout characteristic (as represented by the slope of the S/N 014 curves in Figure 1 to 3). Values of some of the noteworthy contaminants found in the cooler are summarized in Table 3.
THE B5000D COOLER The Life Test The B5000D cooler is a 5 watt linear Stirling cooler.9 The original life test was partially funded and motivated by a customer program called ABL, under the Airborne Laser (ABL) program. One cooler was picked from the production line and put on life test. The cooler has been operating continuously on a bench at room temperature, with acceptance tests performed at scheduled intervals. After successfully meeting the specification of 2,600 hours, CMC decided to continue the life test under IRAD funding to demonstrate life beyond the initial requirement. The exit criteria of the ABL program are listed in Table 4. Figures 4 to 6 show the degradation of the cooler performance as a function of operation time. A total of 9,100 hours have been recorded to date.
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COOLER LIFE AND LEAK RATE As a rule of thumb, the acceptable leak rate for most of the tactical coolers is std cc/sec for a 10 year shelf life. With the extension of cooler life and shell life requirements, it is necessary to revisit this number to ensure that it provides sufficient pressure for the normal operation of the cooler.
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From the conservation of mass, with a constant volume, one can calculate the leak rate as
where V is the total internal volume of the cooler and is the density of the working gas. Expressing the leak rate in terms of the volumetric flow rate, one gets
Integrating both sides of the equation and substituting the ideal gas law, one arrives at the following equation
On the other hand, a simple calculation based on the conversion of the pressurized working gas into std cc and subtracting the leakage yields the following linear equation
Equations 3 and 4 are plotted in Figure 7 for a leak rate of
std cc/sec.
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Within the first 20 years, the linear equation appears to be an adequate tool in predicting the pressure drop in a cooler subject to a leak rate. The leak rate of std cc/sec is compared to that of std cc/sec in Figure 8. At a glance, the latter leak rate, which allows less than 20% pressure drop in 100 years, appears to be adequate. However, due to the extension of cooler life and the stringent requirements of customers as far as the storage time and condition are concerned, one must reexamine the tolerable leak rate carefully. Figure 9 shows the pressure drop in coolers of various volumes. One sees that for the same leak rate, the impact is more pronounced in a smaller cooler (B512C). The effect of long storage time at elevated temperatures must also be evaluated by Eq. 3 to ensure that the anticipated life is reached.
CONCLUSIONS The life test results of the CMC half watt and 5 watt coolers have been presented. All of the CMC coolers have demonstrated a life of ~10,000 hours (Table 1), with the B512B cooler exceeding 20,000 hours. CMC is planning to re-run the life test of both the B602 and B1000 coolers with the lessons learned from the B512B model, with the goal of achieving at least a 20,000 hour life.
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The issue of leak rate on cooler life has also been addressed. For small coolers and/or stringent storage requirements, the adequacy of leak rate must be revisited.
ACKNOWLEDGMENT The authors would like to thank our customers for their support during the early phases of the life test.
REFERENCES 1.
Kuo, D.T., Loc, A.S. and Yuan, S.W.K., “Cryocooler Life Estimate and its Correlation with Experimental Data,” Advances in Cryogenic Engineering, Plenum Publishers, New York (1999), vol. 45A, p. 267.
2.
Kuo, D.T., Lody, T.D. and Yuan, S. W.K., “BAE’s Life Test Result on Various Linear Coolers and their Correlation with a First Order Life Estimation Model,” Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2000), pp. 665-672.
3.
Yuan, S.W.K., Kuo, D.T. and Loc, A., “Cryocooler Contamination Study,” Advances in Cryogenic Engineering, vol. 45A, Plenum Publishers, New York (1999), p. 275.
4.
Yuan, S.W.K., Kuo, D.T. and Loc, A., “Cryocooler Contamination Study: Temperature Dependence of Outgassing,” Cryocoolers 12, Kluwer Academic/Plenum Publishers, New York (2000), pp. 659-664.
5.
Kuo, D.T., Loc, A.S., and Yuan, S.W.K. “Experimental and Predicted Performance of the BEI MiniLinear Cooler,” Cryocoolers 9, Plenum Publishers, New York (1996), pp. 119-125.
6.
Yuan, S.W.K., Kuo, D.T., and A.S. Loc, “Enhanced Performance of the BEI 0.5 Watt Mini-Linear Stirling Cooler,” Advances in Cryogenic Engineering, vol. 43B, Plenum Publishers, New York (1997), p. 1847.
7.
Kuo, D.T., Loc, A.S., and Yuan, S.W.K. “Qualification of the BEI B512 Cooler, Part 1- Environment Tests,” Cryocoolers 10, Plenum Publishers, New York (1998), pp 105.
8.
Pruitt, G.R., “Reliability Growth of Coolers for Advanced Optical Systems and Instruments,” Cryogenic Optical Systems and Instruments IV, SPIE vol. 1340 (1990), pp. 311-324.
9.
Yuan, S.W.K., Kuo, D.T. and Lody, T.D., “Qualification and Test Results of a 5 Watt Commercial Stirling Cryocooler,” Advances in Cryogenic Engineering, vol. 47A, Academic Institute of Physics, Melville, New York (2001), pp. 654-661.
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MTTF Prediction in Design Phase on Thales Cryogenics Integral Coolers J.M. Cauquila and J.Y. Martina P. Bruinsb and T. Benschopb a b
Thales cryogénie SA, Blagnac, France Thales cryogenics BV, Eindhoven, The Netherlands
ABSTRACT A continuous improvement of Stirling integral coolers MTTF is necessary to answer the new market requirements. The time spent in the design phase is critical to get the product ready at the right time and to reduce the development costs. In order to reduce the development risks (be sure to get in the new design at a higher MTTF), we have developed a calculation method for prediction of the expected MTTF of our coolers. Early in the design phase, we use the method of predictive calculations on MTTF. For that, we need MTTF data for each critical single part or function in the cooler. In order to fill the lack of data available on failure rates for mechanical parts (tightness, coating...) in our particular application, we built a list of the different functions inside the cooler from the failure point of view. The data collected from the extensive lifetest results already performed allow us to determine the MTTF for each function. These elements are introduced into the predictive calculation of the new design. The expected MTTF of the new cooler is then available. The results obtained by calculation are still only indicative. These results have to be verified as quickly as possible. To reduce the duration of lifetests, we apply an accelerated test procedure developed and verified over 2 years. This method allows us to obtain results on MTTF equivalent to several thousand hours using only a test duration of a few weeks. This method has been validated on previous cooler models. This MTTF calculation is today applied to all designs of our integral coolers, and in particular, to our new design of the RM2 cooler under development at Thales Cryogenics. This method will be discussed and the results of the calculations performed on the new RM2 design will be presented.
INTRODUCTION Among the performance metrics of a cryogenic cooler, the MTTF is one of the most important. The market requirement for MTTF has been strongly increasing for several years. A continuous improvement of our Stirling integral coolers is essential to remain competitive. Up to now, the MTTF was determined by lifetests conducted on the coolers. The standard test profile applied during the lifetest is representative of operation in a typical application with alternating stops (periods of storage), starts, and running at room and hot temperatures. The duration of such a lifetest is very long.
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In addition, the customer is requiring data on the MTTF of the cooler very early in his system design process and is also interested in having information regarding the MTTF of the cooler in his specific application. Because of the duration of such a standard lifetest (several months), it is generally not possible to wait until the standard lifetest is finished before providing the expected MTTF data to our customers. For costs reasons, it is not possible for us to run lifetests representative of each customer’s application. At the same time, it is essential for us to have a good estimation of the expected MTTF early in the design process in order to reduce the duration of the design phase and to reduce the development risks (design in accordance with the targeted MTTF, verify the design by calculation followed by actual tests). In that context, we have worked on a methodology to allow the calculation of the expected MTTF of our coolers. Lifetests performed after the design are then used as a confirmation of these calculations. Nevertheless, the actual test remains necessary in order to increase our knowledge on the coolers and validate our calculations. In parallel, and in order to reduce the duration of the testing, we developed an accelerated lifetest protocol. This method for calculating the expected MTTF of the coolers will be presented, along with results of the calculations for the RM2 cooler. Two models of the cooler will be addressed, both its configuration as produced in serial production today, and a new design of the RM2 cooler at this point in its development. The Thales cryogenics RM2 cooler, shown in Fig. 1, is a light weight, high efficiency integral Stirling cooler. It provides a total cooling power of 400 mW at 77K and 23°C ambient temperature at 9 Wac input power. Its maximum mass is 275g. This cooler has been produced in a total quantity of over 4000 units in the past years. An upgraded version of this cooler is under development in our facilities. The serial production of this new version is planned to start in the second half of 2002.
PRESENTATION OF MTTF CALCULATION METHODOLOGY General presentation At an early stage in the design process, we identify the parts or functions in a new cooler that are critical from the MTTF point of view. This identification is made through a FMECA (Failure Mode Effect Critical Analysis) based on our past experience with cryogenic coolers. For each part or function, we define the design criteria such as the volume allocated, the stresses applied, etc. According to this, the type of technological solution is chosen for the part or function and its design is made. From the lifetest database we have, or from calculations made on available data from the supplier, we determine the MTTF of each of these elements (or functions). The MTTF of the future machine can then be calculated.
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The actual lifetest is still necessary in order to verify the truthfulness of the calculations and to collect additional data to improve our calculational process.
FMECA The FMECA process relies heavily on our past experience with cryogenic coolers. In order to be as thorough as possible and to remain as close to measured facts as possible, we base our FMECA on the content of our repair database and the results of previous lifetests of similar coolers that are applicable to the new cooler. This allows us, for example, to weight more precisely the criticality of each type of failure mode. The results of lifetests, even if they are restricted to a small number of units, are very interesting in this phase, because they are linked to a perfectly known and mastered profile of use. The output of the FMECA is a list of the identified critical parts or functions in the cooler from the MTTF point-of-view.
Preliminary Design During the preliminary design, on the basis of the required cooler performances (cryogenic performances, volume, weight, MTTF, ...), the design parameters are defined for each part or function. The technologies are chosen according to these parameters and an iterative process allows the optimization of the overall system. The output from the FMECA is taken as input for the design. The design parameters allocated to the critical parts are minimized and the design of these parts takes into consideration an important margin.
Choice of Reliability Distribution For components. When appropriate, we choose the Weibull law1 for the calculation of the MTTF on the components of our coolers. This distribution is well adapted to coolers parts, which are mainly mechanical parts submitted to wear. The failure rate of these components is not constant, so the reliability R(t) at any time t follows the Weibull equation:
where
= scale parameter = shape parameter, characterises the failure mode = life characteristic (set at zero) The use of the Weibull equation requires that one know the three parameters that characterize the component. These parameters are determined by calculation based on previous lifetests conducted on similar components or by finding data from the component supplier (for example, for bearings). Unfortunately, it is often difficult to obtain these parameters for all parts used in our system. For the parts submitted to wear and when we have no access to the parameters of the Weibull law, we take the assumption that the failure rate is constant (random failure). In that case, the test results of the lifetests already performed on similar parts in similar environments are considered. From these results, we determine the MTTF of the part or function with ber of failures. When no failure are observed, we take We also consider the parts having a constant failure rate (random failure) as shown by our experience (seals, electrical components for example). The MTTF calculation is identical as for the parts mentioned above. For cooler assembly. For the calculation of the MTTF of the complete cooler, we assume that the failure rate is constant (random failure). This simplification can be justified by the following considerations:
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All the different components inside the cooler do not follow the same distribution. As a consequence, the system can behave in an unpredictable way (random failure) with a constant failure rate. This is also what experience shows. We also note that the number of failures according to a Weibull distribution is the same as that according to an exponential distribution, if you consider the average number of failures over an appropriate period of time. Therefore, we calculate the MTTF of the cooler as with being the failure rate of each component, and i being summed over the total number of components in the cooler for which a is defined.
Verification The methodology has been verified by applying it to coolers already existing for which a database regarding lifetests is available. The results obtained from the calculations, as mentioned before, are on the same order as the results from the lifetests. However, our approach is to revalidate the calculations with actual lifetests on each cooler under development. The role of the calculations is to help verify the design before going into the prototype fabrication phase and to allow us to give preliminary MTTF figures to customers based on changes made in some components of the design. The standard test profile applied to the cooler during the lifetest is representative for the operation in a typical application. The duration of such lifetests is very long (several months). In order to reduce the duration of the lifetests, we developed an accelerated lifetest protocol with the following guidelines: The accelerated lifetest has to be more severe than the conventional one in order to significantly reduce the duration of the tests. The test must stay representative of the normal failure modes registered in the conventional lifetest. It must not be too severe in order to not exceed the normal capacity of the cooler design and generate failure modes not applicable in the real operational mode. Easy to set, with limited tooling in order to reduce intervention delay and costs. This accelerated lifetest has been implemented and is being verified. It allows us to reduce the duration of the test by a factor of four.
RESULTS OF PREDICTED MTTF CALCULATIONS ON RM2/01 COOLER AND COMPARISON WITH LIFETEST RESULTS This section presents the results of the predicted MTTF calculations conducted for the current production version of the RM2 cooler; this cooler is referred to as the RM2/01. The MTTF of this RM2/01 cooler has also been measured in lifetests, and this value is compared with the calculated value.
Predicted MTTF Calculation for RM2/01 Cooler FMECA. The failure analysis carried out on this cooler has defined the following failure modes as critical from the MTTF point-of-view: bearing failures, helium leakage, electrical defects, internal outgassing, and coating wear. Life time tests database. This database is constituted with the results of lifetests conducted on 42 coolers. Among these 42 coolers, 27 coolers are RM2 coolers, and the 15 remaining are RM5 coolers (rotary integral Stirling coolers, more powerful than the RM2 cooler, but with similar architecture and technologies). This database was used to determine the MTTF of some of the components. Among the 42 coolers, some of the coolers have been tested according to our standard lifetest simulating normal intended use of the product, and the others have been tested according to our accelerated lifetest. MTTF calculations for the bearings. The bearings are components produced in huge quantities. The value (shape parameter) for bearings is given by the supplier
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According to the loads applied on the bearings in the application for the normal operating conditions, the supplier is able to define the L10 value. This value indicates the running duration when 10% of the bearings have failed. This corresponds in the Weibull law to R(L10) = 0.9. This allows the calculation of the value (scale parameter) of the Weibull law. Knowing the Weibull parameters for each bearing, it is possible to calculate their MTTF at 63% according to the Weibull law (MTTF at ). There are five bearings in the cooler. Table 1 presents the L10 figures and associated MTTF values for these five bearings as used in the RM2/01 cooler. MTTF calculation for Helium seals (Helium tightness function). This MTTF is calculated from the failure data available in our lifetest database. Among the 42 lifetests conducted, no helium leakage failures have been observed. Nevertheless, this function is identified as being critical. We chose to take as the MTTF the total of all the running hours measured during the lifetests (71000 hours). This is a conservative approach. The real MTTF for this function is above this. MTTF calculation for electrical components. The assumption made for these components is the same as that for the helium seals. No failures were seen for these components among the 42 lifetests performed. The MTTF for this function with the conservative approach is then 71000 hours. MTTF calculation for outgassing. The approach used for the outgassing phenomena is the same than for the helium seals (see above). Among the 42 lifetests, 2 failures were observed and identified as being caused by internal outgassing. The calculated MTTF for this function is 35500 hours. MTTF calculation for coating wear. The approach used for coating wear is the same as that for helium seals (see above). No failures were observed for coatings among the 42 lifetests performed. Nevertheless, coating wear exists and can generate failures. The MTTF for this function with the conservative approach is then 71000 hours. MTTF calculation for complete RM2/01 cooler. Other parts or functions are not taken into account, as their failure rate is insignificant according to our experience. Table 2 summarizes the overall MTTF prediction for the complete RM2/01 cooler; as noted, the predicted MTTF is 3170 hours.
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Verified RM2/01 MTTF Based on Lifetest Results Measured RM2/01 MTTF in standard lifetest. Five integral cryocoolers RM2/01 have been tested in our standard lifetest. These coolers were taken as samples out of our serial production over a 2-year period of production. Table 3 reports the running hours of the coolers until the first failure occured (according to the failure criteria described above). To compute the associated MTTF we appeal to an iterative graphical method based on the use of Allan Plait paper as shown in Fig. 2. This leads to the indicated classes with their associated failure rates as shown in the accompaning table. For the RM2/01 cooler, the MTTF at 63% for the standard lifetest is 4900 hours. Measured RM2/01 MTTF in accelerated lifetest. Twenty two integral cryocoolers RM2/01 have been tested in our accelerated lifetest. These coolers were taken as samples out of our serial production over a 2-year period of production. During the accelerated lifetest, the coolers were driven at high speed (close to maximum speed) continuously at ambient laboratory temperature.
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Table 4 reports the running hours of the coolers until the first failure. Using the same calculation as displayed in Fig. 2 for the standard lifetest, we determined the MTTF at 63% for the RM2/01 cooler in the accelerated lifetest as being 1100 hours.
Conclusion The predicted MTTF of 3200 h is between the measured MTTF in the accelerated lifetest (1100h) and the MTTF measured in the standard lifetest simulating normal intended use of the product (4900 h). We can see that the predictive method of calculating the MTTF is conservative. This is a good approach, as a lot of assumptions are taken into account in the calculation. This also allows the calculation of the predicted MTTF in a real customer application according to the specific stresses induced by the customer application.
RESULTS OF PREDICTED MTTF CALCULATION ON THE NEW RM2/02 DESIGN AND COMPARISON WITH LIFETEST RESULTS This section presents the results of MTTF predictions being made on a new upgraded version of the RM2 cooler. This upgrade of the cooler is intended to significantly increase its MTTF. This cooler is referred to as the RM2/02. The MTTF of this RM2/02 cooler, as measured by ongoing lifetests, is presented and compared with calculated predictions.
Predicted MTTF Calculation for the RM2/02 Cooler FMECA. The failure analysis carried out on this new cooler has defined the same failure modes as for the previous version as critical from the MTTF point-of-view: bearing failures, helium leakage, electrical defects, internal outgassing, and coating wear. MTTF calculations for the bearings. The bearings of the new version have been fully redesigned as well as their integration into the cooler. According to the new loads applied on the bearings in the application, the supplier has defined the L1 0 value under specified operating conditions. There are still five bearings in the cooler. Table 5 presents the L1 0 figures and the associated MTTF for these five bearings as used in the RM2/02 cooler. The main improvement focuses on bearing No. 1. MTTF calculation for the other critical parts. No major changes were made to these parts. We assume that the MTTF for these parts remains the same as for the previous version. MTTF calculation for complete RM2/02 cooler. As before, the other parts or functions are not taken into account, since, in our experience, their failure rate is insignificant. Table 6 summarizes the overall MTTF prediction for the complete RM2/02 cooler; as noted, the predicted MTTF is 9870 hours. From the calculation, it appears that the new design has a much higher MTTF than the previous version.
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Measured RM2/02 MTTF in Accelerated Lifetest Three first prototypes of the new version RM2/02 cooler are under accelerated lifetest today. Among these three coolers, two are still running as shown in Table 7. A lifetest-based MTTF calculation has been made with the assumption that the remaining coolers failed as of the date of the calculation. As the two coolers are still running, the real MTTF figure will increase in the future. Using the same graphical calculation procedure as for the RM2/01 version, we determined the MTTF at 63% for the RM2/02 cooler as being 2600 hours (minimum) from the accelerated lifetest data.
GENERAL CONCLUSION The predicted MTTF of the new RM2/02 design is significantly higher than the MTTF of the previous RM2/01 design. One means of assessing the expected lifetest MTTF of the new RM2/02 cooler design is to scale its accelerated-test MTTF using the ratio between its accelerated-test MTTF and that of the RM2/01 version (i.e., 2600/1100=2.3) in combination with the correlation factor determined between the accelerated test and standard lifetest (a factor of 4). Using these factors gives the MTTF for the new RM2/02 design as being above 10000 hours for the standard lifetest simulating normal intended use of the product. The new design for the THALES Cryogenics RM2 cooler is already validated as significantly improving the MTTF of the cooler. Complementary testing will be conducted in the near future to confirm this figure and increase the precision of our calculation. The new RM2/02 design is also being qualified with respect to other requirements as well, and will be in serial production by the second half of 2002.
REFERENCE 1. P. Chapouille – R. De Pazzis, F.T., Fiabilité des systèmes, Editions Masson.
An Experimental Study of the Phase Shift between Piston and Displacer in a Stirling Cryocooler S.J. Park, Y.J. Hong, H.B. Kim, D.Y. Koh, B.K. Yu*, and K.B. Lee**
Thermal and Fluid Systems Departments Korea Institute of Machinery & Materials Taejeon 305-600, Korea *Wooyoung, Changdong, Dobong-ku Seoul, 632-39, Korea **Department of Mechanical Engineering Pusan National University Pusan, 609-735, Korea
ABSTRACT Small cryocoolers are being widely applied to the areas of infrared detectors, superconductor filters, satellite communication, and cryopumps. Cryocoolers that utilize the Stirling cycle are characterized by small size, lightweight, low power consumption, and high reliability. For these reasons, FPFD (Free Piston Free Displacer) Stirling cryocoolers are widely used not only in tactical infrared imaging cameras, but also in medical diagnostic equipment. In this study, a Stirling cryocooler actuated by the electric force of a dual linear motor has been designed and manufactured. Next, to understand the operation of the cooler, the displacement of the piston has been measured by LVDTs (Linear Variable Differential Transformers), the displacement of the displacer has been measured via a laser optic method, and the phase shift between the piston and displacer has been analyzed. Finally, when the phase shift between displacement of the piston and displacer is 45°, the operating frequency is optimum and is decided by the resonant frequency of the expander, the mass and cross-sectional area of the displacer, and by friction and flow resistances. INTRODUCTION Stirling cycle coolers were first introduced to the commercial market in the 1950s as a small single-cylinder air liquefier and as a cryocooler for infrared sensors at about 80K. The compressor in the Stirling cryocooler is a valveless type. In order to provide high power densities and keep the system small, the average pressure is typically in the range of 1 to 3MPa, and frequencies are in the range of 20 to 60Hz.1 Free-piston Stirling engines acting as power systems were invented by William Beale in the early 1960s and have been in continuous development since that time at Sunpower. The first development of a linear free-piston Stirling cryocooler was accomplished at the Philips Laboratories in Eindhoven, by Haarhius (1978).2 Later, De Jonge (1979) presented a classic paper related to the theoretical analysis of free-piston Stirling machines.3 Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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The FPFD Stirling cryocooler consists of two compressor pistons driven by linear motors, which make pressure waves, and a pneumatically driven displacer piston with a regenerator. This is the most suitable design for a mechanical cryocooler for use in the night vision environment. In general, the efficiency of a Stirling cryocooler is mainly affected by the efficiency of the linear motor, the resonant frequency of the compressor and expander, the displacement of the piston and the displacer, and the phase shift between the piston and displacer. In this paper, for the given FPFD Stirling cryocooler, methods to measure displacement of the displacer are described, and experimental results are presented on the performance and optimum operating frequency of the Stirling cryocooler according to the variations of the phase shift between the piston and displacer.
DESIGN AND MANUFACTURING OF THE STIRLING CRYOCOOLER Figure 1 shows a schematic view of the FPFD Stirling cryocooler, which consists of two major parts: 1) the linear compressor module, and 2) the expander module. The linear compressor consists of linear motor, inner and outer yoke, permanent magnet, coil, cylinder, piston and spring; the expander module consists of displacer, regenerator in the displacer, displacer cylinder, spring, and heat exchanger. Table 1 shows dimensions and construction materials of the piston, displacer, regenerator, and magnet. Helium is used as the working fluid in the Stirling cryocooler cycle because of its ideal gas properties, its high thermal conductivity, and its high ratio of specific heats. Figure 2 shows a schematic diagram of the compressor as a single piston oscillator; here the coil spring is not shown and the piston’s mass reacts against the stiffness of the buffer volume V. The motion of the moving masses (pistons) may be described by the classical spring-mass-damper equation4,5:
where is the mass of piston, is the damping constant, is the gas spring constant, is the magnetic flux density, i is the zero-to-peak current supplied to the motor, l is the length of wire in the coil, is the angular velocity, and t is time.
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Helium gas at room temperature under moderate pressure can be considered as a perfect gas and the gas spring constant is given by6:
where is the cross section of the piston and is a parameter that depends on the type of compression: 1 (isothermal) (adiabatic). For small strokes and assuming the piston displacement is almost sinusoidal, the resonant frequency without the mechanical spring and with the mechanical spring are respectively given by:
Resonant frequency of the displacer is simply qiven by:
where is resonant frequency of the displacer, is mechanical spring constant of the displacer and is the mass of the displacer. Phase difference of the piston and the displacer is described as equation (6); in general, optimum phase difference is 45° in the FPFD Stirling cryocooler3.
where is phase difference of the piston and the displacer, is resonant angular frequency of the displacer, is operating angular frequency, is constant by friction and flow resistance, and is cross section of the displacer.
EXPERIMENTAL PROCEDURE Figure 3 shows a schematic diagram of the FPFD Stirling cryocooler. Piezo-electric dynamic pressure sensors were used to monitor pressure oscillations at the outlet of the compressor and in the buffers of each end. LVDTs (Linear Variable Differential Transformers) were provided at each end of the compressor for displacement measurement of the pistons. A laser displacement sensor was used to measure displacement of the displacer. A silicon-diode thermometer was attached to the cold head to measure the temperature of the cold end. After attaching those, the apparatus of the Stirling cryocooler except the component of the room temperature region was connected to the vacuum flange. During the experiment, the vacuum chamber was connected to a high vacuum pump with a pressure of Torr. The high vacuum pump system consists of a rotary roughing pump, turbo-molecular pump, and vacuum gauges.7,8
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The following tests were undertaken as part of the experimental analysis of the phase shifts among pressure and displacements of the piston and displacer: (1) The relationship between the output voltage of the laser displacement sensor and position of the displacer (2) Peak to peak displacement of the displacer with different operating frequency (3) Amplitude of the displacer’s stroke and phase shift with different temperature (4) Phase shifts between displacements of the piston and displacer with different operating frequencies
EXPERIMENTAL RESULTS AND DISCUSSION The experimental setup for calibration of displacer’s stroke, as shown in Figure 4, consists of the expander, vernier caliper, optical windows, laser displacement sensor, dewar and digital multimeter. Displacement of the displacer is measured by a caliper vernier with resolution of 0.01mm,
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and the corresponding voltage output is recorded. Figure 5 shows that the relationship between the output voltage of the laser displacement sensor and position of the displacer is almost linear. Figure 6 shows cool down characteristics of the cold end with laser displacement sensor in the expander of the Stirling cryocooler. The charging pressure was and operating frequency was 50Hz. Input power and applied voltage were about 17W and 7V, respectively. In this case, the lowest temperature was 97K. Figure 7 and Figure 8 present real time amplitude and phase shifts of the displacer motion measured by the laser displacement sensor, piston motion, pressure, current and applied voltage. Measured data for the peak-to-peak displacement of the displacer with different operating frequency and peak-to-peak pressure of the compressor are presented in Figure 9. Resonant frequencies of the compressor and expander are 54Hz and 64.5Hz, respectively. As the peakto-peak pressure of the compressor was increased, peak-to-peak displacement of the displacer was increased. The peak-to-peak displacement of the displacer increases in the range of 0 – 64.5Hz, but decreases steeply when the operating frequency is higher than the resonant frequency. Figure 10 shows experimental results for the amplitude of the displacer’s stroke and phase shift between displacements of the piston and displacer with different temperatures of the cold end under the condition of constant operating frequency at 50Hz. The amplitude of the displacer’s stroke and the phase shift between the piston and displacer are seen to decrease as the temperature of the cold end decreases. Figure 11 shows the phase shifts between displacements of the piston and displacer with different operating frequencies. And Figure 12 shows the phase shifts between displacements of the displacer and pressure of the compression space with different operating frequencies. The
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phase shifts and decrease as the operating frequency increases and the temperature of the cold end decreases. As the temperature of the cold end is lower, and converge to 45° and -7°, respectively. And it is clear that and approach 0° and about -40° when the operating frequency coincides with the resonant frequency of the expander, in other words, refrigeration doesn’t occur. Therefore, from the results of Figure 11, 12 and Equation (6), when is 45°, the operating frequency is optimum and is decided by the resonant frequency of the expander, the mass and cross-sectional area of the displacer, and by the friction and flow resistances.
CONCLUSIONS A Stirling cryocooler actuated by the electric force of a dual linear motor has been designed and manufactured. Next, to understand the operation of the cooler, the displacement of the piston has been measured by LVDTs (Linear Variable Differential Transformers), the displacement of the displacer has been measured via a laser optic method, and the phase shift between the piston and displacer has been analyzed. The relationship between the output voltage of the laser displacement sensor and the position of the displacer was almost linear. Cool down characteristics of the cold end were investigated using the laser displacement sensor in the expander of the Stirling cryocooler; in this case, the charging pressure was and the operating frequency was 50Hz. Input power and applied voltage were about 17W and 7V, respectively. The lowest temperature was 97K. As the peak-to-peak pressure of the compressor was increased, peak-to-peak displacement of the displacer also increased. The peak-to-peak displacement of the displacer increased for drive
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frequencies in the range of 0 to 64.5Hz (the resonant frequency of the displacer), but decreased steeply when the operating frequency was higher than the resonant frequency. When the phase shift between displacements of the piston and displacer is 45°, the operating frequency is optimum and is decided by the resonant frequency of the expander, the mass and cross sectional area of the displacer, and by friction and flow resistances.
ACKNOWLEDGMENT This work is supported by the “Dual use technology program” with Wooyoung as industrial partners.
REFERENCES 1.
R. Radebaugh, “Development of the Pulse tube Refrigerator as an Efficient and Reliable Cryocooler,” Proceedings Institute of Refrigeration, London, 2000.
2.
G. Walker, Miniature Refrigerators for Cryogenic Sensors and Cold Electronics, New York, Oxford University Press, 1989
3.
A.K.De Jonge, “A Small Free-Piston Stirling Refrigerator,” American Chemical Society, 1979.
4.
M.K. Heun, et al., “Investigation of Gas Effects on Cryocooler Resonance Characteristics,” Cryocoolers 9, Plenum Publishing Corp., New York, 1997, pp. 421-430.
5.
A.A. Kornhauser, “Dynamic Modelling of Gas Springs,” Transactions of the ASME, Vol.116, 1994.
6.
Ravex, etc., “Development Progress of a Long Life Twin Piston Pressure Oscillator,” Advances in Cryogenic Engineering, 39, 1994.
7.
N. Fujiyama, etc., “Development of Micro-Stirling Cooler”, ICEC16/ICMC Proceedings.
8.
Y. Ikuta, etc., “Development of a Long-Life Stirling Cryocooler,” Cryocoolers 11, Kluwer Academic/ Plenum Publishers, New York, 2001, pp. 95-102.
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Dynamic Analysis of a Free Piston Stirling Refrigerator Y-J. Hong, S-J. Park, H-B. Kim, and D-Y. Koh HVAC & Cryogenic Engineering Group Korea Institute of Machinery & Materials Yu-Sung, Taejeon 305-600, Korea
ABSTRACT Small free-piston type Stirling refrigerators are widely used for the cooling of cryogenic infrared sensors. In this study, a dynamic analysis of a small free-piston type Stirling refrigerator is performed with an isothermal thermodynamic model to understand the characteristics of the refrigerator. In the free-piston type Stirling refrigerator, the pressure wave, which is generated by the dual opposed pistons of the compressor, is the driving force of the displacer. The cooling capacity in the displacer's expansion space is generated by the motion of the displacer and the pressure wave. Therefore, the dynamic characteristics of the linear compressor’s pistons and displacer have a strong effect on the thermodynamic performance of the refrigerator. In this study, the electric equations describing the motor's force generation and the equations of motion for the linear compressor’s pistons and displacer were simultaneously solved using a 4thorder Runge-Kutta method. The pressure drop through the displacer, which contains the regenerator, was included as the driving force of the displacer. The analysis was performed for an example refrigerator, with the key operating variables treated parametrically; these variables included the operating frequency of the linear compressor, the charge pressure of the refrigerator, and the supplied electric voltage to the linear compressor.
INTRODUCTION The small free-piston type Stirling refrigerator has been widely used for the cooling of cryogenic infrared sensors to temperatures in the range of 30-100 K. Figure 1 illustrates a typical small free-piston type Stirling refrigerator. Helium is the typical working fluid in a free-piston type Stirling refrigerator, and the pressure waves that are generated by the dual-opposed pistons of the linear compressor are the driving force of the displacer in the expander. The dynamic characteristics of the linear compressor are determined not only by the forces of the mechanical springs and masses of the moving parts of the compressor, but also by the forces caused by the pressure differences between the compression space and the buffer space. In order to minimize the input power to the linear compressor, the force to actuate the linear compressor should be minimized. Usually the force to actuate the dynamic system can be miniCryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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mized when the system is operated at its natural frequency. Yuan et al.1 has predicted the effects of the diameter of the cylinder, the mass of the moving assembly, the charge pressure, etc. to the natural frequency of the refrigerator. Heun et al.2 has studied the effects of the charge gas to the resonance of the linear compressor. However, the thermodynamic performance of the refrigerator is mainly affected by the dynamic characteristics of the displacer and thermal losses. In this study, an electric equation for the motors and the equations of motion for the linear compressor’s pistons and displacer were simultaneously solved using a 4th-order Runge-Kutta method; the pressure drop through the regenerator was included as the driving force of the displacer. The analysis was performed for an example Stirling refrigerator, with key operating variables treated parametrically; these variables include the operating frequency of the linear compressor, the charge pressure of the refrigerator, and the supplied electric voltage to the linear compressor.
THE GOVERNING EQUATIONS The electromagnetic force is the driving force of the pistons of the linear compressor. The damping force due to viscous friction, the spring force, and the force due to the difference of pressure between the compression space and buffer space are coupled. The equations of motion and the relationship for the electric driving force of the linear compressor are as follows.
where The natural frequency of the linear compressor can be written as follows from Eq. (1).
Therefore, the natural frequency of the linear compressor is dependent not only on the spring stiffness and mass of the moving parts, but also on the diameter of the cylinder, the charge pressure, the temperature of the cold end, the dead volume of the Stirling refrigerator, etc. The pressure difference between the ends of the regenerator is the driving force of the displacer. The pressure of the expansion space is affected by the motion of the displacer and the flow resistance of the regenerator. The equation of motion of the displacer with regenerator is as follows.
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The above equations (1), (2), (3) must be simultaneously solved to get the stroke of the piston and displacer, the applied voltage, etc.; the Kays and London’s correlation3 is used for the flow resistance of the regenerator. The ideal refrigeration power and efficiency of the linear compressor are given as follows.
CALCULATION PROCEDURE The applied voltage to the linear compressor in the analysis is assumed to be a pure sine wave. The iteration process starts with initial guesses for the other unknown variables. Equations (1), (2), (3) are solved iteratively using a 4th-order Runge-Kutta method with a fixed time step. Figure 2 shows the input pure sine waves and the measured voltage and current waveforms. The small disturbances in the measured V and I occur at the turning point of the piston, but there is no disturbance in the pressure curve. Therefore, the assumption of pure sine waves can be used in the calculation. The iterations are performed for several piston compression cycles until the cycle-to-cycle differences between the instantaneous values of the piston and displacer motions are within tolerances. The parameters in the calculation are listed in Table 1.
THE RESULTS The input power to the linear compressor of the Stirling refrigerator can be minimized when the system is operated at the natural frequency of the linear compressor. Figure 3 shows a compari-
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son between the calculated natural frequency and the measured value for a coldend temperature of 270 K. The natural frequency of the linear compressor is higher when the charge pressure of the Stirling refrigerator is higher because the natural frequency is highly dependent on the gas spring of the helium. The discrepancy between the calculated and measured frequencies indicates that the compression and expansion process in the linear compressor is not an isothermal process. Figure 4 shows the instantaneous values of the piston displacement, voltage and current at a charge pressure of when the operating frequency is 50 Hz. The results show the maximum amplitude of piston displacement is about 5 mm for an input power of 34.7 W; other parameters include: the applied voltage is the RMS current is 3.47A, and the power factor is 99.87%. The phase angle difference between piston movement and current is about 117.5 degrees and the natural frequency of the linear compressor is about 46 Hz when the coldend temperature is decreased to 77 K. The amplitude of the piston’s displacement and phase angle are highly affected by the value of the damping coefficient in Eq. (1). In this study, the value of the damping coefficient is adjusted to provide reasonable agreement with the amplitude of the piston’s displacement in the measurements.4 Figure 5 shows the current, input power, and the amplitude of the pistons when the charge pressure is varied from 10 to with and f = 50 Hz. The current is smaller when the charge pressure is increased, and minimized at However, note that the amplitude of the pistons also decreases when the charge pressure is increased. Figure 6 shows the efficiency of the linear compressor and the phase lag between current and piston motion when the operating frequency is 50 Hz. The maximum efficiency of the compressor
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is about 70% at a fill pressure. The calculated efficiencies of the linear compressor have physical meaning compared with the measurement of a moving coil type linear compressor.5 The phase lag decreases with increasing charging pressure, and the resonance aligns with the 50 Hz drive frequency when the charge is and the cold end temperature is 77 K. The calculated stroke of the displacer is very small compared with the measured values.5 The stroke of the displacer is the main parameter for the cooling power. Therefore, the calculated cooling power at 77K is very small. However, the dynamic characteristics of the linear compressor would not be altered greatly for a change of the stroke of the displacer. For a more reliable and accurate calculation, further studies of the flow resistance through the regenerator are needed. Figure 7 shows the characteristics of the stroke of the piston, current, and input power with different applied voltages for a fixed charge pressure of and fixed operating frequency of 50 Hz. Increasing the applied voltage results in a linear increase of the amplitude of the piston’s stroke, current, and input power, but it has no effect on the efficiency of the linear compressor. The amplitudes of the pressure and displacer stroke are also increased when the voltage is increased. It can be concluded that a change in the voltage and current does not alter the dynamic characteristics of the Stirling refrigerator. Figure 8 shows the characteristics of the stroke of the piston, current, and input power for a fixed charge pressure of and a fixed input voltage of Increasing the frequency to 50 Hz results in a decrease of the magnitude of the current, but above 50 Hz the current increases. The magnitude of the stroke is seen to always increase with increasing operating frequency. However, at higher operating frequencies, the linear compressor has much higher input power.
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SUMMARY The electric equations describing motor force and the equations of motion for a linear compressor’s pistons and displacer were simultaneously solved using a 4th-order Runge-Kutta method. Results were obtained and are discussed for an example refrigerator with key operating variables treated parametrically; these variables include the operating frequency of the linear compressor, the charge pressure of the refrigerator, and the supplied electric voltage to the linear compressor.
ACKNOWLEDGMENT This work is supported by the “Dual use technology program” with Wooyoung as industrial partners.
REFERENCES 1.
S.W.K. Yuan, et al., “Prediction of natural frequency of NASA 80K cooler by Stirling refrigerator performance mode,” Cryogenics, Vol. 34 (1994), pp. 383-388.
2.
M. K. Huen, et al., “Investigation of Gas Effects on Cryocooler Resonance Characteristics,” Cryocoolers 9, Plenum Press, New York (1997), pp. 421-430.
3.
W. M. Kays, A. L. London, Compact Heat Exchanger, McGraw-Hill Company (1976).
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Y. J. Hong, S. J. Park, H. B. Kim, J. H. Kim and B.K.Yu, “Study of the Linear Compressor Characteristics of the Stirling Refrigerator,” Journal of KIASC, Vol.3, No. 2, 2001, pp. 49-54 (in Korean).
5.
Y. J. Hong, S. J. Park, H. B. Kim, B.K.Yu and Y. D. Choi, “Study of the Dependency of the Stirling Cryocooler’s Thermodynamic Performance on the Operating Frequency,” Journal of KIASC, Vol.4, No. 1, 2002, pp. 140-144 (in Korean).
6.
S. J. Park, Y. J. Hong,, H. B. Kim, D. Y. Koh, J. H. Kim and B.K.Yu, “An Experimental Study of the Phase Shift between Piston and Displacer in a Stirling Cryocooler,” Cryocoolers 12, Kluwer Academic/Plenum Publishers, New York (2003).
Low Vibration 80 K Pulse Tube Cooler with Flexure Bearing Compressor P.C. Bruins, A. de Koning and T. Hofman Thales Cryogenics b.v. NL 5626 DC Eindhoven, The Netherlands
ABSTRACT A Pulse Tube Cooler and associated cooler drive electronics have been developed at Thales Cryogenics to provide cryogenic cooling for applications that are extremely sensitive to vibrations. Initially, the development focussed on the double inlet design because of its potential high efficiency. The DC flow arising in this design can decrease the performance significantly. Although this DC flow is successfully suppressed in prototype double inlet pulse-tubes, the solution proves to be too complex to be acceptable for large production quantities. It is therefore concluded that due to the DC flow, the double inlet design is not suitable for mass production, and the research further focussed on the development of an inertance type pulse tube. Optimisation of the U-shape inertance-type pulse tube results in a very reproducible cooling system that is easy to produce in large quantities. The cooling performance of 500 mW at 80 K, for 60 W of electrical input, is comparable to that of a double-inlet system without DC flow. Based on previous experience with the vibration reduction of Stirling coolers, a DSPbased cooler drive unit has been designed to reduce the vibration of the dual-opposed-piston flexure bearing compressor. The paper describes the results of a reduction method for DC flow, gives the design tradeoffs for the inertance pulse-tube, and describes the vibration control algorithm, hardware and results. INTRODUCTION The objective of the development is to design and build fit-for-manufacture cooling systems with extremely low vibration, fitting in a pre-described dewar envelope, and meeting various challenging demands required by the customer. The most important technical requirements are a cooling power of 300 mW at 80 K at 45 °C skin temperature, at a maximum electrical input power of 60 W. Furthermore, the level of cooler-induced vibration on the system should be minimised by using vibration supression electronics.
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PULSE TUBE DEVELOPMENT Double inlet pulse tube Because of the limited diameter of the dewar, the development has initially focussed on adaptation of an existing Thales double inlet pulse tube design, of which five prototypes were successfully tested to the specified performance in 1998 in a development program for the French MOD. The secondary orifices of these prototypes were manually optimized, in order to overcome the often observed non-reproducibility in a cooler batch caused by DC flow. The present development has aimed at finding a solution to the DC-flow problem which would be suitable for mass-production. Contrary to the often used assumption that DC flow is caused by asymmetric flow impedance of the secondary orifice, it has been found that DC flow takes place also when the secondary orifice is perfectly symmetrical. Referring to Figure 1, in the secondary orifice gas flows from the warm end of the regenerator (I) to the warm end of the pulse tube (II) during the high pressure part of the cycle. At this part of the cycle, the gas has a density The gas flows back from (II) to (I) during the low pressure part of the cycle, when the density is This implies that for a symmetrical orifice, where the net volume flow over one cycle is zero, there is a net mass flow and thus an enthalpy flow. A solution has been found in creating an asymmetry in the secondary orifice, thereby creating a net volume flow in the direction of the warm end of the regenerator, i.e. from (II) to (I). Calculations have been performed with Thales’ pulse tube simulation program2 in order to find the values of the pressure waves in the spaces on either side of the secondary orifice, and it is found that an asymmetry of 9 % would be sufficient to neutralize the DC flow (see Table 1). Experimental pulse tubes have been built in which the secondary orifice is realized as a replaceable plug which contains the actual orifice, and which is sealed with two small O-rings to prevent gas leak along the orifice (see Figure 1, detail B). The flow resistance of the orifice is tuned with the diameter of the hole (optimum 00.4 mm), and it has been shown that reproducible asymmetries between 4 and 20% can be realized by chamfering the outlet of the orifice with a drill angle of over a depth between 3 and 6 mm.
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The reproducibility of the cooling performance with an asymmetrical secondary orifice has been tested using two identical prototypes. Even though all measurable properties of these prototypes are identical (including their flow resistance for steady flow), it turned out that there is a significant difference in the asymmetry required for their optimal performance (Table 2). It is therefore concluded that manual tuning of each pulse tube would be required in production, making the double inlet concept unsuitable for mass production.
Inertance Type Pulse Tube After dismissing the double inlet pulse tube, the development has focussed on optimizing a pulse tube that uses an inertance to provide the necessary phase shift. The limited diameter of the dewar reduces the degrees of freedom in optimizing the regenerator and pulse tube diameters. The space available for the tubes is further decreased by the space needed for the vacuum-brazed connection between tubes and warm and cold end (see Figure 1). An optimization run performed with Thales’ simulation program has revealed that the optimum configuration within these constraints is a regenerator diameter of ø9 mm and a pulse tube of ø5.5 mm. In other words, the regenerator diameter is maximized, while maintaining enough space to accommodate a pulse tube with a volume that is approximately one third of the total volume of the regenerator. For the first prototypes, it was intended to use stainless steel regenerator gauzes that have the same properties as the gauzes used in the regenerators of Thales’ free displacer Stirling coolers. Calculations performed for the geometry of these prototypes have revealed that the heat capacity of these regenerators is insufficient for efficient operation of the pulse tube at 80 K. The calculations indicate that the optimum filling factor of a pulse tube regenerator is higher than the optimum filling factor of a Stirling regenerator. It was also found that a higher filling factor in the entire pulse tube regenerator would significantly increase the pressure drop over the regenerator, thereby reducing the performance of the pulse tube cooler. An optimization run where the filling factor and wire diameter were varied along the regenerator has indicated that it is optimal to fill the coldest part of the regenerator with gauzes with a high filling factor, and the warm part with gauzes with a lower filling factor. Combining the simulation results with practically available gauzes, an optimum is found with a high filling factor in the cold end, and a low filling factor in the warmer parts (Table 3). Measurements with three differently stacked regenerators have been performed to confirm the predictions of the simulation model. Finally, an inertance optimization has been performed to match the optimum frequency of the pulse-tube with the resonance constraints of the flexure bearing compressor, and secondly to optimize the phase shift in the optimal pulse-tube. The resonance of a linear compressor depends on the filling pressure, cold finger volume and cold finger damping. It is very convenient to have a mechanism of fine tuning the optimum frequency
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of the cold finger, as this makes it possible to drive a large range of cold fingers with a limited number of compressor designs4. For a given regenerator and pulse tube volume, the impedance of the inertance determines the optimum frequency of the pulse-tube5. It is found that the optimum working frequency of the pulse tube can be shifted by approximately 5 Hz without a significant influence on the performance of the pulse-tube. The impedance of the inertance consists of a resistive part R and an inductive part L. As the inductive part is responsible for the beneficial phase change over the inertance5,6, one would expect that it is always beneficial to increase the ratio L/R for the inertance, as this would increase the inertia effect for a given flow resistance. However, an increase in L for the same R means that the overall impedance Z of the inertance increases, thereby decreasing the optimum frequency of the pulse-tube. Furthermore, as L is proportional to 1/A and R is proportional to the ratio is proportional to the surface area A of the inertance. For a large ratio a relatively large surface area A is therefore required, in which case the length should be large to obtain the proper flow resistance. The resulting volume of the inertance causes a storage effect (or capacity C) which gives a phase lead that counterbalances the desired phase lag. Figure 2 shows the measured phase difference between the pressure wave in the pulse tube, and the (small) pressure wave in the buffer. The figure reveals that as the inertia effect in the inertance increases (larger diameter and / or larger length), the phase lead of the pressure in the tube increases, which is the desired effect of the inertance5.
VIBRATION REDUCTION Compressor Vibration The dual opposed piston compressor has an axial vibration level that is specified at which is acceptable for most applications. The axial vibrations originate from an unbalance in the movement of the two opposing pistons. Earlier work7 has indicated that it is possible to reduce the vibrations by a factor of 50 by adjusting the current through the coil of one piston. In collaboration with CILTEC, the Centre for Interfacing Low Temperature Electronics and Coolers of the University of Twente, it was found that with present day electronics, a digital feedforward system is an affordable, reproducible and reliable means of reducing the vibrations of the compressor. A DSP-based vibration control system has been realized in commercially available hardware.
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The heart of the vibration reduction system consists of a DSP, an accelerometer and two analogue amplifiers as shown in Fig. 3. During start-up of the cooler, the DSP measures the transfer functions of the system via a step response. In steady state operation, the transfer functions are used in a feed-forward loop to minimize the vibrations by adapting the phase and amplitude of the current through one of the compressor coils. Meanwhile, a software programmed PID loop maintains a constant cold tip temperature by changing the input power to the compressor. The DSP is capable of minimizing the vibration of the compressor at the drive frequency and the first two harmonics, see Figure 4. The vibrations of the compressor are measured by means of a commercial-of-the-shelf accelerometer. In order to get a large acceleration signal for small force levels, the moving mass should be kept to a minimum. For this reason, the acceleration transducer is attached directly to the compressor, and the compressor is suspended flexibly in order to be able to measure the acceleration. This implies that the moving mass is only the mass of the compressor, and not that of the entire application. Figure 4 shows the results of the vibration reduction.
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Pulse Tube Vibration The absence of moving parts in the cold finger generally makes a pulse-tube cooler very suitable for vibration-sensitive applications. The present research has revealed, however, that a pulse-tube cold finger is not completely vibration free. After de-coupling the compressor vibration there is still some small movement of the pulse-tube cold finger. Significant reduction in this level of movement has been achieved by studying the pulse-tube design and improving it with respect to vibration characteristics.
LIFETIME The design of Thales’ moving magnet flexure bearing compressors4 removes all failure mechanisms commonly identified8 in pulse-tube cryocoolers. Compressor wear is eliminated by the flexure bearings which fully support the piston mass at the front and back side with high radial stiffness. Because of the moving magnet concept, moving current leads and helium-tight current feedthroughs are not needed. The moving magnet concept also moves the synthetic material of the coil insulation outside the working gas, thereby strongly reducing the risk of gas contamination. All parts that are in contact with the working gas are metallic parts, joined together with laser- and electron beam welding techniques. The lifetime expectancy is supported by lifetime tests that are being performed on 10 coolers with the described compressor design. The lifetime coolers have presently gathered a total of 118.000 running hours, with no failures and no performance degradation. The lifetime tests include tests where the cooler is running under high side loads (up to l0g), temperature cycles between the specified extremes (-54°C to 71°C), and cooler on / cooler off cycles.
CONCLUSION Based on an inertance type pulse tube, a U-shape pulse-tube has been designed with a cooling performance of 500 mW at 80 K. The pulse tube is driven by a long-life moving magnet flexure bearing compressor, giving a virtually failure-mode free cooling system. A separate DSP-based vibration control system reduces the vibrations originating from unbalanced motion of the compressor pistons with a sufficient reduction factor. Detailed design work on the pulse tube was needed to reduce the level of vibration induced by the pulse tube cold finger.
ACKNOWLEDGMENT The authors would like to acknowledge Bert Rijpma and Marcel ter Brake (CILTEC / Low Temperature Division) and Rindert Nauta (EAS) of the University of Twente for their support, thoughts, hardware and software.
REFERENCES 1. 2.
3. 4. 5. 6. 7. 8.
Gedeon, D., “DC gas flows in Stirling and pulse tube refrigerators,” Cryocoolers 9, Plenum Press, New York (1997), pp. 385-392. Hooijkaas, H.W.G., “Pulse tube development using Harmonic Simulations,” Cryocoolers 10, Kluwer Academic/Plenum Publishers, New York (1999), pp. 359-367. VDI-Wärmeatlas, VDI-verlag GmbH, Düsseldorf (1984). Meijers, M., Benschop, A.A.J. and Mullié, J.C., “High Reliability Coolers under Development at SignaalUSFA,” Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2001), pp. 111-118. Hooijkaas, H.W.G., “Miniature Stirling-Type Pulse-Tube Refrigerators,” PhD Thesis, Eindhoven Technical University, June 2000. Gardner, D.L. and Swift, G.W., “Use of inertance in orifice pulse-tube refrigerators,” Cryogenics, Volume 37, 1997. Verberne, J.F.C., et al., “Reduction of the Vibration Generated by Stirling Cryocoolers Used for Cooling a High-Tc SQUID Magnetometer,” Cryocoolers 8, Plenum Press, New York (1995), pp. 465-474. Ross, R.G., “Cryocooler Reliability and Redundancy Considerations for Long-Life Space Missions,” Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2001), pp. 637-648.
Development of 40-80K Linear-Compressor Driven Pulse Tube Cryocoolers J. Liang, J.H. Cai, Y. Zhou, W.X. Zhu, L.W. Yan, W. Jing, Y.L. Ju, Y.K. Hou, and K. Yuan Cryogenic Laboratory, Technical Institute of Physics and Chemistry Chinese Academy of Sciences, Beijing 100080, China
ABSTRACT Along with the commercialization of HTS devices in fields such as mobile communications, and the development of far infrared devices for space and military applications, there arises a strong demand for compact and reliable cryocoolers working at 40-60K. The pulse tube cryocooler driven by a linear pressure wave generator has the potential to achieve high reliability and very long lifetime because of the absence of moving parts at low temperature. The ability of this kind of pulse tube cryocooler to achieve efficiencies comparable with Stirling cryocoolers has been theoretically and practically proven1-3. Hence, pulse tube cryocoolers appear a good choice to meet the abovementioned requirements. On the basis of our previous work that focused on miniature co-axial, linear-driven pulse tube cryocoolers with a few hundred milliwatts of cooling power at 80K4, we are trying to develop a new series of pulse tube cryocoolers working at lower temperatures and with larger cooling powers. The goal is to provide about 200mW at 40K or 1 ~2 W at 60K, with input power as small as possible. The pressure wave generators are being developed in our laboratory uses a moving coil linear motor supported by flexure springs. Their maximum swept volumes are 2, 4, 5, and Two types of pulse tube cold head configurations, i.e. coaxial and U-shape, have been adopted to fit different applications. The present status of development for these coolers is presented in this paper,
INTRODUCTION Cryocoolers for cooling traditional infrared devices usually work at 80K or at slightly higher temperatures. The cooling power is usually a few hundred milliwatts. With the development of far and very far infrared devices, lower temperatures such as 40K are needed for these devices to work properly. Larger focal planes also require larger cooling power. The high temperature superconductive devices such as filters, SQUID’s, etc, are finding more and more applications. Although their superconductive transition temperature is near the temperature of liquid nitrogen, they give more optimum performance at temperatures of 60-70K. Based on these demands, we estimate that it is necessary to develop pulse tube cryocoolers working at 40-80K. These pulse tube cryocoolers must be compact, efficient, and reliable in order to compete with Stirling cryocoolers, and a linear driven compressor must be used. The cold head Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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can be either coaxial or U-shape, because each configuration has its own merits and disadvantages. We are developing both kinds of cold heads to give us more flexibility to adapt to various applications. The cooling capacities of the linear driven pulse tube cryocoolers are as follows: 500mW/80K, 1W/80K, 200mW/60K, 1~2W/60K, 200mW/40K. The maximum swept volumes of the linear compressors being developed are: (single piston), (dual piston), (single piston), and (dual piston). These coolers should be able to survive severe environment tests. The goal for their lifetime is greater than 20,000 hours.
SYSTEM DESIGN CONSIDERATIONS The critical point to be considered for system design is the cooler efficiency, which is indicated by the coefficient of performance (COP). The cooler efficiency is the product of the compressor efficiency and the cold head efficiency. The compressor efficiency is the ratio of the output PV power to the electrical power. The cold head efficiency is the ratio of the net cooling power to the PV power input. To increase the cooler efficiency, three aspects should be carefully considered. Firstly, the compressor should be highly efficient. This requires good design of the forces working on the piston. The friction between piston and cylinder should be eliminated by the flexure bearing support structure. The magnetic circuit losses should be reduced to the minimum. The current in the coil and the resistance of the coil should be small enough so that the heating by the Joule effect can be minimized. Secondly, the cold head should be efficient in converting PV power to net cooling power. This converting process happens in the regenerator, with the help of the pulse tube and the orifice5. The geometry of the cold head should be optimized according to theoretical calculation and practical experiences. Two goals are to be achieved at the same time. First, the PV power consumption should be minimized. This means proper regenerator flow channels with a low flow resistance and a proper phase relationship between the mass flow rate and the pressure. Secondly, the cooling power losses should be minimized. The regenerator inefficiency losses and the conduction loss along the regenerator are the principal losses. The above two points may be contradictory and a tradeoff is inevitable to achieve the best overall performance at a given working temperature. Thirdly, good matching of the compressor and the pulse tube cold head is very important to the efficiency of the cooler. Unlike the rotary compressors, the operation of a linear driven compressor is affected by the characteristics of the cold head coupled to it. In other words, the compressor and the cold head interact with each other. The cooler should be optimized as a whole system including a model of the compressor. Beside the efficiency, other key factors such as the lifetime, the environment test conditions, are also considered. At this stage in the development of the principle models we are mainly focused on efficiency. In the next stage, that of the development of engineering models, the focus will be on reliability.
COMPRESSOR DESIGN, FABRICATION AND CHARACTERIZATION The compressor is designed according to the thermal design of the pulse tube cryocooler. The moving coil and the piston are supported with two flexure springs. The flexure spring is one of the key technologies for the compressor. It has been numerically analyzed using finite element models and experimentally tested long before we began to design the compressor. Its long lifetime operation has been demonstrated on a special spring test apparatus. The components are designed so that alignment can be assured during assembly. With this method, the friction between the piston and the cylinder can be eliminated. The magnets are made of neodymium iron boron and the magnetic circuit is made of iron-cobalt alloy. The five forces working on the piston, namely, the electromagnetic force, the spring force, the pressure difference
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force, the inertia force of the moving mass, and the friction force (may be negligible in our compressor) are analyzed in their phases and amplitudes. The calculation of the compressor forces is necessary to ensure that the compressor operates at its optimum condition when coupled with the pulse tube cold head. During fabrication the accuracy of the dimensions is controlled so that damage to the surfaces is avoided. The pieces are also kept clean. After cleaning and baked out, the components are assembled in a clean environment. The radial movement of the piston in one cycle is observed before the piston is fitted into the cylinder. The amplitude of the radial movement should be smaller than the gap between the piston and the cylinder. A high accuracy instrument for measuring cylindricity is used to control the outer surface of the piston and the cylinder. After assembly, the compressor is characterized with a series of standard volumes. The pressure is measured at one end of the volume. The charge pressure, frequency, and input power are varied. The test results are very useful for correctly matching the compressor and the pulse tube cold head. A photo of the prototype of the linear compressor is shown in Fig. 1. We first developed a single piston compressor with a maximum swept volume of then with the experience obtained, the swept volume was enlarged to On the basis of these single piston compressors, dual piston compressors with maximum swept volumes of 4 and will be developed. Presently, the 2 and compressors have been fabricated and tested. If we consider the PV power output as the difference of the input power and the Joule heating power of the coil, the efficiency of the compressor defined as the PV power over input power is greater than 80%; this is very encouraging progress toward our goal to develop high efficiency pulse tube cryocoolers. To improve our compressor design, a compressor will soon be sent to undergo environment tests.
PULSE TUBE DESIGN AND FABRICATION We have many years of experience developing coaxial pulse tube cold head designs. With a coaxial cold head, it is easy to replace Stirling coolers without changing the dewar. This is very desirable in some applications. With the coaxial pulse tube design, it is also convenient to incorporate a multi-bypass and other techniques for improving the performance. However, this design has its disadvantages, such as difficult heat dissipation at the hot end and flow losses at the cold end. Taking into account all these factors, the efficiency of a coaxial pulse tube cold head is lower than its equivalent U-shape or linear one. Beside its better efficiency, the U-shape is convenient for integrating with the devices to be cooled; it is also a strong structure for supporting side forces induced by launch vibration or han-
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dling loads. For space applications where the dewars are specially designed, the U-shape is a good choice. In order to give us more flexibility to satisfy different application requirements, we are developing both coaxial and U-shape cold heads in parallel. To optimize the cold head, regenerator inefficiency loss, pressure drop, and conduction loss should be reduced. These factors are evaluated and balanced in our theoretical model6-8. After the geometry is optimized, it is very important to control the DC flow in the cooler. This can be achieved with nozzles9. A test bench to measure the oscillatory flow characteristics is being set up using a hotwire anemometer and pressure sensors. Presently, coaxial cold heads for coupling with the and compressors have been fabricated and tested. U-shape cold heads for 2, 4, 5 and compressors are in fabrication and will be finished before the end of 2002.
EXPERIMENTAL RESULTS As our own linear driven compressors are under development and are not yet available for coupling with the pulse tube cold heads, we are currently testing the cold heads using commercial linear compressors. Initial testing is being conducted with the compressor of a Leybold Polar SC-7 COM Stirling cooler coupled with our coaxial cold heads. This compressor has a maximum swept volume of A schematic diagram of the coaxial pulse tube cryocooler is illustrated in Fig. 2. The coaxial configuration has advantages of compactness and convenience for integration. A typical cool-down profile of the pulse tube cryocooler is shown in Fig.3. The cooler can reach its no-load temperature of 43K within 30 minutes. In the experiment, the needle valve and the symmetry-nozzle were used for the orifice and double inlet, respectively. The experiments were carried out at a charge pressure of 2.7MPa and an operating frequency of 35 Hz.
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The Leybold compressor is designed for the Stirling cryocooler working at a resonant frequency of 52 Hz. But for the pulse tube cryocooler, the resonant point is changed. It is necessary to find the optimum operating frequency and charge pressure for the coaxial pulse tube cryocooler system. Figure 4 shows the effect of varying the frequency on the performance of the cooler. The lowest temperature varies remarkably with changing frequency. It was found that a frequency of 35Hz is the optimum frequency for the cold head tested. Charge pressure is another important operating parameter that affects the performance of the pulse tube cryocooler. Its effect is given in Fig. 5. The no-load temperature decreases about 3K as the charge pressure increases from 2.5MPa to 2.7MPa, and then increases about 2K as the charge pressure increases to 2.9MPa. The phase shifter at the hot end of the pulse tube cryocooler plays an important role in the performance of the cooler. To investigate the effect of the phase shifter, a variety of different phase shifting devices were tested under the same experimental conditions. Figure 6 shows the experimental results of four kinds of phase shifters. The lowest temperature was achieved using a needle valve as the orifice and the symmetry-nozzle as the double inlet. A cooler using only an inertance tube with an inner diameter of 1. 5mm and length of 1 m as the phase shifter led to a no load temperature of 54K; this is higher than the 47K obtained when both the orifice and double inlet were needle valves. Therefore, the inertance tube could not completely substitute for the orifice and double inlet. With the inertance tube plus the double inlet as the phase shifter, a no load temperature of 45K was achieved.
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The cooling power of a pulse tube is the most fundamental measure of its performance. Fig. 7 shows the relationship between the cooling temperature and the cooling power for our optimum coldend configuration. The cooler achieves a no-load temperature of 43K and produces a cooling power of 2W at 58K and 5W at 83K. This result is quite promising considering the fact that the pulse tube cryocooler is not yet in its optimum configuration. Future work will be carried out to further optimize the pulse tube cryocooler design including better matching of the compressor with the cold head.
CONCLUSION We have been developing a series of linear-compressor-driven pulse tube cryocoolers working in the temperature range of 40-80 K. The prototype compressors have achieved efficiencies higher than 80%. Both coaxial and U-shape pulse tube cold heads are being developed to fit various application requirements. Test results are presented for a prototype co-axial pulse tube cryocooler driven by a linear compressor with a maximum swept volume of The co-axial pulse tube cryocooler has achieved its optimum performance at a charge pressure of 2.7MPa and an operating frequency of 35Hz. The optimum phase shifter was found to be a needle valve as the orifice and a symmetry-nozzle as the double inlet. The cooler reaches a no-load temperature of 43K and produces 2W of cooling power at 58K under its optimum operating conditions.
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ACKNOWLEDGMENT This work is supported by the Natural Science Foundation of China and the Chinese Academy of Sciences.
REFERENCES 1. Tward, E., Chan, C.K., Raab, J., Nguyen, T., Colbert, R. and Davis, T., “High Efficiency Pulse Tube Cooler,“ Cryocooler 11, Kluwer Academic/Plenum Publishers (2001), pp.163-167.
2. Radebaugh, R., “Pulse Tube Cryocoolers for Cooling Infrared Sensors,” Proceedings of SPIE ,vol. 4130 (2000), p.363.
3. Nast, T. C., Champagne, P. J., Kotsubo, V., Olson, J., Collaco, A., Renna, T., and Clappier, R., “Minia4. 5. 6. 7. 8. 9.
ture Pulse Tube Cryocooler for Space Applications,” Cryocoolers 11, Kluwer Academic/Plenum Publishers (2001), pp.145-154. Liang, J., Zhou, Y., Zhu, W., Sun, W., Yang, J., Li, S., “Study on miniature pulse tube cryocooler for space application,” Cryogenics, vol. 40 (2000), pp.229-233. Liang, J., “Thermodynamic cycles in oscillating flow regenerators,” Journal of Applied Physics, vol. 82 (1997), No.9, pp.4159-4156. Liang, J., Ravex, A., and Rolland, P., “Study on pulse tube refrigeration. Part 1: Thermodynamic nonsymmetry effect,” Cryogenics, vol. 36 (1996), p.87-94. Liang, J., Ravex, A., and Rolland, P., “Study on pulse tube refrigeration. Part 2: Theoretical modeling,” Cryogenics vol. 36 (1996), p.95-100. Liang, J., Ravex, A., and Rolland, P., “Study on pulse tube refrigeration. Part 3: Experimental study,” Cryogenics, vol. 36 (1996), p. 101-107. Yang, L.W., Zhou , Y., Liang, J.T., Zhu, W.X., “Analytical study of the performance of pulse tube refrigerator with symmetry-nozzle,” Cryogenics, vol.39 (1999), pp.723-727.
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Performance and System Design of 60K Pulse Tube Coolers Driven by a Linear Compressor for HTS Filter Subsystems Y. L. Ju, K. Yuan, Y. K. Hou, W. Jing, J. T. Liang and Y. Zhou Cryogenic Laboratory, Technical Institute of Physics and Chemistry Chinese Academy of Sciences, Beijing 100080, China
ABSTRACT We report here on a performance study and the system design of a 60 K pulse tube cryocooler driven by a linear compressor. The goal of the development was to achieve the design of a fully integrated, cryogen-free HTS RF filter subsystem for wireless telecommunication. Two different single-stage pulse tube cold finger geometries were designed and analyzed—a U-type and a coaxial. The objective was for 3.5 W of cooling capacity at 65 K with a specific power of 30 W/W. Based on quantitative optimization, the U-type cold finger is predicted to provide 3.5 W at 65 K with a P-V input power of 89 W. The corresponding COP and the specific power are about 3.8% and 25W/W, respectively. In contrast, the coaxial cold finger operated at the same conditions is predicted to require 112 W of P-V power for the same cooling capacity. The corresponding COP and the specific power are about 3.1% and 32W/W, respectively. To improve the overall system performance, we also designed and analyzed a two-stage coaxial pulse tube cooler, driven by the same compressor, to provide two separate temperature stages, each optimized for a different element of the load. This cooler provides 0.5W at 65 K at the 2ndstage while simultaneously providing 3.0 W on the lst-stage cold head at 90K; for this performance the predicted P-V work is 73 W, and the combined COP is 4.8%. The construction of the cooled HTS filter subsystems integrated onto the cold fingers of both the single-stage and two-stage coolers is also described.
INTRODUCTION With the worldwide application of superconducting electronic (SCE) devices, cooling with easy, reliable and compact cryocoolers is highly desirable.1,2 GM and Stirling coolers are reliable machines and are regarded as mature technology and widely used in many areas. However, they have a moving piston at the cold head that inevitably causes mechanical vibration and electromagnetic interference (EMI). The decreased reliability and high-cost of the sliding seals also remains as a severe problem for many applications. In the past decade progress in the development of pulse tube coolers (PTCs) has been impressive. Many new ideas, incorporated with refined thermo-mechanical designs and fabrication approaches, have led to thermal efficiencies as high as Stirling coolers.3,4 The PTC has no moving Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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mechanical parts and no displacer seals in the cold head, so the mechanical vibration and EMI noise can be reduced to negligible levels with higher reliability, longer lifetime, and lower cost than other coolers. All these advantages provide a high degree of design flexibility that allows this cooler to replace Stirling coolers in many applications. It has been found that there is a long-term growth application of cryocoolers for HTS filter subsystems in wireless phone systems. The initial incentive for using HTS RF filters was mainly to provide a reduction in size. However, recently the perceived improvement includes overall performance, such as broader coverage, lower interference, and better quality of service. In order to improve voice transmission quality, the HTS filters are being integrated with an array of Low Noise Amplifiers (LNAs), which reduces the inter-channel EMI of the cellular phone system. The HTS (usually YBCO) RF filters are passive devices and must be cooled below their superconducting transition temperature (usually 65-80K) in order to operate properly. The HTS filters are usually connected through a coaxial cable to an array of LNAs, which are active devices and induce a few hundreds milliwatts of thermal load. In addition, the LNA array is connected to a feedthrough of the vacuum chamber by coaxial cable, which also introduces a few hundreds milliwatts of heat. The third contributor to the heat load is the thermal radiation from the vacuum jacket wall to the cooled surfaces. The heat load is estimated as hundreds of milliwatts. The signal-to-noise ratio of LNAs is improved significantly when cooled to a temperature of 90~110 K. Cooling is currently being performed either by GM coolers with separate compressors or by totally integrated Stirling coolers with linear compressors, both in ground-mounted and tower-top units. The mechanical vibration and EMI noise caused by the moving piston in the cold head of GM and Stirling coolers are a severe problem for the operation of high quality HTS devices and are the main technical obstacles against a more general acceptance of these machines. Acceleration measurements5 have shown that the vibration of PTCs is one order of magnitude smaller than that of GM and Stirling coolers. Fig. 1 presents the frequency responses of a HTS (YBCO thin film deposited on gem substrate) microwave filter cooled by a G-M type coaxial PTC at different temperatures.6 The central frequency moves gradually to lower frequencies with increasing temperatures. The shift of the central frequency becomes obvious at temperature above 75 K, but the bandwidth and insertion loss are nearly constant, except at temperatures up to 86 K. It has been demonstrated that a properly designed PTC has great advantage and is capable of low-noise cooling of highly sensitive HTS devices. This paper gives the performance and system design of a 60K PTC driven by a commercial dual-opposed-piston linear compressor with a swept volume of l0cc. The coolers are specially designed for the purpose of developing fully integrated, cryogen-free HTS filter subsystems. The design objective is for 3.5W of cooling capacity at 65K with a specific power (ratio of PV-work to cooling capacity) of 30W/W.
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THEORETICAL OPTIMIZATION A comprehensive computer model7 has been applied for quantitative analysis and performance optimization and as a guide in the early stage of cooler system design. The model is a 1-D, unsteady compressible flow numerical model that is based on a mixed Eulerian-Lagrangian method developed by the present author. The model is established and updated from the historical developments of finite difference methodology (FDM)8-10 for calculating the time-variations of dynamic parameters and internal processes occurring in a pulse tube cooler. Our design approach, whereby we apply the computer simulation program to the PTCs, involves three stages. The first stage of design consideration is the geometrical arrangement of the pulse tube and regenerator. There are three different arrangements, in-line, U type and coaxial (concentric) type. Obviously, the in-line arrangement has the highest efficiency, as high as 24% of Carnot3, since it avoids losses from curved gas flow and dead volume at the junction between the pulse tube and the regenerator. However, the location of the cold head in the middle between the regenerator and the pulse tube is a disadvantage for connecting to cooled devices. The most compact and convenient geometry for practical application, just like Stirling coolers, is the coaxial type pulse tube. It can replace a Stirling cooler without any change to the dewar or to the connection to the cooled devices. However, this design has several inferior elements that degrade its efficiency. These include the mismatch of temperature profiles between the regenerator and pulse tube, the void space at the cold end, and the reversal of gas flow direction in the cold end space. Some of these issues have been minimized by various techniques, i.e. multi-bypass, symmetric nozzle, inertance tube, low thermal conductivity materials, etc. When the geometrical arrangement of the PTC has been determined, the second stage of computation is to optimize the cooler system with respect to the dimensional layout of the pulse tube, regenerator, cold and hot end heat exchangers, and inertance tube based on the swept volume and input power of the compressor. First of all, the cooler volume must be adjusted to match the swept volume of the compressor in order to gain the proper pressure amplitude in the cold finger. Secondly, the volume ratio of the regenerator to pulse tube must be optimized to achieve a load balance. Thirdly, the arrangement of the regenerator matrix must be optimized based on the gas and matrix temperature profiles along the regenerator to reach a low no-load temperature and high cooling efficiency. Finally, the heat exchange surface area per unit volume of the hot and cold heat exchangers must be maximized to improve the thermal conductivity between the cold end and the cooled SCE devices, and between the hot ends and the external heatsink environment. The third stage of computations is directed at choosing the optimal system mean pressure and operating frequency to maximize the cooling capacity and minimize the power consumption. This is done for the same effective cooling capacity, the optimal size of each component, and the optimal opening condition of the orifice, double-inlet, multi-bypass and inertance tube. Using our computer code, quantitative analyses were conducted of two single-stage PTC cold finger geometries (the U-type and the coaxial). This allowed for the evaluation of the power consumption and cooling capacity, and the main system and operating parameters, like the size of each component, volume ratio, arrangement of the regenerator matrix, the charge pressure, pressure ratio, and operating frequency. Figure 2 shows the predicted performance of the U-type PTC as a function of the volume ratios of the regenerator volume and pulse tube volume to the 10cc compressor volume the average pressure is 3.0 MPa and the drive frequency is 50Hz. The optimal design point of the regenerator and pulse tube is where the COP (cooling capacity at 65 K divided by power consumption, %) is maximized. Figure 2 shows that the optimum ratios of the regenerator and pulse tube volumes to the compressor volume are in the range of 1.0-1.2 and 0.5-0.7, respectively. At its most efficient operating point, the U-type cold finger can provide 3.5W of cooling capacity at 65 K with a specific P-V power of 89 W. The corresponding COP and the specific power are about 3.8% and 25 W/W, respectively. In contrast, the coaxial cold finger operated at the same conditions requires 112 W P-V work to get the same cooling capacity of 3.5W at 65K. The corresponding COP and the specific power are about 3.1% and 32 W/W, respectively, which is slightly lower than the design goal.
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As a lower power alternative to the single-stage design, we proposed and analyzed a two-stage coaxial pulse tube cooler driven by the same linear compressor unit. Table 1 presents the optimization highlights of the main parameters of the single (U-type and coaxial type) and two-stage (coaxial) cold finger configurations. The other parameters (operating and geometric) of the coolers were the same during computer optimization for both cold finger configurations. Table 2 gives the general output parameters after optimization by the computer simulation program at an average pressure of 3.0 MPa and a frequency of 50 Hz. The predicted results show that the two-stage coaxial cold finger can provide a cooling capacity of 0.5 W at 65 K on the 2nd-stage cold head with a simultaneous load of 3.0W on the lst-stage cold head at 90K; the required P-V power is 72 W. The combined COP (sum of 1st and 2 nd stage's effective cooling capacities divided by power consumption, %) and the specific power are improved to 4.8%
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and 21 W/W, respectively. Examination of the general optimization parameters in Table 2 demonstrates that, with the use of the two-stage cold finger instead of the single-stage option, the specific PV power is reduced about 35%, and the heat generated by the compressor is reduced about 30%. In addition to the cooler itself, there are still two problems corresponding to the overall system efficiency: the compressor, and the compressor power and control electronics. For our PTCs we use a commercial double opposed piston linear compressor (Leybold Polar), which is driven by a POLAR DRIVE control unit with specific electronic and control components; this can be easily integrated into the cooler system.
DESCRIPTION OF SYSTEM DESIGN Based on the quantitative optimization of cooler performance as guided by the computer simulations and the above technology developments, two different geometries of cold fingers have been proposed and designed. Both geometries are coaxial pulse tubes, single-stage (type 1) and twostage (type 2), designed to integrate easily into a cryogen-free HTS RF filter subsystem for a wireless telecommunications base station. Figure 3 shows a schematic diagram of the Stirling-type single-stage coaxial PTC. It consists of the following key components: (1) the commercial twin-piston linear motor-driven compressor (Leybold Polar), (2) the flexible connection tube, (3) the integrated gas buffer in combination with one or more flow impedances—symmetric nozzles as orifice and double-inlet for gaining proper phase shift between the gas mass flow and pressure wave at the warm end of pulse tube, (4) the hot end flange, which is also a gas flow control unit on which the hot ends of the pulse tube and regenerator are mounted, (5) the regenerator, (6) the pulse tube, and (7) the cold head. The regenerator is made of thin-walled stainless steel tube with a thickness of 0.15mm and is filled with 400-mesh stainless steel screen as the regenerator matrix. The pulse tube is made of Teflon with a wall thickness of 0.5mm and is placed within the annular regenerator. The copper cold head consists of a cold end heat exchanger with a copper base (platform) on which the superconducting electronic devices can be easily coupled. This design is very compact and could be made into a commercially available pulse tube cooler. Besides the cooler itself, it is also essential to design and provide an entire cryopackage. This cryopackage provides the high degree of integration (thermally, mechanically and electronically) needed between the cold head of the cooler and HTS RF filters and LNAs within an evacuated chamber. Also included are the electrical and/or optical connections through the enclosure and vacuum space. The cryopackage should be reliable, provide adequate integration with the HTS RF filters and LNAs, and be easy to fabricate. The approaches are addressed as follows.
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Figure 3 also illustrates the construction of the cooled HTS filters and their integration with the cold finger of the single-stage PTC. This single-stage configuration is arranged to operate at a temperature of about 65K in order to meet the cooling demands of the HTS filters. The key system components include the array of LNAs (14), RF filters (13) and radiation shield (12) that are thermally mounted on the copper base platform (8), which is anchored to the single-stage cold head (7) of the cooler through thin indium foils with highly thermal conductivity. The whole construction is placed in a vacuum chamber (11), with coaxial cables (9,10) providing electrical interconnection between the RF filters, LNA array, and vacuum feedthroughs. The advantage of such a single-stage design is the simple construction of the cooler and HTS filter subsystem. The disadvantage is the high power required to operate the cooler and cool the radiation shield and the LNA array to a temperature lower than necessary. Unfortunately, these components contribute most of the cooling load. In addition, the cool-down time prior to reaching the operating temperature of the filters is delayed by the process of extracting heat from the LNAs and the radiation shield. In order to overcome the disadvantages of the above single-stage cooler design, the HTS filter subsystem has also been integrated with a two-stage coaxial PTC, as shown in Fig.4. This system is equipped with the same linear compressor unit, but has two cold heads operated at different cryogenic temperatures. The construction allows simultaneous cooling of the HTS RF filters and the LNA array to two different cryogenic temperatures, each tailored for their specific operation. The RF filters (13) are mounted on a copper base platform thermally coupled to the second-stage cold head (15), which is set to operate at a temperature of about 65 K in order to meet the cooling demands of the RF filters. Both the LNA array (14) and the radiation shield (12) are mounted on another copper base platform (8) thermally coupled with the first-stage cold head (7), which is arranged to operate around 90 K. This design of the cryopackage allows the HTS RF filters and LNA array to operate at their optimum temperatures, resulting in reduced heat load and power consumption and increased cooler reliability. In addition, the two-stage configuration is capable of accommodating larger HTS devices and their associated electronic circuits, and the HTS filter subsystem construction is more rigid than the single stage option. It should be pointed out that two specific challenges need more attention: (1) ensuring low losses in the electrical connections without allowing too much heat to transfer from room temperature to the HTS RF filters and the LNA array, and (2) achieving a robust vacuum chamber capable of maintaining the high vacuum required for good thermal insulation over the lifetime of the coolers.
CONCLUSIONS In this paper we presented a performance study and system design for a 60 K pulse tube cooler for the purpose of developing a fully integrated, cryogen-free HTS RF filter subsystem for wireless telecommunication. Two different cold finger geometries of single-stage and two-stage pulse tube coolers were analyzed and designed. The construction of the cooled HTS filter subsystem integrated on the cold finger of the coolers was also described. Comparison of general optimization parameters indicates that the specific PV power of the cooler is reduced about 35% and the heat generated by the compressor is decreased about 30% by using the two-stage configuration rather than the single-stage option. The two-stage cold finger allows the operation of the HTS RF filters and LNA array at two different temperatures, each selected to optimize their specific performance. This also results in reduced heat dissipation by the cooler and increased cooler reliability. In addition, the two-stage geometry is capable of accommodating large HTS devices and their associated electronic circuits, and the HTS filter subsystem construction is more rigid than the single stage option.
ACKNOWLEDGMENT The work is funded by the National Natural Science Foundation (Grant No. 50176052)
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Braginski, A.I., “Superconducting electronics coming to Market,” IEEE Trans. Appl. Supercond. vol. 9 (1999), pp. 91-102.
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Martin, J. L. et al. “Design consideration for industrial cryocoolers,” Cryocoolers 10, Kluwer Academic/Plenum Publishers, New York (1999), pp. 181 -189.
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Tward, E., et al. “High efficiency pulse tube cooler,” Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2001), pp. 163-167.
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Marquardt, E. D. and Radebaugh, R., “Pulse tube oxygen liquefier,” Advance in Cryogenic Engineering 45, Plenum, New York (2000), pp. 629-635.
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H. Li et al., “Demonstration of HTS microwave sub-systems with a pulse tube cryocooler,” Physica C, vol. 282-287, (1997) pp. 2527-2528.
6.
K. Yuan, “Experimental study and optimization of low frequency pulse tube cryocoolers at liquid nitrogen temperatures and their application of HTS filters,” MS thesis, Chinese Academy of Sciences, 2002.
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Ju, Y.L., “Computational study of a 4K two-stage pulse tube cooler with mixed Eulerian-Lagrangian method,” Cryogenics, vol. 41, (2001), pp. 49-55.
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Wang, C. et al., “Numerical analysis of a double-inlet pulse tube refrigerator,” Cryogenics, vol. 33, (1993), pp.526-560.
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Ju, Y.L. et al., “Numerical simulation and experimental verification of the oscillating flow in pulse tube cryocooler,” Cryogenics, vol. 38, (1998), pp.169-176.
10. Ju, Y.L. et al., “Dynamic simulation of the oscillating flow with porous media in a pulse tube cryocooler,” Numerical Heat Transfer, Part A, vol. 33, (1998), pp.763-772.
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High Capacity Pulse Tube Cryocooler I. Charles, J.M. Duval*and L.Duband (1) T. Trollier and A. Ravex (2) J.Y. Martin (3)
(1) CEA, Service des Basses Températures Grenoble, France (2) Air Liquide Advanced Technology Division, AL/DTA Sassenage, France (3) THALES Cryogénie S.A. Blagnac, France * Present address : University of Wisconsin, Department of physics, Madison, Wisconsin, USA
ABSTRACT High capacity pulse tube coolers are presently under development at CEA/SBT. The Development Models of the pulse tube cold finger feature an in-line configuration to ease manufacturing and optimisation of the sizing. A cooling power of 5.2 W at 80 K has been achieved with 200 W of total electrical power, 128 W PV work, and a rejection temperature of 298 K provided by water cooling. This work has been accomplished using a standard wearing linear compressor operating at a fixed frequency of 50 Hz and equipped with a laser transducer to measure the stroke. An ultimate temperature of 38 K has also been achieved. Based on these results, a pulse tube cold finger has been developed for cooling down of HTS filters banks for an UMTS RF base station demonstrator (SUPRACOM project). A standard wearing linear compressor from THALES Cryogenics with an adjustable frequency drive electronics has been used. A performance of 7.7 W at 80 K has been achieved with 200 W of total electrical power and with heat rejection at 298 K; the PV work is about 140 W. This last result is presently being used for the design of a coaxial shape pulse tube cooler. This cold finger will be ultimately supplied with a flexure bearing compressor to increase the reliability. The matching of the compressor and the cold head are discussed.
INTRODUCTION SBT has been involved in the optimisation, design and manufacturing of pulse tube coolers for over 10 years. The needs have been focussed on low cooling capacity coolers (around one watt) for infrared detectors. The development of HTS filters for telecommunications requires coolers with higher cooling power in the 70-80 K temperature range. Since the operating conditions require high
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reliability, pulse tube coolers are a good candidate for this application thanks to the absence of any moving parts in the cold finger. In the framework of a PhD thesis, work on a high-heat-lift pulse tube cooler was initiated using a fixed frequency compressor that allowed a predesign of the cold finger. Further improvements have be achieved by changing the frequency and fill pressure. Finally, measurements of relevant parameters have been performed in order to redesign for a compressor equipped with flexure bearing technology.
TEST SET UP AND OPERATING CONDITIONS Several prototypes have been designed, manufactured, and tested to evaluate the impact of the size of both the regenerator and the pulsation tube on the performance. The in-line configuration was chosen as it allows easy changing of the tube and regenerator lengths. Different diameters and lengths of the regenerator and pulsation tube have been tested. All tests have been performed in the inertance mode1,2, hence the possibility of DC flow3,4 is suppressed. It is foreseen that repeatability trouble on serial devices will be solved using the inertance mode. First, these in-line pulse tube cold fingers were tested with a gas bearing dual linear compressor. This compressor allows a maximum fill pressure of 20 bars and a maximum electrical power of 200 W. The linear motors were driven by a simple adjustable voltage source directly connected to the electrical network (50 Hz). A transfer line of 330 mm was used to connect the compressor to the cold finger. This line, the compressor, the regenerator hot flange and the tube hot end were cooled down by water. The pulse tube and experimental setup are shown in Figures 1 and 2.
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PERFORMANCE OBTAINED A mapping of the optimal performance achieved with the twelve experimental prototypes is plotted in Fig. 3. Each prototype differs from the other with slight changes in the dimensions (regenerator and/or tube, diameter and length), leading to various optimal settings of the inertance. The slope represents the slope of cooling power versus cold temperature. This slope is quite constant in the range 40 - 80 K. It gives the temperature elevation for one watt applied load. The mapping is performed with 298 K heat rejection temperature on the regenerator flange and 200 W of electrical power. The pulse tube cold finger is in a horizontal position without MLI.
Effect of the regenerator length The effect of the regenerator length on the performance was studied. Figure 4 represents the evolution of the ultimate temperature and slope. The data presented have been obtained with the same regenerator diameter and with the same pulsation tube. Decreasing of the regenerator length leads to a more powerful prototype (a lower slope). This can be explained by a reduction of the pressure drop, which allows a better expansion ratio into the pulsation tube. This pressure ratio is directly coupled with the net cooling power produced at the cold end. The effect on the ultimate temperature is less evident. The expected improvement of the no load temperature by increasing the regenerator length was not found. This can be due to the fact that the present regenerator is large enough to offer very good efficiency. The no load temperature is then driven by the pressure drop.
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Effect of the Tube Length A study of the tube length shows that there is a minimum length required to achieved good performance. Three curves are displayed on Fig. 5. The two solid ones correspond to the same tube diameter, but with different regenerator lengths; the dashed one is obtained with a different size tube and regenerator. All the points reported are obtained with an optimised inertance. No degradation of performance was found with a length increase. This shows that in our range of experimentation, the tube capacity effect could be controlled by the right inertance to achieve the required phase shift. Parasitic heat losses, such as tube conduction and wall heat exchange, are negligible. It can be concluded from these results that a double inlet is not needed for phase shift optimisation.
Estimation of the PV work Stroke and pressure measurements can give an estimation of the work given to the gas. This work is important to the determination of the real efficiency of the cold finger. One part of the compressor was equipped with a laser sensor to measure the stroke of the piston. A hole was made on the back of the compressor and a plexiglass window with an elastomer seal was used to ensure gas tightness. In use, the laser is reflected on the back of the piston and allows the determination of the position of the piston. The compressor output is equipped with a dynamic pressure sensor that allows an estimation of the PV work given to the gas. To achieve a precise estimation it is necessary to use a sensor with sufficient time response; for the laser and pressure sensor the frequency response is 915 Hz and 10 kHz, respectively. For operating conditions at 50 Hz, this is enough to have good accuracy. The PV work measured this way was compared with the total electrical power minus the Joule losses in the coils. An example, representative of the measurement, is reported in Figs. 6 and 7. The PV work calculated from the pressure and stroke measurements was found to be 125 W. This PV work corresponds to a stroke of 3.68 mm with a piston cross-section of a pressure variations of 1.15 bar, and with a phase shift between pressure and displacement of 60°. The electrical power was 200 W for a current in the coils of 10.22A, which leads to a equal to 137W. The PV work measured corresponds to 91 % of the This difference could be explained by other compressor losses such as piston rubbing, piston leakage, or eddy currents. This difference could also be explained by the position of the pressure sensor. It is mounted on the transfer line and does not measure the real pressure swing in the compression chamber. Some pressure drop in the compressor output is certainly responsible for the under evaluation of the PV work. Nevertheless, is a good representation of the PV work and it will be used in further work as it is easier to measure.
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Best Results The best prototypes were tested in the vertical position to avoid any convection problems5,6 and were wrapped with MLI to reduce radiation losses. For these operating conditions, the maximum cooling power obtained at 80 K was 5.2 W for 200 W of electrical power. The PV work is estimated to be 130 W. The efficiency of the compressor is 0.65 %. The total specific power at 80 K of the whole system is 38.5 W/W. The pulse tube cold finger has a specific power of 25 W/W. The compressor efficiency is good even if the system is not working exactly at the resonance frequency. The same pulse tube with a longer regenerator has given an ultimate temperature of 38 K. The goal of this study was focused on the cooling power at 80 K, so lower ultimate temperature could be obtained with a specific design.
SUPRACOM PROJECT The new mobile phone generation under development (UMTS) will require better performance for the reception base station. The French Ministry of Industry funded a project to develop a HTS filter. This filter should allow better filtering and lower noise, thus leading to an extended reception area. The cooling specification is around 4 to 5 watts at 80 K for the six filters of the base station. Based on the preliminary study, the in line pulse tube was adapted to a new compressor. This compressor is a linear, wearing compressor from Thalès, which allows fill pressures of up to 35 bars. It is driven by an adjustable frequency set of drive electronics. These operating conditions allowed the optimisation work performed during this PhD research to be extended.
Optimisation with Adjustable Frequency The curves in Fig. 8 represent the no load temperature and ratio of the Joule losses to the total electrical power. These points are obtained with a fixed electrical power of 150 W. The compressor efficiency is better when the Joule losses are minimum, i.e. for a frequency of 45 Hz. The pulse tube cold finger has its lowest ultimate temperature around 50 Hz. The best overall efficiency of the system is obtained when the two optimums match in frequency.
Effect of Fill Pressure In order to achieve a real optimisation of the cold finger and avoid any interaction with the compressor resonance, it is best to perform the test with PV work held constant. It has been shown that this PV work is close to the total electrical power minus the Joule losses. The effect of the filling pressure has been investigated with a constant of 100 W. The results obtained are presented in the Fig. 9. The effect on the cooling power is important up to 30 bars, but is less clear for higher pressures. The measurements done at 30 bars lead to a specific power of 17.7 WAV at 80 K.
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Best Performance In order to achieve the best performance, the main part of the transfer line was removed. Only the valve used to isolate the compressor from the cold finger remained in place. Fans have been added to compensate for the water cooled transfer line. Cold temperature measurements performed with no load and with a 7 W load are given in Fig. 10. A cooling power of 7 W at 77 K was obtained, which leads to an estimated cooling power of 7.7 W at 80 K. The overall specific power of the cooler is 26 W/W. This value could be compared with the specific power of 34.6 W/W reported by L.W. Yang and al.6 and the specific power of 24 W/W reported by S-Y Kim and al.7.
FUTURE WORK AND PERSPECTIVE Based on the experimental results, a coaxial shape pulse tube cold finger has been designed and is under manufacturing (Fig. 11). The use of copper for isothermal locations and stainless steel for gradient areas has been implemented. The cold part is equipped with a support to fix the HTS filter. In parallel, a dual linear compressor is being designed by Thales to match the cold finger operating parameters. This compressor will be equipped with flexure bearing technology and moving magnets8. Motor coils are placed outside the gas cycle to increase reliability.
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CONCLUSIONS Extensive optimisation work performed on various in line pulse tube prototypes has demonstrated the capability to remove over 7 W of heat load at 80 K with 200 W of electrical power. A maximum specific power of 17.7 W/W was achieved for the cold finger. Based on this study a coaxial pulse tube cold finger has been designed and is under manufacturing. This cold finger, coupled with a flexure bearing compressor, could be used for cooling HTS filter banks in future mobile phone base stations.
ACKNOWLEDGMENT The work done on SUPRACOM was performed in partnership with THALES Cryogénic and was supported by the French Ministry of Industry. The PhD of Jean Marc Duval was co funded by Air Liquide and CEA.
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REFERENCES 1. S.W. Zhu, S.L. Zhou, N. Yoshimura and Y. Matsubara, “Phase shift effect of the long neck tube for pulse tube refrigerator,” Cryocoolers 9, Plenum Press, New York (1997), pp. 269-278 2. D.L. Gardner and G.W. Swift, “Use of inertance in orifice pulse tube refrigerators,” Cryogenics, vol. 37 (1997), pp. 117-121. 3. D. Gedeon, “DC gas flows in Stirling and pulse tube refrigerators,” Cryocoolers 9, Plenum Press, New York (1997), pp 385-392. 4. L. Duband, et al., “Experimental results on inertance and permanent flow in pulse tube coolers,” Cryocoolers 10, Plenum Press, New York (1999), pp 281-290. 5. G. Thummes, M. Schreiber, R. Landgraf and C. Heiden, “Convective heat losses in pulse tube coolers: effect of pulse tube inclination,” Cryocoolers 9, Plenum Press, New York (1997), pp. 385 6. L.W. Yang and G. Thummes, “Medium-size pulse tube coolers with linear compressor,” Adv. in Cryogenic Engineering, Vol. 47, Amer. Institute of Physics, Melville, NY (2002). 7. S-Y Kim, W-S Chung, J-J Park, D-K Hwang, H-K Lee, “Reliability test results of LGE pulse tube cryoccoler,” Adv. in Cryogenic Engineering, Vol. 47, Amer. Institute of Physics, Melville, NY (2002). 8. M. Meijers, A. A. J. Benschop and J. C. Mullié, “High Reliability Coolers under development at Signaal-USFA,” Cryocoolers 11, Kluwer Academic / Plenum Publishers, New York (2001), pp 111118.
Development of Single and Two-Stage Pulse Tube Cryocoolers with Commercial Linear Compressors K. B. Wilson Sunpower, Inc. Athens, OH 45701 D. R. Gedeon Gedeon Associates Athens, OH 45701
ABSTRACT Sunpower, Inc. and Gedeon Associates partnered to develop single and two-stage pulse tube cryocoolers under Small Business Innovative Research (SBIR) funding from NASA Goddard Space Flight Center. The development centered around using the high-efficiency, low-cost linear compressor that Sunpower has already taken to the manufacturing level with its M87 Stirling cryocooler. The primary goal was to demonstrate efficient cooling at 77K with a single-stage pulse tube cold head. Both in-line and u-tube configurations of single-stage pulse tube cold heads were fabricated and tested. Cooling at 30 K with a two-stage cold head was a secondary goal undertaken to gain experience with multi-staging. The second stage was designed as an add-on component to the single-stage u-tube configuration. All designs implemented inertance tubes as acoustical tuning devices. The ultimate objective was to establish the commercial potential of linear-compressor driven pulse tube cryocoolers and whether they could be cost competitive with the Sunpower M87 Stirling cryocooler. High efficiency was demonstrated with the single-stage pulse tube cooler achieving 5.2 W of heat lift at 77K with 100 W electrical input power, and the two-stage pulse tube cooler lifting 260mW at 30K, again with 100 W electrical input.
INTRODUCTION Sunpower, Inc. has been developing Stirling cycle cryocoolers for over ten years and has recently introduced the commercial M87 model with its characteristics of low cost, light weight, long life and high efficiency. The M87 makes use of a linear compressor, which is well suited to driving a pulse tube cryocooler. Sunpower’s current factory is capable of producing 2000 units per month of the compressor shown in Figure 1. Gedeon Associates has been developing modeling and optimization software for eighteen years and eight years ago introduced the commercial Sage software. Sage employs a graphical interface Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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that allows the user to assemble a single or multi-stage pulse tube cryocooler model from interconnected component pieces, then to optimize that model interactively. Sunpower and Gedeon Associates combined their resources in an SBIR program funded by NASA Goddard Space Flight Center. The goal of the research was to build and test single and twostage pulse tube cryocoolers (PTC’s) using existing Sunpower linear-compressor technology. The single-stage plan was to build a u-tube configuration after gaining experience with an in-line configuration. The two-stage plan was to design a second-stage “add-on” assembly that could mount to the turning manifold of the single-stage u-tube configuration. Each build was to incorporate inertance-tubes functioning as acoustic tuning devices between the pulse tube and reservoir volume. Performance for all PTC’s was to be demonstrated at 100W electrical input power. The ultimate objective was to establish the commercial potential of linear-compressor driven pulse tube cryocoolers and whether they could be cost competitive with the Sunpower M87 Stirling cryocooler.
CRYOCOOLER CONSTRUCTION The use of a commercial linear compressor as the driver for a pulse tube cold head was a key approach in this development. Sunpower has taken the art of linear compressors to the production level due to its highly efficient yet low-cost and manufacturable design1. The pulse tube cold head design owes much to the efforts of previous researchers. For example, the pulse tube component is tapered to reduce acoustic streaming, based on a streaming formulation developed by Olson and Swift 2 . The decision to use an inertance tube as the phase shifter as opposed to an orifice or double inlet pulse tube was influenced by previous studies such Zhu, et. al.3 The regenerator matrix was stainless steel random fiber, inexpensive and commercially available. The design of the turning manifold/cold flange of the u-tube configuration is an innovative arrangement to allow ease of manufacturing, without excessive braze joints. The assembly process also allows for convenient quality control checks before the cold head is hermetically sealed. The turning manifold itself was aerodynamically designed to reduce convection cells in the pulse tube without introducing flow straighteners that affect thermodynamic performance. Patents are being pursued for the design of the turning manifold itself, along with construction of the cold head. The two-stage PTC uses the same design fundamentals as the single-stage u-tube configuration. The second-stage regenerator is random fiber stainless steel, the turning manifold/cold flange design is innovative as described above, and an inertance tube between the tapered pulse tube and reservoir volume tunes the phase shift between pressure and gas displacement. The second stage was not fully optimized for performance, it was simply meant as an add-on component to gain experience in multi-staging.
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TESTING Component Testing Since our modeling tool for this development effort was the Sage software, we performed a series of individual component and subassembly tests to calibrate simulated results with actual experimental results. This effort of component testing included the following: steady-flow pressure-drop tests of regenerators and inertance tubes; compressor characterization tests; regenerator single-blow heat-transfer tests; and testing of a subassembly consisting of a compressor, inertancetube, and reservoir. We generally found component test results to agree well with Sage modeling predictions.
In-line PTC Testing The achieved load curve for the in-line configuration is given in Figure 2. The test setup of this configuration is shown in Figure 3. For testing, pulse tube orientation was cold-end-down, as shown, and the cold parts were wrapped in radiation shielding within a vacuum Dewar. Table 1 compares test results for the in-line unit against the predictions of Sage. Note that the 77K point in
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the load curve does not match the 77K performance listed in the table. This is because this curve was recorded at a different time than the data point reported in the table, and all other components in the in-line PTC assembly were not necessarily the same. Table 1 shows that Sage predicts cooling power quite closely but under-predicts PV power somewhat. Predicted pressure phasors are also off slightly. The errors could be due to inertance tube coiling effects or turbulence-transition effects. The cooler was tested in various orientations to gage the free-convection loss in horizontal or cold-end-up orientations, with the results shown in Table 2. The pulse tube is the only component likely to be affected by gravitational orientation. The cold-end-down column of Table 2 corresponds to the run presented in Table 1. The horizontal and cold-end-up results were produced by repositioning the entire test rig with all other conditions the same, including radiation shielding. Regarding horizontal pulse tube operation, a simplistic argument suggests that it should have no effect on free convection due to the lack of any horizontal component of the gravitational field. However, this argument neglects any convective cells brought about by density gradients that exist
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transverse to the pulse tube axis (transverse direction now being vertical). In fact, early on we observed slanted frost patterns in our in-line unit running in a horizontal orientation, outside its vacuum Dewar. These patterns strongly suggested the presence of some sort of free-convection cell. So horizontal pulse tube operation should only be viewed as a first approximation to operation in a weightless environment.
U-tube PTC Testing The u-tube cold head construction is shown in Figure 4. Other than the construction of the cold head, the u-tube test setup was the same as the in-line test setup. A load curve for the u-tube configuration is given in Figure 5. Again, the test orientation is cold-end-down and the cold head
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is radiation shielded. Table 3 shows the dependence of the u-tube configuration on pulse tube orientation. It is qualitatively similar to that of the in-line configuration. Performance is best with pulse tube cold-end-down and worst with cold-end-up. Horizontal orientation falls somewhere in between. But quantitatively, the changes are more dramatic. For the u-tube cold head in cold-endup orientation the cooling power at 77 K drops off a full 3.7 W, compared to only 1.6 W for the inline configuration. And the change from cold-end-down to horizontal is 1.1 W, compared to 0.4 W for the in-line configuration. In the u-tube configuration, the turning manifold tends to introduce flow eddies at the pulse tube entrance due to flow separation at the inner bend and an abrupt change in flow area. These eddies propagate along the pulse tube axis creating a thermal convection loss. The pulse tube density gradient, in cold-end-down orientation, reduces the penetration of these flow eddies and reduces the thermal loss. We developed a formula to quantify the buoyant stabilization effect of cold-end-down orientation on the eddy convection loss. The key idea is the following observation: a cold, dense flow eddy introduced at the bottom of the pulse tube must have a minimum kinetic energy in order to overcome the potential energy burden of traveling all the way to the top, through gas layers of diminishing density. We have termed the eddy velocity corresponding to this minimum kinetic energy the eddy escape velocity which is related to the gravitational field strength g, vertical pulse tube height h, cold-end density and mean density by
In the cold-end-up orientation density gradients tend to accelerate flow eddies introduced at the cold end, leading to increased thermal loss. This accounts for the greater sensitivity to pulse tube orientation in the u-tube configuration compared to the in-line configuration. So what might performance be without buoyant instabilities, as when operating in a weightless environment? It is possible to get some idea of the likely performance under these conditions by isolating the gravitational stabilization component of the manifold-induced pulse tube convection loss. We concluded that this gravitational-stabilization component was equal to the performance change from cold-end-down to horizontal orientations (1.1 W in present case), less the observed 0.4 W horizontal free convection in the in-line configuration. This would suggest that gravitational stabilization is worth about 0.7 W (1.1 – 0.4) for the u-tube configuration tested. So we would expect the load curve in weightless operation to be shifted down by about 0.7 W, roughly.
Two-Stage PTC Testing A secondary goal of this program was to gain experience in multi-staging as a stepping stone to get to cooling below 10K. We designed the second-stage cold head simply as an add-on component to the single-stage u-tube PTC. Figure 4 shows the two-stage PTC cold head construction. Again, the cooler was not optimized for cooling in the 20–30K range. However, the results shown below are very promising for our continuing work to achieve efficient cooling below 10 K. Due to constraints on time and its secondary emphasis under our contract, we did not complete exhaustive testing on the two-stage cold head. We installed temperature-measurement diodes on both first and second cold stages and a resistance heater on the second stage. However, we never
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installed a resistance heater on the first stage, so all of our measurements were made with the firststage unloaded. Because the two-stage cold head took several hours to reach equilibrium (after each change in second-stage cooling load), the following load curve is just a linear interpolation between only two points: the no-load minimum-temperature condition and the cooling power at the design point temperature of 30 K. Both points were logged with 100 W electrical input. The main reason for the long time required to reach equilibrium was that there was no way to stop the slow, asymptotic decline in temperature of the thermally massive first-stage components, including the inertance-tube and reservoir for the second-stage pulse tube. A first-stage heater would have helped greatly here, by offering a way to independently stop any slow decline in the first-stage temperature. Figure 6 shows a no-load temperature of 24.4 K at 100 W compressor electrical input. The first-stage temperatures corresponding to the curve ranges from 76 K for the no-load second-stage point to about 80 K for the 260 mW point. The 80 K temperature is an extrapolation because the temperature diode on the first stage failed before the temperature reached equilibrium. Table 4 shows how Sage modeling predictions compare to the experimental test closest to the design point. Agreement is reasonably good, all things considered. The measured first-stage temperature is colder than the design point, but its heat load is zero. Adding the design point heat load of 0.5 W would increase the measured temperature somewhat, but probably not to 95 K. The measured second-stage heat load at 30 K is lower than the design point, but it is at least in the ballpark. Any number of reasons could account for the discrepancy, including imperfections in the second-stage regenerator and unaccounted-for heat loads. Also, the above Sage predictions do not include estimates of the manifold-induced pulse tube loss.
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We also ran the two-stage cold head in horizontal and cold-end-up orientations. Actually, there were two horizontal orientations because there are differences due to the axial non-symmetry of our cold head design. Unfortunately we were never able to achieve the second-stage design point temperature of 30 K in either horizontal or cold-end-up orientation. Instead we measured the minimum achievable no-load temperature as a function of cold-head orientation. Table 5 shows the results of the orientation testing. Note that second stage temperature went up dramatically for anything other than cold-enddown orientation. Also, that the two horizontal orientations perform worse than the cold-end-up orientation. Horizontal orientation I was with the first-stage pulse tube above the regenerators (both regenerators share a common axis) and the second-stage pulse tube below the regenerators. Horizontal orientation II was with the two pulse tubes reversed with respect to the regenerators. There is a third possible orientation with both tubes and regenerators at the same level, but we were unable to rotate the cold head into that position with our current test stand. Based on single-stage testing, we were expecting cold-end-up orientation to perform the worst with the two horizontal orientations falling somewhere between cold-end-down and cold-end-up performance. We were also expecting less of a drastic change in performance. It appears that free convection in the second-stage pulse tube must have been rather large compared to the available cooling power. Also, convection in the first-stage pulse tube was larger than the available cooling power it was designed for (0.5 W at 95 K), with the result that the first-stage temperature warmed up. This tended to de-tune the second-stage inertance tube because its optimal length is directly related to the speed of sound in helium, which increases with increasing temperature. Therefore there is a tendency for the second-stage temperature to increase more than would be expected just from its pulse tube free-convection heat load.
COST ANALYSIS OF U-TUBE PTC COMPARED TO STIRLING M87 The single-stage u-tube PTC that was tested and reported above is very close to a productiondesign model. The major design changes that need to be implemented include: incorporation of a passive balance system, which will be adapted directly from the M87; integration of the reservoir to the compressor; and hermetic sealing of the unit. The brazed and welded joining technology we developed for the cold head appears quite promising for eventual mass production. A preliminary cost study was performed on a production design as represented at the left in Figure 7. The part cost of the single-stage pulse tube cooler is very similar, within a few percent, to that of the M87 Stirling cryocooler. Assembly costs should be similar to, if not lower than, that of the M87. At right in Figure 7 is a rendered representation of a production design two-stage PTC which has not been studied for cost. However, since the entire two-stage concept was based on the add-on component approach, the cost can be logically extrapolated from the single-stage.
CONCLUSIONS We conclude from our SBIR research that both single-stage and two-stage pulse tube cold heads can be driven with Sunpower’s linear compressor technology and that the resulting cryocool-
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ers could be commercially viable, depending on market needs. The u-tube single-stage pulse tube cooler we developed can achieve similar temperatures, cooling powers and efficiencies as the Sunpower M87 Stirling cooler. Its size and weight would be slightly larger than the M87 but manufacturing cost would be similar, possibly lower, with a single close-tolerance fit in the compressor. Theoretically, it would be more reliable in operation as well. The main difficulty of introducing a single-stage pulse tube cooler to market at this time is the large number of cryocoolers already available in this temperature range. This includes the Sunpower M87 cryocooler, where much time and effort has already been devoted to developing its factory production line. It may actually be easier to bring a linear-compressor driven multi-stage pulse tube cooler to market for cooling down around 20K. Multi-stage pulse tube coolers have certain fabrication and reliability advantages over existing competition, for example a multi-stage Stirling-cycle cooler with two close-tolerance moving displacer seals. We have recently begun work under additional funding from NASA Goddard Space Flight Center to develop a three-stage PTC to achieve temperatures below 10K. The only competing commercial cryocoolers in this temperature range are the Gifford-McMahon and related pulse tube coolers, which utilize oil-lubricated refrigeration compressors and rotary valves requiring annual maintenance. They are not well suited to cooling applications requiring compact, maintenance-free and long-life cryocoolers. Applications in this temperature range already exist to some extent in cooling superconducting magnets for MRI refrigerators and scientific instruments. Other applications are emerging in cooling low-temperature superconducting digital circuits, where the right cryocooler would likely be an important enabling technology. Our two-stage pulse tube technology has not yet matured, but we have taken an important step forward in demonstrating the feasibility of driving multi-stage pulse tube cold heads with high-frequency linear compressor technology.
REFERENCES 1. Linger, R.Z., “Linear Compressors for Clean and Specialty Gases,” 1998 International Compressor Engineering Conference at Purdue, Proceedings Volume 1 (1998).
2. Olson, J.R. and Swift, G.W., “Suppression of Acoustic Streaming in Tapered Pulse Tube,” Cryocoolers 10, Plenum Publishing Corp., New York (1999), pp. 307-313.
3. Zhu, S.W., Zhou, S.L., Voshimura, 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.
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Development of a 5 W at 80 K Stirling-Type Pulse Tube Cryocooler L.W. Yang1, N. Rolff2, G. Thummes1, and H.U. Häfner2 1
University of Giessen and TransMIT-Center for Adaptive Cryotechnology and Sensors, D-35392 Giessen, Germany 2 Leybold Vakuum GmbH, D-50968 Cologne, Germany
ABSTRACT Stirling-type pulse tube coolers (PTCs) promise higher reliability, lower mechanical vibrations, and lower manufacturing costs, as compared to conventional Stirling cold fingers. We report the state of development of pulse tube cold fingers that were designed for operation on an existing linear Stirling compressor (Leybold Polar SC 7). Two lab models, a U-shaped and an inline PTC, were designed, fabricated and optimized at Giessen University. So far, the U-shaped PTC has reached a minimum no-load temperature of 32.4 K, and cooling powers of 3 W at 60 K and 6 W at 80 K are available at an electrical input power of 200 W to the compressor. The inline version achieves an even higher cooling capacity of 8.1 W at 80 K with 200 W of input power. On the basis of the U-shaped cold head, Leybold has fabricated three engineering samples, which were then tested for reproducibility. In view of a compact cold head for practical applications, also the effect of buffer volume size on the cooling performance was investigated. The cooling capacities of all engineering models were found to be reproducible in the range 5.3 W ± 0.3 W at 80 K with 200 W of input power.
INTRODUCTION In comparison with the traditional regenerative cryocoolers such as the Stirling- and GiffordMcMahon-cryocooler, the pulse tube cryocooler (PTC) has the advantage of operating without a cold moving displacer. This feature leads to reduced mechanical vibrations, and is also expected to increase the reliability and to lower the manufacturing costs of the cold head. In recent years PTCs have experienced a rapid development with the goal to eventually replace Stirling- and GM-coolers in various applications of cryoelectronics and cryoelectrics.1 A growing market for highly reliable small cryocoolers, either Stirling-coolers or Stirling-type PTCs, can be expected in wireless communication, where the use of high-temperature superconducting devices (e.g. high-quality microwave receiver filters) is moving forward to commercial applications.2 The rapid development of PTCs in the past few years has led to efficiencies of Stirling-type PTCs that come close to or in case of pulse tube coolers for space applications3-4, even exceed that of commercial Stirling coolers.
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Here we report on the state of development of 5 W-class PTC cold heads, which were designed for operation on the linear compressor of the Leybold Polar SC7 Split-Stirling cooler. This compressor is equipped with a classical linear drive motor and it employs dual-opposed pistons for vibration reduction.2 Basic design of the laboratory versions of the cold head and their optimization has been performed at Giessen University/Center for Adaptive Cryotechnology. Some first test results for one of the PTC cold heads that was operated on the Polar compressor, have already been reported recently.5 On the basis of the U-shaped laboratory version of the cold head, three engineering samples were fabricated and tested at Leybold with regard to realization of a prototype for practical applications. This pulse tube cooler development is part of a government supported joint project on "High temperature superconductors and novel ceramics for future communication technology".
DEVELOPMENT OF PULSE TUBE COLD HEADS: LABORATORY VERSIONS Description of Coolers Some details on the design of the PTC cold heads has been given previously.5 For practical reasons we have chosen a U-shaped configuration for the first laboratory version, which allows easy access to the cold tip, as shown in Fig. 1. For basic investigations also an inline cold head was built, since this arrangement avoids losses from curved gas flow and from dead volume at the cold tip. Most of the high-performance space pulse tube coolers3-4 employ such an inline configuration, which has the disadvantage that the cold tip is located in the middle of the cold head, and thus thermal interfacing with the object to be cooled can be rather tricky. The two PTCs differ by the size of pulse tube and regenerator. For the U-shaped and inline version the pulse tube diameter is 9 mm and 12 mm, respectively. Adjustment of phase shift between pressure and mass flow oscillation is accomplished by a second-inlet bypass and a combination of reservoir and inertance6 tubes in series with a needle valve, as illustrated in Fig. 1. In the tests of the laboratory models a reservoir (buffer volume) of about was used. For control of DC flow7, the second-inlet bypass consists of an in-house made needle valve arrangement with adjustable flow symmetry. The length of the transfer line that connects compressor and cold head is about 30 cm. A copper tube for optional water cooling is wrapped and soldered around the transfer line and warm end of the cold head. It was found5 that in the present setup, water cooling provides a higher cooling performance than air cooling only for compressor input powers larger than 100 W. In future, by use of a properly designed air-cooled heat exchanger at the warm end it should be possible to operate the PTCs with only air cooling without a marked degradation in performance.
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All test results below were obtained at an operating frequency of 40 Hz, which was found to be close to the optimum frequency of these cold heads. The helium charging pressure was in the range between 20 and 30 bar.
Cooling Performance of U-shaped Cold Head Fig. 2 shows the cooling capacity of the U-shaped cold head at input powers of 200 W and 100 W to the Polar compressor. The data were obtained after optimizing the cooler either for minimum no-load temperature or for highest cooling power near 80 K. The cooler was operated in double-inlet mode. With 200 W of input power, optimization for lowest temperature yields a no-load temperature of 32.4 K, and cooling powers of 3 W at 60 K and 5.4 W at 80 K are available, as seen from Fig. 2. This cooler optimization would be appropriate for applications that require operation well below 60 K. Replacement of the present stainless steel pulse tube and regenerator tube by those of titanium alloy with low thermal conductance will reduce the heat conduction loss by about 0.4 W, and should then give an even lower no-load temperature. After optimization near 80 K, the cooler provides cooling powers of 6 W and 3 W at 80 K with input powers of 200 W and 100 W to the compressor, respectively. The corresponding coefficient of performance (COP) is 3 % at 80 K, which is about 80 % of the COP of the Polar SC7 Stirling cold finger.
Cooling Performance of Inline Cold Head Like the U-shaped PTC, our most recently fabricated inline PTC was designed for operation with the Polar compressor but has enlarged volumes of pulse tube and regenerator. The cooling performance of the inline cooler is shown in Fig. 3 for operation with and without second inlet, as indicated by full and open symbols. In comparison with the U-shaped PTC in Fig. 2, there is a significant improvement in cooling capacity at higher temperatures at the cost of an increased no-load temperature. In double-inlet mode, with 100 W, 200 W, and 250 W of input power, the lowest temperature is 52.5 K, 43.5 K, and 41.5 K, and a cooling power of 3.7 W, 8.1 W, and 9.7 W is available at 80 K, respectively. The COP at 80 K is now 3.7 % at 100 W and 4.1 % at 200 W input, as compared to 3 % for the U-shaped version. This improvement comes from the larger slope of the load lines in Fig. 3, as compared to that in Fig. 2 at the same input power. The second inlet is not very effective in increasing the cooling performance of the inline PTC, which is seen from the data obtained without double inlet in Fig. 3 (open symbols). This is in contrast to the U-shaped cold head, where previous tests5 revealed a considerable decrease of no-load temperature and corresponding increase of cooling power, when the second inlet was opened. The reason why the inline cold head operates quite effectively without double inlet is tentatively ascribed to the modified regenerator geometry, which results in a different phase shift in the system. From a practical point of view, operation of a PTC without the second-inlet bypass is preferable, because this avoids problems with DC flow and makes the system simpler and presumably more reliable.
ENGINEERING MODEL OF U-SHAPED COLD HEAD Cold Head Description On the basis of the first laboratory version of the U-shaped cold head, an engineering model was designed at Leybold, three samples of which were fabricated and tested. A photograph of one specimen is shown in Fig. 4. As the laboratory version, these engineering cold heads are equipped with water cooling at the warm end, which later will be replaced by air cooling. At the beginning, the same phase shifting components as in the first laboratory version5 were used,
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which consisted of an inertance tube with buffer volume, and a needle valve assembly for the second inlet. Later on, in view of a real product the second inlet valve was replaced by a fixed nozzle and the buffer volume was reduced to a minimum acceptable size, as described below.
Cooling Performance of Engineering Samples Fig. 5 displays the cooling capacity of three samples of the engineering model at 200 W of input power to the Polar compressor. A buffer volume of was used in these tests. A good reproducibility of the cooling performance is found. The no-load temperature varies between 48.2 and 50.6 K, and the cooling power is in the range 5.0 to 5.6 W at 80 K. The corresponding COP of 2.5 to 2.8 % at 80 K is approximately equal to that of the first U-shaped laboratory version5. For a future prototype, all phase shifting components, i.e. inertance tube, second inlet and buffer volume, should be integrated with the cold head to a single compact system. A cold head design without need for manual adjustment of valves is to be preferred.
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As a first step towards a practical cold head, the adjustable second-inlet valve was replaced by a fixed nozzle with suitable flow symmetry. As seen from Fig. 5 for the case of cold head #3, the cooling power with nozzle instead of a valve is even the highest, i.e. 5.6 W at 80 K. This might by due to the somewhat reduced dead volume of the cold head with nozzle. In the next step the effect of buffer size on the cooling performance was tested, in order to find the lowest acceptable buffer volume. Fig. 6 shows the variation of cooling power at 80 K and of no-load temperature with the size of buffer volume. The data were obtained for cold head #3 with nozzle, and with a modified regenerator geometry that gave a lower no-load temperature, as compared to that in Fig. 5. In this test, starting with a buffer of the volume was gradually
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reduced without any other readjustment. As seen from Fig. 6, a good compromise for a compact buffer, without sacrificing too much of cooling performance, will be a volume near At lower volumes the performance is rapidly decreasing.
CONCLUSIONS The present work demonstrates the feasibility of pulse tube cold heads that operate efficiently on a commercial Stirling compressor (Leybold Polar SC). With the U-shaped cold heads, cooling powers at 80 K of 6.0 W for the laboratory model and (5.3 ± 0.3) W for three samples of the engineering model have been achieved at an electrical input power of 200 W to the compressor. The slightly lower cooling power of the engineering samples can be ascribed to the fact that the latest improvements in the U-shaped laboratory versions have not yet been incorporated. With an inline pulse tube cold head even a cooling power of 8.1 W is available at 80 K with 200 W of input power. The COPs of the realized PTCs are in the range 2.6 % - 4%, which comes close to or, in the case of the inline cooler, is even larger than the COP of the corresponding Stirling cold finger. The reproducibility of the performance data for the three engineering samples is rather promising. With regard to a compact and efficient cold head for practical use, it was found that the second-inlet valve of the laboratory version can be replaced by a fixed nozzle without problems, and that the buffer volume size can be significantly reduced without essential loss of cooling performance.
ACKNOWLEDGMENT This work is financially supported by the German BMBF under contract no. 13 N 7393/5.
REFERENCES 1.
Radebaugh, R., “Development of the pulse tube refrigerator as an efficient and reliable cryocooler,” Proceedings of the Institute of Refrigeration vol. 96, London (2001), pp. 11-31.
2.
Häfner, H.-U., Fiedler, A., and Rolff, N., “Long-life Stirling cooler for HTS-electronics qualification and application,” 8th International Superconductive Electronis Conference, -Extended Abstracts-, Osaka (2001), paper P1-E9.
3.
Kotsubo, V., Olson, J.R., Champagne, P., Williams, B., Clappier, B., and Nast T.C., “Development of pulse tube cryocoolers for HTS satellite communications”, Cryocoolers 10, Kluwer Academic/Plenum Publishers, New York (1999), pp. 171 -179.
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4.
Tward, E., Chan, C.K., Raab, J., Nguyen, T., Colbert, R., and Davis, T., “High efficiency pulse tube cooler” Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2001) pp. 163-167.
5.
Yang, L.W. and Thummes, G., “Medium-size pulse tube coolers with linear compressor”, Adv. in Cryogenic Engineering, vol. 47, American Institute of Physics, New York (2002), in press.
6.
Gardner, D.L. and Swift, G.W., “Use of inertance in orifice pulse tube refrigerators”, Cryogenics, vol. 37 (1997), pp. 117-121.
7.
Gedeon, D., “DC gas flows in Stirling and pulse tube cryocoolers”, Cryocoolers 9, Plenum Press, New York (1997), p. 385-392.
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Design and Test of a 70 K Pulse Tube Cryocooler Y. Yasukawa, K. Ohshima, K. Toyama, T. Itoyama, Y. Tsukahara, R. Kikuchi and N. Matsumoto Fuji Electric Corporate Research and Development, Ltd. Tokyo 191-8502, Japan T. Kamoshita and T. Takeuchi Fuji Electric Co., Ltd. Tokyo 141-0032, Japan
ABSTRACT Fuji Electric has designed, fabricated and tested a 70 K pulse tube cryocooler. The cryocooler was designed for commercial use to provide 2 to 3 W at 70 K. The basic concept of this cryocooler is compactness, low cost, high reliability and high performance. A compressor is a key part for providing a compact cryocooler. The compressor is composed of dual-opposed-pistons that move with a coil. A new concept for supporting the moving units with flexure bearings enables a smaller size in the axial direction. The reduction or integration of parts significantly diminishes the cost of the compressor. The pulse tube part of the cryocooler is implemented in an in-line configuration in order to provide high cooling performance. Dimensions of the pulse tube, regenerator geometry and phase shifter are optimized by experimental methods. In addition, the cooling performance has been enhanced by optimization of the compressor. Application of these technologies enables the pulse tube cryocooler to provide a cooling capacity of 2.5 W at 70 K with 100 W electrical input. Reliability is one of the most important requirements, and component reliability tests, which include mechanical fatigue of the flexure bearings and contamination of the helium working gas, were conducted. The lifetime of the cryocooler is expected to be more than 50,000 hours.
INTRODUCTION Pulse tube cryocoolers (PTC) are expected to be utilized for commercial and satellite applications because of their features of no moving parts, low vibration, and potential for long lifetime. Moreover, the simple configuration of a PTC has low cost, unlike that of a Stirling cryocooler. Fuji Electric has been developing long life Stirling cryocoolers for satellite applications for more than 10 years.1 High reliability of those cryocoolers is based on the technologies of the flexure bearings on the moving piston and prevention of contamination of the working helium gas. We have applied these technologies to the new pulse tube cryocooler development.2 Compactness is also a requirement for commercial use. The configuration of the motor and flexure bearings has been redesigned, thereby achieving a smaller PTC compressor size compared to the Stirling cryocooler. Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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DESIGN CONCEPT Pulse Tube The design goals of the 70K pulse tube cryocooler are shown in Table 1. Applications of this PTC are for cooling high-Tc devices (e.g. RF microwave filters) and infrared sensors, and therefore compactness, low cost, high reliability and high performance are important requirements. There are two geometrical arrangements of a pulse tube and a regenerator in a pulse tube cryocooler (PTC). Those are the in-line and the return-shaped (U-tube) configuration. Sometimes the U-tube pulse tube provides more convenient mounting of the cold load. However its performance is not as efficient because of the difficulty of flow in the pulse tube and the increased dead volume in the cold head. On the other hand, the in-line configuration has the advantage of better cooling performance. We have chosen the in-line configuration based on its superior cooling performance. An inertance tube with a buffer tank is utilized as the phase shifter of the pulse tube so that DC flow in the pulse tube is not a factor. Dimensional optimization of the pulse tube and regenerator is also an important issue. At first, we carried out calculations based on a numerical simulation program we have developed. Preliminary test results of a cryocooler fabricated based on those calculations diverged from the calculated results. For a miniature cryocooler, it is very difficult to predict precisely the heat loss into the coldhead. Therefore, dimensions of the pulse tube and regenerator were optimized based on actual cooling performance tests.
Compressor The actual compressor of this PTC development is based on the technologies of our Stirling cryocoolers. The compressor for our Stirling cryocooler is composed of dual pistons that move in opposite directions and are supported by flexure bearings. Between the piston and the cylinder exists a narrow gap, called the clearance seal. There is no lubricant between the piston and the cylinder, which improves the reliability of the compressor. The stress that is generated in the flexure bearings is designed to be less than the allowable stress, considering cyclic fatigue. Therefore, the reliability of the mechanical structure is guaranteed. However, when the flexure bearings are used to support the pistons, it is difficult for the compressor dimensions to be miniaturized by a generic design. Compactness is an important issue. Thus, a new concept of compressor design has been employed to miniaturize the compressor dimensions. The new arrangement of the flexure bearings and linear motors enables the compressor size to be smaller in the axial direction. Strictly limiting contamination of the working gas is another important condition for improving reliability. For the compressor components, we have chosen materials that have a low level of outgassing. Additionally, heat treatment of the components has been used to diminish the moisture that adheres to the surfaces. Reliability against outgassing can be confirmed by accelerated tests. The developed pulse tube cryocooler is shown in Fig. 1.
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DESIGN PROCEDURE Phase Shifter and Method of PV Work Measurements An inertance tube with a buffer tank is utilized as the phase shifter for the pulse tube. The phase shifter is modeled as an impedance of an electrical circuit. The relationship between pressure and flow rate can be calculated by using a distributed parameter system.3 However, our calculations have not agreed well with the actual measurements, and therefore we have chosen to directly measure the PV work by the phase shifter when integrated with the compressor.4 Figure 2 shows the test setup for measuring the PV work. The PV work at the compressor is equivalent to the PV work at the hot end of the pulse tube. Figure 3 shows the PV work at the compressor versus pressure amplitude for various lengths of inertance tube. If an adiabatic state is assumed at the pulse tube, the PV work is constant throughout the pulse tube. So, by measuring PV work at the hot end of the pulse tube using the results of Figure 3, we can obtain the PV work at the coldhead. The total loss can be calculated by subtracting
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the net cooling power from the PV work. Figure 4 shows a sample of the cooling capacity versus effective length of the regenerator. In this figure, the net cooling power is largest at an effective regenerator length of 60 mm while the PV work at the cold head is largest at a length of 50 mm We can obtain the quantity of the heat loss into the coldhead. Thus, the distinction between cooling capacity and heat loss is useful for the process of optimization.
Dimensions of Regenerator and Pulse Tube For miniature cryocoolers, because of the poor predictability of heat loss into the coldhead, it is difficult to optimize the regenerator and pulse tube dimensions by calculation. We prepared test samples of the regenerator and pulse tube having variable inner diameters and lengths. The regenerator geometry also had various test parameters. Figure 5 shows the dependence of cooling capacity on regenerator geometry. We chose #400 S.S. screen mesh for the regenerator material. The pulse tube and regenerator dimensions have been decided by the same procedure. By test results, we can find out the optimum conditions in the pulse tube and regenerator dimensions.
PERFORMANCE TESTS Cooling Load Lines Figure 6 shows the cooling performance for different values of electrical input power. Figure 7 shows the coldhead temperature dependence of the specific power. The main configuration parameters and phase shifters were optimized for a cooling capacity of 2.5 W at 70 K. The lowest temperature achieved at an input power of 120 W was 45.2 K with no heat load. At a coldhead tempera-
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ture of 70 K and input power of 120 W, the specific power is 48 W/W; this is the best value among the different electrical input powers used. The PV work is 67.6 W at 70 K and 66.7 W at 80 K during the 120 W input. These values yield a compressor efficiency of 56 %. On the other hand, the efficiency of the pulse tube is 3.7 % at 70 K and 5.4 % at 80 K. Cooling performance depends not only on parameters of the regenerator, pulse tube, and phase shifter, but also on operating conditions. Because the compressor inefficiency accounts for a large part of the measured cooler performance, it was expected that the cooling performance could be enhanced.
Enhanced performance The results of the performance tests did not completely satisfy the design goals. In particular, the efficiency of the compressor had room to be improved. Important possibilities for enhancing the efficiency were reduction of eddy current losses and better matching of the resonant frequency. Subsequently, we redesigned and fabricated a new compressor. Figure 8 shows the preliminary test results using the new compressor. A cooling capacity of 2.5 W at 70 K with 100 W electrical input power has been achieved. The efficiency of the compressor has been improved to 66%. The COP will be even further improved by producing a resonant system of the compressor.
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RELIABILITY TESTS Reliability is another important issue for the PTC. There are no moving parts in the pulse tube part of the PTC, but the compressor has some moving parts (e.g. pistons and flexure bearings). The flexure bearing is a key component for achieving a long useful lifetime. We have carried out tests to verify the flexure bearing reliability.
Flexure Bearings Cyclic stress is generated on the flexure bearings so that the test is simulated under the same conditions as actual operation. The operating frequency and the amplitude of the piston motion are considered for this test. Since we selected stainless steel for the flexure bearings, cyclic stress of cycles is necessary to evaluate the integrity of the bearings. This is because the tolerant cyclic stress for stainless steel becomes constant over that many cycles. Another important issue to be evaluated is whether the piston remains straight during operation. Figure 9 shows displacement of the piston in the radial direction versus the number of cycles of operation. The test was performed with 6 flexure bearing units. As can be seen the displacements are less than 2 µm of the allowable level of 3 µm, and the
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amount of displacement tends to remain constant. The test results indicate high reliability of the flexure bearing systems. This test is ongoing. The pulse tube cryocooler has an expected long lifetime of 50,000 hours.
Working Gas Contamination Decreasing the working gas contamination is another important condition for improving reliability. The pulse tube part is made of metallic materials so that there is no chance of generating organic outgassing products from these materials. On the other hand, the use of nonmetallic materials in the compressor cannot be avoided. We chose low-outgassing materials for the compressor components. It is well known that outgassing affects the cooling performance of pulse tube cryocoolers. So, we carried out accelerated tests to verify the outgassing effect. Figure 10 shows the effect on cooling capacity of organic contamination in the working gas. The test method evaluates the cooling performance of a PTC that has been deliberately charged by a contaminated working gas. The contaminated gas was produced by another vessel that contains compressor components at a heated condition. The test results predict that the cooling capacity will decrease by less than 10% during 50,000 hours of operation at an ambient temperature of 60°C. That level of degradation is allowable for PTCs.
CONCLUSIONS Fuji Electric has designed, fabricated and tested a 70 K pulse tube cryocooler. The pulse tube cryocooler provides a cooling capacity of 2.5 W at 70 K with 100 W electrical input power. The efficiency of the compressor is less than 70%. So, it is possible for the COP to be further improved. Component reliability tests that included mechanical fatigue of the flexure bearings and contamination of the helium working gas have also been conducted. The lifetime of the cryocooler is expected to be more than 50,000 hours.
ACKNOWLEDGMENT The authors heartily thank professor Matsubara of Nihon University who gave them useful advice on the development of pulse tube cryocoolers.
REFERENCES 1.
Fujinami, F., “Trend of Development of Miniature Cryocooler,” Cryogenic Engineering (in Japanese), vol. 30, No. 2 (1995), pp. 55-60.
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2.
Yasukawa, Y., “Development of A Miniature Pulse Tube Cryocooler,” 65th Meeting on Cryogenic and Superconductivity (in Japanese), (2001), pp. 272.
3.
Hofmann, A., Pan, H., “Phase shifting in pulse tube refrigerators,” Cryogenics, vol. 39 (1999), pp. 529-537.
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, Kluwer Academic/Plenum Publishers, New York (1997), pp. 269-278.
Miniature 50 to 80 K Pulse Tube Cooler for Space Applications T. Trollier and A. Ravex (1) I. Charles and L. Duband (2) J. Mullié, P. Bruins and T. Benschop (3) M. Linder (4) (1) Air Liquide Advanced Technology Division, AL/DTA Sassenage, France (2) Atomic Energy Committee, Low Temperature Division, CEA/SBT Grenoble, France (3) THALES Cryogenics B.V. Eindhoven, The Netherlands (4) European Space Agency, ESA/ESTEC Noordwijk, The Netherlands
ABSTRACT A miniature pulse tube cooler is presently under development in partnership between AL/DTA, CEA/SBT and THALES Cryogenics. The Engineering Model foreseen is aiming at providing 800 mW at 80 K with 40 watts input power to the motors of the compressor. A development phase has been performed with an in-line architecture for the pulse tube cold finger connected to an existing flexure bearing compressor from Thales Cryogenics. Presently, more than 900 mW at 80 K has been achieved at 288 K ambient temperature provided by water cooling, in inertance mode, and with less than 25 watts PV work. The development phase is presented as well as the various trade-offs made, both on the cold finger and compressor side, to cope with the thermal, mechanical and electrical environmental specifications. The impact of the matching between compressor and pulse tube cold finger is also discussed. This work is performed in the framework of a Technological Research Program funded by the European Space Agency. An Engineering Model will be delivered to ESA/ESTEC in February 2003. This coming generation of miniature pulse tube coolers will be used for the cooling down of detectors in future earth observation missions.
INTRODUCTION The overall objective of the work is to optimize, design, and manufacture at pre-qualification level, a 50-80 K Miniature Pulse Tube Cooler (MPTC). The resultant MPTC shall be commercially competitive in performance, mass and cost within the future space cryocooler market. It shall offer significant advantages over the presently available technology and shall require no, or only minor Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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delta qualification for direct use in future spacecraft applications. The technical specifications from ESA/ESTEC are summarized in Table 1.
DEVELOPMENT PHASE A 15 months duration development phase has been performed. During this phase, 9 pulse tube cold fingers Development Models (DM) have been designed, manufactured and tested so far. In order to modify easily the geometry for optimisation of the DMs, an in-line configuration has been used for the cold finger. Common materials such as stainless steel and pure copper have been implemented for the tubes and for the heat exchangers as shown in the Figure 1. Water cooling is provided at both the hot ends of the regenerator and the tube. All the DMs have been operated in an inertance mode. Although the introduction of a secondary orifice usually leads to increased efficiencies compared to the simple orifice pulse tube refrigerator, it also induces some parasitic flow problems. Performance of the double inlet pulse tube refrigerator is not always reproducible within a cooler batch. Researchers1,2 attribute this erratic behaviour to DC flow that takes place in the loop formed by the regenerator, pulse tube and secondary orifice. Asymmetric flow impedance in the secondary orifice can also cause such a DC flow which carries a large enthalpy flow from the warm end to the cold end even for a DC flow of a few percent of the
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AC flow amplitude. As reported in previous work3, operation in orifice mode could also lead to efficient pulse tube cryocooler when used with inertance tube. This mode would by definition eliminate the DC flow. Due to this experience, the inertance phase shift control is our preferred operation mode. The pressure oscillator connected to the pulse tube cold finger Development Models is derived from the new family of tactical Stirling coolers4 manufactured by Thales Cryogenics B.V. It consists of a unique new generation of long-life, flexure-bearing, linear driven dual opposed pistons oscillator. The linear motor configuration uses moving magnets attached to the pistons that reciprocate inside static coils. The compressors used are capable to provide 4.2 cubic centimeters swept volume for 100 watts maximal input power to the motors. The outer envelope of the compressor is 60 mm diameter and 165 mm length. Both compressor and cooler drive electronics are mounted onto a water cooled cold plate as presented in the Figure 2.
PERFORMANCE OF THE DEVELOPMENT MODELS Each DM has been tuned with various inertance length and diameter for 30 and 40 bars filling pressure and for 40 to 60 Hz operating frequency. The pulse tube cold fingers are operated in vertical orientation, with the warm end of the tube in the upright position. MLI has been wrapped around the regenerator and the pulse tube. A mapping of the optimal performances achieved with the nine DMs experimented is plotted in the Figure 3. Each DM differs from the other with slight changes in the dimensions (regenerator and/or tube, diameter and length), leading to various optimal settings of the inertance and the operating frequency. The mapping is performed with 288 K heat rejection temperature, 25 watts input corresponding to the electrical input power minus the Joule losses (thus assumed to be close to the PV work) and 20 cm transfer tube between compressor and cold finger. In the best configuration (DM #8), with 25 W input the no-load temperature is 56.3 K and the temperature increases to 82.3 K with 1 watt heat load applied at the cold tip. This performance corresponds to 907 mW of cooling capacity at 80 K, with a cooling power slope of 26.2 K/W. The stroke of the pistons has been measured in such conditions with an LVDT tansducer, leading to a swept volume of 1.50 cubic centimeters. With 35 W input, the no-load temperature is 53.8 K and the temperature increases to 75.3 K with 1 watt heat load applied at the cold tip. This performance corresponds to 1230 mW of cooling capacity at 80 K, with a cooling power slope of 21.3 K/W. In this case, the swept volume increases
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to 1.76 cc. Both performances were achieved with 40 bars Helium filling pressure and 50 Hz operating frequency. The load curves achieved for 25 W and 35 W input (PV work) are reported in the Figure 5.
ENGINEERING MODEL PRELIMINARY DESIGN Pulse Tube Cold Finger EM The regenerator and the tube are mounted onto a flange in a U-shape configuration. This configuration has been selected to provide good cryogenic performance in a compact, robust and simple design that enhances the integration compared to an in-line configuration. The flange is
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made of a specific aluminium material for thermal heat transfer and mechanical resistance optimisation. The hot flange is designed such as it integrates the buffer volume and the heat exchangers of the regenerator and the tubes. The heat exchangers are manufactured by Electron Discharge Machining (EDM) directly in the flange material. The inertance is wound inside the buffer volume. Both the tube of the regenerator and the pulse tube itself are made of thin walled titanium alloy TA6V4 in order to reduce the parasitic heat leaks. The design makes use of bolted flanges and metallic C-rings to seal the regenerator and the tube to the aluminium hot flange. At the cold side, a high vacuum brazing process is used for the assembly of the titanium tubes onto the pure copper cold block of the cold finger. As represented in Figures 6a and 6b, the pulse tube EM will incorporate a snubber which will be used as a launch bumper stop to prevent any excessive lateral motion of both tubes and consequently to significantly reduce the mechanical stress on the tubes at the flange location. This snubber will be made of Titanium for obvious reasons of mechanical performance and density
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optimisation. A low conductive fibber glass part placed between the cold block and the snubber cylinder is ensuring that low parasitic heat losses are added in case of contact during operation (with loads). In normal condition, there is no contact between the snubber and the cold block of the pulse tube. The design of the EM provides sufficient margins with respect to an internal pressure of 75 bars, equivalent to 1.5 times the maximal operating pressure. The total mass of the EM pulse tube cold finger is predicted to be 700 g including the snubber. Linear Compressor EM The compressor design is built around a moving-magnet linear motor that drives the pistons in dual opposed configuration into the same compression chamber. The moving magnet linear motor offers big advantages over the conventional moving-coil design. This innovative concept allows the coils that are the main source of gas contamination to be placed outside the working gas. Additional advantages are the absence of flying leads and glass feed-throughs to supply current to the coils. Thus, moving magnet technology is applied in our compressor design to improve the reliability of the complete system. The main disadvantages of this configuration are the losses and the EMI, which are higher than in a conventional moving-coil design. High performance axially magnetized NdFeB magnets are used in the motor. Flexure-bearings are used in order to have a radial clearance between the piston and the cylinder. These flexure-bearings are round discs made of spring steel, with 3 arms. With this kind of flexure bearings, a very high radial stiffness can be reached. By changing the shape, the length and the thickness of the arm, the ratio between the axial and the radial stiffness can be changed without increasing the maximum stresses in the flexures. The fatigue limit of the spring steel is To have enough safety margin the design limit for the VonMisses stresses is set to as presented in the Figure 7 below. The EM compressor assembly is represented in the Figure 8. The coils holders are made of titanium alloy in order to reduce the eddy current losses and to combine high mechanical resistance and low density. The two compressor halves are mounted on a dedicated aluminium alloy “centre plate” that contains all the mechanical and thermal interfaces of the compressor and the two cylinders. Bolted flanges are directly machined in the titanium alloy block of the coil holder. The gas containment is achieved by means of aluminium C-rings that provide a leakage rate of The outer diameter of the compressor halves is 63 mm and the total length is approximately 170 mm. The total mass of the Engineering Model compressor is predicted to be 2100 g. The overall efficiency of the compressor has been simulated to 70% at 20°C ambient temperature.
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CONCLUSIONS A compact, lightweight and robust U-shape miniature Pulse tube cooler is under design. Presently, the predicted cryogenic performance of the Engineering Model while implementing the high performance materials depicted herein are 1240 mW @ 80 K with 288 K heat rejection temperature and 43.4 Wac electrical input power (considering 70% compressor efficiency). Some optimisation work is still going-on in order to increase the cooling capacity (pulse tube geometry) and the compressor efficiency (magnetic circuit and coils).
ACKNOWLEDGMENT We acknowledge the financial and technical support of the European Space Research and Technology Centre (ESA/ESTEC, Contract N°14896/00/NL/PA).
REFERENCES 1. L. Duband, I. Charles, A. Ravex, L. Miquet and C. Jewell, “Experimental Results on Inertance and Permanent Flow in Pulse Tube Coolers,” Cryocoolers 10, Plenum Publishing Corp., New York (1999), pp. 281-290.
2. D. Gedeon, “DC gas flows in Stirling and pulse tube refrigerators,” Plenum Press, New York (1997), pp. 385-392.
3. L. Duband and A. Ravex. “Development of a Pulse Tube 50-80 K Cryocooler,” ESA Contract 11331/ 95/NL/FG Final Report, Document CEA Note SBT/CT/99-15, April 1, 1999, Issue 1, Rev 0.
4. M. Meijers, A.A.J. Benschop and J.C. Mullié, “High Reliability Coolers under Development at SignaalUSFA,” Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2001), pp. 111-118.
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Design and Characterization of a Miniature Pulse Tube Cooler A.S. Gibson, R. Hunt Astrium UK Limited Stevenage, Hertfordshire, UK I. Charles, L. Duband CEA/DRFMC/Services des Basses Températures Grenoble, France M. Crook, A.H. Orlowska, T.W. Bradshaw Rutherford Appleton Laboratory Chilton, Didcot, Oxon, UK M. Linder European Space Agency (ESTEC) Noordwijk, The Netherlands
ABSTRACT The design of an advanced Miniature Pulse Tube Cooler (MPTC) for long-life space applications is presented. The cooler system incorporates a balanced compressor, with an in-line pulse tube configuration, yet retains the flexibility of a transfer line. The ‘split’ configuration provides obvious advantages for instrument design, while maintaining economical options for drive electronics and maximising potential for vibration isolation between compressor and detector. It also potentially allows the pulse tube end to be separately placed within a pre-cooled enclosure at temperatures beyond the range of operation for typical compressor mechanisms. The system employs an integrated back-to-back compressor for inherent vibration cancellation. The design draws on RAL compressor experience and uses a modified flexure spring, motor and clearance seal features based on that of the Oxford/RAL heritage coolers, as well as proven methods for alignment and verification of compressor health. The mass of the system has been minimised, with significant mass savings realised in the motor design, with additional challenge presented by the split/in-line configuration. A thorough set of trade-off studies and detailed analyses have been performed to establish the coil, permanent magnet and spring configurations. Baseline design parameters from the CEA/SBT model predictions have been optimised through testing of various pulse tube geometries, with attention to regenerator and inertance tube dimensions, drive frequency and fill pressure. Sensitivity to these parameters has been studied to Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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increase robustness of the finalized EM pulse tube design. The cooler has exceeded the specified 1.2 W heat lift required at 80K, with a capability to lift in excess of 1.6 W at 80 K at full compressor stroke.
INTRODUCTION During the 1980’s and early 1990’s the usage of high frequency Stirling cycle cooler became prominent, selected to provide cryogenic cooling in various satellite instruments for Earth observation and atmospheric monitoring applications. The implementation of linear motors, with clearance seals and diaphragm springs for radial support, proved to eliminate life-limiting wear mechanisms in the coolers. The success of the ISAMS cooler built by Oxford and the ATSR coolers built by Rutherford Appleton Laboratory led to widespread acceptance of the technology.1 With the support of ESA, the technology was transferred from RAL to Astrium UK (formerly MMS and BAe) for industrialisation, resulting in a series of commercially manufactured coolers for a range of flight programs2 including applications in the infra-red and the first HTS3 (high temperature superconductivity) applications. The capability of the cooler was extended by Astrium to provide increased heat lift in the 50-80 K model, with over 25 units manufactured, 10 of which are to be launched in 2002 alone. ENVISAT (launched March 2002) and INTEGRAL4 missions are each responsible for transporting 4 of these coolers into orbit. Despite the heritage accumulated by this class of mechanisms since these early developments, a market trend toward pulse tube coolers has emerged (generally in the range of 40-120 K), with Stirling coolers remaining competitive at lower temperatures.5 Pulse tube coolers in the higher temperature range of applications offer the following key advantages: Potential to eliminate vibration source in detector stage Potential for reduced EMC influence on detector without cold-stage motor Ability to operate the cold-end over wider temperature range (no mechanism)
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Reliability improvement related to no cold-moving parts Drive electronics simplicity (fewer channels, fewer telemetry and launch-lock provisions) Improved side-load capability of the cold-tip Realization of the first three advantages depends directly on the pulse tube configuration implemented. Vibration and electromagnetic interference issues are both driven by the compressor component. The ‘split’ configuration, with the transfer line (refer to Fig.l), maintains all of the above advantages, while offering the ability to reject most of the heat generated via the compressor located outside of the detector region. This is the primary reason for the selection of the split configuration for the design presented herein. In contrast, the compressor unit in the integrated configuration must remain directly coupled to the detector assembly. Mass savings result and a minor improvement in performance is expected with the elimination of the transfer line and the sharing of common structural supports. Such gain is at the expense of some of the inherent advantages that the pulse tube offers (relative to the Stirling), though self-induced vibration can be reduced using more sophisticated electronics. With the space market trending toward pulse tubes, it is recognised that the most critical aspects of the pulse tube cooler technology are directly addressed by existing Stirling technology. Diaphragm springs and clearance seals remain the most critical elements, though no longer required to operate at cryogenic temperatures as with the Stirling displacer mechanism. Adjustments are made for operation at higher fill pressures and to accommodate a larger volume of gas in the cold end. Gas sealing issues, cleanliness constraints, regenerator technologies and structural issues are virtually unaffected from those overcome in the Stirling designs and so the knowledge gained from past experience remains relevant. The pulse tube cooler presented herein builds on past experience gained through development of previous U-shape pulse tubes. Astrium began collaborating with CEA/SBT and RAL on pulse tube development in 1996. Early work culminated in a cooler with ability to lift 1.4 W at 80 K.6 As the understanding of pulse tube phenomena improved, including parasitics related to DC flow in double-inlet systems, the original orifice designs were converted to inertance tube types. This approach proved to be inherently more efficient than the orifice type pulse tube. In the current development program, responsibility for system management, pulse tube and compressor design have been assigned to Astrium UK, CEA/SBT and RAL respectively. The activity is classed as a Technology Research Program (TRP) that is overseen by the European Space Agency. The new miniature pulse tube development specifications summarised in Table 1.
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DEVELOPMENT APPROACH Breadboard Methodology A predictive model was used in an early analysis to determine the swept volume required to achieve the specification. The pulse tube model developed by CEA/SBT is based on considerable test history gained from testing of both low frequency industrial applications and high-frequency coolers. The models have provided a starting point for the investigation of inertance tube and regenerator dimensions, frequency range and fill pressure to achieve the required heat lift with a reasonable margin relative to the ESA specification. The modeled parameters chosen were implemented in a breadboard (BB) model pulse tube and tested with a standard 50-80 K Stirling type compressor. The breadboard unit, pictured in Figure 2, enabled the optimisation of key pulse tube design parameters while the balanced compressor design was being detailed. As the pulse tube operating conditions differ significantly from those of the 50-80 K cooler, the compressor was not expected to operate at an ideal efficiency, but simply to provide a representative swept volume. The lab drive electronics limited the swept volume of the 50-80 K compressor to 83% of the peak design value for the new balanced compressor design. This was deemed to be sufficient to optimise and characterise the breadboard pulse tube design. The thermodynamic specific cooling power was used as the key figure of merit to optimise the pulse tube, separating the issue of compressor design efficiency from this part of the study. The breadboard unit enabled selection of final pulse tube geometry.
Engineering Model Approach In parallel with breadboard testing, the EM compressor design has embraced the same design constraints for critical components as with past Stirling designs, respecting the same
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stress limit guidelines for diaphragm springs in particular. Changes in materials, as well as overall simplification is driven by the goal to meet significantly lower cost targets. These improvements included innovative approaches to improve tolerance stickups, reduce part counts and embrace recent improvements in magnetic materials. The goal was to produce an inherently balanced compressor with low mass, high-efficiency motors and lower cost, suitable for batch production. The EM pulse tube goal is to verify the final geometry with material refinements to optimise mass and regenerator performance, with the inertance tube integrated into the buffer volume. The pulse tube test campaign will focus on performance characterisation over temperature.
BREADBOARD PULSE TUBE TESTS RESULTS Over 350 test cases have been performed to optimise the cooler, varying parameters of regenerator dimensions, inertance tube dimensions, drive frequency, heat load and fill pressure in a systematic manner. Measurements were taken over a range of frequencies (39 Hz, 45 Hz, 50 Hz and 54 Hz) for each geometry, to the optimum parameters. Two filling pressures (20, 30 bar) were tested to see the impact of the mean pressure. As a precaution, the compressor unit was limited to operate at fill pressures up to 30 bar gauge, having been verified to 50 bar gauge in a proof test prior to start of test. Economical measures were taken with the breadboard components used. Stainless steel regenerator and pulse tube shells were used to facilitate quick turnaround times with ease of machining. In addition, multi-layer insulation (MLI) was not optimised, as it was not essential to the thermodynamic optimisation of the study. Therefore, the results obtained are conservative, but are reflective of an optimised geometry despite higher base temperatures. Low conductivity titanium material and improved MLI will be implemented on the EM unit.
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Regenerator and Inertance Tube Optimisation The electrical power injected into the compressor has been adjusted in order to keep constant PV Work in a series of systematic tests. The fixed thermodynamic input power was chosen to be in the region of the specified load point, with being maintained. In some selected configurations the impact of higher power levels up to 35 W have been investigated. The cold temperature has been measured with no load (0 W) and with 1 W. It can be assumed that the variation of the temperature with the applied load is linear so the cooling power available at 80 K is calculated in the range of these 2 test points. Three different regenerator geometries were tested at 25 W PV-work, to understand the effect of diameter and length variation with a constant mesh configuration. For each iteration of the regenerator geometry, the inertance tube was optimised at 2 fill pressures. Following the first iteration, the length of the pulse tube was extended. Finally, the diameter of both pulse tube and regenerator were increased. The progression of regenerator configurations was observed as follows: 1. Original regenerator - 500 mW heat lift at 80 K, reaching 65 K at 0 W heat load 2. Longer regenerator - 730 mW heat lift at 80 K, reaching 61 K at 0 W heat load Longer, wider version - 965 mW heat lift at 80 K, reaching 55 K at 0 W heat load (also at 35 W PV work to demonstrate 1225 mW lift at 80 K) The final configuration was chosen for the baseline geometry. Figure 3 demonstrates the variation of performance with changes in geometry and operating conditions at constant PV work. Clearly, the design is not sensitive to manufacturing tolerances of the inertance tube or fill pressure. The best performance was obtained at the higher of the 2 filling pressures in every case, consistent with modeling predictions. A variation of drive frequency within about 2 Hz of the centre frequency can be tolerated without significant loss of performance. This is obviously important to the design of the compressor in terms of designing to operate at resonance. As predicted, the 50-80 K Stirling compressor operated with a relatively low efficiency of ~60% for a range of pulse tube operating conditions. A PV-specific power of 27.5 W/W at 80 K was achieved in the breadboard tests without the optimisation of fill pressure or parasitic losses. The tests demonstrated the feasibility of a miniature pulse cooler and have provided clear understanding of performance trends. A supplementary test with PV work of 36.7 W demonstrated the cooler’s ability to exceed the specification of 1.2 W at 80 K (ambient rejection).
COOLER SYSTEM DESIGN The Engineering Model Pulse Tube system has been designed to accommodate some additional features for ease of performance monitoring. These include a face-seal type fittings used on the transfer line and a bi-metallic adaptor installed at the pulse tube inlet. The system mass of the EM unit has come reasonably close to the 3.0 kg design goal specified as demonstrated in Table 2 below, considering that some economical choices have been made with fasteners, etc. Removal of face-seal fittings and bi-metallic joints, implementation of titanium fasteners and substitution of a beryllium aluminum alloy thermal-structural support for the pulse tube will allow the design to meet this goal. These are low-risk changes, which can be made without detriment to performance. Note that a low-mass mounting system is offered to the user at an additional mass of ~100 g, and is incorporated into the central flange as shown in Figure 1.
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On the EM unit, face-seal type fittings are attached at either end of the transfer line for the installation of pressure transducers to be used for performance verification.
IMPROVED PULSE TUBE DESIGN An in-line configuration has been selected for the MPTC, with a transfer line of 260 mm maintained to realise the greatest practical benefit of the pulse tube. The benefit is a highperformance system, without the need for sophisticated drive electronics control in the majority of applications. The choice of an inline configuration allows to minimise the dead volume and flow perturbation at the cold end. The result is higher efficiency compared to a U-shape or coaxial configuration. However, this leads to the need for a thermal path between the hot heat exchanger of the tube and the main heat exchanger at the entrance of the regenerator. There is an inherent mass impact related to the performance improvement. The impact is minimised by using the thermal path as a structural/snubber support for the hot end heat exchanger and the cold tip. The hot heat exchanger is connected by a specialised thermal link. This design allows axial displacement of the tube hot end due to differential thermal contraction during cool down. Adjustable screws are used to limit displacement of the cold part. For compactness and integration the inertance tube is wound into the buffer volume, exiting from a small flange and fixed along the thermal/structural support to reach the hot end heat exchanger end. Pressure analysis has been used to verify the design under a conservative proof load. The buffer volume cover makes use of the analysis to ensure that welds are located in low level stress areas (below 100 MPa). The structural design has also been verified by analysis to ensure all fundamental modes of the pulse tube are above 300 Hz. 3 adjustable snubber screws limit the radial displacement of the cold part. Extra-fine threads allow for precise adjustment of snubber gaps. For protection of the copper cold part, a stainless steel ring is installed around the cold-tip. MLI is wrapped around cold components prior to assembly of the hot end thermal link and the snubber points. As mentioned previously, the two thin tubes (regenerator and pulse tube) will be replaced by low thermal conductivity, high strength 6A1-4V Ti parts. This alloy presents a very low thermal conductivity, half that of stainless steel used (80-300 K) for the breadboard. The final regenerator mesh configuration and materials will be confirmed in EM testing, though the breadboard versions have performed exceptionally well. Heat exchangers are used at 3 stages throughout the pulse tube. An EDM method developed by CEA/SBT is used to approximate flat plate heat transfer characteristics, with good thermal coupling between the gas and the main pulse tube body, without a significant pressure drop compared to that of the regenerator. An injection cone allows for a uniform flow pattern at the exchanger entrance. The inlet heat exchanger, which was made from copper in the breadboard unit, has been adapted and integrated with the buffer volume, which is made of aluminium alloy. This change has allowed a low pulse tube mass to be achieved and also reduced the part count. Heat rejected at ambient temperature will depend on the level of the electrical input power and of the cooling of the output gas at the compressor. Using ANSYS™ thermal analysis, assuming 35 W worst-case dissipation, the temperature gradient within the main housing is expected to be < 8 K for with the housing operating near ambient temperature. However, the pulse tube components should be capable of operating over a temperature range far outside the specification (refer to Table 1). The design shows potential to operate below –100°C, enabling radiative pre-cooling options in power-limited systems.
BALANCED COMPRESSOR DESIGN The compressor design draws upon previous linear motor reciprocating mechanism heritage, in particular the flexure bearing technology, for lifetime considerations, and a moving coil for mass and efficiency considerations, consistent with compressors previously qualified by Astrium UK and RAL. The compressor is arranged in a back-to-back configuration to minimise induced vibration, but is made in symmetric modular components attached to a central flange. This
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effectively eases testing and other batch production issues. A thermal model has been used to verify the temperature gradients for a range of operating conditions. Considerable improvements have been made in manufacturability, enabled by the experience of Astrium and RAL from previous programs. An innovative approach to the compressor layout has resulted in low mass, employing a ‘moving cylinder’ and coil design. The compressor is made as compact as possible by replacing the moving piston in previous designs with a moving cylinder and a fixed piston. The moving cylinder allows for greater rigidity, making it consistent with high tolerance manufacturing. The moving coil offers clear advantages to avoid damage and associated generation of debris in assembly, with respect to a moving magnet design. With a moving coil, no side forces are generated during insertion. A moving coil allows a low moving mass to be readily achieved. Despite the polymeric material being located inside the cooler, no gas contamination issues have ever been observed to cause performance degradation of Astrium or RAL coolers, due to rigorous bake-out procedures proven on previous flight programs. The active control loop for the moving piston, which is required to obtain the correct phase between the displacer and the compressors in a Stirling cycle cooler, is not strictly required for a pulse tube cooler. A simplified position pick-off (PPO) design is offered as an option for monitoring or launch lock control, such as that used on INTEGRAL are required.4 Where applicable, simplified electronics results in mass and cost savings for the compressor and drive electronics. The seals between the central flange and the main tubular housings are an Inconel metal-tometal C-type seal with a gold surface plating, of the type previously flown in an opto-mechanical instrument on MOPITT.7 These seals are flexible, providing a seal as they conform predictably within a readily machinable cavity, without polishing. They are also pressure energised. In addition the mating surfaces do not have to be polished in order to achieve the specified leak rate. The piston clearance seal is longer than in previous designs and this, combined with a higher operating frequency, results in less leakage past the seal during operation. This also aids to achieve enhanced accuracy of alignment during assembly, a critical requirement for the correct and reliable operation of the device. Alignment procedures are based on the proven methods used to align the ATSR and Astrium 50-80 K coolers. Diagnostics and verfication methods used to assess performance are based on existing qualified methods for Stirling compressors. The flexure springs are derived from heritage designs, consistent with ATSR and Astrium 50-80 K models (RAL springs tested to greater than cycles), modified to incorporate a larger central hole for the moving cylinder design. The spring configuration has been studied in the process of designing the resonant mechanism for operation in the range of 45 Hz and the number of springs has been increased to reduced radial deflection. This increased mass accordingly and the resonance frequency did not increase significantly. Minor adjustments were made to the spring thickness, respecting the heritage levels of acceptable stress limits. Stress analysis was performed assuming conservative displacements of 5 mm deflection axially and 15 µm radially. These conservative assumptions are beyond the physical limits of axial and radial mechanical clearances. The resulting resonant frequency calculated for the EM compressor at the final fill pressure was confirmed to lie between 45 and 50 Hz. Various coil configurations have been investigated with a trade-off between round, square and ribbon wire having been performed. High-energy NdFeB magnet materials used with optimised permendur pole-pieces, have reduced the motor mass by about 40% from initial designs. To reach a final design, a total of 24 test cases of the permanent magnet system were analysed, as summarised in Figure 4. The magnetic efficiency, is calculated from the resonant model predictions of the input power, W, and the Joule losses, These parameters are compared to data for the compressor. Based on past measurement experience, a value of 50-60 N·s/m for the damping factor gives a reasonable representation. Depending on the actual damping factor achieved, the motor efficiency is predicted to be ~80% efficient, which is exceptional considering the size of the stator and mass reduction of the magnetic components.
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The linearity of force over the range of coil motion has been designed to vary by less than 5%, with a peak at the mid-position of the stroke.
PERFORMANCE OF ENGINEERING MODEL Accounting for improvements in parasitic heat leaks as noted above, the heat lift measurements from breadboard test results have been adjusted to give a conservative estimate of EM cooler performance, as shown in Figure 5. Breadboard data has been corrected to account for parasitic improvement calculated for the implementation of titanium regenerator and pulse tubes, as well as a realistic improvement in the MLI configuration.
CONCLUSIONS The design of a low mass, high performance cooler has been established through extensive breadboard testing, trade-off studies, magnetic and structural analysis. An optimised motor design with higher-energy magnets has contributed significantly to mass savings, while achieving excellent linearity and enhanced efficiency. Mass goals have been met for the challenging split/in-line configuration chosen, thus maintaining the inherent advantages offered by the pulse tube cooler, with respect to the forerunner Stirling technology. Performance has been consistent with predictions of the CEA/SBT models. Breadboard testing of many regenerator and inertance tube geometries have allowed for a systematic optimisation of heat lift and efficiency, while gaining an in-depth understanding of sensitivity to key parameters. With improvements expected due to a further increase in fill pressure, along with reduced parasitics, the cooler has the capability to reach an overall specific electrical power of ~30 W/W at 80 K. In addition to enhancing performance by reducing parasitics, the improvements will be important in terms of application in redundant cooler configurations.
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ACKNOWLEDGMENTS The development work for this technology was carried out under a contract of the European Space Agency (ESTEC Co. 14895/00/NL/PA). It is also acknowledged that Astrium UK has performed additional related research and development activities under internal funding.
REFERENCES 1. Jewell, C., “Overview of Cryogenic Developments in ESA”, Proceedings of the 6th European Symposium on Space Environmental Control Systems , ESA SP-400 (1968), Noordwijk, pp. 447-456.
2. Jones, B.G., Scull, S.R., Jewell, C.I., “The Batch Manufacture of Stirling Cycle Coolers for Space Applications Including Test, Qualification and Integration Issues”, Cryocoolers 9, Kluwer Academic/ Plenum Publishers, New York (1997).
3. McKnight, R.A, Bahrain, M. et al., “On-Orbit Status of the High Temperature Superconductivity Space Experiment,” American Institute of Aeronautics and Astronautics, AIAA99-4486 (1999). st
4. Gibson, A.S., Akhtar, S., et al., “The Flight Model Stirling Cryocooler System for INTEGRAL,” 31 International Conference on Enironmental Systems, Orlando, 2001-01-2380 (2001).
5. Baker, G.R., Féger, D., Gibson, A.S., Little, A., Bradshaw, T.W., et al., “Demonstration of a 10K Stirling Cycle Cooler for Space Applications,” 9th ESMATS Symposium, Liège (2001).
6. Duband, L., Development of a Pulse Tube 50-80K Cryocooler – Final Report, CEA Document No. SBT/CT/99-15, ESA Contract #11331/95/NL/FG, Grenoble (1999).
7. Gibson, A.S, Hackett, J.D., Bailak, G., “Design of a Length Modulated Cell for Remote Sounding of Greenhouse Gases”, Proceedings of NASA 33rd Aerospace Mechanism Symposium, Cocoa Beach (1999).
Low Cost, Lightweight Space Cryocoolers C. S. Kirkconnell1, G. R. Pruitt1, K. D. Price1, B. A. Ross, Jr.2, and W. R. Derossett2 1
Raytheon Electronic Systems El Segundo, California 90245 2 Raytheon Infrared Systems Goleta, California 93117
ABSTRACT Raytheon has developed a concept for compact, lightweight space cryocoolers that merges the existing company expertise in tactical cryocoolers and space cryocoolers. The compressor is an upgrade to the existing Raytheon 705X tactical compressor product line in which the rubbing seals are eliminated through the incorporation of a non-contacting, flexure bearing piston support system characteristic of that used presently on the space cryocooler product line. To minimize cost and weight, the expander is a single-stage pulse tube. A concentric pulse tube configuration is used to simplify system integration by providing a distinct cold tip and radially symmetric structural stiffness. The cryocooler electronics module is essentially a radiation-hardened version of the existing tactical high reliability electronics design. The novel aspect of the proposed concept is the merging of the previously distinct tactical and space cryocooler technologies. The underlying technologies are essentially proven. Over 5,000 linear compressors of similar basic construction (motors, housing, etc.) have been built and fielded by Raytheon over the past ten years. Flexure bearing piston support systems have been employed on many past Raytheon space cryocooler designs (SSC, ISSC, PSC, SBIRS Low) as well as throughout the industry. Similarly, single-stage concentric pulse tubes have been built at Raytheon and elsewhere. More than 100 high reliability cryocooler electronics boxes have been fabricated and delivered to the customer community. The merging of these proven technologies yields a space cryocooler with recurring costs approximately a factor of ten lower than the present industry average of $2M. The projected weight for the combined cryocooler and electronics module is about 3 kg.
INTRODUCTION Space cryocoolers and tactical cryocoolers have to date been viewed as distinct technologies because of design and cost differences driven by the much more stringent space requirements. The price tags clearly substantiate this delineation. Tactical coolers typically have recurring costs in the $3000 to $15,000 range, while present generation space cryocoolers cost over $1.0M for the thermomechanical unit (TMU) alone and another approximately $1.0M for the control and drive electronics, costs based upon the typical industry order quantity of2 to 3 units. Table 1, which is a comparison of the specifications for typical Raytheon space and tactical cryocoolers, contains the key reCryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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quirements that drive the cost difference. These are primarily the lifetime, residual vibration, and radiation hardness requirements. The same factors that drive cost also drive the dramatic size and mass differences between present-generation space and tactical cryocoolers of similar capacity. Table 2 illustrates this fact by comparing the mass of tactical and space cryocooler modules of similar function and capacity. Therefore, the development of a low cost, lightweight cryocooler arose naturally from the dual long-standing goals of continually reducing cost and weight in the space cryocooler product line. The key features of Raytheon’s low cost, lightweight space cryocooler design are provided in the pages that follow.
SPACE CRYOCOOLER COST DRIVERS Marketplace The space cryocooler marketplace is characterized by small quantity orders and unique interface and environmental requirements. The inconsistency in requirements between different customers and payload applications necessitates the tailoring of existing designs for virtually every new program. This is in stark contract to the tactical marketplace’s Standard Army Dewar Assembly (SADA) Cryocooler specifications that, through the thorough qualification of a given design to a well-defined requirements set, allow the incorporation of a single cryocooler design into multiple systems. Regarding the order quantities, tactical cryocooler orders typically involved hundreds, perhaps thousands of units, which naturally enables cost savings in materials, reduced process time through incorporation of specialized subassembly stations, batch unit performance screening, etc. The low production quantities for space cryocoolers preclude many of these opportunities for savings, completing a vicious cycle in which high cost reduces demand for these units.
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Thermo Mechanical Unit (TMU) The requirements for long-life and high reliability have motivated the leading space cryocooler suppliers to adopt what has come to be called the “Oxford class” cryocooler design, that is, flexure bearing suspended pistons utilizing non-contacting clearance seals to separate the working volume from the plenum volume. These suspension systems necessitate precise mating structures to maintain alignment within approximately one hundred micro inches to preclude piston-cylinder contact. In contrast, the tactical cryocooler compressor pistons are supported by much less expensive machined springs and utilize contacting seals that wear over time, but in a generally predictable manner. The long life requirement necessitates stringent gas contamination control, which limits the types of epoxies and other volatiles that can be used. Raytheon has addressed this concern in past designs with hermetic motor enclosures that are effective, but add cost.1 The other significant cost driver for the TMU is the low residual, or operationally generated vibration requirement. This requirement is traditionally met with a closed-loop dynamic control system that requires knowledge and control of piston position; typically this requires a position sensor, such as a linear variable differential transducer (LVDT) for each piston. Vibration level feedback is provided by load cells or accelerometers. In the case of a Stirling expander or a singlepiston compressor design, an additional actively-driven balance mass assembly (mass, piston, motor, position sensor, etc.) must be added to offset the dynamics of the working pistons. The addition of these components adds significant cost and weight to the present state-of-the-art space cryocoolers. Tactical cryocoolers, with the less demanding vibration requirements, do not require piston position sensors and load cells, and Stirling expander vibrations can be sufficiently damped using comparatively small, passive pneumatic or spring-mass balancers.
Electronics The requirement for low residual vibration drives cost on the electronics module as well as on the TMU. Additional circuits are required to process the signals from the load cells or accelerometers and the piston position sensors, and hundreds of lines of code are required to implement the selected vibration control algorithm, such as the adaptive feed forward method used for Raytheon’s PSC.2 The complication of the vibration feedback control, as well as the more stringent temperature stability requirement, requires the use of an expensive microprocessor in the space cooler electronics that is not needed in the tactical electronics. Naturally, increased parts count relates directly to increased weight. The lifetime/reliability requirement results in the mandatory selection of more expensive high reliability components. On the tactical cryocooler line, Raytheon’s high reliability electronics module cost more than ten times the otherwise-similar Low Cost Cryocooler Electronics ($10K vs. <$1K), which provides an indication of how influential this requirement is on the much more complicated space cryocooler electronics modules. Finally, the requirement for radiation hard electronics, typically not requisite for tactical electronics, further increases component cost, and can also increase mechanical design complexity, cost, and weight if the housing is to be relied on for additional radiation shielding. In summary, the low vibration, high reliability, and radiation hardness requirements all contribute to the higher costs and larger size of space cryocooler electronics versus the tactical cryocooler counterpart.
“Hidden” Program Costs Easily overlooked but nevertheless significant contributors to the unit cost of a space cryocooler are the documentation and reporting requirements imposed by the customer. These include conformance to allowable parts and materials lists, which often drive expensive component and process replacement, piece part traceability and planning sheets for each unit, multiple design reviews, and numerous “consent to” meetings preceding key events like the start of manufacture, test, and delivery of each unit. A high level of unit-to-unit oversight is certainly warranted in the space cooler business because of the unit cost, the criticality of the cryocooler to the system performance, and the enormous expense or outright impossibility of fixing repairs or replacing units on orbit. However, these documentation and reporting expenses can easily account for more than 25% of the
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unit cost in a typical space cryocooler program. The hope and expectation is that as the space cryocooler technology and marketplace mature together, a more equitable balance between program management requirements and imposed expense will evolve.
PROGRESS TO DATE ON CONTROLLING SPACE CRYOCOOLER COSTS The most direct method by which to reduce recurring cost is to make design changes that selectively target the expensive features of the preceding design. For example, the next generation Raytheon Stirling/Pulse Tube Two-Stage (RSP2) Cryocooler incorporates many cost saving features such as non-hermetically sealed motors and LVDT coils and an improved design that reduces parts count and simplifies piston alignment. The novel RSP2 was developed in the course of the recent 95K High Efficiency Cryocooler Program and is discussed elsewhere in more detail.3,4 More aggressive design changes to further drive down cost are proposed in the next section. However, significant progress has been made already in reducing the recurring cost for the present build of a single-stage Stirling cryocooler called the RS1, very similar in design to the PSC cryocooler with respect to piston suspension, contamination control, total parts count, and total assembly and process steps.1 Raytheon is presently building fifteen (15) of these flight-qualified cryocoolers for an internal laboratory program, and this production quantity has made cost effective the implementation of a manufacturing facility based upon cost saving low rate initial production (LRIP) techniques such as batch processing, parts kitting, subassembly stations, and in process staging. The space cryocooler manufacturing facility that will be used for the present build, scheduled for completion by August ’02, is depicted in Figure 1. Cost savings have been realized due to reduced costs at the component level by virtue of the higher order quantities of each piece part. Well before the end of the production build, the combined recurring cost of the TMU and the flight-design electronics module is projected to be less than $1.0M compared to the approximately $2.0M industry average for small lot (< 4 units) orders. The accomplishment of this cost target is particularly noteworthy because the design of the RS1 does not incorporate many of the cost saving features of newer designs. Application of these manufacturing techniques and utilization of this state-of-the art space cryocooler manufacturing facility will enable even larger savings when applied to a low cost design such as that described in the next section.
RAYTHEON LOW COST, LIGHTWEIGHT CRYOCOOLER Basic Concept Raytheon is in the singular position of being an established supplier of both tactical and space cryocoolers. Since 1978 Raytheon has delivered over 50,000 rotary and linear tactical cryocoolers,
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and that line continues today with production of cryocoolers for AIM-9X, Maverick, and several other programs. The Raytheon space cryocooler experience also stretches back over 30 years, starting with several Vuilleumier cryocoolers built and delivered in the early 1980’s to the flightqualified Stirling cycle PSC and SBIRS Low cryocoolers of the 1990’s. As discussed previously, fifteen (15) cryocoolers of the flight-qualified RS1 design are presently in production. The basic concept for the low cost space cryocooler evolved from this dual space and tactical cryocooler legacy. The compressor module is based upon an existing tactical linear compressor with the machined spring suspension system replaced with a flexure bearing suspension system with legacy to the space product line. The expander is a single-stage pulse tube, building upon past experience with several IR&D programs5-7 and the commercial Low Cost Cryocooler (LCC) program.8 The integrated compressor-expander assembly (the TMU) is shown in Figure 2. The electronics module is a radiation hardened version of the current high reliability tactical cryocooler electronics package. In short, the low cost space cryocooler approach is to baseline a tactical cryocooler and incorporate only those features from the space cryocooler line that are required to meet the critical mission objectives. As noted earlier, the space cryocooler marketplace requirements are varied and broad in scope. The more the requirements trend towards the demanding end of the spectrum, particularly with respect to cost driving technical features like residual vibration control, the less amenable the design will be to controlling cost. For the present purposes, it was necessary to establish a baseline requirements set to which the design concept must conform so that the applicability of the concept to the general marketplace was evident. Those top level specifications, which were derived from a combination of various tactical and space cryocooler specification documents, are provided in Table 3. Given the focus of the effort on controlling cost and weight, these baseline requirements are representative of the type of space mission more conducive to this low cost approach, that is, moderately stressing in terms of refrigeration temperature, capacity, and vibration control.
Compressor Module The compressor is a dual opposed piston linear compressor with direct legacy to the existing tactical cryocooler product line. The primary difference is that the machined springs and rubbing seals are replaced with flexure bearings and clearance seals to provide the required life and reliability. The change to a flexure suspension system necessitated changes in the motor design. Aggressive size and mass requirements were imposed to make the compressor compatible for various long
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life, high reliability tactical missions of interest. The result is a 2.5 cc swept volume compressor with a mass of 1.8 kg, length of 12.5 cm, and a 6.0 cm diameter. For comparison, the present RS1 compressor design with similar thermodynamic capacity is 8.1 kg. Preliminary analysis has been performed to verify that the flexure stress state meets the targeted 20,000 hour tactical life requirement. Additional analysis is planned to identify what modifications will be required to extend the flexure design life to the 50,000 hour low cost space cryocooler target. The impact on the overall unit size and mass is expected to be small. To control vibration to first order, the moving masses flexure stack spring rates of the opposed piston assemblies are matched. The low cost space cryocooler compressor is essentially a tactical compressor that includes strategically selected upgrades from the space cryocooler line. The flexure design has legacy to the space cryocooler line and the extensive life test data at both the cryocooler level (PSC, >20,000 hours; SBIRS Low, >30,000 hours) and the flexure suspension subsystem level (>8+ years on the SBIRS Low /RS1 compressor flexure suspension system design). Like the existing tactical designs, the motor housings are non-hermetic, but unlike most tactical applications, space cryocoolers tend to operate continuously and are thus more susceptible to long term performance degradation due to freeze out of volatile contaminants. The same material selection and bake out procedures that were established for the non-hermetic motors on RSP2 will therefore be applied to the present effort to minimize volatile evolution. These measures are more extreme than those required on the tactical line. In general, however, tactical cryocooler design practice has been applied wherever possible.
Expander Module The expander is a single-stage pulse tube to save the added cost of the active-drive Stirling displacer and the additional cost and weight of the active-drive balance mass assembly. (The requirement cannot be met with the existing tactical passively-balanced Stirling expander design approach.) The pulse tube cold head is a concentric pulse tube, that is, the regenerator is contained in an annular volume around the pulse tube which yields a radially symmetric cold finger. The concept is illustrated in Figure 3. This approach was taken versus the competing linear and U-tube arrangements to provide optimum system integration features for the user. If cost is a concern at the cryocooler level for a program, it will certainly be a concern at the system level, and complex interfaces increase integration costs. The linear pulse tube design, used extensively throughout industry and often preferred because of the efficiency advantages afforded by its characteristic low void volumes and simple flow paths, is notoriously difficult to integrate because of the location of the cryogenic interface midway down the cold finger, sandwiched between ambient structure at
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both ends. The U-tube and the concentric pulse tube are easier to integrate because, like a Stirling, their cryogenic interface is at the end of the cold finger, providing a distinct “cold tip” that can be conveniently accessed, inserted into tight cryogenic spaces, etc. The concentric pulse tube was selected as the baseline because it is slightly more compact than the U-tube, and it has radially symmetric stiffness. In contrast, the U-tube has a soft axis and a stiff axis, and the impact of this asymmetric load bearing capability must be considered at the system integration level. However, a U-tube configuration can be used, if desired, and all of the first order cost and weight advantages of the basic low cost cryocooler design are still realized.
Electronics The electronics design is a radiation hardened version based on the existing RIO tactical electronics module 416301 and the FPGA controller design of later LCCE and PAWS-II units, which is made up exclusively of high reliability piece parts so the required component upgrade is only with respect to the radiation requirement. Detailed parts selection has not yet occurred, but it is evident from both the legacy RIO 416301 design and the present space cryocooler electronics design that component cost will drive the total module cost. Component-level and box-level shielding will be traded against radiation hard component cost to identify the combination of radiation hard design approaches that optimally balances cost, survivability, and weight. Low cost tactical electronics design, assembly, and test practices will be used to the greatest extent possible, though the performance and reliability screening will obviously have to be more extensive for space applications. The cost of the electronics is significantly lower than present space cryocooler electronics primarily because the vibration level is controlled to first order through design (dual-opposed noncontacting pistons, pulse tube expander with no moving parts, etc.) and manufacture (matched moving masses, carefully screening motor magnets for field uniformity, etc.). Therefore, the electronics simply provide the fine tuning through unit-specific motor drive parameters that are set when the electronics and the TMU are first integrated. This approach precludes the need for the extensive circuitry and software presently used for space cryocooler electronics to monitor and actively control the vibration level. Preliminary analysis indicates that vibration levels below 350 mN can be readily achieved by this approach; achieving the 250 mN target level over the life of the unit may require some limited capability to adjust the motor drive parameters on orbit.
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CONCLUDING REMARKS Raytheon has developed a concept for a low cost, lightweight space cryocooler that meets the requirements for many present and near term missions. The proposed design combines an upgraded tactical compressor with flexure-suspended pistons to a concentric pulse tube expander. The cryocooler electronics are a radiation hardened version of the present high reliability design. A representative requirements set was defined for the purpose of establishing a baseline. The extent to which these cost cutting, light-weighting measures can be applied for a real system will depend on the comparative values of the key cost driving requirements to those assumed herein. For the baseline requirements set, the estimated recurring costs are $80K for the mechanical cryocooler and $60K for the electronics. Given these module costs and projecting that “hidden” program costs will naturally decline as quantities increase and the technology matures, the integrated and tested complete cryocooler system can be delivered for $200K. The projected weight of 3 kg for the combined cryocooler and electronics module is about 1/3 of the present state of the art for a cryocooler of comparable refrigeration capacity.
REFERENCES 1.
Price, K.D., Barr, M.C., and Kramer, G., “Prototype Spacecraft Cryocooler Progress,” Cryocoolers 9, Plenum Publishers, New York (1997), pp. 29-34.
2.
Price, K. Reilly, J., Abhyankar, N., and Tomlinson, B., “Protoflight Spacecraft Cryocooler Performance Results,” Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2001), pp. 35-43.
3.
Price, K. and Urbancek, V., “95 K High Efficiency Cryocooler Program,” Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2001), pp. 183-188.
4.
Kirkconnell, C.S., Price, K.D., Barr, M.C., and Russo, J.T., “A Novel Multi-Stage Expander Concept,” Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2001), pp. 259-263.
5.
Kirkconnell, C.S., Soloski, S.C., and Price, K.D., “Experiments on the Effects of Pulse Tube Geometry on PTR Performance,” Cryocoolers 9, Plenum Publishers, New York (1997), pp. 285-293.
6.
Kirkconnell, C.S., “Experiments on the Thermodynamic Performance of a ‘U-Tube’ Pulse Tube Expander,” Adv. in Cryogenic Engineering, Vol. 43B, Plenum Publishing Corp., New York (1998), pp. 1973-1981.
7.
Kirkconnell, C.S., “Experimental Investigation of a Unique Pulse Tube Expander Design,” Cryocoolers 10, Plenum Publishing Corp., New York (1999), pp. 239-247.
8.
Russo, S.C, and Pruitt, G.R., “Development of a Low-Cost Cryocooler for HTS Applications,” Cryocoolers 9, Plenum Publishers, New York (1997), pp. 229-237.
JAMI Flight Pulse Tube Cooler System J. Raab, R. Colbert, J. Godden, D. Harvey, R. Orsini, G. Toma TRW Redondo Beach, CA 90278
ABSTRACT The Japanese Advanced Meteorological Imager (JAMI) Flight Pulse Tube Cooler System features two integral pulse tube cryocoolers. They are configured to be standby redundant and control the temperature of two focal planes. The JAMI mission acquires multi-spectral imagery for operational weather needs over Japan, East Asia, and Australia. JAMI is scheduled to fly on the Multifunctional Transport Satellite (MTSAT) in summer of 2003. The program delivered the two flight coolers and electronics in December 2001. This paper presents data collected on the flight coolers during acceptance testing. Tests included thermal performance mapping at various power levels, temperature stability tests, and launch vibration testing. Designed conservatively for a 10-year life, the coolers are required to provide 1.75W cooling at 67K while rejecting to 300K with less than 90W input power to the electronics. At these loads and temperatures the cooler has an additional 100% margin. The total mass of each cooler and electronics system is 10.4 kg (20.8 kg as a redundant pair). The radiation-hardened software driven control electronics provides cooler control functions which are fully re-configurable in orbit. These functions include precision temperature control to better than 25mK p-p.
INTRODUCTION The JAMI cooler program delivered two flight cooler systems and one engineering model cooler system, plus ground support electronics (GSE) to interface with the cooler system or mechanical cooler. The program was performed for Raytheon Electronics Systems in Goleta California over a 15-month period. The JAMI mechanical cooler is shown in Figure 1. The cooler provides focal plane array (FPA) cooling via a thermal strap and rejects heat to a radiator. The JAMI cooler system, which is the latest version of the TRW 65 series coolers, consists of the mechanical pulse tube (MPT) cooler with attached accelerometer electronics, and separately, the cooler control electronics (CCE). The mechanical cooler is a build to print of the 95K HEC (High Efficiency Cryocooler). The electronics is the same basic design as the Tropospheric Emission Spectrometer (TES)1,2 and Hyperion (currently in orbit) cooler electronics, except that the input voltage range was modified to 42VDc nominal from 28VDc nominal. Before installation and operation of the cooler on the instrument, both the mechanical and the electronics assemblies together with the operating software underwent flight level acceptance testing, including environmental tests of launch vibration, thermal vacuum cycling, and burn-in. These tests, which are typical for space instruments, are performed to ensure reliability. The cooler perforCryocoolers 12. edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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mance, including load lines, and temperature stability properties, was measured. This paper reports the test data for one of the flight coolers. There was less than 5% input power difference from flight unit to unit at the nominal operating condition of 1.75W at 67K.
COOLER SYSTEM The mechanical cooler (Figure 1) refrigerates via the cold block, rejecting heat at the centerplate of the compressor. Inside the compressor, flexure springs support the moving-coil linear motor, which drives the pistons. The springs maintain alignment for the attached non-contacting piston that oscillates and compresses gas into the pulse tube cold head. A small clearance between the cylinder and the piston seals the compression space. Two opposed compressor halves vibrationally balance the compressor. The compressor is operated at near resonant frequency of 70 Hz. The pulse tube cold head is bolted to the compressor centerplate and is sealed with a metal seal. The centerplate conducts heat to the radiator. The cold head components are arranged linearly: mounting flange, regenerator, cold block, pulse tube, and warm-end heat exchanger body (or orifice block). The cold head is surrounded by an H-bar that supports and provides a thermal path to remove heat from the orifice block. The stainless steel orifice line connects the gas from the orifice block to the reservoir tank. The cooler envelope is shown in Figure 2. The internal wiring in the compressor is stranded, PTFE (cross-linked Teflon), insulated wire. All wiring exits the centerplate through ceramic-insulated pins in a D-shell feedthrough connector for the cooler drive power. A separate connector is used for the redundant platinum resistance thermometers (PRTs) on the cold block and thermistor on the centerplate. An accelerometer is mounted on the compressor centerplate. Together with the signal conditioning electronics, the accelerometer provides a feedback signal to the vibration control algorithm in the drive electronics (CCE).
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The CCE (Figure 3) is based on our high-reliability Hyperion/Tropospheric Emission Spectrometer (TES) flight design modified for 42VDc bus voltage. There are three slice subassemblies, one for control (control slice), one for power amplifiers (power slice), and one for power conversion (converter slice). The slices are housed in a standard subassembly that is 225 mm (L) x 216 mm (W) x 175 mm (H). The bottom of the housing serves as a mounting surface for direct thermal contact. The electronics in the CCE (1) converts the 42 Vdc primary power to the secondary power, (2) drives the cooler, and (3) provides communication with the host and control of the cooler with a processor using software resident in PROM. The software performs the following functions. Transmits spacecraft command and cooler telemetry via the RS422 data bus Collects the cooler state of health data Controls the cold block temperature Actively balances vibration force by controlling the waveform of the pistons Provides safety protection to the cooler
COOLER OPERATION AND CAPABILITIES Table 1 summarizes system weight and capabilities. The cooler electronics provide 70 Hz drive electrical power to the motors in the compressor. The compressor moving coil and piston assembly resonates on the gas and mechanical springs to produce a 70 Hz pressure wave and mass flow to the cold head. The software adjusts the motor drive to maintain the desired cold block temperature. The vibration control algorithm samples the accelerometer signal and determines, by Fourier analysis, transfer gains and error signals for up to 16 harmonic frequencies. The error signal modifies the motor drive waveform to reduce vibration.
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Figure 4 shows the input power for various heat loads at 67K about the nominal heat load requirement of 1.75W. Figure 5 shows the cooling load as a function of cooling temperature for different reject temperatures, and input powers. The cooler exhibits sensitivity in input power for a constant frequency and variable reject temperature. The sensitivity can allow cooler operation that results in minimizing the power input at various reject temperatures over the cooler operating range. The sensitivity is present due to a fixed inertance line tuning that is established at nominally room temperature and optimized for a desired frequency.
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The CCE (Figure 3) plays a critical role in the overall cooler performance. When the output power to the compressor as a function of the input bus power was measured, its efficiency fit the correlation of where the efficiency, and an electronics tare, Hyperion3 incorporated a bus ripple control approach developed on IMAS.4 For the JAMI cooler operating conditions the bus ripple current is 8.7% p-p of the steady state DC current.
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The temperature control algorithm adjusts the stroke level based on the difference between the cold block PRT temperature and the set point temperature value. Figure 6a shows that with a constant heat rejection temperature the cold head temperature stability is within a 20mK bandwith for a constant load at 67K. Similarly, Figure 6b shows that the temperature stability of the cooler operating with a constant load at 67K during a heat rejection temperature change of 0.2°C /min is within a 25 mK band.
ENVIRONMENTAL TESTS AND COOLER ACCEPTANCE The JAMI cooler system acceptance testing included launch random vibration, thermal vacuum testing with operating and non-operating temperature cycles, and burn-in. Levels and ranges for these tests are summarized in Table 1. Repeatable cooler performance after each environmental test is used as an acceptance criterion. The cooler was accepted because no performance change of the load line was detected within experimental uncertainty. The measured helium leak rate was one order of magnitude less than the 10-year-life requirement.
CONCLUSIONS The JAMI cryocooler performance met the program requirements. The coolers were delivered in October 2001 and are being integrated with the imager.
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ACKNOWLEDGMENT The work described in this paper was carried out by TRW under contract to Raytheon Electronic Systems JAMI Project.
REFERENCES 1.
Raab, J., et al., "TES FPC Flight Pulse Tube Cooler System," Cryocoolers 11, Kluwer Academic/ Plenum Publishers, New York, 2001, pp. l31-138.
2.
Collins, S.A., Rodriguez, J.I. and Ross, R.G., Jr., “TES Cryocooler System Design and Development,” Adv. in Cryogenic Engineering, Vol 47B, American Institute of Physics, New York, 2002, pp. 10531060.
3.
Chan, C.K., Clancy P., and Godden J., “Pulse Tube Cooler for Flight Hyperspectral Imaging,” Cryogenics 39, Elsevier Ltd, (1999), pp. 1007-1014.
4.
Chan, C.K., Nguyen T., Colbert R., Raab J., Ross R.G., Jr., and Johnson D.L. “IMAS Pulse Tube Cooler Development and Testing,” Cryocoolers 10, Kluwer Academic/Plenum Publishers (1999), pp. 139-147.
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Performance Testing of a Lightweight, High Efficiency Cooler Salerno, L. J.1, Kittel, P.1, Helvensteijn, B. P. M.2, and Kashani, A.2 1
NASA Ames Research Center Moffett Field, CA, USA 94035
2
Atlas Scientific San Jose, C A, US A 95123
ABSTRACT A lightweight, high efficiency cryocooler has been tested. The design operating point of the cooler is 10 W at 95 K with a rejection temperature of 300 K, and the goal is a 10 yr operating lifetime. The cooler has a mass of less than 4.0 kg. The cooler efficiency at the design operating point is 12.6 W/W. These values represent approximately a one-third decrease in mass and a one-third increase in efficiency over previous designs. The pulse tube is built by TRW, and the compressor is built by Hymatic Engineering, UK. TRW integrates the compressor with the pulse tube and performs final testing of the assembly. Development of the cooler was performed by TRW under a joint NASA-DOD program. Data presented were taken at controlled rejection temperatures of 285 K and 300 K using a cold-water heat exchanger bolted to the cooler. Heat loads were varied between 1W and 10 W by supplying current to a resistor mounted on the cooler cold block. Input power to the compressor was limited to 180 W.
INTRODUCTION Both NASA and the DoD have the need for lightweight, high-efficiency, flight-qualified cryocoolers. NASA is interested in flight coolers for both space transportation systems and scientific instruments. One possible application is zero boil-off (ZBO) cryogenic propellant storage, a method of extending the lifetime and minimizing the size of propellant tanks in longterm space exploration missions. The concept and details of ZBO have been presented elsewhere and will not be covered here1-2. In March of 1998, NASA Ames teamed with the Air Force Research Laboratory/VSSS at Kirtland Air Force Base in Albuquerque, NM to jointly develop a 95 K pulse tube cooler. The goals of the development were to improve cooler efficiency, lower mass and extend lifetime. A contract was subsequently awarded to TRW and two coolers were procured. Development was funded by the Air Force and NASA received a duplicate unit. The coolers were delivered in June of 2001. The cooler delivered to the Air Force was vibration tested to flight qualification levels (17.3 G rms), and the NASA cooler was tested to acceptance levels (8.65 G rms).
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This paper presents the results of testing performed on the NASA cooler (numbered HEC 201) at Ames Research Center. The cooler is shown in Figure 1. It has a mass of less than 4.0 kg. Full details of the cooler have already been presented and will not be repeated here3 however a brief overview will be provided. The cooler is an integral configuration pulse tube type and employs a horizontally opposed flexure bearing compressor and an in-line pulse tube cold head. The radially stiff flexures in the compressor eliminate piston wear. Integration is simplified by the single compressor center plate which also serves as a thermal interface for rejecting the heat of compression. The cold head is supported against launch loads by an aluminum support structure (H-bar). The gold-plated copper cold block is located near the midpoint of the pulse tube. Calibrated platinum resistance thermometers (PRT) are provided in the cold head. The cooler has been shipped to the NASA John Glenn Research Center (GRC) in Cleveland, OH, for hardware design to support ZBO preliminary system testing at GRC, followed by future integration into a ZBO testbed at the George C. Marshall Space Flight Center in Huntsville, AL. The eventual goal is a complete ZBO system demonstration in a flight-like configuration.
EXPERIMENTAL SETUP Figure 2 shows the experimental setup, and Figure 3 shows details of the cryostat. The heater (a 25.6 ohm, 5W resistor), was mounted on the cooler cold block with two M3 bolts,
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torqued to 0.84 ±.08 N-m (7.4 +/-0.7 in-lbs). Within the cryostat, current was supplied to the heater with two 0.018 cm (0.007”) diameter copper wires and heater voltage was sensed with two 0.025 cm (0.010”) diameter manganin wires. Length of the wires within the cryostat was approximately 76 cm (30”). The parasitic heat loss from the wires was calculated at 6.6 mW. After resistor insulation and wiring an MLI blanket was prepared, consisting of 16 layers of
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aluminized mylar. Each layer was crinkled to minimize heat transfer prior to making up the blanket. The blanket was then wrapped around the cooler cold block and resistor, taking care to assure that the ends were sealed as well. The cooler was then mounted onto a water cooled copper heat exchanger with six M5 bolts torqued to 4.0 + 0.3 N-m (35 ± 3 in-lbs) and the assembly was attached to the cryostat and installed in a Dewar of nominal internal dimensions of 34.9 cm (13.75”) diameter and 39.9 cm (15.75”) height. After installation and connection of water lines to the cryostat, the assembly was helium leak checked, water tested, and then leak checked again. In operation, the heat exchanger was connected to an external chiller to maintain a rejection temperature of either 285 K or 300 K, measured at the compressor center plate. An ionization gauge was installed on the cryostat to monitor the Dewar vacuum. Three electrical feedthroughs were provided on the cryostat. Two provided for control and monitoring of the cooler by means of the TRW supplied electronics (communication interface and control unit), and the third was for control and monitoring of the heater. A Valhalla Scientific 21010 Digital Power analyzer measured the power delivered to the cooler. A Leybold turbomolecular pump was used to evacuate the Dewar, and a Leybold Infinicon IG 3 gauge monitored the vacuum. The cooler was tested vertically with the cold end of the pulse tube down to eliminate convection effects during test.
RESULTS Figures 4 and 5 show the plots generated from the test data at the nominal rejection temperatures of 300 K and 285 K respectively. Data of heater power, cold block temperature, and compressor power were taken. At a given rejection temperature (either 285 K or 300K),
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heater powers of 0, 1, 5, and 10 W were applied and compressor power was measured. From these data isotherms were generated at the temperatures labeled in the figures. No extrapolation of data was performed, however the continuous isotherms were interpolated. Figure 6 is a series of load curves at 300 K rejection temperature, obtained by plotting cooling power vs. cold block temperature with compressor input power as a parameter. The lines are linear fits and the points are interpolated. As part of the data taken for developing the plots, TRW’s design point performance was measured and verified. Test conditions for verifying the design point performance were a measured compressor power of 126 W with a 10W applied heat load at a 95 K cold block temperature at a rejection temperature of 300 K. This translates to 12.6 W/W, which isTRW’s previously measured performance. An uncertainty analysis indicated that the uncertainty in the above reported measurement considering the accuracy of reading both the applied heater power and the compressor power is 0.3 W/W, or on the order of 3%. Since the calibration accuracy of the thermometers in the cold head is ± 20 mK, the uncertainty contributed from the temperature measurement is negligible.
SUMMARY Experimental data have been presented for a lightweight high efficiency cryocooler for space flight applications. The cooler represents a significant advance in the state of the art and should find a variety of applications in advanced space transportation systems, including ZBO
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cryogen storage, and science instrument cooling. Future plans call for NASA to demonstrate this cooler as part of a complete ZBO system demonstration in a flight-like configuration.
REFERENCES 1. Salerno, L. J., Valentian, D., Plachta, D., Kittel, P., Hastings, L. J., “Zero Boil off Cryogen Storage for rd Future Launchers”, Proc. of the 3 European Conference on Launcher Technology, Strasbourg, France, Dec 11-14, 2001, pp. 399 2. Hastings, L. J., Plachta, D., Salerno, L., and Kittel, P., “An Overview of NASA Efforts on Zero Boiloff Storage of Cryogenic Propellents”, presented at the Space Cryogenics Workshop, July 2001, Milwaukee, Wisconsin , Cryogenics, vol. 41, no.12, (2001), pp. 833-839 3. Tward, E., Chan, C. K., Colbert, R., Jaco, C., Nguyen, T., Orsini, R., Raab, J., “High Efficiency Cryocooler”, Adv. in Cryo Eng, 47, American Institute of Physics, Melville, New York (2002), pp. 10771084
Development of a Lightweight Pulse Tube Cryocooler for Space Applications T. Nast, J. Olson, P. Champagne, B. Evtimov Lockheed Martin Advanced Technology Center, Palo Alto, Ca. Todd Renna Lockheed Martin Commercial Space Center, Newtown Pa. G. Sarri and C. Gomez Hernandez European Space Agency (ESA), ESTEC, Noordwijk, The Netherlands
ABSTRACT Lockheed Martin has developed a high power version of its mini pulse tube cryocooler to provide an increased cooling capacity (2.7 W at 80 K). This unit was developed under contract to the European Space Center (ESA) as a potential backup for their INTEGRAL experiment. The design approach was to utilize the compressor system from a previously developed mini pulse tube system and increase piston size for higher swept volume and re-optimize the cold head for higher cooling capability. Prior development at LM proved the viability of this approach under company funded and DARPA funded programs. An engineering model of both the cryocooler and its drive electronics was developed and tested at LM, then delivered to ESA for their test program. This paper summarizes the system characteristics and presents the results of test programs at both LM and ESA.
INTRODUCTION Numerous future spaceflight missions require a very lightweight, compact cryocooler system with lifetimes of 10 years or more. Cryocooler customers also want lower costs and shorter delivery times. With the requirement for lifetimes in excess of 10 years, LM ATC selected the no-movingparts pulse tube coldhead driven by a flexure-bearing clearance seal compressor. At LM ATC we have achieved pulse tube efficiencies comparable to Stirlings1-4, such that the cooling requirements of the ESA contract could be efficiently met with a pulse tube. The flexure-bearing compressor has now been demonstrated to be a reliable, robust technology. For this program, LM ATC uses a moving magnet compressor that has simplified assembly, reduced cost and enhanced reliability over the standard Oxford-heritage compressor with moving coil. ESA awarded a contract to LM to develop an EM Cryocooler for potential use in their INTEGRAL program with a requirement of 3 W of cooling at 80 K. The ESA contract calls for the delivery of 1 EM complete with Cryocooler Drive Electronics (CDE).
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The development and the manufacture of the EM CDE were performed by Lockheed Martin’s Commercial Spacecraft Center. An Engineering Model of this system has been delivered to ESA and they have performed system testing in support of their INTEGRAL system, the results of which are reported here.
SYSTEM DESCRIPTION Cryocooler Thermo-Mechanical Unit (TMU) The thermo-mechanical unit is a split system with a short transfer line to the cold head assembly. The compressor utilizes a linear motor, with a dual-opposed configuration for momentum compensation. Figure 1 shows the general configuration of the system, which includes the electronic controller. All components of the cryocooler are packaged within the configuration shown. No additional hardware is required. The system is shown mounted on a fixture that aligns the compressor to the cold head assembly. Both the compressor and the cold head have separate structural mounting and heat rejection. The second generation CDE is also shown. The U-tube coldhead typically simplifies integration to the instrument, reduces overall weight, and eliminates the need for a second heat rejection point at the warm end of the pulse tube as required by in-line designs. The U-tube is also structurally more rigid than the in-line coldhead, making it less susceptible to damage from launch vibrations and instrument side-load forces. The side load capability of our cold tip in the weakest of two lateral axes is predicted to be 6.5 kg. Figure 2 shows more detail of the TMU. The compressor is designed for low manufacturing costs while still maintaining the reliability of the flexure-bearing compressor with non-contacting piston-cylinder seals. The architecture is based on a larger compressor originally developed by LM ATC under NASA funding for a low cost commercial cryocooler.5 This approach utilizes a moving magnet design, with the drive coil external to the working gas space. By placing the coil outside of the pressure vessel, we eliminate the single largest source of contamination of the working gas, the organic coil potting. Likewise, the position sensor’s electrically active element is also outside of the pressure vessel, and thus electrical penetrations through the pressure wall are completely eliminated, removing gas leakage through the electrical feed-through as a potential failure mode. The stationary coil also eliminates breakage of flexing leads of moving coil motors as a failure mode. The compressor incorporates several self-aligning features for the piston/motor/flexure assembly, and a low piece-part count that simplifies the assembly and shortens the assembly time. The critical piston-cylinder seal utilizes a simple alignment adjustment mechanism that rapidly and repeatably performs this task. This mechanism can be computer automated. These compressor
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features will reduce costs in large-volume manufacturing. In small quantities, they enhance reliability by reducing the risk of workmanship defects. Table 1 summarizes these features.
SYSTEM PERFORMANCE AT LOCKHEED MARTIN Thermo-mechanical Unit Performance tests have been conducted on the EM unit at various power inputs and heat rejection temperatures. Figure 3 presents the load lines at several power inputs for heat rejection temperatures, while Table 2 summarizes the EM parameters against the ESA contract specification.
Electronic Controller The cooler drive electronics (CDE) is comprised of two major subassemblies: (1) Control Board Assembly, and (2) Power Board Assembly. Major features of the design are described below for each subassembly. Control Board Assembly. The control board contains the RS422 command and telemetry to the spacecraft (user interface), the temperature control function, and piston position limiting, and signal conditioning circuits required to drive the two PWM motor drive amplifiers. In addition, optional vibration cancellation control circuits (AFFECS – Analog Feedforward Error Correction System) are included in the circuit board design and can be configured in manufacturing depending on the level of vibration cancellation required. Most of the control features are embedded in an FPGA. A UART is included as part of the RS422 interface. Other options include a variable compressor drive frequency for optimization of cooler performance, and to meet specific customer requirements.
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Power Board Assembly Motor Drive Amplifiers. The compressor motors are driven using a high-frequency PWM amplifier. There are two identical amplifiers per assembly. The amplifiers include control circuit interfaces to the control board that enable accurate current control of the compressors in order to eliminate influence of the +28 V bus on the compressor power and also to allow for accurate harmonic cancellation. At the compressor interface, active clamp circuits are included for use during launch to minimize compressor piston excursion by providing magnetic damping. Power Supply. The power supply sections contain a PWM DC-DC converter to provide the necessary bias voltages to the entire CDE. The converter design is optimized for reduced power consumption to help minimize the overall bias power required for the CDE. As part of the power supply, section, the +28 V input bus is filtered using an LC-filter to reduce both differential and common-mode noise within the specified EMC limits. There is an option to include an inrush current limited (may not be necessary due to external bus inrush limiters). Mechanical Packaging. The CDE is housed in an aluminum chassis in an H-Frame configuration. The control board is bolt-mounted into one cavity of the chassis, while the power board is bolted into the opposing cavity. This separation isolates the control circuits from any EMI from the PWM’s and power supply. The two cavities are enclosed using an aluminum cover design specifically to reduce radiated emissions. Each double-sided printed circuit board utilizes both throughhole and surface mount components. The entire assembly was designed to optimize use of board area leading to a minimum overall weight. A summary of the system characteristics is shown in Table 3.
SYSTEM TESTING AT EUROPEAN SPACE AGENCY System Performances at ESA In the reference flight configuration two Pulse Tube Coolers are integrated on the instrument using a common support structure that also provides the thermal and mechanical interfaces to a
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radiator. Via the radiator the heat dissipated by the compressors is rejected to space and a cold heat sink is guaranteed to the pulse tube flange. The cold tips of the two pulse tubes are connected via a common beryllium interface element to the instrument cold bar, which directly interfaces with the detector to be cooled. Figure 4 shows a schematic of the chosen accommodation concept. The purpose of the system test at ESA was to verify the correct mechanical coupling of all the elements (support structure, two compressors, pulse tube interface flanges, beryllium block) and to characterize the thermal performance of the system. In addition, due to the fact that one option was to operate the system in cold redundancy (one cooler ON and the other OFF), a second important objective of the test was to measure the parasitic heat leak induced by the not operative cooler. A flight standard support structure and radiator were manufactured and a pulse tube mock-up was used to simulate the second, cold redundant unit. The interfacing beryllium element was also provided. The EM cooler cold head flange and the compressor foot were bolted onto the support structure. Black Kapton tape was applied to the compressor heads in order to improve the compressor heat rejection. Thermal blankets were also fitted on the radiator backside and inside the cold enclosure where the cold tip of the pulse tube is positioned. Finally, the assembly was bolted on the backside of the radiator. A thermal blanket was also fitted all around with the exception of the radiator heat rejection area. Figure 5 shows the details of the cooler integrated on the support structure. The set up was suspended using stainless steel cables inside a small thermal chamber at ESTEC. In order to simulate the in-orbit environment, the radiator was exposed inside the chamber, in vacuum, to a liquid nitrogen shroud. The radiative coupling between the compressor and the instrument was simulated by keeping the local compressor environment in the temperature range 0 °C to 30 °C. The heat load generated by the detector was provided by a heater mounted on the beryllium
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block area. During the tests the cooler was either operated with laboratory rack-mounted electronics or the cooler drive electronics. The results of the pulse tube system test are summarized in Fig. 6. The data correspond to the compressor driven by the laboratory drive electronics. Table 4 shows the parasitic heat load measured in the various test cases. The total cooler heat lift is the sum of the detector heat load (given by the heater) and the parasitic heat loads. These are due to the second pulse tube head, the radiative leak to the enclosure, the cables, etc. It can be seen that the parasitic heat load is a consistent fraction of the total cooler heat lift. This is mainly the consequence of having the second cooler in cold redundancy. The alternative approach is of course to operate the two coolers simultaneously at a lower compressor stroke (hot redundancy). This would avoid the heat leak via the second cooler. To limit the mechanical stresses on the detector interface the thermal coupling between the pulse tube cold tip and the beryllium block was via a flexible link made of four copper straps. The additional temperature drop associated with this solution was assessed. Figure 7 shows the copper braid conductance versus its average temperature. Cooler performance is normally declared as a function of the input power to the compressor. However, in a real satellite there are two basic parameters that are key elements for sizing and verification of the power bus: 1) the input power, and 2) the input current profile to the electronics. The input current is particularly critical for a cooler because of the sinusoidal shape and therefore the big difference between average input and peak input. The latter has to be considered for the sizing of solar arrays and batteries. The pulse tube cooler peak current at the input of the cooler drive electronics was measured in nominal operation. The results are shown in Figure 8.
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SUMMARY Lockheed Martin’s Advanced Technology Center, under contract from ESTEC has developed a miniature, lightweight pulse tube cryocooler for space applications that has a cooling capability of 2.5 W at 80 K. The contract called for delivery of an Engineering Model of both the mechanical system and the electronics. The system was delivered to ESA for evaluation testing and the system was successfully tested.
ACKNOWLEDGMENTS This work was conducted under ESTEC Contract No. 14533/00/NL/MS. Much help and guidance in system requirements and integration was received from G. Sarri and C. Gomez of ESTEC.
REFERENCES 1.
D.L. Glaister, M. Donabedian, and D.T. Curran, “An Overview of the Performance and Maturity of Long Life Cryocoolers for Space Applications,” Aerospace Report No. TOR-98 (1057)-3 (1998).
2.
V. Kotsubo, J. R. Olson, and T.C. Nast, “Development of a 2 W at 60 K Pulse Tube Cryocooler for Spaceborne Operation,” Cryocoolers 10, Plenum Publishing Corp., New York, 1999, pp. 157-161.
3.
T.C. Nast, P. Champagne, J.R. Olson, and V. Kotsubo, “Development of Pulse Tube Cryocoolers for HTS Satellite Communications,” Cryocoolers 10, Plenum Publishing Corp., New York, 1999, pp. 171-179.
4.
W.W. Burt and C.K. Chan, “New Mid-Size High Efficiency Pulse Tube Coolers,” Cryocoolers 9, Plenum Press, New York, (1997), pp. 173-182.
5.
T. Nast, P. Champagne, and V. Kotsubo, “Development of a Low-Cost Unlimited Life Cryocooler for Commercial Applications,” Adv. Cryo. Eng., 43a, Plenum Press, New York, (1998), p. 2047.
Development of a Two-Stage Pulse Tube Cryocooler for 35 K Cooling T. C. Nast, J. Olson, B. Evtimov and V. Kotsubo1 Lockheed Martin Advanced Technology Center Palo Alto, CA 94304-1191 1
Formerly Lockheed Martin, now with: Praxair, Denver, CO 80239
ABSTRACT Lockheed Martin has developed a unique two-stage, linear motor driven pulse tube cryocooler for cooling at 35 K for space applications. This system, developed under Lockheed Martin Independent Research and Development (IRAD) funding, demonstrates excellent power efficiency, providing 0.5 W of heat lift with 57 W of compressor power. A no-load temperature of 19.8 Kwas achieved. Performance data were obtained at both the design point, 35 K with no cooling at the warm stage, and also with cooling at both stages over a range of temperatures. At the design point (0.5 W at 35 K), the cooler surpassed the original programmatic requirements. The coldhead configuration is readily adaptable to meet specific temperature and load requirements, which is advantageous in that a cryocooler can be optimized for specific applications without the extensive development effort required for compressors. Based on the performance of this cooler, the rapid development time, and high reliability, resized versions of this configuration have been selected for two contracts currently underway for space applications. Other dual stage systems developed at LM-ATC and their relative merits are discussed.
INTRODUCTION Many space cryogenic instruments require cooling at two different temperatures for multiple detectors with different operating temperatures, for simultaneous cooling of detector and optics, or for reduction of parasitic heat loads from support structures and electrical leads. Multiple-stage pulse tube cryocoolers are an excellent way of providing cooling at several temperatures, because it is an inherently simple system: only a single set of the most complex components, the compressor and electronics, is required. The mechanically simple and reliable pulse tube cold head has increased configurational complexity in multiple-staging compared to single staging, but there is a negligible reliability penalty, such that system reliability essentially identical to a single stage pulse tube system. Furthermore, such a system is rapidly adaptable to new temperature and cooling load requirements, as only the coldhead needs to be redesigned, allowing one to develop and qualify a single flight compressor and use it with different coldheads in
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order to meet a wide variety of needs, provided the compressor has sufficient power delivery capacity. Such adaptability is also very important because precise system requirements are often not known until a project is well underway. The pulse tube cold head has many advantages relative to a Stirling coldhead. Flight Stirling displacers are motor driven, and require a tight tolerance, non-contacting displacer clearance seal to be maintained over several centimeters and over a large temperature gradient. Pulse tube coldheads, on the other hand, have no moving parts and no tight tolerance cold moving components. The absence of the displacer greatly simplifies the drive and control electronics, leading to higher reliability and weight reduction in the electronic controller. The pulse tube coldhead is capable of withstanding side loads an order of magnitude larger than typical Stirling coldheads, since the side loads on the Stirling can degrade the displacer clearance seal. This is advantageous in integrating to instruments, since the thermal busbars often require reasonable amounts of mass for high thermal conduction. Lockheed Martin’s Advanced Technology Center (LM-ATC) recognized the desirability of a two-stage pulse tube cryocooler, and in 1998 began funding IRAD supported research. Early work included an investigation into various 2-stage configurations, such as the simple configuration where two separate pulse tube coldheads are driven by a single compressor. A proof-of-concept unit was successfully tested at LM-ATC in 1999, simultaneously providing 0.2 W at 80 K and 0.8 W at 140 K {presented as an oral presentation at CEC-ICMC ’99.1 Also in 1997, LM-ATC had completed a contract with Air Force Phillips Laboratory wherein a two-stage pulse tube was developed to provide cooling simultaneously at 35 K and 60 K.2 The “intermediate bypass” configuration was used for that program, and although the cooling goals were nearly met, the thermodynamic performance was poor. Based on analysis, it was concluded that the intermediate bypass configuration was not the most thermodynamically optimum two-stage configuration. Numerous other configurations were studied on paper before selecting the configuration used in this work. Development for the cooler reported in this paper began in late 1998, as an IRAD funded replacement for a LM-ATC Stirling cryocooler displacer. The Stirling cooler was being developed for a flight program for an external customer, and because of LM-ATC’s successful development of a number of pulse tube coldheads, we considered it prudent to offer the customer a significantly higher reliability cryocooler as an option. The program requirements were to provide 0.5 W cooling at 35 K with 80 W of total system power (compressor plus electronics). The pulse tube was required to be a drop-in replacement for the displacer, and therefore to use the flight-qualified L1710 Stirling compressor, even though the swept volume was less than ideal for a two-stage pulse tube. A single stage pulse tube was not an option as the efficiency would be poor at 35 K, as 35 K is near the lowest temperature achieved for single-stage linear-motor driven pulse tubes.2 The coldhead was also required to have the same footprint, envelope, and interfaces as the Stirling displacer, and to be flight qualifiable. The pulse tube was first tested in 1999, and initially did not perform as well as analytical predictions. Subsequent investigations indicated a manufacturing defect caused the loss of performance. Although the flight program that required the Stirling cryocooler was subsequently cancelled, LM-ATC continued the IRAD program, and the coldhead was repaired and retested in 2000, achieving excellent results, and easily meeting the original specifications.
TWO-STAGE PULSE TUBE CRYOCOOLER The two-stage pulse tube is shown schematically in Figure 1. The compressor used for these tests was our flight-qualified L2010 compressor, an upgrade of the L1710 compressor, with larger swept volume and a higher efficiency motor. The L2010 is a linear flexure-bearing compressor with a mass of 6.3 Kg. The compressor operated near resonance, and was driven by laboratory rack linear amplifiers. At full stroke and nominal gas fill pressure, the compressor electrical power was 100 W.
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The compressor was connected to the pulse tube coldhead with a 17 cm long transfer line. The coldhead mass was approximately 1 kg. The coldhead heat rejection was conductive through a flange with an indium gasket. This flange was bolted to a heat rejection plate maintained at 295K by circulating chilled water. Each stage of the coldhead was instrumented with a calibrated diode thermometer and a wire-wound resistive heater to provide heat loads at those stages. The entire coldhead was insulated with multilayer insulation and inserted into a vacuum can. The compressor and transfer line were outside the vacuum can in air, and convectively cooled with a fan With no heat load into either of the stages, a drive frequency of 42 Hz, and 100 W compressor electrical power, the 1st stage temperature was approximately 110 K, and the 2nd stage temperature was 19.8 K.
Two-Stage Measured Performance: Heat into the 2nd Stage Only Measurements were performed with heat into the 2nd stage only, and with heat into both the 2nd and 1st stages simultaneously. With heat into the 2nd stage only, the cooling power at 35K as a function of compressor power is shown in Figure 2. The heat rejection temperature was fixed at 295K for all the data points in Figure 2. The intermediate stage temperature varied from about 140K at low compressor power to 110K at high power. The pulse tube met the 0.5 W cooling requirement with 57 W of compressor electrical power. This corresponds to a system power of 66 W with LM-ATC’s second-generation electronic controller, The maximum cooling power achieved at 35K was 0.79 W, with 100 W compressor electrical power. The best system power efficiency is near the design point, with compressor electrical power in the range of 40–60 W, and 110 W/W specific power (input electrical power divided by cooling power). Stirling cryocoolers have been able to achieve 85 W/W specific power, albeit for heavier systems than the one presented here. However, the constraint that forced LM-ATC to design this pulse tube for use with the L1710 compressor led to an overall system efficiency penalty. A coldhead redesigned for operation with a compressor with more swept volume can achieve power efficiency as high as or higher than that of Stirling cryocoolers.
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Figure 3 shows the cooling power as a function of cold tip temperature, for two different levels of compressor input power. The cold stage no-load temperature was slightly less than 20 K with both 60 W and 80 W of compressor electrical power.
TWO-STAGE MEASURED PERFORMANCE: HEAT INTO BOTH STAGES To simulate flight applications where cooling at multiple temperatures is required, measurements were made with heat simultaneously applied to both stages. Figure 4 shows the effect of changing the intermediate stage temperature, while fixing the cold stage temperature at 35 K and the compressor electrical power at 60 W. For example, with the intermediate stage at 140 K, the cooling power is 0.42 W at 35 K and 0.8 W at 140 K. As expected, as the intermediate stage
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temperature increases, the cooling power of the intermediate stage rises and the cooling power of the 35 K stage falls. Measurements were also made to map the cooler performance as a function of the two stage temperatures, as shown in Figure 5. The compressor electrical power was fixed at 80 W, and the two temperatures were varied from their no-load values up to 55 K (2nd stage) and 160 K (1st stage). The nearly horizontal lines are for fixed intermediate temperature, and the nearly vertical lines are fixed cold stage temperature. For example, with the 1st stage at 160 K and the 2nd stage at 35 K, the pulse tube simultaneously delivered 0.36 W cooling at 35 K and 2.05 W cooling at 160 K.
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These plots indicate that the two stages are somewhat coupled, such that for applications where the load on one stage is variable, the cooler must be designed to accommodate the maximum cooling loads on the other stage. These plots are also specific to this cooler only. In general, the staging configuration is adaptable to a wide range in the cooling capacity ratios and temperature ratios of the two stages by redesigning the coldhead.
EXTENSIONS TO THIS CRYOCOOLER This staging configuration has proven to be readily and rapidly adaptable to various cooling capacity requirements. LM-ATC has already applied this technology to several other programs, including a high capacity cryocooler described elsewhere in these conference proceedings.3 More significantly, this configuration easily extends to three or more stages. LM-ATC has developed a three-stage cryocooler for operation below 10 K. This cryocooler is the first demonstration of sub-10 K regenerative cooling at linear motor compressor frequencies. Test results from that cooler are also published elsewhere in these conference proceedings.4
CONCLUSIONS LMATC has developed a two-stage pulse tube cryocooler suitable for space applications. The cryocooler used a flight-qualified compressor and demonstrates excellent performance at the design point of 0.5 W of cooling power at 35 K. Simultaneous cooling from both stages was also demonstrated, simulating a spacecraft environment where the cooler is used to simultaneously cool two different temperature zones, e.g. a detector and associated optics. The successful demonstration of good two-stage performance is extremely attractive for use in space, where there are great advantages of the simplicity, high reliability, and robustness of a pulse tube cooler relative to other sorts of cryocoolers. LM-ATC has already applied this staging technology to several new programs.
ACKNOWLEDGMENTS This work was supported by Lockheed Martin Advanced Technology Center IRAD funds.
REFERENCES 1.
Paper CPB-4. Export control prevented publication of the paper.
2.
J.R. Olson, V. Kotsubo, P.J. Champagne, and T.C. Nast, “Performance of a Two-Stage Pulse Tube Cryocooler for Space Applications,” Cryocoolers 10, Plenum Publishing Corp., New York (1999), pp. 163-170.
3.
W.G. Foster et al., “The Development of a High Capacity Two-Stage Pulse Tube Cryocooler,” Cryocoolers 12, Kluwer Academic/Plenum Publishers, New York (2003).
4.
J. Olson, T.C. Nast, B. Evtimov, and E. Roth, “Development of a 10 K Pulse Tube Cryocooler for Space Applications,” Cryocoolers 12, Kluwer Academic/Plenum Publishers, New York (2003).
High Capacity Two-Stage Pulse Tube Cooler C.K. Chan, T. Nguyen, and C. Jaco TRW, Redondo Beach, CA B.J. Tomlinson and T. Davis Space Vehicles Directorate Air Force Research Laboratory Kirtland AFB, NM, USA
ABSTRACT This paper describes the design, construction, and testing of a two-stage pulse tube cooler that can simultaneously provide 2W of cooling at 35K and 20W of cooling at 85K. This cooler, which was designed for space application, is light weight, power efficient and has long life. The two stage cold head is driven by a TRW second generation flexure bearing compressor. The 35K cold block and the 85K cold block are designed in a parallel configuration to minimize interference among the stages. Cooler test data are presented for a range of cooling temperatures and cooling loads.
INTRODUCTION Long life pulse tube coolers have provided significant improvement to space surveillance,1 missile tracking,2 and infrared sensor cooling.3,4 The development of the second generation flexure bearing compressors5 and pulse tube cold head technology,6 has led to lighter weight and more efficient space coolers. As part of an ongoing effort to reduce the total system weight and cost of space missions, particularly when multiple cooling temperatures and loads are required, we have developed the space-qualified High Capacity Pulse Tube Cooler (HCC). This cooler is optimized for two-stage cooling involving 20W at 85K plus 2W at 35K. TRW's second generation pulse tube cooler technology started with the Integrated Multispectral Atmospheric Sounder (IMAS) instrument cooler,7-9 followed by the High Efficiency Cooler (HEC) Program.10,11 The HECs have been applied to several space programs including JAMI and SOFIS.12 The present HCC two-stage cooler was developed to provide a long life, low mass, higher efficiency space cryocooler for surveillance mission such as SBIRS-Low. Because its high efficiency and very low mass/unit capacity, HCC can be applied to simultaneously cool second stage 35K LWIR focal planes, as well as first stage 85K to 150K optics. As shown in Fig. 1, the HCC operation is not limited to the design point of 2W at 35K plus 20W at 85K, but has a wide range of useful first stage cooling temperatures and loads ranging from 80K to 120K for a given input power and heat reject temperature.
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The HCC cooler uses a back-to-back compressor design (Fig. 2) for vibration balancing. Flexure bearing supports provide close tolerance gas gap clearance seals for non-contact piston movement. The flexures themselves are designed for maximum stress levels well below the material's endurance limit. Their reliability has been validated by TRW accelerated tests. The working fluid is dry helium at 500 psig. The drive is a direct voice coil motor similar to a loudspeaker drive, thereby eliminating linkages. The design requirements and the tested capabilities of the compressor are shown in Table 1. The cooler components are shown in Fig. 3. The two parallel cold heads are attached to the centerplate and hermetically sealed with metal o-rings. The cold heads, containing the regenerators, cold blocks, pulse tubes and orifice blocks, are mechanically supported against launch loads by a thermally conductive aluminum supported structure (H-bar). The two 6.45 cm2 copper cold block interfaces are gold plated and located near the midpoint of the cold heads. The 35K cold head is provided with redundant calibrated platinum resistance thermometers used for temperature control. A TRW thermal strap, connected between the 85K cold block and the 35K regenerator sec-
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tion, enhances efficiency and reduces overall system power in a redundant cooler configuration. The centerplate is instrumented with a thermal sensor for use in protecting the cooler against overtemperature operation. Also located on the centerplate is an accelerometer that is used to sense cooler self-induced vibration. The accelerometer preamplifier, which is mounted to the compressor, amplifies the accelerometer signal for transmission to the control electronics. Here it is used as an error signal in a feedback loop to reduce cooler-generated vibration to very low levels. The compressor end caps, which also enclose the reservoir tanks, are hermetically sealed by metal o-rings. The centerplate incorporates simple and effective mechanical and thermal interfaces for attachment to the host payload. The primary mechanical mounting interface of the compressor (as shown in Fig. 4) can adequately remove up to 400W of heat without overheating the aftercooler. Heat removal can either be by direct conductive heat transfer to the mounting surface, or by heat pipes. The secondary interface (Fig. 4b) is an additional heat transfer surface to increase the rejection capability up to 700W. The thermal and the structural design margins of the cooler are listed in Table 2.
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Figure 5 shows the test setup for the two-stage cold head mounted on the High Capacity Compressor. The compressor was operated at 45 Hz with 500 psig helium fill pressure. Figure 6 shows test data taken for 2.25W at 35K plus 17.4W at 85K. The two-stage data is presented as a map of cooling power on the 1st stage (85K) versus cooling power at the 2nd stage (35K), for a fixed input power and fixed heat rejection temperature. The fixed power is the power necessary to achieve the temperatures and the loads. A similar map for the same loads and temperatures at another reject temperature, hence another fixed power, could also be generated.
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The advantage of the parallel cold head configuration in minimizing stage interaction can be seen in Fig. 6. Note that the change of the heat load at the second stage has a very small temperature effect on the first stage, and vice versa. A series of tests involving different cooling loads at 35K and at 85K was performed, and the results for different compressor powers are summarized in Table 3. The cooler is most effective for 1.42W at 35K and 12.0W at 85K using 323W of compressor power with a heat reject temperature of 300K and a compressor drive frequency of 45 Hz. The test data compare well with our design analyses (Fig. 1).
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CONCLUSION Our flight-qualified, two-stage pulse tube cooler has a large cooling capacity, up to 3 W at 35K on the second stage and up to 54W at 120K on the first stage. It weighs only 14 kg, but can accommodate an input power of 600W. Its performance compared to its design requirements is presented in Table 4.
REFERENCES 1.
T.M. Davis, B.J. Tomlinson, J.D. Ledbetter, “Military Space Cryogenic Cooling Requirements for the 21st Century,” Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2001), pp. 1-10.
2.
D.S. Glaister, M. Donabedian, D.G.T. Curran and T. Davis “An Overview of the Performance and Maturity of Long Life Cryocoolers for Space Applications”, Cryocoolers 10, Plenum Publishing Corp., New York (1999), pp. 1-20.
3.
C.K. Chan, et al., “Pulse Tube Coolers for NASA AIRS Flight Instrument,” ICEC 17 Proceedings (1998), p. 77-87.
4.
C.K. Chan, P. Clancy and J. Godden, “Pulse Tube Cooler for Flight Hyperspectral Imaging,” Cryogenics, 39 (1999), pp. 1007-1014.
5.
P.B Bailey, M.W.Dadd, N. Hill, C.F. Clark, J. Raab and E. Tward, “High Performance Flight Cryocooler Compressor,” Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2001), pp. 169-174.
6.
C.K. Chan, “AC Thermodynamics Theory of Stirling and Pulse Tube Cryocooler,” Symposium on Thermal Science and Engineering in Honor of Chancellor Chang-Lin Tien (1995), pp. 531-540.
7.
C.K. Chan, et al., “IMAS Pulse Tube Cooler Development and Testing,” Cryocoolers 10, Plenum Publishing Corp., New York (1999), pp. 139-148.
8.
C.K. Chan and T. Nguyen, “Performance of TRW Second Generation Pulse Tube Coolers,” 45 (1999), pp. 65-74.
9.
C.K. Chan and C. Jaco, “Low Vibrational Performance of TRW Second Generation Pulse Tube Coolers,” Adv. Cry. Eng., 45 (1999), pp. 577-584.
10. E. Tward, C.K. Chan, et al., “High Efficiency Pulse Tube Cooler,” Cryocoolers 11, Kluwer Academic/ Plenum Publishers, New York (2001), pp. 163-168. 11. E. Tward, C.K. Chan, et al., “High Efficiency Cryocooler, Adv. in Cryogenic Engineering, Vol. 47B (2002), Amer. Institute of Physics, Melville, NY, pp. 1077-1084. 12. J. Raab, R. Cobert, J. Godden, D. Harvey, R. Orsini and G. Toma, “JAMI Flight Pulse Tube Cooler System,” Cryocoolers 12, Kluwer Academic/Plenum Publishers, New York (2003).
Development of a High Capacity Two-Stage Pulse Tube Cryocooler W.G. Foster1, J. Olson1, P. Champagne1, B. Evtimov1, E. Will1, A. Collaco1, T. Nast1, R. Clappier2, A. Mitchell3, D. Jungkman3, R. Radebaugh4 and D.G.T. Curran5 1
Lockheed Martin Advanced Technology Center Palo Alto, CA 94304-1191, USA
2
Clappier Consulting
Discovery Bay, CA 94514, USA 3
Northrop Grumman Corporation Electronic Sensors and Systems Sector Baltimore, MD 21203, USA
4
National Institute of Standards and Technology Boulder, CO 80303, USA
5
The Aerospace Corporation El Segundo, CA 90009, USA
ABSTRACT Lockheed Martin’s Advanced Technology Center (LM-ATC) has developed a two-stage engineering model pulse tube cryocooler to provide simultaneous cooling to an instrument at 35 K and 85 K. Both the compressor and the coldhead were scaled up from previous units developed at LMATC. This system is believed to be the highest cooling capacity system developed to date specifically for spacecraft operation. It provides the desired cooling and demonstrates the robust and smooth running characteristics common for this type of linear compressor pulse tube system. This paper presents the test data on cryocooler performance and induced vibration and provides overall system characteristics.
INTRODUCTION The High Capacity Two-Stage Pulse Tube Cryocooler is an engineering model pulse tube cryocooler developed by Lockheed Martin’s Advanced Technology Center (LM-ATC) and Northrop Grumman (NG) to provide simultaneous cooling at 35 K and 85 K for space applications. The system is based on LM-ATC’s efficient pulse tube cold head technology and moving magnet compressor. The cooler was designed simultaneously to provide cooling powers in excess of 1.5 W at 35 K and 15 W at 85 K at 303 K heat rejection temperature, while maintaining low system mass and low induced vibration. Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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This paper reports the results of the program to date, and shows that high cooling powers were achieved with a cryocooler (compressor and coldhead) mass of 23 kg and a maximum induced vibration from the cooler of less than 1 in any axis. During this development phase of the program, LM-ATC designed and built two high power coldheads and one large linear compressor. The first coldhead was the prototype pulse tube, which was tested with an existing laboratory compressor to give early verification of the design approach. The second coldhead was the engineering model pulse tube, which was integrated with the new “Mega compressor” for system tests.
PULSE TUBE COLDHEAD DEVELOPMENT For this program, Lockheed Martin designed and built a two-stage coldhead with a “U-tube” configuration. The coldhead design utilized the results of earlier pulse tube work at LM-ATC that demonstrated the feasibility of the two-stage configuration1. The U-tube configuration is structurally strong and provides easy integration with the instrument package. The cryocooler is shown schematically in Figure 1. The two-stage pulse tube was designed to be very robust, able to sustain a static force of 225 N at the 85 K stage and 9 N at the 35 K stage while in a launch vibration environment.
TWO-STAGE PULSE TUBE PROTOTYPE TESTS Early in the program, the prototype pulse tube coldhead was built in order to demonstrate early verification of the design concept and the ability to achieve large cooling powers. The coldhead was mounted to the warm heat rejection flange, which was cooled with circulating chilled water to 290 K. The compressor was also water-cooled to 290 K. The compressor used for these tests was LM-ATC’s Low Cost Compressor (LCC), an existing laboratory compressor selected because of its relatively large swept volume of This compressor is substantially smaller than the new Mega compressor developed for the engineering model tests, which has a swept volume of but it was sufficient to demonstrate roughly 1/3 of the desired cooling power. The two cold head stages were each instrumented with a calibrated diode thermometer and a resistive heater to provide and measure the heat load at each stage. The entire coldhead was insu-
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lated with multi-layer insulation and inserted into a vacuum can, while the compressor and transfer line remained outside the vacuum in air. Laboratory rack power amplifiers powered the compressor. Figure 2 shows the cooling power for the 35K and the 85K stages as a function of compressor power, with a heat rejection temperature of 290K. At full stroke, there is 200 W of electrical power into the compressor, with the cooler providing 0.65 W cooling at 35 K simultaneously with 6.0W cooling at 85 K. Performance was measured in two orientations, with the cold head oriented downward in gravity (with the gas in the pulse tube in the convectively-stable orientation), and with the cold head oriented 45° relative to gravity. Such orientation characterization is important because a cooler must often be tested in a satellite system on the ground over a range of orientations. Measurements show a negligible degradation in performance at 35K and about 10% degradation at 85K. Figure 3 shows the 45° prototype orientation testing underway. These prototype tests demonstrate that the relative cooling ratio of the two stages is close to that of the goal cooling loads (1.5W at 35 K and 15 W at 85K). The tests were declared a success, and the program moved forward into the engineering model phase.
MEGA COMPRESSOR DEVELOPMENT In order to achieve the large cooling loads desired for this program, it was necessary to design and build a new, large compressor. LM-ATC scaled up an existing moving-magnet flexure-bearing clearance-seal linear compressor to large size, increasing the swept volume to and increasing the compressor mass to 20 kg.
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The moving magnet compressor has a number of attractive features. This design places the motor electric coils outside the working fluid, which has the advantage of removing most of the organics from the working fluid, thus lessening the chance of gas contamination. It also removes the need for electrical feed-throughs into the pressure vessel (a possible source of leakage), and removes the need for the flexible electrical leads required in a moving-coil design. Finally, it allows the motor coils to be conductively heat sunk, which allows much more current to be used to drive the compressor, allowing for higher input power without risk to the compressor. The Mega compressor worked smoothly from the very first time it was turned on, and has accrued over 200 hours of run time without problems, operating from –30 °C heat rejection temperature to +48 °C, with piston motion all the way up to full stroke. This compressor is capable of very high input power, in excess of 600 W, while cooled conductively in vacuum. Consequently, it is suitable for very high cooling load requirements, in excess of 20–50 W cooling at 60–100 K, for example, and is suitable for space applications. We believe that the Mega compressor is the largest flight-type compressor ever developed, both in terms of input power capability and swept volume.
TWO-STAGE PULSE TUBE ENGINEERING MODEL TESTS An engineering model pulse tube coldhead was built nearly identical to the prototype coldhead described above, and mated with the Mega compressor with a transfer line 20 cm in length. This cooler was designed to operate with both the coldhead and the compressor in vacuum. A large copper plate with circulating coolant provided heat sinking for both the compressor and the coldhead in the vacuum chamber. The coldhead warm flange and compressor were bolted to this copper plate and conductively cooled to simulate the cooler interface to a spacecraft. This configuration is pictured in Figure 4. Figure 5 shows the cooling performance with 303 K heat rejection temperature. The top portion shows cooling at 35 K, the middle portion shows cooling at 85 K, and the bottom portion shows the piston stroke (as the percent of maximum allowed stroke), each as a function of compressor electrical power. The maximum allowed stroke includes additional stroke margin to prevent the piston from impacting the soft end stops.
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At 95% of maximum allowed stroke, the cooler simultaneously provides 1.66W cooling at 35 K and 17.0 W cooling at 85 K, with 650 W compressor electrical power. At 55% of maximum stroke, the cooler provided 0.8W cooling at 35 K and 5.55 W cooling at 85 K, with 200 W compressor electrical power Figure 6 shows the load map for the engineering model pulse tube. The temperature of the two stages was varied, while keeping the compressor electrical power fixed at 400 W. The vertical axis is cooling power at the intermediate stage. The horizontal axis is the cooling power at the cold stage. The nearly horizontal lines correspond to constant intermediate stage temperature (80, 85 and 90 K), and the nearly vertical lines correspond to constant cold stage temperature (30, 35 and 40 K). The gray point in the center is the 35 K / 85 K, which is one of the points shown in Figure 5. The effect of heat rejection temperature was investigated. The temperature of the large copper heat sink plate, to which the compressor and coldhead were bolted, was varied between 254 K and 323 K, while keeping the compressor electrical power fixed at 500 W. The system was sealed throughout the testing, so that the operating gas pressure varied as the rejection temperature varied. These measurements are shown in Figure 7. The upper portion of Figure 7 shows the cooling power at 35 K, while the lower portion shows the cooling at 85 K. As expected, the cooling power increases dramatically at lower heat rejection temperature with fixed compressor power, with nearly 50% more cooling power available at 254 K rejection than at 323 K rejection.
EXPORTED VIBRATION TESTS Lockheed Martin has incorporated an analog vibration cancellation method into its flight-type electronic controller. This system, the Analog Feed-Forward Error Correction System (AFFECS) uses real time measurements of the vibration signal to adjust the drive signal to one of the two compressor modules in order to reduce the vibration. It does not rely on position sensors to cancel vibration, which reduces the reliance on such sensors and means that the failure of a position sensor does not make the cryocooler inoperative.
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The vibration was measured with accelerometers, with the cryocooler hanging on a spring stand. The cryocooler was at nominal operating pressure and drive frequency, but the cold stages were warm, because the system was tested in air. The cooler was fan cooled, but still heated up fast enough that measurements could not be performed at a stroke greater than 65% of maximum. Figure 8 shows the measured vibration for this system as a function of piston stroke. The upper portion of the figure shows the axial vibration force, in for the harmonic of the drive frequency with the highest measured vibration. Similarly, the lower portion of the figure shows the lateral vibration force. The highest residual vibration, at 65% of maximum stroke, was 0.5 in the lateral direction.
CONCLUSIONS Lockheed Martin and Northrop Grumman have developed an engineering model high-capacity two-stage pulse tube cryocooler suitable for space applications. The Mega compressor is believed to be the largest long-life flexure-bearing flight-type compressor ever developed. The pulse tube simultaneously provides over 1.6 W cooling at 35 K and over 16 W cooling at 85 K while rejecting heat to a 303 K heat rejection interface. The cryocooler mass (compressor plus coldhead) was 23 kg, and the coldhead is very robust, able to sustain a static force of 225 N at the 85 K stage and 9 N at the 35 K stage while in a launch vibration environment.
ACKNOWLEDGEMENTS This work was developed under Northrop Grumman funding.
REFERENCES 1. T. C. Nast, J. Olson, B. Evtimov and V. Kotsubo, “Development of a Two-Stage Pulse Tube Cryocooler for 35 K Cooling”, published elsewhere in these proceedings.
Two Stage Hybrid Cryocooler Development K.D. Price and C.S. Kirkconnell Raytheon Electronics Systems El Segundo, CA, USA 90245
ABSTRACT Raytheon has demonstrated a two-stage hybrid Stirling/pulse tube cryocooler for long life space infrared (IR) sensor applications. The first expander stage is a conventional Oxford-class Stirling expander. The second expander stage is a U-turn pulse tube mechanically and thermodynamically extended from the first stage Stirling cold end. The combination Stirling and pulse tube expander has higher efficiency, lower weight and size, and lower production costs relative to the current state-of-the-art approaches. The first hybrid experimental unit has now been built and tested. The engineering model delivers 0.5W at 35K plus 2.2W at 80K. Data obtained from these tests have identified a set of cold head modifications that will increase performance up to 1W at 35K plus 7W at 100K for a drive motor input power of less than 170W. One of the motivations behind this development is the versatility of the hybrid technology. In addition to supporting heat loads at two different temperatures, the hybrid has the unique ability to allocate refrigerating power between stages on command by changing the Stirling expander piston phase and/or amplitude. This capability broadens cooler utility in two ways. One, a single cryocooler design can be tuned to perform optimally over an unprecedented broad range of temperature and heat load combinations. This reduces or eliminates development cost for many applications. Two, expander heat lift distribution can be adjusted in real time to optimize performance if heat loads change on orbit or if actual loads are different from predicted. This reduces performance risk on sensors in development where predicted loads may change significantly during design. The cooler has separate compressor and expander modules, weighs less than 7Kg and will be powered and controlled by radiation hard Command and Control Electronics now entering production.
INTRODUCTION Raytheon is developing a long life two-stage hybrid cryocooler, the RSP2, to simultaneously cool the focal plane array, optics, and radiation/thermal shields in modern infrared (IR) space sensors. The various sensors being considered have significantly different temperature and load combinations and some have thermal load cycles that vary substantially during an orbit. We chose to develop the Stirling-pulse tube hybrid cryocooler for these applications because a single design is sufficiently versatile to meet the requirements for most sensors. In addition, the cooler is no more difficult to produce than a single stage Stirling cooler.
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We have built and tested the first breadboard version of the cooler. See Figure 1. This cooler has the demonstrated ability to shift refrigeration capacity from one stage to the other on command. The load shifting ability greatly broadens its performance range and makes the cooler particularly useful in thermally managing sensors that have significantly varying heat loads, as will be described. The cryocooler design is first mechanically described. Next, its ability to shift refrigeration on command is explained. This is followed by performance data at several temperature combinations. Methods for integrating the cooler and using its novel capabilities are discussed. Finally, the next stage of development is described.
THERMO MECHANICAL UNIT DESCRIPTION The TMU is comprised of separate compressor and expander modules connected by a transfer line. See Figure 2. The TMU retains significant legacy to Raytheon’s previously developed “Oxford” class machines 1,2,3 . Compressor swept volume is 7.5cc produced by a pair of pistons working in opposition against a common compression volume. The pistons are driven by linear motors and are suspended on three finger tangential flexures to maintain tight alignment of non-wearing clearance seals. The two-stage expander employs a Stirling first stage and a pulse tube second stage. Helium flows from the compressor into the Stirling first stage and a portion continues on to the pulse tube second stage. The pulse tube stage is configured in a “U-tube” for compactness and structural rigidity. Warm ends of the pulse tube, regenerator tube, pulse tube orifice and surge volume are thermally anchored to the first stage. The breadboard unit reported on in this paper consists of a Raytheon developed expander module driven by an Air Force owned compressor module. The compressor was designed and built by Raytheon on the Air Force 95K High Efficiency Cryocooler program 4,5 .
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LOAD SHIFTING An important feature of the hybrid cooler is its ability to allocate load capacity between stages on command. Shifting capacity between stages, while holding both stage temperatures constant, is possible because each expansion stage is controlled by a separate mechanism. The first stage Stirling expansion piston is motor-controlled. Increasing piston phase angle increases the net refrigeration produced in the stage. (Phase angle refers to the mechanical angle between compressor and expander pistons.) The orifice between pulse tube outlet and surge volume passively controls second stage pulse tube expansion. Refrigeration in this stage is determined primarily by pressure ratio, which is inversely related to first stage expander phase angle. Therefore, when phase angle increases (decreases), first stage net refrigeration increases (decreases) and second stage net refrigeration decreases (increases.) Cryocooler efficiency remains relatively constant over a wide range of piston strokes and phase angles. As in other coolers of this class, increasing compressor piston stroke increases refrigeration in both stages, regardless of capacity allocation. Load shifting broadens the performance range and capability of the RSP2 relative to other coolers for at least two reasons. First, load shifting obtains a much broader range of temperature and load combinations compared to coolers with fixed expansion mechanisms. As a result, the cooler is useable in a wider range of applications. Second, load shifting enables novel thermal management techniques such as: 1. Simultaneously controlling temperature of two FPAs operating at different temperatures. 2. Improving thermal management of sensors with cyclically varying loads. The latter capability will be discussed in further detail below. Figure 3 shows measured performance at stage temperatures of 58K and 110K for a range of Stirling piston phase angles. In the range tested, second stage capacity varies by a factor of 1.42 while first stage capacity varies inversely by a factor of 8.33. The dotted lines indicate expected trend lines as phase angle continues beyond the range tested.
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If the first stage is thermally linked to a temperature-stable heat sink, such as a triple point energy storage system, the cold head can operate when the first stage produces negative capacity, i.e., produces insufficient cooling to maintain temperature. Figure 3 shows first stage performance projected into negative cooling below about 50 degrees phase angle. In this circumstance, the operating time at negative capacity is limited by the capacity of the heat sink, but during that time, the second stage capacity can be further increased as indicated. This has been demonstrated informally with large structural mass substituting for a triple point system, but precise performance numbers require test with a temperature stable heat sink. In addition, the load shifting capability enables the cooler to support an infinite selection of thermal load combinations over a broad range of stage temperatures. The demonstrated temperature range currently extends upward from about 28K on the second stage. Two examples of load point pairs include 35K / 80K carrying 0.45W / 2.2W, respectively, and 112K / 154K carrying 4.5W/4.5W.
INTEGRATION OF THE RSP2 CRYOCOOLER WITH A CRYOGENIC THERMAL STORAGE UNITS The load shifting capability of the RSP2 can simplify thermal management of IR sensors with widely fluctuating FPA thermal loads. This may occur, for example, when the sensor is turned on and off in some prescribed duty cycle. In such applications, FPA heat load will alternate between two levels while optical bench heat loads remain essentially constant. Typically, it is desired to hold FPA temperature constant to within 0.1K at some point between 35 K and 60 K while the optical bench temperature remains constant within several kelvins at a temperature between 80 and 120 K. Traditional approaches for handling duty-cycled loads have been to either size the cryocooler for peak load or to use a cryogenic thermal storage unit (CTSU) at the second stage.
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The former usually suffers size, weight, and power drawbacks. The latter is only volume and mass efficient when the CTSU is a triple point device. Unfortunately, FPA temperature often does not correspond with a triple point. Therefore, a less efficient CTSU, such as a large thermal mass or a solid-solid phase change material, must be employed to stabilize temperature. Load shifting in the RSP2 enables cooler and thermal system design to be based on time average loads instead of peak loads. In addition, the CTSU can be located at the warmer first stage where there is a wide selection of suitable triple point materials. The alternating FPA load can then be managed by load shifting thermal capacity to and from the second stage while the CTSU maintains temperature stability at the first stage. Further, the relatively bulky CTSU is transferred to the optical bench, where there is typically more space for it. Table 1 is a partial list of triple point materials useful at the first stage. There are numerous choices, particularly above 80K. Figure 4 shows schematically how the RSP2 can be integrated to an FPA at the second stage and a CTSU thermally strapped to the first stage. We expect that the CTSU will be thermally mounted to the optical bench and connected to the first stage through a compliant thermal strap. Note that nothing in this arrangement prevents use of a CTSU or added thermal mass at the second stage, if desired. The integrated system works as follows: during periods of peak FPA load, the ”on state,” the Stirling displacer phase angle is reduced to boost refrigeration capacity in the second stage at the expense of capacity in the first stage. The CTSU maintains optical bench temperature during this period with little or no help from the first stage. When the FPA load reduces to its “off-state” value, the Stirling phase angle is increased to increase first stage refrigeration to recharge the CTSU. A temperature control servo operating around the second stage maintains precise FPA temperature throughout the cycle.
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A dynamic simulation model was built to demonstrate the functionality of a combined RSP2-CTSU cryogenic system. The RSP2 experimental data from Figure 3 was adapted into the following system: Second stage controlled to 58 K First stage operated between 115 K and 120 K Peak second stage load = 1.6 W; Minimum second stage load = 0.8 W FPA duty cycle = 20% on (peak load), 80% off (minimum load) Constant 4.2 W heat load at the first stage of the cryocooler CTSU is a 115.8 K krypton triple-point unit with 7800 J capacity Thermal path conductance from first stage cooler interface to CTSU = 1.3 W/K Cycle period = 96 minutes The results of the simulation are provided in Figure 5. Based on the measured performance of the breadboard cooler, including some of the extended range, the defined second stage load variation of 0.8 W to 1.6 W requires a phase angle adjustment between 102 and 56 degrees. The corresponding capacity fluctuation at the first stage is 5.2 W to 0.5 W, respectively. During “FPA on” periods, the phase angle is at 56 degrees, second stage capacity is high and first stage capacity drops below the 4.2 W constant load value. Heat flows from the first stage of the cryocooler to the CTSU, hence the first stage temperature rises a few degrees above the CTSU stabilization temperature. During “FPA off” periods, the phase angle is at 102 degrees, second stage capacity is low and first stage capacity is greater than 4.2W. The excess capacity is used to extract heat from the CTSU to recharge it for the next cycle, so first stage temperature drops below the triple point temperature. The present model does not account for additional thermal inertia in the system that would naturally be present due to the devices being cooled, i.e., the FPA and the optical bench. Hence the relatively rapid step change behavior exhibited in Figure 5 would, in practice, be more gradual. The approach has additional advantages arising from the ability to locate the CTSU at the upper stage. First, the additional parasitic loads on the system due to conducted and radiated heat loads on the CTSU are carried at the first stage instead of the second. This reduces the effective heat burden on the cryocooler. Second, sensor designs are often physically constrained in the vicinity of the FPA and lowest temperature cryocooler stage. There are generally greater volume and location options for the CTSU at the first stage. The analysis was based upon the demonstrated performance of the breadboard unit with some small, well-defined improvements leading to second stage load profiles of 2:1. We believe additional enhancements of load shifting capability can increase this to 3:1.
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DEVELOPMENT PLANS Several ways to further improve performance have been found from either test results or reconsideration of the mechanical design. Some can be implemented in the breadboard and will be done this year. Others require a new cold head, which is now in the planning stages. Breadboard upgrades include a modified pulse tube regenerator and installation of a pulse tube bypass orifice. A new cold head can further reduce dead volume, refine flow geometry, improve heat exchangers, and reduce conducted heat loads. We will also expand the range of load shifting testing to map out performance at different temperature pairs.
ACKNOWLEDGEMENTS The Air Force Research Laboratory, Albuquerque, NM, loaned the compressor used in this effort and Swales provided the triple point data in Table 2.
REFERENCES 1.
2.
3.
4. 5.
Wakagawa J.M., Haque H., and Price K.D. "Improved Standard Spacecraft Cryocooler Life Test for Space-Based Infrared Surveillance", Cryocoolers 8, Kluwer Academic/Plenum Publishers, New York (1995), pp. 69-76. Price, K.D., Reilly J., Abhyankar N., and Tomlinson B.J. "Protoflight Spacecraft Cryocooler Performance Results," Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2001), pp. 35-44. Abhyankar N., Yoneshige C.H., Tomlinson B.J. and Reilly J., “Characterization of Raytheon’s 60K 2W Protoflight Spacecraft Cryocooler”, Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2001), pp. 45-54. Price, K.D., Urbancek, V. "95K High Efficiency Cryocooler Program", Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2001), pp. 183-188. Kirkconnell, C.S., Price, K.D., Barr M.C., Russo J.T. "A Novel Multi-Stage Expander Concept", Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2001), pp. 259-264.
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Development of a 10 K Pulse Tube Cryocooler for Space Applications J. Olson, T. C. Nast, B. Evtimov, and E. Roth Lockheed Martin Advanced Technology Center Palo Alto, CA, 94304-1191, USA
ABSTRACT Lockheed Martin has built and tested a pulse tube cryocooler that achieved a no-load cold tip temperature of 5.35 K while rejecting heat at room temperature. The system is a simple pulse tube cryocooler, with a room temperature flexure-bearing linear compressor as its only moving part. The compressor operates on resonance at high frequency, with the low mass and high efficiency required for space applications. This system was developed with Lockheed Martin Independent Research and Development (IRAD) funds, and proves the concept that a reciprocating cryocooler can work well at low temperature and high frequency. Such a pulse tube is extremely attractive for space applications because the no-moving-parts coldhead and the single compressor and electronic controller have the same high reliability as previous single-stage pulse tube coolers operating at higher cold-tip temperatures. INTRODUCTION In recent years, there has been increasing interest in the science and defense communities for a long-life mechanical cryocooler for temperatures below 10K, capable of being used in space. Applications include pre-coolers for adiabatic demagnetization refrigerators as well as detectors requiring operating temperatures below 10 K. Lockheed Martin’s Advanced Technology Center has developed a cryocooler suitable for lowtemperature space applications. We have built and tested a three-stage pulse tube cryocooler with a flight-type resonant flexure-bearing linear-motor compressor, and achieved a no-load temperature of 5.35 K, and in excess of 125 mW cooling at 10 K, with 240 W compressor electrical power and 290 K heat rejection temperature. With 120 W of compressor electrical power, the no-load temperature was 6.0 K, and the cooling power at 10 K was 75 mW. A multiple-stage pulse tube is an attractive system for space cryogenics, as it has the same high reliability as a single-stage pulse tube because it uses the same high-maturity electronics and compressor. The staging configuration provides natural locations at which to heat sink electrical leads and cooled shields, reducing the cooling load on the lowest temperature stage and possibly removing the need for a cold (80–120 K) radiator often included in a low-temperature system. Furthermore, it is relatively easy to modify a pulse tube coldhead in order to change the various stage temperatures for different system requirements, which leads to short development time between system specification and hardware testing. Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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THREE-STAGE PULSE TUBE CRYOCOOLER The three-stage pulse tube cryocooler is shown schematically in Figure 1. The compressor used for these tests was our Low Cost Compressor, developed with cost sharing between Lockheed Martin and NASA/HQ, and selected because of its relatively large swept volume. It is a linear flexure-bearing compressor with a mass of 12 kg. The compressor operates near resonance, and was driven by laboratory rack linear amplifiers. At full stroke and nominal gas fill pressure, the compressor electrical power was 240 W. The compressor was connected to the pulse tube coldhead with a 10 cm transfer line, though a much longer transfer line could have been used. The coldhead was mounted to the warm heat rejection flange, which was cooled with circulating chilled water to 290 K. Cooling water was also used to cool the compressor to 290 K. The coldhead consists of three stages, each instrumented with a diode thermometer calibrated between 4 K and 325 K. The 2nd and 3rd stages had wire-wound resistive heaters to provide heat loads at those stages. The entire coldhead was insulated with multi-layer insulation and inserted into a vacuum can, while the compressor and transfer line remained outside the vacuum in air. The coldhead mass was approximately 2 kg. With no heat load into any of the stages and 240 W compressor electrical power, the 1st stage temperature was approximately 100 K, and the 2nd stage temperature was approximately 25 K, and the 3rd stage reached 5.35 K.
Three-Stage Measured Performance: Heat Into 3rd Stage Measurements were performed with heat applied to the 3rd stage only, and with heat into both the 3rd and 2nd stages simultaneously. With heat into the 3rd stage only, the cooling power as a function of cold tip temperature is shown in Figure 2, for four different levels of compressor electrical power. The compressor frequency was fixed at 31 Hz, and the heat rejection temperature was fixed at 290 K for all the data points in Figure 2.
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At full stroke, the compressor electrical power was 240 W, and the cooler provided 16 mW at 6 K, 62 mW at 8 K, 127 mW cooling at 10 K and 180 mW at 12 K. The no-load temperature was 5.35 K. The specific power (input power divided by cooling power) is relatively insensitive to compressor power, as shown in Figure 3. This indicates that good performance can be achieved over a wide variation in operating conditions. The lowest measured specific power at 10 K was 1570 W/W, which is equal to 1.8% of Carnot efficiency.
Three-Stage Measured Performance: Heat Into Both 2nd and 3rd Stages Tests were also performed with heat simultaneously applied to both the 3rd stage and the 2nd stage. Such a situation would be common for spacecraft uses, where the upper stages of the cooler would be used to intercept parasitic heat load through electrical leads going to cold detectors, or to cool a radiation shield. It would have been desirable also to use the first stage to intercept heat, but cost and time restrictions prevented us from performing those tests.
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The temperature of the 2nd stage was fixed at 30 K and the 3rd stage at 10 K. The cooling powers were measured as a function of compressor electrical power and are shown in Figure 4a. At 240 W of compressor electrical power input, the cooler simultaneously provided 120 mW cooling at 10 K and 295 mW at 30 K. It is interesting that the cooling power at 10 K dropped only a small amount (127 mW to 120 mW) when the 2nd stage heat went from 0 to 295mW (240W input in both cases). This indicates that the 10K cooling power is relatively insensitive to the 2nd stage heat load, and the cooling power at 30 K can be gotten “for free,” with little additional input power required. Figure 4b shows the cooling powers with the 2nd stage at 30 K and the 3rd stage at 8 K. At 240 W of compressor electrical power input, the cooler simultaneously provided 52 mW cooling at 8 K and 285 mW at 30 K. Of interest in Figure 4c is the frequency range between the peaks in cooling loads, where increasing the frequency from 30 to 32 Hz increases the cooling power at 8K by 80%, while the 30K cooling decreases by 20%. This dependence allows the frequency to be changed “on-orbit” in order to change the ratio of cooling powers between the two stages.
Long-Term Temperature Stability After characterizing the cooler performance as described above, we began a test of the longterm temperature stability of the cryocooler. We started the cooler and left it running continuously for 8 weeks without intervention. The cooler was sealed, the input drive current was set, and the stage heaters were turned off. The temperature traces for this interval are shown in Figure 5. There
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was no temperature control during this time, and the compressor electrical power remained constant at 180 W. The temperatures of all 3 stages remained very steady with no indication of performance degradation, despite the fact that the cooler did not undergo an extended bake-out procedure due to lack of time. The 3rd stage temperature dropped slightly early on, falling from 5.48 K to 5.46 K
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before rising slowly back to 5.48 K. The 2nd stage cooled nearly 0.5 K, from 23.9 K to 23.5 K. The 1st stage temperature cooled slightly from 101.3 K to 100.5 K. We expect that pulse tubes are relatively less sensitive to contamination than are Stirling coolers (with tight seals required at the cold end) and Joule-Thomson coolers, which are prone to plugging at the expansion nozzle. Furthermore, Lockheed Martin’s moving magnet compressor design is inherently cleaner than moving coil compressors, because the electric motor coil is outside the working fluid, which removes most of the organics from the working fluid.
ACKNOWLEDGMENT This work was supported by Lockheed Martin Advanced Technology Center Independent Research and Development (IRAD) funds.
Scaling of Cryocooler Compressors P.B. Bailey and M.W. Dadd Oxford University, Oxford, UK
C.F. Cheuk and N.G. Hill The Hymatic Engineering Co. Ltd, Redditch, UK
Jeff Raab TRW, Redondo Beach, CA, USA
ABSTRACT The successful HEC cryocooler compressor has been used as the basis for two very similar compressors, one smaller and one larger than the original. The new compressors were designed largely by a direct scaling of the original. This paper describes from first principles how some of the design parameters are affected by scaling, and how this will impose limitations on machines at extreme ends of the size range. The three sizes of compressor are compared, and some of the practical problems encountered during the scaling process are mentioned.
INTRODUCTION A new type of compact linear motor and a new flexure spring design were developed by Oxford University for linear compressors with the aim of meeting stringent requirements for high efficiency and low mass. These machines are of the typical ‘Oxford’ design, being powered by permanent magnet, moving coil motors, and having a clearance seal between piston and cylinder. The original balanced pair compressors of this type were developed and made by Oxford University for TRW with TRW internal funding, and two cryocoolers incorporating similar machines were delivered by TRW to NASA/JPL for the New Millennium IMAS project1. Subsequent to this Oxford designed for TRW both a smaller compressor ("800 gram") and a larger compressor ("26cc"), both scaled in size from the original machine, but neither of these were built at the time. The IMAS machine was then developed further in a joint collaboration between Oxford, Hymatic and TRW with the aim of making the compressor more rugged, and also introducing a fully controlled assembly and test process more suitable for repeated and consistent quantity
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production2. These machines, the 6cc High Efficiency Cryocooler (HEC) compressors, are now in production and more than 12 have been made to date, with production ongoing. Following the success of these machines, both the smaller and the larger machines have been redesigned along the lines of the HEC compressor, and one small and three large compressors have now been made and are operating successfully. This paper looks at how various design parameters scale as the size of the machine changes, and compares the resulting machines.
DESIGN PARAMETERS The analysis below is based on the assumption that all linear dimensions of the design, including the stroke, are scaled by a single scaling factor (K).
Mechanical Springs The flexures are a key component in any design, and these machines are based on the classic ‘Oxford’ spiral disc spring. Though the exact behaviour of such a spring is complex, a good approximation can be made by assuming that each spring ‘arm’ acts as a cantilever 'built-in' at both ends. The arguments below are for the deflection of a simple rectangular beam of depth (d) and breadth (b). In practice, springs act in both bending and torsion, but the scaling arguments apply equally to both. Note that the spring thickness (beam depth) is increased pro rata with the spring diameter. The mechanical stiffness of a beam of length (1), elastic modulus (E) and second moment of inertia (I) is given by
As all of the linear dimensions scale according to the linear scaling factor (K), the stiffness is proportional to this factor. The stress (σ) in the beam is related to the Elastic Modulus of the material, the beam depth and the radius of curvature (R) by the formula
and hence the stress is independent of size.
Operating Frequency For maximum efficiency, the compressor is driven at the resonant frequency derived from the moving mass (m), and the mechanical and gas spring stiffnesses
which is by
The gas spring stiffness is typically several times larger than the mechanical spring stiffness, and is found from the piston area the pressure swing ( which is assumed independent of machine size) and the stroke by
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Thus both the gas and mechanical spring stiffness are proportional to the scale factor. The moving mass, however, depends on volume and hence is proportional to the cube of the scale factor. Thus the operating frequency will vary as follows
Ideal Motor Design With a permanent magnet, moving coil motor, providing the magnetic circuits are geometrically similar, the magnetic flux density in the air gap (B) will be the same. The force (F) produced when a current (I) flows through a length (L) of wire in the drive coil is given by
If the mean current density in the coil, constant, then the force is given by
(where
is the coil sectional area) is kept
And if all linear dimensions scale proportionately with a scale factor K then
From equation (5) above, it can be seen that the frequency varies with the inverse of the scale factor, whereas the stroke is proportional to it. These two cancel out, with the result that velocity is independent of scale factor. Hence the motor power varies as the force
One limitation on the motor design is the dissipation of heat within and from the drive coil. Conductive heat transfer within the coil is usually excellent, but convection from the outside of the coil is potentially problematic in larger machines. Given the small radial gap and that the fluid is invariably helium, the heat transfer from drive coils of small compressors is excellent, and it is not until machines are of considerable size that heat transfer from the coil becomes a serious issue.
A Real Motor Though the flux density in the air gap is independent of size, there are constructional problems in designing small motors. If the minimum screw size to be used is M1.6, this puts a lower limit on the size of many parts just in order to accommodate the tapped holes required. Thus the overall size of the motor does not reduce pro rata. In practice, the 'packing factor' (the proportion of material in the air gap that is copper wire) will decrease as the size of motor is reduced. This is due to an increase in the relative thickness of insulation, coil holder and relative increase in running clearances. For both of these reasons, motors become relatively less efficient (or relatively less powerful) as the size is reduced.
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Clearance Seal The power lost by leakage through an ideal clearance seal is related to the piston diameter (D), the radial clearance (t), seal length (L), the pressure swing and the viscosity (µ) by
As the pressure swing tends to be fixed, regardless of size, the power loss will scale according to the cube of the radial clearance. In practice, this clearance increases only slightly as the size goes up, thus the seal loss becomes relatively smaller as the machine gets larger.
Work done by the compressor "PdV" work is done by the piston on the gas in the cylinder. The power delivered to the gas is derived from the frequency, the pressure swing and the swept volume by
As stated before, for many applications the pressure swing will be kept constant, thus the power scales with the square of the length
It is interesting to note that the power that can be delivered by the motor increases by whereas the power that can be absorbed by the gas varies by
Specific Mass For space application, the power per unit mass is important, and a simple approximation to it is
In other words this simple analysis appears to show that smaller machines have a higher power-to-weight ratio than larger ones, which seems to be the opposite of what you would expect. In practice, the interface and the pressure containment do not scale in a truly linear fashion, and these parts are usually a significant proportion of the total. However, the problem with large machines is that as they become larger, the power available from the motor increases faster than the size of piston and cylinder needed to deliver that power, so to avoid the compressor being 'over-motored', the piston diameter (or stroke) must increase by a larger scale factor than the motor. In the present design the piston and cylinder are located within the motor - at some point this configuration will have to change, with the piston having to be cantilevered out beyond the end of the motor (see Figure 1). This point has not been reached yet with the existing “26cc” design, but is probably not far off.
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PRACTICALITIES OF SCALING AS A DESIGN METHOD With the advent of CAD, scaling a design has become a much simpler exercise - an entire design can be quickly scaled up or down at the press of a button. However, life is never that simple. In practice the entire design has to be reworked to make allowances for tolerancing and tolerance stack-ups, which become more significant in small sizes. It is also convenient to have primary dimensions a simple number, and it is not wise to ask the machine shop to cut some M2.17 screw threads. Some of the newer CAD packages may actually make this re-working much simpler by using parametric factors in the design. A huge benefit to the design process is that many of the analyses required to evaluate design stresses can be adapted with the change of just a few dimensions, or even eliminated altogether by the use of dimensional analysis. There are a few areas where simple scaling is not relevant, and these tend to be more problematic as size goes down, rather than up. Pressure Containment. Proprietary seals are usually not available in 'miniature sizes'. Use of screws smaller than M3 on pressure vessels is undesirable. Wall thicknesses less than 1mm need careful manufacture and protection in use. Electrical Interface. Problems with availability of sub-sub-miniature connectors and welding and soldering to them. Screw threads. Using M1.6 screws are difficult; anything smaller is impractical, especially when it is necessary to tap holes into hard materials. In the 800 gram machine the overall diameter of the motor has to be increased from the 'linear scaled' size purely to accommodate fasteners.
MANUFACTURE AND ASSEMBLY With similar components, the ‘learning curve’ for the manufacture of a component is largely overcome the first time round – subsequent similar (but different sized) parts will be easier. In practice, manufacture of smaller parts was easier than larger ones, because some of the smaller parts were given a relative increase in thickness to accommodate fasteners, etc, and these parts thus became stiffer and easier to machine. Conversely, the temptation to reduce the mass of the larger parts, where this is possible, leads to difficulty in achieving the precision required. The assembly process was kept identical for all three machines, and in many cases the same fixturing could be used with adaptation. An additional bonus is that the assembly technicians
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need almost no new training to assemble the newer machines, as the same techniques, assembly and test procedures and standards are used for all. For the large machine, consideration was given to how it was handled during the assembly process, as the mass of both the separate compressor halves and the complete machine was considerably greater than any machine previously assembled.
COMPARISON OF THREE MACHINES A few of the relevant dimensions of the three existing compressors (together with a further proposed design) are tabulated in Table 1, together with a dimensionless scaling factors which have been set to 1 for the 95K compressor. As can be seen, to a first approximation, the parameters listed vary as expected, the most significant deviation being the total mass of the machine. This is largely due to the way in which the interfaces and pressure containment scales, and for this reason the mass has been subdivided. The balanced compressors consist of two identical ‘compressor halves’ which are mounted on to a ‘Centre Plate’, and the mass of these halves is tabulated, together with the mass of the ‘interface’ which is the Centre Plate and pressure containment. An obvious, and deliberate, omission from this table are values for power and compressor efficiency. These are not included for two reasons; firstly, at the time of writing, data for the small and large machines is not available, but more importantly, these values depend on the thermodynamic load (cold head) into which the compressor is driven, and can thus vary over a wide range. Typically a small load will result in a higher efficiency, and a larger load will give a lower efficiency, but figures quoted in isolation from a cold head are meaningless out of context.
CONCLUSION By directly scaling the existing design of the HEC compressor, a smaller and a larger machine have been successfully designed, manufactured and tested. The range over which such a technique can be used is limited in both directions. In small sizes, it is difficult to find space for fasteners, and the size of the interfaces and pressure containment become more significant.
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The efficiency of small machines also drops, due to the smaller amount of copper that can put in the air gap of the motor, and the relatively larger clearance seal loss. In larger sizes, the power available from the motor increases at a faster rate than that at which power can be absorbed into the cylinder, and so a topological change is required so that motor and cylinder are more suitably matched.
ACKNOWLEDGMENTS We acknowledge the help of staff at both TRW and Hymatic for their assistance, and we also acknowledge the strong support of Thorn Davis and B.J. Tomlinson of AFRL.
REFERENCES 1 2
Chan, C.K., Nguyen, T., Colbert, R., Raab, J., Ross, R.G. Jr., Johnson, D.L., "MAS Pulse Tube Cooler Development and Testing", Cryocoolers 10, Plenum Press, New York (1999), pp 139-147. Bailey, P.B., Dadd, M.W., Hill, N., Cheuk, C.F., Raab, J., Tward, E., “High Performance Flight Cryocooler Compressor”, Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2001), pp 169-174.
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The Linearity of Clearance Seal Suspension Systems M.W. Dadd, P.B. Bailey, and G. Davey Oxford University, Oxford, UK T. Davis and B.J. Tomlinson Air Force Research Laboratory, Albuquerque, NM, USA
ABSTRACT The classic ‘Oxford’ cryocooler has a clearance seal between the piston and the cylinder which is maintained by the use of spiral disc springs. In a typical compressor this clearance is about 12 microns, and therefore the spring suspension system must have a linearity of no more than 3 or 4 microns to avoid contact. It has always been assumed that to maintain this linearity, the surfaces between which the springs are clamped must be very flat and very parallel to each other. It has also been assumed that the flatness and parallel-ness of the clamping at the inside of the spring is more important than at the outside. Under a research and development effort with the Air Force Research Laboratory, some work was carried out to investigate how the linearity of motion is dependent on the clamping conditions of the springs. Tests were carried out on a typical suspension system which was deliberately assembled between non-parallel clamping surfaces, and the linearity of the resulting motion was measured. A simple theoretical model was developed which gives good agreement with the experimental results. The results have several useful implications for the manufacture of cryocoolers. It was found that the normal method used to verify the linearity of motion by measuring the run-out at a single axial position could be highly misleading — at least two, and preferably more measurement points are needed. It was also concluded that the clamping surfaces around the outside of the spring are as important as those on the inside of the spring in defining the linearity of motion. INTRODUCTION The defining characteristic of the ‘Oxford’ Cryocooler is the use of flexures to maintain a clearance seal between the piston and the cylinder. These flexures have high radial stiffness and low axial stiffness, thereby constraining the movement of the piston along a predefined axis. Typically, the radial clearance seal is in the order of 12 µm — if they are made much larger than this the seal losses start becoming unacceptably high. It is difficult to make the piston and cylinder with a cylindricity of less than 2 µm. In addition, it is difficult to assemble the piston so that it is exactly concentric within the cylinder — at least another 2 µm should be allowed for this.
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Having taken all these factors into account, it is clear that the flexure suspension system must have a linearity of no more than 3 or 4 µm to ensure that the clearance is maintained under all operating conditions. The aspects that need to be considered in order to achieve accurate linear movement of the flexure bearings can be broadly divided into the following areas: Geometrical accuracy of the flexures. Mechanical properties of the flexures and mounting components Geometric accuracy of the components (especially the surfaces) on which the flexures are mounted. The alignment of flexures and mounting surfaces to the desired datum during assembly. Historically, the flexures have always proved to be satisfactory providing care is taken in selecting material that has uniform thickness, acceptable flatness, and consistent mechanical properties. The flexures themselves are photo-etched to a high precision. The work described herein concentrated on looking at the quality of the mounting surfaces and the alignment achieved during assembly. The approach that has been taken is to build a test rig in which linearity measurements could be made on a well-defined bearing system – i.e. a system where the component and alignment accuracies could be measured. As far as was possible, these measurements looked at the effect of introducing a single geometrical deviation at a time. This was done to help achieve a clear relationship between cause and effect. A model of a flexure bearing system was also developed to form some framework for explaining the results.
DESCRIPTION OF FLEXURE BEARING USED The flexure used for the experiments was typical of the designs used in cryocoolers made at Oxford, and is a flat spiral disc spring similar to that shown in Figure 1. In this geometry the flexing movement is produced by arms that connect inner and outer annuli, these arms being defined by the spiral slots. Flexing allows the planes defined by these annuli to move easily with respect to each other by rotation perpendicular to the flexure axis, displacement along the flexure axis or a combination of both. When the inner and outer annuli are parallel, the displacement is along an axis perpendicular to the annuli. When the inner and outer annuli are not parallel the direction of displacement is not readily defined. The movement that meets the greatest resistance in this type of flexure is a relative radial movement between the two annuli.
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A flexure bearing is formed by spacing two sets of flexures apart and connecting the corresponding inner and outer annuli. If the flexures are mounted so that all the inner and outer annuli are parallel then it will be seen that a single displacement axis is defined. Radial movement or rotation perpendicular to this axis is strongly resisted on account of the radial movement that it causes in one or both of the flexures. In this way the assembly behaves like a sliding bearing in that movement is constrained to a particular direction. Figure 2 shows a simple arrangement for a flexure bearing and identifies the mounting surfaces that ideally ought to be parallel. Ideally, when assembling a cryocooler, it would be useful to measure the alignment of the flexure mounting surfaces (and if possible correct misalignments) before assembling the flexures in place. In practice, though, this is rarely possible, as there is no space in a real machine for the necessary datum surfaces, and access is often difficult for a probe. It is common practice to verify the motion of the assembled device. This is typical done before mounting the cylinder, by measuring the run-out of the piston (or vice-versa) using a submicron displacement transducer such as the Mitutoyu Mu-Checker (Bailey et al1). A schematic of such a set-up is shown in Figure 3. Of course, such a measurement as this not only measures the effects of movement, but also the geometry and surface finish of the piston. Typically a series of measurements will be made around the periphery of the piston, and from this some idea of the complete movement can be obtained.
EXPERIMENTAL METHOD The experimental aim was to determine how the linearity of motion depended on the flexure mounting conditions. To achieve this, several sets of inner and outer spacers were made with the
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faces either parallel or deliberately tapered. The inner spacers have a precisely machined bore to represent the cylinder (or piston), and by taking the flexures through a stroke with a displacement transducer on the bore of the inner spacer, it would be possible to measure the radial movement, and hence linearity, of the flexure bearing system. The test pieces were assembled and the resulting geometry measured using a ‘Talyrond’ roundness measuring machine. This did not give absolute values, but, having established a particular surface as a datum, enabled the relative misalignment of other surfaces, with respect to that datum, to be determined. In a cryocooler, the motion would be provided by a linear motor, but a suitable device was not available, and the manufacture of one was beyond the scope of this project. Instead the flexure bearing assembly was mounted on an Instron tensile testing machine, and this was used both to provide the necessary motion, and to act as a displacement transducer. The test assembly is shown schematically in Figure 4, and photographically in Figure 5. In a normal compressor the outside of the flexures is fixed, with the inside moving, but this was reversed in the Test Assembly, with the inside supported in place by a ‘pusher’ fixed to the base of the Instron, while the outside of the flexures was moved up and down by the Instron ‘crosshead’. To eliminate any backlash, a heavy brass mass was added to the underside of the inner spacer, so that when unsupported, the flexures would droop under gravity by over 4mm. Initially the axial force was transmitted from the ‘pusher’ to the ‘inner spacer’ by means of a precision ball resting in
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a recess on top of the ‘pusher’. It was found that this arrangement could transmit transverse forces and thus affect the radial motion of the system, so a small thrust ball bearing race was added under the single ball. With this new arrangement it was found that a considerable sideways force exerted on the ‘pusher’ would result in only a small radial movement as measured by the MuChecker probe. It was thus determined that only purely axial forces could be transmitted to the inner spacer of the test assembly. The Mu-Checker probe was attached to a bracket above the assembly and could be positioned at twelve different angular positions, and at four heights (See Figures 5 and 6). The radial movement as recorded by the Mu-checker, and the axial displacement taken from the Instron were input to a laptop computer via a data logger, and the results were then viewed and analysed using an Excel spreadsheet. Some typical results are shown in Figure 7, which shows repeatability at the sub-micron level.
TEST RESULTS A variety of inner and outer spacers were manufactured so that the effect of a mis-alignment on both the inner and outer clamp faces could be studied. Four outer spacers were made with the two faces out of parallel by 0.0 (parallel) 0.2, 0.4 and 0.6 mm. Four inner spacers were made with the two faces out of parallel by 0.0 (parallel), 0.1, 0.2 and 0.3 mm. These misalignments were deliberately made much larger than the values normally expected in a cryocooler, so that the resulting run-outs could be more readily detected. The original intention had been to use all of these test pieces to determine quantitively the relationship between the misalignment of the flexure mounting surfaces and the resulting motion. However, problems were discovered in the method of assembling and aligning the test pieces, and this original plan had to be curtailed.
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In the event, only the parallel, and the worst spacer of each type were used (the 0.3 taper inner and the 0.6 taper outer). Having overcome initial problems with the assembly, alignment and measurement of the test pieces, a trial assembly, with the complete assembly parallel and aligned, produced good results, with a very low run-out measured (Figure 7). Some of the early test results indicated that the motion defined by the flexures was not linear, so for one particular configuration (inner parallel, outer with 0.6 taper and misaligned), a complete set of measurements was taken at 30° angles around the perimeter and at 4 different axial positions in the bore. This data was reduced to polynomial form and standardised so that all readings had zero run-out at mid-stroke. These readings were then used to define the motion in each of the twelve angular positions in terms of an offset from the mid-stroke and a change of angle. It could be clearly seen that the motion was curved (Figure 8), and that the maximum run-out was occurring not in line with the maximum taper, but at 30° to it (Figure 9). Using the same components as above, the assembly was then aligned so that the mid-plane of the taper on the outer component was aligned with the inner part. The run-out readings are shown in Figure 10 and clearly shows that the run-out is much lower, and the motion is more symmetric about the upper and lower faces.
MODELLING THE FLEXURE Modelling the movement of the flexures using finite element methods was considered to be beyond the scope of this work and instead a very much simpler approach was adopted. This approach is described below for the spiral spring type of flexure used at Oxford.
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If a single spiral flexure arm is considered in isolation it is found that its stiffness is very different in different directions. Consider such a single arm between a point at inner end of the arm, and at the outer (see Figures 1 & 11). The direction of least stiffness is for relative movement between Pi and Po along the Z-axis (i.e. out of the paper). This is because the movement constitutes the bending of a thin section – the minimum flexure arm width being typically 5 to 10 times the flexure thickness. This direction corresponds to the direction of movement in a flexure bearing. The direction with the next lowest stiffness is along the Y-axis, again a bending movement perpendicular to the arm, but this time in the plane of the flexure. The stiffness is higher because the section is thicker and the stiffness varies according to the cube of the thickness. Finally the direction in which the stiffness is highest is along the arm. This movement corresponds to tension and compression of the arm without any bending. For flexure arms that do not have high angles of curvature, this can be approximated by defining the direction as the line joining the two ends of the arm along the X-axis as in Figure 11. From these considerations, the idea emerges that stiffness along the Y and Z axes can be ignored and that the movement of the flexure arms will be determined solely by considering the strain along the arm i.e. in this approximation, the X axis. Each flexure can then be treated as an assembly of struts. The struts can pivot at their attachment points, they can extend or contract but they are assumed not to bend (Figure 12). In calculating the movement defined by the flexure, the model seeks convergence for the particular case where the strains in these ‘struts’ are minimised In a typical cryocooler, with the flexures at both ends aligned the same way, the piston is free to rotate, and under these circumstances the strain in these struts will be minimised to zero. This approach is clearly a fairly gross simplification and would not be satisfactory if the model was intended to give accurate values of the flexure movement. For the purposes of this study however it was thought that the general behaviour and magnitudes that it generated might be close enough to be useful. In addition this model is a purely strain-based model. A more exact, but far more complex analysis would consider both stresses and strains in the flexure.
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The more detailed assumptions for the model are: A four-arm flexure is modelled for simplicity. It is assumed that characteristics demonstrated will be applicable to the six arm flexures that are more frequently used. To a first approximation the motion of the flexure bearing will be the motion that causes least strain along the flexure arms. The arms are modelled as extendable struts mounted between inner and outer components. The struts are mounted such that they can freely pivot about their mounting points. The struts are assumed not to bend. If the flexures are allowed to rotate then for typical designs a movement can be defined which results in zero extension along the arms. For this simple model, in the ideal case where everything is perfectly aligned, only one set of axes need to be defined to demonstrate the motion. When there is misalignment, it is necessary to use several different sets of axes, each relating to one of the significant surfaces defined in the model. A large part of the modelling process involved defining the relationship between these different sets of axes, in particular. To simplify these relationships, it became clear that the angular alignment of a pair mounting surfaces (e.g. outer or inner) could be expressed as half angles relative to a mean plane.
RESULTS FROM THE MODEL The model was used both to analyse some of the specific experimental cases, as well as being used to look at the flexure behaviour under certain ideal conditions. In particular, the model showed that for three different types of misalignment, three different ‘run-out’ curves resulted. The inner surfaces are parallel and the outer surfaces are not parallel, but the inner and outer are aligned. In this case the run-out varies depending on the position at which it is measured. Taken in the middle of the bore, the run-out is extremely small, but is much higher at the ends (Figure 13). The outer surfaces are parallel and the inner surfaces are not parallel, but the inner and outer are aligned. In this case the run-out also varies depending on the axial position at which it is measured. Taken in the middle of the bore, the run-out is extremely small, but is much higher at the ends (similar to Figure 13). The motion is not linear, but curved. These results are in good agreement with the measured values for this configuration (Figure 10).
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The inner surfaces are parallel, and the outer surfaces are parallel, but the inner and outer have angular misalignment between them. In this case the run-out will be similar wherever it is measured. The motion is linear, but is not aligned with desired datum axis (in this case the bore of the inner spacer) (Figure 14). Inner surfaces parallel, outer surfaces parallel, and inner and outer parallel to each other, but there is a radial offset between inner and outer. In this case, the motion is clearly not linear, being a maximum in one direction at mid-stroke, but also having maxima in the other direction at the extremes of the stroke. In this case, if the run-out is measured half-way up the bore it will be minimum, and it will be worse at each end. The run-out for this case is much smaller than the other cases, and will almost certainly be acceptable for most cryocoolers (Figure 15). It was noted that though the model predicted the maximum run-out not directly in line with the ‘peak’ of the taper, but was some angle away from this.
CONCLUSIONS Considerable difficulty was encountered in assembling, aligning and measuring the test pieces in a way that could be readily defined and compared with the theoretical model. However, there is evidence to show that the model does predict the motion that could be expected from a misaligned flexure bearing system. Were sufficient resources available, some of these tests could have been repeated, and it is likely that more agreement would be obtained. Despite this, some definite conclusions can be drawn from this work:Alignment can be described by angular and offset parameters that define deviations from perfect alignment.
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The alignment between inner and outer mounting surfaces that is achieved during assembly has a significant impact on the final run-out. This alignment can be defined by the angular disposition of the two mean planes with respect to each other. By studying the shape of the curve obtained from run-out measurements, some idea of the geometric cause of the run-out can be determined. If both inner and outer flexure mounting components are not parallel, it is unlikely that good linearity will be obtained. The maximum run-out measured may not occur at the same angular position of maximum taper on the flexure mounting surfaces. For typical assemblies, the model shows that angular misalignments are more significant than offset ones. Tapered mounting surfaces result in a curved motion, and a run-out that is very dependent on the axial position of measurement. The run-out measured half way between the flexures can be very small even for significant tapers. Angular misalignment during assembly will give a run-out that is linear with flexure deflection and independent of axial position. The model appears to predict generally the right qualitative and quantitative behaviour. However it is not very accurate in predicting the right angular orientation of the run out.
ACKNOWLEDGMENTS We acknowledge the strong support of Thom Davis and B.J Tomlinson of AFRL for this project. 1.
Bailey, P.B., Dadd, M.W., Hill, N., Cheuk, C.F., Raab, J., Tward, E., “High Performance Flight Cryocooler Compressor”, Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York, 2001., pp 169-174.
Piston Resonance in the Orifice Pulse Tube P.C.T. de Boer,1 J.-M. Duval, I. Charles and L. Duband Département de Recherche Fondamentale sur la Matière Condensée Service des Basses Températures, CEA-Grenoble 17 rue des Martyrs, 38054 Grenoble Cédex 9, France
ABSTRACT The force exerted on the piston of a linear motor driving an orifice pulse tube is analyzed using linearized theory. First, the case of a pulse tube without reservoir is considered. It is found that there are two possible resonances, corresponding to a very small and a very large orifice conductance, respectively. Such resonances can be used to reduce the amplitude of the current provided to the linear motor, and hence to reduce associated ohmic losses. Use of the resonances thus helps to maximize the power delivered to the pulse tube. An expression is derived for the dissipation in the various chambers associated with heat losses to the walls. Experimental results obtained for the amplitude of the piston force indicate that there is a large influence of turbulent effects on thermal diffusivity. The analysis is extended to the case with regenerator. It is found that large temperature ratios across the regenerator as well as heat losses to the walls tend to decrease the piston resonance effect.
NOMENCLATURE cross sectional area conductance
see Eq. (2) specific heat see Eq. (39) distance between parallel plates force mass flow rate average pressure
Greek Symbols
thermal diffusivity, denotes amplitude phase angle ratio of specific heats characteristic thermal boundary layer thickness,
density frequency, see Eqs.16, 22 and 36 angular frequency
1
Present address: Cornell University, Ithaca, NY 14850.
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pressure rate of heat flow into system heat flux into wall surface area temperature time internal energy volume rate of work done by system distance from center plane; also piston displacement
Subscripts buffer (reservoir) cold driver hot orifice; average pulse tube regenerator wall
INTRODUCTION In the last decade or so, the pulse tube has been developed as an efficient and reliable cryocooler [1]. Many important improvements have been made. These include the introduction of an inertance tube rather than an orifice to provide the required phase difference between mass flow rate and pressure [2, 3]. The inertance tube is useful only for pulse tubes operating at relatively high frequencies. For pulse tubes operating at relatively low frequencies, a major improvement has been the introduction of a second orifice in a tube that short-circuits the regenerator [4]. Current efforts toward further improvements mostly concern attempts to decrease various losses inherent in the operation of the tube. A potential improvement in the operation of pulse tubes driven by linear motors is the reduction of ohmic losses associated with the electric current provided to the motor. The piston driven by the motor can be at resonance with the pressure pulsations in the drive chamber. The resonance corresponds to a minimum of the amplitude of the force exerted by the linear motor to the piston. This force is approximately proportional to the current passing through the motor. Operating the pulse tube at or near piston resonance thus reduces ohmic losses, and is a requirement for achieving high efficiency [5]. Operation at piston resonance is especially important if the amplitude of the current delivered to the linear motor is restricted. The value of the amplitude of the force at resonance is limited by the occurrence of dissipation in the drive chamber and pulse tube, caused by irreversible heat transfer to the walls. Mirels [6] used the linearized result presented by Lee [7] for the rate of this heat transfer. This result is based on assuming molecular diffusivity in the thermal boundary layer as well as in the region outside the boundary layer. The heat transfer is a second order effect ([6], Appendix B), which arises from the phase difference which exists between pressure and first order heat transfer. Theoretical results of this model are in satisfactory agreement with experimental data for a single gas spring without high velocity inflows [8]. Later work carried out in connection with Stirling engine research by Cantelmi et al. [9] showed that in actual engine cylinders, inflow generated turbulence can significantly increase the rate of heat transfer to the walls. To account for this effect, Cantelmi et al. [9] used a turbulence enhanced thermal diffusivity. The present work begins with extending the treatment of [9] from the case of a single wall in contact with a semi-infinite gas to the case of two parallel walls. This extension is of importance in the case of large enhancement of thermal diffusivity by turbulence. Next, the theory presented by Mirels [6] is extended to a pulse tube with heat losses in the reservoir and the pulse tube, as well as in the drive chamber. Expressions are developed for the amplitude and the phase angle of the force exerted on the piston. Frictional effects on the piston are
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left out of account. The resonance phenomenon is illustrated by considering the case of zero heat losses, with the orifice closed. The piston amplitude is assumed to be small enough to justify neglecting the force exerted by the flexure bearings. Piston resonance corresponds to a minimum of the amplitude of the force on the piston. In the case considered, piston resonance occurs at both small and large values of the dimensionless conductance of the regenerator. The power delivered by the piston to the gas does not exhibit a minimum at piston resonance. In cases where the maximum force on the piston is limited by the maximum allowable amplitude of the current provided to the linear motor, operation at or near piston resonance is crucial for maximizing the power delivered to the pulse tube. The expressions derived for the force are compared with experimental data. The latter were obtained in a set-up similar to that of a pulse tube, but with the regenerator replaced by a valve of variable conductance, and without an orifice and reservoir. The experimental heat losses to the wall are found to correspond to a large enhancement of diffusivity by turbulent effects. As a result of this enhancement, the resonance effect tends to be significantly reduced. The last section of the paper considers the complete orifice pulse tube. It is found that the temperature difference across the regenerator also tends to reduce the resonance effect significantly. Following the analyses of several of the works cited, the present treatment uses complex notation. Physical quantities are given by the real parts of the corresponding parameters.
RATE OF HEAT TRANSFER TO THE WALL The rate at which heat is transferred from the gas to the wall can be estimated using the method described by Lee [7] and extended by Mirels [6] and Cantelmi et al. [9]. The method is based on using the simplified energy equation
The quantity represents the ratio of the turbulent eddy diffusivity to the molecular diffusivity. Cantelmi et al. [9] considered the case of a single plate in thermal contact with a semi-infinite gas. They took to be proportional to the distance from the wall. Here, Eq. (1) is applied to the case of two infinite parallel plates. The ratio is taken to be proportional to the distance to the closest wall
The solution of Eq. (1) under the boundary conditions at is
at
and
where and where K and I are the modified Bessel functions. Cantelmi et al. [9] took to be proportional to the ratio of the cross section of the cylinder to the cross section of the inlet tube. They noted that the value of is a measure of the magnitude of the eddy diffusivity in the thermal boundary layer. The heat flux to the wall is given by
where
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ANALYSIS OF ORIFICE PULSE TUBE The model on which this analysis is based is sketched in Fig.l. The analysis is an extension of that presented by Mirels [6]. The three chambers are analyzed sequentially. Each of them is subject to the time-dependent thermodynamic law of energy conservation for a control volume with at most one inlet and one exit
Here,
where use was made of the ideal gas law. The pressures and mass flow rates are written as follows
Application of Eq. (6) to the reservoir (” buffer”) involves setting and This yields, after some algebra
where
It follows that
Next, Eq.6 is applied to the pulse tube. Noting that now and
this yields
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where
It follows that
Application of Eq. (6) to the driver section requires setting Setting furthermore
it is found that
where
It follows that
Taking the value of in the foregoing results is 1/3 for 2 infinite plates. Provided the results may be applied to other geometries by substituting the actual value of the surrounding surface for S, and the value of the actual volume for V [6]. For a cylinder this yields G = 2/3, for a sphere G = 1.
FORCE ON PISTON The force on the piston is found from the equation
Working this out yields
from which
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where use was made of Eq. (26), and where
RESONANCES AT ZERO HEAT LOSS The resonance character of can be illustrated by considering the simplified case of zero heat loss. For the purposes of this illustration, the orifice is taken to be closed and the regenerator temperature is taken to be uniform For this case, This leads to two possible resonances. One of these obtains in the limit of zero regenerator conductance which corresponds to a geometry consisting of the drive chamber, only. In this case, and
The second resonance obtains in the limit of infinitely large regenerator conductance which corresponds to a geometry consisting of a single chamber having the combined volume of the pulse tube and the drive chamber. Now and
Here
is the resonance frequency of the combined volume.
The resonance at corresponds to In considering intermediate values of the dimensionless conductance of the regenerator is defined as
Hence Consequently, can be plotted as a function of for given values of the parameters and Such a plot is shown in Fig. 2 for and for various values of The resonances at and manifest themselves clearly. At values of near unity, the magnitude of is dominated by dissipation in the regenerator, and the resonance phenomenon disappears. Corresponding results for the phase angle are shown in Fig. 3. In the limiting cases and this phase angle jumps from 0 degrees just below resonance to 180 degrees just above. Its value at resonance is 90 degrees. POWER DELIVERED BY PISTON The power delivered by the piston is given by
Using Eq. (32) to substitute for
there results
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where the dimensionless power P equals
As can be seen from Eq. (42), the product contains the frequency-dependent part of while the pre-factor contains the dependence of on A, and for given value of The product is plotted in Fig. 4 as function of for the values of and also used in Figs. 2 and 3. It is seen that the power delivered to the gas at constant pre-factor gradually increases with frequency, and does not exhibit minima corresponding to the minima in at and As noted before, the phase angle equals 90 degrees at these minima. This means that the sine of this angle is at its maximum value. As a result, the product appearing in Eq. 41 for does not pass through a minimum.
ESTIMATE OF In order to assess the influence of heat losses to the wall, the power P representing these losses is shown in Fig. 5 as function of for the case of a single chamber. The single chamber geometry corresponds to setting It follows that in agreement with [6]. The heat transfer to the walls was found from the result for two parallel plates listed previously . The value of G was taken equal to 1. At large values of the results
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are seen to depend strongly on the value of This indicates a large effect of turbulence as compared with the laminar case The results are essentially independent of in the nearly isothermal region The power dissipated goes to 0 in the isothermal limit as well as in the adiabatic limit The previous work by Cantelmi et al [9] has shown that heat losses to the walls in Stirling engines correspond to values of about 0.3 to 0.6. As part of the present work, experiments were carried out to determine a value of appropriate for a geometry characteristic of pulse tubes. The experimental set-up used was similar to that sketched in Fig. 1, except that there was no reservoir. The regenerator was simulated by an adjustable valve. The piston was connected to a linear motor. Measurements were made of the current delivered to the motor as a function of time under zero-load conditions. This current is proportional to the force on the piston. Provided the amplitude of the piston motion is not too large, the spring force exerted by the flexure bearings of our compressor is negligible. The force delivered to the piston then serves only to overcome the inertial force The mass m of the piston is known (m = 0.165 kg). The electronics provided with the motor yielded the value of the amplitude This allowed determination of the inertial force, and hence of the constant of proportionality between current and force. The resulting value was 15.9 N/A. The electronics provided with the motor also allow determination of the phase angle between the force and the displacement x(t) of the piston. The pressure at which the measurements were taken was The volumes of the drive chamber and the reservoir were and respectively. The cross sectional area of the
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piston was the piston amplitude was Experimental results were obtained at several positions of the adjustable valve. These corresponded to 0.5, 1.5, 2.5 and respectively, as determined from pressure amplitudes at the two sides of the valve. The results obtained for the force amplitude are shown in Fig. 6, together with corresponding theoretical curves obtained with It is seen that the theoretical curves generally lie below the experimental data, even though the value used for is very large. Weak resonance effects occur at values of less than 1. While not shown here, the corresponding theoretical values for the phase angles are larger by about 40 degrees than the experimental data. It appears that the theory described can provide an estimate of the order of magnitude of the heat losses to the wall, but does not yield accurate quantitative results. It may be that the heat losses are determined by convective effects not taken into account here, rather than by turbulence enhanced diffusion. APPLICATION TO ORIFICE PULSE TUBE The amplitude of the force on the piston for the case of a complete pulse tube is plotted in Fig. 7. The values used for geometric quantities are the same as those used in obtaining Fig. 6. Additional values used are and The latter value represents a level of turbulence considerably lower than that corresponding to the measurements reported in the previous section. The value represents a regenerator with a low conductance, offering good prospects for making use of the resonance
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in the drive chamber (cf. Fig. 2). Very large values of tend to correspond with poor heat transfer between the regenerator and the gas, and hence are likely to be unacceptable. In carrying out the computations, use was made of the identity
which follows from Eqs. (16) and (22). The value used for
is given by
where As shown in [10], this expression optimizes the enthalpy flux for given amplitude of the piston. While not shown here, the dimensionless power gradually increases with frequency also in this case. Fig. 7 shows that there is a pronounced resonance effect for low values of However, this effect disappears for values of that are of practical interest. Dissipation corresponding to values of larger than 1 would make the resonance effect disappear at still lower values of Conversely, it was found that for values of and much smaller than one there is a pronounced resonance even at high values of It can be concluded that beneficial use of piston resonance in pulse tubes is possible only when the dimensionless conductance of the regenerator is low, and when the heat losses due to turbulence and convective effects are small.
REFERENCES [1] Radebaugh, R., ”Development of the pulse tube refrigerator as an efficient and reliable cryocooler”, The Proceedings of the Institute of Refrigeration, vol. 96, 1999-00 (ISBN 1 872719 14 7) pp. 11 - 31, [2] Zhu, S.W., Zhou, S.T., Yoshimura, N., and Matsubara, Y., ”Phase shift effect of the long neck tube for the pulse tube refrigerator”, Cryocoolers 9, Plenum Press, NY, (1997), pp.269-278. [3] Gardner, D.L. and Swift, G.W., ”Use of inertance in orifice pulse tube refrigerators”, Cryogenics, vol. 37 (1997), pp.117-121. [4] Zhu, S.W., WU, P.Y. and Chen, Z.Q., ”Double inlet pulse tube refrigerators: an important improvement”, Cryogenics, 1990, vol. 30, pp. 514-520. [5] Martin, J.L. and Martin, C.M., ”Pulse tube cryocoolers for industrial applications”, paper presented at the 2001 CEC Conference. Available at www.mesoscopic.com, Cryogenics group. [6] Mirels, H., ”Effect of orifice flow and heat transfer on gas spring hysteresis”, AIAA Journal vol. 32, (1994), pp. 1656-1661. [7] Lee, K.P., ”A simplistic model of cyclic heat transfer phenomena in closed spaces”, Proceedings of the 18th Intersociety Enery Conversion Engineering Conference, IEEE, Boston, (1983), pp.720-723. [8] Kornhauser, A. A. and Smith, J.L., Jr., ”The effects of heat transfer on as spring performance”, Proceedings of the 26th Intersociety energy conversion engineering conference, vol.5, IEEE, Boston (1991), pp. 180-185. [9] Cantelmi, F.J., Gedeon, D., and Kornhauser, A.A., ”An analytical model for turbulent compression-driven heat transfer”, J. Heat Transfer, vol. 120 (1998), pp. 617-623. [10] de Boer, P.C.T., ”Optimization of the orifice pulse tube,” Cryogenics, vol. 40, (2001), pp. 701-711.
Producibility of Cryocooler Compressors C. F. Cheuk, N. G. Hill, R. Strauch The Hymatic Engineering Company Limited, Redditch, UK P.B. Bailey Oxford University, Oxford, UK Jeff Raab TRW, Redondo Beach, CA, USA
ABSTRACT This paper describes the high yield rate, the quality process, and the high performance uniformity among the 12 space qualified HEC (High Efficiency Cryocooler) compressors that have been fully assembled and tested over a period of 15 months. The number of compressors produced allows initial SPC (Statistical Process Control) results of process capabilities to be assessed. 100% yield in final assembly as well as in sub-assembly processes and tests have been achieved. Acceptance tests include compression tests and high temperature friction tests to assure frictionless nonwearing operation of the compressors over their wide operating temperature range. A summary of the manufacturing experience of producing small clearance frictionless compressors is presented. The paper also recommends process enhancements and new testing methods and equipment for future manufacturing.
INTRODUCTION The HEC compressor was developed at Oxford University to supply the pressure source for the TRW High Efficiency Cryocooler1. The compressor is a linear motor, flexure spring, true clearance seal, frictionless configuration designed for high efficiency and low mass for space applications. The compressor has been productionised by Hymatic, with the aim of making the compressor more rugged, and also introducing a fully controlled assembly and test process more suitable for repeated and consistent quantity production2. In a period of 15 months, twelve HEC compressors have been built and tested at Hymatic using new assembly and test processes, and have been delivered to TRW. To ensure high yield, Hymatic has implemented quality control systems at the component manufacturing stage and at the assembly stage of the compressors. The statistical capabilities of the manufacture and test processes are examined and reported in this paper.
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QUALITY PLAN Formulation Based on the performance specification and test requirement of the compressor, and drawing on Hymatic’s experience in manufacturing frictionless cryocoolers, a Failure Mode Analysis was performed on the compressor design. Critical features for each component part and subassemby were identified. In addition to the general drawing dimensions, Key Features were highlighted on the part drawings to focus attention to areas critical to the successful operation of the compressor. Sub-assembly and final test plans were also drawn up to detail the specific measurements required for the success of each assembly operation. 100% inspection of the Key Features was enforced.
Part Quality Some component parts were machined in-house and some were sub-contracted. The quality strategies for the two routes were different. For in-house machined parts, the machine operators were responsible for 100% inspection on all the features of the parts. After machining, the Key Features of the parts were re-inspected 100% by independent inspectors and the results recorded to give objective evidence that the parts comply with the drawing requirements. For sub-contracted parts, the subcontractors were responsible for the quality of the parts and they submitted their inspection reports of the Key Features to Hymatic. At Hymatic, the Goods Inward Inspection department verified that the Key Features inspection reports were complete and performed random re-inspection to audit the sub-contractor’s inspection results.
Sub-assembly and Final Assembly Quality In a batch production environment it was felt unwise to assume that components that passed the Part Quality control procedure were automatically suitable for assembly. Transit and handling damage can and will occur. Features that would affect the quality of the assembly build were reinspected immediately before assembly. The result was that only correct parts were assembled. During each sub-assembly stage specific tests were conducted to measure the degree of success of the sub-assembly. The tests performed at each stage of the sub-assembly are summarized below in Table 1. The result of these inter-operation tests was that only good sub-assemblies were allowed to proceed onto the next stage. Finally, the assembled compressors were put through the performance tests specified by the customer.
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PROCESS CAPABILITY The disciplined data recording procedure and the automatic data logging feature of the test facility provides objective evidence of the capability of the assembly processes. In excess of 800MB of data were collected for the 12 compressors. The capability of critical processes were summarized below. The data shown applies to the compressor pairs and additional test pieces which were tested under similar conditions; some test results were not used because they were tested at conditions before the parameters were finalized:
Production Yield The production yield of the 12 compressor pairs was 100% at the final test as well as at all the sub-assembly tests. No re-work was necessary at any assembly stage. The quality plan, the vigorous process design, and the attention to detail in assembly workmanship have proved to be effective.
Motor Flux Density Magnetic flux density in the linear motor air gap affects the driving force available to the motor. The higher the flux density the more efficient is the compressor. The magnetic flux attainable is a function of the magnet material, pole piece material, dimensional control of the motor components and the magnetizing machine. The distribution of the average flux density in the air gap is shown in Fig 1. The average flux density was measured using a fixed separation ganged search coil averaging the flux density over a fixed distance. The mean of the average flux density was 4.46 mWb-Turn or 0.67 Tesla. The percentage variation of the flux density was expected to be less than 4.45% over +/-3 Sigma or 99.7% of the total population.
Spring Alignment Two sets of flexure springs suspend the piston. Accurate alignment of the springs is required to ensure that the piston moves along its own axis. It is this coaxiality of movement that ensures frictionless compressors. The spring alignment was measured before and after the alignment settings were locked. The distribution of the linearity of the alignment is shown in Fig 2. The data show that the locking procedure has little effect on the alignment. The mean alignment error is 3 microns, and at the +3 sigma limit, the alignment error is expected to be less than 7 microns. The capability of the alignment process compares favorably with the required clearance between the piston and the cylinder of the compressor.
Resonant Frequency in Vacuum Each compressor half was tested for its resonant frequency in vacuum. The results of this test are an indication of the consistency of the mechanical spring mass system.
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The distribution of the resonant frequency is shown in Fig 3. It can be seen from the histogram that there are 2 families of results. This variation was traced to two batches of springs that were used in the 12 compressor pairs. The thickness of the springs within each batch was consistent, but the mean thickness values of the two batches were at either end of the tolerance range. The mean resonant frequencies were 37.33 and 38.77Hz for the weak- and strong-spring batches, respectively. The corresponding expected variation over +/-3 sigma population is 8.3 and 8.1 Hz. The design of the compressor includes a feature to tune the final drive frequency. The small variation of resonant frequency in vacuum is not considered as being significant to the operation of the compressors. Friction Test At Room Temperature At different stages of the build (see Fig 1), the integrity of the sub-assemblies was verified by a set of 3 friction tests. The three tests were Stiction Test, Static Friction Test and Ring Down Test. The Stiction Test checkes the dynamic friction between the moving parts. The Static Friction Test measures the stick-slip characteristics. The Ring Down Test measures the friction as a function of its damping effect on oscillation of the piston in the cylinder. All of the compressors passed the preset limits of the three friction tests at all stages of their builds. No stick-slip was detected at full, quarter, or mid-stroke positions. A sample result of the Static Friction Test is shown in Fig 4. Friction Test At High Temperature To avoid overheating the coil, only the Ring Down Test is used to detect the maximum ambient temperature before friction sets in. This temperature is considered an indication of how even the gap is between the piston and the cylinder, and it sets the maximum operating temperature limit of the compressor.
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The temperature when friction occurred varied from 60°C to 90°C. Considering the maximum operating temperature of 50C, there was a healthy minimum margin of 10°C. However, the large variation was an indication of process variability. Studies were done to correlate the variation with diametrical piston to cylinder clearance and the pressure drop through the clearance as shown in Figs. 5 and 6. No correlation was found. Further work is required to identify the factors that limit the capability of the piston and cylinder gapping procedure.
Pressure Swing of Compressor Half The peak-to-peak pressure swing of the compressor halves when connected to a fixed-size dead volume was tested at 75% and 90% stroke levels as well as at 10W fixed power. The compressor halves were charged to the final charge pressure of the system and the dead volume was initially
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tuned to resonate at the system frequency. This test represents the capability of each half to satisfy the functional requirement of the compressor. Figure 7 shows the consistency of the pressure swing obtained. Over +/-3 sigma of the population, less than 11.7 % variation of the 77.01psi mean pressure swing is expected at a fixed 10W input power.
Pressure Swing of Compressor Pair This test is similar to that of the compressor halves. However, 20W input power and a different representative system test volume were used. Figure 8 shows the consistency of the pressure swing obtained. Over +/-3 sigma of the population, less than 10.5 % variation of the 72.16 psi mean pressure swing is expected at the fixed 20W input power.
MANUFACTURING AND ASSEMBLY EXPERIENCE The quality audit procedure revealed that not all subcontractors’ inspection reports were reliable. It was important not to take data at their face value. Rechecking functional features was key to successful assembly. Subcontractors needed to be educated on the specific requirements of part features for cryocooler applications. It was through this understanding that the manufacturing personnel improved their part quality. The basic rule of manufacturing engineering—to produce the most accurate dimension first and the best finish last — must be enforced. The cleaning and out-gassing procedures for each component must be defined. Particular attention should be paid to blind threaded holes which harbor contamination easily. Electrical insulation breakdown could be a significant cause of failure if sharp edges adjacent to electrical insulators are not removed. Dragged out fraze and loose metal particles may pierce through thin insulation causing an electrical short. Any anaerobic adhesive that is spilt onto an exposed surface must be cleaned off, as it will not cure in the presence of air. Thus, epoxy resin is preferred.
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Automated test and data logging equipment was essential in reducing the test time, improving the test consistency and minimizing human errors.
RECOMMENDED EQUIPMENT Accuracy of dimensional measurement on piston and cylinder can only be ascertained with the use of non-contact measuring instruments. Frequent calibration and temperature stabilization is needed to overcome thermal drift. A means of measuring cylindricity, coaxiality and squareness was required to verify the level of success of assembly processes and part quality before final assembly. Hymatic has purchased a Talyrond 265 roundness checking machine and dedicated it for cryocooler manufacturing inside the cleanroom. Significant benefit had been realized with the use of the machine.
CONCLUSION Statistical evidence confirms that the HEC compressor is a mature and producible design. Through continuous improvement in process design, assembly workmanship, manufacturing equipment and close customer liaison, a batch manufacturing system has been established in Hymatic to produce long life friction free HEC compressors at a consistent quality level to meet the stringent requirement for space application.
ACKNOWLEDGMENT We acknowledge the strong support of Thom Davis of AFRL for this project.
REFERENCES 1.
Tward, E., Chan, C. K., Raab, J., Nguyen, T., Colbert, R. and Davis, T., “High Efficiency Pulse Tube Cooler”, Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2001), pp. 163-167.
2.
Bailey, P. B., Dadd, M. W., Hill, N., Cheuk, C. F., Raab, J. and Tward, E., “High Performance Flight Cryocooler Compressor”, Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2001), pp. 169-174.
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Helium-3 Pulse Tube Cryocooler I.A. Tanaeva and A.T.A.M. de Waele Department of Physics Eindhoven University of Technology NL-5600 MB Eindhoven, The Netherlands
ABSTRACT A three-stage pulse tube refrigerator has been developed for the purpose of reaching temperatures as low as 1.5 K using as the working medium. This work is a continuation of the research on a three-stage pulse tube refrigerator started in 1999 at Eindhoven University of Technology; during that effort a minimum average temperature of 1.78 K was achieved. The size of the three-stage refrigerator is small, compared to other sub-4K pulse tube cryocoolers, in order to have a small amount of gas in the system. The regenerator plays an important role in the performance of the refrigerator. The regenerator was designed to be very flexible in order to be able to test different compositions of materials and their influence on the performance of the cooler. In this contribution we report on the progress in the development of the three-stage pulse tube cryocooler. We also describe our future plans, in which we intend to combine the pulse tube refrigerator with a superfluid vortex cooler in order to achieve temperatures below 0.7 K.
INTRODUCTION The goal of this research is to achieve the lowest temperature possible with pulse tube refrigerators (PTRs). Up until now, the record lowest temperature with PTRs (1.78 K) belongs to the Low Temperature group of Eindhoven University of Technology. This temperature was reached with a three-stage double-inlet PTR. As the lambda transition of was a barrier for reaching temperatures below 2 K (see Fig. 1), was used as the working fluid. As is rather expensive, it is useful to reduce the amount of gas used in the experiment. One more advantage of making the PTR smaller is the fact that less regenerator material is needed. Also, the input power of the compressor can be significantly reduced. Our new three-stage PTR has been designed with a much smaller size than our previous PTR, which was taken as a prototype for designing the new machine. The volumes of all the components of the previous 3-stage PTR were reduced by 50%. This was done by decreasing the diameters of the tubes and keeping the same lengths. The performance of this new machine will be discussed. The last section of this paper describes our future plans that are under development. We plan to combine the PTR with a superfluid vortex cooler in order to achieve temperatures as low as 0.65 K.
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DESIGN OF THE EXPERIMENTAL SETUP It is first useful to discuss why we designed the new PTR the way we did, e.g. by reducing the cross-sectional area of the prototype PTR. According to the thermodynamic model of a PTR, extensively treated in the previous publications of our group1-4, the relations determining the dynamics of the regenerator and the pulse tube are one-dimensional. The cooling power of the PTR can thus be described as where and are the cross-sectional areas of the pulse tube and the regenerator, respectively, is the energy flux in the pulse tube, and is the energy flux in the regenerator. As the energy flux, and the heat flux where is the heat flow due we consider a sum of the enthalpy flux to the thermal conduction. So, in the first order, reducing the cross-section by a factor of two and at the same time reducing the flow rate by a factor of two as well, should keep the lowest temperature the same. However, parameters that do not scale with the surface area, such as external heat leaks and the heat-shuttle effect, complicate this simple picture. The volume of the initial setup was not reduced by more than 50% since the heat-shuttle effect negatively influences the performance of the PTR. Therefore, the diameters of the tubes should be significantly larger than the thermal penetration depth in helium, so that the tubes can operate in the adiabatic mode. The thermal penetration depth in helium at room temperature at different frequencies as a function of pressure is shown on Figure 2. A frequency of 1.2 Hz was found to be optimum for the prototype PTR. Figure 3 shows a schematic diagram of the three-stage PTR. The dimensions of the original prototype and of the new PTR are given in Table 1. As can be seen from the table, the diameter of the 3rd stage tube was already rather small. Therefore, the choice of the dimensions for the third stage tube was a compromise between decreasing the diameter as much as desired and still keeping it at a value larger than the heat penetration depth. All tubes are fabricated from stainless steel. As the regenerator material, we are using stainless steel screens for the first stage regenerator, lead spheres for the second stage regenerator, and a combination (50/50%) of ErNi and for the third stage regenerator. Each stage of the cooler has three adjustable needle valves: the first orifice, the double-inlet, and the minor orifice. Therefore the system contains nine needle valves in total. The cooler is driven by a 4 kW compressor. The new design of the rotating valve requires only 14 V to provide a frequency of 1 Hz. Copper blocks with a few channels inside are used as the cold
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heat exchangers for the 1st and the 2nd stage. The cold heat exchanger for the third stage tube is made from sintered copper powder, filling the copper holder. The hot heat exchangers are watercooled. There are two copper heat shields in the setup, one that is precooled by the first stage (1st heat shield). The second one is attached to the second stage cold heat exchanger. The regenerator was designed to be flexible to provide easy access to each stage. As the regenerator is one of the most important components of the pulse tube system, it is very interesting to investigate different compositions of the regenerator materials and their influence on the performance of the refrigerator. It is very important to fix the material very well and make sure that it doesn’t move during operation, since material movement can degrade the performance of the cooler significantly.
EXPERIMENT At first the system was operated without the second heat shield to reduce the time constant when adjusting the orifices. The optimum frequency was found to be 2.05 Hz. That is the maximum frequency possible with the present motor of the rotating valve. It turned out that the temperature at about ¼ from the hot end of the third stage pulse tube was rather high. After connecting the minor orifice to the lowpressure side of the compressor, the temperature in that part of the tube deceased. Still the minimum temperature at the cold end of the third stage didn’t go below 7.6 K. That could possibly be due to the large surface heat pumping effect in the third stage tube. As we already mentioned in the design description, the large surface heat pumping was expected, as the diameter of the third stage tube was relatively small. Therefore it was decided to change the 3rd stage tube for one with a larger diameter. After changing the tube for a larger one, the high temperature peak, observed with the smaller tube, disappeared. The optimum frequency still remained the same, 2.05 Hz. After adjustment of the needle valves a no-load temperature of 4.03 K was reached. In Table 2 we can see the lowest temperatures reached at each stage under different conditions. As we can see from the first line of Table 2 the temperature of the second stage cold end is relatively high. That means that gas enters the part of the regenerator filled with ErNi having a temperature of 43 K. The heat capacity of ErNi is much lower than that of lead at this temperature. Therefore, a big part of the third stage regenerator is not as effective as it is supposed to be. From Table 2 we also see that the second stage regenerator cools the gas from 84 K to only 43 K. That means that there is poor heat exchange in the second stage regenerator. Therefore, the 0.4-0.48 mm diameter lead spheres were exchanged for lead particles with a diameter of 0.2-0.3 mm. In this way we were able to increase the heat exchanging area of the second stage regenerator. The results of the cool-down with smaller lead spheres in the second stage regenerator are shown in the Table 2. As we can see, the lowest temperature decreased from 4.03 K to 3.04K. The next step was attaching the second heat shield to the cold heat exchanger of the second stage to reduce the radiation heat leak on the coldest point of the cooler. With both heat shields in the pulse tube system, a lowest temperature of 2.31 K was achieved. We investigated the dependence of the lowest temperature on the input mass flow in the system. We reduced the flow into the cooler simply by opening a bypass valve connecting the high- and the lowpressure sides of the compressor. We found a broad optimum (see Fig. 4). The minimum temperature of 2.31 K was reached not with the largest pressure amplitude. As it turned out, the system is not very sensitive to the change in the inlet pressure amplitude. Therefore, a 2 kW compressor would have been sufficient to operate the system at the same level of performance.
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As the conclusion for this part of the paper we would like to point out that the lowest temperature reached with the new PTR is basically the same as the lowest temperature of the prototype PTR. Our initial idea of making the pulse tube smaller by proportionally decreasing the diameters of the tubes and keeping the same length considered only a one-dimensional model for the PTR. Therefore, all the multidimensional effects, for example the radial thermal losses in the regenerator, were neglected. Looking back on the experimental results, we can conclude that scaling the system based on a one-dimensional model for the PTR worked out well. One should take into consideration that the setup is still under development, and with certain optimization, e.g. using different regenerator materials, the temperature of 2.31 K can still be decreased.
EXPERIMENT We have performed a preliminary experiment replacing by In this experiment we reached 2.72 K. This temperature is higher than the lowest temperature achieved with This is due to the fact that the average pressure in the experiment was too low compared to the previous experiments with Therefore, our next step will be to reduce the overall volume of the system by reducing the compressor volume.
SUPERFLUID CRYOCOOLER As the next step in our research work we will study the superfluid cryocooler, which is a combination of a PTR and a superfluid vortex cooler. This cooler is capable of reaching a minimum temperature of 0.65 K, has no moving parts in the cold temperature region, and hardly needs any additional infrastructure. The working fluid for the pulse tube part of the cooler is is the working fluid of the vortex cooler part. The temperature of used in the vortex part, is significantly below the lambda transition (2.17 K). Below this temperature the fluid can be regarded as a mixture of two components, a superfluid component with the density and a normal component with the density The superfluid carries no entropy and has no viscosity, whereas the normal component carries all the entropy of the fluid and behaves as an ordinary viscous fluid. In order to explain the conceptual design of our superfluid cryocooler, we will first treat some theoretical aspects.
Fountain Effect First of all, we will discuss what happens when helium below the lambda point is flowing through a tube. As the normal component carries the entropy and has the viscosity, it will form a
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certain pressure difference over the tube, through which it is flowing. This pressure difference is given by
where is the viscosity of helium, Z is a geometrical factor and sion for the motion of the superfluid component is
is the volume flow. The expres-
where is the chemical potential per mole, M is the molar mass of helium and is the velocity of the superfluid component. We will give a short explanation of the fountain effect. Figure 5 illustrates two reservoirs filled with helium and connected with the superleak. The superleak is just a tube filled tightly with fine powder. The superfluid component has the ability to flow without friction through narrow channels. However the normal component is blocked by its viscosity. Therefore only the superfluid component will be able to flow through the superleak. The temperature in the left reservoir and the pressure are equal to zero. We create a temperature difference between the two reservoirs, applying a certain amount of heat to the right reservoir. The concentration of the superfluid component in the warm reservoir will become smaller than in the cold reservoir. The superfluid component from the cold reservoir will now start flowing through the superleak to the warm reservoir to establish an equilibrium. Under steady state conditions and below the critical velocity the relation (3) can be formulated as where
is the fountain pressure, given by
The essence of the fountain effect is that the superfluid helium has the ability to flow through the superleak in the direction of high temperature and maintains the pressure difference of the superleak. If we calculate the fountain pressure of HeII (He below lambda point), at the lambda temperature we will get a pressure as high as 70 kPa (see Fig. 6).
Vortex cooler The vortex cooler consists of a superleak, followed by a chamber and a capillary. In the beginning all parts of the system have the same temperature and the gradient of the chemical potential is equal to zero. If helium is forced to flow through the superleak, the fluid, leaving the superleak, possesses no entropy. The helium in the capillary is a combination of the normal and the superfluid
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components. If superfluid helium flows through the capillary with supercritical velocity, there are vortices created. These vortices interact with the normal component. The interactions build up a gradient of the chemical potential over the capillary. Due to the created pressure difference over the capillary, the normal component is forced out of the capillary. As a consequence there is now a heat flow from the chamber to the outlet of the capillary. Hence the cooling takes place and a temperature difference is established over the capillary.
The Superfluid Cooler In our research group we combined the fountain pump with the vortex cooler. A schematic diagram of the apparatus is shown on Figure 7. Liquid helium with a temperature of 4.2 K at atmospheric pressure was supplied through the capillary into the bath, the temperature of which can be regulated by means of a pump. By supplying heat to the heater we activate the fountain pump, which consists of the superleak and the capillary Helium flows from the bath through and and arrives at the heat exchanger placed inside the bath with a temperature higher than the bath temperature. It cools down in the heat exchanger and enters the vortex part, having the bath’s temperature and a high pressure, which forces the fluid through the second superleak Through the capillary the fluid returns back to the bath. The coldest point of the setup is between and With such a setup, we managed to reach a temperature as low as 0.76 K with a bath temperature of 1.42 K. An example of the cooling process is represented in Figure 8 as a diagram of the coldest temperature and the temperature in the chamber of the fountain part of the cooler as a function of the heating power As seen from the diagram, one needs less than 7.5 mW of heating power to achieve 1 K.
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Requirements for the Pulse Tube Cooler In our research we are going to combine our three-stage PTR with the superfluid cryocooler to be able to reach temperatures below 0.7 K. We will attach the bath of the superfluid vortex cooler to the cold heat exchanger of the third stage pulse tube. By doing this we avoid using a large helium pump to regulate the temperature of the bath, or so-called base temperature. Now the question is: what performance should the PTR have to be able to operate the vortex cooler? If we assume that the pressure difference over the fountain part of the cooler is equal to the pressure difference over the vortex part of the cooler (see Fig. 9), we can determine the base temperature, i.e. the temperature that the PTR should have to provide the operation of the superfluid cooler. The maximum temperature at which the superfluid cooler can operate is the lambda temperature, 2.17 K. From Figure 6 we see that the fountain pressure corresponding to this temperature is 70 kPa. We now divide this value for the fountain pressure in two, according to the assumption that the pressure difference over the fountain part is the same as the pressure drop over the vortex part, to obtain the fountain pressure at the base temperature. By doing this, we will see that the maximum base temperature should not be higher than 1.92 K. But is the assumption that the pressure drop in the fountain part is equal to the one in the vortex part correct? From Figure 8 we can calculate the pressure drops in both parts of the superfluid cryocooler and plot them as a functions of the applied heating power (see Fig. 10). If our assumption was correct, then the middle line on Figure 10 would be a straight horizontal line. From Figure 10, it is clear that it is not really true. The difference between the pressure drops can be explained as the pressure drops in the capillaries.
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The influence of the base temperature on the coldest temperature of the system can be estimated from Figure 11. A higher bath temperature requires more heating power to be able to achieve the same low temperature. And the heat needed to operate the system is removed at the cold heat exchanger of the pulse tube. Therefore, it is equal to the cooling power of the PTR. The cooling power of the prototype three-stage PTR at a temperature of 1.92 K (our maximum base temperature) was approximately 2.5 mW. That is certainly not enough for our purposes. With our new PTR we intend to reach lower temperatures, e.g. by using new magnetic materials for sub4 K application. So we expect to have enough cooling power to drive the superfluid cooler. An essential difference between our proposed setup and the one described in Figure 7 is that in the new system the overall pressure in the superfluid cooler is not limited to the vapor pressure line. In fact, any pressure up to the melting pressure can be used. Olijhoek5 has shown that the minimum temperature of the vortex cooler decreases with increasing pressure. From 0.75 K at 6 bar it can be decreased to 0.66 K at 24.75 bar.
CONCLUSIONS We have built a new pulse tube refrigerator, which should be able to reach temperatures of 1.8 K and lower. This should be a good starting point to bring the temperature further down by using a superfluid vortex cooler.
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ACKNOWLEDGMENT We would like to acknowledge W.Orbons, J.Oorsprong, L.Habets, M.Verheggen, J.Batenburg, T.Leermakers, J.Duininck and C.Pernot for their contribution to the experimental investigation of the vortex cooler. We also thank P.H. Cappon, L.M.W. Penders and L.C. van Hout for their help to build the setup.
REFERENCES 1.
De Waele, A.T.A.M., Steijaert, P.P., and Gijzen,J., “Thermodynamical aspects of pulse tubes”, Cryogenics, vol. 37, no. 6 (1997), pp.313-324.
2.
De Waele, A.T.A.M., Steijaert, P.P., and Koning, J.J., “Thermodynamical aspects of pulse tubes II”, Cryogenics, vol. 38, no. 3 (1998), pp.329-335.
3.
De Waele, A.T. A.M., Hooijkaas, H.W.G., Steijaert, P.P., and Benschop, A. A.J., “Regenerator dynamics in the harmonic approximation”, Cryogenics, vol. 38, no. 10 (1998), pp.995-1006.
4.
Xu, M.Y., De Waele, A.T.A.M., Ju, Y.L., “A pulse tube refrigerator below 2 K”, Cryogenics, vol. 39, (1999), pp.865-869.
5.
Olijhoek, J.F., “Thermal effects in adiabatic flow of HeII”, PhD thesis, Leiden University, (1973).
Two-Stage Pulse Tube Cryocoolers for 4 K and 10 K Operation C. Wang and P.E. Gifford Cryomech, Inc. Syracuse, New York, USA
ABSTRACT Cryomech has developed large capacity two-stage pulse tube cryocoolers for 4 K and 10 K applications. Both cold heads employ a double-inlet configuration in a compact design. The 4 K model, the PT410, uses rare earth materials in the second stage regenerator. It was improved to provide 1.07W at 4.2 K and 40W at 40.5 K, simultaneously, while consuming 8 kW of electrical power. The 10 K model, the PT810, was developed by modifying the PT410, by using only the lead spheres in the 2nd-stage regenerator. It has 4W at 10 K and 42W at 45 K, simultaneously, and consumes the same electrical power of 8 kW. All the two-stage pulse tubes have been designed to provide a meantime between maintenance > 5 years. INTRODUCTION Some applications of low temperature superconducting devices, such as MRI, NMR, etc. require two-stage cryocoolers to supply large cooling capacity on cooling stations. Presently, a whole body MRI needs cooling capacities of approximately 1 W at 4.2 K with 40 W at 40-45 K simultaneously to eliminate the service of liquid helium. 4 K cryocoolers are in operation either conductively cooling the superconducting magnet or recondensing helium vapor inside the MRI cryostat. 10 K cryocoolers are still being used by some manufacturers to cool the radiation shield(s) of MRI cryostats to reduce the boil off rate of liquid helium. Currently the G-M cryocooler is a workhorse for the cooling of MRI. However, the vibrations, the reliability, and the maintenance interval associated with 4 K GMs have caused problems for the users and manufacturers. These problems can be solved by the use of a pulse tube cryocooler (PTR), because of the lack of the displacer(s). Since the first introduction of a 4 K pulse tube cryocooler from Cryomech in 1999, the advantages of the Pulse Tubes have been approved in many challenging cryo-cooling applications. It is believed that the Pulse Tube will be the next generation refrigeration system for MRI. A few groups1,2,3,4 are working on large capacity two-stage pulse tube cryocoolers for this application. Cryomech had developed a new 4 K pulse tube cryocooler4 in 2001, the Model PT410, with the capacities of 0.83W at 4 K and 38W at 45 K. Recently we have made modifications to improve its performance. In order to meet the requirements of some customers at 10 K, MRI shield cooling and cryopumping, we modified the PT410 for 10 K applications, resulting in the PT810. This paper will introduce the improved two-stage pulse tubes for 4 K and 10 K operations.
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PULSE TUBE REFRIGERATION SYSTEM The two-stage pulse tube cryocooler includes a pulse tube cold head and a helium compressor package. A compressor, Cryomech Model CP980, was designed to drive PT410 and PT810 cold heads. The compressor package supplies the cold head with pressurized helium through flexible metal hoses. A DC stepper motor turns a rotary valve in the cold head to direct the helium gas in and out of the pulse tube system.
1. Pulse Tube Cold Head Figure 1 shows a photograph of the PT410 pulse tube cold head. The PT810 is modified from the PT410, and looks very much the same. The PT410 and PT810 have the same 1st stage regenerator and 1st stage pulse tube, as well as the same warm end parts. The difference between the PT410 and PT810 is that the 2nd stage regenerator and pulse tube of the PT810 is 20 mm shorter than that of PT410. The 2nd stage regenerator is also different. Both the PT410 and PT810 use a double-inlet configuration. Reservoirs, orifices and a rotary valve control the helium flow in the PTR to provide refrigeration. They are integrated inside the motor mount assembly at the warm end, which also performs as the room temperature heat sink for the refrigerator. This configuration and design make the cold head simple and compact and easy to manufacture. The DC stepper motor was chosen to turn the rotary valve because it turns the rotary valve smoothly with low vibration and emits very little EMI. Vibration measurements of our two-stage pulse tube cryocoolers performed by different groups4,5 confirm that the vibrations in the PTR cold heads mainly come from the stretching of the tubes generated by gas compression and expansion.
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Figure 3 shows the packing of the 2nd stage regenerators of PT410 and PT810. The 2nd stage regenerator of the PT410 is packed with three layer materials of Pb, (or Nd) and HoCu2 (Figure 3 (a)). The 4 K Pulse Tube provides the same cooling performance using or Nd in the middle layer. The 2nd stage regenerator of PT810 is filled with only lead spheres (Figure 3 (b)).
2. Compressor Package, CP980 The CP980 compressor package employs a helium scroll compressor made by Copeland. It is the largest of the CP900 Series compressor packages manufactured at Cryomech based on the Copeland scroll, which was designed especially for cryogenic applications. The CP950, CP970 and CP980 all have the same dimensions and components, except for the compressor module. Two important features of the CP900 compressor packages are the high efficiency oil separator and adsorber, which have been designed for an adsorber replacement interval of >5 years. During the early manufacture of the CP950 compressor package, a few oil separators failed. There was too much oil reaching the adsorber. First we developed an improved oil carryover test, which is now part of our standard manufacturing procedures. The oil carryover stems from the fact that the scroll compressor module makes finer oil aerosols than a piston type compressor. To improve the removal of these aerosols, we redesigned and enlarged the oil separator to more efficiently agglomerate the oil out of the helium stream to reduce the amount of oil that reaches the adsorber. We also built into our quality control process the ability to monitor the oil carryover rate before the adsorber. This measurement is taken during the operation of the complete cryorefrigerator, throughout several cool downs and testing. Each system is monitored for a week. Figure 4 shows measured oil carryover rates in some of the CP900 series compressor from our production line. Approximately 96% of the CP900s have oil carryover of less than 35 mg/day, which indicates that a yearly carryover rate of <12 g/year oil is trapped in the adsorber. Only 4% of the CP900s have an oil carryover rate around 70 mg/day (26 g/year). In a test of an in house pulse tube for few months, a total of 120 g oil passed
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through the oil separator and into the adsorber. We found there is no oil passing through the adsorber and no contamination in the cold head. Presently we are testing the adsorber for maximum oil adsorption. In a test made by R. Longthworth6 , it was demonstrated that the test adsorber could adsorb 450 g of oil without passing the oil on to the cold head. Once a maximum amount was adsorbed it would pass the oil on to the cold head. The statistic measurements and adsorber performance give us confidence that we will be able to recommend >5 years meantime between maintenance (MTBM) of the compressor. Currently we request that the adsorber is replaced after 20,000 hours and returned to Cryomech for statistical evidence to support the extended maintenance interval. The rotary valve system in the PT410 cold head has been redesigned for an expected lifetime of >5 years. Without the moving displacers, the PTR does not generate wear particles that travel to the rotary valve and valve plate causing failure. We have attempted to measure the wear of the rotary valve and valve plate and have not been able to measure noticeable wear on the systems that we have run in and tested for up to several thousand hours. We have set up a procedure and process for the purification of the helium gas in the whole system. This process was verified to efficiently decrease the gas contamination to a very low level, which ensures the long-term running of the pulse tube cryocoolers. The Cryomech goal is to supply pulse tube cryocoolers to have MTBM >5 year. The experience with the PT405 pulse tube, which was first delivered in July 1999, has supported our goal. We have not had any requirements for routing service of the PT405 during that time for oil carry over, rotary valve failure, or helium purity issues.
PERFORMANCE OF PT410 The first prototype of the PT410 was tested in May, 2001. After minor modifications to a few of the components, we began production. Table 1 shows the performance of the prototype of the PT410 and four PT410s from the production line. All cold heads were driven by CP980 compressor packages operating at 60 Hz. They demonstrate the normal consistency in capacity normal to the PTRs.
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Since the cooling capacities in Table 1 are marginal for MRI application, we have continued our R&D effort on the PT410. Recently a modification has been made on the PT410, which significantly improved the 1st and 2nd stage performance. Figure 5 shows the cool down curves of the PT410. It takes 44 minutes for the 2nd stage to reach the 4.2 K and 50 minutes to reach the bottom temperature of 2.47 K. The lst stage reaches the bottom temperature of 30.5 K in 110 minutes. Figure 6 gives the cooling load map of the PT410. It provides
[email protected] on the 2nd stage and
[email protected] on the 1st stage simultaneously; or provides
[email protected] and 52W@45K on the 1st stage simultaneously. The input power of it is 8.0kW at 4.2K and 40K on both stages and 8.1kW for 4.2K and 45K cold station temperatures. The cooling capacities of PT410 are also measured from the bottom temperatures to room temperature, which are shown in Figure 7. Two big heaters are mounted on the 1st and 2nd cooling station of the PT410 to supply enough heat load. This curve allows the users of PT410 to estimate the cooling speed of their attachments to the cooling stations on the PT410.
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PERFORMANCE OF PT810 Figure 8 shows the cool-down characteristics of PT810. It takes 50 minutes for the 2nd stage to reach the bottom temperature of 5.9 K, and 65 minutes for the 1st stage to reach the bottom temperature of 32.3 K. The 1st stage of the PT810 has a faster cooling speed than that of the PT410. The cooling load map of the PT810 is given in Figure 9. It can provide 4W@10K on the 2nd stage and 42W@45K on the 1st stage, where it consumes 8.0 kW electrical powers. This data is from a preliminary test, and there is still considerable potential to optimize the PT810 to improve its efficiency. Figure 10 shows long-term operation of PT810. For this test the temperature controller maintained the 1st stage temperature at 45 K by adding heat. The recorded heat load varied from 46 to 47 W at 45 K. The PT810 has the same temperature stability characteristics as the PT410s. Since the PT810 has the same rotary valve system as the PT410, we expect the same MTBM of >5 years. We performed another test on the PT810 to discover the lowest possible attainable temperature with lead in the 2nd stage regenerator. By changing the flow impedances in the PT810, a lowest temperature of 4.6 K was reached, which is shown in Figure 11. A lower temperature for the PT810 could not be obtained because of poor efficiency of the 2nd regenerator around 4.6 K. One should be able to reach a temperature below 4.2 K by using a larger lead regenerator, which should have higher regenerator efficiency. A temperature of 4.23 K was once obtained by a GM/PT hybrid cryocooler7. These tests were for the sake of scientific curiosity. The efficiency of the PTR cryocooler with lead regenerator at 4 K is much lower than that with rare earth regenerative materials.
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CONCLUSION Two efficient and large capacity two-stage pulse tube cryocoolers have been developed at Cryomech for 4 K and 10 K operations. The 4 K model, PT410, was improved to provide 1.07W at 4.2 K and 40W at 40.5 K, simultaneously, for 8.0 kW power input. The 10 K model, PT810, provides 4W at 10 K and 42W at 45 K, simultaneously, also for 8.0 kW of input power. The two-stage pulse tubes developed at Cryomech have expected MTBM >5 years.
ACKNOWLEDGMENT Authors would like to thank R. Dausman and J. P. Cosco for collecting oil carryover rates in the compressor packages.
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REFERENCES 1. Siegel, A. and Haefner, H.U., “Investigation to the Long-Term Operation Behavior of GM-Pulse Tube Cryocoolers,” Advances in Cryogenic Engineering, Vol. 47B, AIP, Melville, New York (2002), pp. 903-910. 2. Gao, J., “IGC-APD Advanced Two-Stage Pulse Tube Cryocoolers,” Advances in Cryogenic Engineering, Vol. 47B, AIP, Melville, New York (2002), pp. 683-690. 3. Zhu, S., Ichikawa, M., Nogawa, M., and Inoue, T., “4 K Pulse Tube Refrigerator and Excess Cooling Power,” Advances in Cryogenic Engineering, Vol. 47B, AIP, Melville, New York (2002), pp. 633-640. 4. Wang, C. and Gifford, P.E., “Development of 4 K Pulse Tube Cryocoolers at Cryomech,” Advances in Cryogenic Engineering, Vol. 47B, AIP, Melville, New York (2002), pp. 641-648. 5. Wang, C. and Gifford, P.E., “Performance Characteristics of a 4 K Pulse Tube in Current Applications,” Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2001), pp. 803-808. 6. Longsworth, R.C., “Helium Compressor for GM and Pulse Tube Expanders,” Advances in Cryogenic Engineering, Vol. 47B, AIP, Melville, New York (2002), pp. 611-698. 7. von Schneidemesser, Almut., “Untersuchungen zum Betriebsverhalten einer 4 K-Pulsrohrstufe mit Seltenerd- und Blei-Regenerator,” Ph.D. Dissertation, University of Giessen, May, 2000.
Development of a 4K Two-stage Pulse Tube Cryocooler M.Y. Xu, P.D. Yan, T.Koyama, T.Ogura, R. Li Cryogenics Department, Sumitomo Heavy Industries, Ltd. 2-1-1, Yato-cho Nishi-tokyo-city, Tokyo 188-8585 Japan
ABSTRACT Sumitomo Heavy Industries, Ltd. (SHI) has been developing 4K pulse tube cryocoolers for cooling MRI Magnet Systems, Small Superconducting Magnets, SQUIDs, X-ray detectors, etc. The valve unit of the cryocoolers is separated from the cold head by a self-sealed coupling. With this configuration, the maintenance for the valve unit becomes much easier and faster than that with a unified one since it is not necessary to warm up the cold head and cool down again. The vibration from the compressor and valve unit can be reduced by this configuration with some simple techniques. A typical cooling capacity is of 0.9W at 4.23 K at the second stage and 40W at 45.9 K at the first stage, simultaneously. The results of the influence on the performance by filling pressure, operating frequency are reported. The vibration of a SHI 4K pulse tube cryocooler is measured with a laser displacement sensor.
INTRODUCTION Many efforts have been made worldwide in developing pulse tube cryocoolers in place of conventional Stirling and G-M cryocoolers since 1984 [1-6]. It is possible to improve reliability, to increase the meantime between maintenance and to extend operating life since there is no moving part at the low temperature region of a pulse tube cryocooler. Since Matsubara, et al. [4] reached a temperature below 4K with a pulse tube cryocooler in 1994, the performance of 4K pulse tube cryocoolers has been improved significantly [4-9]. The 4K pulse tube cryocooler has been a competitor of G-M cryocooler for cooling MRI magnet systems, small superconducting magnets, SQUIDs, X-ray detectors, etc. Since 1999, SHI has been developing 4K pulse tube cryocoolers for the above described applications. The development in SHI on pulse tube cryocoolers has focused primarily on cooling capacity of both the first and second stage, simplification, suitable configuration for mass production and maintenance, as well as long life.
SYSTEM DESCRIPTION Figure 1 shows the system of SHI’s 4K two-stage pulse tube cryocooler. A photo of the cold head and a photo which shows the situation of the cold head connecting with the valve unit are shown in Figure 2 and Figure 3, respectively. As shown in Figure 1, the cold head is separated from the valve unit by a self-sealed coupling. With this kind of configuration, it is possible to decrease Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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the vibration generated from the valve unit and compressor to the cold head with some simple techniques, such as longer connecting tube, vibration restrainer, etc. And also, it is possible to replace the valve unit without warming up the cold head. It takes much less time to replace the valve unit with this kind of configuration than with a unified one. A modified SHI CSW-71C compressor and a valve unit generate the pressure oscillation. The cold head consists of reservoirs, regenerators, pulse tubes and heat exchangers. The first stage regenerator is filled with stainless steel screens disks and lead spheres. The second stage regenerator is filled with the spheres of lead and magnetic material. The two pulse tubes are made of stainless steel tubes. At both ends of the pulse tubes, copper screens and perforated plates are used for flow straighteners and heat exchangers.
PERFORMANCE Load Map The performance of a 4K pulse tube cryocooler can be affected by several factors, such as operating frequency of the compressor and the cold head, the filling pressure, the dimensions of regenerators and pulse tubes, etc. Figure 4 shows a typical load map of a SHI’s 4K two-stage pulse tube
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cryocooler. As shown in Figure 4 the cooling capacity is 0.9W at 4.23 K at the second stage and 40W at 45.9 K at the first stage, simultaneously. Without mention, in the following parts, the compressor is operated at 60 Hz and the cold head is operated at 1.2 Hz. The filling pressure is 1.9 MPa. The flow impedance is optimized for larger cooling capacity of both the first and second stage. The input power of the compressor is about 8.5 kW when the heat load added to the first stage is 40W and the heat load added to the second stage is 0.9W, simultaneously. The temperature of the first stage is measured with a Pt-Co thermometer (Chino Corporation) and the temperature of the second stage is measured with a Germanium thermometer (Scientific Instruments, Inc.). The heat loads of both the first and the second stages are calculated from the voltage and current through the Manganin wire wrapped on the cold stages.
Cool Down Curves Figure 5 shows the cool down curves of a SHI 4K pulse tube cryocooler. The filling pressure is 1 .85MPa. For the second stage, it takes about 77 minutes to reach 4K and 85 minutes to reach the minimum temperature of 2.4K. For the first stage, it takes about 84 minutes to reach 40 K and 145 minutes to reach the minimum temperature of 26.3 K. It is due to the larger copper block of the first stage.
Cooling Capacity of the Second Stage The change of the temperature at the second and first stage when the heat load to the second stage increases is shown in Figure 6. The filling pressure is 1.85MPa. The heat load to the first stage is kept at 30W. The cooling capacity of the second stage is 1.0W at 4.4 K and 4W at 11 K. The cooling capacity of the second stage is almost in proportion to the temperature of the second stage. The first stage temperature with 30W heat load also goes up as the second stage heat load increases. It is due to the loss of the second regenerator becomes larger when the second stage temperature is higher.
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However, the increasing tendency slowed down when the second heat load is larger than 3 W. It is still necessary to investigate the reason.
The Effect of Filling Pressure The effect of the filling pressure on the first and second stage pulse tube is shown on Figure 7. The flow impedance is optimized at the operating frequency of 60Hz/1.2Hz and the filling pressure of 1.9MPa. From Figure 7, it shows that the temperature of the first stage goes up as the filling pressure decreases. The temperature of the second stage also goes up as the filling pressure decreases. As well known, the mass flow into the cold head decreases when the filling pressure decreases. It is shown that from the experiment data, the cooling capacity of the first stage is sensitive to the mass flow rate or the filling pressure. Comparing to that of the first stage, the cooling capacity of the second stage is less sensitive to the filling pressure. However, if the filling pressure is too low, for example, 1.6MPa, the temperature of the second stage also goes up a great degree.
The Effect of Frequency From Figure 7, it is also seen that even with the same filling pressure, the cooling capacity of the first stage is quite different from the operating frequency of 60Hz/1.2Hz to 50Hz/1.0Hz. Naturally, it may be considered that the flow impedance is optimized for the operating frequency of 60Hz/1.2Hz rather than that of 50Hz/1.0Hz. So the following investigation has been done with the same system and the same filling pressure of 1.85MPa as shown in Figure 8. At first, the flow impedance is optimized for 60Hz/1.2Hz. The cooling capacity is 0.9W at 4.3K at the second stage and 30W at 42.4K at the first stage. Then the frequency is changed to 50Hz /1.0Hz. Without changing the flow impedance, the cooling capacity is 0.9W at 4.21K at the second stage and 30W at 52.4K at the first stage. The cooling capacity of the first stage decreases greatly. However, the cooling capacity at the second stage increases to some degree. Finally, the flow impedance is readjusted for 50Hz/1.0Hz. The cooling capacity is 0.9W at 4.22K at the second stage and 30W at 50.2K at the first stage. The cooling capacity of the first stage increases in some degree, but it is still much less than that of 60Hz/1.2Hz. From the
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above described experiment, it is shown that the cooling power of the first stage is sensitive to the mass flow rate rather than the flow impedance—since the input power and mass flow decreases almost in proportion to the operating frequency.
VIBRATION MEASUREMENT Figure 9 shows the test rig to measure the vibration on the cold stage of the pulse tube cryocooler at room temperature. A laser displacement sensor (LDS) measured the displacement of the cold stage. The cold head is operated at 1.0Hz. Figure 10 shows the displacements of the 2nd cold stage. The axes defining the vertical direction are given in Figure 9. The maximum amplitude is around 23µm in the vertical direction. Fig-
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ure 10 also shows the spectrum of the vibration. It shows that the maximum displacement of the vibration is around 1.0 Hz, the operating frequency of the cold head. And also, most of the higher orders of the vibration are under 10 Hz. It means that it is easier to restrain the vibration of a pulse tube cryocooler since there is almost no vibration in the high frequency region.
CONCLUSIONS SHI has been developing 4K pulse tube cryocoolers for MRI and other applications. The valve unit of the cryocooler is separated from the cold head by a self-sealed coupling. The configuration has great advantages for maintenance and reduction of vibration. With about 8.5kW input power, a typical cooling capacity is 0.9W at 4.23K at the second stage and 40W at 45.9K at the first stage, simultaneously. The relationship between the performance and filling pressure, operating frequency, etc has also been investigated by experiments. It is shown that the cooling power of the first stage is sensitive to the filling pressure and/or the mass flow rate rather than the flow impedance. The
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vibration of a SHI 4K pulse tube cryocooler is measured. The maximum amplitude is around 23µm in the vertical direction. The maximum displacement of the vibration is around 1.0 Hz, the operating frequency of the cold head. REFERENCES 1. Mikulin, E.I., Tarasow, A.A., Shkrebyonock, M.P., “Low temperature expansion pulse tube,” Advances in Cryogenic Engineering, vol. 29, Kluwer Academic/Plenum Publishers, New York (1984), p. 629. 2. Radebaugh, R., “Development of the pulse tube refrigerator as an efficient and reliable cryocooler,” Proc. Inst. of Refrig. (London), 1999-2000. 3. Zhu, S. and Wu, P., “Double inlet pulse tube refrigerators: an important improvement,” Cryogenics, vol.30 (1990), p.514. 4. Matsubara, Y. and Gao, J.L., “Novel configuration of three-stage pulse tube refrigerator for temperature below 4K,” Cryogenics, vol.34, no.4 (1994), pp. 256-262. 5. Wang, C. Thummes, G., and Heiden, C., “A two-stage pulse tube cooler operating below 4K”, Cryogenics, vol. 37, no.3 (1997), pp. 159-164. 6. Xu, M.Y., De Waele, A.T.A.M. and Ju, Y.L., “A pulse tube refrigerator below 2K,” Cryogenics, vol.39 (1999), pp. 865-869. 7. Wang, C. and Gifford, P.E., “0.5 W class two-stage 4K pulse tube cryorefrigerator,” Advances in Cryogenic Engineering, vol. 45A, Kluwer Academic/Plenum Publishers, New York (1984), pp. 1-7. 8. Gao, J.L., “IGC-APD advanced two-stage pulse tube cryocoolers”, Adv. in Cryogenic Engineering, Vol. 47, Amer. Institute of Physics, Melville, NY (2002). 9. Zhu, S., Ichikawa, M., Nogawa, M. and Inoue, T., “4K pulse tube refrigerator and excess cooling power,” Adv. in Cryogenic Engineering, Vol. 47, Amer. Institute of Physics, Melville, NY (2002).
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Performance of a 4K Pulse Tube Refrigerator and Its Improvement S.W. Zhu, M. Nogawa, S. Katsuragawa, M. Ichikawa, T. Inoue Second Development Department, Aisin Seiki Co., Ltd. 2-1, Asahi-Machi, Kariya, Aichi 448-8650, Japan
ABSTRACT A pre-cooling two-stage 4K pulse tube refrigerator has been developed for high cooling power. A test machine was first constructed to allow efficient optimization of parameters such as the regenerator and pulse tubes volumes and selection of the regenerator material. The first stage was separately tested before the second stage was installed. About 65 W at 60K was achieved for the first stage—about 9% of Carnot. A new regenerative material (GAP) was also tested by installing it in the bottom of the second stage regenerator in the order an optimum weight ratio of 20-25% of was observed. The improvement of the cooling power associated with the use of GAP was over 20%. As a preliminary result, over 1W of cooling at 4.2K and 30W of cooling power around 45K were achieved with an input power of approximately 7.6kW and a connecting tube length of 1.5 meters. To increase the first stage cooling power, a double-compressor phase shifter was introduced. Its operation is explained, The equivalent PV diagrams, which are derived from numerical simulations, show that it has high potential for increasing efficiency at low temperatures.
INTRODUCTION A 4K pulse tube refrigerator1-7 is a strong potential candidate for cooling MRI, NMR, SQUID and other superconducting instruments due to its low vibration and easy maintenance. There are two kinds of two-stage 4K pulse tube refrigerators. One is the “conventional type” in which the first regenerator is connected to both the second regenerator and the first pulse tube. Another is the “pre-cooling type” in which the first stage and second stage are two independent pulse tube refrigerators thermally connected by a thermal link at the first stage. The authors have developed their 4K pulse tube refrigerator based on the pre-cooling type.3 There are a number of parameters that must be optimized in the pre-cooling type pulse tube. A test machine for optimizing parameters and testing a new phase shifter concept for increasing efficiency and cooling power has been designed and manufactured. In this paper, some preliminary tests and trials of the test machine are described. In particular, some test results are shown for a new regenerative material GAP8, which has been recently proposed as an effective material in the 4K region. For getting higher cooling power for the first stage, a new type of phase shifter (the “double-compressor phase shifter”) is also proposed.
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FEATURES OF THE PRE-COOLING TWO-STAGE PULSE TUBE REFRIGERATOR Figure 1a shows the pre-cooling two-stage pulse tube refrigerator. R11 is the first first-stage regenerator. PT1 is the first stage pulse tube. R21 is the second first-stage regenerator. R22 is the second stage regenerator. PT2 is the second stage pulse tube. A heat link connects each of the first stages thermally. Between the rotary valve unit and the cold head, there are 1.5-meter-long connecting tubes to decrease vibration and simplify integration and use. A detailed explanation of the structure is presented elsewhere.3 For comparison, a conventional two-stage pulse tube refrigerator is shown in Fig. 1b. V1 and V2 are high and low pressure valves, respectively. R1 is the first stage regenerator. R2 is the second stage regenerator. PT1 is the first stage pulse tube. PT2 is the second stage pulse tube. B1, N11, N12, are the buffer tank, orifice, and bypass valve of the first stage, respectively. B2, N21, N22 are the buffer tank, the orifice, and the bypass valve of the second stage, respectively. As shown in Fig. 1a, the pre-cooling two-stage pulse tube refrigerator consists of two independent single pulse tube refrigerators. Since the the performance of the 40-60K single stage pulse tube is well known, the development of the two-stage pre-cooling type could be focused mainly on the second stage. The reliability of the pre-cooling type may also be higher than the conventional type because the first stage refrigerator already has about 10 years manufacturing history and is well proven. The two refrigerator halves can be run phase shifted by 180 degrees to smooth out the mass flow rate of the compressor. This increases the efficiency of the compressor by decreasing the pressure oscillation in the high and low pressure lines of the compressor. About half of the compressor gas flows to the first stage, and half flows to the second stage. The valve opening area can be half compared to a conventional type. The structure of the pre-cooling type two-stage pulse tube refrigerator is rather complex, and there is a thermal connecting loss. The thermal connecting loss includes two parts: one part is the heat conduction loss of the thermal link; the other is the heat transfer loss between the thermal link and the gas at the hot end of the second stage regenerator. Therefore the design of the heat exchanger at the first stage and the thermal link are key points for the pre-cooling type.
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TEST MACHINE A schematic of the test machine is shown in Fig. 1 a. Each part of the test machine is connected by flanges that can be changed without changing other parts. They are for optimizing parameters and confirming the effect of the new phase shifter. The input power of the compressor is limited to less than 8kW. The connecting tubes between the rotary valve and the coldhead are fixed to 1.5m long. Compared to our earlier work3, the first stage and the second stage are enlarged. The rotary valve is also redesigned. At the beginning, only the first stage cooler was installed. Figure 2 shows the cooling power and efficiency vs. refrigeration temperature. A cooling power of about 65 W at 60K with about 9% Carnot efficiency was achieved. This confirmed that the first stage was successfully manufactured. After the test of the first stage, the regenerator material (GAP), which has a high heat capacity peak around 3-4K, was packed in the bottom of the second stage regenerator in the order Figure 3 shows the cooling power at 4.2K vs. the weight ratio of GAP to with a first stage heat load of 30W. The optimum ratio of is
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about 20-25%. Without GAP, the cooling power is about 0.82W. Near the optimum ratio, the cooling power is over 1W at 4.2K. The cooling power is increased over 20%. Figure 4 is a cooling power map with GAP used in the regenerator. The compressor input power is about 7.6kW. Compared to our earlier work3, the cooling power of the first stage and the second stage is increased significantly. As a preliminary result, this also confirms that the design and manufacture of the test machine was successful. It is now ready to be used with altered parameters in a wide program of optimization.
PROPOSAL OF DOUBLE COMPRESSOR PHASE SHIFTER In Figure 1, the two middle-buffer phase shifter is used in the first stage. Figure 5 shows another phase shifting method. A small compressor is connected at the hot end of the first stage
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pulse tube with a buffer connected through on/off valves. We call this a double-compressor phase shifter. It is similar to the 5-valve design.9 The DC gas flow is not introduced perfectly. A typical valve opening process is shown in Fig. 6. Valve V15 is opened until the pressure in PT1 is increased to near the buffer pressure, then closed. Valve V13 is opened. Valve V11 is opened at the same time or a little later. When the pressure in PT1 is near the high pressure of the main compressor, valve V13 is closed, and valve V15 is opened with small opening area to let gas from regenerator R11 flow to the cold end of PT2. After sufficient gas flows into the cold end of PT2, valves V11 and V15 are closed. Valve V15 is opened until the pressure in PT1 is decreased to near the buffer pressure, then closed. Valve V14 is opened. V12 is opened at the same time or a little later. When the pressure in PT1 is near the low pressure of the main compressor, valve V14 is closed, and valve V15 is opened with small opening area to let the gas at the cold end of PT2 flow to R11. After the cold gas flows out of PT2, valves V12 and V15 are closed. This completes one cycle. Figure 7 shows the typical equivalent PV diagrams at both ends of the pulse tube. They are from a nodal analysis numerical program with ideal gas properties. For comparison, the PV diagrams of the two-middle buffer type are also shown in the figure. The bold line is the PV diagram of the double compressor, and the thin line is that of the two-middle buffer. The shape of the PV
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diagram at the cold end of the pulse tube of the double-compressor type is similar to that of the idealized Solvay cycle. Compared to that of the two-middle buffer type, the area of the PV diagram of the double-compressor type is larger. High cooling power and high efficiency could be predicted. The gas displacement at the cold end of the pulse tube can be controlled by the opening area of valve VI5. This is very important for getting high cooling power at low temperatures. At low temperatures, small gas displacement at the cold end of the pulse tube means small mass flow rate and high regenerator efficiency. The input power to the small compressor provides no direct refrigeration effect; it is a necessary loss for getting high cooling power. There are irreversible losses in the valves at the hot end of the pulse tube. Part of this irreversible loss is balanced by the small compressor, and part is balanced by the expansion work at the cold end of the pulse tube. With decreasing refrigeration temperature, the expansion work at the cold end of the pulse tube becomes smaller. More input power is required from the small compressor. This method is also effective for the second stage, though the system becomes more complex. The double-compressor pulse tube refrigerator can also be operated with the following valve opening and closing timing. Valve V15 is opened then closed. Then valve V11 and valve V13 are opened then closed. Valve V15 is opened then closed. Then valves V12 and V14 are opened then closed. This finishes one cycle. With this alternative valve opening and closing timing, the input power of the small compressor is decreased because part of the expansion work at the cold end of the pulse tube is saved in the dead volume of the compressor and is used in the next cycle. The shape of the equivalent PV diagram is the same with that in Fig. 7. In this condition, the gas displacement is controlled by the input power to the small compressor. A simple analysis shows that the double-compressor type has a high potential for the low temperature range, though it is rather complex.
CONCLUSIONS A pre-cooling two-stage 4K pulse tube refrigerator with high cooling power has been developed. A test machine for optimizing parameters and confirming a new phase shifter was manufactured. As preliminary test results, over 1W cooling power at 4.2K and around 30W cooling at 45K were acheived using a new regenerative material GAP with an input power of 7.6 kW. A double-compressor phase shifter has also been proposed for the first stage. Its high potential for increasing the cooling power is predicted based on numerical calculations.
ACKNOWLEDGMENT Thanks to Dr. Mumazawa of Tsukuba Magnet Laboratory, NIMS, Japan, for supplying the regenerator material GAP for the test.
REFERENCES 1.
Thompson, P.S., Ackermann, R.A., and Hedeen, R.A., “A Two-Stage Pulse Tube Cryocooled MRI Magnet,” to be published in Advances in Cryogenic Engineering.
2.
Gao, J.L., “IGC-APD Advanced Two-stage Pulse Tube Cryocooler,” to be published in Advances in Cryogenic Engineering.
3.
Zhu, S.W., Ichikawa, M., Nogawa, M., and Inoue, T., “Two-stage 4 K Pulse Tube Refrigerator,” Cryocooler11, Kluwer Academic/Plenum Publishers, New York (2001), pp. 243-247.
4.
Zhou, S.L., Thummes, G., and Matsubara, Y., “ Experimental Investigation of Loss Mechanism in a 4 K Pulse Tube,” Advances in Cryogenic Engineering, Vol. 45, Kluwer Academic/Plenum Publishers, New York (2000), pp. 81-88.
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5.
Xu, M.Y., De Waele, A.T.A.M., and Ju, Y.L., “A Pulse Tube Refrigerator Below 2K,” Cryogenics, Vol.39 (1999), pp. 865-869.
6.
Wang, C., Thummes, G., and Heiden, C., “Performance Study on a Two-stage 4 K Pulse Tube Refrigerator,” Adv. in Cryogenic Engineering, Vol. 43, Plenum Publishers, New York (1998), pp. 20552062.
7.
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, No.5 (1997), pp. 271-273.
8.
Numazawa, T., Arai, O., Sato, A., Fujimoto, S., Ooda, T., Kang, Y.M., and Yanagitani, T., “New Regenerator Material for Sub-4K Cryocooler,” Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York ( 2001), pp. 465-473.
9.
Yuan, J., and Pfotenhauer, J.M., “A Single Stage Five Valve Pulse Tube Refrigerator Reaching 32K,” Advances in Cryogenic Engineering, Vol. 43, Plenum Publishers, New York (1998), pp. 1983-1988.
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Experimental Investigation of a G-M Type Coaxial Pulse Tube Cryocooler K. Yuan, J.T. Liang, Y.L. Ju Cryogenic laboratory, Technical Institute of Physics and Chemistry Chinese Academy of Sciences, Beijing, P.R. China
ABSTRACT A G-M type coaxial pulse tube cryocooler is first optimized from three aspects: the dimensional layout of the pulse tube and regenerator, the filling materials in the regenerator, and the structure of the cold end heat exchanger. Three different types of phase shifters: needle valves, capillaries and asymmetry-nozzles are then employed at the hot end of the pulse tube for the adjustment of the phase between the gas mass flow and the pressure oscillations. The influences of DC flow are investigated experimentally. It is found that a proper positive DC flow has positive effects on the cooling performance of the coaxial pulse tube cryocooler.
INTRODUCTION Due to the absence of mechanical moving components in the low temperature region, pulse tube cryocoolers (PTCs) have inherent merits in terms of mechanical simplicity, high reliability, low mechanical vibration, and low cost.1 This cryogenic cryocooler has been used for commercial applications such as cooling of infrared devices and sensors, superconducting electronic devices, etc. Among three configurations of the PTC, the coaxial structure is the most compact and convenient for actual applications. The performance of the PTC has been greatly increased by improvements of different phase shifters located at the hot end of the PTC.2- 4 For a coaxial PTC, the phase shifters and the coupling heat transfer between the regenerator and the pulse tube, which is introduced by the mismatched temperature profile along the axial-direction of the regenerator and pulse tube, affect the performance of the cryocooler. In this paper, a G-M type coaxial pulse tube cryocooler is first optimized from three aspects. Then, the effects of three different phase shifters and the coupling heat transfer are investigated experimentally by analyzing the external wall temperature profile of the regenerator. It is found that a proper positive DC flow is benefitial for increasing the cooling performance of the coaxial pulse tube cryocooler.
EXPERIMENTAL SETUP A schematic diagram of the experimental apparatus is shown in Fig. 1. The input power of the compressor is about 1.1 kW. The operating frequency of the rotary valve can be adjusted
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from 1 to 10 HZ. The pulse tube is made of nylon with a wall thickness of 1.2mm and is placed inside the annular stainless steel screens in the regenerator. Three different types of phase shifters, a needle valve, a capillary and an asymmetry-nozzle, are used at the hot end of the pulse tube. For the asymmetrical flow resistance devices such as the needle valve and the asymmetry-nozzle, the flow resistances in the positive and negative directions are different. In this paper, the asymmetrical flow resistance devices are defined as used in the normal direction, as shown in Fig. 2, when the lower resistance is for gas flow directed from the inlet of the regenerator to the hot end of the pulse tube via the double inlet valve; the opposite orientation is defined as the reversed direction.
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Four copper-constantan thermocouples (Type T) are arranged along the external wall of the regenerator. The temperature of the cold tip is measured by a Pt-100 resistant thermometer. Three small quartz differential pressure transducers (KISTLER, Type 601 A), connected to a charge amplifier (KISTLER, Type 5011) having a high natural frequency (150 kHz), are used to measure the transient pressure wave. Another pressure transducer is mounted on the reservoir to measure the charge pressure, as shown in Fig. 1. The cooling power of the PTC is measured by a DC power supply (HP 6634A).
EXPERIMENTAL RESULTS Optimization of the PTC First, the PTC is optimized from three aspects: the dimension of the cryocooler, the filling material in the regenerator, and the structure of the heat exchanger at the cold end. Figure 3 compares the cooling performance of the PTC before and after the optimization. It is clear that both the lowest temperature and the cooling power are improved remarkably. The PTC can reach a no-load temperature below 30K and has a cooling capacity of about 6.0W at 75K. These experiments and all the following experiments are carried out with a charge pressure of 1.45 MPa.
Results with a Needle Valve The orifice valve is used in the normal direction and adjusted to optimum opening (0.89 turns). The double inlet valve is used in the normal direction and adjusted to optimum opening (3.0 turns) first, and then the double inlet valve is reversed with the previous opening. Figure 4 compares the lowest temperatures of the PTC with the double inlet valve used in both directions under different operating frequencies. It is found that the double inlet valve used in the normal direction leads to lower temperatures. The rate of temperature rise per unit cooling power is about the same.
Results with Capillaries The capillaries, which are symmetrical flow resistance devices, consist of a stainless steel tube with an inner diameter of 0.6mm and an outer diameter of 1.0mm. First, the capillary is used as the orifice; we change the length of the capillary to find the optimum length for the orifice pulse tube cryocooler (OPTC) with the double inlet valve closed. Figure 5 gives the relationship between the lowest temperatures of the OPTC and the operating frequency with different capillaries. The results with a needle valve are also shown in Fig. 5 (the curve marked as valve). For the OPTC, the optimum length of the capillary is 14.7cm. Next, the capillary is used as a double inlet with the
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valve at the orifice kept at its optimum opening. With the same procedure, we found that the optimum length of the capillary for the double inlet is 5.1cm, which is illustrated in Fig. 6. The results with the needle valve are also shown in Fig. 6 for comparison (the curve marked as normal and reversed).
Results with the Asymmetry-nozzles Two asymmetry-nozzles were used as the double inlet, with nozzle 1 more asymmetrical than nozzle 2. Figure 7 shows the relationship between the lowest temperatures of the PTC and the operating frequency using the two asymmetry-nozzles. It is found that the asymmetry-nozzles used in the reversed direction lead to a lower temperature. The difference of the temperature under the two directions varies from 0.5K to 6.2K.
External Wall Temperature Profile of the Regenerator The external wall temperature profile of the regenerator with different phase shifters at the hot end is shown in Fig. 8, where the x-coordinate is the ratio of the distance from the thermometer to the hot end of the regenerator (x) to the regenerator length (L), Here, the temperature at the cold end of the regenerator is considered the same as the temperature of the cold tip. The temperature profiles of the regenerator and pulse tube are generally considered as an indirect indication of the DC flow direction in the double inlet pulse tube cryocooler (DPTC).5 Due to
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the difficulties of measuring the temperature profile along the pulse tube in a coaxial PTC, we only measure the external wall temperature profile of the regenerator, which is controlled by both the DC flow and the coupling heat transfer. Fortunately, we can analyze the DC flow by comparing the regenerator temperature profile of the DPTC with that of the OPTC, which has no DC flow component. In the present study, DC flow is defined as positive when directed from the hot end of the pulse tube to the inlet of the regenerator via the double inlet valve, and as negative for the opposite direction (see Fig. 2). We find in Fig. 8 that the external wall temperature profiles of the regenerator are nonlinear for the OPTC due to the coupling heat transfer. Compared with the temperature profiles of the OPTC, the temperature profiles with valve, which are far above that of the OPTC, indicate a positive DC flow in the PTC. Similarly, the temperature profile with the 5.1cm capillary indicates a smaller positive DC flow. Because the temperature profiles of the asymmetry-nozzle are only slightly below the temperature profiles of the OPTC, the DC flow component with the nozzle is very small, although its direction is uncertain.
DISCUSSION Control of the DC Flow Using the asymmetrical flow resistance devices is a practical way to control the DC flow in the DPTC. The performance of the PTC varies remarkably by using the asymmetrical flow resistance devices under different directions as shown in Figs. 4 and 7. We divide the DC flow into two parts: one part is generated by the intrinsic asymmetry of the PTC, and the other part is introduced by the asymmetry of the flow resistance devices in the hot end. In the experiments, we change the latter part of the DC flow by changing the direction of the asymmetrical flow resistance devices.
The Effect of the DC Flow in Coaxial PTC The lowest temperature is achieved by using the needle valve in the normal direction as can be seen in Fig. 8. In this case, the DC flow in the PTC is smaller than that of using the needle valve in the reversed direction, and larger than that of using the 5.1cm capillary. This can be clearly found by comparing the temperature profile curves in Fig. 8. The results with the needle valve are much better than those with the asymmetry-nozzle, which has a much smaller DC flow. This can be explained by considering the structure of the coaxial PTC. In the coaxial PTC, the cold tip is also the connector between the regenerator and the pulse tube. A proper positive DC flow adds an additional gas flow, which is cooled by the regenerator, from the cold end of the regenerator to the pulse tube. This gas flow can cool the cold tip and has positive effects on the cooling performance of the PTC, although it raises the temperatures of the regenerator and the pulse tube at the hot end and deteriorates the heat exchange efficiency of the regenerator at the same time.
CONCLUSIONS A G-M type coaxial pulse tube cryocooler has first been optimized from three aspects: the dimensional layout of the regenerator and pulse tube, the filling materials in the regenerator, and the structure of the coldend heat exchanger. The results indicate that both the lowest temperature and the cooling power of the PTC are improved remarkably. Next, needle valves, capillaries and asymmetry-nozzles have been used as hot end phase shifters, respectively. It is found that using the asymmetrical flow resistance devices is a practical way to control the DC flow in the DPTC, and a proper positive DC flow has positive effects on the cooling performance of the coaxial PTC.
ACKNOWLEDGMENT This work was supported by the National Natural Science Foundation of China (Grant No. 50176052).
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REFERENCES 1. Radebaugh R., “Pulse Tube Cryocoolers for Cooling Infrared Sensors,” Proceedings of SPIE, Vol. 4130 (2000), p. 363.
2. Mikulin E.I., Tarasov A.A., Shrebyonock M.P.P., “Low Temperature Expansion Pulse Tube,” Adv. Cry. Eng., vol. 29 (1984), p. 629.
3. Zhu S., Wu P.P., and Chen Z., “Double Inlet Pulse Tube Refrigerator-an Important Improvement,” Cryogenics, vol. 30 (1990), pp. 514.
4. Kanao K., Watanabe N. and Kanazawa Y., “A Miniature Pulse Tube Refrigerator for Temperature Below 100K,” Cryogenics, vol. 34 (1994) supplement, p. 167.
5. Charles I., Duband L., and Ravex A., “Permanent Flow in Low and High Frequency Pulse Tube Coolers: Experimental Results,” Cryogenics, vol.39 (1999), p.777.
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Experimental Study on Two-stage Pulse Tube Refrigeration with Mixtures of Helium and Hydrogen N. Jiang, Z.H. Gan, G.B. Chen, L.M. Qiu, Y.L. Jiang, Y.L. He, N. Li Cryogenics Laboratory, Zhejiang University Hangzhou 310027, P.R. China
ABSTRACT The effect of mixtures on the cooling power and efficiency (COP) of a 20K two-stage pulse tube refrigerator is studied in this paper. According to our preliminary findings we have concluded that a properly composed mixture can substantially increase the cooling capacity of a pulse tube refrigerator at 20 K. mixtures with 7%-46% were investigated and found to provide 18%-68% more cooling power than with pure helium at 20 K. The above are only preliminary results. In future research, we plan to conduct experiments with mixtures with even higher hydrogen fractions.
INTRODUCTION Our study of the use of gas mixtures in pulse tube refrigerators (PTRs) originated in 1996 and has included both: 1) theoretical analysis of the thermodynamics, heat transfer and fluid flow characteristics of mixtures in PTRs, and 2) experimental study of 80K PTRs with helium and nitrogen mixtures. The theoretical and experimental analysis results1-4 show that adopting appropriate mixtures can improve the refrigeration effect in PTRs. For example, the refrigeration power and COP of PTRs with a mixture(15% ) are higher than that with pure helium as the working fluid at 80K. In particular, taking advantage of the vapor-liquid-solid transform of the nitrogen component during the cooling process, a composition-independent temperature platform around the triple point can be achieved with an additional heat load, which is suitable for low-vibration operation of high Tc superconducting devices.5-6 Similarly, properly composed mixtures can achieve higher refrigeration power and COP at 80K. Thus, the use of mixtures as the working fluid in PTRs can improve the refrigeration power per cycle and corresponding thermodynamic efficiency without increasing the system size, weight or manufacturing cost. In order to explore the feasibility of PTRs with higher refrigeration power and COP at lower temperatures, an experimental study on multistage PTRs with mixtures has been carried out as described in this paper. The only effective working fluid pair at 20K is as theoretical calculations3-7 show that mixtures of helium and other inert gases will lead to an increase of regenerator pressure loss, while the improvement of heat transfer is very little. As a result, there is no improvement of the refrigerator performance. However, for hydrogen, the case is different. Hydrogen has excellent heat transfer and flow characteristics, so properly ratioing mixtures can improve the performance of regenerative coolers. Therefore, we have conducted experiments on PTRs with Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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mixtures, and the preliminary results indicate that a higher refrigeration power is achieved, with an appropriately composed mixture, than that with pure helium as the working fluid. Further improvement of the refrigeration system performance may be possible.
EXPERIMENTAL APPARATUS As shown in Fig. 1, the experimental apparatus used in this investigation comprised of a two-stage pulse tube refrigeration system, a measurement system, and a vacuum system; the solid lines represent the gas circuit, and the dashed lines are the electric circuit. The main part of the experimental device is the refrigeration system, which consists of the gas compressing system, the first and second stage regenerators, the pulse tubes, reservoirs, double inlet valves, heat exchangers and the second orifice, etc. The gas compressing system comprises the compressor, and the high and low pressure switching solenoid valves that are controlled by a computer program. The parameters of the regenerators and pulse tubes are listed in Table 1. The measurement system includes temperature, refrigeration power, and vacuum measurements, etc. Two calibrated rhodium-iron resistance temperature sensors are used to measure coldend temperatures of the first and second stage, respectively. And copper-constantan thermocouples are used to measure temperatures elsewhere. Using the heat balance method, manganin wires are wrapped around the cold heads and powered by a 0-45V constant voltage power supply to measure the refrigeration power.
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EXPERIMENTAL RESULTS In order to explore the feasibility of utilizing mixtures to improve the performance of PTRs at 20K, some experiments were conducted under the same operating conditions. Because of the existence of the vapor-liquid-solid transform of the hydrogen component during the cooling process, the case of mixtures is quite different from that of pure helium. A two-stage double orifice pulse tube refrigerator with 7%, 17%, 26%, 34% and 46% (hydrogen volume percentage) mixtures as the working fluid was first studied. In order to explain the cooling process of mixtures, Fig. 2 gives the cooling power curves corresponding to the different compositions. From the figure, the phenomena of the vapor liquid-solid transform of the hydrogen component during the cooling process can be seen. When the refrigeration temperature approaches the triple-point temperature of the vapor-liquid-solid coexistence of the mixture provides considerable latent heat, which fixes the refrigeration temperature around the triple-point temperature of The cooling capacity is much higher above this temperature because of the considerable vapor-liquid latent heat of the hydrogen component. Two important contributors are: 1) performance improvement of the cold end heat exchanger caused by the existence of liquid, which makes isothermal heat absorbing possible, and thus the cycle more closely approaches the Carnot cycle, and 2) the decrease of the cold end clearance volume. It can be seen that the refrigerator provides more cooling power with the mixtures than with pure helium under 30K. Also, with a larger fraction of hydrogen, this improvement of cooling capacity can be maintained over a wider temperature range. As the temperature increases further and the liquid phase in the mixture is completely vaporized, the slope of the cooling power curve gradually decreases. And the increment of cooling capacity with temperature is observably less than that in the vapor-liquid two phase coexistence region. At some temperature, the cooling capacity of the mixtures will become lower than pure helium. The vapor-liquid phase transition temperature of mixtures is related to the hydrogen fraction, i.e. the partial pressure of hydrogen in the mixtures. The vapor-liquid phase transition temperature of mixture increases with hydrogen content. So with increasing hydrogen fraction, the mixtures can improve the performance of pulse tube refrigerators over a wider temperature range. In the experiments we also found that the effect of the mixtures on the coefficient of performance is essentially the same as their effect on the cooling capacity. Thus, a COP curve similar to that in Fig.2 is also obtained. The cooling capacity at 20 K and the minimum (no-load) temperature acquired are shown in Fig. 3. From the figure, it can be seen that the minimum temperature with mixtures is almost the same as that of pure helium, with an increment of about 0.6-1.2 K. However, the cooling capacity of the refrigerator is highly increased. In fact, the mixtures with 7%-46% used in our experiments yielded 18%-68% more cooling capacity than that achieved with pure helium at 20K. In
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particular, when the mole fraction of hydrogen is larger than 7%, the cooling capacity of mixtures can be at least 1.5 times that of pure helium. Figure 4 shows warm-up curves of the second stage cold head with different refrigerants after the refrigerator was shut down. The figure shows that the warm-up trends with gas mixtures are quite different from that with pure helium. Except for the 7% curve, the other mixtures take a longer time than pure helium to reach 30K. Since part of the hydrogen in the mixture becomes liquid or solid during the cooling process, when the refrigerator stops working, the liquid or solid hydrogen vaporizes and absorbs latent heat, which makes the refrigerator stay under 30K for a longer time. Thus, the time is proportional to the quantity of solid or liquid hydrogen. The time for pure helium is 5 minutes or so, and about 10,10.5,18 and 17 minutes for 17%, 26%, 34% and 46% mixtures, respectively.
CONCLUSIONS According to our preliminary experimental study on pulse tube refrigeration with mixtures we have concluded that a properly composed mixture can increase the cooling capacity of a pulse tube refrigerator at 20K. During the cooling process of a PTR with mixtures, the vapor-liquid and liquid-solid phase transitions of the hydrogen component play an important role. When the refrigeration temperature approaches the triple-point temperature of hydrogen, the vapor-liquid-solid coexistence of the mixture provides considerable latent heat, which fixes the refrigeration temperature around the triple-point temperature. Under the same operating conditions, mixtures with 7%-46% were found to provide 18%-68% more cooling power than with pure helium at 20 K. In particular, when the mole fraction of hydrogen is larger than about 10%, the cooling capacity with mixtures can be more than 1.5 times as much as that with pure helium. The above are only preliminary results. In future research, we plan to conduct experiments with mixtures with even higher hydrogen fractions.
ACKNOWLEDGMENT This work is financially supported by the National Natural Sciences Foundation of China (59976034), the University Doctoral Subject Special Foundation of China (20010335010) and the Foundation for the Author of National Excellent Doctoral Dissertation of P.R. China (200033).
REFERENCES 1.
Chen, G.B., Gan, Z.H., Thummes, G., Heiden, C., “Thermodynamic performance prediction of pulse tube refrigeration with mixture fluids,” Cryogenics, vol. 40 (2000), pp. 262-267.
2.
Gan, Z.H., Chen, G.B., Thummes, G., Heiden C., “Experimental Study on Pulse Tube Refrigerator with Helium and Nitrogen Mixtures,” Cryogenics, vol. 40 (2000), pp. 333-339.
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3.
Yu, J.P., Chen, G.B., Gan, Z.H., et al., “Discussion on regenerator performance improvement with binary gas mixture,” Proceeding of ICEC-17, Institute of Physics Publishing (1998), pp. 117-122.
4.
Chen, G.B., Yu, J.P., Gan, Z.H., Jin, T., “Experimental investigation on pulse tube refrigerator with binary mixtures,” Adv. in Cryogenic Engineering, Vol. 45A, Kluwer Academic/Plenum Publishers, New York (2000), pp. 183-187.
5.
Chen, G.B., Gan, Z.H., Qiu, L.M., Yu, J.P., “Pulse tube refrigeration with a cooling and freezing combined cycle for HTSC devices,” Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2001), pp. 291-300.
6.
Chen, G.B., Jiang, Y.L., Gan, Z.H., “Investigation of a near-63K isothermal in pulse tube refrigerator with mixtures,” Adv. in Cryogenic Engineering, Vol. 47 A, Amer. Institute of Physics, Melville, NY (2002), pp. 855-862.
7.
Daney, D.E., “Regenerator performance with noble gas mixtures,” Cryogenics, vol. 31, no. 10 (1991), pp. 854-861.
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Experimental Investigation of 4K VM Type Pulse Tube Cooler W. Dai, Y. Matsubara, H. Kobayashi Institute of Quantum Science, Nihon University Funabashi, Chiba, Japan 274-8501
ABSTRACT This article presents a new type of 4K pulse tube cooler that we refer to as the VM type pulse tube cooler. What is different from the GM and Stirling type pulse tube coolers is that a thermal compressor is used instead of a mechanical compressor to provide the pressure wave for obtaining temperatures below 4K. Our research focus is on developing more efficient 4K coolers and coolers that can provide temperatures below 2K with The whole system can be divided into two independent subsystems: the VM side, and the GM side. On the VM side, the thermal compressor has no moving components at cryogenic temperatures and is made up of a room temperature displacer, work transfer tube, and one regenerator. This compressor utilizes the temperature difference between liquid nitrogen and room temperature to generate the pressure wave for the 4K pulse tube cooler. The GM side is only used to provide some precooling power at about 20K for gas entering the regenerator in the VM side. Using a temperature below 4K has been obtained with a pressure ratio near 1.3, working frequency of 1 Hz, and charge pressure of 2.1 MPa. The cooling power at 5K is about 12 mW.
INTRODUCTION Temperatures below 4K are very important for many advanced applications. Lower temperature and higher efficiency at this temperature level are two challenging targets for cryocoolers. GM coolers or GM type pulse tube coolers can provide a lowest temperature above 2K and some cooling power at 4K by using But there are two obstacles for further improvement on these 4K coolers: firstly, the thermodynamic properties of limit both the obtainable lowest temperature to being above 2K and efficiency at 4K; secondly, the existence of the gas distribution valves in the GM type coolers causes serious intrinsic losses. To solve the first obstacle, can be used to replace as the working gas. The thermodynamic property of makes it possible to reach higher efficiency at 4K and temperatures below 2K, which has been experimentally proven by researchers in Sumitomo Heavy Industries[1] and Eindhoven University[2]. Due to the extremely high price of reducing the total system volume is important for controlling the cost for future potential applications. The irreversible loss due to the pressure drop through the gas distribution valves is intrinsic and serious in GM type coolers. Valveless compressors can be used to eliminate this loss, but Stirling com-
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pressors are not applicable for these low frequency 4K coolers. The concept of a thermal compressor similar to that used in a VM cooler is one of the possible choices. Based on these considerations, a thermally driven pulse tube cooler has been proposed, and this article is the latest of a series of reports on our progress [3-6]. Our research targets are: 1) more efficient 4K pulse tube coolers by using valveless thermal compressors, and 2) coolers that can reach temperatures below 2K with the least amount of Since our last report [6], we have made some important changes to our system and obtained temperatures below 4K. Although the basic ideas were introduced in that report, we will repeat the important features for the integrity of this article.
EXPERIMENTAL SETUP Figure 1 gives an illustration of our system. In the right side of Figure 1 (surrounded by dashed lines), a liquid nitrogen precooled 4-valve pulse tube cooler driven by a 1.8 kW compressor is used to provide about 2W cooling power to pre-cool the gas entering the 3rd regenerator through the 20K heat exchanger. This is called the GM side. The other parts constitute the VM side. In the following, names such as pulse tube, regenerator, refer only to those in the VM side. The main configuration parameters of the VM side are given in Table 1. Displacer, work transfer tube, liquid nitrogen heat exchanger and 1st regenerator constitute the thermal compressor. In contrast to the thermal compressor in a VM cooler, the displacer in Fig. 1 works at room temperature. When the
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displacer piston moves cyclically, pressure wave will be generated. The mechanism is explained in reference[6]. The 2nd regenerator, 3rd regenerator and pulse tube can be viewed as an ordinary pulse tube cooler which uses the pressure wave provided by the thermal compressor. The expander is the phase shifter at pulse tube hot end. By using an expander as the phase shifter, there is much more freedom in adjusting the mass flow amplitude and phase at the pulse tube hot end. Orifice 1, reservoir, orifice 2 and the check valve are used to introduce and control the DC flow as shown by the dashed line in the VM side. The check valve only allows gas to flow from the reservoir toward 1st regenerator. When pressure inside the reservoir is higher than that inside the 1st regenerator, gas will flow through the orifice and check valve and enters the 1st regenerator. If the pressure inside the reservoir is lower, the check valve will be closed and no gas will flow through it. This will generate a stream of DC flow from the pulse tube cold end to the hot end, which plays a very important role as will be described in the next section. Four pressure gauges are used: at the displacer’s work transfer tube side, at the displacer’s regenerator side, at the pulse tube hot end, and at the reservoir. The temperatures of pulse tube cold end and 3rd regenerator hot end are measured by Cernox thermometers. One Pt thermometer is placed at the middle of pulse tube to monitor the DC flow.
EXPERIMENTAL RESULTS AND DISCUSSION The experiment generally begins by filling with liquid nitrogen. Then both the GM and VM sides start to run. Figure 2 provides typical cool-down curves for different parts of the system.
Thermal Compressor Characteristics After the liquid nitrogen filling is complete, the work transfer tube cold end quickly cools from room temperature down to around 100K within 15 minutes, as can be seen in Fig. 2. The amplitude of the pressure wave also quickly increases. With the present configuration, the high to low pressure ratio is nearly 1.3 when the cold end temperature is higher than 200K, and around 1.25 when the cold end reaches 4K. The equivalent cooling power of the required liquid nitrogen for the thermal compressor is about 20 to 25 watt. Figure 3 shows a typical displacer movement and the generated pressure wave (the lower end of displacer cylinder shown in Fig. 1 is taken as the zero position). The movement and the pressure waves are almost in phase. Due to the small capacity of the thermal compressor, the expander’s movement imposes a large influence on the pressure waves when the cold end temperature decreases below 10K. Figure 4 shows the expander movement’s influence when the cold end temperature is 180K. and 4.2K, respectively. In the case of 180K, there is almost no difference in the amplitude of the pressure waves either with the expander running or stopped. However, when the cold temperature is at 4.2K, running or stopping the expander can make a large difference as shown in the lower two graphs of Fig.4. The cold end temperature will begin to rise quickly above 10K if the expander is not running.
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Active DC Flow Control and Cooling Performance Considering the relatively small pressure ratio inside the system, we use the strong DC flow from the pulse tube cold end to the pulse tube hot end at the beginning period to bring the cooling power of liquid nitrogen down to cool the 2nd and 3rd regenerators faster. This is shown in Fig. 2 by the closeness between the pulse tube cold end temperature and the middle temperature before they reach 50K. When the temperatures of the pulse tube cold end and the 3rd regenerator hot end get close to each other, the strong DC flow will suppress the cooling effect at the cold end and become a loss. Therefore, the DC flow is manually reduced at this point, as shown in Fig. 2 by the abrupt increase in the pulse tube middle temperature and faster cool-down of both pulse tube cold end and GM-side cold end. With this practice of active DC flow control, the pulse tube cold end generally cools down below 8K within one hour and reaches 4K within 100 minutes. At this point, we haven’t yet measured the quantity of the DC flow; we expect to do that in the future. There are many adjustable parameters that can affect the system performance. These include the expander’s phase relationship with the displacer, its swept volume, movement type and the DC flow amount; they are all important for obtaining temperatures below 4K. Up to now, our best result is a no-load temperature near 3.5K under the following conditions: 1 Hz operating frequency and 2.1 MPa charge pressure; expander and displacer pistons nearly 180 degree out of phase (from Fig. 1, this means when displacer piston moves upward, expander piston moves downward) with constant speed; small amount of DC flow, which is controlled by very small opening of orifice 1 and orifice 2; displacer piston at full stroke, while expander’s actual swept volume is about 7cc, 1/3 of the pulse tube volume; 3rd regenerator hot end is cooled to about 20 K by the GM cooler. Figure 5 gives a typical cooling power curve. The cooling power is in the order of several mW at the 4K region.
Energy Flow Inside the System Figure 6 shows typical pressure waves inside the system when the cold end is at 4K. Although there is much noise in the signal of pressure wave inside reservoir, it is clear that its value is much lower than the average value of the other three pressure waves. It is also clear that the pressure drop inside the whole system is so small that the three pressure wave lines nearly overlap each other. Under this condition, it is interesting to observe the amount of work that is done inside both the displacer and expander. Figure 7 gives the PV diagram of both sides of displacer piston and expander piston according to Fig. 6. Due to the unit arrangement, the circle areas are exactly the work
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done by the pistons in one cycle. For the displacer piston, the figure shows that the piston does work on the gas at the work transfer tube side (outer circle) and absorbs work from the gas at the regenerator side (inner circle). Both areas are about 10 J. Subtracting both areas gives the net work done by the displacer piston in one cycle. Because the system’s working frequency is near 1 Hz, the result is that the power required by the displacer piston is much less than 10 watts. This is the characteristic of this type of thermal compressor. For the expander piston, it is very important to notice that the piston does work on the gas with less than 1 watt. In other words, the expander actually works as a small compressor in this 4K pulse tube cooler. A similar phenomenon is also reported in reference[7].
CONCLUSION Our progress to date includes using a thermally driven pulse tube cooler with to obtain a lowest temperature below 4K and about 12 mW of cooling power at 5K. We have called this type of pulse tube cooler a VM type due to the similarity of its thermal compressor concept to that of VM coolers. However, an important difference is that the thermal compressor design in our system eliminates moving components at cryogenic temperatures by using a work transfer tube—similar to what pulse tubes have achieved in other coolers. This new type pulse tube cooler provides an approach for future 4K applications, and 2K applications (by using ). This is the first time we have achieved a temperature below 4K with a small pressure ratio (below 1.3). Because of the many adjustable parameters, there is still considerable potential for further improvement. One of the most important issues is how to build a more efficient thermal compressor. The effect of DC flow will also be studied in more detail, especially the practice of active control.
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REFERENCES 1.
Satoh T., Onishi A., Umehara I, Adachi Y, Sato K and Minehara E.J. “A Gifford-McMahon Cycle Cryocooler below 2K,” Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2001), pp. 381-386.
2.
Xu M.Y., De Waele A.T.A.M., Ju Y.L., “A pulse tube refrigerator below 2K,” Cryogenics, 1999; Vol.39, p.865.
3.
Matsubara Y., Zhou S.L., “Feasibility study of applying thermal compressor to 4K pulse tube cooler,” Proceedings of ICEC 18, 2000; p.535.
4.
Matsubara Y, Kobayashi H, Zhou S.L. “Thermally actuated 3He pulse tube cooler,”Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2001), pp. 273-280.
5.
Dai W., Zhou S.L., Matsubara Y., Kobayashi H, “Thermally driven pulse tube refrigerator,” Proceedings of 64th Meeting on Cryogenics and Superconductivity, 2001, p.23.
6.
Dai W., Matsubara Y., Kobayashi H., “Experimental results on VM type pulse tube refrigerator,” to be published in Cryogenics 2001, Supplementary.
7.
Zhou S.L., “Energy transportation of oscillating flow in pulse tube”, Doctor Thesis of Nihon University, Jan. 2001.
Affecting the Gross Cooling Power of a Pulse Tube Cryocooler with Mass Flow Control A. Waldauf1, T. Schmauder2, M. Thürk1, and P. Seidel1 1
Institut für Festkörperphysik, FSU Jena D-07743 Jena, Germany 2
Leybold Optics GmbH D-63450 Hanau, Germany
ABSTRACT To increase the cooling capacity of a pulse tube cryocooler the gas expansion work (pV-work) at the cold end of the pulse tube needs to be increased and the system losses have to be reduced simultaneously. For a given pulse tube volume the gross cooling power is limited. Since the cooling effect originates from the phase shift angle between the mass flow and the temperature, this angle plays a significant role in designing a pulse tube refrigerator. In conventional pulse tube refrigerators, such as orifice or double inlet refrigerators, the phase shift is thermodynamically limited. That phase shift can be enlarged with an active control unit at the hot end of the pulse tube. The control unit manages the gas flows into and out of the pulse tube. In cases of differing mass flows during the working cycle, a further enlargement of the phase shift is possible. Here we present such a system which reaches a cooling power of 85 W at 80 K with an input power of 4.8 kW.
INTRODUCTION The invention of the Pulse Tube Refrigerator (PTR) has led to intensive research that aims at benefiting from the great advantages of this cooler type. Since the pulse tube cryocoolers do not contain any moving parts in their coldspace, the most important advantages are minimal mechanical vibration, simple construction, and more or less unlimited lifetime. This makes them differ from conventional and widely-used refrigerators, like Stirling or Gifford McMahon refrigerators. Although the basic cooling principle of the PTR has been known for nearly forty years [1], in the beginning this cooler type was of limited importance because of its bad coefficient of performance. Recently, due to increasing improvements in the efficiency, the PTR has become a serious alternative to the well established regenerative cryocoolers. A possible application field of the PTR is the temperature region around 80 K, not only for the cooling superconducting sensors [2], but also for cooling radiation shields as used in systems for magnetic resonance imaging. In such cases a cooling power of 100 W or more at 80 K is necessary. To achieve such a high cooling power we have designed and built a Four-Valve Pulse Tube Refrigerator (FVPTR) in a first stage. The setup and the working principle of this refrigerator is Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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described elsewhere in detail [3]. In a second stage we have developed a powerful tool to increase the p-V area of the cold and the hot space with an intrinsic measurement of the pressure waves and the temperature in the pulse tube [4]. This tool is similar to the thermodynamic analysis of active pulse tube refrigerators by Pfotenhauer et al. [5] or to the isothermal model for the orifice pulse tube refrigerator proposed by Chen, et al. [6]. Typical p-V areas of the cold end for a FVPTR are shown in Figure 1. Obviously, while increasing the exchange volume at the hot end of the pulse tube the shape of the diagram changes from a long and thin one to a more rhombic one But since an additional parameter is missing, it is impossible to get a more rectangular shape of the p-V area. Therefore a higher gross refrigeration power is not available. Kakimi et al. [7] solved this problem by using an active buffer system. To enlarge the p-V work with a FVPTR, the mass flow rate into and out of the pulse tube during compression and expansion needs to be higher than during the displacement of the gas column. By means of two additional valves at the hot end of the pulse tube we are able to control each section in the thermodynamic cycle.
EXPERIMENTAL SETUP AND PROCEDURE A schematical drawing of this “Six-Valve” Pulse Tube Refrigerator (SVPTR) is shown in Figure 2. This is an usual FVPTR additionally equipped with two solenoid valves (V3 and V6) which connect the hot end of the pulse tube with the low and high pressure side of the compressor,
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respectively. Therefore it is possible to switch between a Four-Valve and a Six-Valve mode, in order to compare the cooling power for a chosen setup. To control the mass flow into and out of the pulse tube highly accurate needle valves are series connected to the valves at the hot end of the pulse tube (V3-V6). A small electronic interface enables us to define the opening and closing time of the two solenoid valves, respectively. The working fluid is helium supplied by a CTI-compressor with a maximum electrical power of 4.8 kW. By means of a power measurement system we are able to obtain the real input power at any time during the working period. The whole measuring setup is described elsewhere in detail [8]. A schematic diagram with the main dimensions of the “Six Valve” Pulse Tube Refrigerator is given in Table 1. Alternatively, and in order to make the refrigerator more compact, we developed a new rotary valve made of teflon (see Figure 3). In difference to the FVPTR rotary valve, two additional throttles are integrated into the rotary valve. The rotary valve enables us to gain two different mass flow rates into the pulse tube. Whereas during the compression and expansion phase gas flows into and out of the pulse tube through the conventional valves, the displacement of the gas column is done by the throttles. In this configuration the needle valves are only used to control the mass flow during compression and expansion, as the mass flow coefficient of the throttle is smaller than the one of the needle valve.
Description of the Working Cycle Excluding two important exceptions, the working cycle of the SVPTR is similar to the one of the FVPTR. To affect the size of the cold volume, there is an additional mass flow into and out of the pulse tube during the compression and expansion phase, respectively. The cycle can be divided roughly into six segments. A schematic view of the valve timing was shown in Figure 2.
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1) The working cycle starts with low pressure in the pulse tube. V5 opens for phase shifting in the pulse tube. 2) The main inlet V1 is opened. Gas enters the working space from both inlets, resulting in a compression. Simultaneously V3 opens to affect the size of the cold volume. 3) After V5 and V3 are closed, the pressure continues to increase and reaches the surge pressure. 4) V1 is closed. V4 opens for phase shifting. 5) The main outlet V2 opens. Expansion takes place. V6 opens and the gas column in the pulse tube is shifted to the hot end. 6) After V4 and V6 are closed the pressure returns to low pressure.
Experimental Description While working with the SVPTR, special attention was paid to the fact that for a given setup the total exchange volume of the SVPTR and the FVPTR is almost the same, although for the first the mass flow is higher during the expansion and compression period. Therefore the mass flow of the valves V4 and V5 was reduced from 2.5 mg/s to approximately 2 mg/s at typical pressure differences of 14-15 bar in the pulse tube during one working cycle. All measurements were carried out with the same frequency and the highest possible pressure difference between low and high pressure.
RESULTS AND DISCUSSION The thermal performance of the SVPTR with the two solenoid valves and the two throttle valves at optimized conditions is shown in Figure 4. The corresponding efficiency in %Carnot is shown in Figure 5. Both refrigerator setups provide a minimum temperature of 38 K. The relatively high minimum temperature is due to the fact that the heat exchanger at the cold end is designed for high cooling power. Temperatures below 30 K would be reached for both setups if the heat exchanger was to be designed for minimum temperature. While the cooling power for the setup with the solenoid valves is about 85 W at 80 K, the setup with the throttle, valves reaches only 60 W at 80 K. We explain this degradation in the cooling performance with a decrease of the enthalpy flow at the warm end of the pulse tube. A temperature distribution measurement along the pulse tube showed that applied heat at thecold end leads to a higher average temperature at the hot end of the pulse tube than in the setup with the solenoid valves. The difference in the average temperature was about 20 K.
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To verify our assumption we fixed a water cooled heat exchanger at the hot end of the pulse tube to decrease the temperature. But with this setup the cooling power was only 45 W at 80 K. This proves that the energy of a FVPTR or SVPTR is totally rejected by enthalpy flow at the warm end of the pulse tube. A use of a heat exchanger acts like a regenerator and reduces the gross refrigeration power because of a decrease of the enthalpy flow. Therefore the valve timing has to be further optimized in order to increase the enthalpy flow at the hot end of the pulse tube for this setup. The setup with the solenoid valves showed a much better performance. In comparison to the FVPTR there was an improvent of the cooling power of about 5 % at a working temperature of 80 K. At a usual working frequency of 2.8 Hz the optimal openings for the solenoid valves V3 and V6 were 30 and 40 ms, respectively. If the opening time was further increased there was a decrease in the cooler performance.
CONCLUSIONS In this paper we described the development, test, and performance of a Six Valve Pulse Tube Refrigerator. At present this refrigerator provides a minimum temperature of 38 K and a cooling power of 80 W at 76 K. This cooling power is equal to 6.7 % Carnot. In comparison to the Four Valve Pulse Tube Refrigerator the cooling power is about 5 % better at 80 K. A setup with a throttle valve showed a cooling power of 60 W at 80 K. In this case the valve timing has to be optimized to increase the enthalpy flow at the hot end of the pulse tube.
ACKNOWLEDGMENT This work is supported by the BMBF (No. 13 N 7395). We thank G. Cybik, L. Föllmer and R. Neubert for technical assistance.
REFERENCES 1.
W.E. Gifford and R.C. Longsworth , “Pulse-tube refrigeration,” Trans. ASME J. Eng. Ind. (Ser. B) 86 (1964), pp. 264-268.
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2.
J. Gerster, G. Kaiser, L. Reißig, M. Thürk, and P. Seidel, “Low Noise Cold Head of a Four-Valve Pulse Tube Refrigerator,” in Advances in Cryogenic Engineering 4, Plenum Press, New York (1998), pp. 2077-2084.
3.
T. Schmauder, A. Waldauf, R. Wagner, M. Thürk, and P. Seidel, “Investigation on a Single FourValve Pulse Tube Refrigerator for High Cooling Power,” Cryocoolers 11, Plenum Press, New York (2000), pp. 327-336.
4.
A. Waldauf, T. Schmauder, M. Thürk, and P. Seidel, “Investigation of Energy Transport within a Pulse Tube,” Advances in Cryogenic Engineering, Plenum Press, New York (2001), in press.
5.
J. Yuan, J.M. Pfotenhauer, “Thermodynamic analysis of active pulse tube refrigerators,” Cryogenics, 39, 1999, pp. 283-292.
6.
S.W. Zhu, Z.Q. Chen, “Isothermal model of pulse tube refrigerator,” Cryogenics, 34, 1994, pp. 591595.
7.
Y. Kakimi, S.W. Zhu, T. Ishige, K. Fujioka, and Y. Matsubara, “Pulse-Tube Refrigerator and Nitrogen Liquefier with Active Buffer System,” Cryocoolers 9, Plenum Press, New York (1997), pp. 247-254.
8.
M. Thürk, H. Brehm, J. Gerster, G. Kaiser, R. Wagner, and P. Seidel, “ Intrinsic Behaviour of a Four Valve Pulse Tube Refrigerator,” Proc. of the 16th International Cryogenic Engineering Conference/ International Cryogenic Materials Conference, 1996, pp. 259-262.
9.
J. Gerster, M. Thürk, L. Reißig, and P.Seidel, “Hot end loss at pulse tube refrigerators,” Cryogenics, 38, 1998, pp. 679-682.
Pressure Wave Generator for a Pulse Tube Cooler Y. Matsubara, W. Dai Institute of Quantum Science, Nihon University Funabashi Chiba 274-8501 Japan H. Sugita and S. Tooyama Office of Research & Development, NASDA Tsukuba Ibaragi 305-8505 Japan
ABSTRACT This paper describes the concept of a pressure wave generator for a pulse tube cooler without the use of a mechanical compressor. To understand the basic mechanism, a work amplifier was fabricated and tested. It consists of a work input piston, a regenerator with heat exchanger on both sides, a work transfer tube, a resonator, and a work receiver. These five critical components are connected in series. The pressure wave is generated by the piston at the resonance frequency of the resonator. By heating the heat exchanger, which is located at the work transfer tube side of the regenerator, the input work from the piston is amplified through the regenerator and flows out from the work transfer tube due to the conversion of the heat flow into the work flow. This amplification mechanism makes it possible for the input work to be replaced by a part of the output work through the feedback line. Finally, it becomes a self-actuated pressure wave generator without any external input work. This study mainly discusses the performance of each component to improve the total performance of the system. A simplified analytical method using the equivalent PV work and preliminary experimental results is also given.
INTRODUCTION There are three different types of pulse tube coolers from the view point of the pressure wave generator: the Stirling type uses a directly driven mechanical piston, the GM type uses a circulating gas compressor with pressure switching valves such as a rotary valve, and the thermoacoustic or VM type uses a thermal compressor instead of a mechanical compressor. This third type pressure wave generator can be further classified as follows: (1) Standing wave type based on stack instead of the regenerator. (2) Looped type having the circumference of the sonic wavelength of its resonance frequency. (3) External long resonator with small feedback loop. (4) Mechanical displacer with work transfer tube. (5) Mechanical resonator with a feedback loop. Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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The first concept has been extensively studied as the thermoacoustic pressure wave generator for pulse tube coolers since 1988 [1, 2]. The thermodynamic efficiency of this method, however, is rather lower than the other progressive wave types, which use a regenerator instead of a stack. The second and third concepts, respectively, have been developed by Yazaki and Tominaga [3], and Backhaus and Swift [4] as progressive pressure wave generators. Because of the relatively large resonator size, these concepts are more suitable for relatively large-scale applications. The fourth and fifth concepts have not yet been studied very much. However, the low frequency mechanical displacer, such as that of the VM cycle, has been studied recently. This paper mainly focuses on the basic performance of the resonator system using the solid displacer as a part of the pressure wave generator. BASIC OPERATION PRINCIPLE Figure 1 shows the basic concept of a work flow amplifier. It consists of the work input piston, regenerator having heat exchanger on each side, work transfer tube, resonator, and work output piston. This output piston can be replaced by a set of orifice and reservoir as shown in the figure. If the resonator is not connected, and each piston is linked properly, it becomes a Stirling type pulse tube cooler or a prime mover (without hot piston), which depends on the heat exchanger temperature at the junction of the regenerator and the work transfer tube. In the case of an orifice and a reservoir, it becomes an orifice pulse tube cooler or simple work amplifier. If the resonator is connected as shown in Figure 1, the orifice pulse tube cooler become an inertance type pulse tube cooler, which gives better thermodynamic performance. In general, the phase shifter used in the inertance type consists of a narrow long tube and the reservoir in series. However, this parallel configuration could become an alternative. Here the resonator 1 and 2 are arranged symmetrically to eliminate the mechanical vibration caused by each displacer’s movement. If the heat input at the heat exchanger, Hxh, increases until the temperature exceeds room temperature, it become a work amplifier. The input work from the piston is amplified through the regenerator and flows out from the work transfer tube due to conversion of the heat flow into the work flow. In this case, the function of the resonator becomes more important. To have a better understanding of this effect, the following simplified numerical analysis has been done. Simplified Analysis The ideal gas assumption (gas constant: R) was used for following equations. The notation is given in Figure 2.
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Momentum equation of the displacer: Pressure equation at the work transfer tube: Pressure equation at the backside volume Pressure equation at the work input piston: where the oscillating volume is given as the boundary condition as: Orifice mass flow equation: Pressure in the reservoir: Regenerator mass flow equation is simplified as: where is the equivalent flow coefficient given at the middle of the regenerator. Example calculation in the case of using orifice and the reservoir as the work receiver has been done with the following input parameters in Table 1. The results are given in Figure 3, which clearly shows the existence of the optimum driving frequency, which gives maximum work amplification. The maximum ratio of output work and input work is 1.54 at a frequency of 23.6 Hz. It is noted that the resonance frequency of the displacer does not coincide with this optimum frequency. If the displacer is removed, the work
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amplification ratio is represented by the line of This result indicates the effectiveness of using a displacer as the resonator. Each equivalent PV work at the driving frequency of 24 Hz is shown in Figure 4. The parameters are the same as in Table 1 except and (a) is the input work at the compressor piston, (b) is the output work, which is the dissipated work at the orifice in this case, and (c) is the equivalent work at both ends of the work transfer tube. This example calculation indicates the size of the work transfer tube is large enough. (d) is the dissipated work at the displacer. This narrow PV shape indicates the high quality factor of the resonator to generate the standing wave. The input work of 45.7 watts at the piston is amplified to 85 watts and passes through the work transfer tube to the orifice. The lost work at the displacer is 14.5 watts. This relatively large loss is due to the assumption of a large value of However, the output work is 70.2 watts, and the ratio of 1.54 still remains.
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EMPIRICAL DEMONSTRATION An experimental test device was fabricated with each component sized the same as the values displayed in Table 1. One preliminary test result is given in Figure 5. The driving frequency is 24.5 Hz with The heating temperature Th is 556 K, Ta is kept around 300 K, and the orifice Cv value is 0.2. The measured input work at the compressor piston was 25.5 watts and the output work at the orifice was 33.1 watts. Thus, a work amplification ratio of 1.3 was obtained.
SELF ACTUATED SYSTEM Figure 6 shows a possible arrangement for a self actuated pressure wave generator. Since the output work in Figure 1 is larger than the input work at the optimum driving frequency, the input work could be replaced by a part of the output work through a feedback line. As a result, external work input is no longer required and the whole system becomes a self-actuated pressure wave generator without any external input work. Here, the orifice and reservoir as a work receiver could be replaced by a pulse tube with a regenerator.
Simplified Analysis Instead of the compressor equation:
another momentum equation,
could be applied to solve a system such as the one shown in Figure 6. Here the solid mass used for the convenience of the numerical analysis.
is
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An example calculation has been done using the same parameters as shown in Table 1 with additional parameters: Other parameters different from Table 1 are, Th=800 K, and Co=0.7. The self-actuated oscillating frequency is 21.6 Hz. Work flow within the work transfer tube of 36.7 watts is obtained. The lost work at the displacer is 3.3 watts and the output work is 18.3 watts. The feedback work to the additional solid displacer is 14.7 watts, and then it is reduced to 14.6 watts and supplied to the regenerator.
CONCLUSION This article discussed the possibility of a self-actuated pressure wave generator using a solid displacer. As a first step, experiments using a mechanical compressor to introduce external work input have been conducted, and the work amplification effect has been verified. The results are well explained by the numerical analysis. Based on these, design of a self-actuated pressure wave generator has been presented and numerically verified. The inertance effect of the work flow feedback line is essential for realizing of the self-actuated pressure wave generator. Further experimental study is required to confirm the calculated results.
ACKNOWLEDGMENT We would like to thank Prof. De Waele (Eindhoven) for his helpful discussions.
REFERENCES 1. Hofler, T.J. “Concepts for Thermoacoustic Refrigeration and a Practical Device,” Proceedings of the 5th International Cryocooler Conference, Monterey, CA, August 18-19,1988, Chaired by P. Lindquist, AFWAL/FDSG, Wright-Patterson AFB, OH, pp. 93-101. 2. Godshalk, K.M., Jin, C., Kwong, Y.K., Hershberg, Swift, G.W. and Radebaugh, R., “Characterization of 350 Hz Termoacoustic Driven Orifice Pulse Tube Refrigerator with Measurements of the Phase of the Mass Flow and the Pressure,” Advances in Cryogenic Engineering, 41, Plenum Press, (1996), pp. 1411-1418. 3. Yazaki, T., Iwata, T., Maekawa, T. and Tominaga, A., “Traveling Wave Thermoacoustic Engine in a Looped Tube,” Phys. Rev. Lett. 81,(1998), pp. 3128-3131. 4. Backhaus, B. and Swift, G. W., “A Thermoacoustic Stirling Heat Engine”, Nature 399,(1999), pp. 335-338.
A First Order Model of a Hybrid Pulse Tube/Reverse-Brayton Cryocooler G. F. Nellis1, J. R. Maddocks2, A. Kashani2, J. H. Baik1, and J. M. Pfotenhauer1 1
University of Wisconsin Madison, WI, USA 53706 2
Atlas Scientific San Jose, C A, US A 95120
ABSTRACT This paper describes a cryogenic refrigeration cycle that combines a recuperative lower stage with a regenerative upper stage. The resulting cryocooler avoids the inherent thermal saturation and void volume losses that are associated with a regenerator and therefore has the potential for high performance at low temperature. The hybrid concept also has advantages relative to thermal, mechanical, and electrical integration as well as reliability. The hybrid configuration examined here uses a system of check valves and buffer volumes in order to rectify the oscillating flow within a pulse tube into a quasi-steady flow that can energize a reverse-Brayton stage. This system is analyzed using a first-order model. The governing equations are derived and then reduced to dimensionless form through the selection of logical dimensionless groups that characterize various aspects of the hybrid system. The ideal model is verified by showing that it predicts the correct efficiency in the limit of pure pulse tube operation. The model is used to investigate the effect of the rectification system on cycle performance; both in terms of losses associated with the check valves as well as turbine power fluctuations caused by cyclic variations in the flow. Finally, the model is used to optimize a hybrid cryocooler for a specific set of operating conditions by maximizing the heat lifted per heat transfer in the recuperator and regenerator.
INTRODUCTION Several cryogenic technologies require a low cost, closed cycle refrigeration system that can achieve high efficiency. High efficiency demands the use of a recuperative cycle at low temperatures, yet the reverse-Brayton systems currently available are very expensive and the performance of Joule-Thomson systems is limited by fluid properties. Regenerative systems are available at reasonable cost but suffer from the competing loss mechanisms of thermal saturation and void volume pressurization. These loss mechanisms become dominant at operating temperatures below 10 K, limiting the efficiency that can be realized with Stirling, GiffordMcMahon, or pulse tube cryocoolers. Hybrid cryocoolers such as Gifford-McMahon/JouleThomson systems (e.g. [1,2]) or the Boreas refrigerator [3] are better suited for low temperature Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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operation. The cryocooler technology described in this proposal is another type of hybrid concept in which a regenerative upper stage is integrated with a recuperative lower stage through a system of check valves and buffer volumes. A single, linear compressor can activate this system. The regenerative stage provides a low cost, high efficiency means of achieving an intermediate temperature. The recuperative stage can achieve high efficiency at low temperature and will not have to span the entire temperature range. Furthermore, by activating the recuperative stage at an intermediate temperature with a high pressure difference it becomes possible to use a relatively small recuperative heat exchanger and relatively simple hydrostatic rather than hydrodynamic journal bearings to support the turbine. In the first section of this paper we present a first order model of a hybrid system in which a high temperature pulse tube stage is integrated with a low temperature reverse-Brayton stage. The governing equations for this highly idealized system are derived and then nondimensionalized in order to point out the most important physical groups that govern the system's behavior. The model is verified by showing that it predicts the correct system performance in the limit of pure, pulse tube operation. In the second section, the model is used to examine the effect of the interface components, the check valves and buffer volumes, on the overall cycle behavior. The results of an optimization are presented where the refrigeration load per heat exchanger load is maximized. The final section summarizes this work.
FIRST-ORDER MODEL One of the technical challenges associated with the proposed rectified hybrid cryocooler is converting the oscillating flow within the regenerative cycle into a steady flow for the recuperative system in an efficient and practical manner. In order to examine the fundamental issues associated with this integration process, a first order model of a pulse tube/reverseBrayton system has been developed. The model described here is a substantial modification of the 2-piston, isothermal hybrid Stirling/reverse-Brayton cycle that is described in [4]; which itself is an extension of the well known isothermal Schmidt analysis of a 2-piston Stirling cycle [5]. Here, we consider an orifice pulse tube stage rather than a 2-piston Stirling system interfaced with a reverse-Brayton recuperative stage. This configuration is illustrated in Figure 1. A single linear compressor energizes the system. The orifice pulse tube components are the typical combination of regenerator, pulse tube, orifice, and reservoir. The recuperative stage is activated through the high- and low-pressure check valves that are oriented in opposite directions, as shown. Gas from the pulse tube system passes through high-pressure check valve and into a buffer volume during part of the cycle. The buffer volume stabilizes the inlet pressure seen by the cryogenic turbine in the reverse-Brayton system. A quasi-steady flow of highpressure gas passes through the hot-side of the recuperative heat exchanger and enters turbine where it is expanded, doing work on a rotor. This causes the gas temperature to drop so that the low-pressure gas leaving the turbine can accept a refrigeration load before passing through the cold-side of the recuperative heat exchanger and into the low-pressure buffer. The low-pressure buffer drains back into the pulse tube system when the low-pressure check valve is activated. The volume in the compression space is assumed to vary sinusoidally according to:
where is the total swept volume and is the crank angle. A mass balance on the compression space and the dead volume in the system yields:
Eq. (2) implies that the temperature of the gas contained in the compression space and the dead volume does not change, and that the fluid behaves as an ideal gas with a gas constant R. In Eq. (2), p is pressure, is the mass flow into the cold end of the pulse tube, and
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and are the mass flow rates through the high- and low-pressure check-valves, respectively. Mass balances on the two buffer volumes and the reservoir lead to:
where and are the buffer and reservoir volumes, and are the mass flow rates through the orifice and turbine, and and are the pressures in the high-pressure buffer, low-pressure buffer, and reservoir. In Eq.'s (3) through (5), it is assumed that the temperature within each of these volumes is constant, with the buffer volumes always at the intermediate expansion temperature and the reservoir remaining at the compression space temperature Following the method of other first order pulse tube analyses (e.g. [6]), the tube is divided into three sections. The "expansion space" consists of gas that enters and leaves the pulse tube at the cold end during each cycle; this space is analogous to the expansion volume swept out by a physical piston in a Stirling cycle. The "hot space" consists of gas that enters and leaves the pulse tube at the hot end during each cycle. These two volumes are assumed to be isothermal and are separated by an adiabatic column of gas that acts as the equivalent of a mechanical piston, transferring work from the expansion space to the hot space where it is dissipated in the orifice. These idealizations, together with the assumption that the volume of the adiabatic
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section ( which varies during a cycle) is approximately equal to the total pulse tube volume allow the following equation to be written [6]:
where is the ratio of the specific heat capacities of the gas. Typically, Eq. (6) is rewritten in terms of phasors after making the assumption that all variations are small and sinusoidal. For the hybrid system we don’t have this luxury as the non-linearity introduced by the check valves prevents pure sinusoidal behavior. The orifice is modeled as a linear fluid resistor
A linear resistance relationship is also assumed for flow through the check valves:
where and are the high- and low-pressure check valve resistances, which depend on the direction of the pressure gradient according to:
where is the resistance to flow in the forward direction and is the resistance associated with back-leakage. The flow characteristics of a turbine are very similar to those of a nozzle [7]:
where is the load temperature for the recuperative stage, is the specific heat capacity at constant pressure, and is an equivalent area that characterizes flow through the turbine. After substituting Eq.'s (7) through (12) into the mass balances given by Eq.'s (3) through (6), we obtain four differential equations that govern the behavior of the system. These differential equations are given below in dimensionless form:
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The dimensionless groups used in Eq.'s (13) through (16) have been selected in order to logically characterize various aspects of the system. The dimensionless pressure is the ratio of the instantaneous pressure to the initial charge pressure the pressure in the system when the compressor volume is largest and all of the gas in the system is in mechanical equilibrium.
The operating conditions are characterized by the ratio of the compression to expansion temperatures and the ratio of the expansion to cold load temperatures
The pulse tube system is characterized by the dimensionless orifice resistance dimensionless pulse tube volume dimensionless reservoir volume and the reduced dead volume (S):
where is cycle period. The rectification system is characterized by the dimensionless check valve forward and leakage resistances ( and ) and the dimensionless buffer volume
The recuperative system is characterized by the dimensionless turbine flow area
The four coupled ordinary differential equations are stepped forward in time using a RungaKutta 4th order technique, starting from the initial charge pressure (i.e. ) and continuing until a cyclic steady-state is achieved. Cyclic steady state is defined as a
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condition where the maximum absolute cyclic change in any of the dimensionless pressures is less than 0.0001. The model is implemented in the mathematical software EES - Engineering Equation Solver [8]. The model described above predicts the variation in the four system pressures over a steady state cycle; these pressure waves must be subsequently used to calculate energy transfers within the system. In the limit of a perfect regenerator, the work transfer into compression space is equal to the heat rejected from the compression space These energy flows are made dimensionless and calculated according to:
In the limit of a perfect recuperator, the work produced by the turbine is equal to the refrigeration extracted from the cold reservoir These energy flows are made dimensionless and calculated according to:
In the limit of a perfect regenerator, the work done by the “expansion space” on the adiabatic section of gas is equivalent to the refrigeration provided at the intermediate temperature These energy flows are made dimensionless and calculated according to:
where
is the dimensionless mass of gas in the “expansion space”:
Although the recuperator and regenerator are both modeled as perfect, it is important to determine the total heat load per cycle in these heat exchangers ( and ) as this provides an indication of their size and is therefore a useful means of cycle optimization. These dimensionless heat transfers ( and ) can be calculated from:
where
is the dimensionless mass of gas that passes through the turbine during each cycle.
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The model can be verified in the pure pulse tube limit by “shutting” the check valves (i.e. by setting Figure 2 illustrates the predicted Carnot efficiency as a function of the number of integration steps (n) used to step through a cycle. Kittel [9] showed that the Carnot efficiency of an ideal pulse tube should approach a value of The numerical model also predicts this limit, provided the number of integration steps is sufficiently large. The error between the theoretical and predicted Carnot efficiency is also shown in Figure 2. All subsequent results are obtained with at least 100 integration steps.
MODEL RESULTS The model presented here is very idealized. The expansion and compression spaces are isothermal, the regenerator and recuperator are perfect, and the turbine is isentropic. Therefore, any results obtained with this model are of limited quantitative value. Rather, the utility of the model lies in its ability to qualitatively illustrate how the regenerative and recuperative systems interact and to indicate the dimensionless groups that control different aspects of the system’s behavior. These parameters are related to the operating temperatures ( and ), the working fluid the pulse tube ( and S), the rectification system ( and ), and the reverse Brayton system The performance of the overall system can be expressed in terms of these dimensionless groups:
In this section, we will use the model to examine the effect of the rectification components on the system’s performance and also illustrate how the first order model may be used to select operating parameters that maximize refrigeration power in a meaningful way. The first order model provides a solid starting point for future work; some of the assumptions will be relaxed to create a more predictive tool and other losses can be individually post-calculated and then applied to the first order results; an approach sometimes called de-coupled second order analysis that has been used extensively for Stirling machine development [10]. Figure 3 illustrates the pressures within the hybrid cycle as a function of crank angle. Also shown in Fig. 3 is the pressure in a similar, pure pulse tube system (i.e. one without a
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recuperative system attached, The main effect of the hybrid components is to reduce the overall amplitude of the pressure wave within the regenerative system. The pressure wave also becomes non-sinusoidal as a result of the discontinuous behavior of the check valves. The rectification components (check valves and buffer volumes) are responsible for taking the oscillating flow within the pulse tube and turning a portion of it into a continuous flow that can drive the turbine; this is represented by the pressure difference between the high- and lowpressure buffer volumes in Fig. 3. A “good” rectification system will provide the turbine with a nearly time-invariant driving pressure difference while not introducing significant losses due to pressure drop through the check valves. The ideal rectification system would therefore have check valves that are perfect and infinitely large buffer volumes As the buffer volumes become finite, the pressures in the recuperative stage vary in response to the check valves opening and closing. The turbine power is adversely affected, both in terms of magnitude and stability. The ratio characterizes the buffer volume relative to the flow required by the turbine. When becomes small then the turbine tends to run in two short spurts per cycle; corresponding to a filling and a draining process. This cyclic turbine power fluctuation is undesirable for several reasons. The turbine will be forced to operate off of its design point during a large fraction of each cycle and may also need to be over-designed rotordynamically in order to accommodate over-speed events. Figure 4 illustrates the turbine power fluctuation, the average turbine power, and the normalized power fluctuation as a function of A practical system will have a buffer volume that yields of at least 10. The effect of a finite check valve resistance is to introduce losses associated with pressure drop in the rectification process. The product is proportional to the ratio of the pressure drop across the valves to the pressure drop across the turbine. Figure 5 illustrates the Carnot efficiency of the hybrid cycle as a function of for several values of the dimensionless turbine size, At very high values of check valve resistance, the efficiency limits to the pure pulse tube value described earlier, regardless of turbine size. At very low values of check valve resistance the efficiency limits to a value characteristic of the hybrid system, reflecting losses due to heat transfer in the refrigeration load heat exchanger and the dissipation of the work transfer from the expansion space. Intermediate values of check valve resistance result in larger losses as significant flow passes through the resistive check valves. The effect of the check valve resistance is dependent upon the size of the turbine; however it appears that an efficient system must have a value of that is less than 0.001.
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The first order model allows an initial optimization of the hybrid system. The typical optimization target for a Stirling system is refrigeration power per mass of gas [5], a metric that is proportional to refrigeration power per regenerator heat transfer. In the hybrid system, much of the gas is contained in the buffer volumes and does not directly participate in the refrigeration process. As a result, this metric is not as meaningful and instead we will maximize the refrigeration power at the cold load temperature per unit of heat transfer in the regenerator and recuperator, Equation (35) indicates that the functional dependence of this quantity can be expressed in terms of 10 dimensionless groups. The temperature ratios between the two stages ( and ) are fixed by the particular application under consideration and is fixed by the working fluid composition. Other parameters such as reduced dead volume (S), the dimensionless check valve resistances ( and ), and the reservoir and buffer volume ratios ( and ) are important to the system’s performance but have clear optimal values at a limit (either zero or approaching infinity) and are therefore not “design” parameters. The dimensionless size of the turbine must be adjusted to yield the ratio of mid- to lowtemperature refrigeration loads that is required by the application. The remaining two parameters are the pulse tube volume ratio and the dimensionless orifice resistance The adiabatic section of gas in the pulse tube plays the role of a displacer in a Stirling machine. However the gas column is compliant and therefore its presence has an adverse effect on the amplitude of the pressure variation. Any optimization procedure tends to move towards a pulse tube volume that is just large enough to accommodate the “expansion space” and the “hot space”. The second order losses such as shuttle heat transfer and conduction that promotes a finite pulse tube volume are not included in this first order model. However, we can predict the size of the pulse tube required in this “zero-adiabatic space” limit. By multiplying this minimum volume by an appropriate factor, between 3 and 10 depending on the temperature and power level, it is possible to obtain a realistic pulse tube design. There is an optimal value of the dimensionless orifice resistance, a characteristic of pulse tube systems that has been documented (e.g. in [11]) and is related to the obtaining the proper phase between mass flow rate and pressure at the cold end of the pulse tube. Figure 6 illustrates the variation in the refrigeration load per heat exchanger load, as a function of the dimensionless orifice resistance, for several values of the ratio between the mid- and low-temperature refrigeration loads, In these curves, the pulse tube volume ratio, has been set to three times the “expansion space” volume ratio, as described above. As becomes smaller, the size of the pulse tube is reduced – the compressor is pushing more flow into the recuperative stage and less into the “expansion space”. The optimal dimensionless orifice resistance will
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increase as the pulse tube volume is decreased. This is clearly shown in Figure 7, which illustrates the optimal dimensionless orifice resistance and the associated ratio of heat lifted to heat exchanger load. Figure 8 illustrates the dimensionless pulse tube size and the dimensionless turbine size at the optimal conditions. Notice that the pulse tube becomes smaller, in proportion to the compressor swept volume, as more of the load is taken in the recuperative system; that is as goes down so does This causes the aforementioned increase in the orifice resistance. The size of the turbine, as characterized by tends to increase as in order to supply the required cold refrigeration power.
SUMMARY This paper describes a concept for a hybrid cryogenic refrigeration system combining a regenerative upper stage with a recuperative lower stage. The configuration proposed here uses a simple rectification system consisting of two, oppositely-oriented check valves with buffer volumes to convert the oscillating pressure wave in the regenerative system into a steady flow suitable for a recuperative system. The specific configuration analyzed here is a pulse tube regenerative stage interfaced with a reverse-Brayton recuperative stage. This configuration can potentially provide refrigeration with high efficiency at temperatures below that of a pure pulse tube system. The first order model presented in this paper allows us to identify the important dimensionless groups from the governing differential equations. These various parameters characterize different aspects of the hybrid system and include the operating temperature ratios ( and ), the dimensionless dead volume (S), the ratio of the specific heat capacities of the working fluid the reservoir and pulse tube volume ratios and dimensionless orifice resistance ( and ), the dimensionless check valve forward and leakage resistances and buffer volume ratio ( and ), and the dimensionless turbine size The model is idealized in that it assumes isothermal compression and expansion, perfect recuperation and regeneration, and an isentropic turbine. However, the model is useful in that it allows a qualitative investigation of the effect of the rectification system characteristics. It is shown that an ideal rectification system consists of infinitely large buffer volumes and perfect check valves. The effect of finite buffer volumes is to cause the turbine power to fluctuate during each cycle. The effect of finite check valve resistances is to reduce the efficiency of the cycle. The model is used to optimize the hybrid system by maximizing the ratio of the heat lifted from the cold refrigeration space to the heat transfer in the heat exchangers.
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Our intention is to implement several systematic improvements to this first order model including non-isothermal expansion and compression. The resulting first order model will then be corrected for losses that include imperfect regeneration and recuperation, turbine inefficiency, shuttle heat transfer, turbine transient effects, leakage, and others. In the pure pulse tube limit, the model results will be compared to the predictions of well known pulse tube models such as ARCOPTR [12] in order to validate its results. In parallel with this modeling effort we are developing a proof-of-concept refrigerator by modifying a single-stage pulse tube through the addition of a rectification system and, eventually, the implementation of a recuperative stage. The experimental results will allow verification and improvement of the model.
ACKNOWLEDGMENT This support of Atlas Scientific is gratefully acknowledged.
REFERENCES 1.
Levenduski, R., Gully, W., and Lester, J., 1999, “Hybrid 10K Cryocooler for Space Applications”, Cryocoolers 10, Kluwer Academic/Plenum Publishers, pp. 505-511.
2.
Bradshaw, T. W., Orlowska, A. H., and Jewell, C., 1999, “Life Test and Performance Testing of a 4K Cooler for Space Applications”, Cryocoolers 10, Kluwer Academic/Plenum Publishers, pp. 521-528.
3.
Crunkleton, J. A., 1993, “A New Configuration for Small-Capacity Liquid-Helium-Temperature Cryocoolers”, Proceedings of the Seventh International Cryocooler Conference, A. F. Report PLCP-93-1001, p. 187. Nellis, G. F., and Maddocks, J. R., "An Isothermal Model of a Hybrid Stirling/Reverse-Brayton Cryocooler", submitted to Cryogenics, May, 2002.
4. 5.
Walker, G., 1980, Stirling Engines, Clarendon Press, Oxford, pp. 50-58.
6.
Radebaugh, R., 1990, "A Review of Pulse Tube Refrigeration", Advances in Cryogenic Engineering, Vol. 35, Plenum Press, New York, pp. 1191-1205.
7.
Cohen, H., Rogers, G. F. C., and Saravanamuttoo, H. I. H, 1987, Gas Turbine Theory, 3rd Edition, Longman Scientific & Technical Publishing, New York.
8.
Klein, S. A., and Alvarado, F. L., 2002, "EES-Engineering Equation Software", F-Chart Software, http://www.fchart.com.
9.
Kittel, P., “Ideal Orifice Pulse Tube Performance”, Cryogenics, Vol. 32, No. 9, (1992), pp. 843.
10.
Walker, G., Fauvel, O. R., Reader, G., and Bingham, E. R., 1994, The Stirling Alternative: Power Systems, Refrigerants, and Heat Pumps, Gordon and Breach Science Publishers, Switzerland, pp. 59-69.
11.
Neveu, P., and Babo, C., 2000, “A Simplified Model for Pulse Tube Refrigeration”, Cryogenics, Vol. 40, pp. 191-201.
12.
Roach, P. R., and Kashani, A., 1996, “A Simple Modeling Program for Orifice Pulse Tube Coolers”, Cryocoolers 9.
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The Role of the Orifice and the Secondary Bypass in a Miniature Pulse Tube Cryocooler Y.K. Hou, Y.L. Ju, W. Jing and J.T. Liang Cryogenic Laboratory, Chinese Academy of Sciences Beijing 100080, P.R.CHINA
ABSTRACT This paper focuses on the phase shifting role of the orifice and the secondary bypass in a miniature pulse tube cryocooler. Firstly, the equation of the mass flow rate through the valve is extended into the Fourier series to investigate the phase difference between dynamic pressure in the pulse tube and mass flow rate at the hot end of the pulse tube. The analytical results show that the orifice opening has weak effect on the phase difference between the pressure and the mass flow rate at hot end when the secondary bypass is closed. And the mass flow rate at the hot end of the pulse tube is almost in phase with the dynamic pressure in the pulse tube. The introduction of the secondary bypass will make the pressure in the pulse tube always lead the mass flow rate at the hot end. For the orifice pulse tube cryocooler, the experiments show that with the orifice opening increasing, the phase difference is increasing but less than 3 degree for the opening of the orifice within the range from 0.2 to 2 turns. Based on the optimum orifice opening, the phase difference between the pressure in the pulse tube and the mass flow rate at the hot end increases as the opening of the secondary bypass increases. The experimental result agrees with the analytical result qualitatively.
INTRODUCTION According to the enthalpy flow theory1, the phase difference between the dynamic pressure in the pulse tube and the mass flow rate at the cold end is an important parameter for the performance of the cooler. It is considered at first that the phase difference close to zero will lead to better cooling performance. However, the optimum phase relationship is to have the mass flow rate lags the pressure wave at the cold end2 when the losses in the regenerator are taken into account. For the pulse tube cryocoolers, this phase difference is mainly determined by the phase shift mechanisms such as orifice, reservoir and secondary bypass at the hot end of the pulse tube. Fully understanding of the phase relationship between the dynamic pressure in the pulse tube and the mass flow rate at the hot end is helpful. Cai3 et al investigated the effect of the opening of the orifice and the secondary bypass on the phase difference between the pressure wave and mass flow rate at the hot end of the pulse tube early in 1993, they used the hot wire anemometer and the quartz pressure transducer to measure simultaneously the mass flow rate and the pressure, respectively. When the secondary bypass was closed, their experimental results showed that the dynamic pressure wave lagged the mass flow rate at the hot end from 85 degree to 48 degree with the opening of the orifice
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Figure 1. Experimental setup.
increasing. Inada4 et al also investigated the change of the phase difference between the pressure and mass flow rate at the hot end by using the hot wire anemometer and the pressure transducer in 1996. For the orifice pulse tube cryocooler, they found that the phase difference was almost independent of the opening of the orifice, and the mass flow rate was almost in phase with the pressure wave within the operating conditions discussed in their paper. The difference of their experimental setups is that Cai placed the hot wire anemometer before the orifice and Inada placed it after the orifice, and the geometry of their cryocooler is also different. In this paper, the analytical method similar to what Kuriyama5 applied is used to investigate the phase shifting role of the orifice and the secondary bypass in a high frequency miniature pulse tube cryocooler, and the experiments are carried out to confirm the analytical results.
EXPERIMENTAL SETUP The experimental setup is shown in Fig. 1, which consists of a pulse tube cryocooler, a vacuum system and a measuring system. The pulse tube cryocooler is arranged in the “U” shape. The compressor is of a rotary type with a constant swept volume of 1.66cm3. The regenerator has an outer diameter of 8mm and a length of 60mm, the pulse tube has an outer diameter of 5mm and a length of 70mm. The reservoir has the volume of 55cm3. Two identical fine needle valves are used as the orifice and secondary bypass. They are placed outside of the vacuum chamber for convenience. The needle valve has a total of 10 turns, with each turn being graded in 50 divisions. The vacuum environment is maintained by a vacuum pump. The lowest vacuum of 1 .0Pa can be reached. Temperature is measured at the cold end of the pulse tube with a copperconstantan thermocouple (Type T). Dynamic pressures are measured by small quartz pressure transducers (Kistler, type 601 A) at the inlet of the regenerator, at the hot end of the pulse tube and in the reservoir. The pressure voltage signs are amplified by charge amplifiers (Kistler, type 5011) and collected by the computer. During the experiment, the constant room temperature is obtained by the air conditioning.
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METHOD TO DETERMINE THE MASS FLOW RATE AT THE HOT END6 Fig. 2 is the schematic diagram of a double-inlet pulse tube cryocooler. It becomes the orifice pulse tube cryocooler when the secondary bypass is closed. Neglecting the void volume of the connecting tube and the hot end heat exchanger, the mass flow rate at the hot end of the pulse tube is given by
where is the mass flow rate at the hot end; is the mass flow rate through the orifice and is the mass flow rate through the secondary bypass. Generally, needle valves are used as the orifice and the secondary bypass. If the inertial effect is neglected and the temperature is constant when gas flows through the valve, the mass flow rate through the valve can be expressed as5:
where k is the proportionality constant; and are the pressures before and after the valve, respectively. In Eq. (2), the proportionality constant k is dependent on the opening of the valve and the
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pressure ratio, which is when is greater than or equal to 1 or when is less than 1. It will be determined based on the pressure measurements as follows. According to mass conservation, the mass flow rate through the orifice valve equals to the change of mass in the reservoir, which is6
where subscript “ r ” refers to the reservoir. Assuming that the gas oscillating in the reservoir is adiabatic, we have
where is the volume of the reservoir; is the transient pressure in the reservoir; R is the ideal gas constant; is the ratio of specific heat capacities and is the room temperature. By measuring the transient pressure in the reservoir, the mass flow rate from Eq. (4) can be obtained. With this mass flow rate, the values of k can be determined at different openings from Eq. (2). Because the orifice and the secondary bypass are identical valves, the values k obtained by the orifice valve can be available for the secondary bypass valve. Fig. 3 shows the value of k for the valve with opening of 0.8 turns. The values of k for other valve openings are also obtained.
ROLE OF THE ORIFICE For simplicity, some basic assumptions are made as follows: The pressure oscillation in the pulse tube cryocooler is sinusoidal. There is no pressure drop in the pulse tube. The initial phase angle of the pressure in the pulse tube is zero. The pressure at the inlet of the regenerator (i.e. before the secondary bypass) can be expressed as
where is the average pressure; is the amplitude of the dynamic part of the pressure; the angular velocity and is the initial phase angle of the pressure. The pressure in the pulse tube is:
where is the amplitude of the dynamic part of the pressure in the pulse tube. The pressure in the reservoir is:
is
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where is the amplitude of the dynamic part of the pressure in the reservior and is the initial phase angle of the pressure. Substituting Eq. (6) and (7) into Eq. (2) and rearranging it gives the dimensionless mass flow rate through the orifice
where the subscript “o ” refers to the orifice. Eq. (8) shows that the amplitude and the initial phase angle of the mass flow rate through the orifice are the functions of
and
Eq. (8) is a period function which can be
extended to the Fourier series5. It is reasonable to consider that the amplitude and the initial phase angle of the mass flow rate through the orifice are mainly determined by the main term of the fundamental oscillation. Therefore the mass flow rate through the orifice can be expressed as
where a0 is the constant in the Fourier series; a1 and b1 are the coefficients of the first term in the Fourier series and is the initial phase angle, which is
The positive value of means that the mass flow rate through the orifice leads the dynamic pressure in the pulse tube. The Fourier coefficients can be calculated by numerical integral. When the secondary bypass is closed, it becomes the orifice pulse tube cryocooler. According to Eq. (1), the mass flow rate at the hot end of the pulse tube is equal to the mass flow rate through the orifice for the orifice pulse tube cryocooler. Referring to Eq. (6) and (10), it is known that the initial phase angle of the mass flow rate through the orifice represents the phase difference between the mass flow rate at the hot end and the dynamic pressure in the pulse tube. In the experiment, it is measured that the amplitude of the dynamic pressure in the reservoir is far less than that in the pulse tube because the reservoir is large enough. Table 1 shows the relationship between the initial phase angle and the ratio of
for three cases. Case 1, 2, 3
represent the small, optimum and large opening of the orifice, respectively. The values of for every case are taken from the experimental data. The calculations are performed under the condition of average pressure of 2.2MPa and frequency of 50Hz. Within the orifice opening
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range (less than or equal to 2 turns) considered in the experiment, the phase difference is less than 2 degree no matter how much the initial phase angle of is. The mass flow rate at the hot end of the orifice pulse tube cryocooler can also be obtained by the measured oscillating pressure in the reservoir. Fig. 4 is the experimental result by Eq. (4). In the experiment, the maximum opening of the orifice is 2 turns. With the opening of the orifice increasing, the phase difference is increasing, but is less than 3 degree. The theoretical result and the experimental result all indicate that opening of the orifice has weak effect on the phase difference, and the mass flow rate at the hot end of the pulse tube is almost in phase with the dynamic pressure in the pulse tube. The increasing of the opening of the orifice can reduce the phase difference between the mass flow rate at cold end and the pressure in the pulse tube. The mass flow rate at the cold end can be expressed as
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where is the cold end temperature; is the hot end temperature and is the volume of the pulse tube. Fig. 5 schematically shows the phase angle reducing along with the opening of the orifice increasing. With the opening of the orifice increasing, the first term in the right hand of Eq. (11) increases and the second term decreases, thereby the phase angle reduces.
ROLE OF THE SECONDARY BYPASS Similar to the analysis of the orifice, substituting Eq. (5) and (6) into Eq. (2) and rearranging it, the dimensionless mass flow rate through the secondary bypass is
where the subscript “ d ” refers to the secondary bypass. Writing Eq. (12) with a Fourier series and taking the fundamental term as approximation gives
where
is the initial phase angle, which is
The positive value of means that the mass flow rate through the secondary bypass leads the dynamic pressure in the pulse tube. Referring to the calculation in the prior section, the initial phase angle of the mass flow rate through the secondary bypass is evaluated. Table 2 shows the calculated results for three cases.
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Case 1, 2, 3 represent the small, optimum and large opening of the secondary bypass, respectively. The calculations are performed on the basis of the orifice set to its optimum opening. The calculations show that the initial phase angle of the mass flow rate through the secondary bypass strongly depends on the value of
and
cryocooler considered, the experiments show that the value of increasing of the secondary bypass reduces the value of
For the miniature pulse tube is less than 30 degree; the
but increases the value of
Table 2 also shows that the initial phase angle of the mass flow rate through the secondary bypass is always positive and can be as large as about 90 degree along with the opening of the secondary bypass increasing. From Eq. (1), the mass flow rate at the hot end of the pulse tube is determined not only by the phase angle of and but also by their amplitude. As shown in Fig. 6, the positive value of the initial phase angle of the mass flow rate through the secondary bypass makes the pressure in the pulse tube lead the mass flow rate at the hot end. With the determined proportionality constant k and the measured dynamic pressures at the inlet of the regenerator, in the pulse tube and in the reservoir, the mass flow rate at the hot end of the double-inlet pulse tube cryocooler can be obtained by Eq. (2). The experiment is carried out under the conditions of average pressure of 2.2MPa, frequency of 50Hz and the orifice being set
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to its optimum opening. Fig. 7 shows the relationship between the phase difference (pressure leads the mass flow rate) and the opening of the secondary bypass. With the opening of the secondary bypass increasing, the phase difference between the pressure in the pulse tube and the mass flow rate at the hot end increases. When the opening is larger than 1.2 turns, the slope of the curve is gentle.
CONCLUSION The role of the orifice and the secondary bypass is theoretical and experimental investigated. The mass flow rate through the orifice and the secondary bypass are approximately expressed with the fundamental term of the Fourier series. For the orifice pulse tube cryocooler, calculations and experiments show that the mass flow rate at the hot end of the pulse tube is almost in phase with the dynamic pressure in the pulse tube, and the opening of the orifice has weak effect on the phase angle under the condition that the opening of the orifice is less than 2 turns. Theoretical calculation shows that the initial phase angle of the mass flow rate through the secondary bypass is always positive and can be as large as about 90 degree along with the opening of the secondary bypass increasing, which makes the pressure in the pulse tube always leads the mass flow rate at the hot end for the double-inlet pulse tube cryocooler. With the opening of the secondary bypass increasing, experimental result shows that the phase difference between the pressure in the pulse tube and the mass flow rate at the hot end increases.
ACKNOWLEDGMENT This work is supported by the National Natural Science Foundation of China (Grant No. 50176052).
REFERENCES 1. Radebaugh R., “Advances in cryogenics”, Proc. ICEC16/ICMC, Elsevier Science, Oxford (1997), pp.33-44 2. Radebaugh R., “Pulse Tube Cryocoolers for Cooling Infrared Sensors,” Proceedings of SPIE Vol.4130, edited by Bjom F.Andresen, et al, 2000, pp.363-379. 3. Cai J.H., Zhou Y., Wang J.J., Zhu W.X., “Experimental analysis of the double-inlet principle in pulse tube refrigerators”, Cryogenics, vol.33, No.5 (1993), pp.522-525. 4. Inada T., Nishio S., Ohtani Y. et al., “Experimental investigation on the role of orifice and bypass valves in double-inlet pulse tube refrigerators”, Adv. Cry, Eng., Vol. 41, Plenum Press, New York( 1996), pp. 1479-1486. 5. Kuriyama F. and Radebaugh R., “Analysis of mass and energy flow rates in an orifice pulse-tube refrigerator”, Cryogenics, vol.39 (1999), pp.85-92. 6. Lu Guoqiang, Cheng Ping, “Flow characteristics of a metering valve in a pulse tube refrigerator”, Cryogenics vol.40 (2000), pp.721-727.
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Surface Heat Pumping Loss in a Pulse Tube Refrigerator J. Jung and S. Jeong Department of Mechanical Engineering, Korea Advanced Institute of Science and Technology Taejon, Korea
ABSTRACT Surface heat pumping in pulse tube refrigerator is caused by oscillating heat transfer between the tube wall and the inside gas of the pulse tube. It is equivalent to shuttle heat transfer in GM or Stirling cryocooler and, therefore, acts as loss. In this paper, the surface heat pumping effect is analyzed for idealized orifice pulse tube refrigerator. The linearized fluid equations without thermal diffusion effect and with thermal diffusion effect in the pulse tube were solved analytically. From the solutions the time-averaged enthalpy flows were calculated for both cases. Since the thermal diffusion effect causes the difference in the enthalpy flows, the difference is equal to the net effect of the thermal diffusion effect, or equivalently, surface heat pumping effect. The surface heat pumping effect is heavily dependent on the pulse tube geometry and the operating parameters. It is not negligible in small-size or low frequency pulse tube refrigerators according to the analysis. To reduce this loss, the pulse tube is suggested to have small thermal mass so that the wall temperature can follow the bulk temperature.
INTRODUCTION There have been extensive investigations on how the bulk motion of gas under pulsating pressure in pulse tube generates refrigeration effect in orifice pulse tube refrigerators.1, 2 Gas flow work to lift thermal energy from the cold end of the pulse tube is dissipated at the orifice. The subsequent enthalpy flow along the pulse tube lifts heat from the cold end toward the hot end. The refrigeration capacity would be equal to the work loss at the orifice if the regenerator is perfect and if the near-wall gas effect (gas-to-wall heat transfer effect and viscous effect) is neglibible. However, in actual orifice pulse tube refrigerators the regenerator is not perfect and the near-wall gas effect is not negligible as shown in Fig 1. The refrigeration capacity is, therefore, as follows.
is the work loss, is the regenerator loss and is the loss from the near-wall gas effect. The near-wall gas effect can be divided into two sub-effects.3 One is thermal diffusion effect in the gas near the wall.3,4 The thermal diffusion effect has different names. One of them is surface heat pumping effect.5,6 The oscillatory gas-to-wall heat transfer causes heat flow from hot end toward cold end. The other of the two sub-effects is steady Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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secondary streaming in the pulse tube.3, 7, 8 Viscous effect generates the steady secondary streaming and the corresponding loss in the pulse tube. The loss from near-wall gas effect is the sum of the losses from thermal diffusion effect and from the secondary streaming.
Although such loss mechanisms have been qualitatively discussed and experimental measurement of such loss has been performed10, quantitative evaluation of these effects has not been well studied. In this paper, the thermal diffusion effect of gas in pulse tube is focused on for the analysis. We quantify the surface heat pumping effect (thermal diffusion effect) analytically and investigate the effect of non-dimensional parameters on refrigeration capacity.
ADIABATIC FLOW MODEL OF PULSE TUBE First, the pulse tube of orifice pulse tube refrigerator is analyzed excluding the thermal diffusion effect.
Flow field A spatial coordinate system of the pulse tube is set as shown in Fig. 2. The position x=0 is the hot end of the pulse tube and x=L is the cold end. The tube wall is at r=R. The pressure inside the pulse tube, which is a function of time only, is assumed to vary sinusoidally and is represented by or simply, The bold character denotes complex variable. An orifice and a gas reservoir are connected to the hot end and control the hot end flow. If we assume that the pressure in the gas reservoir is constant as and that the orifice is a linear element although they may not be in actual refrigerator, we can represent the hot end flow as Here is the flow resistance of the orifice. The continuity equation, the energy equation and the state equation of ideal gas are as follows.
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The algebraic manipulation of the continuity equation and the energy equations with the state equation of ideal gas results in Eq. (4) on condition that and
If we have one-dimensional flow assumption and linearize Eq. (4), the following Eq. (5) is obtained.
With the boundary condition that
at x=0, the solution is as follows.
Here, is and is the dimensionless axial position, x/L. The velocity u(x,t) is comprised of two components. One, which is represented by is the velocity component induced by the pressure change inside the pulse tube. The amplitude depends on the axial position in the pulse tube. The other component, which is represented by is solely generated by the orifice and is independent of the axial position. The velocity component of – is in 90 degree phase lag with the pressure and, consequently, does not generate any net enthalpy flow. The velocity component of is in phase with the pressure and generates definite time-averaged enthalpy flow in the axial direction. In orifice pulse tube refrigerator, can be varied only by the change of operating parameters of the compressor. On the contrary, can be varied by adjustment of the flow resistance of the orifice without changing operating parameters of the compressor.
Temperature The sinusoidally pulsating pressure in the pulse tube causes subsequently pulsating temperature of gas. The pulsating temperature can be represented by the sum of the mean temperature and the temperature fluctuation, If we assume no thermal diffusion, is just The oscillation effect is only considered as The energy equation is, therefore, as follows.
The solution of the energy equation is obtained as follows.
Enthalpy flow The time-averaged enthalpy flow in the pulse tube under pulsating pressure is calculated as follows.
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The minus sign denotes the direction from the cold toward the hot end by the definition of the coordinate system of Fig. 2. The enthalpy flow is independent of the cold end temperature and is proportional to the product of pressure amplitude and hot end flow This is consistent with other analysis result.9
THERMAL DIFFUSION EFFECT The thermal diffusion in the gas alters the near-wall gas temperature from the bulk temperature. The enthalpy flow is altered accordingly. Temperature The pulsating temperature T(x,r,t) should be now represented by thermal inertia of the tube wall is much larger than that of the near-wall gas, r=R. The energy equation is written as follows.
Since the is satisfied at
The solution of the energy equation is obtained analytically with Bessel function as follows.
Here, is and designates the degree of thermal diffusion effect. This nondimensional parameter is very important to account for surface heat pumping effect. Large thermal diffusivity of gas results in small The value of is approximately unity except in the vicinity of r=R. The center gas in the pulse tube behaves as if no thermal diffusion effect takes place as we expected.
Enthalpy flow The time-averaged enthalpy flow is calculated as follows.
With the approximation that evaluated as the following equation.
for
the enthalpy flow of Eq. (12) is
Physically, the enthalpy flow should remain invariant with respect to the axial position in the pulse tube because there would be no time-averaged gas-to-wall heat transfer. In Eq. (13), can be represented by and, since can be represented by can be represented by As a result, the following equation relating and is obtained.
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Let and is the ratio of the enthalpy flow with thermal diffusion effect to that without thermal diffusion effect. Eq. (14) is converted into the following form.
The solution of the differential equation is as follows.
Here M is By applying the cold end condition, we can obtain the relation between the dimensionless refrigeration effect X and the cold end temperature with other operating parameters.
Although the amount of the work loss at the orifice is equal to Eq. (9), the enthalpy flow is not. The thermal diffusion effect reduces the enthalpy flow by factor of X of Eq. (17). In other words, 1–X of the work loss is equal to the surface heat pumping loss.
RESULTS AND DISCUSSION The surface heat pumping effect has been analytically found in the previous section. We will examine the impact of and on X in this section.
Value of The values of
which indicates the degree of the thermal diffusion, are calculated for
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several conditions of the pulse tube with helium gas in Table 1. The pulse tubes with small diameter and low operating frequency have low and the pulse tubes with large diameter and high operating frequency have high Practically, is between 30 and 40 in small-size Stirling type pulse tube refrigerator and is between 10 and 20 in medium-size GM type pulse tube refrigerator.
Dimensionless refrigeration effect, X X is plotted in Fig. 3 and 4 to see the effect of M and Fig. 3 is the calculation result of X with respect to M for and and 40. X is highly dependent on For small X is small. In other words, the surface heat pumping loss is considerable for small pulse tube or low frequency pulse tube refrigerator. If increases, however, X approaches to 1. The surface heat pumping loss in this case is not significant. As M increases, X always decreases monotonically. The surface heat pumping loss thus increases. Fig. 4 is the calculation result of X with respect to M for and 80/300, 100/300 and 120/300. As the cold end temperature of pulse tube, decreases, the surface heat pumping loss increases. Even though Fig. 3 does not show the calculation result of high above 120 K, it is true that X becomes 1 or slightly larger than 1 when M is very small and is close to The enthalpy flow is actually enhanced by the thermal diffusion effect, which is the main characteristic of basic pulse tube refrigerator. Experimental result of the ratio of the actual enthalpy flow to the work loss is available in other literature.10 Although the direct comparison of the ratio and X is not possible for lack of the information to evaluate M and we can find out that both have the similar trends
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with respect to the operating frequency. The ratio acquired by the experiment and X in this analysis increase as the operating frequency w increases. Although the value of X can be calculated algebraically from Eq. (17) for small M, the analysis is not quite correct for small M. The axial temperature profile of the pulse tube with and is plotted for M=0.1, 0.5 and 1.0 in Fig. 5. It displays a steep temperature gradient and deflection point in the region near the hot end for M=0.1. It is evident that the thermal diffusion in axial direction in this case is not trivial. The linearization of the energy equation would generate severe errors.
SUMMARY The simple relation of evaluating the surface heat pumping effect in pulse tube is analytically derived. The surface heat pumping loss is readily calculated from the relation and the adiabatic work loss at the orifice. The calculation is possible for various system parameters such as pulse tube geometry, the operating frequency, the work fluid properties and etc. in convenient non-dimensional forms. According to the analysis in this paper, the surface heat pumping loss is not negligible in very small-size or low frequency orifice pulse tube refrigerators.
ACKNOWLEDGEMNTS This research was supported by the Combustion Engineering Research Center (CERC), Korea Science and Engineering Foundation (KOSEF) and the Brain Korea 21 Project.
NOMENCLATURE H h k M P Q
R r T u W X
Specific heat at constant pressure (J/kg.K) Time-averaged enthalpy flow (W) Enthalpy (J/kg) Thermal conductivity (W/m.K) Ratio of to Pressure (Pa) Heat flow (W) Heat generation Radius (m) Universal gas constant (8.314 J/mol.K) Ratio of pressure amplitude to mean Radial position (m) Temperature (K) Velocity (m/s) Work (W) Dimensionless refrigeration effect, Eq. (17)
Greek letters Thermal diffusivity Specific heat ratio Density Period (sec) Angular frequency (rad/sec)
Subscripts 0 cold hot
Cross-sectional mean property Cold end property Hot end property
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REFERENCES 1.
Storch, P.J. and Radebaugh, R., “Development and experimental test of an analytical model of the orfice pulse tube refrigerator,” Advances in Cryogenic Engineering, vol. 33, Plenum Press, New York (1988), pp. 851-859.
2.
Liang, J., Ravex, A. and Rolland, P., “Study on pulse tube refrigeration Part 1: thermodynamic nonsymmetry effect,” Cryogenics, vol. 36, no. 2 (1996), pp. 87-93.
3.
Lee, J.M., Kittel, P., Timmerhaus, K.D. and Radebaugh, R., “Steady secondary momentum and enthalpy streaming in the pulse tube refrigerator,” Cryocoolers 8, Plenum Press, New York (1995), pp. 359-369.
4.
Lee, J.M., Kittel, P., Timmerhaus, K.D. and Radebaugh, R., “Simple two-dimensional corrections for one-dimensional pulse tube models,” Cryocoolers 10, Kluwer Academic/Plenum Publishers, New York (1999), pp. 351-358.
5.
Jung, J., Theoretical study on the heat pumping mechanism in pulse tube refrigerators, MS thesis, Dept. Mech. Eng., KAIST (2000), Taejon, Korea.
6.
Yang, L.W., “Shuttle loss in pulse tubes,” Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2001), pp. 353-362.
Lee, J.M., Kittel, P., Timmerhaus, K.D. and Radebaugh, R., “Higher order pulse tube modeling,” Cryocoolers 9, Plenum Press, New York (1997), pp. 345-353. 8. Jeong, E.S., “Secondary flow in basic pulse tube refrigerators,” Cryogenics, vol. 36, no. 5, pp. 317323. 9. Hofmann, A. and Pan, H., “Phase shifting in pulse tube refrigerators,” Cryogenics, vol. 39, no. 6, pp. 529-537. 10. Rawlins, W., Radebaugh, R., Bradley, P.E. and Timmerhaus, K.D., “Energy flows in an orifice pulse tube refrigerator,” Advances in Cryogenic Engineering, vol. 39, Plenum Press, New York (1994), pp. 1449-1456. 7.
Numerical Model for Pulse Tubes Using Method of Lines A. Schroth1, C. Kirkconnell2 and M. Sahimi1 1
University of Southern California Los Angeles, CA 90007
2
Raytheon Electronic Systems El Segundo, CA, 90245
ABSTRACT An accurate two-dimensional (2D) numerical model is necessary to calculate fundamentally 2D pulse tube loss mechanisms such as momentum streaming and radial gas-to-gas heat transfer. Scale analysis provides clear direction that consideration of nonlinear terms is essential for the accurate determination of the 2D pulse tube equations. The results of the scale analysis provide insight on how to solve the 2D model without having to perform mathematical “tricks” at the pulse tube ends as found with a previous nonlinear 1D model. An enhanced version of that model has been developed in which particular attention is paid to the end effects to improve the energy balance and to improve solution accuracy and convergence. Identified weaknesses in previously performed analyses are addressed with emphasis on viscosity and conductivity effects at the ends. The improved 1D model is an important and necessary step towards the next phase of the research, the solution of the 2D nonlinear pulse tube equations.
INTRODUCTION For many years, cryocoolers have been utilized for space applications to cool devices such as infrared detectors, focal planes, solid-state gamma-ray detectors, spectrometers, low noise amplifiers, superconductivity devices and other scientific devices for atmospheric monitoring and astronomy. Cryocoolers provide the necessary refrigeration such that these devices may be maintained at cryogenic temperatures for stable and effective operation. The dynamic nature of flows within the “pulse tube” itself, require a mathematical model that includes second order effects to predict the dissipative losses. To date, numerical modeling of flows within the “pulse tube” has been very limited and thus the loss mechanisms are not clearly understood, much less quantified. Works by previous investigators have concentrated their efforts on either solving a 1-D control volume formulation (Zhou, Marechal, Radebaugh), or a linearized set of 2-D equations (Lee). Kirkconnell et al. (1995) solved a 1-D set of nonlinear governing equations using the “method of lines” 1 for high frequency (~200 Hz) pulse tubes. Kirkconnell showed that the nonlinear terms are of order one which suggests that linearizing the equations for mathematical simplification is at the expense of accuracy. Subsequently, Kirkconnell (1997) performed a sensitivity analysis2 of his solution to the 1-D, nonlinear model. He found that his Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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solution was sensitive to the amplitude of the mass flow rates, the phase angle between the cold and hot mass flow rates, and compressibility. Shiraishi et al. (1997) performed flow visualization studies using a smoke-wire technique on the working gas inside the pulse tube.3 His team observed secondary flow behavior under conditions of small amplitude oscillations that resulted in large scale streaming in the pulse tube. The existence of this secondary flow behavior further emphasizes the need for 2-D mathematical to qualitatively determine the causes of pulse tube loss mechanisms. In this study, the intent is to extend the Kirkconnell’s 1-D model to include predictions for moderate frequency (~50 Hz) pulse tubes as well addressing resolution sensitivity issues. Performing these exercises should give a good understanding of how the nonlinear model responds to various model inputs. The goal of this paper is to set the foundation for the development and solution of the 2-D, nonlinear, governing equations for flows within the pulse tube portion of pulse tube refrigerators. The model presented by Kirkconnell, in which he solves an explicit set of 1-D nonlinear, timedependent system of equations (mass, momentum & energy), has been extended and modified in order to extend the experimental regime from which experimentally observed data is available. In his model, these governing equations where solved using the “method of lines” approach. In this approach, spatial approximations are resolved using cubic Hermite polynomials with collocation. The break points used for his solution were linearly spaced across the length of the pulse tube. In evaluation of Kirkconnell’s 1-D model, the sensitivity of the end effects was closely investigated since it was in these areas where he found the highest inconsistency in total energy conservation. Consequently, the model was rewritten such that these end effects could be better evaluated. This was done by first evaluating the model using higher spatial resolution, and then by staggering the nodes near the ends of the tube to investigate the effect on the overall solution. Using this model, predictions are made for a new class of pulse tube refrigerators now under investigations at Raytheon Electronic Systems. In addition, several modifications have been made to the program to allow for real-time diagnostic observation of the solution file while being solved. This allows for useful insight on how the solution method is progressing and determines which steps to implement to sustain stability and converge on a solution.
PROBLEM FORMULATION One-Dimensional Formulation The pulse tube itself is merely a cylinder of finite length and thickness. The walls are assumed adiabatic. In the 1-D model, the radial velocity and temperature gradients are assumed zero. With these simplifications, the flow within the pulse tube is modeled by using only information at the ends of the tube. The governing dimensionless equations (shown below) that describe the flow within the pulse tube are the classical mass, momentum and energy conservation equations for a transient, compressible viscous fluid. Figure 1 displays the problem formulation in terms of the independent variables that will be solved. Note the boundary conditions are time-dependent and specified at the
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entrance and exit of the pulse tube section only. The functional form of the boundary conditions was based upon experimental observation. The pressure wave, temperature, and mass flows were assumed sinusoidal, representative of the flows produced by a valveless compressor and having heat exchangers at the ends of the pulse tube. In this study, we focus purely on the flows within the pulse tube component and not the whole Pulse Tube problem in the case of a system model. The positive direction is from left to right (cold end to the hot end). This convention will be used throughout this analysis. Details regarding this problem formulation and applied boundary conditions can be found elsewhere in work previously published by Kirkconnell.2
Note, in the momentum equation, the term (described by Kirkconnell as the numerical dissipation parameter) is shown in Figure 2 which governs the magnitude of the second-order terms. The function was required for numerical stability. As mentioned earlier, the goal of this study is gain insight on how to solve the 2D model without having to perform mathematical “tricks”. The numerical dissipation term produces improved stability, which allows for the calculation of the numerical solution without introducing appreciable error.
NUMERICAL APPROACH Recent numerical experiments with the original 1-D model have revealed that the “numerical dissipation” term, required for numerical stability, results in heat conduction in the gas that is of the same order of the enthalpy flux. In reality, conduction is many orders of magnitude smaller than convection (enthalpy flux). Therefore, the model has been modified to observe if these end effects can be mitigated without sacrificing numerical stability. Finite difference spatial discretization schemes are investigated, coupled with adaptive time relaxation methods. The aim of this work is to produce a scheme that will apply beyond the onedimensional flow equations to the two-dimensional equations. Possible methods will be evaluated with these considerations in mind.
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In the finite-difference approach, the continuous problem domain is discretized so that the dependent variables are considered to exist only at discrete points. Derivatives are approximated by difference, resulting in an algebraic representation of the partial differential equation. Thus, a problem involving calculus has been transformed into an algebraic problem. The method of lines is used as the basic framework for the numerical methods developed in this work. This is a technique for decoupling the spatial and temporal discretization process. A system of partial differential equations is first semi-discretized in space, for example with a finite difference or finite element method, reducing the problem to that of a system of ordinary differential equations in time. This system can then be solved with an appropriate time integration method.
NUMERICAL RESULTS The following results obtained from the numerical model exhibit the differences and improvements made in increasing the number of spatial grids along the pulse tube axis (z*) as well as at the location near the ends of the pulse tube. The direct comparisons of improving the spatial resolution of the current 1-D pulse tube model are reflected throughout this section for density, velocity and temperature. Several considerations determine whether the solution so obtained will be a good approximation to the exact solution of the original system of partial differential equations. Among these considerations are truncation error, consistency, and stability, all of which will be discussed in this section. With regards to the density, the solutions appear similar and without any major differences, except for the location of the max/min points. Figure 3 shows the gas density versus position at nine different points within the tube. The figure clearly indicates that the peaks appear shifted in phase by about 10 degrees. Also, in Figure 3, the depth of the cold penetration is clearly evident. The average density calculations are shown bounded by the maximum and minimum
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density predictions. This linear profile of density exhibited near the cold end is consistent with temperature profile shown later in Figure 5. As for the velocity, similar shifts in phase are evident near the max/min points; however, the changes in amplitudes between the two solutions are negligible. Figure 4 shows velocity at various instants in time as a function of pulse tube position, denoted by different values of the time period over one cycle. In addition, the relative phase-difference in velocity along with their amplitudes are also shown. The temperature profiles appear similar until the max/min/average are examined. Figures 5 and 6 display the temperature as a function of position within the pulse tube. The differences between the maximum and the minimum values become very distinct. The penetration of the cold and hot gases from both sides of the pulse tubes can be clearly seen. The locations of the maximum and minimum temperatures are obtained within the interior of the tube, rather than at the heat exchangers. As compared to the 20-node case, the of the 30-node system shows a linear profile coming out of the cold heat exchanger. The phenomenon of thermal penetration can be clearly seen. The pressure profiles are shown in Figure 7, with no indication of major differences between the two cases as expected. An improvement to total energy conservation (Figure 8) was observed at the ends of the pulse tube as compared to previously results. This is a result of reducing the effect of the numerical dissipation term previously used for numerical stability. Previously normalized results showed approximately 5% error in total energy conservation. Newly modified results have lowered this error to under 1% along the axis of the pulse tube.
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EXTENSION TO TWO-DIMENSIONS Predictions made from the one-dimensional pulse tube model are limited in applicability because the effects along the radial axis are ignored. The nonlinear steady secondary transport that exists inside the pulse tube component create internal effects which include steady secondary flows which manifest themselves as mass, momentum, and energy streaming. In 1995, Lee5 at the Ames Research Center, developed a two-dimensional model to describe these steady secondary flows. His work was based on an anelastic approach, which requires the shock, and acoustic energies to be small compared with the energy needed to compress and expand the gas. In other words, his model applied to flow problems that exhibited small-amplitude fluctuations. Lee’s set of two-dimensional, axisymmetric, differential equations were linearized such that a series-expansion solution could be obtained. In addition, his set of equations were not solved simultaneously (i.e. non-coupled system of equations). Our goal in this approach is to develop a model that is applicable to real world pulse tubes where pressure fluctuations are appreciable. With this is mind, we develop this full, coupled set of partial differential equations. Solving these equations would yield the leading-order equilibrium temperature and allow a better understanding of how mass streaming affects the temperature gradient. The temperature gradients would directly determine the level of entropy generations. In addition, predictions for enthalpy flow and heat flow can be made in parametric form for pulse tube optimization. Extending the formulation to two dimensions was conducted similarly to the one-dimensional formulation where we first began with the classical transport equations that describe flow dynamics and energy (i.e. mass, momentum and energy). The equations were derived using cylindrical coordinate system.
Two-Dimensional Formulation Similarly to the One-Dimension pulse tube problem, we begin by displaying the two-dimensional system shown in Figure 9. Note the emphasis on including the conduction in the pulse tube walls as well as the variations that can occur radially across the pulse tube. The system consists of one long cylindrical tube with heat exchangers at both ends. The assumptions are as follows: 1) The working fluid is helium (inert, single component, ideal gas). 2) Two-dimensional, axisymmetric geometry is assumed. 3) Pulse tube wall has a finite thickness. 4) Outer walls of the pulse tube are adiabatic. 5) No body forces is considered 6) Thermal conductivity, dynamic viscosity and heat capacity are constant properties. 7) Continuity of temperature and heat flux at the gas/wall interface is assumed.
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8) Tube wall will be modeled using planar geometry – valid in the limit of very thin walled tube relative to the inner tube radius. 9) Linear constitutive equations, namely, Newtonian fluid and Fourier’s law of conduction are used.
The dimensionless, two-dimensional, nonlinear, governing equations that describe the flows within the pulse tube component are as follows:
where r is the distance from the center of the tube to the tube wall, and and u represent the radial and transverse velocities. Similarly, the simplified momentum and energy equations are then given by,
where Va is the Valensi number representing the ratio of viscous diffusion time-scale to velocity time-scale, Pr is the Prandtl number representing the ratio of thermal to viscous diffusion timescales, is and is As with previous analysis, the transport properties were held constant (i.e. and ) at reference temperature and pressure. Note, these equations omit contribution of flows in the azimuth direction. Reference velocity was scaled to pulse tube length and operational frequency. After applying the same set of baseline parameters used by Kirkconnell2, we observe that some of the second-order, nonlinear terms, are of significant size as shown below in the table below. The scaled equations contain a variety of dimensionless variables of various orders of magnitudes (smallest values omitted). Table 1 lists the applied baseline reference parameters and scaled parameters. These equations are nonlinear involving products of independent variables and their derivatives. In addition, we have non-homogeneous terms that are functions of independent variables. All the governing equations contain a term that has first-order time dependence in time. These equations are first-order in time and second-order in r and z. In order to solve these equations it will require at least two auxiliary conditions for each ofthese second-order independent variables. These set of equations are suited to be solved using the method of lines. The method of lines has been successfully applied to convection-diffusion problems and seems very promising to obtain numerical solutions of systems with (stiff or non-stiff) source terms.
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Using this approach, the spatial variables r and z would be discretized to obtain a semi-discrete method-of-lines scheme. This is then a system of ordinary differential equations (ODE) in the time variable that can be solved by an ODE solver. A new time discretization technique will have to be developed (similar to the one-dimensional analysis) in order to obtain numerical stability. Gottlieb et al.6 describes various approaches of applying high-order strong-stability preserving (SSP) time discretization for semi-discrete method of lines approximations of PDEs.
CONCLUSION The accuracy of a previous 1-D, nonlinear pulse tube model has been improved by carefully minimizing the effects of the numerical dissipation terms required in the momentum and energy equations for numerical stability. The method of lines solution methodology employed herein for the 1-D model is extendable to the 2-D, nonlinear model. In order to have numerical stability, a new time discretization technique will have to be developed. Pulse tube theory and flow visualization experiments validate the necessity of a 2-D model to accurately compute refrigeration losses, and the scale analysis from the legacy 1-D model and the 2-D scale analysis provided herein provide clear indication that the nonlinear terms are of leading order and must be considered.
ACKNOWLEDGMENT Financial support for this work was provided by a generous gift from Raytheon Electronic Systems Company.
REFERENCES 1. 2.
3. 4. 5. 6.
Kirkconnell, C.S., Numerical Analysis of Mass Flow and Thermal Behavior in High-Frequency pulse tubes, Ph.D. Thesis (1995), Georgia Institute of Technology. Kirkconnell, C.S., “Parametric Studies on Numerical, Nonlinear Pulse Tube Flow,” Journal of Fluids Engineering, Vol. 119, (1997), pp. 831. Shiraishi, M., Nakamura, N., Seo, K., Murakami, M., “Visualization study of oscillating flow inside a pulse tube refrigerator,” Adv. in Cryogenic Engineering, Vol. 43B, Plenum Publishing Corp., New York (1998), pp. 2023-2030. Schiesser, W.E., The Numerical Method of Lines, Academic Press, San Diego, California (1991) Lee, J.M., Steady Secondary Flows Generated by Periodic Compression and Expansion of an Ideal Gas in a Pulse Tube, NASA/TM-1999-208769, Ames Research Center, Moffett Field, California (1999). Gottlieb, S., Strong Stability Preserving High-order Time Discretization Methods, NASA/CR-2000210093, ICASE Report No. 2000-15, Langley Research Center, Hampton, Virginia (2000).
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Pulse Tube Refrigerator Analysis, Including Inertance Tube and Friction in the Regenerator Fang, L., Deng, X., and Bauwens, L. Department of Mechanical and Manufacturing Engineering University of Calgary Calgary, Alberta, Canada T2N 1N4
ABSTRACT An asymptotically consistent regenerator model has been developed, assuming arbitrarily large viscous losses in the regenerator, taken as a porous medium with arbitrary topology. The theory uses the multiple scale method. It is based upon a scaling assumption whereby fluid displacement is of the order of the regenerator length. Other assumptions include a small Mach number assumption, and that the mesh size is small compared with lengths and diameters. A specific distinguished limit (i.e., a specific relationship between these two small parameters) is required for viscous losses to be of the same magnitude as the temporal pressure amplitude. This model is then applied to the entire refrigerator, with amplitudes now restricted to small values (still in a fully asymptotically consistent manner), and with an inertance tube represented by an acoustic model. Results are obtained, including pressure and temperature profiles along the regenerator, and overall performance for various inertance tube designs.
INTRODUCTION A number of different arrangements have been considered, that provide an adequate degree of control of the phase shift between pressure and velocity in pulse-tube refrigerators. These include for instance a moving plug, a check valve, a bypass, multiple bypasses and an inertance tube. The inertance tube is placed between the pulse tube proper and the reservoir. For proper values of its length and diameter, this arrangement conceivably results in improved performance, when compared with other arrangements. This is because the inertance tube offers a good deal of flexibility, potentially resulting in oscillating velocity and pressure with a phase relationship similar to that in Stirling refrigerators. The inertance tube can be traced back to Kanao et al. [1] in 1994, who found experimentally, in a 50 Hz pulse-tube, that a long tube resulted in better performance than a needle valve. Jin [2] implemented an inertance tube in a thermoacoustically driven pulse tube refrigerator operating at 350 Hz. Starting in 1997, Zhu et al. [3] performed numerical calculations. They clearly demonstrated that an inertance tube could lead to the same equivalent PV work in the freezer as in a double inlet pulse tube for less compressor input work. Varying the diameter and length of the inertance tube, the refrigerator can be made to operate in orifice mode, in double inlet and in Stirling mode, as far as the relationship between oscillating pressure and mass is concerned. PV work and enthalpy flux predicted by simulation were found to match quite well the experimental measurements. Gardner [4] explored the effect of the inertance tube impedance, both experimentally and by using a lumped-impedance model obtained from electrical analogy. While lumped-impedance analysis provides an intuitive explanation
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for the phase shift in the inertance tube, it does not account for the dynamic and thermodynamic processes elsewhere. Duband [5] et al. again showed that the inertance tube to provide a significant gain in performance. As expected, as the frequency is increased, the optimum inertance tube length is reduced. A proper analysis of the phase relationship between pressure and velocity throughout the device requires a model including all volumes. Furthermore, if, as usual, large pressure gradients occur in the regenerator, viscous effects in the regenerator will strongly affect the pressure phase angle, as shown in previous study [6]. That study, however, assumed an arbitrary longitudinal temperature profile, when, in reality, pursuing the analysis to higher order allows for its proper determination. The current paper is thus divided into two main parts. First, the regenerator model is developed, based upon assumptions such that spatial pressure differences are of magnitude comparable with temporal fluctuations, and for a fluid motion spanning a length of the same order as the regenerator length. Because of the latter assumption, the frequency is much smaller than acoustic resonant frequencies. Consequently, since the flow Mach numbers are small, the regenerator length is far shorter than the acoustic wavelength, and inertia plays a negligible role in momentum compared with viscous friction. In a second part, a global model is developed of an inertance pulse tube refrigerator, incorporating both the regenerator model and a small amplitude assumption. The reduced problem is characterized by a partial differential equation for pressure and an ordinary differential equation for temperature. For specified rejection temperature and cooling temperature, phases and amplitudes of pressure, temperature and velocity are obtained. COP and cooling capacity are predicted, and the effect of inertance tube is studied in detail.
REGENERATOR AS A POROUS MEDIUM - THEORY It is useful to describe regenerators of the type used in Stirling or pulse-tube refrigerators as porous media because the mesh size is very small compared with length and diameter of the regenerator. Also, the details of the actual mesh topology are typically not known in detail, hence the value of a probabilistic description such as local averaging (or equivalently, a multiple length scale model). The natural approach is to take the size of the mesh as negligible, reducing the device to a continuous one-dimensional model. In the multiple length scale analysis, the size of the mesh is deemed to be negligible compared with larger lengths such as the regenerator length L. Assuming the detailed topology is known, in principle, the problem is fully described by the threedimensional conservation equations of viscous, conductive gas dynamics, assuming the fluid is an idea gas. Bulk viscosity is neglected. Suitable boundary conditions are prescribed on the two ends of the regenerator, and temperature and heat flux are continuous at the interface between matrix and gas. To make the problem dimensionless, velocities are scaled by a characteristic value based upon forcing by the compressor, time by from the assumption that the displacement of the particle is of the same order as the regenerator length L. The thermodynamic state is scaled by some reasonable reference state and the transport properties by their value at the reference state. The goal is then to obtain a formally equivalent one-dimensional problem for the mean local instantaneous values of the longitudinal velocity and the thermodynamic state. To that effect, the following assumptions are introduced in addition to those discussed above: 1) The study considers the case where spatial pressure gradients are of the same magnitude as the amplitude of the temporal pressure fluctuation, which at this stage is taken as comparable with the mean pressure. Since longitudinal pressure gradients due to viscous friction are proportional to the transverse velocity gradients: this assumption requires Thus the distinguished limit is introduced: 2) The thermal mass of the matrix is larger than the thermal mass of the fluid, and the temperature fluctuation is small. More precisely, the ratio of thermal mass is taken as the same order as Replacing by and using two length scales, using the notations and X respectively for the coordinates scaled by and by L, the following dimensionless equations are obtained (in which the symbol refers to differentiation over ):
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Next, perturbation series in are introduced for all variables, such as for instance for temperature: Taking the second order continuity equation, integrating over the void cross section A(X), taking into account that velocity vanishes at the matrix walls, and that the multiple scale analysis allows for no nonzero mean gradient over the short length scale, the expected one dimensional continuity is obtained, using an averaged velocity defined by integration over the void cross section
The problem now is to obtain a similar relationship involving only U, and from momentum. The multiple scale technique requires the indeterminacy in the allocation of the pressure gradients between short range gradients in and long range gradients in to be resolved with no short range mean gradients. Allocating the non-zero mean to the long range, a momentum equation similar to Darcy’s law is obtained for leading order velocity and pressure (See details in [6]).
in which the friction factor is defined by reference to a local incompressible flow problem (on the scale of the mesh size and assuming the detailed mesh topology is known), at the local instantaneous thermodynamic state and the local instantaneous Reynolds number, over an infinite length of mesh and a unit total volumetric flow over the whole void cross section. Calling the resulting pressure field and the local velocity field,
Integrating the energy equation over one period, and over the entire cross-section, the second order enthalpy flux is found to be uniform lengthwise,
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In the expression for the enthalpy flux, five correlations appear, respectively denoted as: and the first of which being the Nusselt number at constant heat flux, while the other ones are new. The second one plays a key role in thermoacoustics; it is thus called the Rott number, in honor of Nicholas Rott. They are formally defined as follows:
where the local functions and which depend upon local instantaneous Prandtl and Reynolds numbers and matrix topology, are determined by the following local flow problems:
This completes the one-dimensional regenerator model. The problem now includes continuity given by Eq. (5), momentum by Eq. (7), and leading order temperature is obtained as the solution of the enthalpy flux equation, Eq. (9) in which the flux, an unknown absolute constant, in effect appears as an eigenvalue. The six correlations that characterize the porous medium will require experimental determination. This problem differs from the standard one-dimensional regenerator formulation. Indeed, the standard one-dimensional energy equation cannot formally be derived from the current procedure. Worse, because it is based upon using an incorrect averaged temperature in the first term in Eq. (3), it demonstrably fails to conserve energy except for uniform velocity profiles in straight passages.
INERTANCE PULSE TUBE REFRIGERATOR The typical inertance pulse tube configuration is shown in Fig. 1. From left to right, the various spaces that the device includes are: the compressor (CP), the aftercooler (AC), the regenerator (Reg), the freezer (F), the pulse tube (T), the inertance tube (In )and reservoir(Res). The compressor, denoted as CP, is the only forcing element in the system, generating a oscillating pressure through the entire device. The compression and expansion of the gas results in a temperature gradient in the regenerator. The profile is determined by the geometry of the cryocooler and the operating conditions. The analysis that follows is based upon the small amplitude assumption. This model originates in the phasor diagrams introduced by Storch and Radebaugh [7]. The following presentation closely follows the precise formulation by Bauwens [8]. All volumes are of the same order, but the compressor
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displacement is taken to be of smaller magnitude compared with the various volumes. The ratio between compressor amplitude and void regenerator volume << 1. Scaling all volumes by the regenerator volume, the compressor motion is given by
Velocities in all spaces including the regenerator (except the orifices) have the same magnitude as the velocity of the compressor, which is of order The solution is sinusoidal, and using the mean pressure as the reference pressure, The equations for the regenerator become:
The correlations K, and vanish as a result of the small amplitude limit. Spaces such as the compressor, pulse tube and aftercooler, can be represented as one-dimensional isentropic but not necessarily homoentropic flow, over a length short compared with the acoustic wavelength, hence negligible pressure gradient. Elimination of density between mass and entropy conservation yields a relationship between pressure and velocity, written, in Fourier space, as:
In the aftercooler and the freezer, the flow is approximately isothermal. For constant temperature, mass conservation yields an expression similar to Eq. (23) but without the factor in the first term. The inertance tube is modeled as an isentropic acoustic tube plus an orifice. The length of the tube is thus taken as of the same order as the acoustic wavelength. Pressure and velocity are then expressed as:
in which is the dimensionless length of the inertance tube, is the wavelength at the operating frequency and and are determined by the boundary conditions on both sides. In the laminar limit, the orifice is a simple ”resistance” with value: [9, 8]:
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and the mass flow rate is uniform along the orifice, thus so is U. At the connection points corresponding to the interfaces between different components, velocity and pressure are continuous, but temperature may not be. For a unit compressor amplitude, After elimination between the various relationships thus obtained, the boundary conditions to the regenerator problem are obtained [6].
RESULTS AND DISCUSSION Configuration and Operating Parameters The model above was applied to a pulse-tube cryocooler with dimensions close to our prototype. The heat rejection temperature is 300 K and the cooling temperature is 50 K. The regenerator volume used as the reference volume is The other volumes are: compressor aftercooler freezer pulse tube 3.0 cm3, reservoir The resulting dimensionless volumes are The working fluid is helium, with The friction factor is taken to be constant, with value 0.1. The reference state is T = 300 K and All results shown below are dimensionless. The effect of frequency is incorporated in the friction factor.
Longitudinal Profiles in the Regenerator For the specific value of the orifice resistance and Lin = 5.0 of the inertance tube length, the longitudinal temperature, pressure and velocity profiles were obtained, and they are shown respectively in Figs. 2 to 4. For comparison, results are also shown for an otherwise identical cryocooler, but with zero inertance tube length. The temperature profile for the orifice pulse tube is closer to a linear profile. There is a significant drop in pressure amplitude across the regenerators for both cryocoolers. But the amplitude is larger when the inertance tube is present, because of the fluid inertia in the long tube. The phase shift in the former one is also quite different from that of the typical orifice design. Instead of monotonously decreasing, the phase of the pressure first decreases to a minimum and then increases again. The oscillating velocity leads the oscillating pressure at the hot end of regenerator and lags at the cold end. For an orifice pulse tube, in contrast, velocity leads pressure at both ends, which reduces the performance. Refrigeration is 20% larger. Velocity hence mass flow rate is larger with an inertance tube.
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Effects of Impedance on Performance The inertance tube impedance has a significant effect on performance. As shown in Fig. 5, for a fixed orifice resistance, there is an optimum tube length maximizing refrigeration. The optimum COP, however, is obtained for a shorter length, although both lengths are close to each other. A smaller orifice resistance yields a larger refrigeration and COP. In other words, the inertance tube should have as small as possible resistance, which conflicts with the requirement that the tube volume should be minimized. The benefit of the inertance tube is that it allows for larger pressure and velocity amplitudes, and for a smaller phase shift. The reservoir volume has different effects on the inertance and the orifice pulse tube; in the latter one, larger reservoir leads to better performance. Beyond a certain point, further volume increase has little effect, COP and refrigeration capacity don’t changer any more. In the inertance pulse tube cryocooler, the volume has an optimum value, after which the performance will deteriorate significantly. This is probably because a large volume acts like a damping force hence attenuating the effect of inertia. The proper value of the volume yields the suitable effect to the phase shift. The key parameter characterizing the regenerator matrix is the friction factor Figure 7 shows its effect on global performance. As expected, both COP and refrigeration decrease with friction because friction results in irreversibility hence a decrease in up in the freezer.
CONCLUSIONS The oscillating compressible flow in the regenerator has been studied using the multiple scale technique. The reduced problem included one-dimensional continuity and suitably simplified momentum and energy equations. Inertia effects in the regenerator can neglected because of the low Mach number
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and because the regenerator length is short compared with the acoustic wavelength. Because of the high thermal mass of the matrix, the local temperature is nearly time-independent and restricted by cycle-steady global energy conservation. Based upon this consistent asymptotic approximation to the exact governing equations, plus the small amplitude limit, a model is proposed for the whole pulse tube device. The physical model includes piston motion, volume distribution, inertance tube and regenerator including viscous losses. A detailed numerical study reveals crucial characteristics of the inertance pulse-tube cryocooler. Overall, the hypotheses are realistic, and the results should thus be meaningful.
ACKNOWLEDGMENTS Work supported by the Natural Science and Engineering Research Council of Canada.
REFERENCES [1] Kanao, K., ”A Miniature Pulse Tube Refrigerator for Temperature below 100K,” Cryogenics, vol. 34, Supplement (1994), pp. 167-170. [2] Godshalk, K. M., Jin, C., Kwong, Y. K., Hershberg, E. L., Swift, G. W., Radebaugh, R., ”Characterization of 350 Hz thermoacoustic driven orifice pulse tube refrigerator with measurements of the phase of the mass flow and pressure,” Advance in Cryogenic Engineering, vol. 41 (1996), pp. 1411- 1418. [3] Zhu, S. W., Zhou S. L., Yoshimura, N. & Matsubara, Y., ”Phase Shift Effect of the Long Neck Tube for the Pulse Tube Refrigerator,” Cryocoolers 9, New Hampshire (1996), pp. 269-278. [4] Gardner, D. L., ”Use of Inertance in Orifice Pulse Tube Refrigerators,” Cryogenics, vol. 37 (1997), pp. 117-121. [5] Duband, L., Charles, I., Ravex, A., Miquet, L. and Jewell, C., ”Experimental results on inertance and permanent flow with pulse tube coolers,” Cryocooler 10 (1999), pp. 281-290. [6] Mayzus, P, Fang, L., Deng, X., Fauvel, O.R. and Bauwens, L., ”Pressure Gradients in the Regenerator and Overall Pulse-Tube Refrigerator Performance,” accepted February 11, 2002, AIAA Journal. [7] Storch, P.J. and Radebaugh, R., ”Development and experimental test of an analytical model of the orifice pulse tube refrigerator,” Advances in Cryogenic Engineering, vol. 33 (1988), pp. 851-859. [8] Bauwens, L., ”Interface loss in the small amplitude orifice pulse-tube model,” Advances in Cryogenic Engineering, vol. 43 (1988), pp. 1933-1942.
[9] Mirels, H., ”Effect of Orifice Flow and Heat Transfer on Gas Spring Hysteresis,” AIAA Journal, vol. 32, no. 8 (1994), pp. 1656-1661.
Cooling Performance of a Small GM Cryocooler with a New Ceramic Magnetic Regenerator Material T. Satoh and T. Numazawa* R&D Center, Sumitomo Heavy Industries, Ltd. Yokosuka, Kanagawa 237-8555, Japan * Tsukuba Magnet Lab., National Institute for Materials Science Tsukuba, Ibaraki 305 -0003, Japan
ABSTRACT A new ceramic magnetic regenerator material (GOS) has been developed for use in regenerative cryocoolers in the 4 K region. Experimental results of the cooling performance of a small Gifford-McMahon cycle (GM) cryocooler with GOS are described in this paper. A small GM cryocooler with 0.l W-class cooling capacity at 4.2 K was used in the experiments. The cryocooler has been developed for superconducting device cooling or test sample cooling for physical property measurements, etc. The cryocooler is driven by single phase 100V electricity and the power consumption is about 1.2 kW (50 Hz). The conventional second-stage regenerator of the cryocooler is composed of lead spheres and spheres. In our experiments, a part of was replaced with a new ceramic magnetic regenerator material GOS and the cooling capacity was measured. The second-stage cooling capacity at 4.2 K was improved by more than 20% by replacing a part of the with GOS.
INTRODUCTION Since liquid helium temperatures were first reached using a three-stage GM cryocooler with the magnetic material Rh in the regenerator1, many kinds of magnetic regenerator materials have been developed and applied to cryocoolers to improve their cooling performance in the 4 K region. is now extensively used as a magnetic regenerator material for regenerative cryocoolers2 because it has good volumetric heat capacity characteristics and magnetic properties. However, the heat capacity of decreases remarkably below 4 K, which causes a rapid falling off of the cooling efficiency below 4 K. To improve the cooling performance of regenerative cryocoolers below 4 K, (GAP) was developed.3 GAP has a very large volumetric heat capacity peak at about 3.8 K coming from a magnetic phase transition to the antiferromagnetic state. The heat capacity of GAP at 3.8 K is more than three times that of and this considerably increases the cooling performance of regenerative cryocoolers below 4 K.4 At the same time, the cooling performance of a regenerative cryoCryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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cooler above 4 K was not particularly improved, since the volumetric heat capacity of GAP decreases very steeply above 3.8 K. Thus, a new ceramic magnetic regenerator material GOS with high volumetric heat capacity around 5 K has been developed to improve regenerator cooling performance in the temperature region immediately above 4 K.4 The results of investigating the effects of GOS on the cooling performance of a 4 K GM cryocooler are described in this paper.
EXPERIMENTAL SETUP Figure 1 shows a schematic diagram of our experimental set up. A two-stage GM cryocooler with 0.1 W-class cooling capacity at 4.2 K was used in the experiments. The compressor is an air-cooled CNA-11 from Sumitomo Heavy Industries, Ltd. It has a rated input power of 1.2 kW at its driving frequency of 50 Hz. The cylinders of the cold head are made of thin stainless steel tubes; the first stage cylinder inner diameter is 44 mm, and the second stage cylinder inner diameter is 18 mm. A copper block is brazed to the stainless steel tube at the end of the second stage cylinder to improve heat exchange. Each cylinder contains a displacer, into which the regenerator material is loaded. The displacer stroke is 16 mm, and the displacers are driven by an AC synchronous motor. The first stage regenerator is composed of 180-mesh copper screens in the higher-temperature region and lead spheres in the lower-temperature region. The second stage regenerator is composed of lead spheres (72 g) in the higher-temperature end, and a magnetic regenerator material in the lower temperature end. Lead spheres of diameter 0.4 mm to 0.5 mm and spheres of diameter 0.2 mm to 0.5 mm are used. Figure 2 shows the temperature dependence of the volumetric heat capacity of GOS, GAP, and lead below 10 K. has a heat capacity peak due to its magnetic phase transition to its antiferromagnetic state around 7 to 10 K. GOS also has a similar heat-capacity peak around 5.2 K due to its magnetic phase transition to its antiferromagnetic state. Since it has a much larger heat capacity around 6 K than as shown in Figure 2, GOS is a good candidate to improve
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the cooling performance in the 4 K region. However, since GOS has a smaller heat capacity than GAP below 4 K, GAP might be better for lower temperature use. As shown in Table 1, seven types of second-stage regenerators were prepared to investigate the effect of GOS on the cooling performance below 5 K. GOS spheres of average diameter 0.4 mm were used in the experiments. The intake/exhaust valve timing was optimized for each experiment, and the helium charge pressure was 2 MPa..
EXPERIMENTAL RESULTS Test data were acquired via a germanium resistance thermometer that was mounted to measure the second-stage temperature, and a platinum-cobalt alloy resistance thermometer was used to measure the first-stage temperature (see Figure 1). An electric heater of manganin wire was installed on each stage to allow known heat loads to be applied independently. To reduce the heat load from room temperature to the second stage, a thermal shield was attached to the first stage.
Cooling Performance of 0.1 W Machine Figure 3 shows a typical cooling-load map of the 0.1W machine using the baseline No.1 second-stage regenerator defined in Table 1. The driving frequency was 50 Hz, and the cycle speed was 60 rpm. The lowest temperature of the second stage with no heat load was 2.17 K, and the cooling capacity at 4.2 K was about 0.2 W. The cooling capacity below 4.2 K at the second stage is only weakly dependent on the first-stage temperature, at least up to 50 K.
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The first and the second-stage temperatures of the cryocooler are slightly dependent on gravity orientation, as shown in Figure 4. The second-stage temperature goes up less than 0.1 K when the orientation is changed from 0 deg. to 90 deg.; this means that the degradation of the cooling performance by changing the orientation is very small.
Effect of GOS on Cooling Performance The heat capacity of the regenerator material is a very important factor effecting the performance of regenerative cryocoolers. Since GOS has a larger specific heat below around 5 K than as shown in Figure 2, it was expected to be an effective regenerator material to improve the cooling capacity below 5 K. As shown earlier in Table 1, seven types of the second-stage regenerators were prepared for the investigation. In these regenerators, 50% of the second-stage regenerator volume (the higher temperature side) was filled with lead spheres and remained unchanged for all of the investigated regenerator configurations. In the lower temperature region of the various second-stage regenerators, either GOS, or a combination of the two was loaded. Figure 5 shows the relationship between the heat input to the second stage and the secondstage temperature. No heat was input into the first stage in the experiments. The lowest temperatures were 2.12 K to 2.19 K. The cooling capacity from the lowest temperature to around 4 K was nearly the same for the regenerators with either or GOS. But above this temperature, more substantial cooling capacity differences were visible. The poorest performance was achieved with either 100% or 100% GOS. When 0.5 W was input into the second-stage, these regenerator mixes led to temperatures of 5.8 K and 7 K, respectively. On the other hand, when or GOS were combined in the regenerator, improved performance was found. The cooling capacities for those regenerators containing both magnetic regenerator materials were similar from the lowest temperature up to around 5 K. Above 5 K, cooling capacity differences again became visible depending on the mass ratio of and GOS. The cooling capacities at 3.8 K, 4.2 K and 5 K are shown in Figure 6 for the seven regenerator configurations listed in Table 1. As noted earlier, 72 g of lead spheres were loaded into the higher temperature region of all regenerators. Figure 6 shows that the cooling capacities at 3.8 K and 4.2 K are improved by replacing a part of the with GOS. However, the cooling capacities are not improved by replacing all of the with GOS. This means that using both materials for a regenerator is important to improving the cooling capacity. It is thought that which has a larger heat capacity than GOS above 5.2 K, compensates for the low heat capacity region of GOS
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above 5.2 K. The cooling capacity was largest for the configuration 25g/38.2g at both 3.8 K and 4.2 K. The largest cooling capacities were 0.189 W at 3.8 K and 0.224 W at 4.2 K, which are respectively 22.7% larger and 22.3% larger than that for the regenerator No.1 of Table 1. The relationship between the cooling capacity and the regenerator configurations at 5 K is similar to that at 3.8 K and 4.2 K, as shown in Fig. 6. But the largest cooling capacity (0.391 W at 5 K), which is 15% larger than that for the configuration 74.5g/0g, was obtained for the configuration 55g/16.1g. The optimum configuration is different from that for 3.8 K and 4.2 K cooling capacities. The amount of for the optimum regenerator configuration for 5 K cooling capacity is larger than that for the other cases; this is because has a larger heat capacity than GOS above about 5 K.
Cooling Performance of 0.1 W Machine with The cooling performance of the cryocooler with as the working fluid was also investigated. The second-stage regenerator No.5 of Table 1 with a GOS mass of 37.3 g instead of 38.2 g was used in the experiment. The charge pressure was 1.96 MPa, and the heat load on the firststage was 0 W. The relationship between the heat input and the second-stage temperature is shown in Figure 7. The cooling load line for the case is also shown; the cycle speed is 60 rpm. The lowest temperature was 1.47 K, which is lower than the no-load temperature obtained by a GM cooler with GAP using The cooling capacity at 2 K was 32.1 mW. As shown in Figure 7, the cooling capacity above the superfluid transition temperature of is also better for because the non-ideality effects are smaller for compared to The cooling capacities at 3.8 K, 4.2 K and 5 K were 0.224 W, 0.288 W and 0.453 W, respectively, which are improved more than 15% comparing to that for
CONCLUSIONS A small GM cryocooler with a cooling capacity around 0.1 W at 4.2 K has been developed. The cryocooler can be driven by wall-power electricity and the power consumption is about 1.2 kW (50 Hz). The compressor and the coldhead weight are 42 kg and 7.2 kg, respectively. The effect of a new ceramic magnetic regenerator material on the cooling capacity has been investigated using a 0.1 W-class GM cryocooler. The cooling performance of the GM
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cryocooler was considerably improved by replacing a part of the with GOS. The cooling capacity at 4.2 K was improved by more than 20% using the optimum regenerator configuration of and GOS. This shows GOS is a good candidate regenerator material for 4 K use. However, a reliability test of GOS is needed before it can be used in a commercial product. The performance of the cryocooler with was also investigated. The no-load temperature of 1.47 K is the lowest temperature ever obtained by a regenerative cryocooler. A 0.1 W-class GM cryocooler with could be a reliable, practical 2 K cooling system without liquid helium for cooling superconducting devices or for measuring physical properties at liquid helium temperatures.
REFERENCES 1.
M. Nagao, T. Inaguchi, H. Yoshimura, T. Yamada and M. Iwamoto, “Helium Liquefaction by a GiffordMcMahon Cycle Cryocooler”, Advances in Cryogenic Engineering, vol. 35, Plenum, NY (1990), pp. 1251-1260.
2.
A. Onishi, R. Li, T. Satoh, Y. Kanazawa, K. Yabe, K. Ueda, Y. Adachi, I. Umehara and K. Sato, “Performance of 4K-GM cryocooler using magnetic regenerator material Proceedings of the Fifth Japanese-Sino Joint Seminar (1997), pp. 95-100.
3.
T. Numazawa, O. Arai, A. Sato, S. Fujimoto, T. Oodo, and M. Kan, “New Regenerator Material for Sub-4K Cryocoolers,” Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2001), pp. 465-473.
4.
T. Numazawa, T. Yanagitani, H. Nozawa, Y. Ikeya and R. Li, “A New Ceramic Magnetic Regenerator Material for 4K Cryocoolers,” Cryocoolers 12, Kluwer Academic/Plenum Publishers, New York (2003).
Improvement of 4K GM Cooling Performance with a New Regenerator Material Y. Ikeya, R. Li Cryogenics Department, Sumitomo Heavy Industries, Ltd. Tokyo 188-8585 Japan T. Numazawa Tsukuba Magnet Laboratory, National Institute for Materials Science Tsukuba, 305-0003 Japan
ABSTRACT Sumitomo's 4K GM cryocooler was commercialized in 1997, and has been successfully applied to a number of cryogenic applications, such as helium recondensing of MRI magnets and direct cooling of superconducting magnets. The rare earth material, and Sumitomo's patented unique seal technology are used in the 2nd stage displacer of the Sumitomo 4K GM cryocooler (SRDK-408D). These two technologies contribute to the large cooling capacity and high temperature stability for long-term and orientation-free installation. Recently, a new regenerator material, which has extremely high volumetric specific heat around the 4K region, has been developed. This regenerator material has promising possibilities for the improvement of the 2nd stage cooling capacity of the 4K GM cryocooler. In this study, GOS is applied to a twostage 4K GM cryocooler and evaluated for its cooling capacity improvement. By replacing 50% of with GOS, the cooling capacity of the 2nd stage at 4.2 K was improved approximately 10%. The most important thing for the industrial use of this new material is its reliability for long term operation. Reliability test data are also reported in this paper.
INTRODUCTION The Gifford-McMahon (GM) cryocooler has been extensively investigated, modified and commercialized since it was invented in 1959[1]. All of these modifications of the GM cryocooler have been based on its original design concept. It has been widely applied to industrial uses such as cryopumping, shield cooling of MRI magnets, and other applications. The achievable lowest temperature for GM cryocoolers has been around 10K, as limited by the specific heat of the lead (Pb) spheres used in the 2nd stage regenerator. After 30 years from its original invention, 4K two-stage GM cryocoolers have recently been developed based upon the use of magnetic regenerator materials for the second stage regenerator [2,3]. Several different kinds of rare earth materials have been developed for 4K cryocoolers [4], and many investigations have been conducted focused on improving the GM cryocooler cooling capacity at 4 K. Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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For the further improvement of the performance of 4K GM cryocoolers, a ceramic magnetic material has been developed for sub-4K cooling; its effect for the improvement of the cooling capacity of 4K cryocoolers was reported previously [5]. The GAP material was also applied to the Sumitomo 1W 4K GM cryocooler by these authors; however, the cooling capacity at 4K was not improved. Recently, a new kind of ceramic magnetic material was developed for 4K cryocooler regenerators. This material has considerably higher specific heat around 4K than other materials such as and GAP. Therefore, this material has the potential for further improving the cooling capacity of 4K GM cryocoolers. In this study, the cooling performance and the reliability of the Sumitomo 4K GM cryocooler, SRDK-408D, has been investigated and evaluated using the new magnetic regenerator material, GOS, applied to the second stage regenerator.
PROPERTIES OF GOS The newly developed ceramic regenerator material, GOS, has the following unique characteristics. (1) The peak of volumetric specific heat is at 5.2K; the peak value is larger than (2) The volumetric specific heat below 5.2K is 2 or 3 times larger than that of (3) The magnetization in the magnetic field is less than half of (4) The thermal conductivity is much higher than that of stainless steel The detailed properties of this regenerator material are reported in another paper [6].
Specific Heat The volumetric specific heat of a magnetic regenerator material is the most important factor to produce cooling capacity for 4K GM cryocoolers. The rare earth material, has been used in the Sumitomo 4K GM cryocooler because it has a relatively high specific heat below 10K. Figure 1 shows the temperature dependence of volumetric specific heat of several magnetic regenerator materials. GAP has a peak of specific heat at 3.8K caused by a magnetic transition. The specific heat of GAP below 3.8K is 3 to 4 times higher than that of but it sharply decreases and has less specific heat than that of above 3.8K. GOS has a similar temperature dependence and has a higher peak specific heat than GAP. The specific heat peak of GOS at 5.2K exceeds and it has 2 to 3 times higher specific heat than below 5.2K. GOS is considered a promising magnetic regenerator material for 4K cryocoolers because of the much higher specific heat between 4K and 5K.
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Magnetization The magnetization of the regenerator material is also important because many 4K GM cryocoolers are used for the cooling of superconducting magnets. A magnetized regenerator material in the magnetic field generates magnetic noise, and the reciprocating regenerator material is subjected to magnetic forces that may decrease the displacer reliability. A ferromagnetic material, was applied to the old type of Sumitomo 4K GM cryocooler. This cryocooler experienced some disadvantages in applications involving superconducting magnets and device cooling, such as magnetic noise or low cooling capacity in a strong magnetic field. To solve these problems, Sumitomo has changed the regenerator material from to an antiferromagnetic material, for the 4KGM cryocooler, SRDK-408D. Figure 2 shows the temperature dependence of magnetization in a magnetic field of 1.0T. The magnetization of GOS at 4K is about 40% of Thus, GOS is considered an attractive regenerator material for the 4K GM cryocooler compared with
Thermal conductivity The cooling capacity of 4KGM cryocooler depends on the efficiency of regenerator. In the helium gas intake and exhaust process, the regenerator material exchanges heat with helium gas. The efficiency of the heat exchange is related with thermal diffusivity that is determined by the thermal, conductivity and volumetric specific heat. The regenerator material with high thermal conductivity allows a deep heat penetration that makes the regenerator more efficient. Figure 3 shows the temperature dependence of thermal conductivity. The GOS has extremely high thermal conductivity compared with stainless steel (SUS) or GAP. The GOS has 4 times to 20
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times higher thermal conductivity than that of stainless steel between 4K and 10K. The GOS is, therefore, expected to be an excellent regenerator material, which has high thermal diffusivity.
4KGM CRYOCOOLER SYSTEM DESCRIPTION The SRDK-408D cryocooler used in the present experiment is shown schematically in Figure 4. It consists of a cold head (RDK-408D), a compressor unit (CSW-71C), interconnecting flexible gas lines and electrical power cables. The cold head is installed in a vacuum vessel, which has a hermetically sealed electrical connector for the lead wires of heaters and temperature sensors. A radiation shield was attached to the first cold stage and it encloses the second stage cylinder to prevent the radiation heat flux from room temperature. The first stage temperature is measured with a Pt-Co resistance temperature sensor and that of the second stage was measured with a germanium resistance temperature sensor. Manganin wires were wound up on the surfaces of the 1st and 2nd stages to apply the heat load. The second stage displacer contains a hybrid regenerator composed of lead spheres and rare earth regenerator material, The detail specification of the SRDK-408D cryocooler is summarized in Table 1.
GOS EFFECT FOR 2ND STAGE COOLING CAPACITY 2nd Stage Regenerator Figure 5 shows the cutaway of the 2nd stage displacer used in the present experiment. The Pb spheres are packed in the warm part of the regenerator. The rare earth material is packed in the cold part of the regenerator. Each regenerator material is separated by felt mat to prevent the both materials from mixing during operation. In order to make clear the effect of replacing the with GOS, the volume of Pb spheres and the total volume of magnetic material (HoCu2 + GOS) were kept constant. The volume ratio was defined as
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Volume ratio effect on 2nd stage cooling capacity The 2nd stage cooling capacity was measured while the volume ratio, “R”, was changed. Figure 6 shows the experimental result of the volume ratio “R” dependence of the 2nd stage cooling capacity at 4.2 K. The cooling performance was obtained for both 50Hz and 60Hz mains power frequency. The initial cooling performance was 1.22 and 1.19W at 4.2 K for 50 and 60 Hz, respectively. This value is normally obtained in the mass production model of the Sumitomo 4K GM cryocooler. The maximum cooling capacities, 1.36 and 1.29W at 4.2 K were obtained with R = 0.5. Thus, the cooling capacity at 4.2 K was improved about 10%. However, in the case of R = 0.25, the cooling capacity was less than R = 0. The reason for this decrement will be investigated by further study. The cooling performance was also decreased for R > 0.5 because of the rapidly decreasing specific heat of GOS above 5.2 K. Figure 7 and Figure 8 show the temperature dependence of the cooling power for the reference (R=0) and for R=0.5, for which the maximum cooling capacity was obtained. Note that the bottom temperature with no heat load for the case of R = 0.5 is almost the same as that for R = 0. However, the cooling capacity is about 10% larger around 4 K for R = 0.5.
RELIABILITY OF GOS Throughout its life cycle the regenerator material is subjected to various forces including mechanical impact during the regenerator packing process, acceleration loads caused by displacer reciprocation, and alternating flow pressures from the helium gas intake and exhaust cycles. The biggest concern for the reliability of the GOS material is surface wear and breakage of the spheres
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that could result in performance degradation and malfunction of the 4K GM cryocooler. For the reliability test, the 4K GM cryocooler with GOS was operated for 670 hours to monitor the temperature stability. The operating frequency was set to 60 Hz, which subjects the GOS to an accelerated environment as compared to 50 Hz operation. Figure 9 shows the temperature transition of the 1st stage, the 2nd stage, and ambient temperature. The lowest temperatures and the temperatures with head loads of 37W for the 1 st stage and 1.0W for the 2nd stage were measured in every data point. The temperature variation of 2nd stage at 1.0W heat load remained less than +/- 45mK during the 670 hours of continuous operation. No significant degradation of the 2nd stage cooling capacity was found. After 670 hours of continuous operation, the GOS material in the 2nd stage displacer was removed and the weight of spheres was measured to compare with the weight before the operation. No weight change was measured. If a weight change was measured, it would mean that some part of the GOS flowed out of the 2nd stage displacer. Figure10 shows the GOS surface after the 670 hours operation compared with that before the operation. The GOS still keeps a polished surface. Additional reliability testing of the GOS will be conducted before future commercialization.
CONCLUSION The new magnetic regenerator material, for 4K cryocooler was developed. This material has attractive properties, such as high specific heat, low magnetization and excellent
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thermal conductivity. The GOS spheres were applied to the 2nd stage regenerator of a Sumitomo 4K GM cryocooler and the cooling performance was investigated. By replacing 50% of existing with GOS, the cooling performance was improved about 10%. The biggest concern for the application of GOS to the 4K GM cryocooler was the mechanical reliability of the GOS. No degradation was observed after 670 hours of continuous operation of the 4K GM cryocooler with GOS spheres, and excellent temperature stability, less than +/- 45mK, was confirmed. The weight and surface condition of the GOS spheres were also compared before and after the 670 hours of operation. No weight change or surface damage was observed. A further reliability test of GOS will be conducted for the commercial 4K GM cryocooler.
ACKNOWLEDGMENT The authors wish to thank T. Yanagitani of Konoshima Chemical Co., Ltd, Japan, who provided the GOS material for this study. A. Yanagida of Sumitomo Heavy Industries, Ltd contributed to the experimental testing of this study.
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REFERENCES 1.
MacMahon, H.O. and Gifford, W. E., “A new low temperature gas expansion cycle, part 1”, Advances In Cryogenic Engineering, vol. 5, Plenum Press, New York (1960) pp. 354-367.
2.
Kuriyama, T., Hakamada, R., Nakagome, H., Tokai, Y., Sahashi, M., Li, R., Yoshida, O., Matsumoto, K. and Hashimoto, T., “High efficient two-stage GM refrigerator with magnetic material in the liquid helium temperature region”, Advances In Cryogenic Engineering, Vol.35, Plenum Press, New York (1990), pp.1261-1269.
3.
Inaguchi, T., Nagao, M. and Yoshimura, H., “Two-stage Gifford-MacMahon cycle cryocooler operating at about 2K”, Proceedings of the 6th International Cryocooler Conference, DTRC-91/002, David Taylor Research Center (1991) pp. 25-36.
4.
Hashimoto, T., Ogawa, A., Hayashi, A., Makino, M., Li, R., and Aoki, K., “Recent Progress on Rare Earth Magnetic Regenerator Materials,” Advances In Cryogenic Engineering, vol.37, Plenum Press, New York (1992), pp. 859-865.
5.
Numazawa, T., Arai, O., Sato, A., Fujimoto, S., Oodo, T., Kang, Y.M. and Yanagitani, T., “New Regenerator Material for Sub-4K Cryocoolers in the 11th International Cryocooler Conference,” Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2001), pp. 465-473.
6.
Numazawa, T., Yanagitani, T., Nozawa, H., Ikeya, T., Li, R., and Satoh, T., “A New Ceramic Magnetic Regenerator Material for 4K Cryocoolers,” Cryocoolers 12, Kluwer Academic/Plenum Publishers, New York (2003).
Thermal Hysteresis at 4 K with a GM Cryocooler G. E. Bonney Advanced Research Systems, Inc. Allentown, PA, USA 18103
ABSTRACT There has been observed a phenomenon of thermal hysteresis with pneumatically-driven two-stage Gifford-McMahon (GM) cycle cryocoolers operating at 4 K. The magnitude of the hysteresis was measured to be a maximum difference of 0.2 K at the minimum-load temperature of the second stage. The effect occurs after a cooldown from room ambient, when the asymptotic minimum-load temperature can be decreased another 0.1 K after subsequently applying a temporary additional heat load. The decreased minimum-load temperature is not transient. The effect occurs after either short or long post-cooldown stabilization periods. However, the effect is not cumulative; additional post-cooldown temporary heat loads do not continually decrease the minimum-load temperature. The effect was observed on all of several expanders of the same design, which were constructed over the course of a year. These expanders have a reciprocating displacer with internal coaxial regenerators, and the second stage regenerator contains layers of spheres of lead alloy and the popular rare-earth composite materials. The effect occurs regardless of all basic operational variables (power, frequency, pressure, stroke, etc.). This hysteresis is an irreversibility; the changes in the refrigerant helium properties are dependent and not causal. The many possible physical, thermal, and pneumatic dynamic conditions that could cause or combine to cause this effect were investigated. The initial hypotheses were: vaporization of a contaminant layer inside the heat station, and/or heat transfer hysteresis due to pressure oscillations in the expansion volume, and/or a shift in the fluid expansion volume due to gas spring hysteresis. A shift in the heat flux in this region of compressed liquid forced convection boiling is considered to be the primary cause.
INTRODUCTION The observed thermal hysteresis effect with a 4 K cryocooler, where the relative improvement of a 0.1-0.2 K colder minimum-load 2nd stage temperature has greater importance, prompted further testing and investigation into the possible causes of this phenomenon. The definition of hysteresis is derived from Greek, meaning “deficiency” or “to lag”.1 The definition is expounded in physics as: “... any of several effects resembling a sort of internal friction suffered by a body subjected to a varying stress or intensity”,2 and “... retardation of an effect when the forces acting upon a body are changed”.3 The definition is expounded in heat transfer as: “... significant drop in when boiling begins is known as a hysteresis effect.”4 The hysteresis effect in this work could be termed “posi-hysteretic”, being an advantage for cryocoolers when the temperature decreases, rather than being a deficiency, lag or retardation.
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TEST SETUP AND METHOD Nine 4 K GM cryocoolers, Advanced Research Systems model DE-204S, were constructed and tested at different times over the course of a year. All were identical in design and all the internal components were within the design tolerances. The expanders are two-stage with a pneumatically-driven reciprocating free displacer. The cycle frequency (f) is a fixed 2.0/2.4 Hz at 50/60 Hz supply power. The displacer has internal coaxial regenerators. The 2nd (coldest) stage regenerator contains layers of spheres of lead alloy and the popular rare-earth composite materials to achieve a minimum net refrigeration capacity of 0.15 W at 4.2 K. Two different model compressors were used; one with a lower operating pressure ratio, and the other a higher ( ratio of average absolute high and low pressures). Repeat test conditions maintained the same compressor operating power (voltage, phase, and frequency), water cooling flow, and ambient temperature range. The same instrumentation was used for all tests, including the cables, temperature monitor (LakeShore Model 340) and heater power supplies. A 4-lead Si diode, calibrated for 1.4-325 K (LakeShore DT-470-SD-13-1.4L), and a 2-lead thin-film heater, were both mounted on a copper block and bolted to the 2nd stage heat station with an In gasket. The manufacturer’s combined accuracy for the temperature sensor and monitor was ± 27 mK. The in situ temperature sensor was calibrated with LHe inside the expander cylinder on two different occasions, measuring 4.214 K and 4.230 at 101.3 kPa, a < ±10 mK total error. All GM cryocoolers have an inherent temperature oscillation, so temperature readings were filtered by sampling at 40 Hz, arithmetically averaged, and recorded at 1 Hz. The range of the average temperature oscillation at 4 K was 0.3 K without filtering, 50 mK with filtering. Stability at the asymptotic minimum-load temperature was determined as a change in the average temperature over a 15 minute period. The minimum-load condition was established the same for all tests by having: 1) the same test dewar and radiant heat shield (mirror-polished nickel-plated copper cooled by the first stage heat station and completely surrounding the colder second stage); 2) an insulating vacuum at the start of cooldown of < 5 mtorr with < 1 mtorr/min rate-of-rise outgassing check, and < 1 µtorr after cooldown. The sequence of each test was: 1) Initiate cooldown with the cryocooler starting at 292-297 K room ambient; 2) Wait until stable at the asymptotic minimum-load temperature; 3) Apply a heat load to the 2nd stage to increase its temperature at least 4 K warmer; 4) Remove the applied heat load; 5) Re-stabilize at the asymptotic minimum-load temperature, now colder. See Fig. 1.
OBSERVATIONS 1. The thermal hysteresis effect was measured to be a 0.1-0.2 K colder minimum-load 2nd stage temperature on all units. This represents a 2-5% improvement at 3.7-4 K. See Fig. 2 for representative data from three units.
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2. The thermal hysteresis effect was not transient. The decreased minimum-load persisted after the applied heat load was removed, and did not gradually revert warmer. 3. The thermal hysteresis effect occurred after either short (< 1 h) or long (> 16 h) postcooldown stabilization periods. The decreased minimum-load T2 resulted only after a temporary applied heat load, and therefor was not an expedited longer-term temperature drift or residual cooldown. 4. The thermal hysteresis effect is not cumulative. Additional post-cooldown temporary heat loads do not continually decrease the minimum-load All of the preceding facts had to be considered when hypothesizing about any possible cause.
HYPOTHESES AND ANALYSES OF POSSIBLE CAUSES
1. Changes in the helium properties? The 4 K GM cycle at the 2nd stage is operating in the compressed liquid region (< 5.2 K and > 0.23 MPa, sometimes referred to as sub-cooled or supercritical). The helium is coldest at the end of the constant volume exhaust portion of the cycle. The helium’s specific heat and density undergo large changes as the temperature drops through the 2nd stage regenerator. The area inside the two curves of Fig, 2 represents dissipated energy or entropy (s). Following the concept of entropy production in a constant heat flux (Q) process5, 6:
The net change in entropy associated with the decreased temperature at minimum-load establishes the hysteresis as an irreversibility. It can also be considered as a change from a less probable to a more probable state.6 Entropy production and irreversibility are state-point dependent and not causal. Therefor, the helium properties don’t cause the temperature shift; the temperature and pressure state points define the properties.
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2. Changes in fluid volume expansivity
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?
If viewed as a property change in density, which is state point dependent, then (sometimes referred to as the thermal expansion coefficient) is also dependent and not causal. However,
such that a change in temperature can be due to a change in specific volume (v) at constant mass and pressure (P). Rearranging Eq. (2),
For example, given a 0.1 K change at 3.8 K and 1.2 MPa average cycle pressure 7 then and Although this change in average fluid density is well within the range of density changes during each cycle and cooldown, it would be transient and would occur without needing an applied heat load. It has been determined to avoid operation at low volume expansivity, which will cause a negative Joule-Thomson coefficient, and temperature rises during isenthalpic pressure drops in the displacer.8
(Z = compressibility, R = gas constant, ). To avoid operating in such regions in a 4 K cryocooler (i.e. T < 10 K, P > 0.8 MPa and low ) is clearly difficult. However, the observed hysteresis effect was not a net temperature increase.
3. Changes in expansion volume? Given an average mass of helium in the expansion space, the physical contraction of the expansion volume materials could contribute to Hypothesis 2, Drift of the neutral position of the displacer stroke was measured on some units as 5% of full stroke. As a further test, displacer stroke was purposely changed > 20%, but it only affected net refrigeration (see Hypothesis 10). Despite these changes in expansion volume, the hysteresis effect consistently recurred. Similarly, if the compressible volume of the expansion space fluid were to be re-configured by the first post-cooldown applied heat load, then the effect should also be transient which it was not. See also Hypothesis 12.
4. Second stage seal leakage? Although a different stratification of gas occurs in the displacer-cylinder gap due to a heat load change, which could effect the amount of flow loss past the seal, such a condition should reverse after the heat load is removed, but it didn’t.
5. Particulate clogging of pneumatic drive controls and/or regenerator beds? The observed effect occurred repeatedly with both used and new components, with and without accumulated valve disc or seal wear dust, on all nine cryocoolers over the course of a year.
6. Regenerator packed bed loosening or compacting? None of the nine 2nd stage regenerators loosened, and none of their measured steady-state “dc” flow rates changed. Variations in porosity, particle size, and particle sphericity were within the range of the available product and assembly tolerances. In addition, the porosity of the 2nd stage regenerator by design was such that (ratio of regenerator void volume to expansion space swept volume) minimized the regenerator loss.9
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7. Influenced by operating pressures, pressure oscillations, or gas supply volumes? The hysteresis effect occurred regardless of all of these. Operating pressures were varied ± 10% by changing the system equalization pressure. Oscillations and gas volumes were varied by changing the compressor-to-expander interconnecting hose diameters and lengths.
8. Warm end supply gas temperature shift? Observations 2 and 3 dispute this as a possibility since tests were conducted over an ambient temperature range of 292-297 K. The observed hysteresis effect was neither dependent on nor proportional to ambient temperature changes in this range.
9. 1st stage temperature shift? The effect occurred regardless of 1st stage temperature, which varied within a 6 K range throughout all the tests. Nor did applied heat loads to the 1st stage influence the effect.
10. Due to net refrigeration capacity (Q’net)? The effect occurred regardless of actual Q’net which simply shifted the post-cooldown asymptotic minimum-load temperature, the starting point before the applied heat load.
11. Influence from damped temperature oscillations, thermal conduction, material thermal hysteresis, thermal diffusion layer conductance, or heat transfer hysteresis? Evaluating the possibility that the measured temperature being damped by the 2nd stage heat station may be shifting, led to the following equation,10
( C, k, x = Cu heat station density, specific heat, thermal conductivity, and conduction path length, respectively; f = cycle frequency; a = constant). The calculated The calculated conduction mK at 0.10 W, 4.0 K, across Le (ref. Fig. 3). These do not equal the observed 0.1-0.2 K shift, and would require the heat station material thermal properties to shift after the initial cooldown. Although material thermal properties are assumed to be only temperature-dependent, and therefor not causal, it has been determined that all material thermal properties may exhibit
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hysteresis when the material is repeatedly thermally cycled.4 No precise measurements were made to verify if any of the thermal properties of the Cu 2nd stage heat stations were shifting. If this were occurring, the observed thermal hysteresis would have been more random with applied heat load. Observation 3 doesn’t support it, as the applied heat load was varied during the tests from 0.2-9 W. Referring to Fig. 3, the total thermal conductance (c), radial and axial, across the thermal diffusion layer and 2nd stage heat station was determined as:
where the heat transfer coefficients are: The total heat transfer is: Similarly, the net heat transfer to an isothermal expansion volume for a real gas is:8 (V = expansion space volume, Combining Eq.’s 7 and 8:
).
Utilizing available equations to determine the heat transfer coefficients,11 the total at V = 4.4 mL, Q’ = 0.10 W, which is similar to the result from Eq. 5. This verifies that temperature gradients can also be set up by pressure variations in the expansion space to cause a heat transfer hysteresis,12 but it would be unlikely to shift and not later revert. Because the prior determinations that expansion volume and operating pressure changes were not the cause of the observed effect, the importance of Eq. 9 is that the hysteresis AT would vary more with conductance than with VdP changes, and that conductance would vary much more with the convective heat transfer coefficient. Kapitza conductance was also considered as a possible contributor.13 The calculations were inexact. There is also no evidence to indicate it too would not be transient with applied heat load.
12. Gas spring heat transfer hysteresis? The amount of damping at the end of each stroke of the free displacer can alter the stroke amplitude, changing the heat flow into the expansion volume. Any irregular damping will cause hysteresis in the total VdP work and Q’net.12 It has been determined that this heat transfer effect is adiabatic at LHe temperatures.14 There is no evidence that this effect would not be transient with applied heat load, as Observation 3 disproves it. See also Hypothesis 3.
13. Magnetothermal hysteresis in the rare-earth composite? A rare-earth intermetallic compound is used in the 2nd stage regenerator for its increased specific heat at 4 K. The increased heat capacity at the material’s magnetic phase transition temperature is due to a magnetic entropy change involved in the magnetic phase transition during its manufacture.15 It is unknown at this writing if this is a contributing cause; further investigation is required.
14. A shift in the flow friction behavior of compressed LHe? It has been determined that both laminar and turbulent regime flow friction factors of supercritical helium (at 5 K, 0.95 MPa) were found to exhibit no significant deviation from the
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normal behavior of a Newtonian fluid.16 There is also no evidence that this effect would not be transient with applied heat load, as Observation 3 disproves it.
15. Condensible gaseous contaminants affecting heat transfer? Although 99.999% purity helium (< 10 ppm total contaminants) is used, and excessive processing is done to desorb and control further outgassing of water vapor, air, and hydrocarbons, some deposition of contamination will occur in the regenerators and on internal 2nd stage surfaces during cooldown. This could impede heat transfer and also cause hysteresis by changing the boiling heat flux by activating a larger density of bubble nuclei.17,18 There are several reasons why this is probably not a sole cause: a) Cryocoolers of this size have a large capacity for contamination. From some testing, ~ 200 ppm in the system could cause a 0.1 K 2nd stage temperature rise. The level of in the system at the start of cooldown during the present tests never exceeded 35 ppm. b) The effect would then be sporadic and/or transient, but Observation 2 disproves it. c) The effect occurred after performing different methods of decontamination, or contaminant migration, to reduce the contaminant level in the 2nd stage while the expander was operating. d) The effect occurred with small applied heat loads, increasing to 5-10 K, which is too low a temperature at these pressures to evaporate any contaminants. e) Although it has been determined that a light dusting of the LHe boiling surface with crystals can cause hysteresis by increased Q’/A (heat transfer per unit surface area or heat flux), 18 it occurs in the transition regime between nucleate and film boiling at Figure 4 shows that is above the region of influence for the observed 0.1-0.2 K effect.
16. Influence from surface finish inside the 2nd stage heat station? This surface finish on all nine test cryocoolers was within 0.8 µm Ra and deoxidized. It has been determined that film boiling is not influenced by surface finish.18 It has also been determined that a Cu surface oxidized > 75% can cause heat flux hysteresis in the transition 19 regime at but that is above the region of influence for the observed 0.1 -0.2 K effect.
17. Influence from orientation? The hysteresis effect occurred with the 2nd stage vertically down or vertically up. In the vertical up orientation gravity pulls LHe away from the 2nd stage heat station, widening the thermal diffusion layer. This only shifted the post-cooldown asymptotic minimum-load temperature slightly warmer and increased the amplitude of temperature oscillations. It has been determined that the heat transfer coefficient for nucleate boiling is not a function of orientation.19 This adds support to Hypothesis 18.
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18. Nucleate boiling heat transfer hysteresis? It was concluded in Hypothesis 11 that a shift in the thermal conductance is predominantly controlled by the heat transfer coefficients in this region of compressed liquid forced convection boiling. The total heat flux during the exhaust stroke, across the end face and wall wetted surfaces, at 0.1 W, 4.0 K, is calculated to range from an average to a maximum . Figure 4 corroborates these values to be appropriate for the observed 0.1-0.2 K effect, albeit the average operating pressure was 0.7 MPa higher than the data in that generalized plot. Maximum heat fluxes will occur at certain pressures. The calculated average 20 at correlates to Such a change in the expansion space average pressure due to a shift in 2nd stage regenerator pressure drop is unlikely because the nominal regenerator pressure drop is less than 28% of that value. Figure 4 also shows nucleate boiling hysteresis, as reported by several sources, to be a very marked phenomenon.18, 21-23 A rapid increase in wall temperature initially induces a smaller heat flux than does the steady state condition, implying a delayed bubble inception in the nucleate boiling region.21 On reducing the heat flux after the phenomenon occurs, the heat transfer coefficients are higher. It has been determined that the temperature drop hysteresis phenomenon is an irreversible process whereby a “froth bubble layer” forms, increasing the heat transfer.22 Furthermore, such hysteresis encountered during small mass flow rates of LHe is similar to other first-order transitions (e.g. boiling curves during liquid-vapor phase changes).23 Figure 5 shows a plot of T vs. applied heater current, I, at a mass flow rate, m’= 0.02 g/s at 0.45 MPa. The hysteresis loop widened at larger applied heater current rates, I’. Although the tested DE-204S cryocoolers have a m’ ~ 100x higher, and the I’ effect was not seen, the phenomenon is similar.
CONCLUSIONS 1. The hysteresis phenomenon is an irreversibility. 2. The hysteresis varies more with conductance than with volume or pressure changes, and conductance varies much more with the convective heat transfer coefficients. 3. A shift in the heat transfer coefficients in this region of compressed liquid forced convection boiling, could be the primary singular cause, based strongly upon similar effects observed by others during boiling LHe experiments. 4. The observed hysteresis is probably a combination of steady-state and transient responses to the temporary applied heat load, such as: a primary conductance shift via nucleate boiling heat transfer hysteresis, plus contributions from material thermal property hysteresis, dispersal of contaminants, regenerator pressure drop change, and expansion volume change. 5. It is unclear why it is not transient. Additional tests are required.
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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.
23.
The Reader’s Digest Encyclopedic Dictionary (1967), p. 662 The American College Dictionary, Random House (1968), p. 596 Encyclopedia Britannica, Micropedia Vol. 5 (1980), p. 261 Rohsenow and Hartnett, Handbook of Heat Transfer, McGraw-Hill (1973), pp. 3-6, 13-22 Arpaci, “Thermal Deformation: From Thermodynamics to Heat Transfer”, ASME Journal of Heat Transfer, Vol. 123, No. 5 (Oct 2001), pp. 821-822 Van Wylen and Sonntag, Fundamentals of Classical Thermodynamics, Wiley (1973), pp. 215, 248, 390 using GASPAK, v. 3.20, Cryodata (1995) Daney, D.E., “Cooling capacity of Stirling cryocoolers - the split cycle and nonideal gas effects”, Cryogenics, Oct 1982, p. 531 Radebaugh R., O’Gallagher, A., and Gary, J., “Regenerator Behavior at 4 K: Effect of Volume and Porosity”, presented at the 2001 Cryogenic Engineering Conference, Madison, WI, pre-published. Jakob, M., Heat Transfer, Vol. 2, Ch. 14, Wiley (1949), pp.292-304 Inaguchi, T., Nagao, M., Naka, K., and Yoshimura, H., “Effects of Thermal Conductance in the Cooling Stage of a 4K-GM Refrigerator on Refrigeration Capacity”, Advances in Cryogenic Engineering, Vol. 43, Plenum Press (1998), p.1807 Urieli, I. and Berchowitz, D., Stirling Cycle Engine Analysis, Ch. 7, Hilger (1984) Barron, R., Cryogenic Heat Transfer, Taylor & Francis (1999), Ch. 3-10 Chafe, J.N. and Smith J.L., “An Experimental Study of Gas Spring Heat Transfer in Reciprocating Cryogenic Machinery”, Advances in Cryogenic Engineering, Vol. 35, Plenum Press (1990), p. 461 Ackermann, R., Cryogenic Regenerative Heat Exchangers, Plenum Press (1997), p. 97 Junghans, “Friction factor for flow of supercritical helium in a straight tube”, Cryogenics, Nov 1980, p. 633 van Stralen, S. and Cole, R., Boiling Phenomena, Vol. 1, Hemisphere Publishing (1979), pp. 376, 379 Cummings, R.D. and Smith, J.L., “Boiling Heat Transfer to Liquid Helium”, Pure and Applied Cryogenics, Vol. 6, International Institute of Refrigeration, Pergamon Press (1966), pp. 85-95 Iwamoto, M., Mito, T., Takahata, K., Yanagi, N., Yamamoto, J., “Heat Transfer From An Oxidized Large Copper Surface To Liquid Helium: Dependence On Surface Orientation And Treatment”, Advances in Cryogenic Engineering, Vol. 41, Plenum Press (1996), pp. 217-224 Klipping, G. and Kutzner, K., “Heat Transfer from Metal to Supercritical Helium”, Pure and Applied Cryogenics, Vol. 6, International Institute of Refrigeration, Pergamon Press (1966), p. 101 Smith, R.V., “Review of Heat Transfer to Helium I”, Cryogenics, Feb 1969, p. 16 Beattie, D.R.H. and Lawther, K.R., “An examination of the wall temperature drop phenomenon during approach to flow boiling crisis”, Eighth International Heat Transfer Conference, Vol. 5, Hemisphere Publishing (1986), pp. 2215-2219 Caspi, S., Lee, J.Y., and Frederking, T.H.K., “Oscillations and Hysteresis of Helium During Lambda Transition Above the Thermodynamic Critical Pressure in the Presence of Heat Flow”, Advances in Cryogenic Engineering, Vol. 23, Plenum Press (1978), pp. 349-357
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Study on the Onset Temperature Gradient of Regenerators Used for Thermoacoustic Prime Movers H. Ling1, E. Luo1, J. Wu1, M. Yang2, X. Li2 1
Technical Institute of Physics and Chemistry Chinese Academy of Sciences, Beijing 100080, China 2 Cryogenic Laboratory Zhejiang University, Hangzhou 310027, China
ABSTRACT Not only for standing-wave thermoacoustic prime movers, but also for traveling-wave thermoacoustic prime movers, the onset temperature gradient of the thermoacoustic stack of the thermoacoustic regenerator is an important parameter for understanding and designing the thermoacoustic prime mover. Due to the inherent advantages of the traveling-wave operating mode over the standing-wave mode, the traveling-wave thermoacoustic prime mover has been getting more and more attention and research. Therefore, this paper presents analysis results for the onset temperature gradient of regenerators used for traveling-wave thermoacoustic prime movers for different structures and operating parameters.
INTRODUCTION A thermoacoustic engine is a machine that can convert heat into mechanical work in the form of an oscillating pressure. In refrigeration and cryogenic engineering, it can be used as the pressure generator for regenerative refrigerators such as pulse tube refrigerators and thermoacoustic refrigerators. Among all parts of such a prime mover, the regenerator is one of the most important components. Whether the acoustic power can be produced or not is mainly determined by the onset temperature gradient of the regenerator. If the applied temperature gradient is larger than the onset gradient, a thermoacoustic self-excited oscillation occurs and a usable acoustic power can be delivered. There are numerous factors that affect the energy conversion from heat to acoustic power. In the following sections, the influence of different parameters on the onset temperature gradient of a regenerator is studied.
ANALYSIS MODEL According to linear thermoacoustic theory1,2, the acoustic power production rate per unit length along the direction of work flow is
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where the nomenclature for the different physical parameters and mathematical operations is defined in Table 1. The first item in the right side of Eq. (1) is produced by sound oscillation and temperature gradient, which is positive for a thermoacoustic prime mover. The second and third terms are caused by finite heat conductivity and finite fluid viscosity, which are always negative because they consume acoustic power. The coefficients are called the factor of power production, the factor of power dissipation by imperfect heat transfer, and the factor of power dissipation by flow friction, respectively. A screen-packed regenerator is usually used in traveling-wave thermoacoustic engines. Therefore, we discuss the performance of the regenerator in this paper.
where
ent
Whether the acoustic power is magnified or attenuated depends on the local temperature gradiThere exists a critical value which makes
is the local specific acoustic admittance. Substituting
into Eq. (1) yields
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where Obviously, when the local acoustic power flow is magnified. Conversely, when the local acoustic power flow is attenuated. Therefore, we should make as small as possible by properly designing the regenerator. There are many factors that influence the critical temperature gradient. In the following section, we examine the influence on the regenerator critical temperature gradient of a number of different structures and operating parameters. Because the local critical gradient is correlated with local temperature, we make the assumption that the temperature distribution along the regenerator is linear.
CALCULATION RESULTS AND DISCUSSION Under the same pressure and velocity oscillations, a pure standing wave does not transfer acoustic power, and a pure traveling wave can transfer an acoustic power of PU/2. If the acoustic admittance Ya is a real number, that means the pressure and velocity are in phase. In the following examples, Ya varies from 0.0005 to 0.5, gas temperature varies from 300 K to 1000 K, and the mesh of stainless-steel screens and the working pressure of the helium gas are changeable parameters. The distributions of average critical onset temperature gradient versus acoustic admittance are presented in Figs. 1 to 4; in these figures one can see that: (1) In Fig. 1, the average critical temperature gradient always decreases first and then goes up. There is an optimum Ya that makes reach a minimum. The phenomenon can be explained from Eq. (5) When Ya is very small, is inversely proportional to Ya and is
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mainly determined by finite heat transfer; when Ya is very large, is directly proportional to Ya and is mainly decided by flow friction. For different mesh regenerators, the critical temperature gradient is almost the same, but the corresponding optimum Ya is different. (2) Figure 2 shows the case for different frequencies. One can see that higher frequency increases the average critical onset temperature gradient; this is particularly visible for Ya to the left of the optimum. Thus, lowering frequency benefits the production of acoustic power. Figure 3 shows the case for different mean pressures. One can see that the minimum average critical onset temperature gradient is not particularly sensitive to mean pressure. In summary, the minimum of the average critical temperature gradient is primarily determined by frequency. (3) Figure 4 shows the distributions of critical temperature gradient versus local temperature. Heat conductivity and viscosity are strongly dependent on temperature, as is the local critical temperature gradient. One can see that curves 1 and 2 are ascending, while curves 3 and 4 are descending. This means the characteristic trendline is decided by Ya. This is because, in the high Ya section, is mainly influenced by viscosity, which goes up with increasing temperature. In the low Ya section, the influence of viscosity may be ignored and the transverse heat exchange strengthens, so goes down with rising temperature. The heat exchange at high temperature is expected to turn into work power, so the trend of curves 3 and 4 is understandable. This demands that Ya be in a low zone. (4) The case of a standing wave under the same conditions has also been simulated. The critical temperature gradient is much higher than that of a traveling wave by a hundred times. This implies that a standing-wave engine should work under different conditions, including geometry and operating parameters. The first step is to reduce mesh number and increase hydraulic radius.
CONCLUSIONS From the above calculated results and analyses, some conclusions are drawn as follows: a. The average onset critical temperature gradient has a minimum value. The minimum value depends mainly on frequency, but there is a corresponding different optimum Ya for different mean pressures and regenerator mesh numbers. b. Although the calculated results are based on a pure traveling wave, they have important significance and provide guidance for general regenerator designing.
ACKNOWLEDGMENTS This research is financially supported by the Natural Sciences Foundation of China under Contract No.59976044.
REFERENCES 1. Swift, G.W., “Thermoacoustic engine,” J. Acoust. Soc. Am, vol.84, no.4 (1988), pp. 1145-1180. 2. Xiao, J.H., “Thermoacoustic theory for cyclic flow regenerators,” Cryogenics, vol.32, no.10 (1992), pp. 895-901.
Study on Thermoacoustic DC-Flow Model for a Cyclic Regenerator due to Nonlinear Effects Ercang Luo Technical Institute of Physics and Chemistry, Chinese Academy of Sciences Beijing 100080, China
ABSTRACT A regenerator is one of the most important elements for many regenerative machines. From numerous experimental observations of regenerators, some dynamic parameters are quite large, and go beyond the assumption of linear acoustic behavior. Thus, some nonlinear effects such as non-zero time-averaged flow can be induced. The non-zero time-averaged flow often produces a significant second-order energy streaming effect in most cases, which can be a serious loss for regenerative machines. Based on a modified Darcy equation and other conservation equations, the paper proposes a model for describing non-zero flows and resulting thermoacoustic effects in a nonlinear regenerator. Moreover, the proposed theory is used to analyze three cases with secondorder, non-zero mass flux or velocity.
INTRODUCTION Nonlinear oscillations with large amplitudes are often involved in the regenerators used for many practical regenerative machines. Higher order harmonic oscillations and time-averaged flows can not be neglected, and consequently obvious thermoacoustic energy streaming can not be neglected either. In many practical cases, these effects heavily affect the global performance of the regenerative machines 1~3. In this paper, we try to propose a comparatively complete mathematical model to describe some nonlinear oscillations and nonlinear thermoacoustic energy effects in the regenerators. In particular, we present a set of comparatively complete equations for describing the thermoacoustic DC-flow effects in regenerators.
MATHMATICAL MODEL Basic Equations We start by making the following assumptions for the regenerator: (1) Flow is one-dimensional and cyclic (2) Working fluid is an ideal gas (3) Axial heat conduction is negligible in the energy equations of the gas and solid matrix, but its effects are included as time-averaged thermoacoustic effects Thus, the following basic equations can be given for the porous regenerator. The continuity equation of the working gas: Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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Eq. (1) is the rewritten form of the common continuity equation of a gas, making use of Maxell’s thermodynamic relations. The momentum equation of working gas:
Eq. (2) is a modified Darcy equation for describing the unsteady flow in a porous medium. The energy equation of the working gas:
In Eq. (3), the dissipation of the porous medium is included. However, the longitudinal heat conduction is neglected. The energy equation of solid matrix:
The longitudinal heat conduction by solid matrix in Eq. (4) is neglected. The equation of state for an ideal gas:
Asymptotic Solutions In linear thermoacoustic theory, time-averaged thermoacoustic energy effects, including total energy flux, work flux and heat flux, are second-order. However, time-averaged mass flux or velocity is also second-order, which inherently contributes a second-order effect to the abovementioned second-order, thermoacoustic energy effects. Thus, in a sense, the linear thermoacoustic theory is not a strict linear theory. To improve the linear thermoacoustic theory, it is necessary to further consider the contribution of the time-averaged mass flux or velocity having second-order. Thus, the following forms of asymptotic solutions for thermodynamic variables with second-order accuracy are assumed.
Thermoacoustic DC-flow Equations Because second-order dynamic oscillations have no contributions to the second-order, timeaveraged, thermoacoustic energy effects, one does not care about how they are coupled with the first-order dynamic oscillations. As mentions above, however, the second-order, time-averaged mass flux or velocity has the same order contribution as the first-order oscillations. Therefore, in the following part, we only give those equations to describe how the second-order time-averaged thermoacoustic effects are coupled with the first-order oscillations and the second-order timeaveraged mass flux or velocity.
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Where, and are the second-order average pressure and volume velocity, respectively. is the time-averaged mass flux. The nomenclature used for the physical properties and coefficients can be found in Reference 6. Eq. (11) to Eq. (16) are six constant differential equations that are coupled together. The first-order oscillation and the second-order time-averaged mass flux or velocity together determine the averaged temperature distribution and total energy flux.
Time-Averaged Thermoacoustic Effects In this section, we give the terms for time-averaged mass and all energy streaming terms with second-order accuracy. Time-averaged mass flowing:
The contribution from linear acoustic oscillation is zero, while the contribution from nonlinear oscillation is
Time-averaged work flux:
The contribution from linear acoustic oscillation is
The contribution from nonlinear oscillation is Time-averaged energy flux:
The contribution from linear acoustic oscillation is:
The contribution from nonlinear oscillation is Time-averaged heat flux:
The contribution from linear acoustic oscillation is
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The contribution from nonlinear oscillation is
CASES STUDY BASED ON THERMOACOUSTIC DC-FLOW MODEL Based on the above-mentioned model and analysis, three cases are studied further in this part. These cases are the onset temperature gradient of a regenerator for a thermoacoustic prime mover, the time-averaged temperature distribution along a cryocooler regenerator, and the calculation of ideal cooling capacity for an orifice pulse tube refrigerator with time-averaged mass flux or velocity.
Case 1: Onset Temperature Gradient For the regenerator used in a traveling-wave thermoacoustic prime mover, the onset temperature gradient is an important criterion. To easily excite thermoacoustic oscillations, the onset temperature gradient should be designed as small as possible. In this section, we discuss the influence of thermoacoustic Dc-flow on the onset temperature gradient. From Eq.(19), the following expression for the local onset temperature gradient of a regenerator can be obtained when time-averaged mass flow exists.
Setting
a dimensionless local onset temperature gradient is achieved as follows.
Where, is the ratio of time-averaged mass flow to the mass flow amplitude of first-order oscillation; is the ratio of oscillation pressure to average pressure. Figures 1 and 2 show the dimensionless local onset temperature gradients under the conditions of with and without thermoacoustic DC-flow, respectively. One can see that the onset temperature gradient for with thermoacoustic DC-flow becomes larger than that without thermoacoustic DCflow. This means that the time-averaged thermoacoustic DC-flow deteriorates the working of a regenerator.
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Case 2: Temperature Distribution of Cryogenic Regenerator Similarly, time-averaged mass streaming also exists in some regenerative cryocoolers such as double-inlet pulse tube refrigerators. Usually, one can estimates the magnitude of the DC-flow by measuring the temperature distribution of a regenerator or a pulse tube. As an example, the temperature distribution of the regenerator of a regenerative cryocooler was studied here under different thermoacoustic DC-flows. In the example, the inlet condition of the regenerator is always kept the same except for the DC-flow. Figure 3 shows the temperature distribution of the regenerator. Obviously, the thermoacoustic DC-flows not only changed the temperature distribution of the regenerator, but also increased the lowest temperature of the cold end of the regenerator. Moreover, one can see that a very small DC-flow heavily degrades the performance of the cryocooler.
Case 3: Theoretical Cooling Capacity of an Ideal Pulse Tube Refrigerator In the analysis of orifice pulse tube refrigerators, one usually estimates the gross or ideal cooling capacity of a pulse tube refrigerator based on the following expression4
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THERMOACOUSTIC REFRIGERATOR INVESTIGATIONS
However, it should be pointed out that this expression has a fatal mistake. The mistake results from not considering the inherent, time-averaged, second-order velocity occurring in an oscillating gas. Although there is no time-averaged mass flow in an orifice pulse tube refrigerator, a second-order time-averaged velocity exists. The second-order velocity is one kind of nonlinear effect. Thus, Eq.(29) based on simple harmonic analysis is questionable, even for the case without time-averaged mass flow. Therefore, the harmonic analysis is not correct for general cases. If the harmonic analysis method is extended to consider the second order timeaveraged velocity, the correct expression for calculating the ideal cooling capacity of an orifice pulse tube refrigerator can be achieved. Considering the factor, the following expression for the ideal cooling capacity of an orifice pulse tube refrigerator is obtained.
For an ideal pulse tube refrigerator, there is no time-averaged mass flow. Thus, a new expression for cooling capacity is achieved finally5
The ideal cooling capacity by Eq.(31) is only 40% of by Eq.(30), and this may explain why the cooling capacity by Eq.(30) is 3 to 5 times larger than measured values.
CONCLUSIONS Based on the modified Darcy model and other conservation equations, a set of nonlinear equations for describing a cyclic regenerator up to second-order accuracy are derived. The timeaveraged velocity in the regenerator is second-order, but its contribution to the time-averaged thermoacoustic effects can be comparable with the first-order oscillations. These phenomena arise from the nonlinearties of the regenerator. A thermoacoustic DC-flow model for the regenerator is derived and can be used to describe our most interesting working performance. The nonlinearties of a regenerator need to be studied in-depth in the future.
ACKNOWLEDGEMENTS The research work is financially supported by the National Sciences Foundation of China under the contract No.59976044.
REFERENCES 1. Gedon, D., “DC gas flows in Stirling and pulse-tube cryocoolers,” Cryocoolers 9, Plenum Press, New York (1997), pp. 385-391.
2. Ju, Y.L., Wang, C., and Zhou, Y., “Dynamic experimental study of the multi-bypass pulse tube
3. 4. 5. 6.
refrigerator with two-bypass tubes,” Advance in Cryogenic Engineering, vol.43, Plenum Press, New York (1998), pp. 2031-2037. Backhause, S. and Swift, G.W., “A thermoacoustic-Stirling heat engine,” Nature, no.399 (1999), pp. 335-338. Storch, P.J., and Radebaugh, R., “Development and experimental test of an analysis model of the orifice pulse tube refrigerator,” Advance in Cryogenic Engineering, vol.33, Plenum Press, New York (1987), pp. 851-859. Luo, E., Liu, H., and Wu, J., “A new model for calculating enthalpy flux of regenerative machines,” Cryogenics and Supercomputing, vol.29,no.4 (2001), pp. 36-40. Xiao,J.H., “Thermoacoustic theory for cyclic flow regenerators. Part I: fundamental,” Cryogenics, vol.32, no.10 (1992), pp. 895-901.
Thermodynamic Analysis of a Traveling Wave Thermoacoustic Device by Use of a Distributed-Parameter Method M. Yang1, E. Luo2, H. Ling2, X. Li1, J. Wu2 and G. Chen1 1
Institute of Refrigeration and Cryogenic Engineering Zhejiang University Hangzhou, China 310027 2 Technical Institute of Physics and Chemistry Chinese Academy of Sciences Beijing,China 100080
ABSTRACT Thermoacoustic refrigerators and engines typically have no moving parts. Thus, these machines have the potential to be both simple and reliable. Thermoacoustic energy conversion is reasonably efficient and should be inexpensive in mass production. The analysis of a traveling wave thermoacoustic engine based on a thermoacoustic distributed-parameter network model is presented. The continuity and momentum equations of an oscillating gas, particularly the oscillating pressure and oscillating volume flow rate of acoustic waves, have analogies in AC electrical circuits. The pressure resembles the voltage, and the volume flow rate resembles the current. A complete distributed-parameter network model has been achieved and the results are presented. The results indicate that the distributed-parameter method provides reasonable and useful simulations capable of being extended to a variety of operating conditions. Our goals include an improved understanding of the fundamental thermoacoustic processes and the development of new thermoacoustic refrigerators and heat engines with increased power density, temperature span, and efficiency, and the commercialization of those devices.
INTRODUCTION Thermoacoustic refrigerators and engines are a relatively new domain with attractive prospects, as they typically have no moving parts. Thus, these engines have the potential to be both simple and reliable. The engine designed by Swift 1–4 delivers 710W of acoustic power to its resonator with a thermal efficiency of 0.30, corresponding to 41% of the Carnot efficiency. There are many factors that influence thermoacoustic machines, such as the structure of the regenerator, the kind of working gas, heat transfer efficiency, etc. Among all these factors, the structure of the regenerator is the most important. Based on the theory of fluid networks5, we have analyzed a traveling wave thermoacoustic engine using a distributed-parameter method. In fact, the analysis results are also applicable to a traveling wave thermoacoustic refrigerator.
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DISTRIBUTED-PARAMETER MODEL Building on Swift’s lumped-element thermoacoustic electrical network model2 , we have chosen to analyze the traveling wave thermoacoustic engine with a distributed-parameter model to allow some additional features to be included. Figures 1 and 2 illustrate the mechanical elements and lumped-parameter analog circuit of the traveling wave thermoacoustic engine.
Cold End of Regenerator Starting first with the thermoacoustic lumped-element model. According to the model:
where the parameters are those defined in Fig 2. This model includes the primary elements of the heat engine; however it does not include consideration of the compliance of the regenerator, or the inertance and compliance of the thermal buffer tube. If the length of the thermal buffer tube is short enough, the error is small. However, to allow for the analysis of these effects in the more general case where these effects are not negligible, we have chosen to include them in our distributed-parameter method. With the introduction of the compliance of the regenerator, and the inertance and compliance of the thermal buffer tube, the thermoacoustic distributed-parameter network model shown in Fig. 3 results. Fig. 3(a) is the traveling wave thermoacoustic network of the whole system, and Fig. 3(b) shows the distributed model of the regenerator that is highlighted in the dashed frame of Fig. 3(a). The network of the regenerator is modeled with the distributed-parameter method. The regenerator of length l is split into segments, each segment spanning a temperature difference With an ideal gas as the working fluid, each segment of the regenerator has an isothermal compliance independent of temperature, where The parameters
ANALYSIS BY USE OF DISTRIBUTED-PARAMETER METHOD
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and S are the volume porosity and cross-sectional area of the regenerator. Across each segment, the volumetric velocity changes due to the temperature increase and the compliance ,i.e.,
Dividing by
and letting
yields a differential equation for
Each segment of the regenerator also has a resistance given by where
is the low-Reynolds-number-limit flow resistance of the regenerator when
its entire length is at temperature At high amplitude, could increase by up to 50%, but this effect is not included in the model. The pressure drop across each segment is given by Dividing by
Setting
and letting
yields a differential equation for
and integrating Eq. (3) yields the result
Similarly, substituting the result into Eq.(5) and integrated again can yield In both integrations, is assumed to vary linearly from to The volumetric velocity at the hot end of the regenerator and the pressure drop across the entire regenerator are given by
Where f and g are given by
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Eq.(7) and (8) are important in the design of a thermoacoustic engine. From Fig. 3 (a) the corresponding relationship between and is analyzed,
Combining the above four equations yields
Comparing Eq. (13) with Eq. (1), the denominator of Eq. (13) has an additional item, and
is multiplied by a coefficient,
items,
Moreover, the numerator has two additional
and
In the lumped parameter model it is sometimes assumed that is smaller than so the influence of in the denominator may be ignored. However, although is small, can be of the same order as ( is about 3 in practice). Therefore, the value of cannot be ignored in some practical systems. The value of
in the numerator is much smaller than the value of
it can be ignored. Comparing the value of order. Then, the value of phase shifting between
and
they are of the same
cannot be ignored and plays an important role in the and
in the cold end of a regenerator. Thus, the numerator is
simplified to From what has been discussed, a simplified equation from Eq. (13) is obtained.
The angle of
The modulus of
is
is
so
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From Eq.(l) and Eq.(14), one can see that the phase angle between and ranges from -90° to +90° in the distributed-parameter method, but from 0° to 90° in the lumped-parameter method. To compare the lumped-parameter method and the distributed-parameter method, consider the example shown in Fig. 4. The parameters of the thermoacoustic engine are described in reference 2. Using the lumped-parameter method for this system, it is found that the phase angle between and is When Eq. (15) of the distributed-parameter method is used, the phase angle between and is The difference between and is 28.69°. Figure 4 plots the phase angle between the as computed using both the lumpedparameter method and the distributed-parameter method. The calculation parameters of the structure are described in reference 4. As seen in Fig. 4, with increasing length of the regenerator, the difference between the lumped-parameter method and the distributed-parameter method increases. Also, with increasing length of the regenerator, the phase angle between and decreases and is always positive in the lumped-parameter method. However, with the distributed-parameter method, the phase angle between and is initially positive ( leading ) and is then goes negative ( lagging ).
Hot End of Regenerator Swift’s traveling wave thermoacoustic network analyzed the relationship between and in the cold end of regenerator. However, the relationship between and in the hot end of the regenerator can also be important.
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Combining Eq.(17), (7) and (13) yields
Giving an example, we can find the importance of the relationship between and in the hot end of regenerator. The parameters of the thermoacoustic engine are described in Reference 4. The phase angle between and is and the phase angle between and is by the distributed-parameter method. Thus, when we design the system, we should not only analyze the phase angle between and but also consider the phase angle between and We hope that both the phase difference should be small.
CASE STUDY As a case study, we have analyzed the traveling wave thermoacoustic network using the distributed-parameter method. (1) If the real part of is large and the imaginary part of it is small, and are approximate in phase from Eq. (14). (2) To decrease the loss of flowing resistance, we should increase the value of To reduce the phase angle between and the value of is larger than that of from Eq. (14), increasing the inertance and compliance of feedback tube, decreasing the compliance and the Resistance of regenerator. On the other hand, because we want to increase the value of the denominator of Eq. (14) has that is required to be large. Therefore, is decided according to the particular circumstances. and have optimum values, so they need to be optimized. (3)When we increase the inertance and compliance of feedback tube, according to where is cross-sectional area, l is length of tube, we can increase the length l of feedback tube to raise the inertance and compliance of feedback tube, but it aggrandize the volume of feedback tube and is disadvantageous in miniaturization. The best method is that one section of feedback tube has small diameter, decreasing the cross-sectional area, which increase the inertance the other section of the feedback tube has large diameter, increasing the cross-sectional area, which increase the compliance Thus, we increase the inertance and compliance of the feedback tube simultaneously under the condition that it does not increase the length of feedback tube. For the particular case study analysis the parameters of the operating conditions are: the working fluid is helium; the frequency f = 50Hz; the working pressure; and the hot end temperature of the regenerator is 725°C. In addition, the parameters of the structure are: the regenerator, the thermal buffer tube, and the feedback tube have a diameter of 50mm and their lengths can be altered; and the regenerator is made from a stack of 120-mesh stainless-steel screen. Figures 5,6,7 and 8 show the results that have been calculated. Figure 5 shows that, with increasing length of the regenerator, the phase angle between and and the phase angle between and first declines and then increases; the phase angle between and and the phase angle between and are first positive ( leading ) and then negative ( lagging ). Figure 6 shows that with increasing length of the regenerator, the modulus of first increases and then declines, while the modulus of continuously increases. We hope that the phase angle between and and the phase angle between and are small and the modulus of and are large. The proper length for the regenerator ranges from 40mm to 80mm.
ANALYSIS BY USE OF DISTRIBUTED-PARAMETER METHOD
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Figure 7 shows that the phase angle between and changes for different lengths of the regenerator under different pressures. When the length of the regenerator is fixed, and the pressure is increased, the phase angle between and gradually declines. Thus, the higher the pressure, the smaller the phase angle between and If the system permits, we may use a higher pressure to operate the system. Figure 8 shows that the phase angle between and is changed by the length of the regenerator under different frequencies. When the length of the regenerator is fixed, and the frequency is decreased, the phase angle between and gradually decreases. Thus, if the system permits, we may use a lower frequency to operate the system.
CONCLUSIONS The results have shown that the distributed-parameter method can expand upon the results of the lumped-parameter network model and provide addition insight into some of the operating parameters of the thermoacoustic refrigerator. In fact, the lumped-element model is one special case of the distributed-parameter model. As an example, the phase angle between and ranges from -90° to 90° in the distributed parameter method while the phase angle between and ranges from 0° to 90° in the lumped-parameter method. In fact, may lead or lag but only leads with the lumped-parameter method. Thus, there are cases where the distributed-parameter model may be preferred. Since the regenerator is only one part of a whole system, the relationship between and in the hot end of the regenerator is as important as it is in the cold end. Thus, we should consider both the phase angle differences in the cold and hot ends of the regenerator. We hope that both the phase angle between and and the phase angle between and are small, so it is close to the traveling wave.
ACKNOWLEDGMENT This research is financially supported by the Natural Science Foundation of China under contract No.59976044.
REFERENCES 1. Swift,G.W., “Thermoacoustic Engines,” J. Acoustic Soc. Am, vol.84, no.4 (1988), pp. 1145-1180. 2. Backhaus,S., Swift,G..W., “A Thermoacoustic-Stirling Heat Engine: Detailed Study,”J. Acoustic Soc. Am, vol.107, no.6 (2000), pp. 3148-3166. 3. Backhaus, S. and Swift, G.W., “A thermoacoustic-Stirling heat engine,” Nature, v.399 (1999), pp.335-338. 4. Liu, H., Luo, E., Liang, J., “The experience research of thermoacoustic engines,” Cryogenics, vol.3 (2000). 5. Luo,Z., The theory of fluid network, Mechanical Press, Beijing (1988).
Investigation of a High Frequency Traveling Wave Thermoacoustic-Driven System Qing Li1, Jihao Wu2, Fangzhong Guo2, Qiu Tu2, Zhibin Yu1 1
The Technical Institute of Physics and Chemistry, CAS Beijing, China 100080 2 Cryogenic Lab, Huazhong University of Science and Technology Wuhan, China 430074
ABSTRACT A traveling wave thermoacoustic-driven system with a designed resonant frequency of 200 Hz has been established. Nonlinear thermoacoustic phenomena, such as the oscillation onset temperature and shifts of the resonant frequency, are presented. Experimental investigations have been conducted by changing wire-screen mesh size, working gases, and mean pressures, and the effects have been noted on temperature profiles, oscillation modes, frequency jumps, and pressure amplitudes. The analysis indicates that the relative penetration depth ratio plays an important role with these nonlinear phenomena.
INTRODUCTION Increasing attention1,2 is being placed on thermoacoustic-driven systems due to their advantage of having no moving mechanical components, except for the working fluid, and their potential for low environmental impact. However, low-efficiency is still an unresolved problem that requires continued research. Based on the research of Yazaki3, Backhaus and Swift4 established a new type of high-efficiency traveling wave thermoacoustic Stirling heat engine at a frequency of 80 Hz with a thermal efficiency of 0.30. This high efficiency corresponds to 41% of the Carnot efficiency and was achieved by taking advantage of the high efficiency of the traveling wave and high amplitude of the standing wave. The reliability and integration of cryo-electronic systems requires miniaturization of the cryocooler system. The use of higher working frequencies (hundreds and even thousands of Hertz) is one possible way to achieve such a goal. Preliminary research on altering the regenerator screen mesh size can help accumulate the experience needed to develop higher frequency thermoacoustic systems and miniature heat engines.
THERMOACOUSTIC DRIVEN SYSTEM Based on a thermoacoustic system in the literature,4 a high frequency thermoacoustic-driven system has been established, designed using an active network model.5 The size of the system is 1.42 meter long by 36 cm tall, and the 25.4-mm inner diameter by 32-mm long regenerator is packed with wire screens. Two pressure measurement points, are located at the upper end of Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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the main cold heat exchanger and the lower end of the second cold heat exchanger, respectively. Four thermal couples are distributed uniformly along the inner wall of the regenerator (black points in Fig. 1). The designed highest fundamental frequency is 200 Hz, with helium as the working gas. On the one hand, high frequency is an approach to realizing miniaturization and high reliability; on the other hand, high frequency brings many problems, such as instability of the oscillation modes, and the phase adjusting mechanism of the acoustic field in the system. The traveling wave thermoacoustic-driven system is configurated by connecting a self-excited loop and a standing wave resonator at one point. Previous calculations and experiments 4,6 indicate that the resonant frequency of the system is defined by the length of the resonator and the properties of the working gas, i.e. f = a/4L, where a is the sound velocity of the working gas, and L is the length of the resonator. If the construction of some components of the loop tube is changed—for example, replacing the cold heat exchanger from a plate-plane type to a shell-tube type, or changing the wire screen mesh number in the regenerator—the onset-of-oscillation performance of the system will change. Instead of oscillating at the lowest fundamental resonant frequency, the system may oscillate at higher frequencies, which can be several times higher than the fundamental mode. When heated continuously, the temperature increases, and the high frequency can jump back to the fundamental mode. Actually, each subsystem has its own eigen frequency. According to the arrangement of the system, the frequencies of the each subsystem can be significantly different from each other. An integrated traveling wave thermoacoustic-driven system has new system characteristics. The selfexcited coupling resonance is a new problem that is closely related to the conditions of the whole system; this evokes some nonlinear phenomena in the system, such as onset-of-oscillation mode, modes shift, and so on. This nonlinear phenomenon is far different from higher order modes, nonlinear saturation, belonging to thermoacoustic instability and weak nonlinear phenomena. The problems and models of this nonlinear mode need to be investigated from a new point-of-view. Additional detailed experiments and discussion of this topic are available in the literature.7
EXPERIMENTAL RESULTS The traveling wave thermoacoustic system has been tested parametrically using a variety of different mesh sizes for the regenerator (i.e. changing the value of where is the hydraulic radius and is the viscous penetration depth, respectively), and different working gases and mean pressures. The resulting temperature profiles, oscillation modes, and pressure amplitudes have been collected and organized in the following sections according to the screen mesh number used.
80-Mesh Screen Regenerator For the case of 80-mesh wire screen, Fig. 2 presents pressure amplitudes at points and temperature profiles along the regenerator for the case of a mean pressure of 0.5 MPa, a heating
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power of 280 W, and with nitrogen as the working gas. Figure 3a-e expand on the observed variations of the oscillating pressure. The oscillating pressure in the traveling wave thermoacoustic-driven system starts at a high frequency (546.4 Hz). When the temperature increases to its maximum value, the pressure oscillation drops to its much lower fundamental frequency (75.66 Hz). After oscillation begins, acoustic power flows into the cold end of the regenerator. The mean temperature rises from to
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(subscripts c and h denote cold end and hot end of regenerator, respectively). The increase of causes a decrease of (mean density of gas). Since to first-order the mass flux (where is volume velocity rate) is nearly independent of x, the result is that the volume velocity increases, namely Because the time-averaged energy flux through the regenerator is small, the acoustic power flowing out of the hot heat exchanger is nearly equal to the heat flowing into the hot heat exchanger. The pressure difference is caused by the viscosity in the regenerator, which is proportional to a weighted average of through the regenerator. The viscous effects are largest at the hot end of the regenerator. Hence, with dominating, leads In the various Figure 3 curves, the phase difference between and varies significantly, from 204 degree at high frequency onset to 22 degree at fundamental frequency onset. The temperature profiles in Fig. 2 show that the temperature at the hot end of the regenerator is still increasing gradually after the onset of high frequency oscillation. At the highest temperature of 750 K, the pressure oscillation jumps to the fundamental frequency. The temperature drops (about 150 K) abruptly, as well as the pressure amplitude (about 60~70%). For gradually increasing mean pressure, the phenomena remain the same as described above. Keeping the experimental conditions constant and changing the working gas to helium, the pressure oscillations at the points of and the temperature distribution along the regenerator are shown in Fig. 4. For this case, the onset of pressure oscillation is directly at the fundamental frequency of 196 Hz, and the high frequency onset mode does not appear. The phase difference between and is 23 degrees in steady state. When the mean pressure in the thermoacoustic-driven system increases, the
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onset of the oscillating pressure is at a high frequency of 1406.25 Hz, which then transfers to a steady fundamental frequency of 196.5 Hz. The phase difference between and is the same approximately as when nitrogen was used. Analyzing the frequency ratio (2.57) at high frequencies and (2.6) at fundamental frequencies for both helium and nitrogen, we find the results are roughly the same, and approximately equal to the ratio of the acoustic velocity of helium and nitrogen under the same working conditions.
120-Mesh Screen Regenerator With wire screens of mesh number 120, the pressure amplitude at the points of and the temperature distribution along the regenerator are shown in Figs. 5 and 6 for nitrogen and helium, respectively. The other conditions are the same as above. With nitrogen as the working gas, the mean pressure is increased step by step. When the mean pressure approaches 0.7 MPa, pressure oscillation initiates at high frequency; the transformation conditions are the same as described above.
150-Mesh Screen Regenerator With wire screens of mesh number 150, the pressure amplitude at the points of and the temperature distribution along the regenerator are shown in Figs. 7 and 8 for nitrogen and helium,
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respectively. The other conditions are the same as above. Increasing the mean pressure in the thermoacoustic-driven system step by step up to 1.5 MPa, the onset of pressure oscillation always occurred at the fundamental frequency.
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RESULTS AND DISCUSSION By comparing the oscillating pressure and the temperature profiles along the regenerator with different mesh numbers of screens, there exists a subtle change in the temperature profiles with increasing mesh number. For the 80-mesh case, at the instant of fundamental-frequency oscillation, the temperatures of all measured points are decreasing, and in particular, the temperature at the hot end drops abruptly, which results in rapid dropping of the oscillating pressure amplitude. With increasing mesh number, at the instant of fundamental frequency oscillation, the temperature at the hot end of the regenerator rises a little and then drops abruptly, but the pressure amplitude changes more evenly. The temperature of the other points changes from 'all dropping' to 'some dropping and some rising'. For different meshes and working gases, the general trend of the temperature profiles is the same (see Fig. 9). Oscillating pressure amplitude increases with increasing mesh number. For helium, the temperature gradient in the regenerator increases, while for nitrogen the condition is the reverse. The reason could be related to the properties of the gases (see Fig. 10). According to the measured temperature and frequency data, the viscous penetration depth has been calculated for different places in the regenerator. Ratios of viscous penetration depth over hydraulic radius are shown in Figs. 11 and 12, based on the temperature profiles of Fig. 2 and Fig. 8. In Fig. 11, the ratios are small at the onset of high frequency oscillation. As the temperature is increased, the ratios rise. When the ratio reaches about 3.3 at the hot end, the resonance fre-
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quency changes to the fundamental mode with an abrupt increase in the longitudinal oscillating pressure; this results in the time-average mass flux becoming large and the temperature dropping abruptly. The system arrives at a stable oscillating state while reaching a balance point. In addition, the ratio of relative penetration depth determines how the longitudinal oscillation is affected by the transverse penetration layer. This results in a feedback mechanism of thermodynamics and hydrodynamics (see Fig. 3), as the phase difference between and is significantly different between high frequency onset and fundamental frequency onset. However, for high mesh numbers (150 mesh), shown in Fig. 12, the ratios are large, about 11.8 at the hot end, and the thermoacoustic-driven system initiates oscillation directly at the fundamental frequency.
CONCLUSION A traveling wave thermoacoustic system, designed with a resonant frequency at 200 Hz, has been established. Its highest fundamental frequency is at 196 Hz. The eigen characteristics of the two subsystems can be significantly different, so the thermoacoustic-driven system contains a nonlinear coupling problem that depends on the system configuration. To explore the effects on system coupling, experimental investigations have been conducted by changing the wire-screen mesh size, the working gas, and the mean pressure in the system. Experimental and theoretical analyses demonstrate that it is advantageous to use higher mesh numbers for the wire screens in the regenerator. This improves coupling of the subsystems, helps suppress high eigen frequencies of the loop subsystem, and promotes oscillation at the fundamental frequency. Nevertheless there are some disadvantages such as higher onset temperature (for helium) with high mesh number. It is observed that higher mean pressure is much more prone to cause nonlinear phenomenon such as high frequency oscillation, and helium is better than nitrogen as a working gas in suppressing the nonlinear phenomenon. The ratio of relative penetration depth plays an important role in these nonlinear effects. The system starts to oscillate with a high frequency at then switches to the fundamental frequency when reaches 3.3 at the hot end of the regenerator. When the ratio reaches 5.5 at the cold end of the regenerator and 11.8 at hot end of the regenerator, the system initiates oscillation directly at the fundamental frequency. With these data, the onset mode and subsystem coupling can be estimated in advance by quantifying experimentally the ratio of relative penetration depth.
ACKNOWLEDGMENTS The authors gratefully acknowledge the support by the funds of “Talents of Oversea” from CAS and by K.C. Wong Education Foundation, Hong Kong.
REFERENCES 1. 2. 3. 4. 5. 6.
7.
Rott, N., “Thermoacoustics,” Adv. Appl. Mech., vol. 20 (1980), pp. 135-175. Swift, G.W., “Thermoacoustic engines,” J. Acoust. Soc. Am , vol. 84, no.4 (1988), pp. 1145-1180. T. Yazaki, A. Iwata, “Traveling wave thermoacoustic engine in a looped tube,” Physical Review Letters, 81 (1998), pp. 3128-3131. S. Backhaus, G.W. Swift, “A thermoacoustic Stirling heat engine,” Nature, vol. 399 (1999), pp.335338. Dong, K., Li, Q., Wu, J., Guo F., “Network modeling of thermoacoustic devices,” Proc. of ICEC 18, India (2000), pp. 705-709. Wu Jihao, Li Qing, Guo Fangzhong, “Experiment research on high frequency traveling wave thermoacoustic engine,” Proc. of the 5th National Cryogenics Conference, Dalian, China (2001), pp. 20-24 (In Chinese). Wu Jihao, Li Qing, Yu Zhibin, Tu Qiu, Guo Fangzhong, “Research of the buildup of oscillations in a traveling wave thermoacoustic prime mover,” ICEC 19, 2002. Grenobel, France (In Press).
The Influence of Thermal Natural Convection on a Traveling-Wave Thermoacoustic Engine H. Liu, E. Luo, H. Ling, J. Wu Technical Institute of Chemistry and Physics Chinese Academy of Science Beijing 100080, China
ABSTRACT A traveling-wave thermoacoustic engine is a new kind of prospective machine that has been studied by many scientists. However, until now, little has been reported on the influence of natural convection on the operation of the thermoacoustic engine. In this paper, the influence of natural convection is examined by running a series of experiments using different orientations with respect to gravity. It is found that with changing orientation, natural convection can cause a variation of the onset temperature where thermal oscillation begins. In addition, the temperature distribution in the regenerator and other important parameters are also affected. In general, adverse natural convection worsened the performance of the system, and in one case even caused the thermoacoustic engine to stop work.
INTRODUCTION Natural convection, a kind of fluid movement caused by buoyancy, has been systematically studied by numerous scientists. Thermoacoustic engines, a type of heat engine with no moving parts, have recently received considerable interest because of their simple construction and promise of high reliability. One such engine is the so-called thermoacoustic-Stirling heat engine that was recently introduced by S. Backhaus.1 In acoustic sense, this engine belongs to the class of traveling-wave thermoacoustic engines. In this machine, there is a so-called thermal buffer with a large temperature gradient. This and other parts of the unit are potentially sensitive to orientation-dependent convection. Since few investigations about the influence of natural convection have been reported, an experimental setup was built to investigate the influence of natural thermal convection on the startup performance and equilibrium operation of a traveling-wave thermoacoustic engine.
EXPERIMENTAL SETUP Figure 1 shows a schematic flow diagram of the studied traveling-wave thermoacoustic engine. In the engine, near the top of the torus, is the main cold-heat-exchanger; this is cooled by chilled water. Below the main cold-heater-exchanger is the regenerator, made from a 7-cm-tall stack of 120-mesh stainless-steel screens. The diameter of the screen wire is 65 µm, and its hydraulic radius is 42 µm; the length of the thermal buffer tube is 20 cm. In our thermoacoustic engine, the diameter of all tubes is 5 cm. Twenty NiCr-NiSi thermocouples were used to measure the Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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distribution of temperatures, and 16 of them were symmetrically attached to the thermal buffer tube, with the others attached to the regenerator as shown as Fig.2. Additional instrumentation included a multimeter (Keithley 2700) to collect all temperatures, and a digital oscilloscope (Tektronix TDS3000) to view oscillating pressures. By changing the spatial orientation of the thermoacoustic engine, the variation of the temperature distributions was observed, both in the thermal buffer tube and in the regenerator; the change in the onset temperature of the thermoacoustic engine was also noted. Four spatial arrangements of the thermoacoustic engine were tested as shown in Figs. 3 to 6. From Case 1 to Case 4, the cold end directions of the thermal buffer tube relative to the gravity are 0°, 45°, 90°,and 135°, respectively.
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EXPERIMENTAL RESULTS AND ANALYSIS In the experiments, nitrogen was used as the working gas, and the operating pressure was about 1.5 MPa. The onset performance for the four cases is given in Table 1. Figures 7 to 9 show the temperature distributions of the thermal buffer tube when oscillation was stable (case 1 to case 3). Figure 10 presents the temperature distribution of the thermal buffer tube for Case 4 after heating for 55 minutes. Figure 11 shows the temperature distribution of the regenerator when the self-excited oscillations just begin. Figure 12 shows the temperature distribution of the regenerator when the oscillation is stable. In Figs. 7 to 10, the abscissa represents the
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distance from the measurement point to the bottom of the hot heat exchanger, and the ordinate represents the measured temperature. In Figs. 11 and 12, the abscissa represents the distance from the measurement point to the top of the hot heat exchanger, and the ordinate represents the measured temperature. Based on the above experimental data, it is clear that spatial orientation can have an important effect on the onset-performance of a traveling-wave thermoacoustic engine. With increasing angle between the cold end of the thermal buffer tube and the gravity direction, the temperature gradient of the regenerator becomes large at the onset of oscillations. However, the temperature gradient of the regenerator became small when the oscillations became stable.
CONCLUSION An experimental setup of a traveling-wave thermoacoustic engine was used to observe the affect of natural convection on the oscillation of a thermoacoustic engine. The experiment showed that there is a modestly strong effect, and that the orientation dependence is similar to that seen in the convective conduction in pulse tube refrigerators.3,4
ACKNOWLEDGMENT This research is financially supported by the National Sciences Foundation of China under contract No.59976044.
REFERENCE 1. S. Backhaus and G.W. Swift, ”A Thermoacoustic-Stirling Heat Engine: Detailed Study,” J. Acoust. Soc. Am., 107 (6), June 2000.
2. D. Geodon, “DC Gas Flows in Stirling and Pulse-tube Cryocooler,” Cryocoolers 9, Plenum Press, New York (1999), pp. 385-392.
3. Thummes, G., et al., Convective Heat Losses in Pulse Tube Coolers: Effect of Pulse Tube Inclination,” Cryocoolers 9, Plenum Press, New York (1999), pp. 393-402.
4. Ross, R.G., Jr., “AIRS Pulse Tube Cooler System-Level and In-Space Performance Comparison,” Cryocoolers 12, Kluwer Academic/Plenum Publishers, New York, 2003.
Experimental Investigation of Thermoacoustically Driven Pulse Tube Refrigerator Using Noble Gas Mixtures G.B. Chen, K. Tang, T. Jin, Y. Shen and Y.H. Huang Cryogenics Laboratory, Zhejiang University Hangzhou 310027, P.R. China
ABSTRACT Based on the estimation of thermophysical properties of binary noble gas fluids, He-Ar mixtures are used as working fluids to improve the performance of a pulse tube refrigerator driven by a thermoacoustic engine. Computed and experimental results are reported in detail, and some discussion is given. With 20% molar fraction of Ar the refrigeration temperature was lowered by 3 K as compared to pure He.
INTRODUCTION A pulse tube refrigerator driven by a thermoacoustic engine, instead of an ordinary mechanical compressor, has no moving parts except for the motion of the working fluids. Mechanical abrasion and sliding seals are eliminated in the system, which should enhance reliability and longevity. Moreover, the machine may be driven by thermal energy, such as flue gas, waste heat, solar energy, etc. The application of low-quality heat energy will not only increase the overall thermal efficiency, but also be attractive in areas that lack electricity, such as deserts and off-shore remote locations associated with the exploration of oil and natural gas. Helium is the typical working fluid for pulse tube refrigerators. In the 1990s, the possibility of using mixtures to improve the performance of pulse tube refrigerators was discussed. Results of both computations and experiments showed that the performance can be improved with some mixed fluids in certain refrigeration temperature zones.1-2 It has also been shown that the thermodynamic, heat transfer and flow properties of binary noble gas mixtures, such as He-Ar and He-Kr, are not comparable with those of pure helium in pulse tube refrigeration, so they do not benefit pulse tube refrigerators.3-5 However, it is possible for a thermoacoustically driven pulse tube refrigerator to improve its performance with noble gas mixtures. The performance of this machine depends on both the pulse tube refrigerator and the thermoacoustic engine. Analysis6 indicates that the thermoacoustic conversion effect can be intensified by using noble gas mixtures and a more appropriate pressure wave can be delivered to the pulse tube refrigerator from the thermoacoustic engine. Under these circumstances, heat transfer and flow losses in the pulse tube refrigeration process resulting from the noble gas mixture may be offset by the improved thermoacoustic performance. Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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The principle of thermoacoustic self-oscillation indicates that the thermal penetration depth and the viscous penetration depth of the working fluid plays an essential role in thermoacoustic phenomena.7 The former is the factor enhancing the thermoacoustic effect, while the latter is the one causing the loss of kinetic energy. Therefore, increasing and decreasing will improve the thermal efficiency of the thermoacoustic conversion. As the ratio of and is the square root of Prandtl number (Pr), a working fluid with a smaller Prandtl number is more beneficial for efficiency improvement. The Prandtl number of helium is about 2/3, and that of binary mixtures of helium and other noble gases is less. Thus, the mixtures can be used to raise the efficiency of the thermoacoustic conversion and to enhance the output pressure wave of the thermoacoustic engine. On the other hand, when the structure and size of the thermoacoustic engine are fixed, the resonance frequency f in the resonant tube is proportional to the sonic velocity a of the working fluid. The sonic velocity of the binary mixtures of helium and other noble gases is smaller than that of pure helium. Therefore, the resonance frequency will be reduced and the frequency matching condition between the thermoacoustic prime mover and the pulse tube refrigerator will be improved with the use of binary mixtures of helium and other noble gases.
THERMOPHYSICAL PROPERTY OF NOBLE GAS MIXTURES The Prandtl number (Pr) and the sonic velocity a of mixtures to be used in a thermoacoustically driven pulse tube refrigerator are very important parameters for the thermoacoustic effect and the lower frequency output of acoustic flow. In order to estimate the Prandtl number (Pr) of binary noble gas mixtures, the dynamic viscosity µ, the thermal conductivity K, and the specific heat need to be computed first. The fundamental equations are as follows:8
where is the specific heat ratio, is universal gas constant, T is thermodynamic temperature, is molar weight of the binary mixture, and and are the molar fractions and molar weights of components, respectively. Equations (1) and (2), which are the approximate formulas for pure gases, can be used to compute the dynamic viscosity and the thermal conductivity of mixtures by using the mixing law. Considering the mixtures as monatomic ideal gases, Eq. (3) can be adopted to calculate the specific heat The Prandtl number of binary gases may be calculated by Eq. (4). As the temperatures of the cold and hot ends of the thermoacoustic stack are about 350K and 750K, respectively, in the experiment, 550K was selected as the computing temperature. The calculation results shown in Fig. 1 denote that the Prandtl numbers of the binary mixtures composed of helium and other noble gases are all smaller than the 2/3 value of pure helium. We can see, that the minimum Prandtl number appears when the molar fraction of the other noble gas is about 35% for each curve in Fig. 1. The minimum Pr decreases with the increase of the molecular weight of the other noble gas (see Fig. 2). The minimum Prandtl number of a helium-nitrogen mixture, which is estimated with extrapolation, is about 0.1; this is the minimum one of this kind of mixtures. The sonic velocity of the mixtures may be estimated by Eq. (5):
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Considering most of the resonator is at room temperature, 300K is adopted as the computing temperature. The resonance frequency may be estimated by means of the fluid impedance method.9 For a resonator with two buffers at the both ends, the frequency equation is:
where L and A are length and flow area of the resonator, respectively, V is the volume of the buffer, a is the sonic velocity, and f is the resonance frequency. e relations of sonic velocity and resonance frequency versus other noble gases’ molar fraction are computed and shown in Fig. 3. Figure 3 indicates that the sonic velocity and the resonance frequency decrease with an addition of the molar fraction of other noble gases. For example, the computed resonance frequency of the mixture of helium (65%) and argon (35%) is 50% lower than that of pure helium. In short, the computed results indicate that the Prandtl numbers of the binary mixtures of helium and other noble gases are smaller than that of pure helium, while their resonance frequencies are lower. Thus binary noble gas mixtures can be used to enhance the thermoacoustic effect and to improve the frequency matching between the thermoacoustic engine and the pulse tube refrigerator.
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EXPERIMENTAL SETUP The experimental setup of our thermoacoustically driven pulse tube refrigerator is shown in Fig. 4; it includes a thermoacoustic engine, pulse tube refrigerator, and measuring system. The thermoacoustic engine mainly consists of a heater, screen stack, cold heat exchanger, resonator, and hot buffer. The dimensions of the main parts are tabulated in Table 1. A coaxial single-stage pulse tube refrigerator is connected to the prime mover. Considering the relatively high temperature of the gas flowing from the prime mover, a water pre-cooler is fixed before the pulse tube refrigerator. Temperature measuring locations are also shown in Fig. 4. T2 & T3 and T4 & T5 are fixed at the cold and hot ends of the stack, each side of the resonator, respectively, and T1 & T6 are located in the hot buffers. Temperatures at the hot buffer and the hot end of the stack are measured with NiCr-NiSi thermocouples, while those at the cold ends of the stack are measured with Cu-Constantan thermocouples. A Rh-Fe resistance thermometer (with 0.1K accuracy) is applied to measure the refrigeration temperature at the cold end of the pulse tube. Pressure measurement is accomplished by a PC-based digital data acquisition system, which includes piezoresistive silicon pressure sensor, data acquisition card (NI product) and PC.
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EXPERIMENTAL RESULTS AND DISCUSSIONS A number of experiments have been carried out to verify the positive effect of utilizing noble gas mixtures as working fluids for the thermoacoustically driven pulse tube refrigeration. Considering its availability, a He-Ar mixture was adopted for the experimental investigation. With the same heating power, 1230W, we set the working pressure to a fixed value, such as 1.8, 2.2, or 2.6 MPa, to observe the relationship between pressure amplitude and pressure ratio versus molar fraction of argon. The results are shown in Figs. 5 and 6, respectively. We can see that the pressure amplitude and the pressure ratio gradually increase with an increase of the molar fraction of argon. This means that the use of binary mixtures of helium and other noble gases can raise the efficiency of the thermoacoustic effect and enhance the output pressure wave of the thermoacoustic engine due to their smaller Prandtl number. The relation between resonance frequency and molar fraction of argon is presented in Fig. 7, which shows that the resonance frequency decreases with an increase of argon fraction. The influence of the working pressure on the resonance frequency is not obvious. Besides, computed results by the fluid impedance method agree with the experimental values. This indicates that the fluid impedance method is acceptable to predict the influence of the mixture on the resonance frequency. The relation between cooling temperature of the thermoacoustically driven pulse tube refrigerator and argon’s molar fraction in the He-Ar mixture under different heating conditions is presented in Fig. 8. Compared with pure helium, a lower refrigeration temperature is obtained with He-Ar mixtures. In this case, an increase of the amplitude and the pressure ratio, or a decrease of the resonance frequency of the pressure wave in the thermoacoustic engine, will offset thermal and pressure losses in the regenerator of the pulse tube refrigerator to a certain extent. As a result, the refrigeration performance is improved. However, when the fraction of argon in the mixture increases further, the refrigeration temperature rises. This means that the increase of the pressure wave in the thermoacoustic engine is not able to make up for the serious deterioration of heat
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transfer and flow in the regenerator of the pulse tube refrigerator. We find that adding a small fraction of argon (less than 20%) into the helium can improve the performance of the thermoacoustically driven pulse tube refrigerator, while a greater amount of argon in the binary mixture makes the refrigeration temperature increase. With a heating power of 2200W, using He (79.7%)-Ar (20.3%) mixture as working fluid, a refrigeration temperature as low as 117.6K has been attained by the thermoacoustically driven pulse tube refrigerator; this is lower by about 3K than that achieved with pure helium.
CONCLUSIONS Binary mixtures of helium and other noble gases are beneficial to enhancing the thermoacoustic effect and can improve the match between a thermoacoustic engine and a pulse tube refrigerator. The experiments show that a He-Ar mixture with 20% argon can improve the performance of a thermoacoustically driven pulse tube refrigerator. With a heating power of 2200 W, using a He(79.7%)-Ar(20.3%) mixture as the working fluid, a cooling temperature of 117.6 K has been achieved from a pulse tube refrigerator driven by a thermoacoustic engine; this is 3K lower than that achieved with pure helium.
ACKNOWLEDGMENT The project is financially supported by the National Natural Sciences Foundation of China (59976034), the University Doctoral Subject Special Foundation of China (20010335010) and the Foundation for the Author of National Excellent Doctoral Dissertation of P.R. China (200033).
REFERENCES 1.
Chen, G.B., Gan, Z.H., Thummes, G., Heiden, C., “Thermodynamic performance prediction of pulse tube refrigeration with mixture fluids,” Cryogenics, vol. 40, no. 4-5 (2000), pp. 261-267.
2.
Gan, Z.H., Chen, G.B., Thummes, G., Heiden, C., “Experimental study on pulse tube refrigeration with helium and nitrogen mixtures,” Cryogenics, vol. 40, no. 4-5 (2000), pp.333-339.
3.
Daney, D.E., “Regenerator performance with noble gas mixtures”, Cryogenics, vol. 31, no. 10 (1991), pp. 854-861.
4.
Yu, J.P., Chen, G.B., Gan, Z.H., et al, “Discussion on regenerator performance improvement with binary gas mixture,” Proceeding of ICEC-17, Institute of Physics Publishing (1998), pp.117-122.
5.
Yu, J.P., Chen, G.B., Gan, Z.H., “Investigation on the regenerator performance using gas mixtures,” Cryogenics and Refrigeration-Proc. of ICCR’98, Hangzhou, China (1998) IAP, pp.409-412.
6.
Tijani M.E.H., Zeegers J., de Waele A., “The experimental study of the influence of the Prandtl number and the spacing in the stack on the thermoacoustic cooler performance,” First Int’l Workshop on Thermoacoustics, the Netherlands, April 2001: A6.
7.
Swift, G.W., “Thermoacoustic engines,” J Acoust Soc Am, vol. 84 (1988), pp. 1145-1180.
8.
Hirschfelder, J.O., Curtiss, C.F., Bird, R.B., Molecular Theory of Gases and Liquids, John Wiley & Sons, New York (1954).
9.
Chen, G.B., Jiang, J.P., Shi, J.L., Jin, T., Tang, K., et al, “Influence of buffer on resonance frequency of thermoacoustic engine,” Cryogenics, vol. 42 (2002) (in press).
Low Temperature Cryocooler Regenerator Materials K.A. Gschneidner, Jr.1,2, A.O. Pecharsky1, and V.K. Pecharsky1,2 1
Ames Laboratory and 2Department of Materials Science and Engineering Iowa State University, Ames, Iowa 50011-3020, USA
ABSTRACT There are four important factors which influence the magnitude of the magnetic heat capacity near the magnetic ordering transition temperature. These include the theoretical magnetic entropy, the deGennes factor, crystalline electric field, and the RKKY (RudermanKittel-Kasuya-Yosida) interaction. The lattice contribution to the heat capacity also needs to be considered since it is the sum of the lattice and magnetic contributions which give rise to the heat capacity maxima. The lattice heat capacity depends on the chemical composition, crystal structure and temperature. As a result, one can obtain large changes in the heat capacity maxima by alloying. Several ternary intermetallic systems have been examined in light of these criteria. A number of deviations from the expected behaviors have been found and are discussed.
INTRODUCTION The use of lanthanide materials as low temperature regenerators has led to some significant advances in low temperature (<20 K) cryogenics since 1990. The first material to be utilized was which replaced Pb in the low temperature stage of a two-stage Gifford-McMahon (G-M) cryocooler.1,2 This modification enabled Toshiba scientists to lower the low temperature limit of the G-M cryocooler from ~10 K to ~4 K. Subsequently, several other lanthanide materials, in particular and (Ref. 4), have been utilized for cooling down to ~4 K More recently Er and Er-Pr alloys (up to 50 at.% Pr) have been suggested as a replacement for Pb5,6 as the intermediate temperature (~10 to ~60 K) range regenerator material. Today research is still being continued on finding improved regenerator materials, especially below 10 K, in order to enhance the performance of Stirling, G-M, and pulse tube cryocoolers. In this quest for new and better magnetic regenerator materials there are no predictive theories or first principles computational method which allows one to foretell the composition of an alloy or intermetallic compound with the desired thermal properties, i.e. magnetic ordering temperature and total heat capacity. There are, however, a few guiding principles which can be utilized to find such materials, and these are described below.
GUIDING PRINCIPLES Magnetic Entropy The high heat capacity of a magnetic lanthanide material is strongly dependent on the magnetic entropy associated with the magnetic ordering process. The theoretical molar magnetic Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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entropy,
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is proportional to the total orbital quantum number J and is given by
where R is the gas constant. A plot of vs. the lanthanide atomic number is shown in Fig. 1. There are two maxima: the minor one in the middle of the light lanthanides centered around Nd, and the major one in the middle of the heavy lanthanides centered around Ho. The magnetic entropy is divided up among several processes which may occur in these elements. Generally, the major one is the consumption of by the magnetic ordering process(es). Many times, however, lanthanide materials will exhibit more than one kind of magnetic ordering, and so, will be divided among them. Usually this is unfortunate for cryocooler applications, especially if the two ordering temperatures are far apart, since some of the entropy will not be available at the desired temperature. Sometimes this may be useful if the two magnetic ordering processes occur only a few degrees apart, such as in see Fig. 2. Also shown in Fig. 2 is the volumetric heat capacity of which exhibits only one magnetic transition. As one can see the maximum value of the heat capacity is nearly the same for the two
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peaks and about 10% larger than that of the peak. But for the heat capacity which is greater than occurs over a wider temperature span (~5 to ~10 K) than for (~6 to ~8.5 K), and therefore, because of these two reasons is a better regenerator for cooling below 10 K. As a matter of fact ~90% of the <10 K cryocoolers utilize as the regenerator material.7 Spin fluctuations of the magnetic 4f electrons above the highest magnetic ordering temperature generally consume about 15% of Thus, one would expect that about 85% of the magnetic entropy is associated with the magnetic ordering process(es).8 But crystalline electric field effects, see next subsection, also compete for this entropy, and thus could lead to a reduction in the magnetic heat capacity at the ordering temperature.
Crystalline Electric Field (CEF) When a lanthanide ion is placed in a solid, the 4f levels are split by the electric point charges of the surrounding atoms (both lanthanide and non-lanthanide). A generalized CEF energy level scheme is shown in Fig. 3 for the trivalent lanthanides. The actual energy levels will vary considerably from compound to compound depending upon the electric point charges and the point symmetry around a specific lanthanide ion. But for a given isostructural compound series, in which only the lanthanide atom varies, the relative crystal field levels across the lanthanide series will have a similar pattern, as shown in Fig. 3 for the cubic case. For a different compound series the energy scale may shift as much as an order of magnitude or more. The entropy utilization occurs as a sample is warmed because the 4f electrons in the lowest level are thermally excited to higher CEF levels. This process shows up as a broad bump in the heat capacity and is known as a Schottky anomaly. In Fig. 4 we compare the broad Schottky heat capacity anomaly for Er3Ni9 with the very sharp magnetic ordering peak of HoSb.10 The Schottky anomaly in accounts for the large heat capacity above the magnetic ordering peak in see Fig. 2, where it is significantly larger than that of above 10 K (and about equal to that of Pb up to ~20 K). As noted by Takahashi et al.9 about 75% of the entropy is utilized in thermal excitations of the 4f electrons from low CEF levels to higher ones, while only 25% is involved in the magnetic ordering of the Er 4f electrons in
Magnetic Ordering Temperature The magnetic ordering temperature in a given lanthanide compound series tend to follow the deGennes factor, F,
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where g is the gyromagnetic ratio and is given by
where S is the spin quantum number and L is the orbital angular momentum quantum number. The variation of the deGennes factor as a function of the lanthanide atomic number is shown in Fig. 5. It should be mentioned that CEF effects can cause anomalies in the sequence of magnetic ordering temperatures expected from the deGennes factor. For example, NdGe is expected to have a higher Curie (ferromagnetic ordering) temperature than PrGe (see Fig. 5), but in fact, NdGe orders at 18 K while PrGe orders at 30 K.11 As a whole, the deGennes factor is a fairly good guide for estimating the ordering temperatures in a given lanthanide compound series if
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one (or two) of the magnetic ordering temperatures is(are) known. In general the deGennes factor cannot be used to estimate ordering temperatures for a given compound series from another series.
The RKKY Interaction Since there is little if any direct overlap of 4f electron wave functions from one lanthanide metal to another, the magnetic coupling arises from the polarization of the 4f electrons on neighboring lanthanide atoms via the conduction (6s) electrons. The spin polarization of the conduction electrons is spatially non-uniform and has an oscillatory nature, i.e. where and is the Fermi wave vector and R is the radial distance from a given lanthanide atom. This is schematically shown in Fig. 6, and is known as the Ruderman-KittelKasuya-Yosida (RKKY) interaction.11 The interaction between adjacent lanthanide ions will be ferromagnetic when the neighboring lanthanide atom lies in a region of where F(x) is positive, and it will be antiferromagnetic when F(x) is negative. The strength of the interaction is given by the amplitude of the function F(x). Because of the lanthanide contraction (i.e. the size of the lanthanide atoms decreases with increasing atomic number), it is possible for the magnetic ordering to change from ferromagnetic to antiferromagnetic in a given compound series, provided the sign of F(x) changes as varies from R = Ce, the first magnetic trivalent lanthanide, to R = Yb, the last trivalent lanthanide. For example, the RGe compounds are ferromagnetic for R = Ce, Pr, and Nd, while those for R = Sm, Gd, Tb, Dy, Ho, and Er are antiferromagnets.11 The variation of the RKKY interaction may not play an important of a role in binary intermetallic compounds, especially with respect to cryocooler regenerator materials. However, in ternary alloys it may play a role when substituting one lanthanide for another R [e.g. or substituting one non-rare-earth metal for another M, to adjust the magnetic ordering temperature or the magnetic heat capacity.
IMPROVING REGENERATOR PROPERTIES A number of materials have interesting magnetothermal properties for their application as magnetic regenerators, but the ordering temperature is not appropriate for their optimum
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performance. Recently, we have shown that the magnetothermal properties for Er metal as a replacement for Pb as a 10 to 50 K regenerator material was significantly enhanced by alloying with Pr metal.5,6 This has also been done for several binary intermetallic compounds3, for example: and
Lattice Heat Capacity One trend that is generally observed in these pseudobinary systems is that the heat capacity maximum increases with increasing ordering temperature. As seen in Fig. 7, the lattice heat capacity increases fairly rapidly with increasing temperature up to (where is the Debye temperature), which is the point where the rate of increase of the heat capacity slows down (i.e. the knee of the heat capacity curve). For example, orders at 6.7 K and orders at 20.6 K, and the heat capacity maximum increases from for to for see Fig. 8. As one notes the magnetic ordering temperature for intermediate
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compositions in the system essentially fall on a straight line connecting the two end members, which is consistent with the deGennes factor (see above). The heat capacity maxima increase as expected with increasing magnetic ordering temperature, but two of the data points fall slightly below the straight line connecting the end members, i.e. and while the heat capacity maximum of is significantly higher. Within experimental error the two points which lie below this line may not be anomalous, but the deviation of the value is probably real. At the present time, the reason for the anomaly is unknown, but several possibilities can be ruled out. The magnetic entropies for Er and Dy are equal, see Fig. 1, so a change in as a function of x could not account for it. Since and have the same crystal structure (the cubic C15 Laves phase structure), and since the metallic radii of Er (1.7566 Å) and Dy (1.7740 Å) are nearly the same (differ by 1%), the CEF is essentially the same from to and it is difficult to see how this could cause a change in the magnetic heat capacity peak at just the x = 0.5 composition. Similarly, because of the small size difference, it is doubtful that a crystal structure change occurs at x = 0.5 (e.g. the x= 0.5 alloy could possibly have the hexagonal C14 Laves phase structure), while the other alloys maintain the cubic C15 Laves phase structure. It is possible, however, that the RKKY interaction is such that at x = 0.5 the F(x) value is a maximum or a minimum, and that for x = 0.4 and 0.6 the absolute value of F(x) approaches zero (see Fig. 6) giving rise to the observed behavior, but this would imply a very short period. Clearly, more detailed studies are required to understand this behavior.
The
System
The system3 is quite interesting, since the observed behavior is counter intuitive of what one would expect. ErNi orders at 13.0 K and has a heat capacity maximum of while for x = 0.1 the ordering temperature drops as expected to 9.5 K on the basis of the deGennes factor, see Fig. 5, but the heat capacity maximum increases by more than 33% to A significant decrease in heat capacity maximum is expected for two reasons: 1) the lattice heat capacity of should be smaller because of the lower ordering temperature, and 2) the magnetic entropy of Yb is significantly smaller than that of Er (by ~25%).
The
System
A third example of the occurrence of some unexpected behaviors when a binary compound is alloyed to modify its properties is found in the system.12 As noted above for the system, the values of Er and Dy are the same and their metallic radii only differ by about 1%. The two end members and all compositions between them have the cubic C15 Laves phase structure, and the lattice parameters essentially obey Vegard’s linear approximation. The magnetic ordering temperature rises in an approximately linear fashion from 13.5 K for to 63.9 K for see Fig. 9. From ~x = 0.4 the observed behaviors are as expected, the maximum heat capacity value rises, as does the lattice heat capacity, with increasing x to pure As noted in Fig. 9, the magnetic entropy also increases because the fraction Er decreases and because there is a CEF contribution from Er at ~8 K which accounts for ~25% of and thus, as the percentage of Er in the alloy decreases more of the becomes available for utilization in the magnetic ordering process. Things, however, become more interesting on the Er-rich side of the system. The large heat capacity peak in which is due to a second order paramagnetic to ferromagnetic transition at 13.6 K, drops drastically by more than 50% when 10% of the Er is substituted for by Dy even though the transition temperature rises to ~18 K, see Fig. 9. This behavior is not quite understood. Even more surprising is the existence in of a low temperature (~10 K) first order magnetic transition where the easy axis changes from the <100> for the high temperature phase to the <111> for the low temperature phase, and a high temperature (~25 K) second order paramagnetic to ferromagnetic transition.11 The heat capacity peak maxima are 0.94 and respectively. For higher Dy contents the behavior is normal as noted in the previous paragraph.
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Non-Rare-Earth Metal Substitutions One can also substitute one non-rare-earth metal by another non-rare-earth metal to raise or lower the magnetic ordering temperature, and one similarly expects the heat capacity maximum to increase as the magnetic order temperature increases, just as noted above. However, only limited number of such systems have been investigated. One of them is the system.3 The magnetic ordering temperature of is 6.7 K and that of is 33 K. As expected, the volumetric heat capacity of is and that of is However, the behavior of intermediate compositions seem to be somewhat anomalous from the expected behavior. The heat capacity of is slightly lower than that of while that of is significantly higher as expected, but it still lies below a straight line connecting the two end members, see Fig. 10. This anomalous behavior could be due to CEF effects and/or a change in the RKKY electron polarization, but more detailed work needs to be carried out on these alloys.
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CONCLUDING REMARKS The low temperature magnetic heat capacity of a given material can be modified by alloying either a rare-earth metal or a non-rare-earth metal. A number of factors apply and need to be considered when carrying out doping studies (alloying substitutions). These include the theoretical magnetic entropy, the deGennes factor, crystalline electric field effects, the RKKY interaction, and the lattice heat capacity. The third and fourth factors (CEF and RKKY) are difficult to predict a priori and are strongly dependent on the crystal structure of the intermetallic compound. Systematic trends are known and reasonably well established but significant and unexpected deviations occur as the alloying agent’s concentration is varied. Thus a blending of an Edisonian approach and systematics is necessary to find improved crycooler regenerator materials for applications below 10 K.
ACKNOWLEDGEMENTS The applied aspects of this research were sponsored by Atlas Scientific, and the basic research portion by the Office of Basic Energy Sciences, Materials Science Division of the U.S. DOE. The authors also wish to acknowledge the assistance of David Kesse in carrying out some aspects of this work.
REFERENCES 1. Sahashi, M., Tokai, Y., Kuriyama, T., Nakagome, H., Li, R., Ogawa, M. and Hashimoto, T., “New Magnetic Material System with Extremely Large Heat Capacities Used as Heat Regenerators”, Adv. Cryogenic Eng., vol. 35 (1990), pp. 1175-1182. 2. Kuriyama, T., Hakamada, R., Nakagome, H., Tokai, Y., Sahashi, M., Li, R., Yoshida, O., Matsumoto, K. and Hashimoto, T., “High Efficient Two-stage GM Refrigerator with Magnetic Material in the Liquid Helium Temperature Region”, Adv. Cryogenic Eng., vol. 35 (1990), pp. 1261-1269.
3. Ackermann, R.A., Cryogenic Regenerative Heat Exchangers, Plenum Press, New York (1997), p. 98. 4. Satoh, T., Onishi, A., Umehara, I., Adachi, Y., Sato, K. and Minehara, E.J., “A Gifford-McMahon Cycle Cryocooler below 2 K” in Cryocoolers 11, Ross, R. G., Jr., ed., Kluwer Academic/Plenum Publishers, New York (2001), pp. 381-386. 5. Gschneidner, K.A., Jr., Pecharsky, A.O. and Pecharsky, V.K., “Ductile, High Heat Capacity, Magnetic Regenerator Alloys for the 10 to 80 K Temperature Range” in Cryocoolers 11, Ross, R. G., Jr., ed., Kluwer Academic/Plenum Publishers, New York (2001), pp. 433-441. 6. Kashani, A., Helvensteijn, B.P.M., Maddocks, J.R., Kittel, P., Feller, J.R., Gschneidner, K.A., Jr., Pecharsky, V.K. and Pecharsky, A.O., “Performance of a New Regenerator in a Pulse Tube Cooler”, Adv. Cryogenic Eng., vol. 47 (2002), pp. 985-991. 7.
Kuriyama, T., private communication (March 2001).
8. Bouvier, M., Lethuillier, P. and Schmitt, D., “Specific Heat in Some Gadolinium Compounds. I.
Experimental”, Phys. Rev. B, vol. 43 (1991), pp. 13137-13144.
9. Takahashi, A., Tokai, Y., Sahashi, M. and Hashimoto, T., “Specific Heat of a Regenerator Material Jpn. J. Appl. Phys., vol. 33 (1994), pp. 1023-1026. 10. Nakane, H., Hashimoto, T., Numazawa, T., Okamura, M., Kuriyama, T. and Ohtani, Y., Adv. Cryogenic Eng., vol. 45 (2000), pp. 397-402. 11. Kirchmayr, H.R. and Poldy, C.A., “Magnetic Properties of Intermetallic Compounds of the Rare Earth Metals” in Handbook on the Physics and Chemistry of Rare Earths, Gschneidner, K.A., Jr. and Eyring, L. eds., North-Holland Publishing Co., Amsterdam (1979), Chap. 14, pp. 55-230. 12. Gschneidner, K.A., Jr., Pecharsky, V.K. and Malik, S.K., “The Alloys as Active Magnetic Regenerators for Magnetic Refrigeration,” Adv. Cryogenics Eng., vol. 42 (1996), pp. 475483.
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Specific Heat and Magnetic Properties of GdSb H. Nakane, S. Yamazaki, T. Yamaguchi, S. Yoshizawa, and T. Numazawa Kogakuin University, Meisei University National Institute for Materials Science Tokyo, 163-8677 Japan
ABSTRACT HoSb and DySb compounds have a very high specific heat peak below 10 K, and the measured specific heat of GdSb has a distinct peak at 25 K. Analyses have been conducted focused on the magnetic properties of these compounds to understand the difference between samples with sharp specific heat peaks and those with broad peaks. The magnetization of the samples was measured with a Vibrating Sample Magnetometer (VSM). The magnetic properties of HoSb and DySb are small above 100 K, while the magnetic property of GdSb exists up to near room temperature. Although further analysis is necessary, the data suggest that the differences in the specific heats between the materials is mainly due to differences in their magnetic properties.
INTRODUCTION The efficacy of heavy rare-earth and antimony (Sb) compounds as regenerator materials has been studied, and the temperature dependence of volumetric specific heat, thermal expansion, and thermal conductivity of HoSb, DySb, and GdSb have been previously measured and reported.1 While HoSb and DySb compounds have a very high specific heat peak below 10 K, the existing measured value of GdSb has a distinct heat peak at 25 K, and its high value steadily increases up to 60 K. In order to find out why such a difference occurs in the specific heat between GdSb and HoSb and DySb, their temperature dependence of magnetization was measured with a Vibrating Sample Magnetometer (VSM), and the obtained data were analyzed. An attempt was made from these data to determine their magnetic specific heats focusing on GdSb, for which a large database of measured specific heat values exists up to a high temperature.
MEASUREMENT OF THE MAGNETIC PROPERTIES The magnetization of HoSb, DySb and GdSb were measured with a VSM. Their temperature dependence of magnetization with an applied external magnetic flux density of 1.0 T is shown in Fig. 1. The temperature dependence of magnetization of nickel (Ni) is also shown in Fig. 1 for the case where the sample is in a state of saturation magnetization.2 Measurement of magnetization at 0 T was also attempted. However, it turned out to be impractical, as the measurement was highly uncertain because the magnetization was unstable. Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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HoSb and DySb samples were found to have a large specific heat peak around 5 and 10 K, respectively, which agrees with their specific heat properties previously reported. Their properties above 100 K were small. As understood from their specific heat, when the materials are below 5 K for HoSb, and 10 K for DySb, they are in a region of antiferromagnetism. A very broad specific heat peak occurs for GdSb at a temperature around 25 K, and the large value continues up to higher temperatures. It can be concluded from the measured magnetization data for GdSb that antiferromagnetism occurs below 25 K, ferromagnetism occurs from 25 K to near room temperature, and paramagnetism occurs at room temperature. The magnetization was measured by supplying an external magnetic field up to 800 kA/m at 12, 42 and 300 K for GdSb, and 292 K for HoSb. The magnetizing curves are shown in Fig. 2. Unexpectedly, the antiferromagnetism curve of GdSb, which occurs at 12 K, was found to be very similar to its ferromagnetism curve at 42 K.
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DETERMINING THE MAGNETIC SPECIFIC HEAT heat
The specific heat C for temperatures over 4 K is expressed as the sum of the lattice specific the electronic specific heat and magnetic specific heat as follows:
is the specific heat caused by lattice vibration. When the Debye temperature is larger than the observation temperature T, at lower temperatures is proportional to In contrast, is the specific heat due to the thermal motion of the conduction of electrons. Free electrons in metals conform to Fermi statistics as limited by the Pauli principle that states that free electrons can occupy only a quantum state. Thus, free movable electrons are limited to approximately where the Fermi temperature has the value of 104 K. Because the electronic specific heat is proportional to the observation temperature T, is negligibly small at higher temperatures in comparison with which is proportional to The magnetic specific heat is defined by:
where the magnetic entropy N is the magnetic ionic number, is the Boltzmann constant, and J is the orbital angular momentum. However, as it is difficult to obtain by using Eq. (2), is estimated using the following variation of Eq. (1):
In the above equation, it is assumed that By substituting the relations of and
and
where a and b are constants. into Eq. (3), we get:
Assuming the values of a and b that correspond to the measured values of C were determined, i.e., a and b that best fit the properties of C were adopted, ignoring The magnetic specific heat is considered to be the features of the measure data that deviate from this curve. In practice, we assumed Y = C / T and and a and b were determined using a linear approximation Y = aX + b along the X-Y axes. The parts deviating from this fitting line were evaluated as Figure 3 was used for the evaluation. The results obtained for a and b for each sample are shown in Fig. 4. The b of GdSb is large in comparison with the b of the other two materials. Therefore, it is assumed that the electronic specific heat of GdSb is larger.
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DISCUSSION The type of crystal structure, melting point, and lattice constant obtained from the analytical results of sample powders by X-ray diffraction, density and Néel temperature were evaluated for HoSb, DySb and GdSb samples. Their values are shown in Table 1. It is confirmed that the evaluated values for density and Néel temperature correspond to the values in the literature. The relation between peak value and half-width of specific heats is shown in Fig. 5. It can be seen from Table 1 and Fig. 5 that some differences in the properties of GdSb and HoSb and DySb exist; however the differences are quite small. Thus, further investigation is required to explain why GdSb has its unique broad specific heat peak that continues up to higher temperatures. At this point, the data suggest that the difference in the specific heats between the samples is mainly due to the differences in their magnetic properties.
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CONCLUSION HoSb and DySb have a very high specific heat peak, while GdSb has a broad peak and its high value steadily increases up to high temperatures. The analyses focused on the magnetic properties of the samples in an attempt to understand the difference between the samples with sharp and with broad specific heat peaks. In this study, the temperature dependence of magnetization of the samples was measured with a VSM with an external magnetic flux density of 1.0 T supplied to the samples. The magnetic properties of HoSb and DySb disappear above 100 K. The magnetic property of GdSb continues up to near room temperature. The difference in the specific heat is considered to be due to the differences between the magnetic properties of the three materials. It is also assumed that the electronic specific heat of GdSb is large in comparison with that of HoSb and DySb.
REFERENCES 1.
H. Nakane, S. Yamazaki, H. Fujishiro, T. Yamaguchi, S. Yoshizawa, T. Numazawa and M. Okamura, “Low Temperature Properties of HoSb, DySb and GdSb,” Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2001), pp. 443-448.
2.
Richard M. Bozorth, Ferromagnetism, D. Van Nostrand (1951), pp.270-271.
3.
M.E. Mullen, B. Luthi and P.S. Wang, “Magnetic-ion-lattice Interaction: Rare-earth Antimonides,” Physical Review B, 10 (1974), pp. 186-199.
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A New Ceramic Magnetic Regenerator Material for 4 K Cryocoolers T. Numazawa1, T. Yanagitani2, H. Nozawa2, Y. Ikeya3, R. Li3 and T. Satoh3 1
Tsukuba Magnet Laboratory, National Institute for Materials Science 3-13 Sakura, Tsukuba 305-0003, Japan 2 Ceramics Division, Konoshima Chemical Co. Ltd. 80 Koda, Takuma-Cho, Mitoyo-Gun, Kagawa 769-1103, Japan 3 Cryogenics Department, Sumitomo Heavy Industries, Ltd. 2-1-1 Yato-Cho, Nishitokyo City, Tokyo 188-8585, Japan
ABSTRACT A new class of magnetic regenerator materials for 4 K cryocoolers has been developed. A hexagonal gadolinium oxisulfide has been found to provide a very high volumetric heat capacity of more than at 5.2 K. Polycrystal ceramic GOS particles have been fabricated with diameters between 0.35 mm and 0.45 mm. The achieved smooth surface on the spherical particles and a Vickers hardness of ~900 make the material very suitable for use in regenerators. The measured data on the thermal and magnetic properties of the GOS also show excellent characteristics as a regenerator material; no change in heat capacity is observed when applying a magnetic field up to 1 T, and the high thermal conductivity of ~0.05 W/cmK at 6 K is twice that of GAP. Magnetization of GOS at 1 T is about half that of between 2 K and 20 K. System-level cooling tests have been conducted with both GM-cycle and pulse tube cryocoolers where a portion of the in the 2nd regenerator was replaced with GOS. At a 50% volumetric ratio of GOS to the cooling power increased up to ~25 % at 4.2 K. The cooling power did not drop above 4.5 K, where it significantly decreases when GAP is used. The new ceramic regenerator material GOS is expected to not only contribute to increased cooling performance at 4.2 K, but also to decreased fabrication cost.
INTRODUCTION Current technology development on conventional mechanical cryocoolers continues to achieve breakthroughs in performance and reliability. However, from the view point of materials research, new regenerator materials are still a key component to improving both the performance and the cost of cryocoolers, especially in the low temperature region below 4 K. We have investigated promising candidate materials focusing on rare-earth oxides, and a ceramic magnetic regenerator material GAP has been developed for 2 K to 4 K cryocoolers.1 Cooling tests using GAP have shown excellent improvement in the cooling performance below 4 K2-4, but the cooling capacity is still somewhat depressed because of GAP's steep decline in heat capacity above 4 K. Therefore, a new magnetic regenerator material, similar to the GAP, is desired for 4 K cryocoolers to provide higher heat capacity in the region immediately above 4 K. Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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This paper describes our approach to solving the requirement for a new 4 K regenerator material. Thermal and magnetic properties of the new ceramic magnetic material will be shown including its fabrication process. Then, experimental results of cooling tests with both GM and pulse tube cryocoolers will be presented to confirm the performance of the new magnetic material.
CERAMIC MAGNETIC REGENERATOR MATERIAL Properties of Rare-Earth Oxides Used in Regenerators Using oxide magnetic materials for low temperature regenerators was first suggested in our previous papers.1,5 The developed GAP material, shown in Fig.1, displays a considerably higher heat capacity than other conventional metallic regenerator materials in the region below 4 K. Such a high heat capacity has proven to be very effective in increasing the cooling performance below 4 K; however, the heat capacity suddenly drops above 4 K because of the antiferromagnet properties of the material. Several important characteristics of rare-earth oxides for regenerator applications are summarized below: 1) 2) 3) 4) 5)
Rare-earth oxides tend to show antiferromagnetic behavior Generally, the magnetic transition temperature is not high, roughly <10 K Heat capacity shows very sharp lambda type anomaly Magnetic field dependence of heat capacity is small for < 1 T fields Magnetization is weak for < 1 T fields
It is noticeable that these properties show both positive and negative aspects as a regenerator material. For example, a sharp peak of the heat capacity is not desirable for a regenerator, but this comes from the anti-parallel alignment of magnetic moments in the oxide crystal structures. On the other hand, this same magnetic interaction provides excellent magnetic characteristics as described in 4) and 5) compared with other metallic regenerator materials.
Approach of Developing a New Ceramic Regenerator Material for 4 K Figure 2 shows the results of a cooling test with a GM cryocooler where 20 % of the volume of was replaced with GAP. It is clear that the GAP material increases the cooling power below 4 K, but there is minimal advantage to using the material above 4.2 K. Then, some questions arise as to what kind of heat capacity properties are most important for a 4 K regenerator. It can be pointed out that the peak temperature and the broadness of the heat capacity in a certain temperature
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range must be considered, but there is no clear answer except real cooling tests with cryocoolers. However, numerical studies will contribute to predict the fundamental aspects of the heat capacity. Torii et. al. showed very useful simulation results on the heat capacity of regenerator material relating to the cooling capacity for 4 K cryocoolers.6 From the view of oxide regenerator material, we summarize their calculated results as shown below: 1) Heat capacity must be larger than to provide a cooling capacity of 2) Peak temperature of heat capacity between 4.2 K and 5 K is effective in increasing the cooling capacity at 4.2 K. 3) Temperature width of heat capacity peak is optimized between 2 K and 4 K to improve the cooling capacity at 4.2 K.
Increasing the peak temperature of GAP from 3.8 K to the temperature between 4.2 K and 5 K seems to be a most simple solution to provide better cooling performance. Thus, we first tried to modify the GAP by changing its composition. Figure 3 shows the heat capacity of 0.5 mol Gdrich GAP, in comparison with that of GAP. Twin peaks appear in and this suggests that two crystals of and coexist in the material. Solubility limit in
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the perovskite structure like GAP is very small and therefore, it is difficult to control the magnetic transition temperature by substituting Gd for the other rare-earth elements. We investigated the alternate way to shift the peak temperature higher than 4.2 K based on the thermodynamic stability of the crystal growing process. Starting from a seed material we identified that a hexagonal gadolinium oxisulfide could be the best candidate for 4 K regenerator material because of the magnetic transition temperature of ~5 K and the high volumetric rare-earth density of the crystal structure. Figure 4 shows the measured volumetric heat capacity of the GOS in comparison with the GAP and The GOS has a considerably large heat capacity of more than at 5.2 K and the temperature width of the heat capacity peak is 2.5 K at while the GAP has that of 1.8 K. Thus, it is concluded that the GOS satisfies the dominant conditions of the heat capacity as a regenerator material to improve the cooling performance of 4 K cryocoolers. Thermal and Magnetic Properties of GOS Figure 5 shows the magnetic field dependence of the heat capacity of the GOS. Zero magnetic field heat capacity shows a typical lambda type anomaly due to the antiferromagnetic material. There is no change in the heat capacity by applying the magnetic field up to 1 T. This
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property is important for the 4 K cryocooler which is used to cool superconducting magnets. Figure 6 shows the thermal conductivity data of the GOS at zero magnetic field in comparison with those of the GAP and a SUS. Note that the measured GAP and GOS samples are the polycrystal ceramics. The thermal conductivity of the GOS is higher than those of the GAP and the SUS, and it will contribute to shorten the thermal penetration time. Figure 7 shows the magnetization data of the GAP as a function of temperature for the magnetic field of 1 T in comparison with that of Magnetization of the GOS at 1 T is about twice smaller than that of between 2 K and 20 K, Some 4 K cryocoolers are also used to cool the sensors which are sensitive to the magnetic noise like a SQUID, so the smaller magnetization of the regenerator material can reduce the magnetic noise coming from the cryocooler.
Fabrication of Ceramic Form GOS In order to use GOS as the regenerator material, sphere form particles between 0.2 mm and 0.5 mm in diameter must be fabricated. The hardness of the particles is another important factor in avoiding the broken particles occasionally caused by the pressurized helium gas flow in the regenerator. Basic scheme of the fabrication method of the ceramic GOS is very similar to that of GAP
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previously reported.1 Starting raw material is very important especially in the GOS because the composition of sulfur is easily changeable. First we prepare the ultra-fine powders with diameter of µm which is produced by a wet chemical processing. Then, the powders are reacted chemically with hydrogen sulfide gas to produce powders directly. This simple method is useful to obtain the uniform fine powders of GOS. The granulator is used to make the spherical particles between 0.2 mm and 0.5 mm in diameter. Then, the particles are calcined at a temperature of 1500 °C in Ar atmosphere. After choosing the proper size particles with screen meshes, the chosen particles are polished to remove the thin and weak surface layer. Precise control technique has been improved to produce more spherical particles with the smooth surface. Currently the distribution range of the particle size can be reduced within 0.1 mm and the larger particle size is obtainable; typical average particle size is 0.4 mm ± 0.05 mm. Figure 8 shows the photographs of the fabricated GOS particles with the smooth surface. The relative density to that of the single crystal is 99 % and Vickers hardness is ~900 which is a high enough value used as the regenerator material.
EXPERIMENTAL RESULTS ON COOLING TEST GM cycle and pulse tube cryocoolers were used to test the cooling performance of the GOS regenerator material. For the GM cryocooler, two kinds of commercial models, Sumitomo SRDK-408D (
[email protected]) and Sumitomo SRDK-101D (
[email protected]) were chosen. The pulse tube cryocooler was a prototype model with a cooling power of 0.5 W at 4.2 K. The cooling tests were done by replacing the with the GOS at the 2nd regenerator. Note that the initial condition of the 2nd regenerator consisted of and Pb and we did not change the volume of
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the Pb. In order to make clear the replacing effect of the GOS, we introduce the volumetric ratio R, which is defined by the ratio of the volume of GOS to the volume of at the lower temperature portion of the 2nd regenerator. Details on the experimental procedure and the measured results will be shown in our other papers.7,8 Figure 9 shows the experimental result on the temperature dependence of the cooling capacity at the 2nd regenerator for the 1 W GM cryocooler. At R = 50 %, the cooling capacity increases ~10 % at 4.2 K. It is noticeable that the cooling capacity keeps increasing even above 4.5 K, where it decreases largely in the case of the GAP. Figure 10 shows the R dependence of the cooling capacity of the 2nd regenerator at 4.2 K for the 1 W GM cryocooler. It is clear that an optimum value of R can be found around 50 % to provide the maximum cooling capacity at 4.2 K. This behavior is different from the case of the GAP in which the optimum R is ~25 %. For R > 60 %, the cooling capacity decreases largely. This is because the heat capacity the GOS steeply decreases above 5.5 K and the compensation for lack of the heat capacity is required by using some amounts of the Figure 11 shows the R dependence of the cooling capacity at 4.2 K for the 0.5 W pulse tube
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cryocooler. The cooling capacity curve shows a local minimum between R = 20 and R = 40, but we guess that it may be caused by the setting error of the experiment. For R > 50 %, the cooling capacity decreases largely less than that of R = 0 and it is clear that an optimum R will be found for R < 50 %. This result is different from the case of the 1 W GM cryocooler; fewer amounts of the GOS will be enough to improve the cooling capacity of the pulse tube cryocooler. For the 0.1 W GM cryocooler, the cooling capacity as a function of R has a broad peak as shown in Fig. 12. The peak value can be found around R = 65 % where the increasing rate on the cooling capacity of 25 % is obtainable. It is somewhat surprising that the cooling capacity does not decrease much at R = 100 % without any Therefore, the GM cryocoolers will require more amounts of the GOS than the pulse tube cryocooler to increase the cooling performance.
CONCLUSIONS A new generation magnetic regenerator material GOS has been developed for 4 K cryocoolers. The fabrication techniques are progressing to produce more uniform sphere particles with the sufficient hardness as the regenerator material, and additionally, to realize less cost compared with the conventional metallic magnetic regenerator materials. Long term tests with the commercial cryocoolers are proceeding. No broken particles have been found and the temperature stability of ± 45 mK at the 2nd stage regenerator with the heat load of 1 W has been measured for the continuous operation at the elapsed time of one month. Variety of other ceramic magnetic regenerator materials are under development to cover the wider temperature ranges. Also new innovative approaches are being studied to make the regenerator configuration much simpler by using the flexible fabrication of ceramics.
ACKNOWLEDGMENT The authors would like to thank Dr. M. Uehara and Dr. K. Kamiya of Tsukuba Magnet Laboratory for the measurements of magnetization of GOS and
REFERENCES 1.
Numazawa, T., Arai, O., Sato, A., Fujimoto, S., Oodo, T., Kang, Y.M. and Yanagitani, T., “New Regenerator Material for Sub-4K cryocoolers in the 11th International Cryocooler Conference,” Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2001), pp. 465-473.
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2.
Fujimoto, S., Kurihara, T., Oodo, T., Kang, Y.M., Numazawa, T. and Matsubara, Y., “Experimental Study of a 4K Pulse Tube Cryocooler in the 11th International Cryocooler Conference,” Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2001), pp. 213-219.
3.
Numazawa, T., Satoh, T., Yanagitani, T. and Sato, A.,”Cooling Performance of Ceramic Magnetic Regenerator Material Used in a GM Cryocooler”, submitted to Adv. in Cryog. Eng., vol.46 (2002).
4.
Qiu, L.M., Numazawa, T. and Thummes, G., “Performance improvement of a pulse tube cooler below 4 K by use of regenerator material”, Cryogenics, vol.41 (2001), pp. 693-696.
5.
Numazawa, T., Okamura, M. and Sato, A., “Magnetic Regenerator Materials for Sub-2K Refrigeration”, Adv. in Cryog. Eng., vol.46 (2000), pp. 1210.
6.
Torii, H., Kurihara, T., Morishita, H. and Miura, K., “Development of two-stage 4K cryocooler with modified Solvay cycle”, Proceedings of ICEC16/ICMC (1996), pp. 367-370.
7.
Ikeya, Y., Numazawa, T. and Li, R., “Improvement of 4 K GM Cooling Performance with a New Regenerator Material in the 12th International Cryocooler Conference,” submitted to Cryocoolers 12, Kluwer Academic/Plenum Publishers, New York (2003).
8.
Satoh, T. and Numazawa, T., “Cooling Performance of a 4 K GM Cryocooler with New Ceramic Magnetic Regenerator Materials in the 12th International Cryocooler Conference,” submitted to Cryocoolers 12, Kluwer Academic/Plenum Publishers, New York (2003).
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Predicted Performance of a Low-Temperature Perforated Plate Regenerator J. B. Hendricks Alabama Cryogenic Engineering, Inc. Huntsville, Alabama USA
ABSTRACT The matrix material and geometry are both critical in regenerator design at low temperatures. Since the release of REGEN3.2 1, it is relatively easy to calculate the performance of regenerators at low temperatures. In this effort, the impact of changing the low temperature regenerator matrix geometry to perforated plates will be studied. The starting point was a cryocooler, using a neodymium packed sphere regenerator, whose performance has been reported earlier 2. The length and diameter of the cold stage regenerator was kept the same, and the geometry was changed. The refrigeration power of the cold stage was then calculated as a function of the perforated plate porosity. For a porosity value of 5% the predicted refrigeration power was more than doubled when compared to the packed sphere baseline. Some parameters, such as the NTU limit, plate spacing, and thermal penetration effect, are not explicitly calculated in REGEN3.2. The effect of these parameters on overall performance will be discussed. In conclusion, the perforated plate regenerator geometry holds promise for an improvement in refrigeration power of at least a factor of two when compared to the packed sphere case.
INTRODUCTION The selection of a matrix material for a cryocooler regenerator is usually driven by material availability. Therefore, matrix materials such as woven wire screens and metal spheres are the usual selections. However, there is no reason to suspect that these are the optimum matrix geometries. For example, the perforated plate geometry, whose heat transfer and pressure drop characteristics can be approximated by internal tube flow, has a higher value of the compactness factor than either screens or packed spheres. This should lead to a higher heat transfer coefficient per unit of pressure drop, and thus result in improved regenerator performance. Earlier “cut-and-try” efforts using perforated plate regenerators have been relatively unsuccessful. It is quite difficult to fabricate perforated plate material, so experimental design data are limited. However, regenerator simulation programs have recently become available, and these can be used to study this regenerator geometry. If the alternate regenerator matrix material shows sufficient promise in these simulations, then a fabrication effort can be initiated which could produce samples that can be fabricated and tested.
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This report covers a study of perforated plate regenerators at low temperatures. In order to function in this temperature range, the plates must be fabricated of a material having a substantial value of volumetric specific heat. Examples include lead, neodymium and several rare earth intermetallics. Neodymium has been chosen as a baseline, since it has the best combination of properties from a fabrication standpoint. The malleability of lead is very good, while that of neodymium is somewhat limited. However, the malleability of neodymium is much better than the typical rare earth intermetallics, such as The specific heat of neodymium is less than that of a number of intermetallics, but it is substantially greater than the value for lead at the lowest temperatures.
BASELINE LOW-TEMPERATURE REGENERATOR In an earlier R&D effort some neodymium perforated plates were tested in a prototype GiffordMcMahon refrigerator2. This machine had also been tested with various packed sphere designs that included combinations of neodymium and lead spheres. As reported earlier, the perforated plates did not improve the performance; in fact, the performance was degraded. These plates did not have uniform pore sizes, and the passage shape was irregular. The plate porosity was approximately 15%, but this also varied over a fair range. For this effort, the same regenerator geometry will be selected, so the “ideal” performance of perforated plates can be simulated. The selected parameters of the low temperature regenerator are given in Table 1. In the simulations the cold end mass flow was held constant, corresponding to a fixed displacer volume. The hot end mass flow was adjusted to give the proper solution. The effect on the overall performance of the refrigerator of this changing hot end mass flow was not calculated. A reduction in the hot end mass flow should improve the performance of the upper stage, which should improve the overall cooling power.
DESCRIPTION OF REGENERATOR SIMULATION PROGRAM The regenerator simulation program was developed at the Boulder, CO laboratories of the National Institute of Standards and Technology (NIST). The latest version of the program is one of an ongoing series that has provided increasing utility for the user. The version used for the simulations presented here was a “beta” version of an improvement over REGEN3.2. The beta version is named RG4. In earlier attempts, REGEN3.2 could not give results for perforated plate regenerators operating between 4K and 40K. However RG4 solved these problems, and the desired results were easily obtained. REGEN3.2 and RG4 is entirely the product of the NIST group. Any questions about the details and availability of the programs should be directed to that group3. The simulations were run on a machine with an 850MHz, AMD K7 processor. The computer operating system was Microsoft WINDOWS 2000 Pro. Older operating systems such as WINDOWS 98 could not run all versions of the code. Run times of several hours were necessary to get “good” solutions.
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RESULTS OF COMPUTER SIMULATION The results for a series of regenerator simulations are given in Table 2. To the extent that was possible, only the porosity and matrix geometry were changed between steps. The hot end mass flow increased as the porosity was decreased. This will have an effect on the upper cryocooler stage, but no attempt was made to calculate this impact. The pressure drop also increased as the porosity decreased, but even for the smallest value (5%) the pressure drop was calculated to be 12.3 kPa. This pressure drop is much less than the average pressure (1.5 MPa) so it should be negligible.
DEAD VOLUME EFFECTS IN PERFORATED PLATE REGENERATORS The amount of dead volume in the regenerator is most critical at the cold end of the regenerator. Above 10 Kelvin the pressurization losses in the helium working fluid are relatively small and can be neglected. There are two variables to be considered. These include the spacer material and thickness and the perforated plate thickness. From experience, the minimum thickness is roughly for stainless steel spacers. Anything thinner is almost impossible to laser cut and then handle during the assembly of the regenerator. If a perforated plate thickness of is selected, then the inter plate volume represents 2.4% of the total volume. If a passage volume of 5% is used, then the total dead volume will be 7.4%. According to the results in the previous section, this will still result in a substantial improvement in the cooling power, when compared to the packed sphere baseline.
LONGITUDINAL THERMAL CONDUCTIVITY EFFECTS IN PERFORATED PLATE REGENERATORS The longitudinal thermal conduction of the matrix material can have a large effect in regenerator performance. REGEN3.2 handles this with a factor called FUDGE. If FUDGE=0.1 then the actual thermal conductivity is 1/10 of the conductivity of the matrix material assuming the regenerator is a continuous bar with no longitudinal interruptions. For the perforated plate system, the total temperature drop can be assumed to occur across the spacers, with no temperature drop across the plates. Using the thermal conductivity values that are included in REGEN, the ratio of the thermal conductivities of neodymium and the helium working fluid (He at 1 MPa) is ~300 at 4K, ~100 at 10K and ~250 at 40K. Note that the stainless steel spacers have a very small width, and their effect can be ignored. Using these ratios, the value of FUDGE should be 0.13 at 4K, 0.4 at 10K and 0.16 at 40K. A value of FUDGE=0.3 was used in the calculations
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reported above. This seemed to be a reasonable compromise number. As a test, for the 5% porosity case, the value of FUDGE was doubled (FUDGE=0.6). The net adjusted cooling power was only reduced by 2%. Even though the temperature drop was assumed to occur only across the spacers, it is clear that the thermal resistance of the plates and spacers is not drastically different. Therefore, there will be some temperature drop across each of the perforated plates. In a perforated plated system, it usually assumed that each plate is isothermal, so the system under consideration is a hybrid between the true perforated plate geometry and longitudinal tube flow. NTU EFFECTS IN PERFORATED PLATE REGENERATORS In a true perforated plate regenerator, each of the plates is an isothermal surface. Under these conditions, by definition, the value of NTU for each isothermal plate is less than or equal to unity Therefore, the maximum value of NTU for the regenerator under consideration is:
This is the maximum value of NTU, and since the system is less than ideal, the actual value will be less. In order to check the result of the simulation for NTU effects, the temperature gradient was taken from the simulation to get the temperature of each plate. The NTU value per plate could then be calculated from: Where A=heat transfer area, U=overall heat transfer coefficient and The flow in the passages through the plate is laminar, so the overall heat transfer coefficient can be easily calculated. The results are somewhat surprising. The NTU values per plate are less than unity only for plate porosities less than 6%. For all higher porosities, the calculated NTU values are greater than unity for at least some of the plates. For a porosity of 38% the calculated NTU values range from a low of 3 to a high of 7. As in the previous paragraph, the system under consideration is not a pure perforated plate regenerator, but is closer to a simple internal tube flow system. Even though the regenerator is not a simple perforated plate system, the gaps between plates still serve a purpose. A simple internal tube flow system will probably not yield good performance because small variations in the pressure drop along the various channels will lead to flow maldistribution. Flow maldistribution is known to severely impact heat exchanger performance. The gaps between the plates will allow the flow to re-distribute in each gap and maintain uniform flow across the plate areas.
THERMAL PENETRATION EFFECTS IN PERFORATED PLATE REGENERATORS If all the regenerator matrix material is to be involved in the regeneration process, then the thermal wave must reach all points on each perforated plate. The center-to-center spacing between the flow passages in a 5% open perforated plate with diameter holes is ~0.2 mm. Therefore, the thermal penetration length must be greater than The thermal penetration length for GdRh is roughly 0.3 mm at 10 Hz for low temperatures. Thus this should not be a problem for pure Nd at 1.42 Hz.
CONCLUSIONS The reduction of the porosity in a regenerator system will reduce the volume of helium working fluid that undergoes periodic pressurization at low temperatures. This should reduce the amount of heating in the working fluid due to non-ideal behavior. However, the reduction of
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porosity also reduces the wetted heat exchange surface that could impact performance. The computer simulation indicates that the reduction in porosity yields a net improved performance. The particular system under consideration compared a packed neodymium sphere regenerator to a neodymium perforated plate regenerator. The external dimensions, cold end mass flows, hydraulic diameters, frequencies, etc. were the same for the two systems. Only the porosity was changed. Therefore, we conclude that the use of perforated plate regenerator systems could have a substantial impact on cryocooler performance.
ACKNOWLEDGEMENTS The computer programs used in the effort were developed at the Boulder Laboratories of the National Institute of Standards and Technology (NIST). I want to thank Ray Radebaugh, John Gary and Abby O’Gallagher for their hospitality during a visit, and for their help and patience during my trip up the learning curve.
REFERENCES 1.
Radebaugh, R., E.D. Marquardt, J. Gary and A. O’Gallagher, “Regenerator Behavior with Heat Input or Removal at Intermediate Temperatures,” Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2001), pp. 409-418.
2.
Chafe, J.N., G.F. Green and J.B. Hendricks, “A Neodymium Plate Regenerator for Low-Temperature Gifford-McMahon Refrigerators,” Cryocoolers 9, Plenum Press, New York (1997), pp.653-662.
3.
Dr. R. Radebaugh, NIST-Boulder, 325 Broadway, Mail Stop 838.09, Boulder, CO 80305 USA.
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LIGA-Fabricated High-Performance Micro-Channel Regenerators for Cryocoolers K. Kelly, A. McCandless, S. Motakef Mezzo Systems Baton Rouge, LA
ABSTRACT The LIGA micromachining process can be used to fabricate sheets perforated with micro channels having length/diameter (L/D) ratios between 10 and 20. The potential benefit of a regenerator composed of an aligned stack of these sheets and fabricated from a material with good cryogenic thermal properties is quantified via analytical as well as numerical models. A cross flow micro heat exchanger is described using a derivative of the LIGA process that has excellent heat transfer/ volume performance. Modifying the process sequence to make these heat exchangers could potentially produce cryocooler regenerators with excellent performance.
INTRODUCTION Cryocoolers are used to maintain the temperature of a device at cryogenic temperatures between 4 and 70 K. The performance of Cryocoolers, in terms of power efficiency and the attainable cold end temperature is strongly influenced by the geometry and cryogenic thermal properties of the regenerator matrix. The ideal regenerator would have the following characteristics: High heat transfer area High wall-to-gas heat transfer Low axial conduction Low pressure drop loss High heat capacity Low dead (void) volume Currently, wire meshes and packed beds of spheres are used as regenerators in cryocoolers [1]. For these systems the pressure drop across the regenerators for a given unit of heat transfer is high, resulting in low overall efficiency of the cryocooler. This problem is particularly acute in low temperature (below 30K) cryocoolers where only packed beds of lead (10-30K) and magnetic intermetallic compound pellets (below 10K) can be used as regenerators. The high pressure drop across these packed beds has a significant effect on the overall efficiency of the cryocooler. This paper describes methods to economically fabricate micro-channel regenerators with length/diameter (L/D) ratios of 10-20 and relatively low porosity that provide significantly superior performance compared to sphere beds and screens in all types of cryocoolers.
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LIGA BMDO SBIR funding is currently supporting an effort to use the LIGA micromachining process to fabricate regenerators such as shown in Figure 1. Each regenerator will be composed of a stack of sheets of lead or lead mixed with a high heat capacity material such as Ni-Er. Each sheet may be coated with a thin electroplated layer of nickel. Each sheet will have a thickness of 1-2 millimeters and will be perforated with a sparse array of holes 25-100 microns in diameter. A stack of these sheets will result in a regenerator with the desired properties listed in the previous section. The sheets will be fabricated using the LIGA micro machining process. The LIGA process, developed at the Institute for Microstructure Technology (IMT-FzK) in Karlsruhe, Germany, is a three-step process comprised of X-ray lithography, electroforming, and molding. Important features of the LIGA process are: microstructures hundreds to thousands of micrometers in height; smooth, flat sidewalls; aspect ratios on the order of 100:1; and economical manufacture of microstructures by molding. The LIGA process (Figure 2) begins by lithographically patterning an x-ray sensitive resist (poly-methyl-methacrylate (PMMA) or SU-8) with a collimated source of x-ray radiation (such as emitted from a synchrotron). After development, polymeric structures are left standing on a supporting conductive substrate. An electroforming procedure is used to fill in the voids between the polymeric features. After the electroforming process is completed, the polymeric resist is stripped, leaving a metal mold tool that can be used repeatedly to reproduce secondary polymer, ceramic or metal micro features nearly identical to those produced by the original x-ray lithography process.
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In the case of regenerators, a mold tool consisting of a field of posts 1-2 mm long and 25-100 microns in diameter is used to mold lead sheets. It may be preferable in some cases to coat the lead sheets with a thin layer of electroplated nickel (to reduce the diameter of the microchannels and to provide strength). An alternative approach would be to mold a polymer piece, electroplate it with nickel, dissolve the polymer to create a nickel shell, and then pour molten lead into the shell, and thereby create a nickel coated lead heat exchanger. This approach would be similar to the fabrication of cross flow micro heat exchangers described below.
MICRO CROSS FLOW HEAT EXCHANGER Phase I of the current BMDO funding focused on fabricating and testing the nickel micro cross flow heat exchanger shown in Figure 3 [2,3,]. To fabricate this heat exchanger, the previously described LIGA micro fabrication process was utilized. A mold insert consisting of a field of posts with hexagonal cross section (Figure 4a) was used to mold polymer parts perforated with holes (Figure 4b). An electroplating process was used to coat the entire surface of the molded polymer part (except two opposing edges) with nickel. After electroplating, the part was immersed in a solvent (acetone) to dissolve the polymer from within the metal shell. Magnified views of the heat exchanger are shown in Figures 4c- 4f. One fluid (Fluid 1) flows through the heat exchanger in a direction perpendicular to the plane of the heat exchanger via thousands of holes with a diameter of 400 µm and length of approximately 1 mm (L/D ratio of 2.5). The second fluid (Fluid 2) flows in the plane of the heat exchanger through a maze of pins with hexagonal cross section. Because of reduced length scale, extremely high values of heat transfer/volume and heat transfer/unit mass can be achieved. A cryocooler regenerator is fundamentally different from the cross flow heat exchanger shown in Figures 3 and 4c-4f. Nevertheless, by making the three modifications listed below, the heat exchanger shown in Figure 3 can be converted into a sheet with attractive properties for use in a cryocooler regenerator: 1. The internal cavity within the plane of the cross flow heat exchanger through which Fluid 2 passes can be filled with a material with appropriate thermal properties at cryogenic temperatures (i.e. Pb, Er-Ni, Er-Co, He). 2. The L/D ratio of the Fluid 1 channels shown in Figure 1 is only 2.5. To achieve an optimal ratio of heat transfer/pressure drop, L/D ratios of 10-20 are needed. In addition, the desired diameter of these channels might range from twenty to few hundred micrometers. The micro fabrication process that was used to produce the cross flow heat exchanger is capable of producing channels with both the desired L/D ratio and the desired absolute channel diameter. 3. In a regenerator, the heat capacity of the solid matrix should greatly exceed the heat capacity of the volume of fluid that passes back and forth through the regenerator. For this reason, the porosity of the sheet is an important variable, and often the desired porosity (5-15%) is much less
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than the actual porosity that can be obtained with wire screens or packed beds of spheres. The density of channels in the heat exchanger shown in Figure 1 can be dramatically decreased, thereby increasing the ratio of solid heat capacitance/fluid heat capacitance to a desired value. By making these modifications, it is possible to produce 1-2 mm thick regenerator sheets that, when stacked together, produce regenerators with very attractive properties at cryogenic temperatures.
EXPERIMENTAL RESULTS FROM MICRO CROSS FLOW HEAT EXCHANGER Heat transfer and pressure drop tests were conducted to quantify the performance of the cross flow heat exchanger shown in Figure 3. Hot water (Fluid 2 of Figure 4) was pumped through the void within the plane of the heat exchanger and room temperature air was pumped through the heat exchanger in a direction normal to the to plane of the heat exchanger. The change in the temperature of water and air, as well as the pressure drop across the heat exchanger, were measured and are plotted as a function of air velocity in Figures 5(a) and (b).
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The experimental results compare very well with the theoretical prediction of heat transfer and pressure drop in these systems. Slightly simplified forms of the equations used to predict the performance of the nickel cross flow heat exchanger are given by Equations (1)- (3) below. Equation (1) predicts the heat transfer to the fluid (air) passing in cross flow through heat exchanger. Here, an assumption was made that the wall of the channel is effectively the temperature of the water, flowing in the plane of the heat exchanger, and the convection coefficient for the gas side is constant along the channel length (strictly true for fully developed flow, which is not the case here).
where: = = L = V = = = = = =
A constant depending upon the Reynolds number, the thermal conductivity of the fluid, the length of the air channel (the thickness of the heat exchanger), the velocity of the air in the heat exchanger, the hydraulic diameter of the air channel, the constant pressure specific heat of the air, the density of the air, the exit temperature of the air from the heat exchanger, consistency of symbols the temperature of the air entering the heat exchanger.
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The rate of heat transfer to the gasper channel is given by Equation (2):
where: Q = =
the rate of heat transfer per area of channel to the gas the mass flow rate of gas through the channel
Equation (3) predicts the pressure drop of the air across the heat exchanger for the case where flow in the channel is fully developed and laminar. The first term on the right hand side accounts for viscous drag within the channel and often represents the dominant term. The second term accounts for inlet and exit losses associated with sudden expansion/contraction.
where: V = µ =
the velocity of the fluid through the channel the viscosity of the fluid
Together, Equations (1)-(3) indicate that for given fluid material properties and velocity, the change in temperature of the gas across the heat exchanger is constant for a given value of Therefore, for a given specific heat transfer/unit volume is proportional to 1/L. The total heat transfer/volume of the heat exchanger will also be proportional to 1/L if the porosity of the heat exchanger (volume of air channels/heat exchanger volume) does not vary with scale. Practical geometric considerations at the micro scale make this difficult to achieve, so actual heat transfer/ volume improvements will not strictly follow the 1/L relationship. Surprisingly, the penalty normally associated with pumping fluid through micro channels is minimal in a cross flow micro heat exchanger since, for a given velocity (assuming the inlet and exit losses are small relative to viscous drag losses), the pressure drop is proportional to the ratio of Therefore, the pressure drop of the crossstream gas across a conventional scale cross flow heat exchanger is the same as the value across a micro cross flow heat exchanger, assuming both heat exchangers have the same value of The experimental measurements of heat transfer and pressure drop (Figures 3a and 3b) compare very favorably with analytical models (solid lines). Test data also shows that the heat transfer/ volume and heat transfer/weight of the heat exchanger shown in Figure 3 is, respectively, 6 and 2-3 times greater than the best conventional scale water-air cross flow heat exchangers for similar design constraints (i.e. the same pressure drop of the air across the heat exchanger and the same difference in the temperatures of the fluids entering the heat exchangers).
THEORETICAL JUSTIFICATION FOR LIGA-FABRICATED CRYOCOOLERS In this section, we draw upon experimental and modeling results to highlight the significant impact that micro-channel regenerators can have on performance of cryocoolers. Specifically: 1. the ratio of pressure drop to heat transfer in micro-channel regenerators is significantly lower than values attainable in currently-used screens and packed beds, implying significant potential improvement in the efficiency of cryocoolers; 2. micro-channel regenerators can provide heat capacity ratios significantly larger than the values possible in current packed bed regenerators, a significant issue for performance of cryocoolers, especially the sub-30K cryocoolers; 3. simulation results using NIST’s cryocooler analysis model predicts significant increases in the efficiency of cryocoolers using micro-channel regenerators.
Results of this section will demonstrate the potential of LIGA-fabricated micro-channels to provide a combination of channel geometry, porosity, and matrix material that will offer substantial advantages
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over existing systems. In addition to performance factors, these new regenerators should offer advantages regarding ease of assembly, and environmental safety (lead micro spheres are considered dangerous to work with).
A. Pressure Drop/Heat Transfer Ratios in Regenerators Improving performance of regenerators requires maximizing heat transfer from the gas to the regenerator, while at the same time minimizing the pressure drop across the regenerator. The heat transfer versus pressure drop feature of a regenerator is primarily a function of the regenerator type and the flow rate of the gas. Ruhlich and Quack have proposed the ratio of “Number of Pressure Heads pressure drop (NPH) to Number of Heat Transfer Units (NTU)” as a metric for regenerator performance, ref. [4]. Their study observed that sphere beds have the highest ratio of NPH/NTU, followed by screens. More importantly, NPH/NTU for sphere beds and screens is much higher than that for square channels and parallel plates. That is, for the same heat transfer characteristics the pressure drop penalty for sphere beds is about 6-7 times, and that for screens is 3-6 times higher than the penalty for square channels and parallel plates. According to Radebaugh [4] the performance loss attributable to pressure drop in regenerators is comparable to the net performance of the cooler, especially for pulse tube coolers with high mass flow rates through the regenerator. Therefore, replacing screen and sphere bed regenerators with one consisting of parallel square-channels, such as the one being proposed, can be expected to lead to a substantial increase in the efficiency of cryocoolers.
B. Matrix Specific Heat/Fluid Specific Heat Ratios in Regenerators In addition to the pressure drop/heat transfer ratio (NPH/NTU), a second crucial performance ratio is the matrix capacity ratio, the ratio of heat capacity of the regenerator to that of the working gas occupying its open space,
where: p = = = = =
the porosity of the regenerator the density of the matrix the specific heat of the matrix the density of the gas the specific heat of the gas
A large value of implies that heat is transferred into and out of the regenerator with very small changes in the temperature of the regenerator, resulting in higher performance of the cryocooler. In the above equation, is a strong function of the regenerator porosity, increasing rapidly with decreasing porosity. The porosity of currently available regenerators cannot be lowered significantly: the minimum porosity of packed sphere beds is 38% and that of screens is generally in the 40-50% range. The inability to reduce the porosity of these regenerators significantly impacts the performance of cryocoolers, especially as the cold-end temperature is reduced to below 20K. This is in contrast to micro-channel regenerators whose porosity can be reduced to fairly low values. For example, compare the cases described below. In all cases, the pressure of the working fluid (He) is 10 bars. In Case 1, the regenerator consists of a packed bed of lead spheres (porosity of 38%). Case 2 is a Pb micro-channel regenerator with a porosity of 10%. Case 3 is an Er-Ni compound micro channel regenerator with porosity of 10%. In all three cases, the value of has been calculated at both 20 K and 10 K. Table 1 shows the compares the values for each Casetemperature combination, as well as the volumetric specific heat values assumed for both matrix and helium.
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From Table 1, the value of can be dramatically increased using a combination of decreased porosity and matrix materials with higher specific heat. The relative importance of porosity versus matrix material varies with temperature. The LIGA technology and its derivatives described in this proposal can easily produce geometries with low porosity structures of the order of 5-10%. Thus, the work proposed here has the potential to produce regenerators with substantially higher than those currently available
C. Specific Example Demonstrating Potential of Micro-channel Regenerator In this section we investigate the potential impact of replacing the regenerator in a cryocooler of Ball Aerospace and Technology with a LIGA-manufactured micro-channel regenerator. The regenerator in Ball’s Stirling cycle cryocooler consists of Pb spheres. It occupies an annular region around a cylindrical piston with an outer diameter of 0.574 inches, an inner diameter of 0.362 inches, and a length is 0.89 inches. The volume is filled with lead spheres of .005-in diameter and packing factor 0.62 (porosity 38%). Helium flows back and forth through the regenerator with a frequency of 30 Hz at a mean pressure of 10 bars and at a flow rate of 0.3 grams/second. A preliminary Mezzo Systems’ design to replace the packed bed of lead spheres involves stacking approximately 20 washer-shaped regenerator sheets, each with a thickness of 1.0 millimeter and perforated with 20-40 µm diameter holes. The stack of washers occupy the annular volume presently filled with the lead spheres. The porosity of each washer can vary from 10% to 20%. The holes of all the washers are assumed to be aligned well enough so that, to first order, continuous channels run through the length of the regenerator. An analysis focusing on the impact of micro-channel regenerators on the overall performance of the cryocooler has been performed. These calculations were performed using the simulation software REGEN3.2 developed at NIST. The simulations software models the total performance of a cryocooler by capturing the complex coupling between heat transfer and fluid dynamics within the cryocooler. It calculates the Coefficient of Performance (COP) of the regenerator, defined as the ratio of net refrigeration power to the work input into the cryocooler. This figure of merit essentially captures the efficiency of the cryocooler. Results of this analysis are shown in Table 2. The analysis is made for Ball’s cryocooler operating between 10 and 30 K. The matrix material is assumed to be lead. The table clearly shows that replacing the existing regenerator with a microchannel one can increase the performance of the cryocooler significantly, by a factor of 3.5-4. This
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scale of improvement is not related to some particular attribute of Ball’s cryocooler. Rather it is innately related to the low-pressure drop, high heat transfer, and high heat capacity attributes of micro-channels with low porosity and small channel diameters.
CONCLUSIONS Sheets having excellent cryogenic thermal properties that are perforated with a sparse array of micro channels having small diameters (20-200 micrometers) and large L/D ratios have the potential to improve greatly the performance of cryocoolers. The LIGA process is capable of molding parts that have the desired range of properties. Mezzo Systems will use funds recently awarded by the Ballistic Missile Defense Organization (BMDO) to explore a number of methods to fabricate highly efficient cryocoolers based on derivatives of the LIGA process.
REFERENCES 1. Ackerman, R., “Cryogenic Regenerative Heat Exchangers,” Plenum Press, New York, 1997. 2. C. Harris, K. Kelly and M. Despa, “Design and Fabrication of a Cross Flow Micro Heat Exchanger,” Journal of MEMS, Vol. 9, no. 4, December 2000.
3. C. Harris, Kevin Kelly, Shariar Motakef, Andrew McCandless, Tao Wang, “Fabrication Modeling and Tesing of Micro Cross Flow Heat Exchangers,” submitted to Journal of MEMS, November 2001. 4. I. Ruhlich, H. Quack, “Wound profile-wire regenerators – fabrication and test,” in Advances in Cryogenic Engineering (45), 2000.
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Improved Flow Patterns in Etched Foil Regenerator M. P. Mitchell1 and D. Fabris2 1
Mitchell/Stirling Berkeley, California, USA 2
Santa Clara University Santa Clara, California, USA
ABSTRACT Cryocooler regenerators can be fabricated by rolling strips of etched metal foil on a mandrel, or on a pulse tube, as described in earlier papers and in a U.S. patent.1,2,3 Early versions of regenerators of that type exhibited far less pressure drop than comparable stacked-screen regenerators.2 However, performance of cryocoolers equipped with regenerators bearing those early etch patterns proved to be inferior to performance of the same cryocoolers equipped with stacked screens.3 Possible explanations included irregularities in the etching process that created the flow passages; defects in the assembly process resulting in uneven distribution of flow over the cross section of the foil regenerators; porosity of the foil so great that the thermal mass of the regenerator was inadequate; and short, straight flow passages through the regenerator that greatly reduced heat transfer as well as pressure drop. Those hypotheses led to the design of patterns that would lengthen flow passages and induce secondary flows in etched passages that were oriented normal to the overall direction of flow. Porosity of the regenerators was also reduced. Tests with a low-frequency G-M type coaxial pulse tube cooler demonstrated major differences in performance with three different patterns of foil. Best performance was obtained using the regenerator with lowest porosity, highest pressure drop and, presumably, the largest induced secondary flows. A low temperature of 54.2 K and a maximum spread of 250.1 K between warm and cold ends of the pulse tube were achieved.
INTRODUCTION The research underlying this paper was conducted under a Phase II SBIR contract to build and test a coaxial pulse tube cooler incorporating a vortex heat exchanger at the warm end of the pulse tube.4 In the course of the project, three annular etched foil regenerators were fabricated and tested. The three regenerators tested were interchangeable; all were about 100 m long, with ID of about 22 mm and OD of about 29 mm. All were etched from 300 series stainless steel nominally .0508 mm thick. All were etched from both sides of the foil. All had unetched strips normal to the overall direction flow on the back side of the foil. In each, the unetched strips were held together
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with unetched spacers on the front side of the foil. Where the foil was etched from both sides, holes appeared. The depth of etch on each side was about .035 mm at the points of deepest penetration. Regenerator #1 weighed 76.1 g, #2 weighed 79.6 g and #3 weighed 109.0 g. Since all were essentially the same size; the variation in weight reflected variation in fill factor; the third regenerator was about 47% metal and 53% void. Regenerators #1 and #2 were rolled from strips that had been welded together end-to-end. Regenerator #3 was assembled by a new method using tab connections that eliminated the expense (and potential scrap loss) of the welding process. The regenerators were held in place between a serrated ring or “spider” at the warm end of the pulse tube and the cold heat exchanger at the other end. Springs forced the spider against the warm end of the regenerator, thus pressing the other end of the regenerator firmly against the cold heat exchanger. The areas of the regenerator that contacted the surfaces of the spider and cold heat exchanger were blocked off from flow. The spider, shown in Figure 1, blocked about 40% of the cross section of the warm end of the regenerator, and the final version of the cold heat exchanger blocked about 85% of the cold end.
THE ETCH PATTERNS Etched foil regenerators were first tested in cryocoolers in the mid-1990s. At first, results were positive2. However, additional tests at Lockheed were disappointing.2, 3 Foil of the type tested by Lockheed is shown in Figure 2. It had wide flow channels oriented in the overall direction of flow,
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(i.e. top to bottom on the page), thus passing straight through the regenerator. The light-colored horizontal strips on the “front side” view are stainless steel, half-etched from the front side. The dark-colored horizontal strips are slits completely through the foil. The vertically-oriented ovals are stainless spacers, unetched on the front side, but etched part way through on the back side. The whole front area between the spacers is flow channel. (In the orientation shown, flow is vertical). The “back side” view shows the wide spacing between horizontal strips and the relatively equal size of strips and slits. The fill factor of the regenerators made with that foil was around 20% (i.e. 80% porosity). The pattern shown in Figure 3 was used in regenerator #1. The etch pattern was generally similar to the pattern that Lockheed had tested in the sense that the flow passages throughout most of the regenerator were parallel to its axis (vertical to the page). However, the spacers at both ends of the regenerator were angled in a zigzag pattern in hopes of generating some lateral flow in the channels on the back side of the foil. The purpose was to overcome the flow blockage at the points of contact between the regenerator and the surfaces of the spider and cold heat exchanger. The etch pattern of regenerator #2, shown in Figure 4, was similar to the etch pattern of regenerator #1 except that all of the flow channels were angled in a zigzag pattern relative to the overall direction of flow in the regenerator. The flow channel areas between spacers was still wide.
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As shown in Figure 5, the channels on the back side of the #3 foil, normal to the overall direction of flow, were relatively narrow and were spaced more widely from each other than in prior designs. The flow channels on the front side of the foil were likewise narrower than in previous versions. The foil was perforated at points where the zigzag flow passages on the front side of the foil intersected the straight passages on the back side, but the perforations were more like round holes than the slits that had characterized the earlier patterns. Thus, the pattern in regenerator #3 tended to channel flow more tightly than either of the earlier versions. When fluid flowing in a slanting channel on the front side of the #3 foil reached an intersection, it could bend 90 degrees into the next slanting channel on the front side of the foil or it could bend 45 degrees, pass through the hole to the other side of the foil, and proceed in a direction normal to the overall direction of flow through the regenerator. In practice, some of the fluid is presumed to have taken one path, and some the other. That presumably happened repeatedly, at each intersection as fluid moved through the regenerator. Fluid driven into a channel normal to the overall direction of the flow must eventually have gone somewhere. Presumably it was swept up in the flow moving through zigzag channels offset from the zigzag channels in which it had entered. Thus, the regenerator presumably redistributed flow in both directions normal to the overall flow direction.
COOLING PERFORMANCE Regenerators #2 and #3 were increasingly radical attempts to ensure good distribution of flow around the circumference of the regenerators. The improvement in performance, particularly with #3, was substantial. The improvement, #3 over #1, is shown in Table 1. All numbers are Kelvins. The results labeled “comparable conditions” were for tests run with all other components of the cooler the same. The results labeled “”best conditions” were selected from the entire universe of tests with many different combinations of pulse tubes, cold heat exchangers, warm heat exchangers, double inlet arrangements and reservoirs.
FLOW REDISTRIBUTION To test the effectiveness of the zigzag patterns in redistributing flow, the cold tube of the cooler was placed in a clear acrylic tube in which small (13 mm) holes had been drilled at both ends. The holes permitted some air and water vapor to enter the acrylic tube but the tube itself suppressed breezes and controlled convection. The cooler was then run until a frost ball had formed and the
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outline of the frost ball was observed. Since the annular regenerator was immediately inside the cold tube, any maldistribution of flow in the regenerator was expected to show up in the location of the frost boundary. The results are shown in Figure 6. With regenerator #1, the boundary of the frost ball varied greatly around the circumference of the cold tube, reaching almost to the warm end of the tube on one side. With #2, the frost boundary was again irregular, but less so than with #1. The frost boundary with #3 was almost perfectly regular, and confined almost entirely to the portion of the cold tube that surrounded the cold heat exchanger. Those results appear to confirm the importance of circumferential flow and may supply at least one reason for the superior performance of regenerator #3.
FILL FACTOR AND PRESSURE DROP Regenerators #1 and #2 both have fill factors below 35%. The fill factor in #3 is about 47%, giving it substantially more thermal mass than either of the others. Conversely, the dead volume in #3 is also substantially lower than in #1 or #2. Both differences give #3 an advantage. All three patterns have larger fill factors than the foil tested by Lockheed3. The larger fill factor, and differences in the flow regime, give #3 far more pressure drop than the other two regenerators. Results of pressure drop tests are shown in Figure 7. The implications of the higher pressure drop in #3 are not necessarily adverse. With all types of regenerators, pressure drop and heat transfer go hand in hand. The design objective is simply to obtain the smallest pressure drop commensurate with the desired heat transfer. Early work with a pattern similar to pattern
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#1 demonstrated very low pressure drop relative to comparable screen regenerators.2 It appears quite possible that the pressure drop in that foil was simply too low to permit adequate heat transfer. That hypothesis appears to be confirmed by the results shown in Table 1.
SENSITIVITY TO FREQUENCY The pulse tube cooler in which these regenerators were tested was optimized at between 2 and 3 Hz. Over a range of frequencies from 2.25 Hz to 3.25 Hz, performance changed very little. Figure 8 is a composite of three runs with regenerator #3. It needs some interpretation. Best perfor-
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mance appears to be at about 2.3 Hz. However, during that run, frequency was being ramped down at about 10 minute intervals, starting about 40 minutes into the run. Normally, it took about two hours to reach maximum between the warm and cold ends of the pulse tube, so if the progression had been reversed, the slope would have been somewhat flatter, or possibly even reversed.
CONCLUSIONS Regenerator #3 worked substantially better than regenerator #1. The improvement ranged from 47.5 K at no-load conditions with the other components of the cooler the same to 23.5 K at 10 W load with the best overall arrangement of other components. Regenerator #2 also worked substantially better than regenerator #1, although the improvement was smaller. Progressive improvement appears to have resulted from changes in the patterns etched into the regenerator foil. The degree of improvement exhibited by regenerator #3 suggests that regenerators of that type may now surpass screen regenerators, but no direct comparison was attempted.
ACKNOWLEDGMENT The work described was performed under a Phase II Small Business Innovation Research contract funded by the Missile Defense Agency and Managed by the Air Force Research Laboratory, VSS, Kirtland AFB, NM.
REFERENCES 1. U.S. Pat. 5,429,177; other patents pending. 2. Yaron, R., Shokralla, S., Yuan, J., Bradley, P. and Radebaugh, R.“Etched Foil Regenerator,” Adv. in Cryogenic Engineering, Vol. 41B, Plenum Publishing Corp., New York (1996), pp. 1339-1346.
3. Olson, J. R., Kotsubo, V. Champagne, P. J., and Nast, T. C., “Performance of a Two-Stage Pulse Tube Cryocooler for Space Applications,” Cryocoolers 10, Plenum Publishing Corp., New York (1999), pp. 163-170. 4. Mitchell, M. P., Fabris, D., Sweeney, R.O., “Blind Vortex Tube as Heat-Rejecting Heat Exchanger for Pulse Tube Cryocooler,” Adv. in Cryogenic Engineering, Vol. 47B, Amer. Institute of Physics, Melville, NY (2002), pp. 1093-1100.
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Compact High Effectiveness Parallel Plate Heat Exchangers E.D. Marquardt1 and R. Radebaugh2 1
Ball Aerospace & Technologies Corp. Boulder, CO 80301
2
National Institute of Standards and Technology Boulder, CO 80303
ABSTRACT Compact high effectiveness heat exchangers are required for recuperative cycles such as Joule-Thomson and Brayton coolers. Higher exchanger effectiveness translates directly to system efficiency. A quick survey of the theoretical heat transfer performance of different geometries leads designers to parallel plates as the most compact surface. Unfortunately, parallel plates in practical exchangers have never lived up to their theoretical performance. The primary difficulty with any multiple flow path exchanger is one of mass flow imbalance between the channels. In order to achieve very high effectiveness, over 98%, the mass flow must be balanced to within 2% or better in each channel; even if the heat transfer is ideal. To reach these high levels of effectiveness, the geometry of each channel must be nearly identical. We have developed a method of manufacturing parallel plate exchangers to produce more uniform channels. It is based on photo etching and diffusion bonding. This results in very uniform channel geometry, limited by the thickness variations in the original raw material used in photo etching. An exchanger has been manufactured and tested with a measured effectiveness of 97.3%. While this was short of the design goal of 99.5%, it shows great promise. Future work includes improved measurements, long-life hermetic sealing, header design, pressure drop predictions, and improved modeling.
INTRODUCTION Recuperative heat exchanger effectiveness plays an important role in cryogenic systems. Effectiveness is defined as the ratio of the amount of energy transferred within an exchanger to the maximum possible amount of energy that could be transferred. In many process fields, an effectiveness of 80% is considered high and the difficulties and costs associated with achieving higher effectiveness is not considered worth the effort. In the cryogenics field, 80% is considered a low effectiveness and it is common and often required that effectiveness be at least 90%. Performance of lower effectiveness exchangers is dominated by the heat transfer coefficient and enhancements are largely aimed at its improvement. At higher values of effectiveness, there are additional loss terms that become important factors. These losses include: axial conduction, radiation, flow maldistribution, and wall thermal resistance. Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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In cryocoolers with refrigeration capacities ranging from a few milli-Watts to several 10’s of Watts, exchanger effectiveness becomes even more important and effectiveness greater than 98% is required. Figure 1 plots the required input power verse the effectiveness of the exchanger for a typical Brayton cycle cooler. The input power requirements quickly increase with lower exchanger effectiveness. In general, exchangers with high effectiveness can be achieved in very large systems simply with an increase in surface area. For small cryogenic systems, this can lead to difficulty with mass, volume, and radiation heat loads. The desire for compact size causes the hydraulic diameter to be small or to have high surface to volume ratios. Small hydraulic diameters have low Reynolds numbers, typically less than 200, resulting in laminar flow. It is common in heat exchangers to try to increase the turbulence as a way to increase heat transfer. This is not the case for compact cryogenic exchangers. The small hydraulic diameters create very short entry lengths before the flow is fully developed. Turbulence has very little effect on the heat transfer while the price in pressure drop can be high. Current state-of-the-art in compact high effectiveness exchangers is slotted plate exchangers1 or perforated plates2,3,4. These types of exchangers tend to have small flow passages resulting in high heat transfer coefficients and large surface area to volume ratios, up to While these exchangers can reach effectiveness values of over 99%, to reduce the pressure drop caused by the expansion through each plate, they tend to be very large and heavy. With the more favorable heat transfer and friction factors of parallel plates5, mass can be reduced by 60% and volume by over 70% compared to the most advanced slotted plate exchangers. This offers significant advantages for system integration particularly for space applications but also for terrestrial applications where volume and mass dictate other system requirements. Parallel plates in practical exchangers have never lived up to their theoretical performance. The primary difficulty with any multiple flow path exchanger is one of mass flow imbalance between the channels6. In order to achieve very high effectiveness, over 98%, the mass flow must be balanced to within 2% or better in each channel; even if the heat transfer is ideal. Since energy is transferred between the streams and is not stored, if the mass flow is imbalanced between the streams, there is no place for the energy to go. To reach these high levels of effectiveness, the geometry of each channel must be nearly identical. Consider two sets of identical channel pairs, each set transfers its energy only to its mated pair, If the flow is distributed 50% in each set and the heat transfer is perfect, we can have an effectiveness of 1 if longitudinal conduction is ignored. Now let the cold side of the exchanger perfectly split the flow
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between the two channels but let the hot side split the flow 49%/51%. With perfect heat transfer, we can only transfer 49% of the energy in one set and 50% of the energy in the other set of channels, resulting in a 99% effectiveness. This problem of flow maldistribution is the primary difficulty with multi-passage exchangers. Other methods have been tried to create parallel plates. Spirally wrapped foil has been tried, for regenerators7, to create a parallel plate geometry but has end effect issues and a problem with maintaining the coils tightly and uniformly packed. Exchangers where the plates are machined or depth-etched suffer from a lack of uniform flow channel thickness resulting in flow maldistribution. In addition, the pressure differential between the flow streams can cause the plates to bow. A new method of manufacturing parallel plate exchangers has been developed that produces more uniform channels. It is based on photo etching and diffusion bonding. This results in very uniform channel geometry, limited by the thickness variations in the original raw material used in photo etching.
HEAT EXCHANGER MODEL The design of a heat exchanger requires several parameters determined from a higher level system model. The required parameters include the overall effectiveness, inlet temperatures for both streams, and the working fluid and pressure drop for each side of the exchanger. Given these parameters, there are an infinite number of solutions for the geometry of the exchanger, so we can search the solution space and find the smallest size exchanger. It can be shown that
Where is the gas cross-section flow area, is the mass flow rate, is the number of transfer units, is the Prandtl number, is the density, is the ratio of the product of the Stanton number and the Prandtl to the 2/3 power to the friction factor, and is the pressure drop. Equation 1 provides a fundamental base for heat exchanger design showing that the gas flow area is linearly proportional to the mass flow rate. It shows that a high value of meaning a high value of the ratio of the heat transfer to the coefficient of friction, results in the most compact heat exchanger. For parallel plates, the gas cross-sectional area, is equal to the product of the gap width, w, and thickness, therefore the required gap width, in terms of the design variables, can be written as
Where w is the width and is the thickness of the gas flow channel as shown in Figure 2. Equation 2 is used for both the hot and cold sides of the exchanger. The used is per side so it is higher than the design Since the width must be the same for both sides, the larger value is used for the exchanger width. This is the total width of the exchanger and can be divided into a number of parallel flow channels. A typical example might find the total gap width to be 10 cm, using 10 parallel channels, the width of the exchanger is 1 cm. The length of the exchanger is determined by balancing the heat transfer and the pressure drop between the two sides and is determined using
where L is the length and µ is the viscosity. Since the length must be the same for both sides, the pressure drop cannot be independently set for both sides using fixed values of the channel thickness. An iterative method is used. The length is calculated for each side of the exchanger using the same pressure drops used in Equation 2. The lengths are compared and the pressure
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drop on the side with the longer length is reduced and Equations 2 and 3 are repeated until the length of each side is equal. Finally, the pressure drop can be calculated using the determined geometry and verified against the value used in equation 1. The equation for the pressure drop in a parallel channel is
EXCHANGER DESIGN Optimization A heat exchanger was designed for the Ball Aerospace 10 K cooler described by Glaister et al8. Equations 1-4 were solved for a large number of different values of the gap thickness. The results of the predicted exchanger mass are plotted in Figure 3. A mass of under 140 grams is found at 0.0048” for the cold-side and 0.0031” for the hot-side. Since the contours are fairly flat as the gap thickness is increased, the cold-side gap of 0.005” and hot-side gap of 0.004” was selected. This provided a mass estimate of 150 grams, not including the header mass. The header mass includes the manifold to direct the flows into and out of the exchanger. Figure 4 shows a photograph of the exchanger and the cross-section.
Expected Performance Figure 5 shows the expected sample exchanger performance operating under the design conditions. The figure also includes the two parts that lead to the overall effectiveness, the heat transfer and the conduction loss. This shows how at low mass flow rates, where the capacity rate of the flow is low, the conduction loss, which is constant for any given temperature difference, dominates the overall effectiveness. At high mass flow rates, the conduction is a small term and the heat transfer losses dominate the overall effectiveness.
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EXPERIMENTAL TESTING Apparatus A schematic diagram of the experimental system is shown in Figure 6. The system was designed to control the inlet temperatures of both flow streams as well as the pressure and mass flow rate. Measurements are made of the temperatures and pressures at both inlets and outlets of the exchanger. The system includes a Gifford-McMann (GM) cooler which provides the
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temperature sinks. Various other temperature sensors are used to monitor the complete system operation. A mass flow controller is used which also controls the backpressure. The flow path starts at the helium gas bottle and the pressure is controlled by a pressure
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regulator. The gas is then passed through a liquid nitrogen trap to remove any contaminates from the gas stream. The gas then flows through the recuperative precooler heat exchanger. It is next cooled to the GM first-stage temperature in the exchanger built into the radiation shield which runs slightly lower in temperature than the desired high-temperature exchanger inlet. Next, the gas flows through the hot sink that controls the gas temperature to the test exchanger highpressure stream. After passing through the high-pressure side of the sample exchanger, the gas flows through the GM cold sink that is maintained at the desire low-temperature stream inlet. After passing through the cold side of the sample exchanger, the gas enters the recuperative precooler. Finally, the mass flow controller is used to control the mass flow rate before the gas exits to the atmosphere. The gas pressure and temperature are measured at all four inlets and outlets to the sample exchanger.
Effectiveness Measurement Techniques The primary goal of the experimental system is to verify the thermal performance of the new heat exchanger. There are two methods used to measure the effectiveness. The simplest way is to measure the amount of cooling required to make the hot stream exit gas temperature equal to the cold stream inlet temperature. This method is limited to using the same pressure on both sides of the exchanger, not allowing effectiveness measurement under the design pressure. The second is the measurement of all the inlet and outlet temperatures. This method works at any temperature or pressure. The first simpler method can be accomplished by measuring the change in the heat input to the second-stage of the GM cooler as a function of the mass flow rate. If the input power to the GM cooler is constant, then to hold the temperature of each stage constant requires a fixed load heat load. Since the temperatures are held constant, the radiation and conduction losses are also constant. Therefore, the only variable heat load to the second-stage of the GM cooler is the energy required to cool the gas to the cold-stream inlet temperature. The inlet temperature to both flow streams can easily be controlled, but the disadvantage of this method is that the pressures cannot be independently controlled between the two streams. To control the pressures, there must be a pressure drop between the hot-stream exit and the cold-stream inlet. The pressure drop will cause the enthalpy to change and adds another unknown heat load to the GM secondstage. The second method to measure the effectiveness requires the measurement of all four temperatures. This is a direct measurement of the effectiveness if the exchanger is balanced. If the exchanger is not balanced, the measurement can only be made at one end of the exchanger, the hot end if the hot flow stream has the minimum capacity or the cold end if the cold flow stream has the minimum capacity. This is a difficult measurement to make since the temperature difference between the inlet and outlet temperatures at each end are less than one degree. It is not possible to get an accurate measurement of the gas temperature by measuring the outside of the tube since the tube also includes a conduction term. The mass of the exchanger will dominate the tube temperatures close to the exchanger, providing a lower measurement of the true temperature difference. Although the temperature gradient is small between the exchanger and the temperature sink, the tube wall temperature includes a conduction and convection term and the temperature at any location is dependent on the mass flow rate. To measure the small temperature difference, a differential thermocouple was placed directly in the flow stream. A three-wire arrangement was used, two constantan and one chromel wire were run from room temperature. By using this arrangement, it is possible to measure the small temperature difference and also the absolute value of the temperature.
Results A number of sample runs were made, Figures 7 and 8 show typical results. As can be seen in Figures 8, there is a discrepancy of 2% between the effectiveness measurements using the two different methods. It is difficult to see how the GM heat load measurement could be wrong.
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Since the conduction and radiation loads don’t change, the additional heat load with varying mass flow must be from the ineffectiveness of the exchanger. The differential thermocouple measurement is a new technique but is a fundamentally sound measurement. The explanation of the different measurements was found by examining the longitudinal temperature profile along the exchanger. We would expect a linear temperature profile, but the measured was nonlinear. The hot end of the exchanger is similar to the hot fluid inlet temperature but the cold end is ~30 K above the cold fluid inlet temperature.
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The exchanger was built with 10 cold flow channels and 11 hot flow channels. The two outside hot flow channels were designed to carry ½ the mass flow compared to the other hot flow channels to allow for a uniform heat flux. Since the cross-axis thermal conduction is very poor, this profile indicates that the flow in the outside channels is too large. The flows would start hot and transfer as much energy as possible to the adjacent cold flow stream. The cold steam is limited in its capacity so it is unable to completely cool the outside hot stream. By estimating the remaining heat capacity in the outside streams, the mass flow can be estimated. Several assumptions must be made to estimate the mass flow rate in the outer flow channels. First, it should be noted that the differential thermocouples were placed very close the outlet of the exchanger. The tip of the thermocouples is less than five hydraulic diameters from the exchanger, it is unlikely that the flow has had time to fully develop or mix thoroughly. If it is assumed that the thermocouples measure the effectiveness of the central channels, not including the two outside channels, the heat lost in the outer flow streams can be calculated. The total mass flow in two outer streams can be calculated from the following equations.
Where is the heat not transferred in the exchanger, is the heat transferred from the outer flow passages, is heat transferred between the inner flow passages, is the maximum possible heat transferred between the hot and cold stream, is the effectiveness measured with the differential thermocouple, is the mass flow in the outer channels, is the specific heat at constant pressure, is the outer flow channel outlet temperature, and is the cold sink temperature. Conduction will modify the outlet temperature so the highest mass flow rates will provide the best estimate of the mass flow rate in the outer streams. At 100 mg/s, is 2.362 W, is 88.4 W, is 0.992, is 5.2 J/g-K, is 67.5 K, and is 40 K. This results in a mass flow rate of 11.6 mg/s compared to the desired flow of 10 mg/s, 16% high.
FUTURE WORK There are a number of areas that require additional work. With the promising initial results, additional funding has been made available to continue this work at Ball Aerospace. The most important task is to develop a process to hermetically seal the exchanger. The epoxy used on the test exchanger posed a number of problems and would not be acceptable for any long term testing. Two additional exchangers have already been manufactured where the edge of the exchangers were silver soldered. This process also requires some development, three attempts were required to seal the exchangers. A more consistent process is required, The mass flow in the outside end flow channels must be controlled to ½ the primary channel flow rate. This was the most significant factor in the lower effectiveness of the test exchanger. The strength of the diffusion bond needs to be tested and the process better understood. This is a critical safety issue and affects the support structure required for the exchanger to survive the high vibrations levels during launch.
REFERENCES 1.
Swift, W., “Recent Developments in Components for Reverse Brayton Cryocoolers,” in Proceedings of the Fourth Interagency Meeting on Cryocoolers, David Taylor Research Center, Bethesda, MD (1991), pp. 51-60.
2.
Venkatarathnam, G. and Sarangi, S., “Matrix Heat Exchangers and Their Application in Cryogenic Systems,” in Cryogenics, Vol. 30, Butterworth-Heinemann Ltd. (1990), pp. 907-918.
3.
Fleming, R.B., “A Compact Perforated-Plate Heat Exchanger,” in Advances in Cryogenic Engineering, Vol. 14, ed. K. Timmerhaus, Plenum Press, New York, NY (1969), pp. 197-204.
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Vonk, G., “A New Type of Compact Heat Exchanger with a High Thermal Efficiency,” in Advances in Cryogenic Engineering, Vol. 13, ed. K. Timmerhaus, Plenum Press, New York, NY (1968) pp. 582-589. 5. Kays, W.M. and London, A.L., Compact Heat Exchangers, Third Edition, McGraw-Hill, Inc., New York, NY (1984). 6. Fleming, R.B., “The Effect of Flow Distribution in Parallel Channels of Counterflow Heat Exchangers,” in Advances in Cryogenic Engineering, Vol. 12, ed. K. Timmerhaus, Plenum Press, New York, NY (1967), pp. 352-372. 7. Rawlins, W.C., The Measurement and Modeling of Regenerator Performance in an Orifice Pulse Tube Refrigerator, submitted for partial fulfillment for Ph.D. degree, University of Colorado, Chemical Engineering (1992). 8. Glaister, D.S., Gully, W.J., Wright, G.P., Simmons, D.W., and Tomlinson, B.J., “A 10 K Cryocooler for Space Applications,” Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2001), pp. 505-511. 4.
Measurement of Heat Conduction through Bonded Regenerator Matrix Materials M.A. Lewis and R. Radebaugh National Institute of Standards and Technology Boulder, Colorado, USA 80303
ABSTRACT Regenerative heat exchangers have had a significant influence on the development of small refrigerators for cryogenic applications. The optimized design of these regenerators takes into account the axial thermal conduction of the matrix. Until recently this thermal conduction has been unknown even for the commonly used screen or packed sphere matrices. Research at NIST on the thermal conduction through such matrices has shown that the thermal conduction is best represented by a thermal conductivity degradation factor. We have given this factor previously for stacked metal screens of various mesh and porosities and for packed spheres of various metals. This factor is important in optimizing the geometry of the stacked screens or packed spheres. In this paper we discuss the measurements of the thermal conduction in regenerator matrices when they are bonded either by sintering or with the use of thinned epoxy. Such bonded matrices offer some advantages in the fabrication of regenerators. For example, the uniform stacking of large diameter screen matrices with negligible gaps around the circumference can be difficult to achieve. Also, the containment of fine metal powders can be difficult. The bonding of these matrices can solve many of these fabrication problems, but could possibly be a disadvantage because of enhanced thermal conduction. Experimental results with diffusion-bonded 325-mesh stainless steel screen and epoxy-coated lead spheres are presented in this paper. The results show only a small increase in thermal conduction, which does not significantly affect the overall cryocooler performance.
INTRODUCTION The use of stacked stainless steel screens and packed beds of lead spherical material are commonly used in the design of cryocooler regenerators.1 Research has been done at NIST on the thermal conductivity characteristics of some of the more commonly used materials for both screen and spherical matrices. Some regenerator fabrication utilizes the techniques of bonding the regenerator material together for more simplified installation, maintaining the integrity of the material or containment of loose particles from system contamination. This research examines the effects of diffusion bonding stainless steel screen stacks and the manufacturing technique to epoxy bond lead spheres into a monolithic yet porous bed. The results of the experiments are compared with the previous data of comparable non-bonded materials.2,3
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EXPERIMENTAL APPARATUS AND PROCEDURE Figure 1 shows the experimental apparatus used for the study. The apparatus consists mainly of a test section, a two-stage Gifford-McMahon (GM) cryocooler, a heat flow sensor, and a vacuum vessel (not shown in Fig. 1). The details of the test section are shown in Fig. 2. Two identical regenerators are used in this apparatus. Regenerator cylinders are made of fiberglass-epoxy, with an inner diameter of 24.4 mm and a length of 55 mm. The wall thickness of each cylinder is 1 mm. Heat conduction along the length of the cylinder wall of a single regenerator from room temperature to 80 K is estimated to be 0.21 W from published thermal conductivity data.4 Helium gas lines are connected to the regenerator to change the filling pressure in the regenerators. The pressure can be varied from vacuum to 2.0 MPa. Multi-layer insulation was wrapped around the test section to reduce radiation heat loss. The cold ends of both regenerators are connected to a cold plate, which is cooled by the GM cryocooler by the heat flow sensor. The hot ends of both regenerators are capped by piston-shaped water jackets. Flowing water maintains the hot end temperature at room temperature. A bellows is attached to the lower water jacket. The cold plate and the two regenerators are free to move with respect to the water jackets, so the force exerted by the bellows is applied equally to both regenerator columns of bonded matrix materials. The heat flow sensor is mounted between the cold plate and the first stage of the GM cryocooler. A flexible thermal link between the cold plate and the heat flow sensor allows for movement of the cold plate when the bellows pressure is changed. The heat flow sensor consists of a copper bar and two silicon diode thermometers. The copper bar is made of oxygenfree copper with a cross-sectional area of and a length of 135 mm. The distance between the two thermometers is 91.3 mm. The relationship between heat flow through the copper bar and temperature difference was calibrated before these experiments by using a heater attached to the cold end. The experimental procedure is as follows. After pumping the vacuum vessel, the twostage GM cryocooler is turned on. Both the cold plate and the heat flow sensor are cooled by the first stage of the GM cryocooler. The cold plate temperature is kept at a constant temperature by the temperature controller using a silicon diode thermometer at the cold plate and an electric heater mounted on the GM cryocooler first stage. The cold plate temperature can be varied over a wide temperature range but for these tests was maintained at 80 K. The temperature stability could be maintained to within ±0.05 K. Once the cold plate temperature is set
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and the temperature difference at the heat flow sensor is measured, an additional heat load is supplied to a heater mounted on the cold plate, and its effect on the temperature difference at the copper bar is measured. The heat loads, the temperature differences, and the calibration curve obtained previously are used to calculate the heat flow through the heat flow sensor. The calculated heat flow here includes the heat conduction through both bonded matrix materials, the two fiberglass-epoxy cylinders, and other heat losses, such as radiation loss and heat conduction loss through instrument wires. In a separate experimental run, the heat flow through the regenerators without any matrix material and with a vacuum inside was measured to provide information needed to determine the heat flow through the columns of matrix material only.
REGENERATOR SAMPLES The materials used for the experiments were sintered stainless steel screens and epoxy bonded lead spheres. The 316L stainless steel screens were 325-mesh with 35.6µm (0.0014”) wire diameter. The screens were cut into a square pattern and stacked to a specified height. The material was heated in a vacuum furnace to diffusion bond the contact points of the stainless steel. This procedure produces a block of stainless steel screen material. The diffusion bonded (sintered) block of screen was commercially fabricated. Our experience has shown that the diffusion bonding process takes place at a temperature of approximately 1100 °C. The stainless steel block was then machined using a wire electrical discharge machining (EDM) process. The cylinders were a sliding fit inside the G-10 cylinders used in the experimental apparatus. This sample was shorter than the existing G-10 cylinders so copper discs were made to fill the volume in the G-10 regenerator. Knowing the geometry as well as the mass of the sample allows us to calculate the porosity of the material, which for this sample was 0.6043. This porosity is lower than the standard value of 0.64 for 325-mesh screen. This lower porosity of the sintered screen is caused by the preparation and diffusion bonding procedure for the material. The electrical resistivity of a similar sintered screen cylinder was measured and found to be when the sample porosity is taken into account in the calculation of the crosssectional area. When compared with solid stainless steel with a resistivity of the ratio suggests that according to the Wiedeman-Franz Law the thermal conductivity degradation factor will be about 0.025 under vacuum conditions. The epoxy coated lead spheres were cast into a thin-wall G-10 housing utilizing a cryogenic epoxy thinned with acetone.4 The mass of epoxy remaining after being cured is only 0.11 % of the mass of the bonded spheres. The small amount of epoxy used has negligible influence on the gas flow characteristics of the regenerator. The spheres were made of 95% Pb + 5% Sb and had a diameter of 102 ± 13 µm. The cured sample is a monolithic regenerator material that provides a solid and porous test section for our experiments. Knowing the geometry and the mass of the material allowed us to calculate the porosity of the material, which was an average of 0.3881 for both regenerator sections. Characteristics of the epoxy packed bed regenerators with the G-10 housing are given in Table 1. The G-10 housings had a 0.25 mm wall thickness that fit precisely into the existing G-10 regenerators of the experimental test apparatus. The length of the G-10 housings matched the existing G-10 regenerator sections.
EXPERIMENTAL RESULTS AND DISCUSSION The first measurement was of the heat leak through the G-10 regenerator cylinders without matrix material inside and under high vacuum conditions. The cold plate temperature was 80 K
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and the hot plate temperature was 285 K. The measured total heat leak was 1.06 W, with 0.42 W the calculated heat conduction through the cylinder wall of the two regenerators. These measurements gave a total background heat leak of the test apparatus and this value would be subtracted from all heat leak measurements for the materials tested. The regenerator test matrices for each material were installed into the experimental test section and the heat leak was measured from vacuum condition up to a helium gas pressure of 2 MPa. The effects of the helium gas pressure over the operating pressure range can be seen in Figure 3 and Figure 4. The heat leak increases with the increasing helium pressure until there is little significant change for pressures above 0.5 MPa. This can be explained by the heat transfer mechanism in the regions of molecular flow, viscous flow and intermediate flow along with the pressure dependence of mean free path and thermal conductivity of helium gas. The heat leak rises in the molecular flow region where the heat transfer is proportional to gas pressure and levels out in the viscous flow region where the heat transfer is proportional to the thermal conductivity and is independent of pressure for an ideal gas. The effects of mean free path indicate that most of the heat is transferred by the helium gas in a region very close (~3µm) to the individual contacts between the wires or the spheres. Figure 3 shows the heat leak values of the sintered 325-mesh stainless steel screen. These results are plotted with previous research done at NIST with standard 325-mesh stainless steel screens packed into the same G-10 regenerators and using the same test apparatus. The heat leak values of the screens were measured under various porosities over the same He pressure range. The data shows how the heat leak values of the sintered screens are approximately 2.5 times higher than that of the average value of the standard packed screens at 2 MPa. This higher heat leak is a result of the enhanced thermal contact at the diffusion-bonded metal-tometal contacts as well as a result of the higher cross-sectional area of the metal due to a lower porosity. In addition the sintered screen samples were shorter than the packed screen samples. These differences in porosity and length will be factored out in the calculation of the thermal conductivity degradation factor. Figure 4 shows the heat leak values of the epoxy bonded lead spheres. These data are plotted with previous data taken at NIST using standard packed beds of lead spheres utilizing the same experimental test section and apparatus. The data show that the heat leak value for the epoxy bonded spheres is approximately 1.35 times higher than the value for the standard packed sphere bed at 2 MPa. This heat leak increase can be contributed to the enhanced thermal contact of the boundaries due to the epoxy. The porosity values for the epoxy bonded lead spheres and the standard lead sphere packed bed were 0.3881 and 0.38 respectively. Because of the geometry of the stainless steel screen and the lead sphere matrix, the thermal conduction is reduced compared to a solid bar of the same material and same crosssectional area as the metal in the regenerator cylinder. Therefore to estimate this heat leak through the matrix columns, a conduction degradation factor is applied. The actual conduction through the regenerator matrix is then given as:
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where Heat flow through matrix material Conductivity degradation factor Total cross-sectional area of matrix Length of matrix Porosity of material Temperature at cold end of regenerator Temperature at hot end of regenerator Thermal conductivity of regenerator matrix material In the calculations for these measurements the integrated values of thermal conductivity between 80 K and 290 K for the stainless steel and the 95% Pb + 5% Sb were 25.26 W/cm.K and 55.07 W/cm.K, respectively.5,6 Figure 5 and Figure 6 show how the conductivity degradation factor varies with increasing pressure within the regenerator for the 325-mesh stainless steel screen and the packed bed lead sphere material of both the bonded matrix and the packed bed regenerators. This degradation factor begins to level out at a pressure of 0.5 MPa, which is explained by the effects of pressure relative to mean free path and thermal conductivity. The data show that the conductivity degradation factors for both of the bonded materials are higher than that of the non-bonded screen and sphere material. This is consistent with the heat leak data as shown previously. The conductivity degradation factor for the sintered stainless steel screens was 0.18, and for the epoxy bonded lead spheres the factor was 0.12 for pressures above 1 MPa. The data show that the comparison of the heat leak data to the conductivity degradation factor has a consistent difference between the lead sphere materials but that the comparison is quite different for the stainless steel material. This can be explained by the length differences of the matrix material and by the different porosity values. The lengths of the lead material were consistent and the porosities were about the same.
CONCLUSIONS Measurements were performed to observe the thermal conduction properties of sintered stainless steel and epoxy bonded lead spheres and the results compared to measurements taken on stacked stainless steel screens and packed lead sphere beds. These bonding procedures are helpful for some regeneration construction and the effects of performing these bonding procedures would be helpful in evaluating loses associated with regenerator performance.
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Measurements as a function of helium gas pressure indicate that most of the conduction between the individual wire or spheres is through the helium gas within a few microns on the contact point rather than directly through the contact. Bonding the layers by sintering or with thinned epoxy increases the conduction across the contact points, but that the conduction remains small compared with the conduction through the helium gas near the boundary. Thus, the increased conduction caused by epoxy bonding or sintering is small enough so as not to significantly degrade the performance of regenerative cryocoolers that use the bonded regenerators. The thermal conductivity degradation factors presented here for the bonded regenerators allow their geometry to be optimized by using REGEN3.2 or other design software.
ACKNOWLEDGMENT The authors would like to thank John Barclay of CryoFuel Systems for the fabrication of the epoxy bonded lead regenerator samples.
REFERENCES 1.
Walker, G., Cryocoolers, Plenum Press, New York (1983).
2.
Lewis, M.A., Kuriyama, T., Kuriyama, F., and Radebaugh, R., “Measurement of Heat Conduction Through Stacked Screens,” Advances in Cryogenic Engineering, Vol. 43, Plenum Press, New York (1998), pp. 1611-1618.
3.
Lewis, M.A., and Radebaugh, R., “Measurement of Heat Conduction Through Metal Spheres,” Cryocoolers 11, Kluwer/Plenum Publishers, New York (2001), pp. 419-425.
4.
Mérida, W.R., and Barclay, J.A., “Monolithic Refrigerator Technology for Low Temperature (4K) Gifford-McMahon Cryocoolers,” Advances in Cryogenic Engineering, Vol. 43, Plenum Press, New York (1998), pp. 1597-1604.
5.
Radebaugh, R., Gary, J., Marquardt, E., Louie, B., Daney, D., Arp, V., and Linenberger, D., Measurement and Calculation of Regenerator Ineffectiveness for Temperatures of 5 to 40 K, Flight Dynamics Directorate, Wright Laboratory, March, 1992.
6.
Childs, G.E., Ericks, L.J., and Powell, R.L., Thermal Conductivity of Solids at Room Temperature and Below, U.S. Department of Commerce, NBS Monogram 131.
Regenerator Loss Measurements at Low Temperatures and High Frequencies J.M. Pfotenhauer University of Wisconsin – Madison Madison, Wisconsin, USA 53706 P.E. Bradley, M.A. Lewis, and R. Radebaugh National Institute of Standards and Technology Boulder, Colorado, USA 80305
ABSTRACT We report the design of an experiment characterizing the loss mechanisms in a regenerator operating with a cold end temperature between 10 and 20 K, and in a frequency range of 30-60 Hz. The regenerator is designed as part of a two-stage pulse tube system, incorporating 400mesh stainless steel screens and 100-micron lead shot in the first and second stages respectively. The sinusoidal pressure wave generated by the linear flexure-bearing compressor provides an average pressure of 2.5 MPa and a pressure ratio of 1.4. Flow rates at both the warm and cold ends of the regenerator are measured as well as the pressure and temperature at the ends and intermediate locations. A laminar flow element has been designed to measure flow at the warm end, while the pressure oscillation in a cold reservoir permits calculation of the flow rate at the cold end. A cold inertance tube provides the desired phase shift. Loss measurements on the regenerator alone are achieved without the use a pulse tube by anchoring the intermediate and cold end temperatures to the first and second stages, respectively, of a Gifford-McMahon cryocooler. Loss mechanisms including conduction, radiation, PV degradation, and regenerator ineffectiveness will be measured as a function of cold end temperature and frequency. The experimental apparatus allows future measurements to be taken of the pulse tube losses in the same temperature and frequency range.
INTRODUCTION The REGEN code developed at NIST1,2 is routinely used to optimize the design of regenerative heat exchangers over a wide array of thermal, material, and geometric parameters. Confirmation of its accuracy has been demonstrated at temperatures down to 4 K for low frequencies (1 – 2 Hz)3, however, very little data is available to check the accuracy for frequencies above 20 Hz and for temperatures below 25 K Because of the significantly higher efficiencies of pulse tube refrigerators operating between 30 – 60 Hz as compared to the 1-2 Hz variety, and the continuing development of low temperature pulse tubes and regenerative materials4,5, there is a growing interest to understand and quantify regenerator loss mechanisms in the low temperature, high frequency regime. Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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An experiment has been designed to characterize losses in the respective temperature and frequency range of 10 – 20 K and 30 – 60 Hz for a regenerative heat exchanger and simple orifice pulse tube refrigerator utilizing 1 kW of room temperature acoustic power from a linear compressor. Separate attention is given to losses in the regenerator associated with a) thermal conduction and radiation, b) heat exchanger ineffectiveness, and c) acoustic power loss. The composite of losses in the pulse tube from conduction, radiation, shuttle heat transfer, flow transition, and other factors will be measured as a function of temperature, frequency, and orientation angle. In the initial phase of the experiment regenerator losses are measured without a pulse tube and temperature control is achieved with a two-stage GM cryocooler. The component designs of particular interest that are reported here include those of the aftercooler, the inertance tube, and the flow meters at the warm and cold end of the regenerator. Future measurements and characterization of the pulse tube will appear in a subsequent publication.
EXPERIMENT DESIGN The components used to characterize the regenerator losses are shown in Figs. 1a and 1b. The regenerator assembly, show in detail in Fig. 1a, is comprised (in order from bottom to top) of the aftercooler, a laminar flow meter, the 1st stage regenerator, the ‘80 K’ thermal intercept and attached radiation shield, the 2nd stage regenerator, the ‘10 K’ thermal intercept, a flow transition piece, the inertance tube, and the reservoir. Figure 1b displays the regenerator assembly mounted directly on the top of a linear compressor and connected to the 1st and 2nd stages of a GM cryocooler. Not shown are the water-cooling lines to the compressor and aftercooler. The complete assembly of the regenerator, compressor, and GM cryocooler is contained within a vacuum vessel that can be rotated about its bottom horizontal axis. Relevant dimensions and characteristics of the various components are given in Table 1.
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In the paragraphs that follow, the design approaches for the various components are described in the order that they are associated with the three loss categories for the regenerator: 1) thermal conduction and radiation, 2) heat exchanger ineffectiveness, and 3) acoustic power loss.
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Conduction & Radiation loss Thermal losses associated with conduction through the regenerator walls, the regenerator matrix and the interstitial helium gas, and radiation exchange between the cold components and the warm surroundings are governed by the axial temperature profiles of these components. A close approximation of such losses realized during operation with a pulse tube will be obtained through temperature control at the warm, intermediate, and cold locations through the use of the aftercooler, and the thermal intercepts connected to the GM cryocooler. During this phase of the test, the compressor and regenerator assembly will be pressurized to their average operating pressure of 2.5 MPa, and cooling water flowing through the compressor and aftercooler will be used to control the ambient (warm-end) temperature, but the compressor will not be operating. The heat flow to each stage of the GM cooler will be measured by pairs of calibrated thermometers mounted at precise separations on their respective bus bars. ‘Baseline’ heat flow measurements associated with the GM cryocooler and copper buses will be initially gathered with the thermal buses disconnected at the thermal intercepts. The various conduction and radiation contributions to the measured heat flow will be determined by taking measurements at a variety of controlled temperatures and through the use of energy balance relations:
Here the respective baseline and total heat flows to the 1st and 2nd stage of the GM cryocooler are and the respective radiation load to the components between the ambient and 1st stage temperatures and between the 1st and 2nd stage temperatures are and me respective conduction load through the walls of the 1st and 2nd stage regenerator are and and the respective conduction load through the matrix and helium gas in the 1st nd and 2 stage regenerators are and Equations (la) and (1b) can also be used to define the effective emissivity in if the helium-pressure-dependence of is known7,8, or to determine the helium-pressure-dependence of if the emissivity in is known. Heat is intercepted at the 1st and 2nd stage regenerator temperatures through 9.6 mm thick by 34.8 mm diameter stacks of 100 mesh copper screen, each of which are diffusion bonded to solid copper tabs connecting to their respective thermal buses (see Fig. 2). The estimated heat flows
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during full operation to the 1st and 2nd stages are 13 watts and 1 watt respectively. Combining these values with the dimensions and conductivities of the thermal paths between the regenerator and GM cryocooler results in respective temperature drops along the 1st and 2nd stage paths of 20 K and 2.5 K. The expected operating temperatures of the GM cryocooler (Ebara ICP200) with such loading are 50 K and 11.5 K at the 1st and 2nd stages respectively, so that the minimum associated intercept temperatures expected at the regenerator are 70 K and 14 K.
Heat exchanger ineffectiveness Ineffective heat exchange through the regenerator will be quantified by comparing the increased heat flow at the 1st and 2nd stages of the GM cryocooler when the compressor is on with the same values measured when the compressor is off. In this test, temperatures at the warm, intermediate and cold ends of the regenerator will be held constant. It is assumed that no cooling will be effected by the inertance tube and reservoir that are attached to, and isothermal with, the cold end of the regenerator. The aftercooler is designed using the ISOHX code6 to maintain near-ambient temperature at the compressor outlet. When properly optimized, the aftercooler should minimize both the amplitude of temperature oscillations and the average temperature difference between the gas and ambient conditions at the heat exchanger outlet. The first requirement ensures minimal enthalpy flow into the regenerator, and the second minimizes an efficiency loss associated with increased compressor temperature. The goals for the present design were: a) to maintain the amplitude of temperature oscillation at the aftercooler outlet below 3% of the inlet amplitude, b) maintain the outlet average temperature at less than 30 K above ambient, c) minimize the ratio of pressure-loss/average-pressure to less than 3% of the ratio of dynamic to average pressure, and d) limit the void volume to less than 10% of the regenerator void volume. As shown in Fig. 3, the aftercooler design satisfying the above constraints is comprised of a 6.35 mm thick by 50.8 mm diameter stack of 80 mesh copper screen penetrated near the mid-plane by six 3.2 mm diameter water-cooled copper tubes. The tubes and water flow are designed to extract up to 1.4 kW of heat while maintaining the water pressure drop below 241 kPa. The mean spacing of the six tubes in fact minimizes the horizontal temperature variation in the screens to less than 5 K above ambient. The outlet temperature amplitude is calculated to be less than 1.5% of the inlet amplitude, the pressure loss is 5% of the dynamic to average pressure ratio, and the void volume is 9% of the regenerator void volume.
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Acoustic power loss The acoustic power flow is defined by the equation
where and are the dynamic and average pressures respectively, and are the volumetric and mass flow rates respectively, R is the ideal gas constant, and is the average temperature. and represent the vector magnitudes of pressure, volumetric and mass flow rates at a given location in the system, while is the phase angle between the pressure and either the volumetric or mass flow rate. Viscous losses in the regenerator cause a decrease in the dynamic pressure amplitude. The loss of acoustic power through the regenerator will be determined by measuring both the pressure and mass flow waves at each end of the regenerator. Pressure taps at each end will record the pressure waves, while the mass flows will be measured by the laminar flow meter at the warm end, and through the pressure oscillation in the reservoir at the cold end. The inertance tube between the regenerator and reservoir is included to produce the desired phase angle between the mass flow and pressure waves at the regenerator cold end. Laminar flow meter. The constraints imposed on the warm-end flow meter included a) maintaining a dead volume below 10% of the compressor swept volume, b) maintaining a Reynolds number below 2000, and c) producing a ‘measurable’ pressure drop - on the order of 10 kPa. Two different flow geometries have been investigated – a set of parallel tubes, and a set of parallel slots. In the parallel tube geometry, the fixed Reynolds number and peak mass flow determine a fixed product of the number (n) and diameter (d) of the tubes. Since the volume scales as it can be minimized with a large number of very small diameter tubes. Unfortunately, for the compressor swept volume of the volume constraint requires n > 5000 and d < 0.5 mm – a significant fabrication challenge. The parallel slot approach provides a less challenging fabrication scenario. The hydraulic diameter for a slot whose cross section is long by ‘w’ wide is 2w when The Reynolds number is then given by
which, when limited to 2000, and combined with a peak flow limit of 30 g/s requires that The slot geometry has the added benefit that a smaller slot width provides both a smaller volume and a larger pressure drop. The capabilities of electron-discharge-machining (EDM) place a practical limit on w to ~ 0.15 mm, but this provides a reasonable solution for the three constraints. An end-view of the laminar flow meter satisfying the constraints is shown in Fig. 4. Each of the six symmetric grooves have been machined using EDM by starting at the outer diameter, cutting the circumferential groove for ~ 60°, moving in to a smaller radius, cutting back along the ~60°, moving in to a smaller radius, and so on. The length of the EDM wire extends the full 38.1 mm length of the piece, and is cut when the groove reaches the inner radius. The total cross-sectional flow area provided by the slots of combined with the flow length of 38.1 mm produce a total dead volume of and with a 25 g/s flow rate, provides a pressure drop of 8 kPa. Cold end mass flow rate. If one assumes that the pressurization associated with the mass flow into and out of the reservoir occurs adiabatically, then the energy balance associated with this process is given by
where the internal energy (u) and enthalpy (h) of the mass (m) are represented in the initial (i), final (f), and incoming (in) states by the corresponding subscripts. From equation (4), one can
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show that the final temperature (T) and pressure (P) associated with adiabatic changes in the reservoir are given respectively by
where is the specific heat ratio The measured pressure oscillations in the reservoir will be used along with equations (5a) and (5b) to determine the associated mass flow rates. Intertance tube. Because both the ineffectiveness and pressure loss (the dominant regenerator losses) are proportional to the average mass flow rate in the regenerator, it is desirable to minimize the average amplitude of mass flow through the regenerator. This can be accomplished by fixing the phase angle at the cold and warm ends of the regenerator to be of equal but opposite sign with respect to the pressure phasor. The changing mass in the regenerator volume provides the connection between the warm and cold end mass flow rates through the vector mass balance equation:
Thus, adjusting the phase angle at the regenerator cold end allows one to minimize the average mass flow rate. The volume of the regenerator in the present experiment, determined using REGEN 3.2 for the 1 kW of room temperature acoustic power, is approximately From equation (2), and with a first-order approximation of equal and opposite phase angles, the acoustic power at 15 K can be estimated as 50 watts, which combined with an estimated phase angle of –35° determines a cold end mass flow rate of 24.5 g/s. The possibility of achieving the desired phase angle must be verified through analyses associated with the complex impedance of the inertance tube. Using analogous relations from transmission line theory it can be shown that the impedance introduced by the inertance tube and reservoir are given by
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where is the load impedance determined from the compliance of the reservoir, is the diameter-dependent characteristic impedance of the inertance tube, x is the length of the inertance tube, and is a propagation function that is dependent on the complex resistance/length, impedance/length, and compliance/length of the inertance tube. By matching the acoustic power flow where the inertance tube meets the regenerator, it is possible to define a required length and associated phase shift as a function of inertance tube diameter. For the desired phase shift of –35°, the analysis defines a diameter and length of 2.1 mm and 105 mm respectively.
NEXT STEPS Construction and assembly of the components shown in figure 1 is in progress. Measurements of the regenerator losses will be gathered following assembly and instrumentation calibrations. A subsequent re-assembly with an appropriately sized pulse tube will then allow the final set of pulse tube loss measurements to be gathered.
SUMMARY An experiment has been design to characterize regenerator and pulse tube losses in the range of 10 – 20 K, and 30 – 60 Hz. In the initial phase of development, the experiment has been designed to measure conduction, radiation, regenerator ineffectiveness, and acoustic power losses. The associated test rig incorporates a linear compressor providing 1 kW of room temperature acoustic power at an average pressure of 2.5 MPa, a water-cooled aftercooler and a 2-stage GM cryocooler to control warm, intermediate, and cold end temperatures, a laminar flow meter to measure the warm-end mass flow, and a combined inertance tube – reservoir to measure the mass flow at the cold end and optimize its phase angle with respect to the pressure oscillation.
REFERENCES 1. R. Radebaugh and B. Louie, “A Simple, First Step to the Optimization of Regenerator Geometry,” Proceedings of the Third Cryocooler Conference, NBS Special Publication 698, (1984), pp. 177-199. 2. J. Gary, R. Radebaugh, A. O’Gallagher, and W. Rawlins, “An Improved Model for the Calculation of Regenerator Performance (REGEN3.1),” in Proceedings of the Fourth Interagency Meeting on Cryocoolers, David Taylor Research Center Technical Report DTRC91/003 January 1991, pp. 165-176.
3. R. Radebaugh, A. O’Gallagher, and J. Gary, “Regenerator Behavior at 4 K: Effect of Volume and Porosity,” Advances in Cryogenic Engineering vol. 47, American Institute of Physics, Melville, NY (2002), pp. 961-968. 4. A. Lang, H.U. Håfner, and C. Heiden, “Systematic Investigation of Regenerators for 4.2 K Refrigerators,” Advances in Cryogenic Engineering, vol. 43, Plenum Press, New York (1998), pp. 1573-1580. 5. K.A. Gschneidner, Jr., A.O. Pecharsky, and V.K. Pecharsky, “Ductile, High Heat Capacity, Magnetic Regenerator Alloys for the 10 to 80 K Temperature Range,” Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2001), pp. 433-441. 6. R. Radebaugh and V. Arp, “Isothermal Heat Exchanger – version 3,” NIST – Boulder, (1992). 7. T. Kuriyama, F. Kuriyama, M. Lewis, and R. Radebaugh, “Measurement of Heat Conductions through Stacked Screens,” Cryocoolers 9, Plenum, New York (1997), pp. 459-464.
8. M.A. Lewis and R. Radebaugh, “Measurement of Heat Conduction through Metal Spheres,” Cryocoolers 11, Kluwer Academic/Plenum Publishers, NY (2001), pp. 419-425.
Regenerator Characterization under Oscillating Flow and Pulsating Pressure S. Jeong, K. Nam and J. Jung Cryogenic Engineering Laboratory Department of Mechanical Engineering Korea Advanced Institute of Science and Technology Daejon, Korea
ABSTRACT An experimental apparatus has been developed to investigate the thermal and hydrodynamic characteristics of regenerators. Instantaneous gas temperature, pulsating pressure and mass flow rate have been measured at both ends of the regenerator. First, pressure drop characteristics were investigated by comparing numerical simulations with experimental data. When the frequency was 4.6 Hz, the time variation of the pressure drop was predicted quite well by calculations based on conventional steady-flow friction factors. At 60 Hz operation, the amplitude of pressure drop was underestimated when the mass flow rate became high. Unlike the amplitude of the pressure drop, the phase angle was poorly predicted by the steady-flow friction factor at both 4.6 Hz and 60 Hz. Second, the ineffectiveness of the regenerator was obtained from the experimental data for various cold-end temperatures. For low cold-end temperatures the ineffectiveness of the regenerator increased. The main reason for this increase appears to be a decrease in the interstitial heat transfer coefficient caused by the reduced thermal conductivity of the gas at low temperatures.
INTRODUCTION The regenerator is a key component for good performance of regenerative cryocoolers because it must efficiently cool and heat of the gas as it alternately flows between the compressor and the cryogenic expansion space. To carry out this function effectively, the ineffectiveness of the regenerator must be minimized. In addition, the pressure drop across the regenerator must be minimized to achieve maximum P-V work in the expansion space. During the last several decades, regenerators for cryocooler have been extensively studied both theoretically and experimentally. Numerical simulation of regenerators involves empirical coefficients such as friction factors and interstitial heat transfer coefficients for the porous medium. Several authors have investigated friction factors under oscillating flow with no pressure pulsation.1,2,3 They selected the maximum value of the pressure drop and the flow rate in their experimental data and presented the correlations for the maximum friction factor for low frequencies (< 9 Hz). In recent papers, research on regenerator friction factors has been extended to oscillating flow and pulsating pressure, which simulates the real operating conditions of cryocoolers.4,5,6 It was found that the maximum friction factor of oscillating flow at high frequency can be significantly different from the friction factor determined for steady flow. Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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The single blow test has been a popular method for determining the ineffectiveness of regenerators.7 However, this method cannot be applied to the measurement of ineffectiveness in cyclic operation because the heat transfer coefficient is likely to be different, and the pressure oscillation may affect the regenerator performance. Recently, regenerator ineffectiveness was directly measured in oscillating flow and pulsating pressure.6,8 In this paper, pressure drop and thermal performance of the regenerator was measured in a carefully designed apparatus to simulate the actual operation of the cryocooler; the results are then compared with numerical simulation results. The applicability of the conventional steady friction factor is thoroughly examined by comparing both amplitude and phase data for the pressure drop with calculation results. The ineffectiveness of the regenerator was experimentally obtained with different cold-end temperatures. The physical cause of the cold-end temperature dependence of the ineffectiveness is discussed and compared with simulation results.
EXPERIMENTAL SETUP Fig. 1 shows the experimental apparatus used to measure regenerator performance under oscillating flow and pulsating pressure. Instantaneous gas temperature, pulsating pressure, and mass flow rate were measured at both ends of the regenerator. From these experiments, the hydrodynamic and thermal behavior of the regenerator can be fully characterized. Details of the experimental setup were described in the author’s previous work.6
NUMERICAL SIMULATION A numerical simulation model of the regenerator was developed to provide theoretical calculations to compare with the experimentally measured pressure drop and ineffectiveness. The onedimensional governing equation set of conservation laws is summarized as follows.
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where, is the free flow area, A is the total heat transfer area of the regenerator matrix, is the hydraulic diameter of the regenerator matrix, is the porosity, is the interstitial heat transfer coefficient between the regenerator matrix and the gas, is the Fanning friction factor, k is the thermal conductivity of the gas, is the effective thermal conductivity of the regenerator matrix, and is the regenerator length. In the numerical simulation we used the correlations for and based on the correlations given by Kays and London9 for steady flow through stacked screens. The effective thermal conductivity was assumed to be 10% of the regenerator matrix material conductivity.10 The second term in Eq. (2) is the axial variation of gas momentum and is usually negligible in a porous medium.11 To simplify the calculation procedure, we assumed that the gas temperature in Eq. (1) and Eq. (2) is constant with time and varies linearly through the regenerator. This assumption was validated, as the calculated temperature from Eq. (3) and Eq. (4) had little effect on the calculated pressure and mass flow rate. Therefore, the calculation of mass flow rate and pressure was decoupled from the two energy equations. The calculation procedure is described as follows. First, Eq. (1) and Eq. (2) were discretized in the axial direction of the regenerator. Eq. (1) and Eq. (2) were solved for the given cyclic boundary conditions at the warm-end of the regenerator. Then, the two energy equations were solved subsequently. The finite difference method proposed by Patankar12 was utilized in the numerical simulation.
PRESSURE DROP CHARACTERISTICS Dynamic pressure and mass flow rate were measured at the ends of the regenerator at room temperature. The tested regenerator was composed of # 200 stainless steel wire-screens stacked in a stainless steel tube. The inner diameter and length of the regenerator were 9.1 mm and 70 mm, respectively. The amplitude of the mass flow rate was controlled to investigate the effect of the hydraulic Reynolds number on the pressure drop characteristics at low (4.6 Hz) and high (60 Hz) frequencies. Figures 2 to 5 show the experimental measurements of mass flow rate and pressure together with the simulation results. In these figures, the subscripts 1 and 2 represent the warm-end and the cold-end, respectively. A positive sign of the mass flow rate means that the flow is from the warm end to the cold end, and vice versa. The uncertainty of the mass flow rate measurement at the warm end and cold end was estimated to be about 7 % and 2 %, respectively. To validate the steady friction factor under oscillating flow and pulsating pressure, experimental measurement results were compared with the simulation results obtained by Eq. (1) and (2). Filtered experimental data at the warm end were inserted as the boundary conditions, and the mass flow rate and the pressure at the cold end were calculated in the numerical simulation. First, consider the experimental results for 4.6 Hz as shown in Figs. 2 and 3. As the amplitude of the mass flow rate was increased, the predicted pressure drop became slightly different from the
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measured data. Note that there exists a phase difference as well as an amplitude discrepancy in the pressure drop comparison. This deviation was more severe for the 60 Hz case as shown in Figs. 4 and 5. As the amplitude of the mass flow rate or the Reynolds number increased, the pressure drop difference and phase shift between the calculation and the experiment were increased. Prediction of the mass flow rate at the cold end was not satisfactory as the oscillating frequency increased from 4.6 Hz to 60 Hz. The friction factor obtained from steady state flow experiments may not be adequate for predicting pressure drop of oscillating flow and pulsating pressure. Not only the amplitude, but also the phase of the pressure drop in the simulation needs to be corrected. A fundamentally different form of the friction factor is required for oscillating flow and pulsating pressure to account for the pressure drop in the regenerator at high frequency.
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INEFFECTIVENESS FOR DIFFERENT COLD-END TEMPERATURES The ineffectiveness of the regenerator is usually defined by the following equation.13
where h is the enthalpy of the gas, and the time interval from 0 to is the duration of the warm flow period, and from to represents the cold flow period. stands for the enthalpy flow loss for one cycle at the cold end of the regenerator. means the maximum possible amount of heat transfer in the regenerator. If the instantaneous mass flow and gas temperature are measured, the ineffectiveness of the regenerator can be directly determined from Eq. (5). The tested regenerator was # 200 stainless steel wire screens stacked in a stainless steel tube of thickness 0.25 mm. The inner diameter and length of the regenerator were 18.6 mm and 130 mm, respectively. The cold-end temperature of the regenerator was cooled by a heat exchanger in a liquid nitrogen bath, The cold-end temperature was changed by adjusting the amount of liquid nitrogen around the heat exchanger. To confirm the cyclic steady state, the cold-end temperature was changed very slowly by natural evaporation of liquid nitrogen around the heat exchanger. The warm-end of the regenerator was maintained at about 290 K, and the operating frequency was fixed at 4.6 Hz. The amplitude of the mass flow rate at the warm end was maintained near 1.5 g/s. Fig. 6 shows that the ineffectiveness increased as the cold-end temperature was decreased. Here, the coldend temperature represents the average temperature of the oscillating gas at the cold end. The maximum uncertainty of the ineffectiveness calculation from the measured quantities was estimated to be about 12 %. In general, the NTU (Number of Transfer Units), the heat capacity ratio of the matrix to the gas, and the axial conduction through the matrix are the main parameters that have an influence on the ineffectiveness of the regenerator.7 The NTU is a non-dimensional variable proportional to the interstitial heat transfer coefficient of the regenerator matrix. Conceptually, as the gas temperature becomes lower, the interstitial heat transfer coefficient drops due to the lowering thermal conductivity of the gas. The heat capacity of the regenerator matrix is also reduced in the lower temperature region, and it is easy to understand that the axial conduction through the matrix is increased when the temperature difference across the regenerator is large. To clarify the effect of the cold-end temperature on the ineffectiveness, numerical analyses were performed for different conditions. Experimental data such as the pressure, the mass flow rate at the warm end, and the in-flow gas temperature at both ends, were applied as the boundary conditions of the simulation. Table 1 shows the four conditions of the simulation to investigate the quantitative impact of the cold-end temperature on the ineffectiveness. Simulation 1 included all the temperature variation effects that were discussed in the above paragraph. In simulation 2, the heat capacity of the matrix was fixed at its 300 K value; with stainless steel as the regenerator matrix material, the specific heat at 77 K is about 40 % of the value at 300 K. The axial conduction through the regenerator matrix was neglected in simulation 3. Finally, the gas thermal conductivity in the interstitial heat transfer coefficient was fixed at its 300 K value in simulation 4. All of the above simulation results are plotted in Fig. 6.
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From simulation 1, the predicted ineffectiveness was about 75 % of the experimental value. The results of simulations 2, 3, and 4 represent three temperature-variation effects on the ineffectiveness calculation. As can be seen in Fig. 6, the calculated ineffectiveness from simulation 4 was nearly constant for the different cold-end temperatures. This means that the decrease of gas thermal conductivity has a strong influence on the ineffectiveness of the regenerator. The thermal conductivity of helium at 77 K is about 45 % of that at 300 K. The interstitial heat transfer coefficient is, therefore, reduced and the heat exchange performance is noticeably degraded in the low temperature region. From this simulation study, it appears that the reduced interstitial heat transfer coefficient due to the lower gas thermal conductivity at low temperatures is the main reason for degraded regenerator thermal performance at low temperatures.
CONCLUSION Instantaneous gas temperature, pulsating pressure, and mass flow rate were measured at both ends of the regenerator under oscillating flow and pulsating pressure. When the frequency was 4.6 Hz, the amplitude of the pressure drop was predicted quite well by the numerical simulation with the conventional steady flow friction factor. For 60 Hz operation, the amplitude was underestimated at high Reynolds numbers. The amplitude of the calculated pressure drop could be equal to the experimental one by adjusting the magnitude of the steady flow friction factor in the simulation, but the phase angle mismatch was not improved by this magnitude correction of the friction factor. A new friction factor correlation for oscillating flow and pulsating pressure is required to account for the realistic pressure drop in a regenerator at high frequency. The ineffectiveness of the regenerator increased as the cold-end temperature decreased. From the numerical study, it was apparent that the main reason for this low temperature effect was a drop in the interstitial heat transfer coefficient due to the reduced thermal conductivity of the gas at low temperatures.
ACKNOWLEDGMENT The authors are grateful to the Ministry of Science and Technology (MOST) and the Agency for Defense Development (ADD), Korea for the support of this work through the Dual Use Technology project (grant no. 99-DU-04-A-02). The authors also acknowledge the support of LG Electronics Corp.
REFERENCES 1.
Hsu, C. T., Fu, H. and Cheng P., “On Pressure-Velocity Correlation of Steady and Oscillating Flows in Regenerators Made of Wire Screen,” Trans. ASME J. of Fluids Engineering, vol. 121, issue 1 (1999), pp. 52-56.
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2.
Tanaka, M., Yamashita, I. and Chisaka, F., “Flow and Heat Transfer Characteristics of the Stirling Engine Regenerator in an Oscillating Flow,” JSME International J. Series II, 33 (1990), pp. 283-289.
3.
Zhao, T. S. and Cheng, P., “Oscillatory Pressure Drops through a Woven-Screen Packed Column Subjected to a Cyclic Flow,” Cryogenics, vol. 36 (1996), pp. 333-341.
4.
Helvensteijn, B. P. M., Kashani, A., Spivak, A. L., Roach, P. R., Lee , J. M. and Kittel, P., “Pressure Drop over Regenerators in Oscillating Flow,” Advances in Cryogenic Engineering, 43B (1998), Plenum Press, New York, pp. 1619-1626.
5.
Ju, Y., Jiang, Y. and Zhou, Y., “Experimental Study of the Oscillating Flow Characteristics for a Regenerator in a Pulse Tube Cryocooler,” Cryogenics, vol. 38, no. 6 (1998), pp. 649-656.
6.
Nam, K. and Jeong, S., “Experimental Study on the Regenerator under Actual Operating Conditions,” Adv. in Cryogenic Engineering, Vol. 47, Amer. Institute of Physics, Melville, NY (2002).
7.
Ackermann, R. A., Cryogenic Regenerative Heat Exchangers, Plenum Press, New York (1997).
8.
Rawlins, W., Timmerhaus, K. D., Radebaugh, R. and Daney, D. E., “Measurement of the Performance of a Spiral Wound Polyimide Regenerator in a Pulse Tube Refrigerator,” Advances in Cryogenic Engineering, 37 (1992), Plenum Press, New York, pp. 947-953.
9.
Kays, W. M. and London, K. L., Compact Heat Exchangers, 2nd ed., McGraw-Hill Book Company, New York (1964).
10. Kuriyama, T., Kuriyama, F., Lewis, M. and Radebaugh, R., “Measurement of Heat Conduction through Stacked Screens,” Cryocoolers 9, Plenum Press, New York (1997), pp. 459-464. nd 11. Kaviany, M., Principles of Heat Transfer in Porous Media, 2 ed., Springer-Verlag, New York (1995).
12. Patankar, S. V., Numerical Heat Transfer and Fluid Flow, Hemisphere Publishing Corp (1980). 13. Radebaugh, R., Linenberger, D. and Voth, R. O., “Methods for the Measurement of Regenerator Ineffectiveness,” NBS Special Publication, 607 (1981), pp. 70-81.
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Oscillating Flow Characteristics of a Regenerator under Low Temperature Conditions K. Yuan, L. Wang, Y.K. Hou, Y. Zhou, J.T. Liang, Y.L. Ju Cryogenic Laboratory Technical Institute of Physics and Chemistry Chinese Academy of Sciences, Beijing, P.R. China
ABSTRACT An experimental system was designed and constructed to investigate the oscillating flow characteristics of a regenerator under low temperatures. Experimental data on the pressure drops of a regenerator subject to oscillating flow under low-temperature conditions have been obtained. It is found that the value of the cycle-averaged pressure drop of the oscillating flow in the regenerator under liquid nitrogen temperatures is 5~6 times higher than that of a steady flow at the same Reynolds number based on the cross-sectional mean velocity. A correlation equation for the friction factor at liquid nitrogen temperatures has been obtained, and a comparison is made with that obtained at ambient temperatures. Because the test conditions are very close to the operating conditions of practical high frequency cryocoolers, the experimental data should be useful for the prediction of performance and design of cryogenic regenerators.
INTRODUCTION The ability to accurately predict pressure drops in a regenerator subject to an oscillating flow is of crucial importance in the optimum design of cryocoolers. In the past, the regenerator was normally taken to be a kind of high efficiency heat exchanger with properties close to those measured for unidirectional steady flow. The correlation equation of the friction factor given by Kays and London1 has been widely used. There are also some correlations with pressure drop in regenerators2, 3, but most of them are based on a unidirectional steady flow through packed screens. Considering the fact that most regenerative cryocoolers operate under periodically reversing flow conditions, it is evident that the correlation equations based on the steady flow cannot accurately predict the pressure drons and phase shifts in the regenerator under oscillating flow. Zhao and Cheng4 and Helvensteijin et al.5 found a higher friction factor in the regenerator under oscillating flow than that under steady flow at the same Reynolds number based on the cross-sectional mean velocity. Since 1996, we have been working on an experimental study of the flow resistance characteristics of regenerators under oscillating flow conditions. We previously reported detailed experimental data of the pressure drops, and predicted that the oscillating flow characteristics of the regenerator demonstrate not only pressure drops, but also phase shifts6-8. However, all of the above work was focused on the flow characteristics of the regenerator under ambient temperature conditions.
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The regenerator in a practical cryocooler operates under low temperature conditions, and there exists a large temperature gradient along the regenerator length. Therefore, it is necessary to study the oscillating flow characteristics of the regenerator under low temperatures. In this paper, we will focus on the effects of low temperatures on the oscillating flow characteristics of the regenerator.
EXPERIMENTAL SYSTEM A hot wire anemometer was used in the experiments for the velocity measurements. However, the hot wire anemometer must be calibrated firstly when used under low temperatures. A calibration system was set up to calibrate the hot wire anemometer under liquid nitrogen temperatures. Fig.l shows the calibration data, which is fitted by:
here, E is the voltage output of the hot wire and U is the velocity of the working fluid. The schematic diagram of the experimental apparatus is given in Fig. 2. The test section of the regenerator consists of a thin wall stainless steel tube (48 mm in length and 9 mm inner diameter) packed with stainless steel screens. A linear compressor is connected to one end of the test section by the velocity straightener (21 mm in length and 8 mm inner diameter) to provide desirable oscillating flow. The swept volume of the compressor is about 2cc, and its operating frequency can be adjusted from 20 to 80 Hz. A reservoir with an adjustable needle orifice valve is connected to the other end of the test section through a U-shaped tube. To maintain the low temperature, the test section is immersed into a tank filled with liquid nitrogen. Three small quartz differential pressure transducers (KISTLER, Type 601A), connected to a charge amplifier (KISTLER, Type 5011) having a high natural frequency (150 kHz), are used to measure the transient gas pressure wave. They are placed at the outlet of the velocity straightener, at the inlet of the U-shaped tube and at the reservoir, as shown in Fig. 1. A hot wire anemometer (DANTEC, Model 90N10) is used to measure the instantaneous cross-sectional mean velocity. Two small hot wire probes (DANTEC, Model 55P11) are placed at both ends of the test section of the regenerator, details of the principle of the hot wire anemometer was given in Ref. [7]. Analog-to-digital conversions are carried out by an A/D conversion board (KEITHLEY, DAS 1610), which is plugged into a 486 personal computer. A 4-channel simultaneous sample and hold front ends are employed to ensure that both the dynamic pressures and the velocity voltage signals are sampled simultaneously. An oscilloscope (HP 5402B) is also employed to simultaneously observe the pressure wave signals.
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EXPERIMENTAL CONDITIONS In the present experiment, the stainless steel screens with three different mesh sizes are tested under various working conditions. The properties of the number of screens n, wire diameters mesh distance porosity and hydraulic diameter for the three mesh sizes of the wire screens are listed in Table 1. Among of the parameters, the wire diameter the pitch and the mesh distance are provided by the manufacturer, respectively. The hydraulic diameter and the porosity of the regenerator are determined from the equations given below. The working medium is helium gas. The operating frequency is 50 Hz and the system mean pressure varies from 0.5 to 2.0 MPa. The raw measured experimental data include: (1) the pressures at the inlet and outlet of the regenerator, and at the reservoir, (2) the temperatures at the inlet and outlet of the regenerator, and at the reservoir. Due to operating difficulties of using the hot wire anemometer in the liquid nitrogen temperatures, the velocity of the working gas at the outlet of the regenerator is obtained by measuring the instantaneous pressure oscillation at the reservoir. Assuming the relationship between the gas mass of helium and the pressure in the reservoir is given by 9, 10 .
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here, P is the pressure oscillation in the reservoir, and are the steady and oscillating part of the pressure in the reservoir, respectively. Both the pressure ratio and the oscillation amplitude of the temperature are so small in the reservoir during the experiments that the temperature in the reservoir can be approximated as the environmental temperature T. Thus the mass flux into the reservoir can be rewritten as:
here, is the local sound speed. With outlet of the regenerator can be simply expressed as:
the velocity of the helium at the
The averaged relative deviation between the velocity obtained by the hot wire anemometer and Eq. (4) is about 3.35%.
RESULTS AND DISCUSSIONS Pressure drop Performance predictions of the cryocoolers involve estimating pressure drops to determine regenerator efficiency under oscillating flow condition. Therefore, it is important to know the correlation to calculate pressure drops and pressure drop factor. We have obtained the experimental data for the pressure drops across the regenerator subjected to oscillating flow under different experimental conditions. To compare the pressure drops over the regenerator in oscillating flow with those in steady flow, we use the following correlation equation for predicting the friction factor of a steady flow through a stack of wire screen1:
where
where and is the pressure drop and cross-sectional fluid velocity in the packed column under steady flow; n is the number of screens; is the distance between meshes, and is the Reynolds number based on and . We use the cycle-averaged velocity instead of and equations (5), (6) and (7) to predict the steady flow pressure drops. To prediet the pressure drop across the regenerator under the oscillating flow, the maximum friction factor is defined as follows6:
and the cycle-averaged friction factor is
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Defining the Reynolds number:
where is the maximum pressure drop in one cycle, is the cycle-averaged pressure drop, is the maximum cross-sectional mean fluid velocity in the regenerator, L is the length of the regenerator, and is the hydraulic diameter of the screen, which is defined as follows6
with
being the screen porosity which is defined as:
We also define the dimensionless distance as:
Based on the experimental data, the cycle-averaged friction factor is evaluated according to Eq. (9). Fig. 3 represents the experimental data of oscillating flow of 50 Hz, which is obtained in the liquid nitrogen temperatures. It is shown that the experimental data are well fitted by the following correlation equation:
This equation can be used to predict the 50 Hz oscillating flow pressure drop in the liquid nitrogen temperatures. Table 2 is the comparison of the pressure drops of steady flow and oscillating flow with operating frequency of 50 Hz at ambient temperatures and liquid nitrogen temperatures. It shows that the ratio of to increases with decreasing temperature. It was of 2~3 at ambient temperatures 7, while in present work, we find that the ratio is of 5~6 at the liquid nitrogen temperatures. We conclude that the lower the temperature is, the larger the ratio of to is. The effect of the temperature on the pressure drop of the oscillating flow is useful to evaluate and design the regenerator.
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Friction factor In this section, we compare the cycle–averaged friction factor under the liquid nitrogen temperatures with which under the ambient temperatures. Fig.4 illustrates the comparison results with different mesh sizes. We can see clearly that the friction factors have similar tendency under different temperature range with the Reynolds number in the range
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of 5-100. They decrease with increasing Reynolds number and reach a constant value ultimately. The cycle-averaged friction factor under the liquid nitrogen temperatures is less than that at ambient temperatures. It means that the oscillating flow resistance in low temperatures is smaller than that in ambient temperatures. Fig. 5 demonstrates the influence of the mesh size on the cycle-averaged friction factor . At ambient temperatures7, the cycle-averaged friction factor with 250 mesh size has the largest value. However, at liquid nitrogen temperatures, the cycle-averaged friction factor increases with increasing mesh size. Under the condition of ambient temperatures and small Reynolds number, the effect of mesh size on the friction factor is smaller, while the effect increases with increasing Reyholds number. However, the effect of mesh size on the friction factor under the liquid nitrogen temperatures decreases with increasing Reynolds number. Therefore, the friction factor is nearly independent of the mesh size under the condition of large Reynolds number and liquid nitrogen temperatures.
CONCLUSIONS Experimental measurements on the oscillating flow characteristic of the regenerator under low temperatures have been carried out in this paper. The experimental data have been compared with that under ambient temperatures. Correlation equations for the maximum and cycleaveraged friction factors in terms of Reynolds number and dimensionless distance X are obtained. It is found that the value of the cycle-averaged pressure drop of the oscillating flow in the regenerator under the liquid nitrogen temperatures is 5~6 times higher than that of a steady flow at the same Reynolds number based on the cross-sectional mean velocity. The friction factor decreases with decreasing temperature.
ACKNOWLEDGMENT This work is supported by the National Natural Science Foundation of China (Grant No. 50176052).
REFERENCES 1.
Kays WM, London AL., Compact Heat Exchanger, 2nd edition, McGraw-Hill, New York (1964).
2.
Tong LS, London AL., “Heat Transfer and Flow Friction Characteristics of Woven-screen and Cross-rod Matrices”, Trans. ASME (1957), pp.1558-1570.
3.
Walker G, Vasishta V., “Heat Transfer and Friction Characteristics of Wire-screen Stirling Engine Regenerator”, Adv. Cry. Eng., vol.16 (1971), pp. 324-332.
4.
Zhao TS, Cheng P., “Oscillatory Pressure Drops Through a Woven-screen Packed Column Subjecked to a Cyclic Flow”, Cryogenics, vol. 36 (1996), pp. 333-341.
5.
Helvensteijin BPM, Kashani A, Spivak AL et al., “Pressure Drop Over Regenerators in Oscillating Flow”, Adv Cry Eng vol.43 (1998), pp.1619-1626.
6.
Ju YL, Jiang Y, Zhou Y., “Experimental Studies of the Oscillating Flow Characteristics for a Regenerator in a Pulse Tube Cryocooler”, Cryogenics, vol. 38 (1998), pp. 649-656.
7.
Jiang Y, Ju YL, Zhou Y., “A Study of the Oscillating Flow Characteristics of the Regenerators in a High Frequency Pulse Tube Refrigerator”, Adv. Cry. Eng., vol.43 (1998), pp.1635-1641.
8.
Ju YL, Zhou Y, He GQ., “Investigation on the Phase Shift of the Oscillating Flow Regenerator for the Pulse Tube Cryocooler with High Frequency”, Proceedings of ICEC 17 (1998), pp. 93-96.
9.
Nishitani T, Nakano K, Maruno Y, Yanai M, “Investigation of Acoustic Streaming and Steady Flow in the Orifice of Single Stage Pulse Tube Refrigerator”, JSJS-5, Osaka, Japan, pp.66.
10. Lu GQ, Cheng P., “Flow Characteristics of a Metering Valve in a Pulse Tube Refrigerator”, Cryogenics vol.40 (2000), pp. 721-727.
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A Comparative Evaluation of Numerical Models for Cryocooler Regenerators J.P. Harvey1, P.V. Desai1, and C.S. Kirkconnell2 Georgia Institute of Technology1 Atlanta, GA 30332-0405 Raytheon Electronic Systems2 El Segundo, CA 90245-0902
ABSTRACT Various models for predicting the flow and heat transfer in a porous cryocooler regenerator have been proposed in the literature. One such model utilizes a semi-implicit set of equations after making some simplifying assumptions, resulting in a momentum equation that is decoupled from the energy and continuity equations. This work addresses concerns with a semi-implicit model based on scale analysis. An important result is that the pressure gradient term in the energy equation, which has been neglected in the semi-implicit model, is leading order.
INTRODUCTION The generalized governing equations for 1-D flow and heat transfer in a porous medium may be derived using volume averaging of the conservation equations to give
where is the gas density, u is the gas velocity, P is the gas pressure, is the gas viscosity, T is the gas temperature, e is the gas specific internal energy, h is the gas specific enthalpy, is the gas thermal conductivity, and is the matrix temperature. The Darcy permeability, K, with units of the Forchheimer inertia coefficient, which is dimensionless, the porosity, and the volumetric interfacial convection coefficient, with units of are four quantities which characterize the pressure drop and heat transfer in the porous media. The surface area per Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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unit volume, has units of and is frequently referred to in terms of the hydraulic diameter, such that Pressure, internal energy, and enthalpy can be eliminated from these equations through the use of the exact thermodynamic relations
and
where is the volume expansivity, is the isothermal compressibility, and volume specific heat, to result in a simplified set of equations as
is the constant
In these equations, the Gruneisen parameter (1) and the Reynolds number based on the Darcy permeability length scale are given as
respectively. These equations in terms of density, velocity and temperature of the gas and matrix provide a common starting point for a variety of modeling efforts in the literature. Solutions range from control volume analysis, linearized solutions for small amplitude fluctuations, to fully nonlinear numerical solutions. It should be noted that in the present form, these equations represent one-dimensional flow of a real gas in a porous medium. It is assumed that the kinetic energy is small in comparison to the internal energy and that the viscous dissipation is negligible.
FULLY IMPLICIT MODEL Equations (8)-(11) define the fully implicit math model. These equations represent a system of parabolic-hyperbolic partial differential equations. The scale analysis that follows will justify the reduction of these equations to hyperbolic, requiring less boundary data. The continuity equation represents a balance between three terms (when the spatial gradient is expanded) to give
where is the velocity scale, is the length scale, and is the time scale which has been chosen as the inverse of the angular frequency, Scales with tildes indicate temporal scales
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while overbars distinguish spatial scales. Assuming a balance between the second and third terms leads to a definition for the spatial velocity scale given as
This in turn requires that
The ideal gas law can be used to relate the density and temperature scales. To proceed, the gas law is expanded to give
Scaling this result produces three scales given as
which requires that
and similarly
These relations developed from the continuity and ideal gas law will be useful in scaling the momentum and energy equation. Proceeding to the momentum equation (equation 2), scaling leads to a possible balance between four terms with scales given as
where the “ref” subscript on the Reynolds number indicates that this is a reference Reynolds number based on the velocity scale. The leading order balance is clearly between the third and fourth terms which can be used to define a suitable pressure scale as
The scaled energy equation represents a balance between six terms given as
where the Peclet number is given as
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is the specific heat ratio, is the gas dynamic temperature scale, is the gas temperature gradient scale, is the scale of the temperature difference, is the hydraulic diameter, and the Stanton number is given as
In these developments, the reference velocity scale,
is defined as
It should be noted that the scale of the volume expansivity was assumed to be the inverse of the temperature scale. The interesting development which can be noted here is that the first four terms are of order one since the Gruneisen parameter is order one. These terms resulted from scaling the accumulation and advection terms. In many analyses, the pressure gradient term (represented here in terms of a density and temperature gradient) is eliminated from the energy equation based on scaling assumptions, but it is clear from this analysis that it is leading order. The temperature difference scale can be evaluated by assuming that
which leads to
This result makes intuitive sense. As the Stanton number increases, the temperature difference between the gas and matrix decreases. A similar scaling of the matrix energy equation gives a balance between three terms as
where
and is the dynamic matrix temperature scale. The appearance of the gas Peclet number in the matrix energy equation is a bit of an artifact, but this leads to a more simplified scale. Using the definition for the temperature difference scale, the matrix energy equation scales can be written
as
Assuming a balance between the first and third terms leads to a scale for the dynamic matrix temperature as
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To close the scale analysis, the spatial temperature scale can be defined intuitively as the difference between the hot and cold end reference temperatures as
where and are the hot and cold end reference temperatures, respectively. Additionally, the reference temperature scale can be defined as the average of the end temperatures as
The density scale can be defined as the reference density at the temperature scale and the charge pressure, i.e.
where is the charge pressure and R is the ideal gas constant. Finally, the length scale, which has yet to be defined, is most accurately described as a penetration length scale. As the gas flows into the regenerator, the particles only travel a fraction of the overall length before reversing. An estimate of this penetration length can be calculated based on the swept volume of the hot piston as
where is the hot piston stroke, is the hot piston diameter, and is the regenerator diameter. The test apparatus that is being used to correlate this model can be used to evaluate these scales. A representative set of data follows in Table 1 for the specific regenerators being tested. The order one terms in the momentum equation represent macroscopic acceleration effects. For flows in the regenerator, the pressure gradient is several orders of magnitude larger than one. It should also be noted that the scale of the conduction term is small for large Peclet number, which in this case is O(100). Thus the conduction terms appear to be negligible. The cyclic regenerator loss can be corrected by calculating the conduction rates in post processing. This results in a system of hyperbolic PDEs with the matrix energy equation a zeroth-order PDE, The limiting solution of negligible acceleration in the momentum equation leads to an algebraic equation for the velocity in terms of the temperature and density gradients. When the velocity is eliminated using this relation, the continuity and energy equations become second order. Therefore, it seems reasonable to expect boundary conditions for temperature and density at both ends. Equivalently, the pressure can be imposed without directly specifying the temperature or density. To correctly calculate the pressure such that mass is conserved, additional modeling is needed for the end conditions. Conservation of mass and energy written for the compression and expansion volumes results in an additional set of four equations which are coupled to the mass flow rates at the ends of the regenerator. To solve these equations, the Numerical Method of Lines is being used. This numerical scheme converts the system of four partial differential equations into a large system of ordinary differential equations (ODEs) by discretizing the spatial derivatives. The writers have found that second order central finite differences are optimally suited. The system of ODEs are integrated in time until a periodically stable solution is obtained. The key data from the model is the cycle averaged enthalpy flow rate, which should be constant along the length of the regenerator. This takes several hundred cycles typically. The model has been implemented using Matlab, but is being converted to Fortran to improve computational efficiency. Currently, it is not feasible to run the model until periodic steady state due to time limitations. As a result, the matrix energy
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REGENERATOR PERFORMANCE ANALYSES AND TESTS
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(and the gas energy) are not steady. The resulting enthalpy flow is seen to decrease monotonically from the warm end to the cold end.
REGEN Gary et al. propose a semi-implicit model, known as REGEN, by assuming that the pressure gradient is negligible, which results in the continuity equation being transformed into a second order ODE for velocity (2, 3). This problem is closed by imposing velocity boundary conditions. The effect of the pressure drop is re-introduced as a correction. The main difference in the REGEN model and the fully implicit model proposed here is the manner in which the velocity is calculated, In the fully implicit model the velocity is calculated primarily from the pressure gradient. In the REGEN model, the velocity is calculated to satisfy the continuity equation since the pressure gradient is assumed to be small. The pressure gradient is also neglected in the energy equation. In effect, the fully implicit model is driven by pressure while REGEN is driven by mass flow. The REGEN model allows the user to select the matrix geometry which directs the program to use a predefined set of correlations for the friction factor and heat transfer coefficient, which for screens are taken from Kays and London (4). It has become obvious that the friction factor correlations are inadequate for 325 and 400 mesh screens based on direct measurement. Some authors have attributed this to the effect of flow oscillations, but steady flow measurements differ quite significantly, making it difficult to assess the effect of flow oscillations. There is also some ambiguity in how the flow oscillation conclusions have been made. From preliminary calculations using the fully implicit model using the measured friction factors, we find that the pressure gradient field is not uniform, making it necessary to evaluate the friction factor locally rather than on a control volume sense. This is being currently investigated. The authors believe that the friction factors can accurately be represented without any corrections for oscillations if the model is implemented correctly. The heat transfer coefficient has not been measured directly, but several researchers have reported on measurements for some regenerator materials. Initially, it is assumed that the Kays and London correlations for heat transfer coefficient would also be inaccurate. One issue that has been noticed with REGEN is a non-uniform cycle averaged enthalpy flow rate, which is expected to be the correct result. Simulations for typical operating conditions for the apparatus being used result in an enthalpy flow rate that increases by approximately 20 percent at the ends after 10,000 cycles. The implicit model has only been simulated for 100 cycles due to the computational inefficiency, so comparisons are not possible at this time.
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FUTURE WORK The near-term work plans are to perform one-to-one calculations with these two models. These calculations will be based on an experimental apparatus which has been used to measure the comparative performance of a variety of different regenerator matrices (5). The model being developed will be correlated using the regenerators described, in addition to other possible designs and materials.
CONCLUSIONS These two problems represent approximate solutions for the fully nonlinear governing equations. The REGEN model has gained wide acceptance, but there have been little published data evaluating its accuracy. From the scale analysis, it appears that the pressure gradient is important in the energy equation and should not be neglected. Additionally, the model is somewhat restrictive on the type of boundary conditions that can be used, requiring sinusoidal mass flow rates. More realistic boundary conditions can be imposed with relative simplicity by modeling the warm and cold end volumes with piston/cylinders of prescribed motion. These boundary conditions lead to near-sinusoidal mass flow waves, but with the added benefit of the pressure wave being modeled directly. In modeling the regenerator, the user should evaluate the accuracy of the friction factor and heat transfer correlations for their specific application. It is expected that future tests will examine the effect of flow oscillations on friction factor with the anticipation that they can be modeled directly from steady flow measurements if implemented correctly. At the time of this writing, the implementation of the fully implicit model is still in progress. ACKNOWLEDGMENT The work described herein was performed with funds provided by Raytheon Electronic Systems to the Georgia Tech Foundation.
REFERENCES 1.
Gary, O’Gallagher, Radebaugh, and Marquardt, REGEN 3.2: User Manual (Draft), NIST, Boulder, Colorado (2000).
2.
Gary, Daney, and Radebaugh, “A computational model for a regenerator,” Proceedings of the Third Cryocooler Conference, NIST Special Publication 698, (1985).
3.
Gary and Radebaugh, “An improved model for the calculation of regenerator performance (REGEN3.1),” Proceedings of the Fourth Interagency Meeting on Cryocoolers, David Taylor Research Center Technical Report DTRC91/003 (1991).
4.
Kays and London, Compact Heat Exchangers, McGraw-Hill, Inc, New York (1964).
5.
Harvey, Kirkconnell, and Desai, “Regenerator performance evaluation in a pulse tube cryocooler,” Advances in Cryogenic Engineering, Vol. 45, Kluwer Academic/Plenum Press, New York (2000), pp. 373-382.
Periodic Porous Media Flows in Regenerators T. Roberts US Air Force Research Laboratory-Space Vehicles Directorate Albuquerque, NM, USA 87122 P. Desai G. .W. Woodruff School of Mechanical Eng., Georgia Institute of Technology Atlanta, GA, USA 30332 ABSTRACT Flow through porous media can be modeled and measured by focussing on the way that it differs from flow in an unobstructed regime. If this flow occurs within a highly transient periodic regime, the resulting flow begins to exhibit a dependency on the imposed transient pressure gradient and mass flow through the media. A model is offered to express this time dependency and the implications of this dependency on the resulting system of partial differential equations that represent periodic porous media flow in regenerators. Taking note of the resulting hyperbolic nature of this system and the boundary-condition-imposed temporal discontinuities in the entrance regions of the regenerator, a solution by the Method of Characteristics is proposed and the resulting characteristic lines and their compatibility equations are derived.
INTRODUCTION Transport processes in regenerators can be modelled on both an overall, macroscopic level as has been done by Organ1, or Radebaugh, et al.2, as well as on a pore size or local level as is common in the various Computational Fluid Dynamics methodologies. Reconciliation of the results of these two approaches has posed a serious theoretical, experimental, and modeling challenge. In an attempt at such a reconciliation, Local Volume Averaging of the microscale fluid dynamic and transfer processes seeks to portray the small local effects as a phenomenon dependent on the overall flow and heat transfer regime and the physical nature of the porous media. The intent of this paper is to show how the general equations for mass, momentum, and energy balance for a generally isotropic medium experiencing transient compressible gas flows that varies from the temperature of the solid matrix can be depicted using the Method of Characteristics.
ANALYSIS METHODOLOGY The continuum to be studied is assumed to be comprised of phases and with being the gas phase and being a solid phase. The conventions and assumptions behind the volume averaging method are as presented in Gray and O’Neill 3 and are depicted in Fig. 1 using the nomenclature of Table 1. Subscripts are tensor indices or their derivatives, while superscripts refer to phases or coefficient labels. Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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Using the above nomenclature, the general mass, momentum, and energy balance equations applicable to this flow are
PERIODIC POROUS MEDIA FLOWS IN REGENERATORS
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After application of the volume averaging methods outlined in Gray and O’Neill and eliminating terms that are zero for isotropic media with axially oriented entrance flows, these balances can be restated as
which can be normalized to determine their relative magnitudes and simplified using scaling arguments. (Note that all properties from this point forward are normalized):
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The applicable equation of state for a helium working fluid above 30 K would be the Ideal Gas Law, with its relationship of acoustic speed
or by using the general approach of Shapiro4 the momentum equation and the energy equation can be restated as
Using the definition of the total derivative for density, velocity, and pressure, the general system of equations can be stated in matrix form as
This particular system is hyperbolic in nature, and therefore its solution is solely dependent on the initial data. This allows for any discontinuity in that data to be propagated along the Characteristic Lines of the solution. Consequently, these lines are unique in propagating property slopes, which can be undefined, or equal to 0/0 due to discontinuities in any initial data for any relevant period of time. Presumably, in realistic thermal systems such discontinuities would be rare or nonexistent. But this mathematical system can then be used to find where using Cramer’s Rule (see Abbott 5 for theoretical implications of hy-
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perbolic system propagation of initial data discontinuities, and Shapiro 4 for detailed practical application to fluid flows).
For the bottom determinant to be equal to zero one of three conditions must be met, namely
These lines, illustrated in Fig. 2, are the characteristic lines of the solution and have a particular significance as they refer to where a solution point’s relevant initial data originates from for any timestep. The lines are called Mach lines as they are related to how pressure waves propagate through the fluid, while the line is called the pathline as it expresses how the massflow propagates through the media. Figure 2 shows how these lines are related to the solution of the partial differential equation system over a time step For any discontinuity to be possible in the solution space, the denominator must also be equal to 0, which gives the Compatibility Equations (Riemann differential equations) that hold along those Characteristic Lines. Along the Mach Lines:
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Using the same system denominator to find the pathline compatibility equation gives no useful information along the pathline, but the system gives along the pathline
This set of equations can be used in a predictor-corrector method using Newton’s method in four variables: and which solves a discretized form of the algebraic-partial differential equation system. It should be noted that while a rough solution to the system could be obtained using only the characteristic lines and compatibility equations, this system is quite nonlinear and approximation of actual experimental data will require a fully convergent corrector step that reflects the changes to the state properties along the characteristic lines over any timestep, The implications of this mathematical model on describing the flow through regenerators are twofold: 1) The flows through the regenerator follow that of the d’Alembert solution to the wave equation, on a first order approximation. The consolidation of the thermal and viscous diffusive terms in the momentum equation qualify this approximation, and show the flow being closer to a nonlinear propagation problem. 2) The effect of the microscale transfer phenomena on the macroscale flow and transfer properties are incorporated directly into the overall flow equations; this allows the transient effects of variations in enthalpy, massflow, and pressure through the regenerator to be expressed. If the system of equations were to be linearized, in the mode of thermoacoustic models, these complex-plane effects would be expressed as complex impedances following a Fourier transform of the equations. The thrust of these modeling measures have been loosely asserted as observations in the prior literature. Practically, the overall validity of this approach can be seen in Figs. 3-6, which show how pressure propagates through an operating regenerator (400 mesh stainless steel, 50.2 mm long by 14.28 mm diam). As expected, the pressure wave is attenuated, but it is also shifted in phase by a time lag much greater than would be expected by the speed of acoustic propagation through the working fluid. Moreover, this shift’s magnitude appears to be directly related to the frequency of operation. The measurement of this effect, not only on pressure propagation but also on massflow and enthalpy flow, is the subject of ongoing research today.
CONCLUSIONS By considering the additional terms affecting momentum diffusion and heat transfer within porous media, the ability of a macroscale model to depict porous media oscillatory flows has been enhanced. This assertion leads towards the theoretical unity of such flows with the field of nonlinear acoustics in which complex acoustic impedances are an integral theoretical component. That such impedances should be present in a conjugate momentum balance-heat transfer situation for the energy portion of the overall model can be seen as being theoretically justified, as well as having been noted saliently in various prior experimental studies in this area. As was noted in Gray and O’Neill3, the coefficients characterizing the terms ( and ), which delineate this porous media flow from its single phase theoretical antecedent, can be
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experimentally measured. This measurement requires sufficient experimental data and the model’s incorporation into a numerical model to depict how estimates to the system coefficients alter the modeled flow from flows modeled using only steady state impedance data. Using experimentally measured estimates for these coefficients allows for their effect on the propagation of this predominantly hyperbolic system of equations to be modeled in a non linear implementation of the Method of Characteristics. REFERENCES 1. Organ, A. J. Thermodynamics and Gas Dynamics of the Stirling Cycle Machine, Cambridge University Press, 1992. 2. Radebaugh, R. et al., Measurement and Calculation of Regenerator Ineffectiveness for Temperatures of 5 to 40 K, USAF Wright Laboratory (WL-TR-92-3074), 1992. 3. Gray, W. & O’Neill, K. On the General Equations for Flow in Porous Media, Water Resources Research, April 1976. 4. Shapiro, A.H., The Dynamics and Thermodynamics of Compressible Fluid Flow, Ronald Press Co., NYC, 1954. 5. Abbott, Michael, The Method of Characteristics, Thames and Hudson, London, 1966.
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Initial Operation of the NICMOS Cryocooler on the Hubble Space Telescope W. L. Swift, J. A. McCormick, J. J. Breedlove, F. X. Dolan and H. Sixsmith Creare Incorporated Hanover, NH, USA 03755
ABSTRACT The NICMOS Cooling System (NCS) was installed on the Hubble Space Telescope (HST) in March 2002 during HST Servicing Mission 3B (SM3B). It is designed to cool the detectors for the Near Infrared Camera and Multi-Object Spectrometer (NICMOS) to a temperature of approximately 75 K. The system consists of a single-stage turboBrayton cryocooler connected to a cryogenic circulator loop through a gas to gas heat exchanger at the cold end of the cryocooler. Three high-speed miniature turbomachines with self-acting gas bearings and all-metal heat exchangers are used to provide the necessary refrigeration and heat transport with low input power and negligible vibration. The circulating loop transfers heat from the NICMOS detectors to the cold end of the cryocooler through flexible tubes that connect to the NICMOS dewar. Provisions for contamination control of the circulator loop prior to on-orbit integration with the NICMOS dewar were an important part of the pre-launch operations. Following the integration of the NCS with the NICMOS dewar, the system was subjected to a series of tests to verify its operation and to characterize its performance. A separate series of tests was conducted to assess the impact of NCS operation on spacecraft pointing accuracy. This paper describes the NCS, the pre-launch procedures to minimize contamination of the circulator loop, and the performance of the system on orbit.
INTRODUCTION On March 8, 2002, members of the crew of STS-109, Columbia, installed the NCS on the Hubble Space Telescope. The cooling system replaces the solid nitrogen that had depleted from the NICMOS instrument in late 1998. The three main elements of the cooling system include a cryogenic circulator loop, a closed loop turboBrayton cryocooler, and capillary pumped loop (CPL) that transports heat to a radiator.1 The installation involved the following major elements: Mounting the radiator on the external surface of the telescope, Fastening the cryocooler to the structure inside the telescope aft shroud, Routing and fastening fluid lines between the NICMOS coolant line interface and the cryocooler, Routing and fastening CPL loop fluid lines between the radiator and the warm end of the cryocooler, and Installing the electronics module for control, including the harnessing between the telescope, the control box and the cryocooler module. Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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The work was completed within about eight hours of extra vehicular activity during the flight day seven and eight of the 11-day servicing mission to the telescope. The cryocooler and its underlying technology had been developed at Creare over a period of about 15 years. The flight version of the system, including the cryogenic circulator loop was designed, fabricated and flight tested in October 1998 during an 18-month effort. Only a single system was fully fabricated, though spare components were also produced to assure that developmental issues could be addressed during this aggressive schedule. Three such issues evolved. First, modifications to the design of the turboalternator thrust bearings were implemented to address the broad range of operating conditions that occur during system start-up and cool-down. Limitations in the bearings were encountered and corrected by retrofitting new components during ground tests prior to the initial flight demonstration. Secondly, contamination of the non-hermetic circulator loop during ground tests required hardware and operational modifications to assure protection of the loop. Thirdly, the power electronics driving the compressor were modified following the October 1998 flight demonstration to improve the efficiency of this component. Following the servicing mission, a series of verification tests were performed on the cooling system and the NICMOS instrument to assess its behavior and to establish initial operating conditions. The NCS cool-down started on March 18, 2002. A stable temperature of 70 K was achieved on April 11, 2002. The tests were conducted during cooldown, and extended through several stable temperature set points. They included operations to assure that the cooling system controls were functioning properly, perturbations of set point temperature to verify the response of the cooler and stability of temperature control, and measurements to assess the impact of vibration from the cryocooler on the stability of the pointing accuracy and jitter of the telescope. The initial measurements produced the following results: There was no detectable leakage from the cryogenic circulator loop, There had been no detectable change in the focus or resolution of the NICMOS instrument, A detector temperature of about 77 K would increase the observation efficiency (signal to noise ratio) by about 30%, The vibration from the cryocooler was negligible, and Performance targets for the cooling system had been met. The instrument was returned to service on May 18, 2002. The following sections describe the cryocooler, the steps taken to address contamination of the non-hermetic portion of the system, and results obtained during the initial testing of the cooler after integration with NICMOS.
THE NICMOS COOLING SYSTEM (NCS) Figure 1 is a schematic of the NICMOS Cooling System (NCS). Heat from the NICMOS detectors is transported by conduction from the detector support structure through an aluminum matrix within the dewar to a cooling coil attached to the dewar shell. The tubing had been used to maintain the nitrogen as a solid prior to launch of the instrument. Valved flanges with bayonet fittings at the tube ends were accessible for external cooling loops. The NCS was designed so that flexible lines with matching bayonet fittings could be attached to the NICMOS tube connections during the servicing mission to the telescope. After the connections are made, the circulator loop is charged with neon from a pressurized gas fill bottle contained in the NCS assembly. A centrifugal circulator produces a flow of neon through the loop, exchanging heat with the cryocooler loop at the cold load interface (CLI).
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The closed loop cryocooler consists of a centrifugal compressor supplying a pressure ratio of about 1.6 at a speed of 7,000 rev/s. Neon in the cryocooler is pre-cooled by a recuperator as it flows to and expands through a turbine, producing refrigeration for the cycle. The gas returns to the compressor inlet by first passing through the CLI where it absorbs the heat from the circulator loop, then through the recuperator where it cools the high pressure stream from the compressor. The flow is continuous at about 1.8 g/s and the machines use gas bearings to support the very low mass, high speed rotors. Vibration is barely detectable—during the initial on-orbit vibration tests on the telescope, induced jitter from the operation of the cryocooler was well below acceptance levels. At the warm end of the cryocooler, the cycle rejects heat to a CPL through a heat rejection interface (HRI). The CPL provides some temperature control and modulation between the radiator surface and the heat rejection interface. The telescope has an orbital period of about 90 minutes, and seasonal and orientation variations that produce a range of heat rejection temperatures. In general, the temperature difference between the radiator surface and the HRI is about 20°C. The expected variation in HRI temperature during the year is about 15°C. During cooldown, the compressor is driven at maximum speed. After the control temperature reaches the desired set point, the NCS is controlled using PID algorithms to maintain a cold well temperature within +/- 0.1 °C. This is done by increasing or decreasing compressor speed to maintain the average of the NICMOS circulator loop inlet and outlet temperatures at the control point temperature. Tests during the verification phase established the gradient between these values and the detectors—a constant at about 4.8 K. The difference is caused by the thermal resistance between the remote coil attachment and the detector support.
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The main cryocooler assembly includes the compressor mounted on the heat rejection plate, the recuperator, the CLI that supports the circulator, the turboalternator, and the tubes that attach to the NICMOS dewar. Figure 2 shows the cryocooler assembly during mechanical spin testing. In this figure the compressor is fastened to a support fixture. Figure 3 shows the cryocooler assembly in its frame as it is being instrumented, prior to being wrapped with multi-layer insulation. In addition to the cryocooler components, the framed assembly carried several other items. The power conversion electronics (PCE) were mounted on the same heat rejection plate as the compressor. Six pressure transducers were supported on the frame. They provided cryocooler low pressure and high pressure data, pressure in the circulator loop, differential pressure across the circulator and pressures in the neon fill and refill bottles for the circulator loop. Temperature sensors were located in the cryocooler and circulator loops to provide inputs for control and health monitoring. Manual and solenoid actuated valves were also included in the assembly. The manual valves were used for ground operations including bakeouts. Solenoid valves in the circulator portion of the loop facilitated the initial on-orbit fill of the loop, and permit refills in the event of excessive leakage through the bayonet seals.
CIRCULATOR LOOP CONTAMINATION CONTROL The cryocooler portion of the system was hermetic, and the removal and control of contaminants was straightforward.2 Special procedures and auxiliary hardware were developed to assure that the non-hermetic circulator loop remained free of moisture prior to ground tests and prior to launch. The main steps involved periodic vacuum bakeouts of the circulator loop, supplemented by continuous external purges with dry nitrogen maintained in the vicinity of the bayonet fittings. In addition to ground operations, the valves on the NICMOS bayonet manifold had been opened during the prior servicing mission in late 1999. Thus, the fluid lines in the NICMOS dewar had been exposed to the vacuum of space for nearly 20,000 hours assuring that the amount of moisture remaining in the NICMOS loop was negligible.
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Bakeout operations were performed on the circulator loop prior to each thermal vacuum test. In the bakeout procedure, a continuous flow of dry and heated nitrogen gas was maintained through the circulator loop while the components in the loop were heated to about 70°C. This process continued for several hundred hours while the moisture levels in the inlet flow and in the exhaust flow were measured using dewpoint hygrometers to determine the extent of the bakeout. This procedure proved effective in reducing the adsorbed moisture to an acceptable level for testing. A final bakeout was conducted before the NCS was shipped to Kennedy Space Center for the SM3B launch. This bakeout proceeded slightly differently since only the NCS portion of the circulator loop was involved and the bayonets were inserted into the flight holster assembly. In this final bakeout, a turbomolecular vacuum pump was connected to fittings on the bayonet holster assembly. The circulator loop temperature was elevated to approximately 70°C for 14 days while the circulator loop was evacuated. The effectiveness of the bakeout was monitored continuously by using moisture sensors to measure the partial pressure of water in the gas at the inlet to the vacuum pump. The bayonets were connected to the bayonet holster assembly as shown in Figure 4 from the beginning of the bakeout until the NCS was deployed during SM3B. Each bayonet relies on a pair of silicone O-rings to form a seal with a corresponding bayonet receptacle. The circulator loop was susceptible to contamination during ground operations since water and other atmospheric contaminants can permeate through elastomer O-rings. Special procedures were implemented to minimize contamination between the end of the bakeout and SM3B launch. A dry nitrogen environment was maintained around the exterior of the O-rings from the end of bakeout until SM3B liftoff. In addition, the circulator loop and the small volume between the O-ring seals was evacuated continuously with a turbomolecular vacuum pump from the end of bakeout until four days before liftoff. The circulator loop evacuation process was halted temporarily each time the NCC was transported. When this occurred, the circulator loop was filled to slightly more than one atmosphere with 99.999% pure neon. Isolation valves were closed and the evacuation system was disconnected from the holster assembly. Before the evacuation process was restarted, the evacuation system was re-connected to the holster assembly and the hermetic integrity of the system was verified to below std cc/s with a helium mass spectrometer. A vacuum bakeout of the evacuation system was then performed for 8 hours at 100°C. These steps ensured the complete integrity of the circulator loop during ground operations leading up to launch.
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SERVICE MISSION OBSERVATORY VERIFICATION A series of operations was initiated seven days after the installation of the cooler. Table 1 provides a brief chronology of these operations. The tests were designed to assess the behavior of the cooler as well as assessing the performance of NICMOS. The circulator loop was filled by activating a solenoid valve that connected the main fill bottle with the circulator loop. The charge pressure in the bottle had been pre-set so that when the valve opened, the bottle and loop remained connected, and the charge pressure was at the appropriate value (4.9 atm.). A cold-cathode pressure sensor in the HST aft shroud was used to monitor the vacuum level in this area. The sensor was designed to react to the presence of neon, a notification that the bayonets or the cryovalves might be leaking. This sensor remained active during cooldown to provide constant monitoring as the temperature of the tubing in the circulator loop was reduced from ambient. Heaters were used locally to maintain the valves at acceptable temperatures. No leakage was detected during any of the verification tests.
Cooldown Cooldown was initiated by first starting the circulator, then the compressor. The control temperature (weighted average of NICMOS inlet and outlet temperatures) was set to 70 K, and the control system drove the compressor to maximum cooldown speed—about 7,000 rev/s. During cooldown, the turbine output varied from about 15 W (warm) to about 10 W as the cold end approached 70 K. The rate of cooldown is determined by the heat load applied at the cold end, parasitic heat loads, and the thermal mass of the NCS and the NICMOS dewar assembly. Of these, the thermal mass of the dewar dominated. During the initial stages of cooldown, the NICMOS detectors were on to allow for characterization of the detectors during some phases of the cooldown. The heat load from the on-state was about 0.5 W, and they were turned off for a period of time to help speed the cooldown. The detectors were activated again after the cold end temperature had reached 70 K. Figure 5 shows key temperatures during the cooldown. The NICMOS inlet and outlet temperatures are measured on the outer surfaces of the tubes between the CLI and the bayonet attachments to the NICMOS dewar. The turbine inlet temperature is measured on the turbine housing near the inlet port to the machine. The control temperature is a calculated set point value representing a weighted average between the NICMOS inlet and outlet temperatures. The detector temperature is one of four temperature measurements on the detector support. The vibrations induced by the operation of the cryocooler were monitored during cooldown by analyzing the disturbances to the spacecraft pointing control system. The results of these tests verified that the cooler induced vibration level was well below allowable levels.
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Temperature Stability After reaching a cold end temperature of 70 K, several control tests were performed to assess the ability of the PID control system to maintain the set point temperature, and to evaluate the resolution of the detectors. These tests were conducted by changing the set point temperature, waiting for thermal equilibrium, and monitoring temperatures and system parameters as the HST was exposed to a range of radiation environments. The set point was first raised to 72 K, then increased by 0.5 K, decreased by 1 K, and reset to 72 K. Following these operations, results of the detector performance were assessed by astronomers and a final set point was determined for the first year of observation. The final set point (control) temperature was 72.4 K. This results in a detector temperature of 77.1 K. The difference in temperature between the control set point and the detector temperature is a result of the relatively poor conduction path between the circulator loop and the detector support structure. Figure 6 shows the temperatures of the gas into the NICMOS dewar, out of the dewar, and temperatures at the detector and detector support temperature during these operations. The tests verified that temperatures were held well within the desired range of +/- 0.1 K.
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Following the temperature stability tests, the NCS and the NICMOS instrument were put into normal operation. Table 2 summarizes the operating characteristics of the NCS that were obtained during orbital tests.
Conclusions The NICMOS cooling system was successfully installed in the HST during SM3B in March 2002 and has cooled the NICMOS detectors to the first-year set point temperature. All goals for the cooling system have been met. The NCS represents the first generation of cooling systems for space-borne detectors to broaden our vision of the universe.
ACKNOWLEDGMENT Goddard Space Flight Center, the Hubble Space Telescope Project Office, the Air Force, and the Ballistic Missile Defense Organization have supported the development of technology used in the NICMOS cryocooler. The authors gratefully acknowledge this support. We also acknowledge and thank the many engineers and scientists at NASA, their subcontractors, and at the Space Telescope Science Institute who contributed to the successful installation and operation of this system on the HST.
REFERENCES 1.
Nellis, G.F., Dolan, F.X., McCormick, J.A., Swift, W.L. and Sixsmith, H., “Reverse Brayton Cryocooler for NICMOS,” Cryocoolers 10, Kluwer Academic/Plenum Publishers, New York (1999), pp. 431-438.
2.
Breedlove, J.J., Zagarola, M.V., Nellis, G.F., Gibbon, J.A., Dolan, F.X. and Swift, W.L., “Life and Reliability Characteristics of Turbo-Brayton Coolers,” Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2001), pp. 489-498.
Development of a Turbo-Brayton Cooler for 6 K Space Applications M. V. Zagarola, W. L. Swift, H. Sixsmith, J. A. McCormick and M. G. Izenson Create Incorporated Hanover, NH USA 03755
ABSTRACT Future missions are planned by NASA to investigate the structure and evolution of the universe, and search for the origins of galaxies, stars, planets, and life. These missions include the Terrestrial Planet Finder, Constellation-X, and the Next Generation Space Telescope. A critical need for these missions is the capability to reliably cool detectors and associated components to temperatures around 6 K for space missions of five to ten years. Turbomachinebased, reverse-Brayton cycle (or turbo-Brayton) cryocoolers are ideal candidates to provide this cooling because they are essentially vibration-free, have high reliability, long life, low mass, and high efficiency. Developments are in process at Creare to: (1) reduce the size of the components to attain high efficiency and low mass at the temperatures and power levels needed for these applications, (2) demonstrate the performance of these advanced components, and (3) demonstrate the performance and operational features at the system-level. This paper discusses the current status of these development efforts. In particular, the paper reviews the development of an advanced low temperature expansion turbine; a compact, lightweight recuperative heat exchanger; and compressors for this system. The results of a system-level performance test of a 2-stage turbo-Brayton cryocooler at temperatures at and below 20 K are also presented.
INTRODUCTION The turbo-Brayton cryocooler has been proven as an attractive candidate for providing reliable, vibration-free refrigeration. Recently, a single stage version of the cooler was flight qualified and then installed on the Hubble Space Telescope to cool the Near Infrared Camera and Multi-Object Spectrograph (NICMOS).l The NICMOS Cooling System (NCS) was designed to supply about 7 W of refrigeration at 70 K. It requires just under 400 W of electric input power. Future space applications require lower cryogenic temperatures, and the input power available to several of them will be limited to 100 W or less. The developments that are described below aim to preserve the flight-qualified features of the NCS components, while extending the technology to lower temperatures and power levels.
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The turbo-Brayton cryocooler may be configured several different ways to achieve low temperature operation. The primary performance drivers involve trades between size, mass, power, and complexity. By choosing centrifugal turbomachines in order to maintain low vibration, one is limited to constraints on pressure ratios that can be achieved in these devices. The total system pressure ratio is therefore determined by the number of compressors that are used in series in the cycle. For neon cycles with load temperatures above about 30 K, a single stage centrifugal compressor will produce a pressure ratio of about 1.7. If helium is used (for temperatures below 30 K), the maximum expected pressure ratio in a single stage is under 1.2. Low pressure ratios across the cycle limit the isentropic, and therefore the actual work and temperature change that can be extracted by the turbine(s). Very high effectiveness recuperators are then needed to permit as much of the total refrigeration as possible from the turbines to be available to cool the load. Compressors or compressor stages (on a single shaft) may be arranged in series to increase the total cycle pressure ratio. This will reduce the impact of ineffectiveness in the recuperator, allowing these components to be smaller and lighter. However, additional compression stages increase the complexity of the system, and may increase the input power requirements for a given load. Trades between mass and power limitations, component performance characteristics, and system complexity are key factors in selecting the final cryocooler configuration. Other cycle features, such as the availability of intermediate heat rejection sites, also must be taken into account. An intermediate heat rejection stage will have roughly the same effect as an expansion turbine at the equivalent temperature, without requiring the compressor work associated with the work of expansion. However, for space applications, radiators and heat transport paths are required to provide for this “benefit.” The additional mass and structure to support the radiator must be taken into account when evaluating such options. We are focusing on two low temperature cycle configurations at present to guide component developments. These configurations are shown in Figures 1 and 2. Approximate performance characteristics for these cycles are listed in Table 1. The first is designed to be applicable to a range of future NASA space instruments. It provides cooling at two temperatures, has a primary heat rejection temperature near ambient, and an intermediate temperature rejection site at somewhere between 55 K and 100 K. The second cycle is designed to address higher loads at a different set of temperatures. No intermediate temperature rejection is used.
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COMPONENTS Compressors Centrifugal compressors are being developed to provide higher efficiency than the 400 W compressor used in the NCS. An early brassboard version of the centrifugal compressor was tested to verify the mechanical and aerodynamic performance of the design.2 The compressor incorporated a permanent magnet motor in place of the original induction type motor that was used in the NCS compressor because the permanent magnet design yields higher efficiency at low power levels. The rotor, impeller and motor size were also reduced to provide a better match to the flow requirements of low temperature applications. The compressor assembly is shown in Figure 3. The brassboard version is cooled by water flowing through flow passages in the outer housing. This machine provided initial performance data and verified important mechanical operating characteristics such as maximum speed and start-up behavior. The tests showed that efficiency, pressure ratio, and operating speed targets for low temperature systems could be achieved. Following these tests, advanced versions of this machine were built. The internal mechanical and electromechanical features of the design were maintained. The aerodynamic design was modified somewhat to shift the best efficiency point and the pressure coefficient curve to slightly higher flow rates. The housing was redesigned to replace the water cooling by conductive cooling to a baseplate. An aftercooler was also incorporated in the assembly. Heat from the motor and the compression process is transferred by conduction through the housing and aftercooler to a suitable heat rejection interface through a bolted flange. The compressor assembly, mounted in a system test configuration, is shown in Figure 4.
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This advanced compressor was tested at room temperature with helium to verify its performance. Figure 5 shows the comparative performance of the brassboard and advanced designs. The data are plotted in terms of dimensionless pressure rise (head coefficient) versus dimensionless flow rate (flow coefficient). The head coefficient characteristic has been moved to higher flow rates as desired. Table 2 summarizes physical features of the compressor.
Turbines Two designs are being developed: a Low Specific Speed (LSS), isothermal unit that is similar to the design used in the NCS, and a High Specific Speed (HSS) turboalternator that operates with a temperature gradient within the assembly. The LSS Turboalternator (LSSTA) development focuses on reducing the size of the NCS design by about 33%. This will result in a better match for the flow rates in the low temperature cycles, reducing the leakage penalty associated with the larger design. The LSSTA uses selfacting gas bearings that operate at approximately the same temperature as the cycle fluid expanding through the turbine rotor. The shaft in the new design is about 2 mm diameter, significantly smaller than the NCS TA rotor. New self-acting gas bearings are being developed to assure that the smaller rotor can operate at the required speeds in the low-viscosity gas. The bearings use a proprietary pad geometry that is machined into the bearing, reducing the complexity of the fabrication process. At present, the initial stages of bearing development are complete. A 2 mm diameter turbine rotor has been successfully operated at ambient temperature at a speed just over 12,000 rev/s. The next phases of development involve extending the range of operation to cryogenic temperatures. Table 3 lists characteristics of the LSSTA.
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In addition to the bearing developments for the smaller shaft diameter, the turbine is also reduced in size. This reduction required the development of a new Electric Discharge Machining (EDM) device to produce the necessary precision in the turbine flow channels. Turbines of the required size have been successfully machined. The HSSTA is a derivative of a turboexpander design used in a 5 W, 65 K Single Stage Reverse-Bray ton (SSRB) cooler.3 The earlier turboexpander incorporated a 3 mm diameter turbine and a brake wheel on the same shaft. The brake operated at ambient temperature and absorbed the shaft work produced in the turbine. In the HSSTA, an alternator replaces the brake wheel that had existed in the SSRB design. This reduces system complexity and control issues associated with the brake fluid lines between the warm end of the turbine assembly and the heat rejection site. The alternator and bearings operate at a temperature significantly above the temperature of the fluid expanding through the turbine rotor. This permits higher rotational speeds and higher aerodynamic efficiency than is achieved in the LSSTA. The assembly is designed to minimize thermal conduction from the bearings to the expanding fluid at the turbine. The net loss in refrigeration from an isentropic expansion is a combination of aerodynamic losses, leakage, and heat conduction. For a given set of operating conditions, trades are performed to evaluate the merit of the LSSTA in comparison to the HSSTA design. Rotor and bearing dimensions for the HSSTA design are comparable to the earlier turboexpander. The only technology modification is the addition of the permanent magnet in the hollow shaft of the alternator. A brassboard version of the HSSTA has been successfully demonstrated at the required speeds and temperatures. Measured efficiencies during tests in helium have varied from about 25%–40%. Results are described in a later section. Table 4 lists the important features of this design.
Recuperators Two recuperator designs are being pursued for use in the low temperature systems. The first is a design incorporating slotted plates arranged axially in a stainless steel shell. This design has been flight qualified and was used in the NCS and the SSRB systems. It can be adapted to lower temperature applications without technology development. However, in low temperature space systems, input power is at a premium, and in order to achieve high cycle efficiency, the length of the slotted plate heat exchanger is unattractive. A new design, using primary heat transfer surfaces is being developed.4 This design is lighter and more compact for equivalent performance than the slotted plate configuration.
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The new design, a Radial Flow Heat Exchanger (RFHX), is a primary surface heat exchanger comprising a brazed stack of precision-formed metal plates. Laminar flow between parallel microchannels provides very high heat transfer coefficients with low pressure losses, while the convoluted flow path minimizes thermal penalties due to streamwise heat conduction. The design models for thermal effectiveness and pressure drop have been validated in tests of prototypical heat exchanger stacks. An all-metal developmental model of the RFHX is being fabricated using Hastelloy X, which was chosen for its excellent formability, ideal thermal properties, and high strength. The heat exchanger is designed for operation as a second-stage recuperator with a flow rate of 0.20 g/s of He-4 between 65 and 6 K. This heat exchanger is predicted to achieve a thermal effectiveness of 0.998 with a total fractional pressure drop of only 1.2%. The mass of this RFHX will be less than half the mass of an equivalent slotted plate heat exchanger designed for the same performance. Figure 6 shows a brazed assembly making up one module of an RFHX. Several such modules may be joined to achieve the desired thermal effectiveness.
SYSTEM TESTS System tests have been conducted in a serial manner to assess the performance of individual components as they are added to the closed loop system. Initial testing was performed to assess the operational behavior of the facility and to verify the functioning of controls and instrumentation. The initial tests were followed by a second test series in which the cold end of the system reached thermal equilibrium. Several performance test data points were recorded at this condition. During the first two series of tests, two different compressor designs were used to produce the required flow and pressure ratio. Brassboard versions of the PMMC and an older induction motor compressor (of the type used in the NCS) were plumbed in series. These compressors did not provide optimum conditions for the cycle. However, they demonstrated that there were no operational problems associated with using the two machines in series. They also provided a comparison in the efficiency of the two machines.
Initial Low Temperature Tests Tests were conducted in a low temperature test facility in June 2001. The tests were conducted to verify the proper functioning of the components, instrumentation, and controls in the facility. Two compressors were used to pressurize and circulate helium through the system.
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One compressor was an NCS-type 400 W induction motor design. The second was a brassboard version of the permanent magnet 100 W design. A single HSSTA and a low temperature recuperator were incorporated in the test loop. The temperature of the warm end of the cold stage recuperator and the alternator in the HSSTA was controlled by using liquid nitrogen in a separate flow loop. The test configuration is described in reference 2. Figure 7 is a plot of some of the test parameters during the test. The tests showed that for the same flow rate, the input power to the Permanent Magnet Motor Compressor (PMMC) was about 70% of that to the induction motor, verifying the expected increase in efficiency of the new design. The tests also verified that the HSSTA would operate at the required temperatures. Turbine exit temperatures as low as 5 K–6 K were attained. No steady performance data were collected during this quick cooldown cycle.
Performance Tests A second series of tests were conducted in the same test facility in September 2001. These tests were designed to produce performance data for the compressors, the turbine, and the low temperature recuperator. Results from these tests are shown in Figure 8. The cooldown of the system was accelerated by using a bypass loop that allowed liquid nitrogen and then liquid helium to cool the components at the cold end of the facility. After the cold end had reached about 5 K, the liquid bypass loop was isolated and the cold end was allowed to move to thermal equilibrium. The only refrigeration being supplied during the steady temperature tests was by the HSSTA. Measurements were taken at several test conditions during a period of continuous operation of about 100 hours. During these tests, the thermal load at the cold end was varied up to about 100 mW using an electric heater. System pressure and turbine speed were also varied within a limited band. The performance data were all collected around a narrow range of turbine temperatures of about 20 K. Excessive radiation heat loads through the multi-layer insulation in the chamber did not permit lower temperatures without the use of the liquid helium bypass loop. Following these tests, the system was reconfigured to reduce the radiation heat leak. Perforated multi-layer insulation was used to replace the continuous sheet type that had been originally installed. Subsequent tests were performed to evaluate the modification. These tests showed that a parasitic heat load of about 200 mW had been removed by the modification. The advanced design compressors will be installed in the system for the next sequence of performance tests.
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CONCLUSIONS Progress has been made in the development of components for low temperature turbo-Brayton cryocoolers. Operational and performance targets for a new high efficiency compressor have been demonstrated in a prototype version of the machine. Operational and performance targets have been demonstrated in a new low temperature HSSTA. Substantial progress has been made in the development of a new reduced size LSSTA. Initial system tests have verified that temperature targets can be achieved. Future tests are planned to demonstrate the complete system under closed loop operation.
ACKNOWLEDGMENT Support for these technology developments have been provided by NASA/GSFC and the U.S. Air Force.
REFERENCES 1. Swift, W.L., McCormick, J.A., Breedlove, J.J., Dolan, F.X. and Sixsmith, H., “Initial Operation of the NICMOS Cyrocooler on the Hubble Space Telescope,” Presented at 12th International Cryocooler Conference, Cambridge, Massachusetts Jun 18 2002. 2. Swift, W.L., McCormick, J.A. and Zagarola, M.V., “A Low Temperature Turbo-Brayton Cryocooler for Space Applications,” Advances in Cryogenic Engineering, vol. 47B, (2002), pp. 1061-1068. 3. Swift, W.L., “Single Stage Reverse Brayton Cryocooler: Performance of the Engineering Model,” Cryocoolers 8, Plenum Press, New York (1995) pp. 499–506. 4. Valenzuela, J., “Radial Flow Heat Exchanger,” U. S. Patent 6,170,568 B1, Jan 2001.
A Hybrid Multistage 10K Cryocooler for Space Applications R. Levenduski1, J. Lester1 and E. Marquardt2 1
Redstone Engineering Consulting, Inc. Carbondale, Colorado
2
Ball Aerospace & Technologies Corp. Boulder, Colorado
ABSTRACT Redstone Engineering Consulting Inc. is under contract with the Air Force Research Lab to develop and test a 10K Joule-Thomson cold head and associated thermal storage units.1 These components will be integrated with a Ball Aerospace rotary vane compressor2,3 and a commercial Gifford-McMahon refrigerator to create a complete 10K cryocooler. This hybrid cryocooler will be tested at Ball Aerospace in August 2002. The hybrid cryocooler has the potential of providing cooling at 10K and below with high power efficiency and low mass. The baseline approach combines a multi-stage Stirling cryocooler with a low pressure Joule-Thomson stage. This approach allows each of the two coolers to operate in their efficient regimes. A thermal storage unit is incorporated into the Joule-Thomson cold head to provide variable load cooling. Peak cooling of six times the low load can be achieved by the Joule-Thomson stage with a Stirling cryocooler sized for the average load. This approach allows the Stirling cryocooler to operate at nearly constant load and takes advantage of the Joule-Thomson stage’s ability to change cooling capacity by changing compressor speed.
INTRODUCTION Long-life cryocoolers are needed to cool focal plane arrays to 10 K. The focal planes in some systems will likely be operated intermittently and have an operating load several times greater than the non-operating load. In addition, simultaneous cooling at several higher temperatures will be needed for optical and thermal components that support these focal planes. Multiple cryocoolers are currently needed to meet these cooling requirements because no single 10K cryocooler can simultaneously provide adequate cooling at multiple temperatures. Nor is there a cryocooler currently available that can provide constant temperature cooling at 10K with a significantly varying heating load. Development of a multi-stage 10K cryocooler with these capabilities is an enabling technology for future systems and would greatly enhance the system’s reliability and reduce its size, weight, and power consumption. Review of these requirements led us to conclude that a hybrid cryocooler that incorporates a unique thermal storage unit (TSU) could meet these requirements. The hybrid cooler under development combines a multistage precooler with a 10K Joule-Thomson (J-T) stage. Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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CRYOCOOLER DESCRIPTION The cryocooler under development is a hybrid arrangement of a multistage precooler and a 10K J-T stage. A schematic of the cryocooler is shown in Figure 1. The precooler for this project is a commercially available two-stage Gifford-McMahon (G-M) refrigerator. A thermal standoff is used to create a third stage of cooling to meet the needs of the J-T stage. This will be replaced in a flight cryocooler by a three-stage Stirling, pulse tube or Turbo-Brayton machine. The J-T stage consists of four subassemblies: cold head including TSU, rotary vane compressor, connecting plumbing and electronics. This arrangement has the potential of being highly efficient because it allows the precooler and J-T stage to operate in their most efficient regimes. The cold end of the precooler operates at 15K, which is at the lower end of its efficient operating regime. The J-T stage provides additional cooling to achieve 10K, and below if necessary. The J-T stage operates efficiently at relatively low pressures below 20K with helium as the working fluid. This combination also has other system integration advantages, including configuration flexibility and ease of redundancy. This hybrid cryocooler combines the best features of the precooler and J-T stage to minimize input power and to provide other operational benefits. The TSUs are added to allow the cryocooler to perform efficiently for a highly variable heat load application. The TSUs perform the function of averaging the heat load so that a cryocooler designed for the average load can be used. During the low load part of the duty cycle, the cryocooler has more cooling capacity than needed to cool the sensor. Temperatures in the precooler decrease by several degrees during this time. Thermal storage units attached at two places on the precooler cool down, and, in effect, store a reserve of cooling. During the high heat load period when the heat load is higher than the capacity of the cryocooler, the TSUs warm up and use their stored cooling to supplement the precooler capacity. All during this time the feedback control loop on the J-T stage controls the temperature of the cooled sensor to 10K. The design requirements and goals for the hybrid 10K cryocooler are shown in Table 1.
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Precooler A commercially available two-stage Gifford-McMahon (G-M) refrigerator will be used for this project. It has enough cooling capacity at its first and second stages to meet the precooling requirements of the J-T stage. A third precooling stage is created by using a thermal standoff attached to the second stage. The operating temperatures of the first, second and third precooling stages are 15K, 40K, and 180K, respectively. This precooler will be replaced by a three-stage Stirling, pulse tube or turbo-Brayton cryocooler in the flight configuration.
J-T Stage Compressor. A steady flow compressor with variable speed capability is used to circulate helium in the J-T stage. The compressor is a rotary vane type designed and built by Ball Aerospace. It uses lubricated graphite vanes running against a metal cylinder. The lubrication has a very low vapor pressure to enable long-term, contamination-free operation. A process has been developed to precisely and uniformly fill all the pores in the graphite with lubricant to ensure long life. The vane’s wear rate has been characterized and the vanes are conservatively designed to achieve a 10-year life for the compressor. Oversized filters are placed at the inlet and exhaust ports to compressor to trap any particles before they enter the cold head. This is a 2-stage compressor designed to have an overall pressure ratio of two. The delivered engineering unit does not meet the design goals for the pressure ratio and mass flow. However, it produces adequate pressure and flow to assess the cryocooler’s overall performance. The compressor was delivered with its own drive electronics. Cold Head. The cold head contains four recuperative heat exchangers, four interfaces operating at different temperatures, two TSUs, a J-T valve, two thermal shrouds and multilayer insulation. All of these components are packaged within a vacuum housing as shown in Figure 2. Parasitic heat leaks must be minimized because they tend to increase the power requirement of the cryocooler. This leads to the requirement for compactness in the cold head design to minimize radiating surface areas. The cold head design is compact so as to anticipate a highly efficient, next generation flight cryocooler design. The internal support structure has been minimized for this version to ease assembly. Internal supporting schemes are well known and design specific, so it was reasonable to forgo them in this project. The two warmest heat exchangers are critical to the overall performance of the cold head. They must be highly effective and compact. These heat exchangers were developed by NIST and delivered to Redstone for assembly into the cold head. These two exchangers are a NIST parallel plate design. The parallel plate exchangers use photo etching and diffusion bonding to create uniform flow channels for high effectiveness. They also incorporate a mass balancing technique to further improve performance. The first exchangers built of this type, which were delivered to Redstone, have a pressure drop that is higher than expected on the high-pressure side
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of the exchanger. The pressure drop is expected to be in the range of 8 psi at operating temperatures. It is possible to reduce this pressure drop down to the range of 1 to 2 psi in future heat exchangers. The delivered heat exchangers are expected to be over 99% effective. The performance requirements for the two coldest heat exchangers are not as stringent as the two warm heat exchangers. Consequently, a simple tube-in-tube design is used to achieve the required performance. Redstone designed and built these heat exchangers and the remainder of the cold head. Thermal Storage Units. The cold head incorporates two TSUs. One operates at 40K and the other operates at 15K. The 15K and 40K TSUs are quite different in construction. This is because the thermodynamic processes employed and the thermal capacities of available TSU materials are quite different at these two temperatures. The 40K TSU is a simple copper plate mounted on the 40K stage of the precooler. Adding this TSU takes advantage of the fact that there is also excess cooling capacity in the precooler at 40K during times of low load. This TSU also provides a store of reserve cooling for high loads but its purpose is different from the 15K TSU. It allows simplification and size reduction of the two warmest recuperative heat exchangers in the J-T loop. These two heat exchangers must be highly effective to prevent serious loss of efficiency in the cryocooler. They must be capable of recovering over 99% of the cooling in the outgoing fluid stream. Furthermore without the warmer TSU, these exchangers must be highly effective under all load conditions from low to high load. This constraint results in large and complex heat exchangers. When the 40K TSU is added, the heat exchangers can be sized for the low load conditions The 15K TSU is a unique component developed by Redstone. Materials with high thermal capacity are needed to build a lightweight, small and effective TSU. However, the choice of thermal storage materials that operate well at very low temperatures is extremely limited. Solid
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materials such as pure lithium, lithium/aluminum alloys or solid nitrogen work well at higher temperatures, but are not practical for 10K applications because of their low heat capacity. Helium and hydrogen are the only materials with acceptably high heat capacities. The use of charcoal adsorbent to increase heat capacity was determined to be a good enhancement for temperatures below 20K. The 15K TSU is a small pressure vessel filled with charcoal adsorbent. The vessel is connected in the high-pressure side of the J-T stage, which allows gas to flow through it during operation. It works in tandem with the compressor to maintain the sensor interface at 10K in the presence of a changing heat load. During the high load period, heat flows into the TSU as the compressor increases the flow rate in response to the increasing sensor interface temperature. The TSU temperature begins to rise as a result. However, the temperature rise liberates some of the adsorbed helium, thereby creating a cooling effect that counteracts the heating from the increased flow rate. The net result is a slow increase in the TSU and precooler temperatures. This allows the J-T compressor to provide enough cooling in the short term to keep the sensor interface temperature constant at 10K. An additional benefit of having the charcoal in the working gas is that it is a very effective gas purifier. The TSU’s location upstream of the J-T valve is most beneficial because it can capture the contamination before it reaches the J-T valve, which is the most contaminationsensitive component in the cold head. Table 2 shows the primary features and benefits of this TSU concept.
Electronics Laboratory electronics provide power to both the precooler and the J-T stage. Proportionalintegral-derivative (PID) controllers are used to maintain the interfaces at the proper temperatures because the G-M refrigerator produces more cooling than is needed by the J-T stage. The 180K, 40K and 15K interface temperatures are controlled with heaters. The sensor interface temperature is maintained at 10K by controlling the speed of the J-T compressor. Increasing the compressor speed provides more cold helium flow, and thus more cooling, to the sensor interface. Decreasing the speed reduces the flow and cooling. The ability of the compressor to operate over a wide range of pressure and flow is one of the main reasons the rotary vane compressor was chosen for the 10K cryocooler. The power and control functions for the J-T stage are very simple. The compressor has a brushless D.C. motor that can use standard PWM drive technology. A properly matched precooler would eliminate the need for active interface temperature control. It is likely that the J-T stage and precooler electronics will be combined into a single unit for a flight system.
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PERFORMANCE PREDICTIONS A simulation model was created to analyze the performance of the hybrid cryocooler. It contains performance estimates for a three-stage flight cryocooler and modules for each component of the J-T stage. The model was used to conduct a parametric analysis for the J-T stage to determine the best combination of pressure and flow to meet requirements. Then the model was used to determine the required TSU size and maximum cooling capacity achievable for the J-T stage for a given load profile. The load profile used in the analysis was a repeating 90-minute cycle with a high load period of 15% at the beginning of the cycle and a low load of 100 mW for the remainder of the cycle. The results show that a TSU of reasonable size and weight significantly increases the peak cooling capacity of the J-T stage. The results are shown in Figure 3. This assumes that the JT compressor speed can be increased 50% over the nominal operating speed. The expected performance of the cryocooler is substantially less than the model predicts because the compressor does not meet its pressure and flow requirements. However, using commercially available compressors that meet the pressure and flow requirements is being investigated as a means of demonstrating the benefits of the TSUs. Testing will be conducted in August 2002.
SUMMARY Redstone Engineering Consulting Inc. is developing a hybrid 10K Joule-Thomson cold head and a unique thermal storage unit that can significantly increase the peak cooling capacity of the cryocooler for short durations. These components will be integrated with a Ball Aerospace rotary vane compressor and a commercial Gifford-McMahon refrigerator to create a complete 10K cryocooler. Testing will be conducted in August 2, 2002.
ACKNOWLEDGMENT This work is sponsored by the United States Air Force Research Laboratory, Kirtland Air Force Base, under a SBIR Phase II contract. The authors would also like to recognize the support given by NIST and Ball Aerospace and Technologies Corporation in the execution of this project.
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REFERENCES 1.
Lester, J.M., unpublished work discussed at 11th International Cryocooler Conference, June 2000.
2.
Glaister, D.S., Gully, W.J., Wright, G.P., Simmons, D.W. and Tomlinson, B.J., “A 10K Cryocooler for Space Applications,” Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2001), pp. 505-511.
3.
Levenduski, R.C., Gully, W.J. and Lester, J.M., “Hybrid 10K Cryocooler for Space Applications,” Cryocoolers 10, Kluwer Academic/Plenum Publishers, New York (1999), pp. 505-511.
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Development of a Medium-Scale Collins-Type 10 K Cryocooler C.L. Hannon and J. Gerstmann Advanced Mechanical Technology, Inc. Watertown, MA 02472 M. Traum, J.G. Brisson and J.L. Smith Jr. Cryogenic Engineering Laboratory, MIT Cambridge, MA 02139
ABSTRACT The feasibility of a compact, reliable, low-cost, and efficient cryocooler capable of delivering 2W of cooling at 10 K using less than 1 kW of input power has been demonstrated analytically. The technical approach is to apply a high-efficiency thermodynamic cycle to a compact and reliable small-scale system by implementing a modern microprocessor into a mechanically innovative machine. The innovations of the design include “floating” piston expanders and electromagnetic “smart” valves, which eliminate the need for mechanical linkages and reduce the input power, size, and weight of the cryocooler in an affordable modular design. It is predicted that a three-stage cryocooler operating with 15-bar helium could produce 2 W of cooling at 10 K while requiring less than 1 kW of compressor power. A laboratory prototype is currently being designed and built for development testing in the Summer and Fall of 2002.
INTRODUCTION Many current cryocoolers, most notably Stirling and pulse-tube types, have achieved compactness and reliability by adopting mechanically simple cold head configurations at the expense of thermodynamic efficiency. Large multistage terrestrial cryogenic refrigerators are able to achieve higher thermodynamic efficiencies, but do so by employing mechanically complex designs that are not feasible at a small scale. The ideal 10K spacecraft cryocooler would have an efficiency comparable to that of large terrestrial machines, i.e., a power requirement less than 1 kW per Watt of cooling at 10K, with the compactness and reliability of a pulse-tube or Stirling cryocooler. In addition, for applications in space, a cryocooler requires continuous multistage cooling capability with wide load-capacity, good temperature-stability, low vibration, and low EMI. An efficient multistage 10K cryocooler with many of the above attributes is currently under development by AMTI in conjunction with the Cryogenic Engineering Laboratory at MIT. This design achieves compactness and reliability by using modern microelectronics to enable a complex, but inherently efficient, cold head design. The cryocooler is based upon a novel modification of the Collins cycle which is used in many high-efficiency terrestrial cryogenic machines. InnovaCryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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tions of the proposed approach are several-fold, but all serve to reduce the input power, size, and weight of the cryocooler in an affordable modular design. Size, cost, and complexity are reduced in the proposed concept by employing a modular design whereby each stage is of identical construction (except for expander and heat exchanger length), and where the heat exchanger and expander are constructed as an integral unit. The expanders are of extremely simple floating piston construction that requires no seals or mechanical power transmission devices (e.g., connecting rods or crankshafts) to extract power from the cold expander. Piston motion is controlled by electromagnetically actuated “smart” valves that require no mechanical valve linkages or mechanical timing mechanisms. This further reduces system complexity, improves reliability, and eliminates thermal leakage paths. Expander power is dissipated in the warm end of the cryocooler using smart-valved throttles connecting to the compressor suction and discharge, and to an accumulator volume. This cryocooler is expected to be significantly more efficient than existing cryocoolers, so its compressor will be proportionally smaller. The smaller compressor, its smaller power supply, and the reduced size of the necessary cooling equipment will result in the most significant reduction in the overall system weight. Modular construction allows the construction and assembly of each stage as a heat exchangerplus-expander assembly that slides into a cylindrical stage enclosure from the warm end. With this arrangement the complete stage can be withdrawn without disturbing the vacuum system. The stage enclosure is filled with low-pressure helium. The low-pressure stream that pre-cools the next stage leaves the stage through a tubing connection near the bottom of the stage enclosure. The stream passes through a tube connected into the side of the enclosure of the next lower temperature stage. As the stream flows into the enclosure of the next stage, it joins the flow in the low-pressure passage of the heat exchanger of the next stage. With this configuration there is only a single cold tube crossing through the vacuum between stages.
CRYOCOOLER DESIGN Design Paradigm The modified Collins cycle cryocooler currently under development employs a design paradigm that emphasizes the use of an inherently efficient thermodynamic cycle to meet required performance goals, and then optimizes an efficient design for low cost, compactness, light weight, and high reliability. This is opposed to the currently prevalent paradigm that chooses a candidate cycle with low mechanical complexity (and hence low cost) without sufficient regard for the inherent efficiency of the cycle, and then attempts to optimize a design for better performance (higher efficiency). The current design paradigm has been very successful for cryocoolers that operate above 20K. However, it has been less successful for small-scale refrigeration at temperature of the order of 10K and below, due to the very large power required by inefficient refrigerators of the current design for this temperature range. One of the major reasons for the poor efficiency of current designs are the inherent characteristics of the periodic-flow regenerative cycles employed, and the very low heat capacity of regenerator packing at extremely low temperatures. Large-scale machines, such as those used in the production of liquid helium, are generally based on the Brayton or Collins cycles. These cycles may be characterized as employing constantpressure, quasi-steady, recuperative heat exchange between the high and low-pressure gas streams, thereby requiring a two-stream heat exchanger. This compares with the variable pressure, periodic, regenerative heat exchange of the Stirling, pulse-tube, or Gifford-McMahon (G-M) cycles typically used in small-scale machines, which utilize a less expensive single-stream regenerator. The Brayton and Collins cycle machines also require valved or turbo expander/compressors, whereas the simpler regenerative cycles are valveless, or have only warm valves (G-M cycle). The larger steady flow machines utilize pressure ratios on the order of 10 to 1 up to about 15 to 1, whereas the oscillating flow regenerative cycle machines usually utilize pressure ratios of 3 to 1 or lower that require a significantly lower temperature difference for heat transfer. A comparison of the thermodynamic performance of refrigeration at 4K clearly shows the potential for successful cryocooler design using the new paradigm. The large-scale, high-effi-
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ciency machines based on the conventional Brayton and Collins cycles routinely operate with input powers of about 740 watts per watt of refrigeration, which is only 10 times the ideal Carnot power at 4K. Small commercial sub-10K G-M cryocoolers capable of about 1W of cooling at 4K have been available for several years (Daikin, Sumitomo, Leybold). These devices typically require about 5kW - 7.5kW per Watt of cooling at 4K, which is about 70 to 100 times the Carnot requirement.1 The modified Collins cycle under development is particularly well suited for cryocoolers operating from below 4K (liquid helium temperatures) up to about 30K, above which temperature the thermodynamic losses of the mechanically simpler cryocoolers become less dominant.
Thermodynamic Cycle The refrigeration cycle is based upon the Collins cycle, which employs a single stack of heat exchangers in which the compressed gas is first cooled. A portion of the gas is further cooled by expansion, after which the cold expanded stream returns to the heat exchanger stack. There it joins the low-pressure stream returning from the next lower temperature stage. The combined streams cool the remaining portion of the high-pressure stream to near the exit temperature of the expander. The effect is to reduce the temperature difference in the exchanger to a pinch point where the two streams join at the exit temperature of the expander. Tapering the heat exchanger stream-to-stream temperature difference with the pre-cooling expanders allows the heat exchanger surface to be concentrated in the lowest temperature exchangers where it is most effective in improving the performance of the refrigerator. The present cycle configuration differs in that an individual heat exchanger is associated with each stage, rather than a stack of heat exchangers in series supplying high-pressure gas at staged temperatures to each of several expander stages. With this configuration the high-pressure stream is split at room temperature into sub-streams for each stage, as shown in Figure 1. A heat exchanger for each stage pre-cools the high-pressure sub-stream to the inlet temperature for the expander of the stage. Except in the coldest stage, the exhaust stream from the expander is split. The larger stream returns through the low-pressure passage of the heat exchanger, and the smaller stream flows to the next colder stage to enhance the pre-cooling of the next stage. When two pre-cooling stages are employed in this manner, the third 10K stage is very effectively pre-cooled, with the majority of its cooling power available to cool the load and a smaller fraction required for heat exchanger loss. A rendering of the assembled device is shown in Figure 2. The objective in dividing the gas flow among the stages is to approach, to the extent possible in a small number of modules, minimum entropy generation. Minta and Smith have shown2 that the
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optimum performance of a cryocooler with a specified total heat exchanger surface is achieved when each of the several stages has the same temperature ratio. Furthermore, Minta and Smith have shown that the optimum performance is achieved when the temperature differences between the two streams in the heat exchangers are scaled in proportion to the absolute temperature. Furthermore, this objective can be achieved by causing a heat exchanger flow imbalance, equal to The condition for constant in the upper section of the heat exchanger is used to determine the required pre-cooling flow for each stage. Note also, that this requires that the “temperature pinch point” of a heat exchanger should occur at the cold end, which generally requires that the flow rate of the low-pressure (colder) stream be higher than that of the high-pressure (warmer) stream. This ideal is approached, more or less, in the system of Figure 1, except for the heat exchanger of the first stage, which of necessity has the opposite flow ratio necessary for a coldend pinch point. Based on the analysis of Minta2, the for practical heat exchanger sizes will be in the range of 0.05 to 0.1. An optimization of the loss due to pressure drop versus (for a given heat transfer surface) usually gives the total to be about 1/2 of The cycle utilizes a pressure ratio of 15 to 1, which results in a relatively small mass flow rate required for a given refrigeration load, and consequently reduces the size of the heat exchangers. A thermodynamic model has been developed to predict the performance of the modified cycle. The model accounts for finite component efficiencies and process stream temperature differences. Additional assumptions used in the analysis are: (1) The compressor has an isothermal efficiency of 50%, which is consistent with the state-ofthe-art of a two-stage intercooled terrestrial compressor. (2) The expansion process is adiabatic and the expander has an adiabatic efficiency of 75%. This takes into account piston blowby, friction, and valve breathing losses, and is consistent with the performance of similar floating and mechanically-actuated piston expanders measured in the Cryogenics Engineering Laboratory of MIT in the recent past. (3) The system is adiabatic except for conduction of heat from the warm end along the heat exchanger shell and cylinder (which forms the inner shell of the annular heat exchanger), and “shuttle” heat transfer from the warm end as a result of heat transfer between the piston and cylinder walls as the piston shuttles between warm and cold ends. Axial conduction along the coiled heat exchanger tubing is negligible. All of the axial conduction and shuttle heat is intercepted by the low-pressure gas in the heat exchanger that surrounds the cylinder. (4) A 100% factor of safety is applied to the heat transfer correlation3. Other heat losses from the surroundings, such as radiation, heat conduction in the electrical leads to the valve actuators, and electrical dissipation in the power supply, current leads, and actuators are expected to be small and were not included. Pressure losses in the heat exchangers are not explicitly included in the thermodynamic analysis, but the heat exchangers are sized to limit the pressure loss to less than 5% of the stream pressure. The model predicts that a cryocooler capable of producing 2 watts of cooling at 10K would require only 528 W of compressor power, which is only about 9 times the Carnot input. This is comparable to the efficiency achieved by large-scale machines. Even allowing for significant additional losses, it appears reasonable to conclude that the design should be capable of achieving significantly less than 1kW of compressor power per Watt of cooling at 10K. A multi-temperature level cooler has also been analyzed. Only 368W of compressor power was required to provide 0.5W of cooling at 10K, 3W at 30K, and 6W at 77K. This performance is about 10 times the equivalent Carnot input.
Floating Piston Expander The floating piston expander, first studied by Jones4, illustrated conceptually in Figure 3 is a key element in the cryocooler concept. By employing a floating piston, the expander provides a highly effective expansion at a high-pressure ratio without the size, weight and geometric constraints of a mechanical crosshead to stroke the piston and operate the cold valves via pull rods. The expander consists of a piston (or displacer) that floats with the gas in a closed cylinder. At the
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cold-end of the cylinder, high-pressure gas is admitted through the smart electromagnetic inlet valve, and low-pressure gas exits through the smart electromagnetic exhaust valve. A microprocessor controls the opening and closing of the valves to achieve efficient expansion of the gas, and to minimize losses in the valves. The piston floating in the cylinder moves to keep the pressure in the warm-end displacement volume nearly the same as the pressure in the cold-end volume. The analytical design of the three expanders for a 2W case is shown in Table 1. The piston diameter is selected to achieve the required mass flow with an approximately square bore/stroke while keeping the stroke frequency between 1 and 4 Hz. However, there is considerable latitude to apply other sizing criteria that would result in different dimensions and/or frequencies. The configuration for the warm end of the expander is also shown in Figure 3. The warm end has three electromagnetic valves and a single warm gas reservoir. The three valves are flowcontrol throttling valves that control the velocity of the floating piston by controlling the rate of gas flow into and out of the warm cylinder volume. With this configuration, the helium pressure on the warm end of the piston is always essentially the same as the pressure on the cold end. The piston floats quasi-statically on the gas in the cylinder with low velocities set by the throttling of the gas in and out of the warm cylinder volume. The cold valves always open when there is essentially no pressure difference across the valve. This reduces the actuating force required of the cold valve actuators. It is important to differentiate the floating piston from what is commonly known as a “free piston.” Free-piston expanders and compressors operate in a resonant mode in which the kinetic energy of the piston is exchanged periodically with the potential energy of a gas spring. It is difficult to achieve small cylinder clearance volumes with free pistons due to the danger of collision with the cylinder head. Piston displacement and frequency are coupled and only loosely regulated by the working fluid. In comparison, the floating piston displacement is quasi-static. Its motion is well below resonant frequency, and except for brief periods during the intake event, the pressure is balanced on either end of the piston. The piston displacement sensors are able to sequence the valves so as to precisely guide the piston gently to top and bottom dead-center without fear of impact, while independently controlling piston frequency to any desired value within a reasonable range. An important feature of the floating-piston modular cycle configuration is that the floating pistons do not require close-clearance seals. Since the piston is of low mass and floats with the gas in the cylinder, the pressure difference between the warm and cold volumes is limited. In addition, at steady cyclic conditions, the average total mass at the warm end is constant. Any leakage from
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the cold end to the warm end through the piston-to-cylinder gap is balanced by the leakage from the warm end to the cold end. The oscillating flow in the gap does not create significant heat leak because the gap acts as an effective balanced flow regenerator. Also the floating piston remains properly centered in the cylinder since it reaches the zero volume condition at both ends of every stroke as a result of the controlled valve timing.
Smart Electromechanical Valves The modular cryocooler will utilize “smart” electromagnetic valves in the multiple expanders. The design for these valves is based on U.S. Patent 5,211,372, but the specific configuration of these valves has evolved significantly. “Smart” valves of this design have been employed successfully in the Boreas cryocooler. These valves operate at 4.5K and have been shown to have dissipation levels of only a few milliwatts. An improved valve configuration has evolved from the need to fully utilize the available diameter of the expander for active magnetic components, coils and cores. As shown in Figure 4, the new valve arrangement is a concentric design, with the exhaust valve located concentrically outside of the inlet valve. The moving valve element is a flat ring that closes the valve ports in a flat valve plate that is also the cold-end cylinder head of the expander. The valve ring is made of magnetic steel. The valve is opened by a magnetic force that lifts the valve off of the seat. The force is from the magnetic attraction of the valve ring to the magnetic steel yoke that is around the ring coil of conductor carrying the valve actuation current. The valve port area for each valve is evenly divided between multiple ports through the valve plate. This subdivision of the valve port area significantly reduces the required lift for the valve. Having a small lift for the valve ring significantly reduces the current required to generate the force to lift the valve. This design for reduced lift is widely utilized in pressure operated gas compressor valves. The ring disk valve also matches the geometric requirement that the magnetic gap area is significantly larger than the gas port area. The improved configuration for the valves provides the valve sealing force by the pressure difference across the valve. A smaller initial closing and sealing force is provided by a permanent magnetic valve spring.
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A unique feature of the new valve configuration is that the coil and magnetic yoke for the exhaust valve are located inside the cylinder of the expander. This is acceptable since the yoke and coil are filled or potted to be solid with no gas volumes to add to the clearance volume of the cylinder. With the yoke and coil on the cylinder side of the valve plate, the exhaust valve has the gas pressure difference providing the sealing force. The valve opens by having the magnetic force pull the valve ring into the cylinder. The only additional clearance volume is the swept volume of the valve ring motion. Since the cold-end valves do not have to open against a significant pressure difference, the required magnetic force and thus the electrical dissipation associated with opening the cold valves is significantly reduced. The electronic circuit that drives the valve is designed to match the force-position characteristics imposed by the gas pressure and spring force of the valve. The driving circuit delivers a shaped current pulse that provides a high force for a few milliseconds to rapidly lift the valve off the seat and to overcome the gas pressure-difference holding the valve closed. Once gas pressure has equalized across the valve, the driving circuit delivers a much lower level of force to continue the motion of the valve-piston to the wide-open position, and to hold the valve open. The shaping of the current pulse thus provides adequate opening force but avoids high velocity impact at the wideopen position that would result if a constant voltage were applied to the valve. Similarly, a shaped current pulse is applied to decelerate the valve at closing to avoid heavy impact, while permitting a relatively high closing velocity up to the point of contact. The shaped current pulse used to operate the valve is another feature that provides long life and very low electrical dissipation. The function of the smart valve control is not only to provide a desirable valve stroke profile, but also to control the piston displacement and frequency. Consequently, valve actuation will be guided by piston position sensors, and pressure transducers. Hall-effect sensors will be used to determine piston location by measuring the radial magnetic field strength of a permanent magnet located within the piston. Heat Exchangers Two significant features of the heat exchangers in the proposed design are their use of a simple two-passage design, and unidirectional flow in each passage. A two-passage heat exchanger simplifies manifold design while permitting unidirectional (continuous) flow in each passage. Unidirectional flow offers the advantage of significantly lower thermal losses than are typically experienced with oscillating pressure flow in the regenerative (one-passage) heat exchangers typical of G-M and Stirling cycles operating below about 20K, where the regenerator heat capacity is very low. The modular cryocooler will have cross-counter flow heat exchangers in an annular space formed around each expander cylinder. Small diameter finned tubing will be wound around the expander cylinder. High-pressure gas from the compressor will feed into the finned tubes at the warm end and will manifold directly into the inlet valve plenum of the expander at the cold end. The low-pressure passage between the ID tube wall and the cover tube will connect directly to the compressor suction at the warm end. The flow split for the pre-cooling flow is adjusted with lowpressure valves (not shown) at the warm end of each exchanger. At the cold end, the low-pressure passage connects directly to the expander exhaust valve plenum. Pre-cooling gas for the next colder stage will feed directly from the exhaust plenum to the pre-cooling port in the low-pressure passage of the next stage. For the coldest stage the expanded gas will feed from the exhaust plenum to the cold plate of the heat load, and will then return to the low-pressure passage. The results of a heat exchanger sizing analysis for a 2W design are presented in Table 2. The heat exchanger design is matched to and coordinated with the dimensions of the associated expander. The fin spacing and the annular width of the low-pressure passage are matched to give the required NTU and (pressure loss) for the low-pressure side. The tube diameter for the highpressure passage is selected to give the required for the high-pressure side. The length of the heat exchanger annulus is matched to the length of the expander cylinder. The axial location of the pre-cooling port in the outer shell of the heat exchanger annulus is determined to provide the required ratio of NTU’s for the upper and lower portions of the heat exchanger.
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CONCLUSION Successful development of this cryocooler concept will result in a compact, reliable and lowcost cryocooler with efficiency several times better than current technology. This will benefit military and civilian space missions by meeting the need for 10K cooling while reducing the payload burden associated with thermal management. Terrestrial applications will benefit from the size reduction and improved efficiency of this device with respect to current technology of similar capacity and cost.
ACKNOWLEDGMENT This work has been supported by an SBIR Grant from the Missile Defense Agency (formerly BMDO), Contract Number DASG60-01-C-0051 administered by the U.S. Army Space and Missile Defense Command. REFERENCES 1.
L. D’Addario, “Review of Gifford-McMahon Refrigerators for 4K,” Millimeter Array (MMA) Project MMA Memo # 185, http://www.alma.nrao.edu/memos/html-memos/alma185/mma185.txt, September 25, 1997.
2.
Minta, M., A Study of Helium Liquefaction Cycles at Elevated Pressures, MS Thesis, MIT (1981).
3.
J.M. Geist and P.K. Lashmet, “Miniature Joule-Thompson Refrigeration Systems,” Advances in Cryogenic Engineering, Vol. 5, Plenum Press, New York (1960).
4.
R.E. Jones and J.L. Smith Jr., “Design and Testing of Experimental Free-Piston Cryogenic Expander,” Advances in Cryogenic Engineering, Vol. 45, Kluwer Academic/Plenum Publishers, New York, 2000.
Efficiency of the ARC and Mixed Gas Joule Thomson Refrigerators A.Alexeev*, D.Goloubev** and E.Mantwill** * **
Messer Cryotherm GmbH, 57548 Kirchen, Germany TU Dresden, Lehrstuhl für Kryo- und Kältetechnik, 01062 Dresden, Germany
ABSTRACT Several different configurations of mixed gas refrigerators are described in the literature. The mixed gas Joule Thomson cooler (JT, also known as mixed gas throttle cooler) and the Auto-Refrigerating Cascade (ARC, also known as one-flow cascade, mixed refrigerant cascade, or Kleemenko refrigerator) are the most known. The main advantage of the mixed gas JT system is its simplicity; the benefit of the ARC is better oil management, and consequently higher reliability at temperatures below 100 K. The goal of the present work was to obtain a deeper insight into the behavior of the ARC system with a single separator. The efficiency of this system is investigated and compared with the efficiency of the JT-cycle. Calculations for cycles with nitrogen-hydrocarbon mixtures for cooling temperatures in the 80-100 K range are presented and discussed.
INTRODUCTION Two different positions concerning the efficiency of the mixed gas systems are now presented in the available literature: 1.
The group around Mr. Boiarski found that the mixed gas Joule Thomson system is more 1 efficient than the ARC or has at least the same efficiency:
2. The calculation results made by the group from the Chinese Academy of Sciences show 2 that the ARC is better:
Unfortunately, direct measurements are not available at the moment. Therefore, this question was discussed passionately at the CEC conference in Madison in 2001 during the “mixed gas cooler” poster session. The participants could not reach an agreement concerning which cycle has the better efficiency. However, they were sure that the optimal mixture composition of the JT cycle differs from the optimal mixture composition of the ARC, and that the efficiency of an ARC depends on the separator temperature.3 Indeed, very little information concerning the thermodynamics of the ARC cycle is available: Pfotenhauer, et al. reported a calculation method based on a simple thermodynamic model4; another calculation method was contributed by Gong, et al.5 Therefore, the first step in our study was to investigate the ARC system with a single separator.
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INVESTIGATIONS OF THE ARC CYCLE WITH SINGLE SEPARATOR
ARC with Single Separator Figure 1 show the basic flows and heat exchanger arrangements for an ARC system with single separator. In this system a compressor circulates a mixed refrigerant, and an aftercooler cools the high pressure vapor mixture after compression. Next, heat exchanger 1 cools the stream further and partially condenses the mixture at progressively lower temperatures. The phase separator removes the condensate from the vapor stream. A throttling device controls the exiting liquid flow. The expanded liquid mixes with the returning low pressure stream, flows through the heat exchanger 2 and cools the vapor stream from the phase separator. This stream flows to the conventional Joule Thomson stage and produces cold at the required cooling temperature. The mixture composition in the vapor stream after the separator differs from the total mixture composition circulated by compressor.
ARC Simulation Model A model for the simulation of the ARC with single separator has been developed. It is based on the following assumptions:
Steady-state operation Pressure losses in the heat exchangers are neglected Mixture composition does not change in the cycle No maldistribution in heat exchangers No heat leak from ambient
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The input parameters are the following: Mixture composition: n-number of components in given mixture Total flow rate of refrigerant: High and low pressures: HP and LP Ambient temperature (corresponds to the temperature of the warm end of the heat exchanger, high pressure stream): Cooling temperature separator temperature and subcooling temperature Minimal internal temperature difference in the cycle or value for heat exchangers (U = heat transfer coefficient, A = heat exchanger surface) Separator efficiency: The output parameters are the following: Cooling capacity of the cooler: Temperature distribution in the heat exchangers Exergy losses for all elements (heat exchangers, throttle devices, and mixer) The Peng-Robinson equation of state was used for prediction of the thermodynamic properties of mixtures. The software program PROVISION6 was used for the calculation.
CALCULATION RESULTS Numerous calculations were made for cycles with nitrogen-hydrocarbon mixtures for cooling temperatures high pressures = 14-20 bar, separator temperature = 240-295 K, and separator efficiency
Subcooling Temperature First, the subcooling temperature was varied with other parameters fixed. We found that the efficiency of the cycle does not depend on the subcooling if the subcooling temperature is more than 20 K lower than the separator temperature (see Figure 2). This is because the temperature drop during expansion in the throttle device is about 20 K in the given pressure range. If the is smaller than 20 K, the temperature difference between the high and low pressure streams in the heat exchanger 2 is too large and causes additional thermodynamic losses. If the is larger than 20 K, all the cold produced during expansion in the throttle device will be optimally utilized.
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In further calculations, a fixed heat exchanger 2 is optimally sized.
was used. This value meets the requirement that
Separator Temperature As a next step, the influence of the separator temperature was investigated. Various series of calculations was made. Each sequence included calculation of the cooling capacity and temperature distributions for a set of fixed input parameters, thereby allowing the separator temperature to be varied. The simulation of the ARC showed that the behavior in the system initially depends on the mixture composition. All relevant mixtures can be sorted into three groups according to the kind of influence they have: on the separator temperature, on the cooling capacity, or COP, respectively. group I:
does not depend on the separator temperature and
group II,
depends on the separator temperature and
group III,
depends on the separator temperature and
Figure 3 shows the typical behavior of the COP for an ARC system with single separator, and its dependence on the separator temperature. Typical mixture from the group I (pinch point at the warm end of the HE): nitrogen 28 mol methane 24 mol ethane 14 mol propane 10 mol i-butane 20 mol
Typical mixture from the group 11 (pinch point in the middle of the HE): nitrogen 30 mol methane 23 mol ethane 12 mol propane 10 mol i-butane 21.5 mol
Typical mixture from the group III (pinch point at the cold end of the HE): nitrogen 20 mol methane 20 mol ethane 12 mol propane 10 mol i-butane 21.5 mol
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Calculations made with non-ideal separator show the similar behavior, but the difference between the ARC and JT system is less pronounced.
DISCUSSION Concerning the Influence of Mixture Composition on the COP During an earlier investigation of the mixed gas Joule Thomson system7 it was found that all relevant mixtures can be sorted into three groups (group 1, group 2 and group 3) according to the kind of temperature distribution in the heat exchanger: Group 1: mixtures with pinch point at the warm end of the heat exchanger (low fraction of high-boiling components like propane and butane); Group 2: mixtures with pinch point in the middle of the heat exchanger (low fraction of medium boiling components like ethane and methane); Group 3: mixtures with pinch point at the cold end of the heat exchanger (low fraction of lowboiling components like nitrogen and methane). Surprisingly, we realized that our new groups (I, II and III) are identical with the known groups (1,2 and 3), i.e. group I equals the old group 1, group II equals group 2, and finally group III equals group 3. Based on this fact, it is possible to explain the different influence of the separator temperature on the cooling capacity for different mixtures, namely: 1. If the pinch point is located at the warm end of the heat exchanger (mixtures group 1 or I), the cooling capacity of the system is primarily defined through the enthalpy difference at the warm end for both the JT-system as well as the ARC: The mass flow and the composition of the mixture (and consequently enthalpies) at this point for both cycles are identical Therefore the cooling capacity of the ARC is equal to the cooling capacity of the JT cycle. It does not depend on the separator temperature. 2. If the pinch point is located at the cold end of the heat exchanger (mixtures group 3 and III), the cooling capacity is primarily defined through the enthalpy difference at the cold end for the JTsystem as well as for the The enthalpy difference in this case depends mainly on the mass flow of the low-boiling components (like nitrogen, methane and ethane) in the mixture at this point: The mass flows of the low boiling components for both cycles are near equal. The enthalpy differences varies a little. But at the optimal separator temperature It is because the composition of the mixture at this point in the ARC system differs from the composition of the JT cycle—the fractions of the low boiling components are higher (because the high-boiling components like propane and butane are already separated in the separator). Consequently, the partial pressures of these components in the mixture are higher, and it means the higher enthalpy difference. Therefore, the cooling capacity of the ARC in this case is higher than the cooling capacity of the JT cycle. The mixture composition depends to some extent on the separator temperature, therefore the cooling capacity and the COP also depend on the separator temperature.
3. In the case of the “pinch point in the middle of the heat exchanger” (group 2 or II), the cooling capacity is defined primarily through the enthalpy difference at this point And, the enthalpy difference depends on the mass flow of the medium-boiling components (like methane, ethane and propane): The separator temperature is usually very close to the boiling temperature of the propane and ethane, therefore the mass flows of these components in the ARC are less than in the JT cycle. This is the main reason why the cooling capacity of the ARC-system is lower than that of the JT-system. The cooling capacity and the COP also depend on the separator temperature because the mixture composition after the separator depends on the separator temperature.
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4. For the ARC with a non-ideal separator, the differences between the mass flows and mixture compositions for ARC and JT cycles are smaller; therefore, the difference between the COP of the ARC and the JT-systems is less conspicuous.
Concerning the Optimization Methods A further important message is that the separator of the ARC does not cause new pinch points if the separator temperature and subcooling are optimal (heat exchanger 2 and precooler are optimally sized). Consequently, the mixture composition optimized for the JT cycle is also the optimal mixture for the ARC. This message is very important in practice because it is possible to use the relative simple optimization methods developed for the JT cycle for optimization of the ARC, without any modifications. The optimization methods developed specially for ARC, which are essentially more complicated, are no longer necessary.
Concerning the Calculation Results with Fixed U*A-Value The discussion above concerns the calculations made with a fixed minimal internal temperature difference in the cycle. It means that the ARC and the JT systems with disimilar heat exchangers were compared, but the heat exchangers used had similar efficiencies. Practically, it means that the heat exchange surface of the ARC-heat exchangers was smaller than the heat exchanger surface of the heat exchangers of the JT system. Calculations with fixed U*A-values for heat exchangers were also made. This condition is more realistic for many cases and means that the ARC-heat exchangers have heat transfer surfaces similar to those of the JT-system (with similar heat transfer coefficient). The results of these calculations differ a little bit from those discussed above in favor of the ARC (because a larger heat exchange surface means a higher heat exchanger efficiency) and look as follows: group I: does not depend on the separator temperature and group II, group III,
depends on the separator temperature and depends on the separator temperature and
This means that the ARC is always more efficient than the JT-cycle, if you do not use mixtures from Group 2 (II).
CONCLUSION A simulation model of the ARC cycle was developed to allow the investigation of this system. Generally the efficiency of the ARC does not differ from the JT system essentially. The simulation of the ARC system shows that the behavior in the system initially depends on the mixture composition. All relevant mixtures can be sorted into three groups according to the kind of influence the separator temperature has on the cooling capacity. If the mixtures from groups I or II are used, the ARC system is better than the JT system. If the mixtures from group III are used, the JT system is more efficient. The mixture composition optimized for the JT cycle is also the optimal mixture for the ARC cycle. Therefore, it is possible to use the relative simple optimization methods developed for the JT cycle for optimization of the ARC without any modifications.
REFERENCES 1. Boiarski M., Khatri A. and Kovalenko V.N., “Design Optimization of the Throttle-Cycle Cooler with Mixed Refrigerant,” Cryocoolers 10, Plenum Press, New York (1999), pp. 457-465.
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2. Gong M.Q., Luo E.C., Liang J.T., Zhou Y, Wu J.F., “Thermodynamic Analysis of a Mixed-Refrigerant Auto-Cascade J-T Cryocooler with Distributed Heat Loads,” Cryocoolers 11, Kluwer Academic/ Plenum Publishers, New York (2001), pp. 523-530. 3. Conversation with B.Yudin, O.Podtcherniaev, D.Goloubev and M.Gong during CEC-conference in Madison, July 2001. 4. Pfotenhauer J.M., Veka E.C., Klein S. A., Thermodynamic Modeling of a Missimer-Type Refrigerator Using Ideal Solution Relationship, Private communication, 2000. 5. Gong M.Q., Luo E.C., Wu J.F., Zhou Y., “Thermodynamic Design Principle of Mixed-Gases Kleemenko Refrigeraion Cycles,” Adv. in Cryogenic Engineering, Vol. 47A (2002), Amer. Institute of Physics, New York, pp. 873-880. 6. Provision, User Manual, Simulation Science Inc., Brea, California 92621, USA.
7. Alexeev A., Haberstroh Ch., Quack H., Thiel A., “Study of Behavior in the Heat Exchanger of a Mixed Gas JT Cryocooler,” Adv. in Cryogenic Engineering, Vol. 45 A (2000), Plenum Publishing Corp., New York, pp. 307-314.
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Further Development of the Mixture Refrigeration Cycle with a Dephlegmation Separator M.Q. Gong, J.F. Wu, E.C. Luo, Y.F. Qi, Q.G. Hu, and Y. Zhou Technical Institute of Physics and Chemistry Chinese Academy of Sciences Beijing 100080, P.R. China
ABSTRACT Mixed gas refrigeration cycles with phase separators provide a wealth of configuration options for the design of the cycle. This paper focuses on the development of the cycle with a simplified dephlegmation separator. Both theoretical analysis and experimental investigation were carried out to optimize the refrigeration cycle configuration, including the arrangement of the heat exchangers, phase separator, mixer, etc. Particular attention is placed on the position of the mixer. Experimental results confirm the conclusions drawn from the thermodynamic analysis of the configuration of the dephlegmation cycle. A Carnot efficiency of 11.7% was achieved in an experimental test at 125 K with the cycle driven by an air-conditioning compressor with a nominal input power of 1.1 kW. INTRODUCTION There have been numerous configurations of the mixture throttle refrigeration cycle since it was first presented by Podbielniak1. Each configuration of the cycle has its own purpose, either to improve efficiency or to get higher reliability. The mixture refrigeration cycle with phase separators has higher reliability than the single-stage cycle without phase separators, especially in the temperature range lower than 130K.2 With proper design of the phase separators, oil is removed with separated liquids in the relatively higher temperature range, and thereby lubricant management is essentially inherent. The mixture refrigeration cycle with phase separators has much more freedom in the selection of system configuration options. However, this also brings some difficulties in the design of these systems. There are several articles concerning the design of such mixture refrigeration cycles3,4,5,6, A new type mixture refrigeration cycle with a simplified dephlegmator acting as the phase separator was first developed by the authors7 to build a refrigeration system with the consideration of high efficiency, high reliability, and simplicity. Further development of this new cycle is presented in this paper including both theoretical analysis and experimental investigation to optimize the refrigeration cycle configuration.
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DIFFERENT CONFIGURATION DESIGNS In the mixture refrigeration cycle, the arrangement of heat exchangers, phase separators, and mixers is very important to achieve optimal performance. In addition, the separation temperature is also an important parameter for the performance of the system5, which is determined by the position at which the separator is placed. Figure 1 shows two configurations of the separation cycle that incorporate a simplified dephlegmation separator instead of a phase separator based on mixture equilibrium characteristics. In Figure 1, the letter “E” denotes that the separated liquid refrigerant is returned to the low pressure stream close to the separator. The letter “F” denotes that the separated liquid (after passing through the throttle valve) is returned to the low pressure stream at an appropriate position between heat exchangers 4 and 5 in Figure 1. The “E” pattern is the type of configuration developed by the authors7. Good performance was achieved with this configuration. However, detailed measurement of the mixture composition showed that the separation in this refrigeration configuration was not complete, that is, there was liquid partially brought into lower temperature section in the refrigerator. What will happen if there is a complete separation? In the remainder of this paper, both configurations “E” and “F” will be studied theoretically and experimentally. The simulation and optimization were made to determine the influence of the configuration on the thermodynamic performance. In the calculation, the inlet temperature of the dephlegmation separator was also optimized. The simulation results of the cycle with different configurations of patterns “E” (with an abbreviation of DSC-E) and “F” (with an abbreviation of DSC-F) are illustrated in Figures from 2 to 4. The mixtures and other parameters in these simulation cases are listed in Table 1. In each simulation case, a comparison of the performance is illustrated for the singlestage cycle without phase separators (with an abbreviation of NON-SEP), the separation cycle with phase separators based on mixture equilibrium characteristics (Flash at given Temperature and
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Pressure with an abbreviation of FLTP), and the separation cycle with a simplified dephlegmation separator (DSC-E, or F). Figure 2 shows the performance comparison of different cycles and configurations with the calculation conditions of Illustration 1 in Table 1. The abscissa axis of Figure 2 is the separation ratio of the component of isobutane which is the mole fraction (of isobutane) in the separated liquid of the initial composition (of isobutane) in the total mixture. From Figure 2, it can be seen that the performance of the single-stage cycle (NON-SEP), the separation cycle with FLTP separator, and the best performance of the separation cycle with the dephlegmation separator in pattern F (DSC-F) are approximately equal to each other. However, the performance of DSC-F is higher than that of DSC-E in all separation ratio ranges. From the results of Illustration 1 in Figure 2, it is found that the configuration of DSC-F is better than DSC-E. From Figure 1, the cooling capacity of the refrigeration system is determined by the enthalpy difference of the incoming and return streams. For the DSC-E configuration, the cooling capacity is bounded by: where h is specific enthalpy, subscripts A and B denote the state points after the separator in sequence, G is the total flow rate, a (a < 1) is the separation ratio based on flow rate. When the separation is complete in the separator, the state of the high pressure stream coming from the separator is the saturation vapor state (Point A in Figure 1). At this time, the state of Point B of the return stream must also be the vapor state. It is well known that the enthalpy difference is small of gases in vapor state with same temperatures but different pressures, which is within the limit of the pressure difference caused by a single stage compressor. The temperature difference of points A and B (the hot section of heat exchanger 4 in this system) is not large. Then the enthalpy difference of points A and B is small. Therefore from Equation (1), the final cooling capacity of this configuration of DSC-E is low. For DSC-F configuration, the cooling capacity can be calculated with the following equation: In fact, for a certain mixture, the specific enthalpy of the mixture can satisfy the following inequalities: From above equations, it is obvious that the cooling capacity of DSC-F is larger than that of DSC-E. In fact, the configuration of DSC-E increases the temperature difference of the heat exchanger, and also increases the temperature before throttling. Then the total exergy losses increase compared with DSC-F. The difference of DSC-E and DSC-F calculated above is only the return position of the separated liquid (after throttling). Figure 3 shows the results of the DSC-F with a mixture that has an increased fraction of highboiling components. Comparison of the performance of the DSC-F, non-separation cycle, and the separation based on FLTP are illustrated in Figure 3. It is found, for the best performance, that the
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DSC-F is the highest of these three cycles. At this time, the separation ratio of DSC-F is higher than the traditional separation cycle of FLTP. Figure 4 shows another optimization illustration of the DSC-F cycle, as well as the FLTP cycle. It can be found that the best performance of the two cycles is approximately equal to each other. It is also found that in a relatively large separation ratio range the DSC-F cycle maintains a high performance. This is very important for practical applications. For the best performance of the DSC-F, the separation ratio is close to the separation cycle based on FLTP in which the separation temperature is near ambient temperature. Then, for practical application, the heat exchanger before the dephlegmation separator is not necessary. The optimal configuration of the dephlegmation separation cycle is illustrated in Figure 5.
EXPERIMENTAL TESTS An experimental apparatus was developed based on the configuration of the dephlegmation separation cycle illustrated in Figure 5. Several mixtures were studied with this prototype. The experimental results with a mixture obtained from Illustration 3 are listed in Table 2. The performance comparison of the two cycle configurations of DSC-E and DSC-F are also presented in Table 2 with the same mixture. From Table 2, it is found that the performance of DSC-F is larger than that of DSC-E with the same mixture. Because of incomplete separation in the simplified dephlegmation separator, the performance of DSC-E is larger than the results of the theoretical calculations. When the throttling valve of the separated liquid is closed, the cycle is changed to a single stage cycle. However, the separator, which still exists in the system, becomes a depositing volume for refrigerant. This changes the circulating composition and decreases the effective flow rate, then finally degrades the performance.
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The cooldown curve of the experimental test of the DSC-F is shown in Figure 6. The comparison of the theoretical and experimental result is illustrated in Figure 7. In the theoretical calculation, the efficiency of the compressor and the pressure drops are obtained from the experimental test. It can be seen that the results from the calculations and the measurements agree well.
SUMMARY Detailed analyses have been presented on different configurations of mixture refrigeration cycles, especially on the newly developed dephlegmation separation cycle. Two types of configurations of the dephlegmation separation cycle were studied, both theoretically and experimentally. The experimental results verify the conclusions drawn from the thermodynamic analysis of the configuration of the dephlegmation cycle. A Carnot efficiency of 11.7% was achieved at 125 K when the cycle was driven by an air-conditioning compressor with a nominal input power of 1.1 kW.
ACKNOWLEDGMENT This work is financially supported by the National Natural Sciences Foundation of China under the contact number of 50076044.
REFERENCES 1. 2. 3. 4.
Podbielniak, W.J., US patent 1426956, (1936). Missimer, D. J., “Refrigerant conversion of Auto-Refrigerating Cascade (ARC) systems”, Int. J. Refrig. Vol.20, No.3, (1997), pp201-207. Boiarski, M., et al. “Modern trends in designing small-scale throttle-cycle coolers operating with mixed refrigerants,” Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2001), pp.513-533. Podtcherniaev, O., Boiarski, M., Flynn K., “Performance of throttle-cycle coolers operating with mixed refrigerants designed for industrial application in a temperature range 100 to 190 K,” Adv. in Cryogenic Engineering, Vol. 47 (2002), Amer. Institute of Physics, Melville, NY (in press).
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5.
6.
7.
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Gong, M.Q. Luo, E.C. Wu, J.F. and Zhou, Y., “Thermodynamic design principle of mixed-gases Kleemenko refrigeration cycles,” Adv. in Cryogenic Engineering, Vol. 47 (2002), Amer. Institute of Physics, Melville, NY (in press). Luo, B.C., Gong, M.Q., Zhou, Y., Liang, J.T., “Experimental comparison of mixed-refrigerant JouleThomson cryocooler with two types of counter flow heat exchanger,” Cryocoolers 10, Plenum Publishing Corp., New York (1999), pp.481-486. Wu, J.F., Gong, M.Q., et al. “A new type mixture auto-cascade refrigeration cycle with partial condensation and separation reflux exchanger and its preliminary experimental test,” Adv. in Cryogenic Engineering, Vol. 47 (2002), Amer. Institute of Physics, Melville, NY (in press).
Research on Adiabatic Capillary Tube Expansion Devices in Mixed-Refrigerant J-T Cryocoolers Y.F. Qi, Y. Cao*, M.Q. Gong, E.C. Luo, J.F. Wu, and Y. Zhou Technical Institute of Physics and Chemistry Chinese Academy of Sciences, Beijing, China, 100080 *Refrigeration and Cryogenic Engineering Institute Zhejiang University, Hangzhou, China, 310027
ABSTRACT This paper presents a study of capillary tubes used as expansion devices in mixed-refrigerant J-T cryocoolers. A numerical model is presented for predicting capillary tube performance with multicomponent refrigerants, and the results are compared with experimental data. Because of the multicomponent refrigerant used in the system, the flow characteristic is different from that in a household refrigerator or freezer. The experimental results are discussed, and the performance of the capillary is analyzed.
INTRODUCTION A capillary tube is an important element and is commonly used as an expansion device in vapor-compression cycles because of its simplicity and low cost. The capillary tube functions as an automatic controller for the refrigerant flow rate when operating conditions of the refrigeration system vary. It is very important to select an appropriate capillary tube, including the inner diameter and the total length. Many articles found in the open literature have focused their attention on experimental and analytical investigations describing the characteristics of capillary tubes used in domestic refrigeration; this application has a working temperature range from about 243 K to room temperature, with either a pure refrigerant, or an alternative binary or ternary refrigerant mixture. In recent years, there has been a remarkable development of the mixed-refrigerant Joule-Thomson cryocooler, which covers a large temperature range from below liquid nitrogen temperature (77 K) to 230 K with nitrogen and hydrocarbon mixed refrigerants. Fig. 1 shows a schematic of the single stage throttle cycle. For the different working temperature ranges, especially different working fluids, the capillary tube behaves differently. In this paper, an effort is made to dig further into thermodynamic process of the expansion in the capillary tube with multicomponent mixtures as working fluids. Because of the different configurations of the low-temperature-mixture refrigeration cycle, the inlet conditions of the capillary tube are dependent on the heat exchanger that connects the after-cooler and the expansion device, rather than on the condenser in a domestic refrigeration system. The performance of the heat exchanger greatly influences the capillary tube's behavior; thus we have carried out the research on the capillary tube in the context of the complete integrated system. Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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THEORETICAL MODEL AND NUMERICAL SIMULATION In the theoretical model, the flow in the capillary tube is divided into subcooled single-phase liquid and two-phase flow regions. Before the numerical simulation of the performance of capillary tube, there are several assumptions presented as follows: 1. Straight capillary tube with constant inner diameter and roughness. 2. Homogeneous and one-dimensional flow through the capillary tube. 3. Capillary is fully insulated. 4. The metastable flow phenomenon is neglected. 5. The gravity effect is negligible. 6. Refrigerants are free of oil. 7. Thermodynamic equilibrium through the capillary tube. 8. Phase equilibrium through the capillary tube. Based on these assumptions, the refrigerant fluid flow conservation equations according to the control volume in Fig. 2 are described as follows.
Continuity Equation
where the subscripts 1, 2 are the inlet and the outlet of the control volume, respectively, A is the cross-section area of the capillary tube, is the average velocity of the refrigerant, and v is the specific volume of the refrigerant.
Momentum Equation
where P is the pressure at the cross-section, L is the length of the control volume, G is refrigerant mass flux flowing through the capillary tube, d is the inner diameter of the capillary tube, and f is the friction factor which is listed as1
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Energy Equation
where h is the enthalpy of the refrigerant. Equation of State The Peng-Robinson equation of state is used to predict the properties of the mixtures, which is listed as
where R is the general gas constant, a and b are equation parameters which are functions of the critical pressure and critical temperature. For the mixture refrigerant, the mixing rules are
where and are the mole fraction of component i and j respectively, and interaction coefficient between component i and j .
is the
Numerical Simulation Based on the theoretical model, a program2 for predicting capillary tube performance using the multi-component refrigerant is presented. In the simulation, some parameters are already known which are listed as follows: The geometric parameters of the capillary tube. The composition of the refrigerant mixture. The mass flow of the refrigerant. The pressure and temperature at the inlet of the capillary tube. The pressure at the outlet of the capillary tube. The capillary tube is spatially discretized by defining the control volume according to a given pressure drop. The temperature at the outlet of the control volume can be calculated with Eq. (4), and the length of the control volume can be calculated with Eq. (2) and Eq. (3). The calculation continues until the pressure at the outlet of the control volume equals to the pressure at the outlet of the capillary, then the temperature at the outlet of the capillary and the length of the whole capillary can be calculated. The properties of the mixed-refrigerant can be calculated with Eq. (5) and Eq. (6). Table 1 is the simulation results on the 150 K-temperature range cryocooler. In Table 1, the “Time” column refers to the time after the cryocooler is running, “Pin” is the pressure at the inlet of the capillary tube, “Pout” is the pressure at the outlet of the capillary tube, “Tin” is the temperature at the inlet of the capillary tube, “Tout” is the temperature at the outlet of the capillary tube, “T out” is the calculated temperature result at the outlet of the capillary tube, and “Overall Absolute Deviation” is calculated with the Eq. (7).
The numerical results agree reasonably well with the experimental data and the model fairly predicts the capillary tube behavior.
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FLOW CHARACTERISTIC OF REFRIGERANT IN THE CAPILLARY TUBE The process of a refrigerant flow through the capillary tube is a flashing process, in which the refrigerant changes from a subcooled liquid to a vapor-liquid blend. Because of different temperature ranges, especially the different working fluids, the flow characteristic of refrigerant in the capillary tube is different between in household refrigerators and in mixed-refrigerant J-T cryocoolers.
Household Refrigerators and Freezers In household refrigerators and freezers, the refrigerant used is pure such as R-22 or alternative binary refrigerant mixtures such as R-410a. A typical refrigerant R-22 flow process in the capillary tube is shown in Fig. 3. The refrigerant experiences four processes: subcooled liquid region, metastable liquid region, metastable two-phase region, and equilibrium two-phase region. In subcooled liquid region (from point a2 to point b), the temperature of the refrigerant is nearly constant, and its pressure decreases linearly due to the friction. At the intersection point b, the pressure is equal to the saturation pressure; therefore, the refrigerant is in the saturation state at the intersection point. However, it can be seen that vaporization does not take place at the intersection point. The refrigerant remains liquid state until to the point c1. In the metastable liquid region (from point b to point c1), the refrigerant is in superheated liquid state. At the point cl, the vaporization takes place. In the metastable two-phase region (from point c2 to point d), the refrigerant is in non-equilibrium vapor-liquid state. Because of the vapor acceleration, the friction increases. Therefore, the temperature and pressure decrease suddenly in the equilibrium two-phase region (from point d to point e). The two curves of the temperature and the pressure join together, which indicates the thermal equilibrium state is reached after point d. After flowing through the capillary tube, the refrigerant changes from the high pressure and high temperature state to the low pressure and low temperature state because of friction in the capillary tube and the isothermal throttle effect of the refrigerant.
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Mixed-Refrigerant J-T Cryocoolers The refrigerants in the mixed-refrigerant J-T cryocoolers are mixtures that may optionally consist of methane, ethane, propane, i-butane, and nitrogen. Generally there are four to six components in the mixture. The selection of the component depends on the boiling point of the refrigerant and the required refrigeration temperature of the cryocooler. As shown in Fig. 1, the cryocooler is different from the household refrigerator. Therefore the operating condition of the capillary tube depends heavily on the operating condition of the heat exchanger. For the different refrigeration temperature ranges, especially the different working fluids, the flow process in the capillary tube reveals different characteristics. As shown in Fig. 4, when the system achieves at steady state, the pressure drop from the inlet to the outlet of the capillary tube is about 1.0 MPa. However, the temperature difference between the inlet and outlet of the capillary tube is very small. This phenomenon indicates that phase-equilibrium has been reached within the capillary. In this temperature range, the refrigerant mixture contains methane, ethane, propane, and i-butane. In the beginning, the refrigerants that flow through the capillary are all in the vapor phase. As the system runs, and the heat-exchanger gradually cools down, the component with the highest boiling point, such as i-butane and propane, changes to the liquid phase first. Until the system reaches steady state, the gases are all basically liquid phase in the capillary. That is to say in the 150 K-temperature range cryocooler, the refrigerants in the capillary are all in the liquid phase. This indicates why the temperature difference between the inlet and outlet of the capillary is smaller.
CONCLUSIONS A homogeneous flow model has been developed to study the performance of an adiabatic capillary tube for mixed-refrigerant J-T cryocoolers, and it has been compared with experimental data. For the multicomponent refrigerant in the cryocooler, the characteristics of the flow in the capillary are different from those of the pure refrigerant in a household refrigerator. Based on this analysis, an experimental setup is being fabricated, and further experimental testing is ongoing.
ACKNOWLEDGMENT The authors acknowledges with gratitude the continued support received from The National Natural Sciences Foundation of China under the contract number of 50076044.
REFERENCES 1. Qi, Y.F., J.F. Wu and M.Q. Gong, “Theoretical analysis and numerical simulation of capillary in mixed-refrigerant J-T cryocoolers,” Cryogenic Engineering, vol. 124, no. 6 (2001), pp. 6-12 (In Chinese).
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2. Bittle, R.R. and M.B. Pate, “A theoretical model for predicting adiabatic capillary tube performance with alternative refrigerants,” ASHRAE Transactions, vol. 102, no. 2 (1996), pp.52-64. 3. Chen, S.Li, and Y.R. Cheng., C.H. Liu, and C.S. Jwo, “Simulation of refrigerants flowing through adiabatic capillary tubes,” International Journal of HVAC&R Research, vol. 6, no. 2 (2000), pp. 101-115.
Study of a Vortex Tube by Analogy with a Heat Exchanger Y. Cao1, Y.F. Qi2, E.C. Luo2 , J.F Wu 2, M.Q. Gong 2, G.M. Chen1 1
Institute of Refrigeration and Cryogenic Engineering Zhejiang University, Hangzhou, China, 310027 2 Technical Institute of Physics and Chemistry Chinese Academy of Sciences, Beijing, China, 100080
ABSTRACT Based on the models of Scheper, Lewins, and Bejan, a new model has been established to study the influence of the cold mass flow fraction on the temperature separation effect in a vortex tube. The model is based on making an analogy between the vortex tube and a counterflow heat exchanger. The results show the model can accurately explain the correlation of cold mass flow fraction to the temperature separation effect. INTRODUCTION The Ranque-Hilsch vortex tube is an intriguing device in which a stream of high pressure gas swirls to split into two low pressure streams, one is higher than the entry temperature and the other is lower. It consists of a tube open at both ends and having one or more nozzles contained in a plane perpendicular to the tube’s axis. Air expands through these nozzles, which have tangential outlets entering the main tube. The air revolves inside the main tube and its angular velocity can be several million revolutions per minute as it moves along the tube towards the throttle valve where a proportion of the air is released. The remaining air moves along the center of the tube in the reverse direction, and is discharged through the orifice which forms the cold air outlet, while the air being discharged through the throttle valve is hot. The vortex tube is a very simple mechanical device without any moving components and is also easy to construct. Although its structure is very simple, this temperature separation has baffled researchers ever since it was first discovered by Ranque and studied by Hilsch.1 Many different qualitative explanations have been presented: internal friction theory2; turbulent heat transfer of thermal energy in an incompressible flow;3 Goertler vortices theory;4 acoustic streaming processes leading to increased flow velocities which therefore enhance the kinetic energy available for conversion into heat;5 and the secondary flow theory, 6,7 which shows that the thermal and fluid dynamics of the vortex tube are like the heat-pump cycle. However, none of these explanations so far has led to a quantitative model for the Ranque-Hilsch effect. The performance of the vortex tube varies with the fraction of the stream diverted to the cold side. Hence, for a general discussion it may be useful to have an elementary model of the relationship of how performance varies with the cold mass flow fraction. In this paper we have formulated a correlation of cold mass flow fraction to the temperature separation effect based on making an analogy between a vortex tube and a heat exchanger. Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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THEORETICAL MODEL Scheper8 first studied the vortex tube by making an analogy with a concentric-pipe counterflow heat exchanger, and Lewins and Bejan 9 developed it further. As shown in Fig. 1, one can imagine that there is a zero thickness tube wall, a diameter approximately equal to that of the cold orifice, which is rotating at every axial position at the same angular velocity as the vortex. The heat transfer across this imaginary tube surface is proportional to the bulk static temperature difference and to the overall film coefficient of heat transfer occasioned by the relative axial velocity of the oppositely moving streams. With a difference from Scheper’s theory, we can suppose that the gas expands in the vortex and produces an internal temperature difference. There will be a heat exchange between the two streams formed at different temperatures. This exchange will be by thermal diffusion so that we would hope that the vortex tube is a poor heat exchanger. There are some assumptions that can be made, such as the gas obeying ideal gas laws, having constant specific heat capacity, and considering a well insulated tube. To satisfy the necessary overall heat balance, from the First Law of thermodynamics, there are:
As in the case of conventional heat exchangers, a cooling effectiveness can be defined with the ratio of the actual heat transfer to the maximum possible heat transfer presented as follow10:
In equation (5), is the lowest available internal temperature that can be produced. If an ideal expansion is assumed, it is function of and There is an active length L along the heat exchanges with an effective thermal conductivity k, if the radial length scale is r0, then conventional heat exchange theory may be applied. The UA value of the tube as a heat exchanger is and we can set a modified nondimensional heat exchanger size, in NTU unit as notably independent of the radial size9. The expression for the effectiveness of a counter-flow heat exchanger with unbalanced streams is:
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When
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It is obtained:
From the definition of equation 4, it can be obtained:
Combined equation 2, 3, 5, 9 and 10, it can be obtained:
ANALYSIS AND DISCUSSION Several calculations have been made using various conditions such as the nondimensional size of the vortex tube, inlet temperature, the pressure ratio of the inlet pressure, and the cold outlet pressure. When the working medium is air and the inlet temperature is 300 K, the results are illustrated in Figs. 2 to 5. Figures 2 through 4 show the results.of the influence of the cold mass flow fraction on the temperature separation effect for pressure ratios of 3, 4 and 5, respectively. Figure 5 shows the influence of the cold mass flow fraction on the temperature separation effect compared to the experimental results for a No.2 Hilsch tube.
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From the results, it can be concluded: 1. When the cold mass flow fraction is between 0.3 - 0.4, the cooling effect achieves a maximum, and this finding agrees well with the results of experiments published in the open
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11,12,13
literature. When the pressure ratio increases, the cold mass flow fraction of maximal cooling effect gradually decreases. 2. The maximal cooling effect of the vortex tube will increase with a reduction of “M”, or slightly increase with an increase of the pressure ratio. 3. In the Fig. 6 comparison with the experimental results, the result (M= 0.07, = 4) agrees well with the ( two spray nozzles) for the Hilsch N0.2 vortex tube.1 4. Our results agree with the abnormal phenomenon observed in the literature,12,14 whereby heat is sometimes produced in the center of vortex tube. From Figs. 3, 4 and 5, it can be seen that when M is larger than 0.1, and the cold mass flow fraction is smaller than 0.15 and larger than 0.85, this heating phenomenon appears in our results.
CONCLUSION A useful temperature separation effect model of a vortex tube has been established by using an analogy with a heat exchanger. The model has been used to obtain a correlation between the cold mass flow fraction and the temperature separation effect. The model agrees well with experimental results published in the open literature, and illuminates the large role that heat transfer in the vortex has on the temperature separation effect; it also points out that the cold mass flow fraction is an important factor of the heat transfer. However, because the model is established from a macro view, without consideration of the actual complicated heat and mass transfer, it has limited ability to explore deeper into the physics of the vortex tube operation. For this one would need models based on detailed computational fluid dynamics, which may be a very considerable challenge, and would require further effort.
ACKNOWLEDGMENT The authors acknowledges with gratitude the continued support received from The National Natural Sciences Foundation of China under the contract number of 50076044.
Nomenclature specific heat capacity, J/kg total mass flow rate, kg/s inlet pressure, MPa cold outlet pressure, MPa inlet temperature, K cold outlet temperature, K hot outlet temperature, K the lowest available static temperature, K µ cold mass flow fraction, k effective conductivity, W/m.K L active vortex tube length, m NTU nondimensional heat transfer units M tube size, dimensionless N smaller flow-side tube size, dimensionless heat exchanger effectiveness Subscripts C cold H hot
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REFERENCES I.
Hi1sch, R., “The use of the expansion of gases in a centrifugal field as a cooling process,” Review Sciences Instruments, vol. 18, no.2 (1947), pp. 108-113.
2.
Deissler R.G., Perlmutter M., “Analysis of the flow and energy separation in a turbulent vortex,” International Journal of Heat & Mass Transfer, vol. 1 (1960), pp. 173-191.
3.
Linderstrom-Lang C.U., “The three-dimensional distributions of tangential velocity and total temperature in vortex tubes (part I) , ” Journal of Fluid Mechanics, vol. 45 (197 1), pp. 161-187.
4.
Stephan K., Lin S., “An investigation of energy separation in a vortex tube”, International Journal of Heat & Mass Transfer, vol. 26, no. 3 (1983), pp. 341-348.
5.
Kurosaka M., “Acoustic streaming in swirling flow and the Ranque-Hilsch (vortex tube) effect,” Journal of Fluid mechanics, vol. 124 (1982), pp. 139-172.
6.
Ahlborn B.K.,Groves Stuart, “Secondary flow in a vortex tube,” Fluid Dynamics Research, vol. 21 (1997), pp. 73-86.
7.
Ahlborn B.K., Keller J.U., Rebhan E., “The heat Pump in a Vortex Tube,” Journal of Non Equilibrium Thermodynamics, vol. 23 (1998), pp. 159-165.
8.
Scheper G. W.,”The vortex tube: internal flow date and a heat transfer theory,” Refrigeration engineering, October 1951, pp. 985-989.
9.
Lewins Jeffery, Bejan Adrian, “Vortex tube optimization theory,” Energy, vol. 24, no. 3 (1999), pp. 931-943.
10. Helmut Wolf, Heat Transfer, New York (1983), pp. 309-323. 11. Yang Zeliang, Yang Cheng, “Experimental Study on the Strongly Swirling Flow of Compressible Fluid in Vortex Tube,” Journal of Experimental Mechanics, vol. 13, no. 3 (1998), pp. 399-403 (in Chinese). 12. Cockerill T.T., Thermodynamics and fluid mechanics of Ranque-Hilsch vortex tube, [MSc thesis] Cambridge University, UK, 1998. 13. Bruun H.H., “Experimental investigation of the energy separation in vortex tube,” Journal Mechanical Engineering Science, vol. ll, no.6(1969),pp. 567-582. 14. Da Zhong Lao, Study on the Performance Characteristics of Vortex Tube, [PhD thesis], Dalian University of Technology, China, 1995.
Thermodynamic Prediction of the Vortex Tube Applied to a Mixed-Refrigerant Auto-Cascade J-T Cycle Y.Cao 1, M.Q.Gong2 , Y.F.Qi2, E.C. Luo2, J.F.Wu2, G.M.Chen1 1
Institute of Refrigeration and Cryogenic Engineering Zhejiang University, Hangzhou, China, 310027 2 Technical Institute of Physics and Chemistry Chinese Academy of Sciences, Beijing, China, 100080
ABSTRACT A new hybrid refrigeration cycle of the Mixed-refrigerant Auto-Cascade J-T cycle combined with a vortex tube is introduced. This cycle, which holds the advantage of a mixed refrigerant Auto-Cascade J-T cycle and the vortex tube, is expected to achieve a temperature lower than 65 K. A thermodynamic model is used to investigate the vortex tube energy separation. The exergy method is applied to analyze the thermodynamic performance of each unit and the whole refrigeration cycle. A comparison is made between the Auto-Cascade J-T cycle and the new hybrid refrigeration cycle for the same conditions. The total exergy efficiency achieved in the new hybrid refrigeration cycle is 78.9% better than the Auto-Cascade J-T cycle. The results show that using the new type of compound refrigeration cycle can improve exergy efficiency of the whole cycle. It is completely possible to achieve a temperature lower than 65 K.
INTRODUCTION In recent years, there has been a remarkable development of the mixed-refrigerant Auto Cascade refrigeration cycle which is also called the Kleemenko cycle.1,2 There are many people interested in this cycle. Without any modification of the hardware, it can produce cryogenic and ultra-low refrigeration covering a large temperature range from 80 K to warmer than 230 K — the conventional vapor-compression cycle temperature range. In this cycle one or more intermediate phase separators (liquid-vapor separators) are employed; this avoids the compressor lubricant and high boiling components plugging problem in the coldest section. In addition, driven by a commercial air-conditioning compressor, this type of the mixed-refrigerant Auto-Cascade throttle cryocooler is more reliable and flexible in many applications such as with infrared detectors, cryosurgical devices, HTS devices, material studies, biomedical storage, etc. However, because there are no substances in nature that have a boiling point between nitrogen (77.4K) and neon (27K), the efficiency of the Mixed-refrigerant J-T cryocooler decreases rapidly at temperatures lower than liquid nitrogen temperature (77.4K). Therefore, a mixed-refrigerant J-T cryocooler has trouble achieving a temperature lower than 65 K. In order to achieve a temperature
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lower than 65 K, this paper discusses a new hybrid refrigeration cycle: a Mixed-refrigerant Auto-Cascade J-T cycle combined with a vortex tube. The vortex tube, as used, is to be regarded as a J-T component, and is not limited by the inversion temperature. In general, for a vortex tube, the well-balanced expansion pressure ratio is 3-8, which is just the operating pressure ratio of the cycle. The exergy method is applied to analyze the thermodynamic performance of components and the whole refrigeration cycle.
VORTEX TUBE A French physicist whose name is Georges Ranque first discovered the vortex tube in 1930. However, Hilsch was the first engineer to develop this phenomenon into a practical, effective cooling solution for industrial applications, and to generate a theoretical model and calculations.3 Therefore, the vortex tube is also called Ranque-Hilsch vortex tube. The Ranque-Hilsch vortex tube is an intriguing device in which a stream of high pressure gas swirls to split into two low pressure streams, one higher than the entry temperature and the other lower. It consists of a tube open at both ends and having one or more nozzles contained in a plane perpendicular to the tube’s axis. Air expands through these nozzles, which have tangential outlets entering the main tube. The air revolves inside the main tube and its angular velocity can be several million revolutions per minute as it moves along the tube towards the throttle valve where a proportion of the air is released. The remaining air moves along the center of the tube in the reverse direction, and is discharged through the orifice which forms the cold air outlet, while the air discharged through the throttle valve is hot. The efficiency of the vortex tube is lower than the adiabatic expansion of gas with external work, but higher than isenthalpic throttling. Though the mechanism of the temperature separation has not been explained up to the present, the vortex tube is a very simple mechanical device without any moving components and also easy to construct. Hence, vortex tubes have been applied in refrigeration, air conditioning, liquefaction of gases, purification of natural gas, air supplied suits, cutting tools, etc.4 HYBRID REFRIGERATION CYCLE DESCRIPTION Figure 1 shows a basic flow diagram of the two stages of the Auto-Cascade cycle. The number of the heat exchangers and phase separators in the Auto-Cascade cycle varies flexibly based on the required cooling temperature. Figure 2 shows a basic flow diagram of a two-stage Auto-Cascade cycle combined with a vortex tube. It consists of a compressor, an after-cooler, four counter-flow heat exchangers, two vapor-liquid separators, two throttle devices that are usually a capillary or an orifice, an evaporator, and a vortex tube in the last stage. Any compressor lubricant entrained in the circulating refrigerant mixture requires proper management. The employment of an oil separator in the cycle may avoid plugging problems in the coldest section.
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In this system, after each cooling step in a counter-flow heat exchanger, a phase separator removes the condensate from the vapor stream; a throttling device controls the exiting liquid flow. The liquid passing through the throttling device usually becomes a two-phase blend. The twophase blend, mixed with the low pressure return stream, cools the incoming high pressure stream. Finally, the high pressure stream passes through the vortex tube to produce the lowest refrigeration temperature stream through the evaporator. The other stream mixes with the low pressure return stream of the counter-flow heat exchanger III. The only difference between Fig. 1 and Fig. 2 is the last stage. In Fig. 1, the last stage is a throttling device; however, in Fig. 2, the last stage is a vortex tube. Special attention must be paid to the heat stream of the vortex tube entering into the low pressure return stream of the counter-flow heat exchanger III.
THERMODYNAMIC PROCESS ANALYSIS The conventional energy analysis considers only heat balance. An exergy analysis, based on irreversibility, includes the losses due to heat transfer and pressure drop. Therefore, an exergy analysis was employed in the performance of the hybrid refrigeration cycle of Mixed-refrigerant Auto-Cascade J-T combined with vortex tube. Gong has analyzed the thermodynamic performance of the mixed-refrigerant Auto-Cascade refrigeration cycle in detail. The thermodynamic model of the compressor, after-cooler, counter current flow heat exchanger, liquid-vapor separator, throttle device, blender process, and the evaporator have been previously established.5 These models are not illustrated in this paper. Only the thermodynamic model of the vortex tube is discussed. The overall-cycle thermodynamic model is based on our previous models.5 For the L-V separator, the emphasis is placed on the calculation of the phase equilibrium of the mixture. In this calculation, the exergy loss of the L-V separator is not considered. The exergy equation of the hybrid refrigeration cycle is presented as follows:
where: W is the input power, is the exergy of the cooling capacity, is the exergy loss of each unit of the cycle, is the exergy efficiency of the refrigeration cycle. The following task is to find out the exergy loss of the each unit of the hybrid refrigeration cycle.
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Vortex tube The conditions of flow in a vortex tube can be assumed adiabatic, steady state having no heat and work interaction with environment, illustrated in Fig. 3. The entropy equation of the vortex tube is described as follows:
The common method for calculating total irreversibility is to use a Gouy-Stodola relation:
where :
is the mass flow; Subscripts i, c and h denote the inlet stream, cold
stream and heat stream of the vortex tube, respectively. With the assumption of ideal gas, it can be obtained7:
THERMODYNAMIC PERFORMANCE SIMULATION Based on the model of energy analysis described above, a program has been developed for the simulation of the thermodynamic process of the Auto-Cascade cycle and the hybrid cycle. The main task of this article is focused on the difference between the Auto-Cascade cycle and the hybrid cycle, especially on the thermodynamic parameters coupling in the two cycles. Therefore, the separating process is considered as ideal processes in the calculation, that is, no exergy loss occurred. The thermodynamic performances of the ideal refrigeration cycles, are simulated at the same given calculation condition. The calculation conditions are presented as follows: Mixture: (25%), (30%), (15%), (10%), (10%), (10%); Temperature: The cold mass flow fraction of vortex is 0.68. The simulation results are presented in Table 1. From Table 1, it can be easily found that at the same condition the thermodynamic efficiency of the new type of hybrid cycle is larger than that of the Auto-Cascade J-T cycle. The improvement of the thermodynamic performance is mostly as a result of the exergy loss reduction in the last stage of throttle device. Shown in Table 1, the exergy loss of the vortex tube is up to 15.6% of the total exergy input, and 16.8% of the total exergy loss of the Auto-Cascade J-T cycle. Compared to the Auto-Cascade J-T cycle, the exergy loss of vortex tube is 6.2% of the total exergy input, and 7.1% of the total exergy loss of the hybrid refrigeration cycle. The total exergy gained in the new hybrid refrigeration cycle is 78.9% better than the Auto-Cascade J-T cycle.
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CONCLUSION In this paper, a new hybrid refrigeration cycle made up of a Mixed-refrigerant Auto-Cascade J-T cycle combined with a vortex tube is presented. A thermodynamic model of the hybrid refrigeration cycle has been established, and an exergy analysis method has been used to analyze the hybrid refrigeration cycle. The calculation results show that a Mixed-refrigerant Auto-Cascade J-T cycle combined with a vortex tube can improve the thermodynamic performance of the cycle compared to the standard Auto-Cascade J-T cycle. In addition, the study suggests that the vortex tube has special advantages in achieving temperatures lower than 65K, because the heat stream exergy of the vortex tube can be used. It can also provide a new way to apply the vortex tube. From the above thermodynamic analysis of the hybrid refrigeration cycle, it has been shown that it is completely possible to achieve temperatures lower than 65 K. An experimental setup is being fabricated, and further experimental tests are ongoing.
ACKNOWLEDGMENT The authors acknowledges with gratitude the continued support received from The National Natural Sciences Foundation of China under the contract number of 50076044.
REFERENCES 1.
A. P. Kleemenko, “One flow cascade cycle,” The proceedings of Xth International Conference of Refrigeration, I-a-6 (1959), pp. 34-39.
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2.
E.G. Luo, M.Q. Gong, Y. Zhou, “Experimental Investigation of a Mixed-Refrigerant J-T Cryocooler Operating from 30 to 60K,” Advances in Cryogenic Engineering, Vol.45 (2000), pp. 315-322.
3.
Hilsch R., “The use of the expansion of gases in a centrifugal field as a cooling process,” Review Sciences Instruments, vol. 18, no.2 (1947), pp. 108- 113.
4.
Thomas T Bruno, “Laboratory application of the vortex tube,” Journal of Chemical Education, vol. 64, no. ll(1987), pp.987-988.
5.
M.Q. Gong, et al., “Therodynamic Analysis on the Mixed-Refrigerant Auto-Cascade J-T Cryocooler with Distributed Heat Loads,” Cryocooler 11, Kluwer Academic/Plenum Publishers, New York (2001), pp. 523-530.
6.
M.H.Saidi, M.R.Allaf Yazdi, “Exergy model of a vortex tube system experimental with result,” ENERGY, Vol.24 (1999), pp.625-632.
Evaluation of Hydride Compressor Elements for the Planck Sorption Cryocooler R.C. Bowman, Jr.1, M. Prina1, D.S. Barber1, P. Bhandari1, D. Crumb2, A.S. Loc1, G. Morgante3, J.W. Reiter2, and M.E. Schmelzel1 1
Jet Propulsion Laboratory, California Institute of Technology Pasadena, C A 91109, US A 2 Swales Aerospace Pasadena, CA 91107, USA 3 IASF/CNR - Sezione di Bologna Bologna, 40129, Italy
ABSTRACT Hydrogen sorption cryocoolers are being developed for the European Space Agency Planck mission to provide nominal 19 K cooling to instruments for measuring the temperature anisotropy of the cosmic microwave background with extreme sensitivity and resolution. The behavior of the metal hydride sorbent beds used in the compressor dominates both the performance and reliability of these sorption cryocoolers. The compressor elements have been designed to minimize their input power requirements and to enhance durability during extended temperature cycling while in operation. The Lanthanum-Nickel-Tin alloy in the sorbent beds circulates and compresses the hydrogen refrigerant gas while the ZrNi alloy is used to provide variable pressure in the gas-gap heat switches for each compressor element. Characterization tests have been performed on the compressor elements built for an Engineering Bread Board (EBB) cooler to evaluate the behavior of both the sorbent bed and gas-gap switches under conditions simulating flight operation. These results provide a basis for predicting EBB cooler performance and to identify any design deficiencies prior to fabrication of the flight compressor elements. In addition, experiments were done on compressor elements that had been operated up to several thousand cycles to assess degradation in the sorbent hydride and reduction in the effectiveness of the gas gap switches in reducing parasitic heat losses
INTRODUCTION The Jet Propulsion Laboratory (JPL) is developing hydrogen sorption cryocoolers to provide cooling below 19 K to the instruments on the Planck spacecraft, which is scheduled to launch in 2007. Descriptions of the cryogenic systems for the Planck mission have been published1,2. Laboratory testing of a full-scale Engineering Bread Board (EBB) version of the Planck sorption cooler was started3 in January 2002. The present paper presents a comprehensive description of the design and performance of the hydride compressor elements that were utilized in the EBB cooler3. After final iterations, the components will be built for integration into the flight coolers. Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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DESCRIPTION OF COMPRESSOR ELEMENTS The compressor assembly for a Planck sorption cooler2 has six identical compressor elements (CEs) each containing 615 grams of the hydride alloy in its sorbent bed. The CEs are independently heated and cooled through a series of heat-up, desorption, cool-down, and absorption steps to provide compression and circulation of the hydrogen refrigerant gas during closedcycle Joule-Thomson (J-T) process that generates liquid hydrogen at ~ 18 K in the cryostat. The CEs are directly mounted to a radiator that is sized to reject the heat from the input power and exothermic hydrogen absorption by the alloy at 270 K +10 K/-20 K. Gas gap heat switches4,5 are used to facilitate heat transfer between the sorbent bed and the radiator while minimizing thermal mass during the heat-up and desorption phases. Although the first version of the Planck compressor element was described previously2,6, several modifications have been made to make the unit more robust during launch vibrations, reduce thermal gradients in the sorbent bed, etc. An exploded view of components for the compressor elements used in the EBB cooler is given in Fig. 1. The central portion of the outer tube is aluminum metal type 6061-T6, the foam is also 6061 aluminum at ~11% of its bulk density, the button is A286 stainless steel (SS), and all other components in contact with hydrogen are made from 316L stainless steel. To minimize sources of contamination and enhance impurity removal, all surfaces of the 316L SS were electropolished giving bright surfaces. The outer surface of the inner tube assembly and both surfaces of the aluminum portion of the outer tube were electroplated with a gold film of nominal thickness of 0.75 µm to reduce thermal emissivity in the gas gap volume. A cross sectional view of the assembled compressor element is shown in Fig. 2 and illustrates the 0.75 mm gas gap separation between the inner and outer tubes. The thickness of the wall for the inner tube is 1.22 mm except at the weld zones on the ends where it is 1.52 mm. The porous filter tube ensures that hydride powder contained in the Al foam does not migrate from the sorbent bed during the temperature and pressure cycling. Tight physical contact of the Al foam with the inner surface of the tube wall provides heat transfer from the sorbent bed to the gas gap. The sorbent bed is attached to the outer housing only at the ends using the supports shown in Fig. 1. All assembly joints are made using automatic orbital-tube welding under argon/3% gas.
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Eight compressor elements were fabricated and assembled: Six for the EBB sorption cooler, one for vibration/shake testing, and one spare. A photograph of the spare CE with a gas gap actuator (GGA) attached is shown in Fig. 3. The assembly of the CEs occurred in three phases after the fabrication of the various components and processing of the and ZrNi alloys for the sorbent bed and gas gap actuator bed, respectively. First, the inner bed is filled with alloy powder and a closure weld performed with all processing occurring in an argon atmosphere (<1 ppm and <2 ppm ) glove box. After a uniform powder distribution is achieved in the sorbent bed, the alloy is activated by vacuum baking above 525 K and reacted with purified hydrogen gas yielding hydride compositions in agreement with independent isotherm measurements. The inner bed assembly was next sent to Epner Technology (Brooklyn, NY) for gold electroplating. The gas gap actuators, which are shown in Fig. 1, were assembled following previously described procedures4,5,7 using strips of high purity ZrNi cut with a diamond saw. Their external heaters were directly brazed to the GGA cap to provide stable thermal contact7. The resulting stability and reproducibility of each GGA unit was verified by performing a minimum of ~1000
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temperature cycles over the expected operation range. Typical results obtained for one of the GGA units is shown in Fig. 4. No degradation in the upper and lower pressure values or the absorption and desorption kinetics was found. Similar results were obtained from all GGAs that were installed on EBB compressor elements. A manual valve was used to isolate activated and charged GGA with its initial hydride content of The third phase was sequential integration of goldplated inner bed assembly to the outer tube assembly and a GGA unit, which was followed by extended hydrogen gas fill and evacuation cycling to clean the gas gap volume. During this process, the inner bed was heated to ~670 K with its internal heater while distilled water heated to ~360 K was circulated through a chiller plated mounted on the outer shell, which was covered in an insulation blanket. The gas gap volume was alternatively filled with purified hydrogen at ~1 bar pressure and evacuated with a turbomolecular pumping station. These fill-evacuate cycles were repeated until the mass spectrum from a residual gas analyzer (RGA) indicated no hydrocarbon species above instrumental background levels (i.e., between 50 - 100+ cycles were required to obtain sufficient cleaning) in the vacuum state.
PERFORMANCE TESTING OF INTEGRATED COMPRESSOR ELEMENTS In order for the Planck sorption cryocooler to meet its performance goals2, numerous requirements are imposed8 on the compressor elements. The heated sorbent beds need to provide hydrogen gas at 50 bar pressure and an average mass flow rate of 6.5 mg/s with a total (i.e., heat-up, desorption, and gas gap) input power of <410 W at end-of-life (i.e., two years of operation during ground tests and flight). During absorption, the beds must maintain hydrogen pressure below 0.59 bar (445 Torr) for a radiator temperature of 270 K. The ON-state pressure for the gas gap switch must exceed 800 Pa (6 Torr) and its OFF-state pressure go below 0.7 Pa (5.2 mTorr) with switching times below 250 s. Past studies6,9 with the first prototype Planck CEs demonstrated that the sorbent bed met its beginning-of-life (BOL) requirements using an external
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gas manifold and vacuum pump for the heat switch. In addition, life cycling tests showed4,7 that actuators composed of also satisfied the requirements for heat switches. However, the behavior of the complete Planck compressor elements with hydride gas gap actuators had not been evaluated. These measurements have now been performed on seven of the EBB compressor elements. An individual EBB compressor element was installed into a dedicated vacuum test chamber equipped with separate gas lines for the sorbent bed and gas gap volume, a circulation loop to an external refrigerator capable of cooling the CE chiller plate to 278 K, and hermetic electrical connections for the heaters and pressure and temperature sensors. The sorbent bed gas line was attached to the gas manifold described by Pearson, et al.9 for controlled gas absorption and desorption that closely simulates behavior in the operating cryocooler. Lab View software controlled the heaters, electropneumatic valves and pressure control valves as well as recorded the data from the sensors. The relationship between the pressure and temperature in the gas gap with the parameters for the CE sorbent bed is illustrated in Fig. 5. The change in thermal conductance across the gas gap as a function of pressure is shown in Fig. 5(a) while the time constants for temperature variations during the ON-OFF and OFF-ON switches are given in Fig. 5(b) along with sorbent bed pressures during two heating/cooling cycles. The hydride-actuated gas gap heat switches clearly meet their performance requirements during both transitions. A detailed comparison of the sorbent bed temperature and hydrogen mass flow rate during the desorption step is given in Fig. 6. The initial flow oscillations at the start of desorption are artifacts from the outlet flow controller system and were not seen during cooler operation3 using the J-T expander. The input power during the desorption process in Fig. 6 was 158 W and outlet pressure was held at a constant 50.7 bar. While the flow varied during the desorption process, the average was close to the goal of 6.5 mg/s (i.e., 4.34 sl/m). The maximum temperature did not reach 465 K at the end of desorption. Representative hydrogen absorption behavior observed from the same CE is shown in Fig. 7 where the shell temperature is maintained within 2 K of the chiller plate value of 278.5 K throughout the absorption. The hydrogen flow rate is around 1.42 sl/m (i.e., 2.1 mg/s) as specified2,8 for boiloff gas from the liquid hydrogen reservoirs from Planck cooler operation at its nominal heat load. The temperature measured by the thermocouple in the hydride sorbent bed is 7-8 K above the shell temperature due to the limited thermal conductance of the powder/Al foam matrix to remove the heat from the exothermic absorption process. While the bed temperature remains fairly constant the pressure rises noticeably during the last third of the absorption. The pressure and temperature
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spikes at the beginning and end are again artifacts from the control system. These effects are not seen during EBB cooler operation . Seven compressor elements were each subjected to 50 - 100+ cooler simulation cycles to assess their initial behavior. Although there were some variations in the flow rate, temperature, and pressure profiles due to differences in starting hydride compositions and chiller plate temperatures, overall agreement in average mass flows, input power, pressure values, and gas gap parameters was observed. The behavior was found to be much more consistent among the seven EBB units than was previously described by Pearson et al.9 from the three Planck Life-cycle compressor elements (LCEs). This improvement is mainly attributed to a more uniform distribution of hydride powder in the EBB sorbent beds, where much greater efforts were taken to avoid segregation and separated regions (as confirmed by x-ray radiographic imaging) prior to hydride activation.
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VIBRATION TESTING OF A COMPRESSOR ELEMENT After completion of its initial performance cycling tests, the CE designated for vibration testing was transferred into the glove box where the gas gap actuator assembly was removed and replaced with a VCR plug. This assembly was mounted on a custom vibration test fixture and instrumented with several force transducers/accelerometers. The hydrogen pressure in the sorbent bed was 0.42 bar and the hydride stoichiometry was determined to have been from quantitative desorption measurements following the vibration tests. The EBB compressor element was subjected to sine sweep and random vibration in each of its three principle axes at qualification levels expected during system random testing and launch on an Ariane V rocket1. The vibration tests were completed with no structural failures occurring although metal-to-metal impacting was detected during the high level runs and some minor modal frequency shifts were noted in the sine surveys. Various post-vibration tests confirmed that no damage has been produced by these shake tests. X-ray radiography indicated no redistribution of the hydride in the sorbent bed and no displacements or cracks of the support structures at either end of inner tube assembly. A series of helium leak tests were performed of the gas gap volume (i.e., external leaks with the unit in a helium filled bag, the leaks from the inner bed when it was pressurized to 52 bar at ambient temperature, and with the inner bed heated 470 K containing helium pressure of 52-55 bar) where no leaks were detected above the background level of under any test condition. Further performance and cycling tests are planned to see whether there are any other changes although preliminary tests appear consistent with pre-vibration behavior.
LONG TERM OPERATION DEGRADATION AND CONTAMINATION The Planck compressor elements will undergo ~20,000 cycles between 270 K and 470 K during ground testing and flight operation2. The degradation of the sorbent and gas gap hydrides has been a concern since the conception of the Planck sorption cryocooler and considerable efforts have been made to evaluate their behavior during accelerated aging studies3-7,9,10. The CE design includes storage margin that accounts for anticipated2 rates of hydride degradation. The stoichiometric alloy (i.e., [Ni + Sn]/La ratio = 5.0) used in the EBB-CE sorbent beds has recently shown7,10 smaller rates of degradation than the material used3,7,9 on the earlier Life-cycle compressor elements (LCEs). To date there has been no indication of hydride degradation on the performance of the EBB compressor elements during their characterization tests or initial operation of the EBB cooler2. Three LCE units were thermally cycled up to 5000 times as reported previously9. The hydrogen gas from these sorbent beds has been recently analyzed by RGA and mass spectrometry. The dominant impurity was methane in the concentration range of 200-370 ppm relative to the total hydrogen content. There was also evidence of moisture and a species at mass peak 28 amu (i.e., either or CO) present at variable levels in the different beds. The beds had been initially filled with research grade hydrogen gas (i.e., 99.999+% purity) that was further purified by flowing it through a Nanochem chemical purifier and a carbon cold trap cooled in liquid nitrogen prior to admitting this gas into the hydrogen filling station. Hence, the impurities in the LCE gas seen after cycling were probably generated from residual hydrocarbons on surfaces of filters, foam, and other components even though vacuum and purge gas cleaning was performed during activation. The hydride may act as a catalyst for the conversion of condensed impurities into methane, water, and CO. Since all these molecular species will form solids well above the temperature of liquid hydrogen, they would likely cause plugging at the J-T expansion valve. The Planck sorption coolers contain chemical purifiers/getters and a carbon trap cooled to 50 K to removed these condensable species from the hydrogen gas before entering the cold system and the J-T valve region2,9. However, the cycling of the LCE beds and individual EBB CEs was done without any filtering or gettering to remove impurities from the hydrogen. The remaining issue is how large to size these devices for the flight coolers to ensure the effective removal of the impurities during the nominal two years of
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flight operation? A methane level of 43 ppm was observed in the EBB CE used for the vibration tests after only 68 cycles. It is thus likely methane formation is most rapid at the beginning of cycling when the residual hydrocarbon content is expected to be the highest. A combination of initial cycling, evacuation and refilling with pure hydrogen may prove to be a viable means of eliminating long-term creation of methane and the other species. This supposition will be examined during future tests of both individual EBB compressor elements and the EBB cooler during extended operation. During the time period between the completion of the characterization tests of the individual EBB compressor elements and start of the initial cooler operation, the two manual valves for the gas gap volume (see Figs. 1 and 3) remained closed isolating the ZrNi hydride from the rest of the gas gap. When the pressure was checked for each CE, the pressures in the isolated gas gap volumes range from 70 Pa to 700 Pa (i.e., 0.5 - 5.0 Torr) following ambient temperature storage between two and six months. Several tests, including RGA determination of gas composition, identified that hydrogen was responsible for this entire pressure rise. Since the OFF state pressure in the gas gap must lie below 1.3 Pa (~10 mTorr) to minimize excessive parasitic heat leaks4,5 to the outer shell and radiator, the magnitude of hydrogen pressure increase is a serious issue. Because the gas gap sorbent was configured to work in the middle of its plateau region4,7 during heat switch cycling, it can accommodate a certain amount of additional hydrogen without significant performance impact. However, its capacity is limited and increasing the mass and size of the actuator will require additional power to activate the heat switches. Consequently, quantitative assessment of amount of hydrogen has become imperative before the design of the compressor elements is finalized and the flight units fabricated. Measurements of the rate of pressure increase in the gas gap volumes of two of the EBB compressor elements have been done under various conditions. The outer shell was maintained between 290 K and 296 K while inner bed was held at fixed temperatures from 293 K to 540 K and the pressure rise with time was recorded. Tests were made with the sorbent bed under vacuum, helium (0.6 - 35 bar), and hydrogen (0.5 - 55 bar). The observed rates for the quantity of gas produced in units of standard cubic centimeters per second (scc/s) are nearly independent of the pressure and gas composition in the sorbent bed, but are strongly temperature dependent. The rates range from at 293 K to at 525 K. These values are consistent with hydrogen outgassing and permeation rates reported11-14 for stainless steel and other metals. These tests are still in progress and additional measurements are planned to determine the total hydrogen contents in the electroplated gold films, the nickel under layers, and host 316L SS and Al metals. The results will be reported elsewhere in the future. The information on hydrogen contents in the structural materials and outgassing/permeation rates will be used to properly size the gas gap actuators and make any other modifications to provide efficient heat switch performance during the planned operational life of the Planck sorption cooler.
CONCLUSIONS The second generation of compressor elements for the Planck sorption cryocoolers has been successfully fabricated for use in the EBB system. The hydride gas gap actuators have been found to work efficiently and reliably. The CE sorbent beds deliver and absorb hydrogen within the performance specifications. The results from the vibration tests and post shake measurements confirm that the basic structural design and component fabrication processes are sufficiently robust to withstand the expected launch conditions. Formation of methane has been observed during the temperature cycling of the CEs and this impurity must be removed from the hydrogen gas stream before entering the region of the J-T expansion valve. While the coolers have cold carbon traps for this purpose, improvements in cleaning processes are also being investigated for use on the flight units. Outgassing of hydrogen gas into the gas gap volumes has been detected and alternative methods to minimize its impact through use of larger actuator beds or incorporation of supplemental valves will be considered after the total magnitude of gas has been established.
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ACKNOWLEDGMENTS The research described in this paper was carried out at the Jet Propulsion Laboratory, California Institute of Technology, under a contract with the National Aeronautics and Space Administration. We thank P. A. Barrett, M. R. O’Connell, and the personnel at the TRW Space Park dynamic test laboratory for performing the vibration tests. We also thank J. G. Kulleck for the x-ray radiography measurements.
REFERENCES 1. Collaudin, B. and Passvogel, T., “The FIRST and PLANCK ‘Carrier’ Missions. Description of the Cryogenic Systems”, Cryogenics, vol. 39 (1999) pp 157-165. 2.
Wade, L.A. et al, “Hydrogen Sorption Cryocoolers for the PLANCK Mission”, in Advances in Cryogenic Engineering 45A, edited by Q-S. Shu, et al., Kluwer Academic/Plenum, New York, 2000, pp. 499-506.
3.
Pearson, D., Prina, M., Borders, J., Bowman, R.C., Schmelzel, M.E., Hardy, J., Sirbi, A., Bhandari, P., Loc, A., Wade, L.A. and Nash, A. “Test Performance of a Closed Cycle Continuous Hydrogen Sorption Cryocooler”, presented at 12th International Cryocooler Conference, Cambridge, MA, June 18-20, 2002 (Submitted to These Proceedings).
4.
Prina, M., Bhandari, P., Bowman, Jr., R.C., Paine, C.G., and Wade, L.A., “Development of Gas Gap Heat Switch Actuator for the Planck Sorption Cryocooler”, in Advances in Cryogenic Engineering 45 A, edited by Q-S. Shu, et al., Kluwer Academic/Plenum, New York, 2000, pp. 553-560.
5.
Prina, M., Kulleck, J.G., and Bowman, Jr., R.C., “Assessment of Zr-V-Fe Getter Alloy for Gas-gap Heat Switches”, J. Alloys Comp., vol. 330-332 (2002) pp. 886-891.
6.
Paine, C.G., Bowman, Jr., R.C., Pearson, D., Schmelzel, M.E., Bhandari, P., and Wade, L.A., “Planck Sorption Cooler Initial Compressor Element Performance Tests”, Cryocoolers 11, Kluwer Academic/ Plenum Press, New York (2001) pp. 531-540.
7.
Bowman, Jr., R.C., Prina, M., Schmelzel, M.E., Lindensmith, C.A., Barber, D.S., Bhandari, P., Loc, A. and Morgante, G., “Performance, Reliability, and Life Issues for Components of the Planck Sorption Cooler”, in Advances in Cryogenic Engineering, Vol. 47, edited by S. Breon, et al. (Am. Inst. Phys., New York, 2002) pp. 1260-1267.
8.
Prina, M., Bhandari, P., Bowman, R.C., Wade, L.A., Pearson, D.P., and Morgante, G., “Performance Prediction of the Planck Sorption Cooler and initial Validation”, in Advances in Cryogenic Engineering, Vol. 47, edited by S. Breon, et al. (Am. Inst. Phys., New York, 2002) pp. 1201-1208.
9.
Pearson, D., Bowman, Jr., R.C., Schmelzel, M.E., Prina, M., Bhandari, P., Paine, C.G., and Wade, L.A., “Characterization and Lifecycle Testing of Hydride Compressor Elements for the Planck Sorption Cryocooler”, in Advances in Cryogenic Engineering, Vol. 47, edited by S. Breon, et al. (Am. Inst. Phys., New York, 2002) pp. 1209-1216.
10. Bowman, Jr., R.C., Lindensmith, C.A., Luo, S., Flanagan, T.B., and Vogt, T., “Degradation Behavior of (x = 0.20 to 0.25) at Elevated Temperatures“, J. Alloys Compounds, Vol. 330-332 (2002) pp. 271-275. 11. Young, J.R., “Outgassing Characteristics of Stainless Steel and Aluminum with Different Surface Treatments”, J. Vac. Sci. Technol., Vol. 6 (1969) pp. 398-400. 12. Perkins, W.G., “Permeation and Outgassing of Vacuum Materials”, J. Vac. Sci. Technol., Vol. 10 (1973) pp. 543-556. 13. Le Claire, A.D., “Permeation of Gases Through Solids”, Diff. Defects Data, Vol. 34 (1983) pp.1-35. 14. Ishikawa, Y., Koguchi, Y., and Odaka, K., “Outgassing Rate of Some Austenitic Stainless Steels”, J. Vac. Sci. Technol., Vol. A9 (1991) pp. 250-253.
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Initial Test Performance of a Closed-Cycle Continuous Hydrogen Sorption Cooler, the Planck Sorption Breadboard Cooler M. Prina1, G. Morgante2, A. Loc1, M. Schmelzel1, D. Pearson1, J. W. Borders1, R.C. Bowman1, A. Sirbi1, P. Bhandari1, L.A. Wade1, A. Nash1 1
Jet Propulsion Laboratory California Institute of Technology Pasadena, CA 91109 2 CNR-IASF Sez. Bologna Bologna, 40129 Italy
ABSTRACT The Jet Propulsion Laboratory (JPL) is developing a continuous hydrogen sorption cryocooler for the ESA Planck mission, which will measure the anisotropy in the cosmic microwave background. The sorption cooler is the only active cooling for one of the instruments and it is the first of a chain of three coolers for the other instrument on Planck. The cooler has been designed to provide a cooling capacity of 1.1 W at a temperature below 20 K with a temperature stability requirement of 100 mK over a compressor cycle (667 s). The performance of these coolers depends on many operating parameters (such as the temperatures of pre-cooling thermals shield and the warm radiator and their fluctuations) and compliance can only be assessed through a detailed testing of the whole cooler and its interfaces. A breadboard sorption cooler (EBB) is undergoing testing to verify the flight cooler design performance in terms of input power, cooling power, cold end temperature and cold end temperature fluctuations, heat load on the pre-cooling stages, and heat flow to the warm radiator. We present initial test data compared to predictions based on previously performed component tests.
INTRODUCTION Planck is a European Space Agency (ESA) mission, whose main objective is to image the temperature anisotropy of the Cosmic Microwave Background (CMB) at high angular resolution. Planck will produce high sensitivity maps over 95% of the sky in a wide range of frequencies that have never before been studied at such high resolutions and sensitivities. The telescope will measure temperature fluctuations in the CMB with a precision of ~ 2 parts per million and an angular resolution ~ 10 arc-min. The analysis of these fluctuations will determine to a precision of few percent the fundamental cosmological parameters (Hubble constant, density of the Universe, cosmological constant etc.). Planck will carry two instruments: the High Frequency Instrument (HFI) and the Low Frequency Instrument (LFI). Together these instruments will observe and image the full sky in nine spectral bands between 30 and 857 GHz. Both the LFI and the HFI instrument sensors need to be cooled to cryogenic temperatures to optimize their signal to noise ratio. The detector cooling system has also to minimize the mechanical vibration to reduce the spurious signal generation on the ultra-sensitive detectors. Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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The LFI radiometers need a temperature of 20 K reached through a combination of passive cooling to about 50 K and active cooling using a hydrogen sorption cooler to reach lower temperatures. The HFI uses bolometers cooled to 100 mK through a combination of passive cooling (radiator at 50 K), the 20 K sorption cooler, a 4.5 K Mechanical Joule-Thomson cooler and a Benoit style open cycle dilution cooler (3He-4He). The use of an open cycle dilution cooler will limit the mission life to 1.5 years for the HFI. The description of the whole cooling chain has been previously provided by Collaudin et al.1 and by Wade2. The sorption cooler operates by compressing the refrigerant hydrogen through the compressor to the high pressure stabilization tanks which are maintained at 50 ATM. The refrigerant than travels from the tanks through a series of heat exchangers and radiators, which provide pre-cooling to approximately 50 K (see Fig. 1), through the JT expander at the Front End Unit (FEU). When the refrigerant (H2) expands through the JT flow restriction valve, hydrogen forms liquid droplets whose evaporation provides the cooling power in the liquid reservoirs. Each of the reservoirs is filled with a wicking material in order to retain the liquid in the reservoirs without gravity. The third reservoir is maintained above the hydrogen saturated vapor temperature, to wick and then evaporate any liquid that reaches it, thus providing an even gas flow back to the sorbent bed. The functional requirements of the Planck Sorption Cooler are summarized in Table 1.
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FACILITY FOR THE EBB COOLER TEST The flight sorption cooler design is currently being validated by operating an Engineering Bread Board unit (EBB) over the whole parameter space defined by the thermal, mechanical and electric interfaces provided by the spacecraft to the cooler. The EBB cooler sub-components, such as the compressor elements3,5, gas gap heat switches6,7, the check valves and the filters8, the JT valve and the liquid reservoirs9 have been previously tested to validate the component design but EBB operation is the first test of the whole cooler as a unit. Hence, it is providing information on the mutual interaction of the components and a validation of the cooler subsystem flow-down requirements assumed in the design2,10. It should be observed that, contrary to the case for the Planck flight coolers which will be thoroughly integrated using only weld joints, all the subsystems of the EBB cooler could be interchanged with similar units in case of a component failure or of a substantial redesign. The EBB cooler compressor is composed of six compressor elements with their own gas gap actuator as shown in Fig. 2. To verify the effect of the high pressure stabilization tank size on the performance, a system with multiple tanks was built to possibly connect to the high pressure manifold a 2, 3, 4, 5 liter volume. For the same reason two identical low pressure stabilization beds were connected to the low pressure manifold. Each compressor element is chilled by attaching it to
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a plate cooled independently by an external chiller. The chiller provides a 3-4 gpm (gallon per minute) flow to a common manifold inside the chamber from which the feed lines to the individual chiller plate leave (see Fig. 2). The chiller plate temperature fluctuations during cooler operation are within the present requirements based on a simulation of the spacecraft radiator. The whole sorption cooler was placed into a thermal vacuum chamber for the test, also providing a verification of the test facilities required to operate the sorption cooler. As previously described2, the sorption cooler has complicated mechanical and thermal interfaces with the spacecraft V-Groove shields (see Fig. 1) needed to efficiently pre-cool the high pressure hydrogen gas below its inversion temperature. The 100 K and 50 K pre-coolers designed for the EBB tests are shown in Fig. 3, where the two 50 K and the 100 K pre-coolers can be observed below the facility radiative shield containing the cold end. Each thermal interface where the cooler attached has been designed to provide a heat flow measurement at the. interface. The cold end assembly is the part above the interface “50 K shield” and is composed of the JT valve, two liquid reservoirs, a particle filter, a continuous tube-in-tube heat exchanger and a discrete heat exchanger (reported by Sirbi et al.9) that stabilizes the liquid-vapor interface in the cold-end assembly. The cold end assembly flight design is not complete and a separate design validation is currently ongoing at JPL. The main difference between the EBB and the flight sorption cooler is the flow allowed by the JT porous plug. The JT valve used for these tests was not characterized at operating temperatures and it was observed that the flow was reduced to 5.1 mg/s compared to the design value of 6.5 mg/ s. Based on this, different system performance is expected as shown in Table 2. The expected cooling power has been obtained by simply scaling the design cooling power by the ratio of the measured and design hydrogen flow. Total cooler power for this reduced flow does not scale with the flow since more than half of the power supplied to the compressor is used to pressurize the compressor element to 50 ATM while the rest of the power is used to supply the mass flow. For this reason the estimated power for the EBB cooler is not linearly increased, as for the cooling power, but is obtained by adding the heat up power, the gas gap actuator power (that did not change), and the new desorption power for a total of 390 W.
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INITIAL TEST RESULTS Initial testing of the EBB sorption cooler was performed in January 2002. This testing demonstrated a continuous sorption cooler capable of producing liquid hydrogen at a temperature less than 20 K with a cooling power greater than 1 W for a 50 K pre-cooling temperature. The tests conducted were also the first for the test facility and some uncertainty in the measurements are due to improper calibration of some instruments and by thermal design limitations of the facility during these first tests. The first test started with the coldest pre-cooler temperature at 86 K and cooling at a rate of 0.5 K/hour. This is shown in Figure 4, where the pre-cooler and cold-end temperatures are displayed as a function of time. After 6 hours, the pressure in the high pressure tanks reached the operating value of 50 ATM. The first accumulated liquid hydrogen in the liquid reservoir was observed 49 hours after the beginning of the test. The cold-end temperature stabilized at17.7 K with an applied power of 900 mW. After six hours the sorption cooler was shutdown to allow for the pre-cooler to reach 50 K. The large amount of noise observed in the pre-cooler temperature is due to faulty read-out electronics. When 50 K was reached on the pre-cooler, a second test was started. During this test the JT valve plugged when it reached a temperature of 37 K due to what we believe to be methane contamination. The plug was removed by heating the JT to a temperature of 80 K. During this period, the temperature of LR1 reached 50 K while the temperature of LR2 did not change. When the
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temperature of the JT reached 48 K the sorption compressor was restarted. Liquid was observed in the liquid reservoirs 5 hours and 20 minutes after the JT was observed to be plugged. The cooler produced 1.2 W (pre-cooling shield at 50 K) of cooling power at a temperature of 17.7 K on both LR1 (the HFI interface) and LR2 (the LFI interface). The temperature between the two liquid reservoirs differed by less than 40 mK.
CONCLUSIONS The EBB sorption cooler has provided first demonstrations of the design of the Planck 20 K sorption cooler by operation of a closed loop hydrogen sorption cycle. The results reported in the present paper are only the initial test and a fully detailed description of the EBB test results will be reported at the end of the complete test program. Expected contamination problems were observed during the test when the JT plugged before running smoothly probably due to a residual methane gas component generated by the compressor elements (Bowman et al.5). Currently, the cooler has been operated for almost 750 hours, and these tests are providing essential information for the full validation of the flight design.
ACKNOWLEDGMENTS This research was carried out by the Jet Propulsion Laboratory, California Institute of Technology under a contract with the National Aeronautics and Space Administration.
REFERENCES 1. Collaudin, B. and Passvogel, T., “The FIRST and PLANCK ‘Carrier’ Missions. Description of the Cryogenic Systems,” Cryogenics, vol. 39 (1999) pp 157-165. 2.
Wade, L.A. et al., “Hydrogen Sorption Cryocoolers for the PLANCK Mission,” Advances in Cryogenic Engineering, 45A, Kluwer Academic/Plenum, New York (2000), pp. 499-506.
3.
Paine, C.G., et al., “Planck Sorption Cooler Initial Compressor Element Performance Tests,” Cryocoolers 11, Kluwer Academic/Plenum Press, New York (2001), pp. 531-540.
4.
Pearson, D., et al., “Characterization and Lifecycle Testing of Hydride Compressor Elements for the Planck Sorption Cryocooler,” Advances in Cryogenic Engineering, Vol. 47, Am. Inst. Physics, New York (2002), pp. 1209-1216.
5.
Bowman, Jr., R.C., et al., “Evaluation of Hydride Compressor Elements for the Planck Sorption Cooler,” Cryoeoolers 12, Kluwer Academic/Plenum Publishers, New York (2003).
6.
Prina, M., et al., “Development of Gas Gap Heat Switch Actuator for the Planck Sorption Cryocooler,” Advances in Cryogenic Engineering, 45A, Kluwer Academic/Plenum Publishers, New York, 2000, pp. 553-560.
7.
Prina, M., Kulleck, J.G., and Bowman, Jr., R.C., “Assessment of Zr-V-Fe Getter Alloy for Gas-gap Heat Switches,”J. Alloys Comp,, vol. 330-332 (2002), pp. 886-891.
8.
Bowman, Jr., R.C., et al., “Performance, Reliability, and Life Issues for Components of the Planck Sorption Cooler,” Advances in Cryogenic Engineering, Vol. 47, edited by S. Breon, et al. (Am. Inst. Phys., New York, (2002) pp. 1260-1267.
9.
Sirbi, A., Bowman, Jr., R.C., Wade, L.A., Barber, D.S., “Cryogenic System Design for a Hydrogen Sorption Cooler,” Advances in Cryogenic Engineering, Vol. 47, Am. Inst. Physics, New York (2002), p. 1217.
10. Prina, M., et al., “Performance Prediction of the Planck Sorption Cooler and initial Validation”, Advances in Cryogenic Engineering, Vol. 47, Am. Inst. of Physics, New York (2002), pp. 1201-1208.
Construction and Operation of a 165 K Microcooler with a Sorption Compressor and a Micromachined Cold Stage J.F. Burger, H.J. Holland, H.J.M. ter Brake, M. Elwenspoek and H. Rogalla University of Twente, Faculty of Applied Physics, 7500 AE Enschede, The Netherlands
ABSTRACT This paper presents the integration and testing of a 165 K microcooler that operates with a sorption compressor and a micromachined cold stage. Attractive features of this combination are the lack of vibration and a long lifetime for a potentially very small cryocooler. The developed cold stage works with the vapor compression cycle and consists of a silicon condenser and flow restriction/evaporator, which are combined with two miniature glass-tube counterflow heat exchangers. The thermal compressor consists of five miniature cells filled with adsorption material that are sequenced between ad- and desorption – thus providing a continuous pressure difference. The system was tested with ethylene gas operating between 14 and 1.5 bar, and produced a cooling power of 200 mW around 170 K with an input power of about 20 W.
INTRODUCTION Sorption compressors are basically thermal compressors, which have no moving parts.1 Therefore, they have the potential of an extremely long lifetime and can operate with an absolute minimum vibration level. Furthermore, they can be scaled to small dimensions without deteriorating the efficiency. This paper describes a microcooler that applies a sorption compressor, which drives a miniature Linde-Hampson cold stage. The sorption compressor as well as the cold stage utilizes micromachined silicon components, by which we try to point out a new direction towards really small, reliable and integrated cryocoolers. We have shown that a number of essential cryocooler components can be fabricated on a very small scale and in a welldefined manner by means of MEMS technology, which brings the development of even smaller integrated cryocoolers into reach. An additional advantage of fabricating these structures in silicon technology is the possibility to closely integrating these coolers with electronic components and their cryogenic packaging. Furthermore, attractive thermal and fluidic scaling properties can be utilized that cannot be achieved with conventional fabrication technologies, e.g. enhanced heat fluxes in miniature heat exchangers, integrated capillary structures, integrated spring structures for micro-Stirling coolers, etc. Also, batch-fabrication of micromachined cryocoolers may lead to cost reduction. Further background information, as well as the thermodynamic modeling and development of the sub-elements for this cooler was described at previous meetings1-6; in this paper we will Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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describe the integration and testing of the complete cooler. In the following two sections first a summary of sorption cooler operation will be presented as well as a brief description of the developed cooler components.
SORPTION COOLER OPERATION Operation of a sorption compressor is based on the principle that large amounts of gas can be adsorbed on certain solids such as highly porous carbon. The amount of gas that is adsorbed is a function of temperature and pressure; it increases with a reduction of the temperature and an increase of the pressure. If a pressure cylinder is filled with a sorber material and gas is adsorbed at a relatively low temperature and pressure, then a high pressure can be created inside the closed vessel by an increase of the temperature of the sorber material. A constant gas flow out of the vessel can be maintained at a high pressure by further increase of the temperature until most of the gas is desorbed. A sorption compressor in its basic configuration consists of a combination of two sorption cells: one delivering high-pressure gas at a relatively high desorption temperature and another adsorbing low-pressure gas at a lower temperature. Ad- and desorption of these two cells is periodically swapped when they reach their saturated adsorption levels. A continuous pressure difference can be created from this intrinsic intermittent system by combining at least two pairs of sorption cells that are connected via check valves6 and that are sequenced out of phase. A schematic diagram of such configuration is given in Fig. 1, which shows a combination of a sorption compressor with a cold stage. The figure also shows a schematic plot of the timedevelopment of one complete cycle of a sorption cell; the plot shows the amount of gas adsorbed, the compressor temperature, the pressure and the mass flow. Compressed gas coming out of the compressor unit is cooled to the environmental temperature after which it is fed into the counterflow heat exchanger. A (thermoelectric) precooler may be applied to improve the system performance. Next, the compressed refrigerant is expanded in the JT valve to provide refrigeration in the evaporator. The low-pressure refrigerant then returns through the recuperative heat exchangers to the compressor unit.
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DEVELOPED COOLER COMPONENTS Sorption compressor cells Fig. 2a shows a photograph of one integrated sorption cell. It consists of an inner pressure cylinder with adsorption material (upper part in Fig. 2a), which is placed inside another cylinder that acts as a heat sink. Heating of the sorption cell is done with an electrical heater and cooling can be done by turning on a heat-switch between the sorption cell and the heat sink. Fig. 2 shows a compressor cell with a gas-gap heat switch between the two cylinders. Adjustment of the gas pressure in the gap between the cylinder and the heat sink can vary the heat conduction through the gap. This pressure adjustment can, for instance, be done with another tiny sorption pump7. In the experiments described in the nest sections, however, the gas-gap was not used as a heat switch, but instead adjusted to a fixed thermal conductance by filling the gas gap with xenon gas. Five instead of four sorption cells were used to allow the compressor cells more cooling-down time via the fixed (relatively low conducting) thermal link. This somewhat simpler configuration results in an efficiency of about 1% for our ethylene cooler. This efficiency is limited by the increased heat losses through the gas gap during heating of the sorption cells. It can be increased to 3% by application of the gas-gap heat switch.
Check valve unit Fig. 2b shows a photograph of a part of the check valve unit. One valve consists of a thick plate (boss) suspended by four thin springs6. The thin springs behave like single clamped beams,
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thus facilitating high deflections that are required to obtain a low pressure drop in the forward direction. When the valve is in forward direction, the gas is able to flow through the holes surrounding the springs. The entire valve construction, including the interfacing gas lines, is made out of two silicon wafers that are covered by two glass wafers, all bonded together. The design is such that the gas lines on both sides of the valve can cross in the upper and lower wafers. This is required for the construction of an integrated valve manifold that has only six connections to the outside world: four to the compressor cells and two to the cold stage. The valve seat is made out of the polished surface of the wafers, facilitating a perfect fit and alignment of both sides of the valve seat. Stiction of the fragile spring beams to the bottom wafer is prevented by etching a cavity in the bottom wafer under the beams. Stiction or bonding of the valve seat is prevented by application of a nitride coating. The maximum allowed deflection of the boss can be selected by choosing an appropriate thickness of the boss relative to the wafer thickness. A 4 µm sieve is constructed in the inflow line of the valve to trap possible contamination. The sieve is made of a row of pillars standing in the channel. Fluidic and mechanical modeling was applied to find the proper dimensions of the valve construction6.
Cold stage Fig. 2c shows the cold stage, which is made of three micromachined silicon components with two glass-tube counterfiow heat exchangers in between. All three silicon parts are constructed by fusion bonding of two 500 µm thick silicon wafers in which channels and spaces are etched by KOH etching. After processing and separating these silicon samples, the glass tube heat exchangers are glued into the samples, after which integration with a small vacuum flange follows. The applied glass tubes have inner/outer diameters of 0.25/0.36 mm and 0.53/0.67 mm, respectively. Two glass support tubes are added parallel to the two counterfiow heat exchangers to add more stability to the system. The left silicon part is called the ‘splitter’, and makes it possible to supply separate connection lines to the high and low pressure channels of the first counterflow heat exchanger. The middle part is the condenser, in which the high-pressure fluid is able to condense in a long meandering channel, which is etched in the silicon. From the condenser, the precooled highpressure fluid enters the restriction/evaporator (the right silicon part) where it flows through an etched channel to the entrance of the flow restriction, which typically consists of a 4 mm wide, 1 µm shallow channel with a length of about 3 mm. The low-pressure liquid that exits the flow restriction is collected in the liquid bath of the evaporator, which connects to the low-pressure return line of the second counterflow heat exchanger.
COOLER SET-UP Fig. 3 shows two photographs of the assembled cooler. To facilitate a simple connection and exchange of the sub-components, Swagelock fittings were used between most of the different elements of the cooler. An efficient heat sink with fan was directly mounted on the compressor cells, resulting in a 0.1 K/W thermal resistance to the room-temperature environment. The cold stage was mounted in a small vacuum chamber that was covered by a glass plate. This glass plate facilitates visual observation of the two-phase fluid flow in the glass-tube counterflow heat exchangers. Seven small pressure sensors8 were mounted in a connector-block, which interfaces the seven lines of the check-valve unit to the five compressor cells and the high and low pressure lines of the cold stage. Two flow sensors9 were used to measure the gas flows into and out of the cold stage. Labview software10 on a laptop PC with two data acquisition cards10 was used to measure the important parameters of the cooler, and control the heater input into the compressor cells. Different control algorithms for the compressor could easily be programmed and tested in Labview. Electrical connections of the compressor heaters, temperature sensors, pressure sensors, flow sensors, thermoelectric cooler and the heater on the cold stage were made via one shielded cable to a custom fabricated electronics board. In total 32
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signals were constantly monitored by the data acquisition electronics, while simultaneously the five input powers to the compressor heaters were controlled. Purging and filling of the system with clean (99.95%) ethylene gas was done via seven gas lines that were connected to the connector block in which also the pressure sensors are mounted. Seven instead of one or two pump lines were used because evacuation with one or two pump lines would require pumping via the check valves, which would result in a very poor vacuum. These gas lines were fabricated of soft copper and were pinched off after purging and filling of the system, resulting in a closed gas-system.
MEASUREMENTS Fig. 4 shows the start-up and subsequent cyclic operation of the five sorption cells, which results in cooling-down of the cold stage. The heaters of the compressor cells are used to measure the cell temperatures as well; the electronic switching of these heaters between heating and sensing causes the rapid variations in the measured input powers and temperatures. In the experiment of Fig. 4, a control algorithm was used that maintains the high pressure of the system at a constant value – by adjustment of the heater input of the cell that delivers gas at this high pressure. Furthermore, there is always one cell adsorbing the low-pressure gas. The other three cells are either building-up or lowering the pressure, or waiting for the next phase of the cycle. The algorithm switches to the following phase when the cell delivering high-pressure gas reaches a certain temperature-limit (i.e. 540 K in this measurement). For each cell, the cyclic operation that was schematically depicted in Fig. 1 can be distinguished. The ‘waiting’ periods that occur before the high-pressure gas flows out of the cell and before the low-pressure gas flows into the cell can be identified as the ‘horizontal’ periods in the compressor temperature and pressure graphs. In those ‘waiting’ periods, the cell pressure is controlled slightly below the system high pressure or above the system low pressure so that the corresponding check valve remains closed. As a consequence, no gas can flow out of or into the cell. At t = 2950 s, the thermoelectric precooler is started and temperature-controlled at 230 K, 2 K below the condensation temperature of ethylene at 14 bar. At t = 2990 s, the ethylene starts to condense in the condenser. From this moment onwards, the evaporator starts to cool more rapidly because of the increased cooling power of the liquid ethylene that now flows from the condenser to the evaporator. As long as the temperature of the evaporator is above the saturation temperature of ethylene at 14 bar (232 K), the produced liquid will evaporate upon reaching the silicon evaporator – thus providing cooling power. At t = 3550 s, the high-pressure fluid starts to enter the restriction as a liquid, which increases the mass flow because of changing fluid density
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and viscosity. The short reduction in the mass flow between t = 3700 s and 3750 s is caused by the filling of the evaporator with liquid ethylene. After t = 3750 s, the cooling power exceeds the applied thermal load and a two-phase fluid exits the evaporator. Capillary effects associated with this two-phase fluid are responsible for the variations in the outgoing mass flow. The larger flow peaks are caused by switching the low pressure side of the cold stage to the following sorption cell. The variation of the low temperature of the cold stage is directly related to the variation of the low pressure in the system via the Clausius-Clapeyron relation that governs the phasetransition in the evaporator. A control-algorithm was successfully tested (not shown in Fig. 4) to
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stabilize the pressure of the sorption cell adsorbing the low-pressure gas – and by that the low temperature of the cold stage.
CONCLUSIONS A very small cryocooler was fabricated and tested that combines a miniature micromachined cold stage with a thermal sorption compressor that operates without moving parts. These two combined concepts are useful for applications in which very small and integrated Cryocoolers are needed that operate without vibrations for a long period of time. The developed cold stage shows that essential cryocooler components of very small dimensions can be constructed using MEMS technologies. The described cooler shows the feasibility of the concepts using ethylene as the refrigerant that can produce temperatures below 170 K. Other temperatures can be reached with this type of cooling cycle by replacing ethylene with a different refrigerant and subsequent tailoring of the individual cooler components11,12.
ACKNOWLEDGEMENTS This research was supported by the Dutch Technology Foundation (STW) and the Research Institute MESA+ of the University of Twente.
REFERENCES 1.
Burger, J.F., Cryogenic Microcooling: A micromachined cold stage operating with a sorption compressor in a vapor compression cycle, Ph.D. Thesis, Twente University, The Netherlands (2001). 2. 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. 3. 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 (1999), pp. 553-564. 4. Burger, J.F., Holland, H.J., van Egmond, H., Elwenspoek, M., ter Brake, H.J.M., Rogalla, H., “Fast gas-gap heat switch for a microcooler”, Cryocoolers 10, Plenum Press, New York (1999), pp. 565574. 5. Burger, J.F., Holland, H.J., Seppenwoolde, J.H., Berenschot, J.W., ter Brake, H.J.M., Gardeniers, J.G.E., Elwenspoek, M., Rogalla, H., “165 K microcooler operating with a sorption compressor and a micromachined cold stage”, Cryocoolers 11, Plenum Press, New York (2001), pp. 551-560. 6. Burger, J.F., van der Wekken, M.C., Berenschot, E., Holland, H.J., ter Brake, H.J.M., Rogalla, H., Gardeniers, J.G.E. and Elwenspoek, M., “High pressure check valve for application in a miniature cryogenic sorption cooler”, Proc of IEEE MEMS 99 (1999). 7. Gardeniers, J.G.E., Burger, J.F., van Egmond, H., Holland, H.J., ter Brake, H.J.M. and Elwenspoek, M., “ZrNi thin films for fast reversible hydrogen pressure actuation”, Proc. of Aktuator 2000, Bremen, Germany (2000). 8. Kulite Semiconductor Products, Inc., One Willow Tree Road, Leonia, NJ 07605, USA. 9. Bronkhorst High-Tech B.V., Nijverheidsstraat 1 A, Ruurlo, The Netherlands. 10. National Instruments, Inc., 6504 Bridge Point Parkway, Austin, TX 78730-5039, USA. 11. Wade L.A. et at., Hydrogen sorption Cryocoolers for the PLANCK mission, Adv. in Cryogenic Engineering, vol. 45A (2000), pp. 499-506. 12. Burger, J.F., ter Brake, H.J.M., Rogalla, H., Linder, M., Vibration-free 5 K sorption cooler for ESA’s Darwin mission, Cryogenics, vol. 42, no. 2 (2002), pp. 97-108.
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Automated Closed-Cycle Cooling to 250 mK for the Polatron R.S. Bhatia, V.V. Hristov, B.C. Keating, A.E. Lange, P.V. Mason, B.J. Philhour, G. Sirbi & K.W. Yoon California Institute of Technology, Pasadena, California 91125, USA S.T. Chase Chase Research Ltd., Sheffield S7 1LB, United Kingdom
ABSTRACT We have integrated a 4 K mechanical cryocooler with a helium sorption refrigerator to achieve closed-cycle cooling to a base temperature of 239 mK. The cryocooler consists of two Gifford-McMahon stages which pre-cool a Joule-Thomson expansion stage to below the inversion temperature of This in turn provides cooling to below the 5.2 K critical temperature of The first two stages of the sorption cooler consist of a and a refrigerator, with the providing cooling to below the 3.3 K critical temperature of This subassembly buffers the heatload on the final refrigerator. We achieve an operating temperature of 250 mK at the expected 0.5 µW heat input from parasitic conduction plus the optical load. The duty cycle of the sorption cooler is 83 %. The sorption cooler is automatically cycled using digitally commanded power supplies. The system will allow for unattended cooldown from 300 K to 250 mK for cooling of the detectors on the Polatron, a ground-based receiver which will be fielded at the 5.5 metre telescope at the Owens Valley Radio Observatory.
INTRODUCTION The cosmic microwave background (CMB) radiation is the remnant signature on the sky of the Big Bang which occurred 15 billion years ago. We have designed and built the Polatron receiver to measure the undetected polarisation of the CMB, which can give information on the origin of gravity waves and on the ionisation history of the Universe. Because the 2.7 K CMB has an expected polarised level of only a few mK, we need extremely sensitive detectors which (for incoherent systems) require cryogenic cooling to ~ 0.3 K. Several cryogenic instruments with 0.3 K sorption refrigerators have been built, but none are suitable for unattended operation because they all rely on replenishment of liquid cryogens. The necessity to perform astronomical observations from remote, high sites such as the South Pole means that the supply of these liquid cryogens can be logistically demanding. It is therefore advantageous to cool an instrument using a completely closed cycle cooling system which can be automatically and remotely cycled. This design and operational philosophy has been adopted for the Polatron, which will be fielded at the Owens Valley Radio Observatory.
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POLATRON INSTRUMENT An assembly drawing of the instrument is shown in Figure 1. The incoming polarised radiation is modulated by a rotating quartz half-waveplate. A series of optical filters allow radiation only at 96±10 GHz to pass through into an orthomode transducer, which splits the radiation into two orthogonal linear polarisation components. Feedhorns couple each of these two polarised components to bolometric detectors. The cryogenic stages of the instrument are summarised in Figure 2 and consist of a three-stage helium cryocooler precooling a three stage helium sorption refrigerator. The work reported here is an update of the development of the cryogenic system reported in Bhatia et al 1 .
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CRYOCOOLER The Polatron uses a two-stage Gifford-McMahon (G-M) cryocooler to precool to below its inversion temperature. A third stage of the cryocooler uses Joule-Thomson (J-T) expansion to cool to below the critical temperature of for operation of the sorption cooler and to provide 4 K cooling of the optical filters. The cooler specification is a heatlift of 1 W at 4.2 K at the J-T tip. One important advantage of this cryocooler is its insensitivity to gravity, which is not the case for commercially available 4 K pulse tube coolers. Disadvantages of the cryocooler compared to a cryostat are the point cooling available at 4 K, the reduced cooling power, and the requirement to supply power (and cooling) for the compressors. Both the G-M and J-T cycles are inherently inefficient and the coolers have moving parts so that the reliability is reduced compared to a cryostat, although scheduled cryocooler maintenance consists solely of replacement of the J-T filters every 10 000 hours.
CRYOCOOLER – DETECTOR INTEGRATION ISSUES The mechanical cryocooler can potentially degrade the performance of the detectors in a number of ways. The cryocooler provides cooling of optical filters and radiation shields so that any variation in the temperature of these components leads to generation of instrumental noise. The cryocooler generates both conductive and emitted electromagnetic interference (EMI) and it is necessary to use a variety of techniques to reduce these sources of noise2. The detectors are electrically isolated from the cryocooler using kapton to reduce the susceptibility of the detectors to conducted EMI. This requirement for electrical isolation is at odds with the requirement for as high a thermal conductivity as possible. Because the Kapitza boundary thermal impedance3 scales as we have designed in the ability to perform this isolation at the 4 K cryocooler-
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sorption interface rather than the 0.25 K sorption-detector interface. The detectors are electrically biased at 225 Hz to avoid the 1/f noise associated with the JFET preamplifiers which comprise the first stage of the detector readout circuit. To reduce the mechanical vibration transmitted to the detectors, flexible thermal links made from OFHC copper are used at the cryocooler 4 K and 80 K stages. A pneumatically driven vibration isolator4 is used at the interface to the cryocooler vacuum shell to reduce transmitted vibration at 300 K. We have measured the acceleration at the detector stage using a cryogenic accelerometer. The acceleration levels in the vertical axis are shown in Figures 3 and 4 respectively for the signal bandwidths and the bandwidth about the bias frequency. The cryocooler harmonic at 2.4 Hz is designated as the first harmonic. The acceleration levels in the lateral axes have been shown to be comparable to the vertical levels.
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SORPTION REFRIGERATOR DESCRIPTION A schematic of the three stage sorption refrigerator is shown in Figure 5, and an assembly drawing in Figure 6. The detailed design and manufacture of the sorption refrigerator was undertaken by Chase Research. The cryocooler precools the condensation point to below the 5.1 K critical temperature of The sorption liquid is then pumped upon to provide a temporary heatsink at 0.85 K for condensation below the 3.2 K critical temperature of the two volumes of Finally, the liquefied in the Intercooler buffers and reduces the heatload on the final Ultracooler still which cools to 250 mK. The specifications for the Polatron sorption refrigerator design were as follows. The required cooling power at the Ultracooler still was 250 mK for 12 hours with 0.5 µW load (over and above parasitics from pump tube), and at the Intercooler still was 500 mK for 12 hours with 60 µW load (over and above parasitics from pump tube). It will be necessary to tilt the instrument at the telescope to make observations, particularly for calibration which can be done at relatively low elevation angles. Therefore, the orientation requirements were that the refrigerator should meet the above cooling specifications within 55 degrees of vertical when tipped in any direction. The duty cycle should be > 83 % (i.e. total cycle time should be less than 2 hours from start of cycling to final stable operation). The total allowable physical envelope was defined as 275 mm high with a semicircular footprint of radius 95 mm.
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Room temperature charge pressures for the refrigerator are 90 bar with 99.95% purity and 99.999% purity The Intercooler contains 11 S.T.P. litres of and 5 S.T.P. litres of The Ultracooler contains 2 S.T.P. litres of All three pumps use activated charcoal as the sorbent. Each pump is cooled using a gas gap heat switch filled with The pumps and heat switches are instrumented with metal film resistors for heaters and silicon diodes5 for thermometers. The Intercooler and Ultracooler stills and the condensation point are instrumented using germanium resistance thermometers6 and read out using an AC resistance bridge 7 . A heat exchanger is located between the intercooler still and the baseplate. It extracts the enthalpy of the cold gas flowing from the still to the baseplate and suppresses most of the parasitic load which would otherwise find its way up to the stills. To meet the cooling and hold time specifications with the refrigerator within 55 degrees of vertical when tipped in any direction, a fine copper mesh is silver soldered into both stills. The surface tension of small droplets of the liquid within the mesh contains the liquid against the pull of gravity. The sorption refrigerator is cycled with the instrument pointing vertically upwards. This is because the pump tubes are then vertical, and gravity can be used to help condensation and to prevent convective thermal loading on the stills. In addition, the minimum optical loading on the cold stages of the instrument will be achieved with the instrument pointing at the zenith such that there is minimal atmospheric optical power loading. Successful operation of the cooler depends critically on condensing enough liquid to serve as a heatsink for condensation of the liquid. Correspondingly, new features of this system are the use of a separate thermal link (Link 1 in Figure 5) directly from the condensation point to the 4 K thermal ground, together with a separate intermediate baseplate. The intermediate baseplate intercepts the majority of the enthalpy of the gas desorbed from each pump. The sum of the residual gas enthalpy and the heat of condensation is then sufficiently low that the condensation point temperature in each case does not rise significantly above its steadystate value. This is particularly important with the relatively large pump and the limited cooling power of the cryocooler. Link 2 thermally connects the intermediate baseplate to the 4 K baseplate, and Link 3 the 4 K baseplate to the cryocooler 4 K thermal ground. The thermal conductivity of Link 1 to thermal ground is 39 mW resulting in a peak power dissipation at the condensation point of 12 mW.
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SORPTION REFRIGERATOR CYCLE PROCEDURE Figure 7 plots the temperatures of the three pumps and three heatswitches during the cycle, and Figure 8 shows the temperatures at the beginning of each of the different steps of the cycle. By cycling both pumps together instead of individually, we allow much of the heat of desorption and enthalpy of the Intercooler to be dissipated only to the thermal ground (the 4 K stage of the cryocooler) rather than partially to a cold Ultracooler still. Further, by heating the pumps before condensing the the stills have more time to cool back towards 4 K before condensation is begun. This means that less liquid is boiled away in cooling the stills. The simultaneous condensation also leads to a significantly reduced refrigerator cycle time. The Intercooler pumps are heated to 45 K and the Ultracooler pump heated to 40 K. During condensation, the cryocooler J-T tip warms to 4.9 K and the intermediate baseplate warms to 10 K. The pump is maintained at 45 K for ~ 15 minutes then allowed to cool by switching off power to the pump heater and powering the heat switch heater. However, it is vital to ensure that this heat switch does not transition into its on state too soon. This would dissipate too much of the enthalpy of the hot gas and pump immediately to thermal ground, and boil away all that has been liquified by the J-T system. The heat switch temperature is therefore stabilised about 15 K in an intermediate state to cool the pump slowly enough that the condensation point temperature can be kept below 4.3 K. Once the stills have cooled to below 0.85 K and the liquid has all evaporated, the supply of current to the pumps is stopped and the two pump heat switches are switched fully on. Because the ultimate hold time is limited by the Intercooler liquid running out before the Ultracooler liquid, we switch on the Ultracooler pump heatswitch a few minutes before the Intercooler pump heatswitch. The Ultracooler still cools first and therefore precools the Intercooler still further, increasing the residual which remains in the Intercooler still once final cooldown has been achieved. Also, the Intercooler pump has greater thermal mass and therefore takes longer to cool than the Ultracooler pump. Once the both pumps have cooled to below 17 K, pumping on the liquid commences and the required base temperatures in the stills are achieved. Astronomical
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observations are made once the temperature of the detectors is stable at a level of (with the help of active PID temperature control), and observations are continued until the liquid in the Intercooler still has all evaporated. The pumps are then heated and the cycle repeated.
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SORPTION REFRIGERATOR PERFORMANCE The Ultracooler still has achieved a base temperature of 239 mK with a hold time below 260 mK of 37 hours. The corresponding Intercooler temperature was 340 mK, and it is the Intercooler which limited the hold time of the refrigerator. The cycle time to cool back down to 260 mK was 2.5 hours, resulting in a duty cycle of 83 %. The cycle was repeatable and easy to perform, where the only procedures that needed to be carefully monitored were operation of the heat switch in an intermediate state, and allowing only just enough time for the liquid to evaporate before commencing condensation. A load line for the Ultracooler still is shown in Figure 9. The still temperature for the expected load of 0.5 µW was 250 mK. For a load of 2 µW the still temperature was 265 mK, a temperature sufficiently low to permit operation of the bolometers with very high sensitivity and fast speed of response. This cycling procedure was performed using programmable power supplies and a Lab VIEW1 data acquisition system. Testing ofthe refrigerator has shown that the cycling procedure is sufficiently robust that the cycling can be automatically performed using a combination of LabVIEW, a GPIB bus and the programmable power supplies. The cryocooler itself does not need attention during cooldown to 4 K with the exception of throttling of the J-T orifice which can be done mechanically and remotely commanded. The Polatron therefore has the important advantage ofhaving an automatic cooldown from 300 K to 0.25 K and automatic cycling between 4 and 0.25 K, which will facilitate astronomical observations needing long integration times. Finally, tilting of the instrument to 55 degrees of vertical does not change the final base temperature which is achieved.
SUMMARY We have designed and built a cryogenic system which achieves a base temperature of 239 mK and a heatlift at 250 mK of 0.5 µW without the use of expendable liquid cryogens. The cryogenic system is adapted for use in the Polatron instrument for observations at submillimetre wavelengths of the polarisation of the cosmic microwave background.
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ACKNOWLEDGEMENTS This work was partly supported by NASA Innovative Research Grant NAG5-3465 and NASA NAG5-6573 for US involvement in Planck. The construction of the Polatron is funded by a Caltech/ JPL President’s Fund Grant PF-414, NASA Grant NAG5-6573 and NSF Grant AST-9900868. We thank Kathy Deniston for administrative support.
REFERENCES 1. Bhatia, R.S., et al., “Closed-cycle Cooling of Infrared Detectors for the Polatron,” Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2001), pp. 577-586. 2. Bhatia, R.S., et al., “The Susceptibility of Incoherent Detector Systems to Cryocooler Microphonics,” Cryogenics 39 8 (1999), pp. 701-715. 3. Lounasmaa, O.V., Experimental Principles and Methods Below 1K, Academic Press (1974). 4. National Electrostatics Corp., Middleton, Wisconsin, USA. 5. Type DT470 Silicon Diode, Lake Shore Cryotronics, Inc., Westerville, OH 43082, USA. 6. Type GRT- 200A Germanium Resistance Temperature Sensors, Lake Shore Cryotronics, Inc., Westerville, OH 43082, USA.
7. AVS-47 AC Resistance Bridge, RV-Elektroniikka Oy Picowatt, Vantaa, Finland, Distributed by Oxford Instruments Ltd., Witney, Oxfordshire, UK. 8. National Instruments Corporation, 11500 N. Mopac Expressway Austin, TX 78759-3504, USA.
Progress in the Development of a Continuous Adiabatic Demagnetization Refrigerator P.J. Shirron1, E.R. Canavan1, M.J. DiPirro1, J. Francis1, M. Jackson1, T.T. King2, and J.G. Tuttle1 NASA/Goddard Space Flight Center 1 Code 552, 2Code 541 Greenbelt, MD 20771
ABSTRACT In this paper we summarize recent progress in the development of an adiabatic demagnetization refrigerator (ADR) that operates continuously at sub-Kelvin temperatures. The ADR uses multiple stages, one of which cools a load while the others periodically transfer heat to a heat sink. The architecture is very flexible, allowing stages to be added at the low end to achieve lower operating temperature, or at the high end to increase the heat rejection temperature. The present design goals are to achieve high cooling power (on the order of 10 µW) at 50 mK or lower, and heat rejection at 6 K or higher. The latter reflects the performance goals of various cryocoolers being developed for use in space through NASA’s Advanced Cryocooler Technology Development Program1. Over the past year we assembled and tested a 3-stage ADR that used a superfluid helium bath as a heat sink. We have also recently assembled a 4-stage ADR that can operate with a 4.2 K helium bath. Details of the design, operation and performance of these systems are discussed.
INTRODUCTION To meet the growing need for more capable low temperature refrigerators for space astronomy missions, we are developing an adiabatic demagnetization refrigerator that operates continuously at low temperature2,3. As solid state systems, ADRs are well-suited to the space environment. In fact, they are the only type of refrigerator that can cool to temperatures below 200 mK presently qualified for flight. The drawback to traditional single-shot ADRs, due to the low energy density of magnetic refrigerants, is their relatively large mass and limited cooling power. The continuous ADR (CADR) under development improves the situation dramatically. Because the internal recycling operations that transfer heat up to the heat sink do not interrupt cooling at the cold stage, the CADR can be operated on a short cycle time (~1 hour). By avoiding the need to store heat for prolonged periods of time, the amount of refrigerant needed (per stage) is reduced by 1-2 orders of magnitude. Consequently the CADR achieves an order of magnitude higher cooling power per unit mass. We project that a system with a cooling power of 10 µW at 50 mK using a 6-10 K heat sink will have a cold mass less than 10 kg. This is actually a 5-stage ADR which includes a continuous 1 K stage as a thermal shield for telescope or detector components, or as a stable platform for detector amplifiers such as SQUIDs. Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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CADR CONCEPT The CADR, shown schematically in Figure 1, uses multiple stages connected in series to a heat sink. Each stage consists a superconducting magnet, a “salt pill” containing the magnetic refrigerant, and a heat switch linking to it the next higher stage or, in the case of the highest stage, to the heat sink. The lowest stage (stage 1, also called the continuous stage) is connected to the load to be cooled and is magnetized or demagnetized to maintain constant temperature. The upper stages shuttle back and forth in temperature, picking up heat from the next lower stage and transferring it to the next higher stage. By performing these transfer operations sequentially, heat can be cascaded from stage 1 up to the heat sink without disturbing stage 1 ’s temperature. The number of stages needed to span a given operating temperature range depends, roughly, on the ratio of the heat sink and operating temperatures. Although there are trades one can make to expand the operating of any one ADR stage, heat switch and magnet considerations tend to set a practical limit of about a factor of 5 in temperature. Thus a CADR operating between 50 mK and 6 K will require one continuous stage and three additional stages, with upper temperature limits on the order of 0.25 K, 1.25 K, and 6 K. The operating range of the ADR can be expanded, either to achieve lower base temperature or higher heat reject capability, simply by adding stages to the cold or warm ends. In a system that is already optimized, no re-optimization of the existing stages is necessary to do this. This has allowed us to proceed from an initial demonstration of heat transfer between two ADR stages to construct a 3-stage and, more recently, a 4-stage CADR simply by adding components. At present, we are just beginning performance tests of the 4-stage system, therefore this paper we focus on the performance of the 3-stage CADR and conclude with an overview of the 4-stage system’s design and expected performance.
3-STAGE CADR The relevant parameters for the 3-stage CADR are given in Table 1. Stages 1 and 2 use chrome potassium alum (CPA) refrigerant because its entropy capacity is higher than ferric ammonium alum (FAA) below about 60 mK. Its low ordering temperature of 9 mK will also allow the ADR to operate in the 15-20 mK range. The third stage, using ferric ammonium alum (FAA), is an engineering unit ADR produced for the X-Ray Spectrometer mission4. It has much
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larger cooling capacity than necessary, but this proved very useful for optimizing the recycling parameters. For the 4-stage CADR, this stage is replaced by a much smaller CPA stage. Figure 2 is a picture of the 3-stage CADR in the test dewar. The third stage is located inside the dewar’s helium tank, below the bottom of the picture. All of the ADR’s magnets are shielded to minimize fringing fields. Fringing fields are problematic for two reasons. First, they can affect the temperature of the refrigerant in an adjacent stage, making temperature control of the ADR more difficult. Second, some experimental components, like SQUIDs, are extremely sensitive to magnetic fields and require low background levels in order for their own shielding to be effective. In most cases, reducing fringing fields into the millitesla range is sufficient. From a mass perspective, passive ferromagnetic shielding is the most effective option for the small magnets used on the first and second stages. The third stage, however, is relatively large and achieves better attenuation with active shield coils on the magnet. Measurements with a hall probe show all fringing fields are reduced below 1 mT in the experiment space. OPERATION The CADR has two operational modes. The first involves cooling down from the heat sink temperature and establishing temperature control. The second is the periodic cycling of the upper stages to maintain constant temperature. To begin, all stages are magnetized to full field with the heat switches turned on. Starting with the third, each stage is sequentially demagnetized (after turning the appropriate heat switch off) to the low end of its operating range. The rate is not critical but should be slow enough to efficiently cool the lower stages. In the final state, the continuous stage is fully recycled and is actively cooling the load. The process takes less than 1 hour. Control is then passed to a routine that automatically recycles each stage as needed. Recycling consists of demagnetizing the adjacent upper stage to a lower temperature and closing the heat switch. The temperature controller automatically magnetizes the lower stage as the upper stage absorbs its heat. When the magnetic field reaches an upper threshold, the heat switch is opened, and the upper is magnetized to the high end of its range to reject the heat to the next stage.
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PERFORMANCE The performance areas of greatest interest for a low temperature cooler are operating temperature, temperature stability, cooling power and efficiency. These have been characterized extensively for the 3-stage CADR. For operation with a mechanical cryocooler, heat rejection rates are also very important. We present this for a typical cycle, but emphasize that we can reduce the peak heat flux by regulating the upper stage at a lower temperature while it is recycled. In principle we can accommodate any heat flux limitation, but beyond a point it will impact the ADR’s cooling power.
Operating Temperature To date, the 3-stage CADR has operated continuously at temperatures down to 35 mK and up to 200 mK. Higher temperature operation is possible, but because of minor constraints imposed by the passive gas-gap heat switch, some software changes would be necessary in order to preserve high efficiency. These will be implemented as the need arises.
Cooling Power The CADR’s cooling power shown in Table 1. We define cooling power as the maximum sustainable heat load that can be applied to the first stage. Although the CADR can tolerate larger momentary loads, if a higher load persists, the first stage eventually runs out of cooling capacity and loses temperature control. The time period for this to occur will depend on the excess heat load. As an example, for a 1 µW excess at 50 mK, the ADR will continue to function for 10 hours or more. But we also note that the system requires a similar amount of time to regain its cooling reserve after the heat drops below the cooling power limit. The cooling powers in Table 2 have improved considerably over the last year. This is due to improvements in certain components and in the control software. The two most important component changes were to increase the thermal conductances of the first stage thermal bus and of the superconducting heat switch linking the first and second stages. These resulted in more efficient heat transfer and greater cooling capacity of both stages. The control software was also tuned to accelerate the recycling process. This translates directly to higher cooling power by allowing the ADR to reject heat more frequently. As a result, we are close to meeting our original cooling power goal of 10 µW at 50 mK. It also has enabled us to achieve continuous operation at 35 mK with good cooling power. As expected, the cooling power is approximately linear in temperature above 50 mK. A linear relationship results from the nearly constant entropy capacity of the first stage as a function of temperature (using a fixed magnetic field range), and therefore a linearly increasing cooling capacity. The non-linearity below 50 mK is due to several factors, the most important of which appear to be the rapidly increasing thermal boundary resistance between the salt and the thermal bus (which significantly reduces heat transfer efficiency), and the diminishing entropy capacity of CPA as its ordering temperature is approached.
Thermodynamic Efficiency The CADR’s efficiency (relative to Carnot) was measured at both 50 and 100 mK for the
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applied heat loads given in Table 2. The measurements include all parasitic effects (such as eddy current heating, magnet and magnetic shield hysteresis, heat switch and suspension system parasitics, and thermal gradients within the salt pills), but does not include heat that is dissipated in the third stage magnet or heat needed to turn on the third stage heat switch. We also do not account for dissipation in the room temperature electronics. The reason is that these components are not yet representative of a final, optimized refrigerator, and the dissipation is consequently much higher man we eventually expect. With this in mind, we regard the measurements of 50% efficiency at 100 mK and 14% at 50 mK as being the right order of magnitude for a complete system. Granted the efficiency will decrease when the neglected terms are added, but this will be offset to some degree by other changes such thermally anchoring all magnets and shields to the heat sink rather than to the warmest ADR stage. We attribute the marked drop in efficiency at 50 mK to the higher thermal boundary resistance between the thermal bus and the refrigerant in the first and second stages. Boundary resistances generally vary as making the thermal gradients with the salt pills an order of magnitude larger at 50 mK than at 100 mK. The solution is to use a more finely divided thermal bus to increase the contact area, but we are presently at the limit of the wire EDM technique used to manufacture our thermal buses.
Temperature Stability Temperature control tends to be more of an issue for the CADR than for conventional ADRs because of the fairly large changes in heat flow that occur during cycling. Considerable effort has gone into developing control routines that minimize temperature disturbances. For example, the superconducting heat switch that links the first and second stages is turned on and off only when the temperature gradient across it is close to zero. It is also important for the temperature controllers to have fairly high bandwidth (~10 Hz) so they can respond quickly to heat impulses. With have used both analog and digital controllers, and find both are able to control the base temperature of the CADR to the noise of the temperature readout. Figure 3 shows the temperature and magnetic field of the continuous stage during one cycle conducted at 100 mK. The magnetic field has the characteristic sawtooth pattern as heat is alternately absorbed and rejected, and the temperature is constant to within the 8 µK rms noise level of the resistance bridge. Although this level of control can be achieved, it is not necessarily typical, especially at lower temperature. There, higher thermal boundary resistances lead to long thermal time constants within the salt pills, and it becomes difficult to recycle the first stage disturbing its temperature. As seen in Figure 3, with the ADR operating at 50 mK, the temperature experiences spikes by about 200 µK. for 30-50 seconds whenever the superconducting heat switch is turned on or off. These can be reduced, ultimately to negligible levels, by slowing down the recycling operations, but this will also reduces the ADR’s cooling power. Conversely, one can obtain higher cooling
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power if temperature stability can be sacrificed. For space astronomy missions, detector temperature is invariably a critical parameter. Stability requirements for microcalorimeters, which measure photon energies through changes in detector temperature, are typically at the level of a few µK. Therefore we are still working to improve stability without sacrificing cooling power, and will focus on both software and component modifications to do so. For software modifications, we intend to implement feedforward control using the second stage temperature and ramp rates as a way to anticipate changes in heat flow as the first stage is recycled.
Heat Reject Rate A plot of typical heat rejection rate for two complete cycles is shown in Figure 4. In this case, the third stage was regulated at 1.4 K as it rejected heat to the superfluid helium bath, which itself was at 1.19 K. With a heat switch conductance of about 150 mW/K, the peak heat load was approximately 35 mW. The duration was less than 100 seconds, resulting in a transfer of 2.2 J every 2600 seconds. The above figures can be scaled to estimate the heat loads for different heat sinks. For example, for a 6 K heat sink, the total heat rejected per cycle will be about 11 J. If this occurred in the same time frame, the peak heat load would be 175 mW. This could overwhelm the cooling power of some cryocoolers and would therefore need to be reduced. This can be done two ways. The first is simply to regulate the third stage closer to the heat sink temperature as it rejects heat (i.e. impose a smaller gradient across the heat switch). The second option would be to reduce the size and conductance of the heat switch. The choice will depend largely on the tolerance of the cryocooler to fluctuations in heat load that will be present in some measure due to fluctuations in the ADR’s temperature. In either case, slowing the heat rejection may lengthen the ADR’s overall cycle time, and thereby reduce the cooling power. This kind of trade will have to be made once the characteristics of both systems are known.
4-STAGE CADR The 4-stage CADR is a direct extension of the 3-stage system. Conceptually it involved adding a fourth stage that could operate between 1 and about 5 K. In reality, the existing FAA stage became the fourth stage, and a new CPA stage was inserted as the third stage. The parameters of all four stages are given in Table 3. We continue to use the XRS ADR as a fourth stage more for convenience than for its operating capabilities. In the 1 -5 K temperature range, there are many rare earth compounds, such as gadolinium gallium garnet (GGG), that have much higher entropy density and lower field requirements than FAA. However, the arrangement of the
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FAA stage in the test dewar allowed it to be quickly integrated into a 4-stage system, and, more importantly, the system’s performance is not compromised by its use. That will not hold true if a cryocooler operating at 6 K or higher is use as a heat sink, and it will be important in the near term to replace the FAA stage with one having higher temperature capability. In fact, the focus of our R&D effort is now shifting toward the development and characterization of new refrigerants tailored for the 1-10 K range5. A picture of the 4-stage ADR is shown in Figure 5. The only significant change to the configuration was to thermally anchor the magnets and shields for stages 2 through 4 to the helium bath, rather than to an ADR stage. Although the suspension systems will increase the parasitic heat loads on the lower stages, there is a net benefit to absorbing the hysteresis heat in these components at higher temperature. We therefore expect to obtain overall efficiencies and cooling powers comparable to those measured for the 3-stage ADR.
SUMMARY We have constructed a 3-stage CADR that operates continuously at temperatures down to 35 mK with high cooling power and high efficiency. A superfluid helium bath at 1.2 K serves as the heat sink. The system can achieve 8 µK rms temperature stability throughout its cycling operations, and cooling powers of 1.5 µW at 35 mK rising to 15 µW at 100 mK. We have
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recently reconfigured the ADR to add a fourth stage which will increase its heat rejection capability to above 4.2 K and allow it to operate with an unpumped helium bath. Testing is in progress. We are also working on designs for suspension systems and magnetic shielding so that the system can be packaged more compactly. One layout for a four-stage cooler operating with a 4.2 K heat sink fits within a 2-liter volume and weighs less than 7 kg.
ACKNOWLEDGMENT This work has been supported by NASA’s Cross-Enterprise Technology Development Program and by Goddard Space Flight Center’s Commercial Technology Development Program.
REFERENCES 1.
Ross, R.G., and Boyle, R.F., “NASA Space Cryocooler Programs - An Overview,” presented at the 12th International Cryocooler Conference, submitted for publication in Cryocoolers 12, Kluwer Academic/Plenum Publishers, New York (2003).
2.
Shirron, P.J., Canavan, E.R., DiPirro, M.J., Tuttle, J.G., and Yeager, C.J., “A Multi-Stage Continuous Duty Adiabatic Demagnetization Refrigerator,” in Adv. Cryo. Eng. 45, Kluwer Academic/Plenum Publishers, New York (2000), pp. 1629-1638.
3.
Shirron, P.J., Canavan, E.R., DiPirro, M.J., Jackson, M., King, T., Panek, J., and Tuttle, J.G., “A Compact, High-Performance Continuous Magnetic Refrigerator for Space Missions,” Cryogenics, vol. 41 (2002), pp. 789-795.
4.
Serlemitsos, A.T., SanSebastian, M., Kunes, E., “Design of a Spaceworthy Adiabatic Demagnetization Refrigerator,” Cryogenics, vol. 32 (1992), pp. 117
5.
King, T.T., Rowlett, B.A., Ramirez, R.A., Shirron, P.J., Canavan, E.R., DiPirro, M.J., Panek, J.S., Tuttle, J.G., Shull, R.D., Fry, R.A., “Rare-Earth Garnets and Perovskites for Space-Based ADR Cooling at High T and Low H”, Adv. Cryo. Eng. 47, Kluwer Academic/Plenum Publishers, New York (2002).
Sub-Kelvin Mechanical Coolers A. Ravex1, P. Hernandez1 and L. Duband2 1
Air Liquide – DTA (Advanced Technology Division) Sassenage, 38360, France 2 Atomic Energy Committee, Low Temperature Division, CEA/SBT Grenoble, France
ABSTRACT Double stage pulse tube crycoolers (PTC) capable of lifting about 0.5W of cooling power at 4K are commercially available. Single stage and sorption coolers have been developed for both ground and space applications. A precooling temperature below 2.5K is required to operate a sorption cooler efficiently. This is commonly achieved with a pumped liquid bath. However for some applications (for example bolometers cooling in astrophysics) the vibrations induced by the pumping system are problematic. In addition the fraction of liquid lost during the bath pump down significantly affects the bath hold time. To overcome this difficulty, a double stage sorption cooler has been developed by CEA/SBT that can be operated either from an atmospheric bath or from a 4K PTC, and which provides temperature down to 260 mK. This paper presents the performance obtained with such a double stage sorption cooler using a 4 K CRYOMECH PT405 Pulse tube for precooling. A similar approach is under development with a dilution fridge being coupled with a 4K pulse tube.
INTRODUCTION Helium sorption coolers provide a wide range of heat lift capability at Sub-Kelvin temperatures [1]. They have no moving parts, are vibrationless and can be designed to be self contained and compact with a high duty cycle efficiency. These coolers rely on the capability of porous materials to adsorb or release a gas when cyclically cooled or heated. Using this physical process, one can design a compressor/pump that, by managing the gas pressure in a closed system, can condense liquid at some appropriate location and then perform an evaporative pumping on the liquid bath to reduce its temperature. Consequently, it requires a precooling stage at a temperature lower than the helium liquid-vapor transition, i.e. 3 K, with enough cooling power. This is in general a pumped helium bath, which is a potential drawback, since it leads to a reduced autonomy for the cryostat and induced vibration generated by the mechanical pumps. To overcome this problem CEA/SBT has developed a double stage helium sorption cooler that can be operated from any precooling stage at a temperature below 5K. Experimental results on this cooler using an helium bath have already been reported [2]. We present in this paper the results obtained with the association of this cooler and a 4K commercial pulse tube cooler developed by CRYOMECH [Syracuse, NY USA]. Air Liquide (AL) is distributing these 4K PTC in Europe and CEA-SBT has transferred this technology to AL for industrial manufacturing and commercialization of the overall system referred to as “SoCool” [3]. Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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DOUBLE STAGE SORPTION COOLER DESCRIPTION Figure 1 shows the general architecture of the double stage unit, while the specifications are listed in Table 1. Each stage comprises a sorption pump, a condenser and an evaporator, with the exception of the stage (2nd stage) which also include a heat exchanger. This heat exchanger is thermally coupled to the condenser (1st stage). The evaporator is thermally connected to nd the condenser (2 stage). In addition, to increase the thermal efficiency of the cooler, both stages feature a gas gap heat switch [4]. These switches thermally couple or decouple the sorption pumps to the precooling stage. The cooler does not require any mechanical or vacuum connections and is fully controlled by electrical heaters. The details of operation of this cooler have been reported elsewhere [2]. One specific feature of this architecture is the presence of two separate evaporators to allow for more efficient operation as described later. Once the cooler has been thermally coupled to the pulse tube cold head and cooled down to below 5 K operation can begin. Two operating modes are possible : In mode A both stages have been recycled – the stage cooling capacity has been used to cycle the 3He stage and the cooler cold tip drops to its ultimate temperature where it remains stable until the liquid is in turn exhausted. In mode B, a full recycling has been carried out as in mode A, but once the stage has run out, it is recycled again. This stage provides then a heat intercept for the stage which consequently operates at a significantly lower temperature. Once the evaporator cold tip has run out of liquid helium the cooler can be recycled.
EXPERIMENTAL RESULTS As mentioned before, the double stage sorption cooler has been previously characterized on a 4.2 K helium bath cryostat and on a 4K pulse tube cooler developed at CEA-SBT [5]. The double
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stage cooler has then been integrated on a CRYOMECH PT405 pulse tube cold head. Two pictures of the overall system (excluding the compressor) are shown in Fig 2. The initial cooldown from room temperature down to 4 K takes approximately 3.5 hours (see Fig. 3). During this initial cooldown both sorption pumps temperature will decrease at a slower rate than the condenser temperature, and it is thus possible to perform a first cycle of the sorption cooler. In this case after about 4.5 hours since startup one can reach 300 mK at the cold tip. Subsequent recycling of the double stage sorption cooler takes about 1.5 hours. A particular feature of the pulse tube is its availability to operate at a temperature significantly lower than 4.2 K (ultimate temperature of about 2.4 K), leading to a reduction of the parasitic loads on the sorption cooler. As a consequence and contrary to the 4.2 K helium bath, once both stages have been recycled, the parasitic load on the stage is substantially lower and it is possible to use it to cycle the stage and to take advantage of the remaining liquid to directly access Mode B (both stages cold).
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Once the sorption cooler has been recycled and is operating at low temperature, the pulse tube cold head drops to about 2.4 K and the sorption cooler ultimate temperatures obtained in mode A and B are respectively 265 mK and 291 mK. We have noticed that the vibrations generated by the 4K pulse tube contribute to a slight increase of the parasitic load on the evaporator cold tip. This is clearly seen if the pulse tube is turned off for a couple of minutes and the temperature of the evaporator drops very quickly from 265.5 mK down to 255.7 mK. The cooling power curves for both modes are reported on Fig. 4. This measurement has been repeated and the reason why the two curves cross each other is not clearly understood and is being investigated. The lower operating temperature of the pulse tube cold head leads to an increase of autonomy for the sorption cooler. Indeed, as mentioned before, after a standard recycling (both stages recycled once) there is enough liquid to recycle the stage and then directly access mode B (both stages operating). In effect we have measured 26.5 hours at 265 mK in mode B followed by over 28 hours at 291 mK ( stage out of liquid helium – mode A). This is substantially better than with a helium bath where mode B can only be accessed after an additional recycling of the stage, to then provide typical hold time of 8 hours at 260 mK followed by 13 hours at 308 mK.
DILUTION FRIDGE AND 4K PULSE TUBE COUPLING The coupling of a dilution fridge with a 4K Pulse Tube has been recently published [6]. AL is presently developing a commercial product based on a similar architecture. As shown on Fig. 5, the design is already performed. For the cold part (T<4K) we are using our standard MiniDil dilution fridge design and components. The expected ultimate temperature is about 20mK with a cooling power of about 80 µW at 100mK. The prototype is under assembly and optimization tests will be performed in the next coming months.
CONCLUSION A double stage sorption cooler has been coupled with a 4K PT405 CRYOMECH pulse tube cooler as the precooling stage. This system is the first cryogen-free system able to cool down three orders of magnitude in temperature. It provides ultimate temperatures down to 265 mK and has a typical cooling power of 20 µW at 290 mK. Hold times in excess of 2 days have been measured for a service temperature of 300 mK.
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A drive electronics unit is under development at CEA-SBT and in a near future the complete system will be fully automated. This technology has been transferred to Air Liquide for industrial manufacturing and commercialization. A dilution fridge is presently being coupled to the same 4K Pulse Tube. This cryogen free system will provide 4 orders of magnitude in temperature with a continuous operation.
ACKNOWLEDGMENT. AL is grateful to CEA/SBT for the Sorption Cooler technology transfer and Lionel Duband and Laurent Clerc for their helpful advise during industrialization.
REFERENCES 1.
Duband, L. and Collaudin, B. Cryogenics, 39 (1999), p. 659.
2.
Duband, L. Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2001), p. 561.
3.
For commercial and technical informations contact:
[email protected].
4.
Duband, L. Cryocoolers 8, Plenum Press, New York (1995), p. 731.
5.
Poncet JM., Ravex A. and Charles I. Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2001), p. 229.
6.
Uhlig, K. Cryogenics 42 (2002), p. 73.
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Preliminary Performance of a Superfluid Compressor F.K. Miller and J.G. Brisson Cryogenics Engineering Laboratory Massachusetts Institute of Technology Cambridge, MA 02140
ABSTRACT A preliminary test of a compressor developed to compress superfluid mixture in a superfluid Joule-Thomson refrigerator has been conducted using gas at 4.2 Kelvin. The results of this test and a description of the compressor design are presented in this paper.
INTRODUCTION The superfluid Joule-Thomson refrigerator, which uses a mixture as the working fluid, is shown in Figure 1. A high concentration mixture is discharged from the compressor that is kept at 1.2 K by a evaporation refrigerator. The fluid enters a counterflow heat exchanger where it is cooled to low temperature by the counterflowing stream. The component then enters the throttle valve and “expands” to a low concentration mixture. After this, the enters the low temperature heat exchanger where the mixture absorbs heat from the low temperature reservoir. The enters the counterflow heat exchanger where it is warmed by the opposing flow. Finally, the enters the compressor where it is recompressed to high concentration and
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rejects the waste heat of the cycle to a 1.2 K refrigerator. A superleak at low temperature maintains a zero chemical potential difference across the Joule-Thomson valve.1 Models for the SJTR have been developed to predict the performance of a superfluid Joule-Thomson refrigerator.2 A SJTR with a concentration ratio of 10:1, a compressor temperature of 1.2 K and a molar flow rate equal to 0.3 mmoles/s will have a cooling power on the order of 0.1 mW at 0.7 K.1
Here we discuss a bellows-piston compressor with a superleak-bypassed suction valve that should give a concentration ratio equal to 10:1 with molar flow rate greater than 0.3 mmoles/s when cycled at a rate of two cycles per minute. This is followed by a discussion of the results of a preliminary compressor test at 4.2 K.
COMPRESSOR DESIGN Figure 2 shows the four steps of the compressor cycle. In the suction stroke the suction valve is opened and the compressor volume is increased to draw dilute mixture into the compressor. During the compression part of the cycle, the suction valve is closed and the piston compresses the while the component flows out through the superleak and into the dilute phase reservoir. The discharge valve then opens and the concentrated mixture is swept out of the compressor. The discharge valve closes and the remaining is expanded back to its original concentration as the flows back through the superleak, diluting the mixture. Then the suction valve opens as the compressor volume continues to increase taking in the next charge of helium mixture.
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The compressor, shown in Figure 3, consists of a copper compressor body, brass top and bottom flanges that are connected to the body with edge welded stainless steel bellows, three brass stay bars connecting the two flanges, low dissipation suction and discharge valves and three 1 cm diameter superleaks through the body of the compressor connecting the two chambers created by the bellows. The internal volume of the bottom bellows is the compression chamber and the internal volume of the top bellows is the dilute phase reservoir. Both bellows have a 6.5 cm outer diameter and a 4.4 cm inner diameter with a stroke length equal to 1.8 cm. A volume block is attached to the bottom flange so that the dead volume in the compression chamber is minimized; therefore, the concentration ratio is maximized. The superleaks allow the superfluid component of the mixture to flow freely around the suction valve when it is closed. The top and the bottom flanges are connected by three brass stay bars, so that a change in the volume of one chamber is balanced by an equal but opposite change in the other chamber. Therefore, the dilute phase reservoir expands during the compression process to make room for the superfluid component that flows through the superleaks from the compression chamber. The suction and discharge valves are based on a design we have previously developed which includes opposed bellows, a polished stainless steel stem and a PCTFE seat to minimize sealing forces and energy dissipation.3 The thermal dissipation for each valve is estimated to be 2 µJ/cycle and leak rates are less than 0.3 µmoles/s for the operating pressures in the SJTR. This leak rate is negligible compared to the anticipated 300 µmoles/s circulation rate for in the SJTR. The overall dimensions of the compressor are 8.6 cm in diameter by 8.9 cm tall. The valve assemblies are separate from the compressor and are 4.1 cm in diameter by 6.4 cm tall.
TEST APPARATUS Figure 4 shows the test apparatus we constructed to measure the compression ratio of the compressor. In addition to the compressor and valves described above the apparatus consists of a suction and discharge volume, two calibrated carbon resistors that measure temperature and two capacitance type pressure gauges that measure the pressure on each side of the compressor. For the preliminary test we pumped helium gas at 4.2 Kelvin from a volume connected to the suction side into a volume connected to the discharge valve of the compressor. The suction volume was and was maintained at 4.2 Kelvin. The discharge volume was 70 liters and was maintained at room temperature. Both the volume and the 70 liter volume at room temperature were filled with helium gas at a pressure equal to 320 torr at the beginning of the test. As
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the compressor operated at a speed of one cycle per minute, the pressure on the suction side dropped and the pressure on the discharge side rose. We calculated the compression ratio by dividing the pressure on the discharge side by the pressure on the suction side. The final steady state ratio of these pressures was the maximum compression ratio for the compressor. Because we used helium gas at 4.2 Kelvin for this test, there was no superfluid in the system; and hence, there was no flow through the superleaks during the test.
PRELIMINARY RESULTS Figure 5 shows the results of the preliminary test. The pressure on each side of the compressor was 320 torr when the test was started. The pressure on the suction side dropped off smoothly until a steady state pressure of 89 torr was reached after 45 minutes.
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The pressure on the discharge side rose rather steeply initially, due to a flow restriction between the discharge valve and the 70-liter volume at room temperature. Since the pressure was measured right after the discharge valve at 4.2 Kelvin and the discharge valve was connected to the discharge volume by a small capillary there was a significant pressure difference between pressure we were measuring and the pressure in the discharge volume. When the compressor was first started, the flow rate was large so that this effect was significant. As the pressure on the suction side dropped the flow rate became smaller and this effect became negligible. The final steady state pressure in the discharge volume was 410 torr. The final steady state compression ratio was 4.6:1. This compression ratio is high enough to run a superfluid Joule-Thomson refrigerator, but an analysis of the compressor volume and the dead volumes predicted a compression ratio greater than 10:1. We looked for possible sources for the lack of compressor performance and found that the discharge valve seat was damaged and had a significant leak rate when the valve was closed. We expect to get a compression ratio closer to the predicted value once the valve seat is replaced.
CONCLUSIONS We have designed and constructed a cold compressor that is suitable for use in a superfluid Joule-Thomson refrigerator. Our calculations indicate that this compressor is capable of producing a compression ratio greater than 10:1. However, preliminary test results using helium gas at 4.2 Kelvin showed that the maximum compression ratio was 4.6:1. We determined that the discharge valve was leaking during the test due to a damaged valve seat. We expect to achieve a higher compression ratio when we test the compressor after the valve is repaired.
ACKNOWLEDGMENT We gratefully acknowledge the National Science Foundation for their support of this work.
REFERENCES 1.
Brisson, J.G., “Superfluid Joule-Thomson Refrigeration, a New Concept for Cooling Below 2 Kelvin,” J. Low Temp. Phys.,vol. 110, no. 1/2 (2000), p. 151.
2. Radebaugh, R., “Thermodynamic Properties of 3He-4He Solutions with Applications to the 3He-4He Dilution Refrigerator,” NBS Technical Note No. 362 (1967), p. 6. 3. Miller, F.K. and Brisson, J.G., “Development of a Low-Dissipation Valve for Use in a Cold-Cycle Dilution Refrigerator,” Cryogenics, vol. 39 (1999), pp. 859-863.
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Preliminary Experimental Results Using a Three-Stage Superfluid Stirling Refrigerator C. Phillips and J.G. Brisson Cryogenic Engineering Lab Department of Mechanical Engineering Massachusetts Institute of Technology Cambridge, MA 02139
ABSTRACT This paper describes the first operation three-stage superfluid Stirling refrigerator (SSR) and reports on its preliminary experimental performance. Previous SSR’s were single or twostage machines. The three-stage SSR has a total internal volume of and uses two and one Kapton heat exchangers. Operating from a high temperature of 1.07 K, and with a 3.0% mixture, this SSR achieves a low temperature of 338 mK. The single successful operation of this machine did not surpass the best low temperature results of previous two-stage SSR’s. This lower performance is attributed to a low level helium leak into the vacuum insulation space.
INTRODUCTION The superfluid Stirling refrigerator (SSR) is a Stirling cycle refrigerator that uses as the working fluid to cool to sub-Kelvin temperatures. The basic components of a single stage Stirling refrigerator are a hot (compressor) piston and a cold (expander) piston connected by a regenerator. The cyclic compression and expansion of the ideal gas within these pistons pumps heat from the cold temperature reservoir to the high temperature reservoir. For temperatures below 1 K, the component of the mixture behaves as an ideal gas in an inert background of superfluid Superleak bypasses in each piston allow the superfluid component to flow freely through the pistons while the is expanded and compressed within the piston cylinders. Kotsubo and Swift demonstrated the first single stage SSR in 1990.1,2 In 1992, Brisson and Swift further developed and improved the single stage SSR performance by using a recuperative SSR design.3-6 In the latter design, two refrigerators are operated 180 degrees out of phase with each other and a counterflow heat exchanger is used as the regenerator. This first recuperator was made of CuNi tubes. Brisson and Swift achieved a low temperature of 296 mK while operating from a high temperature of 1.05 K. Using the same machine, Watanabe, Swift, and Brisson later reached 168 mK while operating from a compressor temperature of 383 mK.7
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In 1997 a new single stage SSR was built by Patel and Brisson which used a counterflow heat exchanger recuperator manufactured from plastic that delivered substantially higher cooling power. Using a small plastic heat exchanger, they achieved an ultimate temperature of 344 mK from a high temperature of 1.0 K.8 Using a large plastic heat exchanger, they achieved a low temperature of 291 mK.9 Patel and Brisson also developed a two-stage SSR and using a large upper plastic recuperator and a small lower recuperator, the SSR achieved a low temperature of 248 mK.10 In 2000, Patel and Brisson proposed that a three-stage SSR could improve upon the performance of the two-stage SSR and potentially cool to 100 mK.10 As shown in Fig. 1, a threestage SSR consists of a single stage SSR thermally connected to the bottom platform of a twostage SSR. A three-stage SSR is comprised of a single compressor platform, two expander platforms, and a single compressor and expander platform. This paper describes the first three-stage SSR and reports its preliminary experimental performance. This three stage SSR has a total internal volume of and uses two Kapton heat exchangers having devoted to recuperative heat transfer and one Kapton heat
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exchangers having devoted to recuperative heat transfer. Operating from a high temperature of 1.07 Kelvin and with a 3.0% mixture, this SSR achieves a low temperature of 338 mK. We attribute this rather high ultimate temperature to a low level helium leak into the vacuum insulating space around the SSR.
DESCRIPTION OF THE THREE-STAGE SSR Figure 1 shows a schematic of the three-stage SSR. This refrigerator uses the counterflow recuperator configuration. This SSR consists of four isothermal platforms; the hot, first and second intermediate and cold platforms are connected by Kapton recuperators as shown. The compressor pistons on the hot platform are held at approximately 1 Kelvin by a evaporation refrigerator. The third stage of the SSR is thermally connected to the second stage of the SSR through the second intermediate platform. However, there are no fluid connections between the third stage and the second stage of the SSR. The third stage is filled separately from the first and second stages. The hot, first and second intermediate, and the cold platform of the SSR are made of solid blocks of OFHC copper on which the pistons are mounted. The pistons are made with edge welded stainless steel bellows, 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, for the first intermediate platform pistons, for the second intermediate platform (second stage), for the second intermediate platform (third stage) and for the cold platform. The hot platform pistons are rigidly connected together and driven sinusoidally using a push rod from a room temperature drive. The first and second intermediate (second stage) platform pistons are similarly connected and driven together using a common push rod. The third stage compressor pistons (on the second intermediate platform) are rigidly connected to each other. The cold platform pistons are also rigidly connected. These two sets of third stage pistons are independently driven by two push rods actuated from room temperature. Within each set of pistons, there are superleaks made from porous Vycor glass, that 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 first intermediate platform, the superleaks are three Vycor cylinders 10.63 cm in length with diameters of 0.74 cm. In the second intermediate platform (second stage), the superleaks are three Vycor cylinders 15.16 cm in length with diameters of 0.74 cm. On the second intermediate platform (third stage) the superleaks are four Vycor cylinders 13.09 cm in length with diameters of 0.74 cm. On the cold platform, the superleaks are three Vycor cylinders 12.06 cm in length with diameters of 0.73 cm. The total volume of the Vycor glass in the first and second stage of the SSR is The total volume of the Vycor glass in the third stage is Since 28% of the Vycor glass is void space, the glass contributes to the mixture volume of the first and second stage of the SSR and to the mixture volume of the third stage. The that diffuses into this volume does not participate in the operation of SSR. Within each piston platform, there are also isothermal heat exchangers made from nested OFHC copper cylinders press fit into the piston platforms. A 76 mm gap exists between the inner walls of an outer cylinder and outer walls of an inner cylinder. At the top of each cylinder is a flow distributor 0.635 mm deep and 0.317 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 65.94 cm2. Each half of the first intermediate 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 second intermediate (first and second stage) platform contains four cylinders 6.07 cm in length providing a total heat transfer area of The diameters of the cylinders are 4.43 cm, 3.91 cm, 3.39 cm, and 2.88 cm, respectively.
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Each half of the second intermediate (third stage) platform contains four cylinders 5.02 cm in length providing a total heat transfer area of The diameters of the four cylinders are 4.94 cm, 4.43 cm, 3.79 cm, and 3.16 cm, respectively. Each half of the cold piston platform has seven cylinders 4.77 cm in length providing a total heat transfer area of The diameters of the seven cylinders are 5.83 cm, 5.32 cm, 4.82 cm, 4.30 cm, 3.80 cm, 3.29 cm, and 2.78 cm, respectively. The recuperators used in this SSR are of a plastic design type. The recuperative portion consists of alternating layers of 127 mm Kapton film and 25.4 mm Kapton film glued together using Stycast 1266. Each 127 mm layer has five passages 2.38 mm in width and 20 cm in length. The recuperators from the hot platform to the first intermediate and from the first intermediate to the second intermediate platforms (first and second stage of the SSR) are large recuperators with forty 127 mm layers and thirty-nine 25.4 mm layers. Each large recuperator has a total volume of of which is dedicated to recuperative heat transfer. The recuperator from the second intermediate to the cold platform is a small recuperator with ten 127 mm layers and nine 25.4 mm layers. The total volume of the small recuperator is of which was dedicated to recuperative heat transfer. Calibrated ruthenium oxide, germanium, and Speer resistor thermometers mounted on the outside of the piston platform are used to monitor the temperature. The precision of the ruthenium oxide and germanium thermometers, which were used on the first and second intermediate and the cold platform, temperature measurements are ± 0.67 mK at 1.0 K and ± 1.02 mK at 350 mK. The total volumes of the SSR’s are given minus the void spaces and the volume of the fill lines because these volumes are inactive during the operation of the SSR. The total volume of the first and second stage is The total volume of the third stage SSR is As can be seen in Fig. 1, a total of four fill lines run from room temperature to the SSR. The two fill lines to the two halves of the first and second stage SSR were designed to be sealed at low temperature by valves mounted on the hot platform. Due to technical difficulties, the valves in the fill lines to the two-stage SSR were removed. The flow resistance of the fill lines, which have an inner diameter of 0.1 mm and have a length of approximately a meter between the hot piston platform and the platform at 4.2 K, was sufficiently high to prevent significant helium flow in the fill lines during the operation of the SSR. The two fill lines into each half of the third stage are sealed at low temperature by valves mounted on the second intermediate platform. The valves are also actuated manually from room temperature and act to prevent the mixture from moving up and down the fill capillaries during SSR operation. These valves were designed and tested to be superfluid tight in case it became necessary to pump out the fill lines to prevent heat loss. Although not shown in Fig. 1, the third stage capillary lines are wrapped around the evaporation refrigerator to help condense the mixture when it is loaded into the SSR third stage. Each fill line to SSR third stage passes through a void volume that is thermally pinned to the evaporation refrigerator. These void volumes insure that there is enough in the fill lines to prevent the heat flush effect from flushing all the to the second intermediate platform. If this were allowed to occur, a slug of high-thermal-conductivity pure would extend from the hot platform to the second intermediate platform in the capillary, essentially thermally “shorting out” the first and second stages of the SSR.
PROCEDURE AND RESULTS The SSR was prepared for operation by first cooling the refrigerator to 1.07 K, then centering the pistons on each platform to ensure equal volumes of working fluid in each SSR half. The first and second stages were filled first with a 3.0 % mixture. The two-stage SSR was test operated with a hot piston stroke of 1.03 cm ( volume displacement and clearance volume) and first/second intermediate piston stroke of 0.96 cm ( and volume displaced, and clearance volume).
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The operation of the two-stage was paused while the third stage was filled with a 3.0 % mixture. The fill lines to the third stage of the SSR were then closed and the SSR was operated a hot piston stroke of 1.03 cm ( volume displacement and clearance volume) and (second stage) first/second intermediate piston stroke of 0.96 cm ( and volume displaced, and clearance volume). The third stage was operated with a (third stage) second intermediate piston stroke of 1.03 cm ( volume displaced, clearance volume) and a cold piston stroke of 1.03 cm ( volume displaced, clearance volume). Combinations of two-stage cycle periods and third stage cycle periods were explored and their effect on the cold piston temperature. The temperature was measured by recording the minimum temperature of the platform during the cycle. In Fig. 2, the cold piston temperature is plotted against the cycle period of the cold piston for different cycle periods of the two-stage SSR. Effective operation of the third stage SSR required it to have a cycle period approximately ten times longer than the two-stage SSR. In Fig. 3, the cold piston temperature and intermediate temperature are plotted for each of the two-stage cycle periods. The slow third-stage cycle period, small variance in the second intermediate piston temperature and high cold to second intermediate piston temperature ratio indicate that the thirdstage recuperator is undersized.
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The data collected provides far from a complete map of the third stage behavior. In this experimental run, we were hampered by three major technical problems. The first was a very small superleak from the SSR into the insulating vacuum. This caused an ever increasing heat leak to the SSR cold platform. After a few days of operation, the heat leak got so large even the evaporation refrigerator could not keep up with the heat load. There was also some uncertainty in how balanced the mass of the was on each side of the SSR. An imbalance in the concentrations on each side of the SSR leads to imperfect recuperation, reducing the net cooling power of the SSR. Ordinarily, in situ electric switches provide fixed reference points that are used to center the pistons prior to the introduction of the helium mixture. Four out of the five sets of pistons had to be centered without the reference to the electric switches. Finally, the heat rejection in the hot piston may have been compromised by helium mixture flow in the fill lines to the two-stage SSR.
CONCLUSIONS This paper describes the first operational three-stage SSR and reports its preliminary experimental performance. The three-stage SSR has a total internal volume of and uses two and one Kapton heat exchangers. Operating from a high temperature of 1.07 K, and with a 3.0% mixture, this SSR achieves a low temperature of 338 mK. We believe that the refrigerator’s performance was limited by a heat leak due to a small helium leak to the vacuum insulation space. Previous work showed that SSR’s could achieve temperatures of 170 mK when combined with a evaporation refrigerator. This work SSR’s can be thermally chained. Consequently, we fully expect that a three-stage SSR should be able to achieve temperatures of the order of 100 mK.
REFERENCES 1. V. Kotsubo and G.W. Swift, “Superfluid Stirling refrigerator: A new method of cooling below 1 Kelvin,” In Proceeding of Sixth International Cryocoolers Conference, vol. 2, Bethesda Maryland (1991), David Taylor Research Laboratory, pp. 59-70.
2. V. Kotsubo and G.W. Swift, “Superfluid Stirling-Cycle refrigeration below 1 Kelvin,” J. of Low Temperature Physics, 83 (1991), pp. 217-224.
3. J.G. Brisson, V. Kotsubo, and G. W. Swift, “The superfluid Stirling refrigerator, a new method of cooling below 0.5 Kelvin,” Physica B, :45-46 (1994), pp. 194-196.
4. J.G. Brisson and G.W. Swift, “A recuperative superfluid Stirling refrigerator,” Adv. in Cryogenic Engineering, 39B (1994), Plenum Publishing Corp., New York, p. 1393.
5. J.G. Brisson and G.W. Swift, “Measurements and modeling of a recuperator for a superfluid Stirling refrigerator,” Cryogenics, 31 (1994), pp. 971-982.
6. J.G. Brisson and G.W. Swift, “High temperature cooling power of the superfluid Stirling refrigerator,” J. of Low Temperature Physics, 98(3/4), (1995), pp. 141-157.
7. A. Watanabe, G.W. Swift, and J. G. Brisson, “Measurements with a recuperative superfluid Stirling refrigerator,” Adv. in Cryogenic Engineering, 41 (1996), Plenum Press, New York, p. 1527.
8. A.B. Patel and J. G. Brisson, “Experimental performance of a single stage superfluid Stirling refrigerator using a small plastic recuperator,” J. of Low Temperature Physics, vol. 111 (1998), pp. 201-212.
9. A.B. Patel and J.G. Brisson, “Experimental evaluation of a single stage superfluid Stirling refrigerator using a large plastic recuperator,” J. Low Temp. Phys., 118 (2000), pp. 189-206.
10. A.B. Patel and J.G. Brisson, “Progress in Two-Stage Superfluid Stirling Refrigeration,” Adv. in Cryogenic Engineering, 45 (2000), Plenum Press, New York, pp. 1647-1652.
Dielectric Mirror Leakage and Its Effects on Optical Cryocooling G. Mills, J. Fleming, Z. Wei Ball Aerospace and Technologies Corp Boulder, CO USA J. Turner-Valle Optical Engineering Associates Longmont, CO USA
ABSTRACT Optical cooling using anti-stokes fluorescence in solids and liquids appears to have several advantages over more conventional techniques and has been the topic of recent analysis and experimental work by numerous organizations. Significant cooling of as much as 65 K below ambient temperature has been reported for isolated samples of cooling material. In this paper, we address the problem of thermally connecting the optical cooling element to the cooling load, without having the load heated by absorbed fluorescence. To date, it has been assumed that the load could be shielded from the fluorescence by the use of high performance dielectric mirror stacks in series with the thermal connection. We have done extensive analysis of these mirror stacks, using commercially available software, and have found that significant leakage occurs at large angles of incidence. Since the fluorescence is emitted from the cooling material isotropically, the leakage and resulting absorption at the load can be large enough to negate the cooling effect. We have investigated several variations in mirror design and have concluded that the leakage appears to be inherent to all dielectric mirror stacks and cannot be significantly mitigated through mirror design. We have formulated a solution to this problem and incorporated the solution into our laboratory optical cryocooler.
INTRODUCTION The basic principle of cooling by anti-Stokes fluorescence was suggested as early as 19291, but it was not until 1995 that the actual cooling of a solid was first demonstrated by Epstein et al. at Los Alamos National Laboratory (LANL) using Yb doped Zirconium Fluoride (Yb:ZBLAN) glass.2,3 In 1996, Clark and Rumbles reported cooling in a dye solution of rhodamine 101 and ethanol.4 A collaborative effort by LANL and Ball Aerospace resulted in an isolated cylinder of Yb:ZBLAN cooling 48 °C below the ambient temperature.5 Gosnell has reported cooling of 65 °C in a Yb:ZBLAN fiber.6 The fundamental refrigeration cycle of fluorescent cooling is simple. In the case of the Yb: ZBLAN material, the presence of the internal electric fields of the host ZBLAN material cause the ground and first excited states of the ion to be split into multilevel manifolds as shown in Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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Fig. 1. A photon from a laser tuned appropriately will be absorbed only by an ion that has been thermally excited to the highest level of the ground-state manifold, and will promote that ion to the lowest level of the excited-state manifold. When that ion decays radiatively, it can fall to any of the four ground-state levels. On average the outgoing fluorescent photon will therefore carry slightly more energy than the pump photon absorbed. By selectively “picking off” the “hottest” ions, this process depletes the population of the highest ground-state level. Thermal equilibrium is reestablished when another ion is promoted to that level by absorbing a phonon from the host material. The absorption of this phonon constitutes the refrigeration. In summary, a dopant ion absorbs a pump photon and the photon is re-emitted slightly bluer (higher energy). This energy difference comes from thermal vibrations (phonons) of host material. The simplest implementation of a cryocooler based on this principle is a simple Yb:ZBLAN cylinder (cooling element) with high-reflectivity dielectric mirrors deposited on the ends as shown in Fig. 2. The pump beam is introduced through a small feed hole in one mirror, and then bounces back and forth until it is absorbed. A key feature of this arrangement is that the pump light is confined to a nearly parallel beam, while the fluorescence is emitted randomly into steradians. This makes it possible to allow the fluorescence to escape while trapping the pump light inside. The fluorescent photons that are nearly parallel to the pump beam are also trapped. They are reabsorbed and then simply try again to escape with a small and calculable degradation to the overall efficiency. In previous work7, we performed a detailed design study of a complete cryocooler to verify feasibility and to determine the advantages and disadvantages of this technology. We chose to design an optical cooler for cooling a generic focal plane. Figure 3 shows our design concept for an optical cryocooler detector dewar. The cooling element is bonded directly to the focal plane structure in order to absorb the heat. The fluorescence
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is absorbed by a heat sink with high absorptivity at this wavelength. The cooling element and focal plane are supported by a folded tube of low thermal conductivity material, in a manner similar to dewars that have been used for focal planes cooled by mechanical cryocoolers. Note that the optical cooling element and heat sink are small compared to the structure that is needed to support a focal plane at cryogenic temperatures. An optical fiber is aligned and focused to transmit the light from the laser into the cooling element. For the 400 mW at 80 K cooler shown, a 20 W diode laser package would be required. This would be a package made up of a number of individual diode laser modules, plus optical fiber coupling, with dimensions of approximately 14 cm by 40 cm by 5 cm and a mass of 1.4 kg. It could be located remotely from the dewar to optimize heat transfer from the laser modules, and reduce EMI or magnetic fields, at the focal plane or for other system reasons.
DIELECTRIC MIRROR LEAKAGE In the work done to date in optical cooling, it has been assumed that the load could be shielded from the fluorescence by the use of dielectric mirror stacks in series with the thermal connection as is shown in Fig. 3. In the laboratory, we observed fluorescence coming through the dielectric mirror at large escape angles. To understand this, we analyzed the performance of mirror stacks as a function of incident angle, using commercially available thin film software.8 Dielectric mirrors are most commonly used (e.g. laser cavities) with the incident light on the vacuum or air side and with near normal angles of incidence. Figure 4 shows that in this case the performance of the mirror is quite high until 50 degrees angle of incidence where it begins to fall off for the P polarization. The S polarization is not affected by the angle. The situation is different when the light is incident on the mirror from the high-index substrate side. Figure 5 shows the result for a dielectric mirror in the arrangement required for optical cooling, where the incident medium is the ZBLAN glass cooling element. The mirror is in contact with a YAG substrate, which acts as the thermal conductor to the load. This results in much greater mirror leakage for both polarizations. The fluorescence is emitted within the cooling material isotropically and will be incident on the mirror at all angles. The mirror leakage show in Fig. 5 and the resulting absorption of energy at the load can be large enough to negate the cooling effect. The implications of these results were significant and unexpected enough that we confirmed them using a commercially available ray trace software package9. As can be clearly seen in Fig. 6 the results are identical to those in Fig. 5.
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In an effort to mitigate the mirror leakage, a layer of silver was added to the mirror stack. The results are shown in Fig. 7. Including an optically thick layer of silver on the exit face of the quarter wave dielectric stack does not reflect all of the light leaked at oblique angles of incidence; instead all of the light transmitted by the dielectric coating is absorbed by the silver layer, making the layer essentially “black.” This occurs because the dielectric multilayer acts as an anti-reflection coating on the silver, allowing more of the light to penetrate the silver layer where it is efficiently absorbed. In another effort to mitigate the mirror leakage, stack materials with higher refractive index were considered. For the visible-NIR spectral region, the greatest available dielectric material refractive index contrast comes from zinc sulfide and magnesium fluoride. With dielectric mirrors a large refractive index contrast produces a large region of high reflectance as a function of incident angle. Note that even the index contrast of the ZnS and is insufficient to reflect all of the Sand P-polarized light between 40 and 90 degrees at 975 nm as shown in Fig. 8. This behavior is typical of dielectric mirrors in the visible-NIR spectral region because of the limited refractive indices of dielectric materials available in these spectral regions. We also investigated the omni-directional dielectric mirror designs suggested by Mansuripur10. We found that this approach would allow for high reflectance over broad angles of incidence at a single wavelength. However, to extend this to wavelength range of the fluorescence that creates the cooling effect requires layer materials with very low dispersion, which currently do not exist.
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EFFECT ON OPTICAL COOLER PERFORMANCE The effect of mirror leakage on optical cooling performance is significant. Using the ray trace software, we have calculated that 27% of the light emitted by the cooling element leaks through a typical mirror stack and might be incident on the cooling load. Since the available heat lift ranges from 1 to 6% of the fluorescence, depending on temperature, the energy from the leakage has the potential to reduce or even completely negate the cooling effect.
CONCLUSIONS Based on our design and analysis, we conclude that optical cryocooling is a feasible method for cooling focal planes and has a distinct niche in extended solid-state cooling to those temperatures that cannot be achieved efficiently (or at all) by thermoelectric coolers (TECs). We have analyzed the dielectric mirror stacks, which are required for such a cooler using commercially available software, and have found that significant leakage occurs at large angles of incidence. We have investigated several variations in mirror design and have concluded that the leakage appears to be inherent to all dielectric mirror stacks and cannot be significantly mitigated through mirror design. This leakage has the potential to negate the cooling effect. We have formulated a solution to this problem and are currently testing it in our laboratory at Ball Aerospace.
ACKNOWLEDGMENT This work was performed as part of NASA contract NAS1-00115.
REFERENCES 1.
Pringsheim, P., Z. Phys 57, (1929), p. 739.
2.
Epstein, R.I., M.I. Buchwald, B.C. Edwards, T.R. Gosnell and C.E. Mugan, “Observations of LaserInduced Fluorescent Cooling of a Solid,” Nature 377, 500 (1995).
3.
Epstein, R.I., et al., “Fluorescent Refrigeration,” U.S. Patent. No. 5,447,032 (University of California).
4.
Clark, J.L., G. Rumbles, Phys. Rev. Letters, 76, 2037 (1996).
5.
Edwards, B.C., J.E. Anderson, R.I. Epstein, G.L. Mills, and A.J. Mord, “Demonstration of a SolidState Optical Cooler: An Approach to Cryogenic Refrigeration,” Journal of Applied Physics, 86 (1999).
6.
Gosnell, T.R. “Laser Cooling of a Solid by 65 K Starting from Room Temperature,” Optics Letters, vol. 24, no. 15 (1999).
7.
Mills, G.L., Mord, A.J., Slaymaker. P. A., “Design and Predicted Performance of an Optical Cryocooler for a Focal Plane Application,” Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2001), pp. 613-620.
8.
“The Essential Macleod,” version 8.3, Copyright Thin Film Center Inc. 1995-2002, Tucson, AZ.
9.
“ASAP” version 7.0.6, Copyright Breault Research Organization, Inc., 1982-2001, Tucson, AZ.
10. Mansuripur, M. “Omni-Directional Dielectric Mirrors,” Optics & Photonic News, Sept. 2001.
Advanced Components for Cryogenic Integration D. Bugby, B. Marland, C. Stouffer, and E. Kroliczek Swales Aerospace Beltsville, MD 20705
ABSTRACT This paper describes the development status of four advanced components that help solve important problems in cryogenic integration. The four components are: (1) an across-gimbal cryogenic loop heat pipe (CLHP); (2) a miniaturized CLHP; (3) a differential thermal expansion (DTE) cryogenic thermal switch (CTSW); and (4) a dual-volume cryogenic thermal storage unit (CTSU). The across-gimbal CLHP, which is baselined for operation from 80-100 K with nitrogen, provides a low weight, low torque, high conductance solution for gimbaled cryogenic systems wishing to position one or more cryocoolers off-gimbal. The miniaturized CLHP, which is baselined for operation near 35 K with neon, combines localized thermal transport, flexibility, and thermal switching into one device that can be directly mounted to both the cryocooler cold head and the cryogenic component. The DTE-CTSW, which was designed and successfully tested in a previous program using a stainless steel support tube and beryllium end-pieces, was redesigned with a polymer support rod and high-purity aluminum end-pieces to improve performance, manufacturability and structural stiffness while still providing a highly miniaturized design. Lastly, the CTSU, which is baselined for operation at 35 K with nitrogen (via its solid-solid transition at that temperature), was designed with a 6063 aluminum heat exchanger and integrally welded, segmented, high purity aluminum thermal straps for direct attachment to both a cryocooler cold finger and a beryllium component whose peak load exceeds its average heat load by 2.5 times. Test data is included herein only for the CTSU.
INTRODUCTION Cryogenic systems on future spacecraft will require a variety of advanced integration components to meet performance goals. These components are needed to achieve lower parasitics, better vibration isolation, tighter temperature control, improved reliability, higher heat transport capability, increased transport distances, across-gimbal transport, and longer life. To address these needs, the DoD and NASA have funded several research programs to develop novel cryogenic integration technologies. This research has led to the development of a number of innovative cryogenic devices, four of which are described herein including an across-gimbal CLHP, a miniaturized CLHP, a differential thermal expansion (DTE) cryogenic thermal switch (CTSW), and a dual-volume cryogenic thermal storage unit (CTSU). The technical problems that can be solved with these components have been long-standing targets of cryogenic system engineers. The across-gimbal CLHP will enable the cryocoolers on a cryogenic telescope like SBIRS-Low to be mounted off-gimbal thereby improving spacecraft pointing Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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agility due to the much lower on-gimbal weight. The miniaturized CLHP can simultaneously function as a flexible conductive link (FCL), cryogenic heat transport device, and CTSW, which enables the implementation of a vibration-isolated, thermally-switchable cryogenic system, obviating the need for a CTSW and two FCLs. The highly reliable DTE-CTSW enables multiple cryocoolers to be thermally linked to a cryogenic payload with minimal parasitics, opening up the possibility of using several low cost cryocoolers in a space system instead of one or two very costly long-life cryocoolers. Lastly, the dual-volume CTSU enables focal plane duty-cycling without sacrificing focal plane isothermality, so that the cryocooler can be sized for the average load rather than the peak load, potentially saving hundreds of watts of spacecraft power. Past programs that have helped set the stage for the work described herein include the STS flight programs CRYOHP, CRYOTP, CRYOFD, and CRYOTSU as well as the on-ground research and development programs CRYOBUS, 35KCTSU, CLHP, and CRYOTOOL. The key cryogenic (and ambient) technologies demonstrated on these programs are listed in Table 1. The remainder of the paper is organized into four stand-alone sections that describe each of the cryogenic components. The information is organized under the following main heading titles: AcrossGimbal CLHP, Miniaturized CLHP, Improved DTE-CTSW, and 35 K Dual-Volume CTSU. Within each section is described the component’s development rationale/objective, operating principles, heritage, requirements, design/construction details, and test results (where available). At the time this paper was written, test results were only available for the 35 K DV-CTSU. Test results for the other three components will be described in future papers.
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ACROSS-GIMBAL CLHP Development Rationale/Objective The objective of this effort was to develop an across-gimbal cryogenic heat transport system (HTS) that would enable gimbaled cryogenic components (e.g., optics) to be cooled by cryocoolers mounted off-gimbal. The advantages of off-gimbal mounting are: (a) improved gimbal pointing agility due to the lower on-gimbal weight; and (b) simplified thermal management of cryocooler waste heat due to the closer proximity of the cryocoolers to spacecraft radiators. Figure 1 illustrates cryogenic and ambient across-gimbal HTS concepts. This effort builds on the recent work by Marland8, who demonstrated that the on-gimbal mounting of cryocoolers (without an on-gimbal ambient radiator) was feasible using an across-gimbal ambient HTS comprised of an ammonia loop heat pipe (LHP) with coiled rigid tubing (small diameter, self-supporting coil). Obviously, the gimbal motors with this type of across-gimbal system must be able to handle the on-gimbal LHP weight and line torque. Off-gimbal mounting of cryocoolers using a CLHP with coiled rigid lines represents the optimum solution from an on-gimbal weight standpoint, but parasitics are higher and transport line torque is still an issue. Moreover, with any across-gimbal HTS, the coil diameter needs to be large enough to accommodate the optics. However, beyond a coil diameter of about 50-75 times the tubing OD (0.3 cm typical), the coil is not self-supporting. Thus, the purpose of this effort was to design, build, and test an across-gimbal CLHP with large diameter externally supported coils low parasitics, and low line torque.
Operating Principles A CLHP is a passive two-phase fluid loop that utilizes the capillary pressure developed in a fine pore (typically sintered metal) wick to circulate the working fluid. CLHP operational details and key features, along with those of the capillary pumped loop (CPL), LHP, and cryogenic capillary pumped loop (CCPL), are described by Yun7. Figure 2 depicts simple flow diagrams for the CPL, LHP, CCPL, and CLHP. As depicted in Figure 2, a CLHP is actually comprised of two loops: the primary loop and the auxiliary loop. During start-up, fluid is circulated in the auxiliary loop by applying power to the secondary pump/reservoir attached to the cooling source. When the primary evaporator is cold enough to sustain liquid, load can be applied there and the start-up heater can be turned down. After start-up is complete, heater power to the secondary pump must be maintained at a level greater than
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the sum of the transport line parasitic load and “reverse-conduction” across the primary wick. Thereafter, the net load to the primary evaporator is carried by the primary loop.
Heritage The design of the across-gimbal CLHP described herein builds on the designs of the following three devices: (1) ethane CLHP with a cryo-radiator interface described by Yun7; (2) ethane CLHP with the cryocooler interface described by Bugby9; and (3) across-gimbal ambient thermal transport system described by Marland7. O’Connell10 has also demonstrated a hydrogen CLHP using a design similar to that utilized herein.
Requirements The requirements for the system are as follows: (1) temperature (80-100 K); (2) heat load (2-20 W); (3) thermal stability (+/- 2 K); (4) azimuth motion (+/- 180°); (5) elevation motion (+/- 45°); (6) elevation/azimuth velocity, acceleration 2rad/s, ; (7) number of cycles (500,000); (8) line torque (0.07 N-m, 0.035 N-m goal); and (9) coil diameter (> 25 cm). These requirements implied the use of nitrogen as the working fluid. Nitrogen freezes at 63.15 K, boils at 77 K, and is supercritical above 126 K. From the programs described by Yun7 and Marland8, each of these requirements was easily achievable except for torque. In the latter program, torque was measured to be 0.17 N-m/line (0.32 cm OD, 0.025 cm wall) for +/- 200° range of motion.
Design/Construction To meet the requirements, a two-axis across-gimbal CLHP with coiled rigid tubing was developed as illustrated in Figure 3. The key features of this design are as follows: (1) azimuth coils (+/- 180° motion, 10 m line length, 25 and 38 cm diameter coils); (2) elevation coils (+/- 45° motion, 2 m length, 15 and 19 cm diameter coils); (3) line separators (slots formed by Delrin rod stacks placed at 45° increments around periphery); (4) evaporator (Al body, 1 -2 µm Ti wick, 1.6 cm diameter, 16 cm length, resistance heating source, mounted on elevation gimbal); (5) condenser (parallel channels, bolted interface, dual [1.5 cm diameter, 18 cm length] tube design formed by swaging axially grooved Al heat pipe extrusions around solid Al rods); (6) reservoir/secondary pump (1-2 µm Ti primary wick, stainless steel wire mesh secondary wick, shunted to condenser); (7) hot reservoir (stainless steel, 2000 cc volume, not shown); (8) transport lines (stainless steel, vapor line and secondary vapor line are visible, liquid line is coaxial within secondary vapor line; and (9) electrical connections (flat ribbon cable coils and uncoils in small cylindrical cell at center of azimuth gimbal axis). The line diameters are as follows: (a) vapor line (0.32 cm OD, 0.29 cm ID); (b) secondary vapor line (0.40 cm OD, 0.37 cm ID); and (c) liquid line (0.24 cm OD, 0.21 cm ID). Most cold surfaces will be Au-plated.
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Testing At the time this paper was written, the across-gimbal CLHP had not been fully assembled or tested. However, to verify the design, a breadboard test unit was constructed to evaluate the line separators and line durability. The breadboard evaluation unit was cycled successfully through about 50,000 cycles and the line separators were found to function acceptably.
MINIATURIZED CLHP Development Rationale/Objective The objective of this effort was to develop a miniaturized neon CLHP to replace an FCL and/ or a CTSW in linking a cryocooler to a vibration-sensitive component operating near 35 K. Figure 4 illustrates the concept. The potential benefits of this technology are fewer interfaces, better controllability, lower parasitics, smaller and easier integration. Use of a CLHP for this purpose requires a working fluid that is two-phase over the component operating range. The areas where CLHP miniaturization can be realized (and the type of design changes that are anticipated) include the evaporator (reduced diameter/length), condenser (multiple parallel microchannels vs. larger single-channels, integral vs. separated condenser for primary/auxiliary loops), transport lines (minimum diameter, liquid line coaxial within the secondary vapor line), and secondary pump/reservoir (minimum volume).
Operating Principles The operating principles of the miniaturized CLHP are identical to those of the across-gimbal CLHP described earlier. To achieve FCL-like operability, flexibility was designed-in by modeling the CLHP as an extension spring for axial flexibility, as a torsion spring for rotational flexibility, and as a combined curved/straight cantilevered beam for cross flexibility.
Heritage The design heritage for the miniaturized CLHP includes the two ethane CLHPs mentioned earlier and two previous CCPL designs. Bugby11 describes the small-scale CCPL flown on STS-95 as part of the CRYOTSU flight experiment. This device, denoted as CCPL-5, weighed just 200 g and occupied 300 cc of volume (not including its hot reservoir). With 0.16 cm OD tubing and nitrogen as the working fluid, CCPL-5 was able to transport 0.25-13 W a distance of 0.25 m with a conductance of roughly 1 W/K while operating over the range 70-110 K. In addition, a previous design iteration (CCPL-3) was used to demonstrate CCPL operation with neon as described by Bugby12. The miniaturized CLHP described herein will be about half the mass and one fourth the volume of CCPL-5.
Requirements The primary requirements for this device are as follows: (1) temperature (35 K); (2) heat load (0.3-2.5 W); (3) thermal switch performance (>1000 K/W “OFF”, <1 K/W “ON”, <10 minute
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actuation time); (4) mass (<< 0.5 kg); (5) transport length (> 10 cm); (6) axial flexibility (< 1000 N/m); (7) cross flexibility (< 4000 N/m); (8) rotational flexibility (< 5 N/degree); and (9) side load survivability (> 1 N). The temperature requirement implies the use of neon as the working fluid. Neon freezes at 24.5 K, boils at 27 K, and is supercritical above 44.5 K. Based on design calculations as well as experience gained in previous programs described by Yun7, Marland8, Bugby11,12, and Cullimore13, each of these requirements was thought to be achievable. Design/Construction To meet the requirements, the miniaturized CLHP illustrated in Figure 5 was designed. The key features of this design are as follows: (1) coiled tubing section (30 cm line length, 5 cm coil diameter); (2) straight tubing section (8 cm line length); (3) primary evaporator (1-2 µm Ti wick, 1 cm square bimetallic [aluminum-stainless steel] housing with 2.5 cm wide bottom flange, 3.8 cm length, resistance heating source); (4) condenser (OFHC Cu, photo-etched micro-channels, direct condensation, clamps directly to cold head, 3.8 cm diameter, 0.7 cm thickness, formed by HIPbonding); (5) reservoir/secondary pump (1-2 µm Ti primary wick, evaporator section same as primary evaporator, stainless steel reservoir, stainless steel wire mesh secondary wick, moderate conductance shunt to condenser); (6) hot reservoir (stainless steel, 100-200 cc volume, not shown); and (7) transport lines (stainless steel, visible lines are the vapor line and the secondary vapor line, liquid line is coaxial within the secondary vapor line. The line diameters are as follows: (a) vapor line (0.16 cm OD, 0.13 cm ID); (b) secondary vapor line (0.20 cm OD, 0.17 cm ID); and (c) liquid line (0.08 cm OD, 0.05 cm ID). The entire unit will be gold-plated to minimize radiative parasitics. To illustrate a comparison between the miniaturized CLHP and the CCPL flight unit mentioned above, an appropriately sized photo of CCPL-5 (sized to roughly match the scale in Figure 5) is provided in Figure 6.
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Testing At the time this paper was written, the miniaturized CLHP had not been fully assembled and tested. A future paper will describe the test results.
IMPROVED DTE-CTSW Development Rationale/Objective The objective of this effort was to modify the design of the first generation DTE-CTSW to include a polymer support tube (or rod) and high purity aluminum end-pieces. The first generation device, described by Marland14, utilized a stainless steel support tube and beryllium end-pieces. This initial choice of materials was made to enable direct mounting of the DTE-CTSW to beryllium components (e.g., FPAs) to insure zero-DTE. However, this initial material choice also meant that the gap thickness had to be quite small (0.005 cm or 2 mil) and precluded its use when system surroundings were colder than about 220 K. To eliminate those shortcomings, new construction materials were needed. After investigation, it was found that a polymer rod in conjunction with aluminum end-pieces would do the following: (a) allow the gap to be at least 4 times larger (0.02 cm or 8 mil) because polymers have a much higher coefficient of thermal expansion (CTE) than stainless steel; (b) allow the surroundings to be as cold as 80 K because polymers continue to contract to well below 80 K whereas most metals do not; and (c) make manufacturing easier because beryllium machining, which can only be done at specialized facilities, would not be needed. Thus, the improved DTE-CTSW could be utilized on lower temperature applications (perhaps 10 K) whose surroundings invariably need to be cooled to below 100 K to minimize radiative parasitics. Combining the new design with integrally welded, segmented FCLs (described in the 35 K DV-CTSU section to follow) would still allow near zero-DTE mounting to beryllium components. Performance of the improved design is expected to be comparable to or exceed that of the first generation unit, which was 1400 K/W “OFF” and 1.2 K/W “ON” at 50 K. The primary application for the DTE-CTSW is in thermally coupling redundant cryocoolers to a cryogenic component while ensuring low “OFF” cooler parasitics and high reliability.
Operating Principles Figures 7a and 7b show the design of the first generation DTE-CTSW. As indicated, the device is comprised of a large beryllium end-piece, a small beryllium end-piece and a thin-walled stainless steel tube. A heater on the stainless steel tube is needed for “OFF” actuation. The small end-piece is attached to the cryocooler and the large end-piece is attached to the cooled component. To actuate the DTE-CTSW “ON”, the cryocooler is powered on, which causes the stainless steel tube to contract, bringing the end-pieces into contact. Since the beryllium contracts less when cooled than the stainless steel, the DTE-CTSW will stay “ON” even after the beryllium has cooled, with a force determined by the initial gap, the DTE, and the stainless steel modulus.
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To actuate the DTE-CTSW “OFF”, a small heater at the center of the stainless steel tube is powered on to temporarily raise the tube midpoint temperature to 300 K. This action expands the tube enough to break contact, causing the cryocooler end of the DTE-CTSW to warm. The DTECTSW will stay “OFF” as long as the cooler cold head is above about 220 K. The “trick” in staying “OFF” is that the large beryllium end-piece, which is still cold, contracts more than the stainless tube with one end cold and one end warm. This design has been patented by Swales.
Heritage The idea for the first generation DTE-CTSW came about during a project to redesign the gasgap CTSW of Johnson15. The redesign involved changing the end-pieces (that initially formed a conical gap separated by G-10 pins under compression) to create a flat gap that was maintained by a central cantilevered tube. Although this change did what it was supposed to do (simplify construction, improve reliability and performance), continuing difficulties with the metal hydride pump pointed to a need for further simplification. It was postulated by Marland16 that the metal hydride pump and the hermetic shell surrounding the end-pieces could be eliminated completely and that end-piece actuation could be accomplished by DTE between the end-pieces and the central tube. Thus, the first generation DTE-CTSW was born.
Requirements The primary requirements for this device are as follows: (1) temperature (35-95 K cold end, 300 K warm end); (2) heat load (0.3-2.5 W); (3) switching ratio (1000:1 with a goal of 2000:1); (4) switching time (<10 minutes); (5) mass (<< 0.5 kg); (6) switching cycles (>100, no cold welding); and (7) interfaces (ability to integrate high conductance FCL into design). Based on design calculations as well as experience gained in the previous DTE-CTSW program, each of these requirements was thought to be achievable.
Design/Construction To meet the requirements, the improved DTE-CTSW illustrated in Figure 8 was designed. The key features of this design are as follows: (1) polymer rod (Ultem, 5 cm length, 1.5 cm diameter); (2) small end-piece (5-nines Al, 3.2 cm diameter, 0.5 cm thickness); (3) large end-piece (5-nines Al, 2.8 cm diameter [5 cm flange], 5.4 cm length); (4) heater (Kapton resistance heater); and (5) coatings (end-pieces bare Al, one layer of double-aluminized Mylar around rod).
Testing At the time this paper was written, the improved DTE-CTSW had not been fully assembled and tested. A future paper will describe the test results. Typical applications that might benefit from DTE-CTSW technology are illustrated in Figure 9.
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35 K DUAL-VOLUME CTSU Development Rationale The objective of this effort was to determine the ability of a nitrogen dual-volume (DV) CTSU to maintain precise thermal control of a duty-cycled cryogenic component at 35 K. This technology is applicable to cryogenic components such as focal plane arrays of remote sensing satellites that need to be ON only part of the time to fulfill their mission requirements. Although duty-cycling reduces cooler capacity and input power irrespective of whether a CTSU is used, the real difficulty is in minimizing component temperature variation during duty-cycling. Thus, a 35 K CTSU that minimizes component temperature variation during duty-cycling was the goal. The desired test configuration, as shown conceptually in Figure 10, placed the CTSU heat exchanger (HX) between the component and the 35 K cryocooler. One side of the CTSU-HX was to be thermally linked to the cooler via an FCL. The other side was to be thermally linked to a dutycycled beryllium component via one of the following: (a) direct mounting; or (b) FCL attachment. The design goal in selecting either (a) or (b) was to minimize the DTE between the CTSU-HX and the duty-cycled beryllium component while still meeting program requirements. Direct mounting
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implied constructing the CTSU-HX out of beryllium while using an FCL enabled a lower risk construction material such as aluminum to be used. Both options were feasible using techniques described by Bugby9,17,18. Ultimately, option (b) was selected.
Operating Principles When configured as Figure 10 indicates, the CTSU-HX must intercept and store excess heating from the component when the component is ON and intercept and store excess cooling from the cooler when the component is OFF. To do so isothermally, the CTSU-HX must have a high effective heat capacity and a high thermal conductance. Those two attributes can be obtained by filling a thermally conductive HX with a cryogenic working fluid (normally a gas at room temperature) that has a phase transition near the component operating temperature. To reduce the fill pressure, minimize HX wall thickness, and maximize PCM mass (hence energy storage capacity), a dualvolume system is preferred to a single-volume system. Component isothermality is maximized by ensuring the HX core has a high surface area to volume ratio and that all resistances in the system (interfaces, FCLs, HX) are minimized. HX core configurations may include metal foam, honeycomb, drilled holes, parallel plates, or combinations thereof.
Heritage The program that set the stage for this effort was the CRYOTSU flight experiment that flew on STS-95 in 1998 (see Bugby4). The goal of that program was to demonstrate 3000 J of phase change at or near 60 K with beryllium CTE compatibility. During the program, several single and dualvolume CTSU test units were designed/constructed and many PCMs were evaluated to investigate CTSU performance. The design ultimately chosen had a drilled-hole core beryllium HX, a stainless steel storage tank, and nitrogen PCM (63 K melt). During the flight, the 63 K performance tests were all highly successful. In addition, due to the exceptional performance of the thermal isolation system, one test was carried out that cooled the HX to 33 K, which successfully demonstrated the nitrogen 35 K solid-solid transition in zero-g. The program described herein provides additional characterization of that transition. Requirements The requirements for this effort consisted of four primary thermal requirements, several secondary requirements, and additional derived requirements. The cooled component in this system was a 12.7 cm x 12.7 cm x 1.27 cm O-30 beryllium plate designated as the “simulated instrument bar” or SIB. The primary thermal requirements, arranged in order of priority, were as follows: (1) SIB dT/dt during peak heating (< 0.02 K/min); (2) between SIB and cooler cold head (< 8 K); (3) SIB spatial (< 0.25 K); and (4) SIB temperature variation (36.5 K +/- 2 K). The heat load was a recurring square-wave with a peak of 3.7 W for 27 minutes and a minimum of 0.7 W for 73 minutes (total period of 100 minutes). The ability of the CTSU to meet the SIB thermal requirements was to be evaluated over at least 14 consecutive square-wave cycles (which is roughly 24 hours of continuous cycling). The environment surrounding the HX was to be held at 80 K. Figure 11 illustrates the recurring heat load profile and analytical/graphical determinations of the minimum energy storage capacity of 3548 J and the minimum cryocooler capacity of 1.51 W. An additional requirement was to minimize the DTE between the CTSU-HX and the SIB.
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The derived requirements consisted of constraints placed on the following intra-system resistances: (a) SIB FCL and its two interfaces (b) CTSU HX and (c) cooler FCL and its two interfaces Figure 12 graphically illustrates the situation and the computational methodology used to arrive at the derived resistance requirements of (to meet Requirement 1), (to meet Requirement 4), and (to meet Requirement 2). The meeting of Requirement 3 was not mentioned above because doing so depends solely on the intrinsic thermal conductivity of the SIB material, hence the choice of 0-30 Be (k of 210 W/m K at 35 K) rather than S-200F Be (k of 150 W/m K at 35 K). Design/Construction The design of the 35 K DV-CTSU HX is illustrated in Figure 13. As indicated, the HX is formed by welding together two drilled-hole 6063 aluminum halves. The SIB FCL is comprised of 20 high purity (99.999% or “5-nines”) aluminum foil S-links welded to one side of the HX. The SIB-end of each of these S-links is comprised of a small cube of high purity aluminum that bolts separately to the SIB to minimize DTE effects on the SIB. This type of FCL attachment scheme has been designated as a “segmented FCL”. The cooler FCL is comprised of 4 high purity (5-nines) aluminum foil S-links welded to the other side of the HX. The opposite end of each of these four S-links is welded to a 5-nines aluminum plate to which an operational cooler cold head would be bolted. In the tests to be described later, the top end of the Q-meter simulated the operational cooler cold head interface. Figure 14 illustrates additional system details including charge pressure (2.24 MPa or 326 psi), charge mass (630 grams), energy storage capacity (5670 J), HX void volume (760 cc), and storage tank volume (23.07 liters).
Testing Testing of the 35 K DV-CTSU was carried out with the set up illustrated in Figures 15 and 16. As indicated, the HX was suspended within a gold-coated OFHC Cu shroud by Kevlar fibers.
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Cooling was provided by connecting the cooler FCL to the top end of a Q-meter whose bottom end was cooled by an L-shaped OFHC Cu bracket connected to the 2nd stage of a 2-stage GM cryocooler. An OFHC Cu bar was used to couple the shroud to the GM cooler 1st stage. The testing involved the following sub-elements as described below. Q-Meter Calibration. In this test, the Q-meter, illustrated in Figure 16, was calibrated by keeping its bottom temperature constant (29.0, 31.0, 32.9, and 35.0 K) while varying its top temperature by applying heater powers between 1.25 and 3.3 W in 0.25 W increments. The Q-meter was a 6061 aluminum spool surrounded by an OFHC Cu shroud. During calibration an OFHC Cu “cap” closed out the top of the shroud. When the Q-meter was bolted to the cooler FCL during later tests, the calibration curve enabled heat flows to be resolved to within 10 mW. Parasitics. This test was carried out to measure the total parasitic load into all DV-CTSU elements (SIB, FCLs, and HX) at shroud temperatures of 36, 80, and 120 K. The results indicated
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respective parasitic loads at those temperatures of 0.14, 0.20, and 0.38 W. All tests were carried out with the Q-meter cold end (bottom) at 32.91 K, which resulted in respective warm end (top) temperatures of 35.32, 35.38, and 35.62 K. Conductance. This test was carried out to attain a preliminary idea of how well the system would function in later testing. The test was carried out by applying a load of 1.51 W to the SIB heater and measuring the change in system values. Resistances were computed by dividing the change in across a given portion of the system by the applied heat load. The results were as follows: (a) which was over 10 times better than the derived requirement of 0.94 K/W; (b) which was nearly 2 times better than the derived requirement of 0.18 K/W; and (c) which was about 7 times better than the derived requirement of 2.46 K/W. So, after carrying out this test, there was a high level of confidence that the system would easily exceed all thermal requirements. Sensible Heat. This test was carried out to assess system response without the PCM (HX evacuated) to illustrate the benefits of the PCM. A heat load of 3.7 W was applied to the SIB heater resulting in an initial slope of the SIB temperature about 25 times higher than the requirement of 0.02 K/min. This result agreed with an analytic calculation (m = Q/[Cp dT/dt]) that indicated that 175 kg of aluminum (about 25-30 times more than the mass of the HX) would be needed to provide the required SIB thermal stability. Full-Discharge Cycling. A full-discharge cycling (FDC) test consists of a complete phase transition on heating (of measured duration ) followed or preceded by a complete phase transition on cooling (of measured duration ). The phase transition durations and applied heat loads can be used to compute the “measured” energy storage capacity and the net cooling rate The net cooling rate is defined as the cooler capacity less the parasitics which are assumed to be constant during both heating and cooling. The requirements are: (1) a constant heat load in excess of the net cooling rate during heating; (2) a constant (possibly zero) heat load that is less than the net cooling rate during cooling; and (3) a near constant net cooling rate during both heating and cooling. Equations (1) and (2) provide the relationships. The theoretical energy storage capacity can be computed using charge mass as indicated in Equation (3). should agree with to within a few percent. Figure 17 and Table 2 illustrate the results of FDC testing as well as the results of applying Equations (1-3).
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Partial-Discharge Cycling. A partial-discharge cycling (PDC) test consists of a series of temperature cycles each comprised of a partial phase transition on heating followed or preceded by a partial phase transition on cooling. Ideally, the PCM during PDC testing should stay within its twophase region. So, if the working fluid were a freeze-thaw PCM, it should never fully melt and never fully freeze. From experience, margin of at least 15% of the minimum charge is needed on either side of the two-phase region. Hence, for a freeze-thaw PCM, 15% needs to remain unfrozen after cooling and 15% needs to remain frozen after heating. To achieve this margin, the charge ratio (actual/minimum) needs to be at least 1.3. The 35 K DV-CTSU was given additional margin to a charge ratio of 1.6 (5670 J vs. 3548 J). A PDC test is supposed to demonstrate how a duty-cycled component with a CTSU will operate. But an actual duty-cycled component will probably have many more cycles than can be reasonably tested in the laboratory. As such, small cycle-to-cycle variations may cause a system to drift out of the two-phase region and operational temperature limits may be exceeded. To alleviate this problem, an autonomous control scheme to eliminate drift was needed. The control method that was developed depends on the following; namely, that as long as the heat load, cooling rate, and the environment are relatively constant, CTSU temperatures are uniquely related the charge state of the system. This relationship means that during cooling a thermostatically controlled heater can be used to maintain the temperature of an arbitrary reference point in the system (the CTSU bottom temperature, for example) when it reaches a level corresponding to 85% charged. This method thus eliminates drift by readjusting the charge level during each cycle. To implement the control method, the cryocooler capacity needs to be slightly higher (5-10%) than the minimum. Figure 18 illustrates the method using the parameters of this problem. Figure 19 shows how well the method works in controlling actual PDC tests. Ambient-Cryogenic Cycling. The ambient-cryogenic cycling test was carried out to determine if repeated 300 K to 35 K to 300 K cycles would degrade system performance. In all, 30 such cycles were carried out over a period of 45 days. No degradation occurred. System Performance vs. Requirements. After thorough evaluation of the test data, the performance of the 35 K DV-CTSU is summarized in Table 3 and compared to requirements. Also
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shown are results from thermal modeling efforts (not described here) that were carried out before testing (see Model: As Built) and after testing (see Model: Test Verified). As indicated, all requirements were met by a considerable margin and the model predictions agreed with the data.
SUMMARY This paper has described the development status of four advanced components that help solve important problems in cryogenic integration. The across-gimbal nitrogen CLHP will enable gimbaled cryogenic optical systems to mount cryocoolers off-gimbal for optimum pointing agility and easier integration with ambient heat dissipation systems. The miniaturized neon CLHP will enable a cryogenic cooling source to be directly linked to a nearby cryogenic component with vibration isolation and thermal diode action. The improved DTE-CTSW will enable cryocoolers to be linked to a cryogenic component with negligible parasitics and high reliability. Lastly, the dual-volume nitrogen CTSU will enable a high power-dissipating component with a need for isothermality to be dutycycled without sacrificing temperature stability. In testing of the DV-CTSU, an autonomous control methodology for duty-cycled, CTSU-endowed systems was developed and that method was described herein. In sum, each of these components can potentially solve or help solve a range of cryogenic integration problems of interest to DoD, NASA, and the commercial sector. Test results were provided herein for the DV-CTSU. Future papers will address the testing of the other components.
ACKNOWLEDGMENT The authors would like to acknowledge and thank B.J. Tomlinson and Thom Davis of AFRL for their support and guidance on this work. The authors would also like to acknowledge Ted Swanson, Jentung Ku, and their co-workers at NASA/GSFC for their support of the CRYO-series of test-bed flight experiments upon which the work described herein is based.
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REFERENCES 1.
Beam, J., P. Brennan, M. Bello, “Design and Performance of a Cryogenic Heat Pipe Experiment (CRYOHP),” 27th AIAA Thermophysics Conference, Nashville, TN, July 6-8, 1992.
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Brennan, P., M. Buchko, T. Swanson, M. Bello, and M. Stoyanof, “Cryogenic Two-Phase Flight Experiment: Results Overview,” 1995 Shuttle Small Payloads Symposium, NASA/GSFC, 1995.
3.
Thienel, L., P. Brennan, M. Stoyanof, and C. Gerhart, “Design and Performance of the Cryogenic Flexible Diode Heat Pipe (CRYOFD) Flight Experiment,” 28th International Conference on Environmental Systems (ICES-1998), Danvers, MA, 1998.
4.
Bugby, D. and C. Stouffer, “Flight Results from the Cryogenic Thermal Storage Unit (CTSU) Flight Experiment on STS-95,” SAE Paper No. 1999-01-2085, 29th International Conference on Environmental Systems (ICES-1999), Denver, CO, 1999.
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Bugby, D., Integrated Cryogenic Bus (CRYOBUS) Program, Final Report prepared for the Air Force Research Laboratory by Swales Aerospace, AFRL-VS-TR-2001-1059, March, 2001.
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Bugby, D., B. Marland, and C. Stouffer, “Development and Testing of a 35 K Cryogenic Thermal Storage Unit”, Briefing presented at the 2002 Spacecraft Thermal Control Symposium, The Aerospace Corp., El Segundo, CA, March, 2002.
7.
Yun, J., E. Kroliczek, and L. Crawford, “Development of a Cryogenic Loop Heat Pipe (CLHP) for Passive Optical Bench Cooling Applications”, 32nd International Conference on Environmental Systems (ICES-2002), SAE Paper No. 2002-01-2507, San Antonio, Texas, 2002.
8.
Marland, B., J. Yun, D. Bugby, C. Stouffer, B. Tomlinson, and T. Davis, “Across-Gimbal Ambient Thermal Transport System”, 31st International Conference on Environmental Systems (ICES-2001), SAE Paper No. 2001-01-2195, Orlando, FL, 2001.
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Bugby, D., B. Marland, C. Stouffer, and E. Kroliczek, “Advanced Components and Techniques for Cryogenic Integration,” Briefing presented at the 2002 Spacecraft Thermal Control Symposium, The Aerospace Corporation, El Segundo, CA, March, 2002.
10. O’Connell, T., T. Hoang, D. Khrustalev, S. Norin, J. Ku, “Hydrogen Loop Pipe Design and Test Results,” Spacecraft Thermal Control Symposium, Aerospace Corporation, El Segundo, CA, 2002. 11. Bugby, D., E. Kroliczek, J. Ku, T. Swanson, B. Tomlinson, T. Davis, J. Baumann, and B. Cullimore, “Design and Testing of a Cryogenic Capillary Pumped Loop Flight Experiment,” Space Technology and Applications International Forum (STAIF-99), Albuquerque, NM, 1999. 12. Bugby, D., E. Kroliczek, B. Cullimore, and J. Baumann, “Experimental Investigation of a Neon Cryogenic Capillary Pumped Loop,” 34th IECEC, Paper 97272, 1997. 13. Cullimore, B., E. Kroliczek, and J. Ku, “Cryogenic Capillary Pumped Loops: A Novel Cryocooler Integration Technology,” Cryocoolers 8, Plenum Press, NY, 1995. 14. Marland, B., D. Bugby, and C. Stouffer, “Development and Testing of a High Performance Cryogenic Thermal Switch,” Cryocoolers 11, Kluwer Academic/Plenum Publishers, NY, 2001. 15. Johnson, D. and J. Wu, “Feasibility Demonstration of a Thermal Switch for Dual Temperature IR Focal Plane Cooling,” Cryocoolers 9, Plenum Press, NY, 1997, pp. 795-805. 16. Marland, B., D. Bugby, and C. Stouffer, “Development and Testing of Advanced Cryogenic Thermal Switch Concepts,” Space Tech. Appl. and Int’l Forum (STAIF), Albuquerque, NM, 2000. 17. Bugby, D., R. Bettini, C. Stouffer, et al., “Development of a 60 K Thermal Storage Unit,” Cryocoolers 9, Plenum Press, NY, 1997, pp. 747-764. 18. Bugby, D., “Cryogenic Thermal Storage Unit and Flight Test Results,” Final Report for the Air Force Research Laboratory by Swales Aerospace, AFRL-VS-TR-1999-1044, February 1999.
Cryogenic Loop Heat Pipes as Flexible Thermal Links for Cryocoolers Dmitry Khrustalev Thermacore, Inc. Lancaster, Pennsylvania, USA 17604
ABSTRACT Loop heat pipes with flexible transport lines are needed in some cryogenic applications as thermal links between cryocoolers and the cooled components for a variety of reasons such as vibration isolation, increased thermal transport distance, thermal diode function, multiple components, etc. This paper presents an innovative configuration of a cryogenic loop heat pipe designed to operate in the temperature range between 70 K and 140 K. The experimental data obtained for the loop heat with oxygen as a working fluid show that it reliably and predictably started in multiple tests and operated with the heat load range at the main evaporator from 0.5 W to 9 W with the main evaporator elevated up to 5 cm versus the condenser.
INTRODUCTION Thermal links between cryocoolers and cooled components ideally should satisfy multiple requirements: flexibility for the installation convenience and vibration isolation, heat transport with low end-to-end temperature differential, thermal diode function, increased transport distance, completely passive operation, etc. Conventional heat pipes and loop heat pipes (LHPs) can be used for that purpose. Conventional cryogenic heat pipes typically have rigid casings with capillary structures inside and cannot operate in ground testing with a significant elevation of the evaporator versus the condenser due to a relatively large pore size of the capillary structure, as explained, for example, by Rosenfeld1. Loop heat pipes have capillary wicks, usually with the pore radii from 1 micron to 3 microns, only inside the evaporator, which is linked with the LHP condenser with two small diameter smooth-wall tubes up to several meters long. It is due to the smaller pore radius of the wick, which is capable of maintaining high capillary pressure, that LHPs can transport much higher amounts of thermal energy compared to conventional heat pipes. LHPs are also thermal diodes due to their configuration. Long small diameter wickless transport lines allow bending, flexibility and can provide vibration isolation when coiled. One additional technical requirement for cryogenic LHPs is that they should be capable of cooling down from a room temperature, which is above the critical temperature of the working fluid, to the operating cryogenic temperature with only the condenser end being cooled by a cryocooler. This requirement results in a difference in the configuration of cryogenic LHPs compared to room-temperature LHPs. Various configurations of cryogenic LHPs were discussed by Khrustalev2, Khrustalev3, and Hoang4. LHPs with longer heat transport distance require three transport lines for reliable operation. Test data obtained for a LHP with three transport lines with Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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hydrogen as a working fluid are presented by Hoang4, where the LHP operated at about 30 K and the lines were 2.5 m long. The LHP configuration presented in this paper has two transport lines, which is beneficial for the applications in terms of ease of installation, flexibility, mass, vibration, etc.
CONFIGURATION AND OPERATION OF THE CRYOGENIC LHP The LHP configuration and components are shown in Fig. 1. The LHP has three additional components versus conventional LHPs: a secondary evaporator, additional condenser, and a pressure reduction reservoir. The secondary evaporator is needed to force the flow of the cold liquid out of the additional condenser towards the main evaporator to cool it below the critical temperature of the working fluid at the initial stage. The pressure reduction reservoir (PRR) helps to keep the pressure inside the LHP within reasonable limits, while the LHP is stored at a temperature exceeding the critical temperature of the working fluid. PRR can be, in principal, eliminated by minimizing the component volumes. Additionally, the top portion of the main evaporator envelope has a void called the compensation chamber. A mesh-screen wick (so-called secondary wick) is enclosed inside the compensation chamber to hydraulically link the liquid in the compensation chamber with the surface of the main evaporator wick. The initial cooling down and the normal operation modes are explained below. The pressure inside the LHP at room temperature is reduced using the pressure reduction reservoir so that it drops below the critical pressure of the working fluid when the main and additional condensers are cooled by the cryocooler. As soon as the condenser temperatures are below the critical temperature, the secondary evaporator is heated by an additional heater to initiate vaporization from the porous metal wick, shown in Fig. 2. The porous wick has a blind concentric bore inside, where the cold liquid resides. Evaporation of liquid takes place on the wall-wick interface, when the wall is heated, and the vapor flows through the grooves on the surface of the wick out of the evaporator into the tube leading to the additional condenser. The vapor is being pushed by the capillary pressure developed by the heated porous wick. The vapor
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displaces the cold liquid from the additional condenser pushing it towards the main evaporator through the liquid return tube, thus cooling the main evaporator to the operating temperature. After the main evaporator is sufficiently cold, the normal LHP operation can start. The thermal energy to the main evaporator causes vaporization on the wick-wall interface and the vapor flows through the vapor tube to the main condenser, where it condenses. The cooled liquid flows through the porous wick of the secondary evaporator, which is not heat loaded in the normal operation mode, then through the additional condenser, returning to the main evaporator through the liquid flow tube. The two-phase circulation is supported by the capillary pressure developed as a result of heating of the main evaporator, which thus serves as a capillary pump. Significant parameters of the tested LHP with oxygen as a working fluid are given in Table 1.
TEST SETUP AND PROCEDURES The main and additional LHP condensers were bolted to the cold aluminum block attached to the cryocooler head, as shown in Fig. 3. The LHP was filled with oxygen to the pressure of 4.53 MPa (657 PSI) at room temperature of 23°C and instrumented with eighteen T-type thermocouples calibrated in the temperature range between 23°C and – 200°C using highprecision Lake Shore DT-670 silicon diodes. The accuracy of the thermocouples was better than ±1°C. Thermocouples were spot-welded to the LHP components, except for the small diameter vapor and liquid lines, which were gold plated to reduce the parasitic heat input due to the radiation from the shrouds enclosing the LHP. Several thermocouples were taped to the gold plated lines using several layers of Kapton tape. Thermocouple locations and numbers are shown in Fig. 4. The two temperature-controlled copper shrouds were positioned inside a 75 cm ID cylindrical vacuum chamber with the CGR-511 cryocooler cold head between the top and the bottom shrouds. The cryocooler cold head temperature was controlled by a temperature controller with accuracy of 0.1°C at steady state. The LHP was initially put in the horizontal orientation (the vapor line orientation was verified by a laser level). The main evaporator of the LHP was hanging on a stainless steel chain. The pressure inside the chamber was typically
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between and torr depending on the temperatures of the shrouds and the cryocooler head. Thermocouple readings were recorded using a Fluke Data Logger system. Each test started after the chamber was under vacuum conditions, from room temperature, by turning on the cryocooler and pumping the coolant through the shrouds. It took about 1.5 hours for the cryocooler to cool the aluminum block to 140 K, at which time the heater on the secondary pump was turned on to initiate cooling down the main evaporator to the operating temperature. The main evaporator then was heat loaded and the steady state was maintained for at least one hour to make sure the LHP was able to operate continuously for long time periods. The test data are presented in the next section.
TEST DATA OBTAINED FOR THE LHP FILLED WITH OXYGEN Multiple thermal tests were performed for the LHP all with the cryocooler cold head temperature of 70 K. The tests were performed with three different temperatures of the shrouds: 79K (-192°C), 163K (-110°C), and 290K (+17°C), however only test data for the first two temperature levels are presented in this paper. The test data for the room-temperature shrouds are being obtained with and without MLI, for completeness of this investigation. Figure 5 shows temperatures of the LHP components, pressure reduction reservoir (PRR) and that of the shrouds versus time as well as the heat loads on the main and additional evaporators.
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The shroud temperature was maintained close to –110°C. The additional evaporator was loaded with 4 W 2.5 hours after the test started. Increasing the heat load to 6 W noticeably accelerated cooling down of the main evaporator. The main evaporator was loaded with 3 W at 5.4 hours and the LHP successfully started and continued to operate with zero power on the additional evaporator. Note that the main evaporator was elevated by 5 cm versus the condenser and the PRR temperature was about +10°C throughout the all test. An enlarged fragment of Fig. 5 is given in Fig. 6, where the temperatures of the components can be read conveniently. Increase of the heat load to 6 W at 7.6 hours resulted in elevation of the temperature on thermocouple 4 (main evaporator) to –198.5°C by less than one degree. When the power was increased to 9 W at 8.8 hours, the evaporator temperature went up to –194°C by almost four degrees, signifying that the LHP approached the evaporator maximum heat flux capability, while its operation was absolutely stable for more than one hour. In fact, the evaporator was designed to minimize the so-called “heat leak” to the compensation chamber, while operating below the power level of 10 W. Abrupt change of the heat load to 1 W at 10.4 hours resulted in the reduction of the evaporator temperature to –200°C (70K) and increase of the temperatures read by thermocouples located on the transport lines. The vapor line thermocouple temperature went up to –184°C (86K) due to the radiation from the shrouds, which were almost 100 degrees above the LHP evaporator temperature, and decreased mass flow rate of the cold oxygen inside of the LHP lines. Moreover, the vapor flow inside the transport lines does not cool the metal lines as well as the liquid does. The actual temperature given by a thermocouple attached to the outer surface of the metal line does not give the temperature of the vapor inside that line. The lowpower operation of the LHP at 1 W continued for more than 4.5 hours. Due to the importance of the low-power operation of LHPs in cryogenic applications, where the heat loads can be 0.5 W or less, another test was conducted with the LHP operating at 0.5 W, as shown in Fig. 7. The LHP operated in horizontal orientation with the shrouds at – 192°C and the heat load at the main evaporator of almost 9 W till 12.3 hours, when the heat load was reduced to 0.5 W. At that time the evaporator temperature (thermocouple number 4) went down to –202°C and the LHP operated in a steady state- steady flow regime for more than one hour. At 13.6 hours the coolant flow in the shrouds stopped and their temperature started to slowly increase. The LHP operation continued after that moment for 1.5 hours and stopped due to a dryout in the evaporator when the shroud temperature reached –120°C. Operation of cryogenic LHPs with the shrouds at room temperature is a desirable feature that would make them more reliable and suitable for a wider range of applications. Figure 8 shows test data for the LHP with the evaporator elevated versus the condenser by 2.5 cm. The LHP operated for several hours with the main evaporator heat loaded with 7 W and the shroud temperature of –110°C. The coolant flow in the shrouds stopped at 29.5 hours and the shrouds started to warm. The LHP continued to work with slightly increasing operating temperature due to the increasing parasitic radiation from the shrouds to the LHP components. The operation was interrupted by a dryout in the evaporator when the shroud temperature reached –25°C. Further minimizing the parasitic radiation to the LHP components by protecting them (for example, with MLI) and making the surface areas of the components smaller will allow the LHP to operate with the shrouds at room temperature.
SUMMARY AND CONCLUSIONS 1. An innovative configuration of a cryogenic LHP with a secondary evaporator to facilitate cooling down from the supercritical conditions prior to startup of the main evaporator has been successfully demonstrated using oxygen as a working fluid. 2. The LHP reliably and predictably started in multiple tests and operated with the heat load range at the main evaporator from 0.5 W to 9 W with zero power on the secondary evaporator. 3. The LHP operated with the evaporator temperature of about 75 K (-198°C) with the shroud temperature of approximatelyl70K(-1100C±10°C) and at 71 K (-202°C) with the shroud
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temperature of 81 K (–192°C) while the cold block temperature was 70 K (-203°C). That means that the evaporator and condenser thermal resistances of the LHP are sufficiently low. 4. The LHP had no problem transporting 9 W with the main evaporator elevated versus the condenser by 5 cm. 5. The LHP operated in a transient regime with the subcooling of more than 25 degrees with the shrouds approaching the room temperature. That indicates that the radiation parasitic heat input to the LHP components is a major factor affecting the LHP operation with unprotected components (no MLI on the components). More thermal tests will be conducted with the shroud temperatures at +20°C and the LHP components covered with MLI.
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ACKNOWLEDGMENT The work presented in this paper was funded under BMDO DASG60-99-C-0061 Contract. Appreciation is also expressed to Mr. Peter Wollen for his efforts in assembly and testing of the cryogenic LHP.
REFERENCES 1. Rosenfeld, J., Wolf, D., and Buchko, M., “Emerging Technologies for Cryocooler Interfaces,” Cryocoolers 8, Plenum Press, New York (1995), pp. 743-753. 2. Khrustalev, D., Rosenfeld, J., Stanley, S., and Gerhart, C., “Cascaded Oxygen/Ethane Loop Heat Pipes,” Proceedings of the 11th International Heat Pipe Conference, Tokyo, Japan, September 1999. 3. Khrustalev, D., “Development of Cryogenic Loop Heat Pipes,” 11th Annual Spacecraft Thermal Control Technology Workshop, March 2000, El Segundo, CA. 4. Hoang, T., O’Connell, T., Khrustalev, D., and Ku, J., “Cryogenic Advanced Loop Heat Pipe in Temperature Range of 20-30 K,” Proceedings of the 12th International Heat Pipe Conference, Moscow, May 2002.
A Thermally Conductive and Vibration Protective Interface for Linear Cryogenic Coolers in Applications for Gimbaled Infrared Devices A.M. Veprik, V.I. Babitsky Wolfson School of Mechanical and Manufacturing Engineering, Loughborough University, Loughborough, Leics, LE1 13TU, UK S.V.Riabzev, N.Pundak Ricor Ltd, En Harod Ihud, 18960, Israel
ABSTRACT Modern electro-optic applications often contain a cryogenically cooled infrared imager placed upon a stabilized platform which is further connected to an outer housing by low-friction gimbals. Since the active system of gimbal stabilization is dedicated primarily to maintaining the steady line-of-sight control by eliminating the relatively slow effects of yaw, pitch or roll, it may have insufficient resources to suppress an excessive high-frequency vibration exported from the internal active components such as a cryogenic cooler. The vibration protective and thermally conductive interface developed allows the use of the split Stirling cryocooler, relying on highly efficient, cheap and durable linear single-piston compressor within the infrared imager mounted upon the gimbaled stabilized platform.
INTRODUCTION Sophisticated infrared (IR) imagers, which rely on a new generation of cryogenically cooled high-resolution focal plane arrays, enhance tremendously the ability to detect and track ground, sea, air and space targets and also to navigate at night time [1,2]. Modern military electro-optic applications often contain such an IR imager placed upon a stabilized platform which is connected to an outer housing by low-friction gimbals, thus eliminating the effects of yaw, pitch or roll for steady line-of-sight control. Among those applications are so-called forward-looking-infrared (FLIR) devices [3], which find use in aircraft and helicopters for detection/recognition/identification/laser designation of surface targets, accurate delivery of smart weaponry, performance of low level night flights, etc [4,5, see also examples at http://www.flir.com]. In another example, the space based infrared sensors (SBIRS) are utilized for space surveillance, missile warning and tracking missions [6]. Similar systems have also been adapted for use in ground and naval vehicles. In recent years some manufacturers have developed hand-held units.
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The above stabilization in a fixed inertial orientation relies typically on a closed-loop operation of motorized rotation stages, the servomotors of which are driven by the system controller. This control is based on the information provided by the sensors directly measuring the rotation of the gimbal axes. The new generation of stabilized IR imagers requires an essential reduction in the levels of overall power consumption and heat generation along with a further improvement in the pointing accuracy within very fine micro-radian tolerances through the application of advanced, integrated active and passive structural controls [6]. The above IR imagers often rely on closed cycle linear split Stirling cryogenic coolers. Such a cooler normally comprises two separate components: a linear compressor and a pneumatically driven expander, which are interconnected by a flexible gas transfer line (thin-walled stainless steel tube of small diameter) providing for maximum flexibility in a system design [7]. The simplest linear single-piston compressor is an inexpensive, reliable and highly efficient component of the above cooler. Such an unbalanced compressor is commonly used when the level of vibration export is not critical. However, it typically produces excessive vibration export which is simply inappropriate for direct use in the above stabilized sensor platforms. Since the active system of gimbal stabilization is dedicated primarily to maintaining the steady line-of-sight control by eliminating the relatively slow effects of yaw, pitch or roll, it may have insufficient resources to suppress excessive high-frequency vibration exported from the components of a cryogenic cooler. This may lead to an intolerable increase in the electrical power consumed by the servomotors, the controller over driving and essential degradation in the overall performance of the IR imager. The solution, which is presently adopted by industry, relies on “should-be-balanced” twinpiston compressors. In a simplistic approach, the twin-piston design comprises two identical linear compressors which are mounted back-to-back and have oppositely wound coils driven by a single AC power supply. In theory, such a device should show zero vibration export. In practice, however, the above two linear compressors are not perfectly identical and, therefore, balanced. For example, the vibration export from the twin-piston compressor of the Ricor Ltd. 1 Watt Linear Cooler (model K539, see http://www.ricor.com) in extreme case might be as high as 2.5 Nrms at driving frequency. Similar cryocoolers, such as the 1 watt linear cooler of DRS Technologies (see http:// www.drs.com), Hughes Linear Tactical Cooler (model 7050, see [8]), and the Litton Systems Inc. LLS cooler (see http://www.littonls.com) show almost similar levels of vibration export which are quite capable of upsetting the stabilization performance of the typical stabilized platform, as shown in this article. Further well-known drawbacks of the twin-piston devices are that they are extremely sensitive to spatial orientation, the action of external vibration and g-loading. Moreover, since they contain two identical linear compressors, they are at least twice as expensive to manufacture, typically consume 30% more electrical power and are less reliable and less protected from overstroking as compared with similar single-piston rivals. The more sophisticated active vibration cancellation in twin-piston cryogenic compressors relies on using two separate AC power supplies and an enhanced controller. The error signal is typically provided by motion sensors mounted upon the moving pistons or by an externally mounted load-cell which directly measures the dynamic force transmitted to the base [9, 10, 11]. In this approach, one of the compressors (slave) may be thought of as a counter-balancer with respect to the second one (master). The vibration export from such a compressor is normally very low, the price, however, is too high. In fact, the actively balanced cryogenic compressors find use in the critical space-borne applications where price is not a primary concern. The designers of electro-optic applications are currently looking for cheaper and more reliable approaches to controlling the vibration export from cryogenic compressors down to levels which would be acceptable for use in the design of the gimbaled stabilized sensor platforms. Vibration isolation is the simplest, most widespread and well-studied method of vibration export control [12, 13, 14]. Unfortunately, the severe vibration requirements cannot be met solely with a passive mechanical isolation. In [13,14], it has been shown that reasonable restraints imposed on the peak deflections of the vibration isolated cryogenic compressor under typical harsh
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wideband random vibration, shock and g-loading calls for a vibration isolator with a natural frequency close to or even higher than the driving frequency. This makes such an isolator inapplicable for the purpose of vibration control at the above driving frequency. The desired suppression of the self-induced force produced by the vibration-isolated compressor at driving frequency might be achieved by utilizing the principle of the tuned vibration absorber [12]. This approach is viable even if the natural frequency of the primary vibration isolator is higher than the driving frequency [14]. In this case, the primary flexural suspension is required, generally speaking, not for vibration isolation at driving frequency, but mainly for dynamic de-coupling of the compressor from the base and for the close control of peak deflections under the worse combination of wideband random vibration, shock and g-loading. On the other hand, in the high-frequency range, such a suspension behaves as a vibration isolator and produces the effect of essential attenuation of the higher-order harmonics. The authors present a practical implementation of the above approach to the design of a vibration protective interface which is based upon an all-metal vibration isolator and a tuned dynamic absorber allowing for maintaining persistent dynamic properties over a wide temperature range and lifetime. As is known, vibration isolators are typically poor heat conductors. Since the heat sinking from the cryocooler mounted inside the electro-optic application by natural or forced convection is not viable, the authors also developed a novel approach to heat sinking by using a commercially available and cheap self-lubricated bronze-backed and maintenance-free metallic plain bearing (patent pending). Finally, the authors show by example that the above approach yields sufficient control of vibration export from a linear single-piston compressor. From numerical simulation, robust and accurate stabilization has been obtained at practically no extra power consumed by the servomotor.
DESIGN OF VIBRATION PROTECTIVE AND THERMALLY CONDUCTIVE INTERFACE The authors in [14] have studied the dynamics of the vibration protective interface comprising an optimally damped and stiffened vibration isolator in combination with a tuned dynamic absorber. Following the developed approach, the optimal problem was formulated in terms of minimizing the vibration export from the linear compressor whilst maintaining its peak deflection within pre-specified tolerances under the worst combination of typical environmental vibration and constant acceleration. In this article, we deal with the vibration protective arrangement for the single-piston compressor of the 1 watt Ricor model K529H cryocooler operating at45Hz. Figure 1 and Figure 2 show the design schematics and pictures of the major components of the above mentioned arrangement. The tuned dynamic absorber, comprising a package of counterbalance masses tightened by screws and
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positioned upon the outer regions of two identical metallic flexure springs, is attached axially to the linear compressor, as shown in Figures 1 and 2. The above absorber was designed to have a mass of 100 g; this is 20% as compared with the total mass of compressor. The dynamic properties of the above dynamic absorber have been estimated using SDOF curve-fitting of experimentally measured absolute universal transmissibility, as shown in Figure 3a. The estimated natural frequency and loss factor are 45 Hz and 0.18%, respectively. The self-induced force developed by the linear compressor is unidirectional. This call for an application of a single-axis vibration isolator, which is implemented using a combination of flexure spring and two identical wire mesh rings manufactured from stainless steel, see Figure 1 and 2. The calculation of the optimal vibration isolator is based on the above mass ratio and the experimentally estimated loss factor of the dynamic absorber. Following [14], for a typical vibration profile, constant acceleration and allowed maximal peak deflection of 1 mm, the solution of the above optimal problem yields the optimized vibration isolator with undamped natural frequency of 45 Hz and loss factor of 15%. The details of this calculation, however, are beyond the scope of the article. In application, the above vibration isolator comprises the flexure spring and two wire-mesh rings, as shown in Figure 1 and 2. The dynamic properties of the vibration isolator have been estimated using a single-degree-of-freedom (SDOF) curve-fitting of the experimentally measured absolute universal transmissibility, as shown in Figure 3b. The estimated natural frequency and loss factor are 46 Hz and 11%, respectively; this is fairly close to the above desired values. The tuning of these parameters has been achieved by varying the flexure spring stiffness and squeezing the wire mesh rings. Figure 3c shows the transmissibility of the combined system, where both resonant peaks are closely controlled by damping in the primary vibration isolator [14] helping to prevent the excessive dynamic responses under external wideband vibration and shock. Further, the module of universal transmissibility of the combined system is smaller than unity in the frequency range above 70 Hz, thus showing that this vibration protective arrangement might be used for attenuating the vibration export at higher-order harmonics (see experiment below). The above arrangement maintains effective heat sinking from the cryogenic cooler, due to permanent contact between the linear compressor housing made of aluminum (hard anodized surface finish) and the metal plain bearing being thermally attached to a “cold wall” of the device. Such a design relies on a standard industrial INA Permaglide® PAP 4050 P11 (see http://www.ina.com) plain bearing, which is manufactured of bronze backing covered inside by an anti-friction and thermally conductive layer comprising porous bronze filled with a mixture of polytetrafluoroethylene and lead.
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The bearing is sized to provide a close sliding fit and guide for the cylindrical casing of the linear compressor. A heat conductive plain bearing housing, typically made of aluminum, provides a bore into which the plain bearing is inserted and retained by a press fit.
Experimentation on Attainable Performance The steady-state self-induced force developed by the linear single-piston compressor of the 1 watt Ricor cryogenic cooler (model K529H) operating under full heat load (power consumption 31.5W) was measured in the axial direction using the Kistler four-axis dynamometer Type 9272A and the Data Physics SignalCalc ACE Dynamic Signal Analyzer [14,15]. In Figure 4, reference case (a) corresponds to the rigid mounting, and case (b) corresponds to the flexural mounting of the above compressor. From Figures 4a and b, the fundamental harmonic component undergoes a 170-fold suppression (from 12 N rms to 0.07 N rms) due to the action of the dynamic absorber (this is comparable with the performance of the active vibration suppression system). The second harmonic component undergoes a 2-fold suppression due to the action of the vibration isolator. Similarly, we observe essential suppression of the higher-order harmonics, as predicted above. The thermal interface was experimentally tested when the cryocooler operated at full thermal loading while consuming 31.5W of electrical power. From experiment, the steady-state tempera-
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ture gradient between the compressor and the plain bearing housing skins was measured as 4.3 deg, indicating the overall thermal resistance of the above thermal interface to be as low as 0.14 deg/W. Figure 4c shows the reference spectrum of the steady-state self-induced force developed by the 1 watt linear dual-piston compressor of the Ricor model K539 cryogenic cooler operating at full heat load and, therefore, consuming 45 W of electrical power.
DYNAMIC MODEL OF THE MOTORIZED STAGE From the above experiment where the vibration export was measured (see Figure 4a,b) using a dynamometer mounted upon a heavy isolated base, we observed that the above vibration protective arrangement yields essential reduction of the vibration export from the single-piston linear compressor. However, the stabilized gimbaled platform is a relatively light and actively controlled device. Introduction of additional degrees of freedom related with the vibration isolated compressor and dynamic absorber alters the dynamic system and can interfere with the operation of the active stabilizer. Therefore, at this stage we need to enhance the dynamic model of the system in [14] by considering the multiple degree of freedom system of active stabilization where dynamic disturbance and control torque are applied at different locations. The primary goal of the further analysis is to ensure that the proposed vibration protective interface (i) does not cause any negative interference with the existing system of active stabilization and also (ii) yields better performance in terms of stabilizing accuracy and power consumption as compared to the reference case of the dual-piston rigidly mounted compressor. Figure 5 shows the simplified dynamic model of the actively stabilized gimbaled platform which is mounted in frictionless bearings The moment of inertia of the platform about the rotation axis is J and its absolute angular deflection is Q(t). The platform supports the linear cryogenic cooler, the linear compressor of which has mass and is vibration isolated from the platform, where the dashpot and spring represent the viscoelastic properties of the above isolator. The absolute linear deflection of the compressor from the position of static equilibrium is The self-induced force, F(t), originated from an unbalanced motion of internal parts of the above linear compressor [14,15] is directly applied to the compressor housing, as shown in Fig. 5. The tuned dynamic absorber of mass is supported from the compressor housing by a flexural member represented as a combination of the dashpot and spring The absolute linear deflection of the dynamic absorber from the position of static equilibrium is For suppressing the above self-induced force, the natural frequency of the dynamic absorber should be chosen to be equal to the driving frequency, and the damping should be as small as possible [12, 14]. The active closed-loop stabilizing system relies on the angular velocity transducer controller and servomotor We further assume that the velocity transducer has infinite bandwidth and neglect the friction in the gimbal bearings along with the noises produced by the transducer, control system, servomotor, structural resonances and mass imbalance.
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The above active stabilizing system contains a negative feedback loop for maintaining (i) the prescribed line of sight independent of the motion of the base and (ii) for suppressing the jitter caused by the environmental vibration and due to vibration export from the active system components (e.g. cryogenic cooler).
EQUATIONS OF MOTION The dynamics of the system in Figure 5 are governed by the set of differential linear electromechanical equations. The motion of the platform is given by the equation: in which the arm r is the distance between the force action axis (see Figure 5) and the rotation axis; the control torque developed by the typical DC servo-motor is: where is the motor constant, and I is the motor current. The motion of the compressor housing is governed by the equation: and the equation of motion of the dynamic absorber has the form: The electrical equation for a typical DC servo-motor is: where R and L are the active resistance and inductance of the servo-motor winding, and the driving voltage U is produced by the negative feedback in the system controller in the form: where are positive differential, proportional and integral coefficients of the PID controller, and G is the amplification ratio in the power amplifier. The problem of the steady-state stabilization is stated as follows:
where is the peak voltage available in the controller. The case of current limitation may be also considered. The analysis of stability may be performed on the basis of Eqs. (1) – (6). The design of the PID controller should be aimed primarily at providing the best available performance in maintaining the prescribed line of sight independent of the motion of the platform. More sophisticated controllers may also be considered. These issues are, however, beyond the scope of the present article.
SIMULATION is: is:
Figure 6 shows the system diagram based on Eqs. (1) – (7). The transfer function of the gimbal and the transfer function of the compressor with the dynamic absorber attached
where s is a complex (Laplace) variable. For reference purposes, we simulated the dynamics of the actively stabilized gimbaled platform carrying the rigidly mounted Ricor model K539 twin-piston compressor producing self-induced force, the spectrum of which is shown in Figure 4c. In our simulations, we considered further numerical values: r = 0.06 m, R = 10.2 Ohm, and L = 22 mH. The numerical values for stiffness and damping of the vibration isolator and dynamic absorber are calculated using experimentally estimated natural frequencies and loss factors using the formula:
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where and are experimentally estimated natural frequencies and loss factors of vibration isolator and dynamic absorber, respectively. For reference purposes, we simulated the dynamics of the actively stabilized gimbaled platform carrying the rigidly mounted Ricor model K539 twin-piston compressor producing self-induced force, the spectrum of which is shown in Figure 4c. To simulate rigid mounting with no dynamic absorber attached we temporarily assumed the stiffness of the vibration isolator to be very high and the mass of dynamic absorber to be very small. In simulations, the properties of the simplest PID controller were chosen in such a manner as to provide the settling time after typical disturbance to be less than 30 ms and minimized overshooting during typical transient process.
Rigidly Mounted Dual-Piston Compressor, Voltage-Limited Control In this simulation, the compressor was subjected to the action of the dynamic force, the spectrum of which is shown in Figure 4c and the peak voltage produced by the controller was limited to a typical value of 4V. Figure 7 shows the fragments of the simulated time histories of voltage and current developed by the controller along with the angular vibration of the stabilized platform in the steady-state mode. As the feedback voltage was limited, we clearly see that the voltage applied to the servomotor takes a rectangular shape. From simulation, the controller was over driven, and at steady-state mode consumed 28 W of electrical power. As a result of such a bang-bang control, the magnitude of the steady state vibration was 25 µrad rms. This accuracy is not sufficient for proper operation of the particular IR device.
Compressor, Voltage-Unlimited Control Further, we assumed that there were no limitations at all imposed on the values of voltage developed by the controller in the simulation. As in the preceding case, the compressor was subjected to the action of the dynamic force, the spectrum of which is shown in Figure 4c. Figure 8 shows the fragments of the simulated time histories of voltage and current developed by the controller in the steady-state mode. From Figure 8, the required peak voltage was in excess of 100 V, which is far above the available value. In the course of simulation it was estimated that in the steady-state mode the power consumed by the servomotor was 103.5W, and the residual angular vibration was 7 µrad rms.
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Figures 9a,b show the power spectral densities (PSD) of angular vibration in the two above cases, respectively. From the above simulations, we have found that the standard twin-piston compressor produces such a level of vibration export that the existing active control system, which has inherently limited resources, is not capable of precise stabilizing of the gimbaled platform. Increasing the supply voltage available for operation of the above active stabilization system yields nearly suitable stabilization, however, the level of power consumption is so high that heat sinking becomes a concern. The above two controls, evidently, are not viable.
Flexibly Mounted Single-Piston Compressor As an alternative to the above two cases, let us consider now the proposed vibration protective arrangement where the suppression of the self-induced force is due to me combined action of the optimally stiffened and damped vibration isolator and tuned dynamic absorber. In simulation, such a compressor produces a gross steady-state self-induced force, the spectrum of which is shown in Figure 4a. Figure 10 shows the simulated fragments of the time histories of the voltage and current developed by the controller in the steady-state mode, where the coefficients of the above PID controller remained the same as in previous simulations. In the course of simulation, it was estimated that the steady-state power consumption is as low as 2W, and the residual angular vibration is 0.8 µrad rms. Figure 11 shows the corresponding PSD of angular vibration (compare with Figures 9a and 9b). From the above simulations, the proposed vibration protective interface for a single-piston compressor yields a 30-fold and a 9-fold improvement in terms of stabilizing accuracy as compared with the above two reference cases of voltage-limited and voltage-unlimited control of gimbaled platform carrying a standard twin-piston compressor. Simultaneously, we obtained a 14-fold and a 52-fold improvement in terms of power consumption as compared with the above reference cases.
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CONCLUSIONS The developed and thoroughly studied novel vibration protective and thermally conductive interface allows a highly efficient, cheap and reliable linear single-piston compressor to be used within the infrared imager mounted upon the gimbaled precisely stabilized sensor platform without redesigning the existing systems of active stabilization and heat sinking.
REFERENCES 1. 2.
3. 4. 5. 6. 7. 8. 9. 10. 11. 12. 13. 14.
15.
Miller J.L., Principles of IR technology, A practical guide to the state of the art, Van Nostrand Reinhold, LA (1994). Coleman, “Infrared Sensors: The Eyes of the Digital Battlefield,” Military & Aerospace Electronics, July 1994, pp. 29-31. “F-117A vs. Radars,” see http://www.geocities.com/CapeCanaveral/Hangar/3558/pgradar.htm. “LITENING, Advanced Airborne Targeting and Navigation Pod,” see http://www.fas.org/man/dod101/ sys/smart/litening.htm. “Joint strike Fighter Sensor Suite,” see Cover Story at Raytheon web site http://www.raytheon.com/ coverarchive/cover008 (updated: November 13, 2000). “Space Based Infrared System,” see http://www.fas.org/spp/military/program/warning/sbir.htm Walker G., Cryogenic coolers, Part 2 – Applications, Plenum Press, New York (1983). Flint E, Flannery P, Evert M, Anderson E, “Cryocooler disturbance reduction with single and multiple axis active/passive vibration control systems,” published electronically at http://www. csaengineering.com/techpapers/techpapers.html, CSA Engineering, Inc. Collins S.A., Paduano J.D., “Multi-axis vibration cancellation for Sirling cryocoolers,” Proceedings of Cryogenic Optical Systems and Instruments Conference VI, April 4-5,1994, Orlando, FL, Proceedings of SPIE, Volume 2227 (1994), pp. 145-155. Johnson, D.L., Collins, S.A., Ross, R.G., Jr., “EMI Performance of the AIRS Cooler and Electronics,” Cryocoolers 10, Plenum Publishing Corp., New York (1999). Yeoung-Wei A Wu, “Stirling-cycle cryogenic cooler using feed-forward vibration control,” UK Patent 2279770,1995. Harris C.M., Crede C.E., (Eds.), Shock and Vibration Handbook, New York: McGraw-Hill Book Company (1976). Babitsky V.I., Veprik A.M., “Universal bumpered vibration isolator for severe environment,” Journal of Sound and Vibration, 218(2) (1998), pp 269-292. Veprik A.M., Babitsky V.I., Pundak N, Riabzev S.V., “Passive vibration protection of linear split Stirling cryocooler for airborne infrared application,” Shock and Vibration, vol. 7, no. 6 (2000) pp. 363-379. Riabzev S.V., Veprik A.M., Pundak N, “Technical Diagnostics of Linear Free-Piston Split Stirling Cryocooler through the Analysis of the Self-Induced Forces,” Proceedings of CEC/ICMC 2001 Conference, Madison, Wisconsin (2001).
Cryocooler Load Increase due to External Contamination of Cryogenic Surfaces R.G. Ross, Jr.
Jet Propulsion Laboratory California Institute of Technology Pasadena, CA 91109
ABSTRACT An important challenge for the cryogenic system designer of space-instruments is achieving stable emittance values for low-emittance cryogenic surfaces that may be affected by condensed water films or other outgassing products over the life of a mission. Typically, low-emittance cryogenic surfaces getter water films that increase in thickness over time and may lead to significant increases in the cryogenic heat load. If excessive load levels are reached, the cryogenic surfaces must be "defrosted" by raising them to elevated temperatures to evaporate the built-up contaminant films. As a help to those designing and conducting future long-life missions with cryocoolers, this paper summarizes the applicable physics associated with surface contamination and compiles example flight data on contamination effects experienced during multi-year space missions and ground tests. The flight and ground test data are then compared with the physical parameters involved with contamination transport rates and the dependence of emittance on contaminant film thickness. Although the test sample is small, the data indicate that around a 10-20% load increase may be expected on-orbit due to contamination of surfaces. This level is found to be consistent with the sensitivity of emittance to film thickness and the expected contaminant levels in space.
INTRODUCTION The cryogenic heat loads on a cryocooler can be roughly divided into three types: 1) Electrical power dissipation such as from a focal plane detector, 2) conductive parasitics down supports and electrical cables, and 3) radiation loads dependent on surface emittances and view factors of warmer objects that surround the cryogenic surfaces. In general, the electrical and conductive loads are relatively stable and predictable over the mission. However, the radiation loads are strongly dependent on surface temperatures and emittances that can change over time in less predictable ways. Key drivers for radiation loads are: The effective emissivity of the cryogenic surface area. Because the load is directly proportional to the emittance, very low emissivity levels can result in smaller loads. However, very small changes in the emittance will have a very large effect. The temperature of the warmer background that is radiating to the low-emissivity cryogenic surface. The radiation load varies as the fourth power of this background temperature. Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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The total cryogenic surface area subject to radiation; the load is directly proportional to the total cryogenic surface area. Given the above, the most sensitive parts of a cryogenic system will be high-surface-area parts at cryogenic temperatures that view high temperature surrounds, and utilize very low emittances to reduce the load. A difficult challenge faced by the cryogenic system designer is trying to achieve stable long-term emittances for these low-emittance surfaces as they are affected by contamination by condensed water films and outgassing products over the life of a mission. THE PHYSICS CONTROLLING CONTAMINATION SENSITIVITY As a precursor to examining measured rates of load increase in space, it is useful to first review the physics involved with contamination transport rates and the dependence of emittance on contaminant film thickness. Effect of Contaminant Film Thickness on Emittance Very thin deposited films of contaminant gases can significantly influence the emittance of low emittance surfaces such as polished gold or aluminum. Figure 1 describes the strong sensitivity of the emissivity of polished stainless steel to deposited films of various substances.1 Note that thin films of water-ice have a particularly pronounced effect, increasing the emittance of the polished metal surface at an initial rate of around per micrometer of water-ice thickness. Note also, that the emittance of water-ice films continues to increase and approaches unity at film thicknesses greater than 20 µm. One can use these data to estimate the allowable film thickness for water-ice on a low emittance cryogenic surface. For example, for a polished gold surface with a nominal emittance of 0.05, a water-ice film thickness of approximately 1.0 micron would be expected to increase the emittance by two or three times, and thus greatly increase the radiative load. Contaminant Film Transport Dynamics Given that very thin deposited films of contaminant gases can significantly influence the emittance of low-emittance surfaces, the next issue is the rate that such films build up—in particular the rate dependence on the partial pressure of the contaminant gas that surrounds the target surface, and the surface temperature.
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Saturation Vapor Pressure. The saturation vapor pressure of a contaminant gas is the pressure at which the gas is in equilibrium with its solid or liquid phase at the same temperature. As noted in Fig. 2, the saturation vapor pressure of gases is a very strong function of temperature—spanning many orders of magnitude over the range of cryogenic temperatures. Of particular interest to the problem of surface contamination, a contaminant gas will build a meaningful film on a surface only if the gas' existing partial pressure is greater than or equal to its saturation pressure at the temperature of the cryogenic surface. By meaningful, we mean the film will develop enough thickness to have a measurable effect on the surface emittance. Virgin surfaces will adsorb and hold many monolayers of gas because the water-metal bond is much stronger than the water-water bond. However, the equilibrium thickness of these few monolayers is not sufficient to have a measurable effect on the surface emittance. As an example, note in Fig. 2 that for a cryogenic surface at 145 K, a meaningful water film will deposit if the partial pressure of the surrounding water vapor is greater than likewise, no significant film will develop if the partial pressure is less than Or, from a cryopumping point-of-view, if one has an instrument enclosure at 145 K, the partial pressure of water vapor in the enclosure will be reduced to by the cryopumping action of the enclosure walls. Surface Residence Time. Another physical parameter associated with film development is surface residence tune. This is the average time between when a molecule arrives on a surface and when it leaves the surface. In other words, it is a quantification of the ability of a molecular species to outgas from a particular surface at a particular temperature. Figure 3 illustrates the surface residence time involving two types of molecular bonds: 1) water-water bonds, and 2) water-metal bonds.3 Note that water is extremely tightly locked onto virgin metal surfaces below 250 K, but also to surface films of water below 100 K. For such surfaces the residence time is several years, and the surface film of water is essentially immobile. Thus, once water arrives on a cryogenic surface, it is very difficult for it to leave and migrate to another surface. Predicting Typical Film Deposition Rates. For most applications involving cryocoolers, any cryogenic surfaces will be typically enclosed in a larger volume substantially warmer than the cryogenic surface itself. An example would be a cryocooler cold-tip enclosed in a room-temperature vacuum chamber, or a 60 K cryocooler cold-tip enclosed in a 150 K thermal shield. For such a case, the temperature of any residual gas in the volume will be approximately equal to the temperature of the chamber or shield wall, and any deposited film will be at the temperature of the cryogenic surface. Note from Fig. 3, that once a gas molecule hits a cryogenic surface, it is
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unlikely to return to the surrounding gas. The rate at which molecules stick to the cryogenic surface and contribute to the film growth is referred to as the cryogenic pumping rate; this rate has been extensively modeled in the literature based on the kinetic theory of gases.4,5 For our case, the residual gas pressure is typically > 10x higher than the saturation gas pressure on the cryogenic surface, so the net mass flux collected by the cryogenic surface simplifies to: where A P M T
= rate of mass collected (g/sec) = area of the cryogenic surface = external residual gas pressure (torr) = molecular mass of the residual gas (e.g. M=18 for ) = external gas temperature (same as external wall temperature) (K)
Equation 1 also applies to the rate of sublimation from a surface where the saturation pressure of the surface is much higher than (> l0x) the residual pressure in the surrounding volume. However, for this case T is the temperature of the cryogenic surface, and P is the film saturation pressure (from Fig. 2) corresponding to the temperature of the cryogenic surface. To estimate the rate of film thickness buildup, or loss, we must divide by the film density and area; thus where = =
rate of film thickness buildup (cm/sec) density of film (typical value for water ice is
)
To quantify the issue for our application, consider a polished gold surface at 80 K surrounded by water vapor at a partial pressure and temperature of 300 K. For a cryogenic surface at 80 K, Fig. 2 indicates that the saturation pressure is many orders of magnitude below the residual gas pressure of so Eq. 2 is applicable, i.e., the rate of film deposition is completely determined by the external gas temperature and pressure. Substituting these values into Eq. 2 gives Thus, around 5 minutes would be required to build up a 1 µm film of water with a corresponding emittance increase of Note that the rate of buildup is relatively independent of the cryogenic surface temperature once the saturation pressure at the surface temperature is well below the external vapor pressure.
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As a second example, one can compute what external vapor pressure of water would be required to keep the deposited film thickness growth to 1µm/year Using Eq. 2 and solving for P gives:
Thus, to achieve a long period without contamination, requires a very low residual partial pressure of water, or a high level of isolation of the cryogenic surfaces from the residual gas source. Predicting typical film deposition rates where the gas source is isolated from the cryogenic surface. For many applications involving cryocoolers, the cryogenic surfaces are carefully sealed off from sources of external contaminants using multilayer insulation or physical barriers. For such applications it is useful to understand how well the sealing must be to achieve a desired film deposition rate. For the typical case where the pressure outside the restriction is much greater than (> l0x) that near the cryogenic surface (the low pressure volume), we find that Eq. 1 again applies, but with the following change in definitions where A P M T
= = = = =
rate of mass passing through an orifice into the low-pressure volume (g/sec) area of the orifice gas pressure outside the orifice (external source pressure) (torr) molecular mass of the gas (e.g. M = 18 for ) gas temperature outside the orifice (external source temperature), K
Comparing Eq. 5 with Eq. 1 indicates that the mass flow rate and film buildup rate are reduced proportional to the area ratio of the orifice area to the cryogenic surface area. Thus, a x100 rate reduction can be achieved by restricting access to the volume containing the cryogenic surface via an orifice with an area of 1% of the cryogenic surface area. Emittance Degradation with Multilayer Insulation (MLI). For many applications involving cryocoolers, the low emittance surface is achieved using blankets made up of multi-layers of aluminized Mylar with internal spacer scrims to minimize direct film-to-film conduction. The various layers of the MLI thus operate as a stacked series of radiation shields at progressively increasing temperatures — from the cryogenic surface temperature to the surround temperature. Analyzing the effect of surface contamination on MLI is beyond the scope of this article, but the physics involved is the same, with the considerably added complexity of quantifying the water vapor migration rate through the MLI to get to the layers that are below the vapor's saturation temperature. Thus, the MLI acts as both a series of low emittance surfaces, plus a gas migration barrier to slow the deposition rate of contaminant films on the layers that are sufficiently cold. Unfortunately, the higher temperature layers of MLI blankets also serve as a considerable source of trapped surface moisture, and thus serve as a source of the water vapor capable of condensing on cold instrument optical and low-emittance surfaces.6
FLIGHT EXPERIENCE Although the above theoretical background illuminates the physics behind the contamination of low-emittance cryogenic surfaces, there are a number of parameters that remain unquantified in practice. These include such things as: What is the background partial pressure of water in space in the interior of an instrument, and how does it vary over the course of a long-term mission? What sort of typical emittance changes or cryogenic load increases are seen over a longterm space mission?
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To shed some light on these questions this section examines available data on the contamination effects experienced during multi-year space missions and laboratory life tests carried out over several months.
ISAMS Flight Experience One of the first large cryocooled instruments in space was the Improved Stratospheric and Mesospheric Sounder (ISAMS) instrument that was launched aboard the Upper Atmospheric Research Satellite (UARS) platform in September 1991. This application utilized two Oxford 80 K cryocoolers running at ~83% stroke to hold a number of cryogenic detectors near 80 K. Figure 4 highlights the cryogenic temperature increase experienced by this instrument.7 Immediately after turn-on in orbit, it can be seen that the temperature of the detectors increased by around 5 K (from 80 to 85 K) over a three-week period. This temperature increase, which is considered to be the result of contamination gettered onto the cold plumbing's low-emittance surfaces, necessitated conducting a number of high-temperature decontamination cycles. Over tune, the level of contamination slowly subsided, and after a year in orbit, the time required to reach a five-degree gradually increased to 2-3 months. This decrease over time confirms that the contamination was external-to and not internal-to the cryocooler. To estimate the level of load increase associated with the 5K one can appeal to the performance curves8 for the 80K British Aerospace (BAe) cooler shown in Fig. 5. From these curves it is seen that a temperature change from 80 to 85 K at a constant piston stroke (83% of 8mm = 6.64mm) equates to a load increase of around 100 mW per cooler, or a 12% increase in the total cryogenic load. Understanding the actual level of emittance increase would also be useful, but the data needed for this calculation were not available to this author.
MOPITT Flight Experience As a second example, the Measurements Of Pollution In The Troposphere (MOPITT) instrument initiated operation in March 2000 aboard the NASA EOS Terra spacecraft.9 This instrument uses two back-to-back BAe 50-80 K coolers to cool two detectors: #1 at around 92 K, and #2 at around 85 K. The instrument uses closed-loop control of the cooler stroke to maintain the #1 coldtip at a constant 78 K to control the #1 detector temperature, while the second cooler is forced to match the stroke of the first to maintain vibration control; the second cooler's cold-tip temperature thus varies somewhat.
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Figure 6 describes the stroke and temperature history of the MOPITT coolers during the first year of mission operation.10 Note the decontamination warm-up cycle that was carried out around day 190 to recover gradual deterioration of the instrument's science quality. During the first part of the mission it can be seen that the cooler stroke increased from 6.4 to 6.9 mm p-p to carry the gradually increasing cryogenic load. After the decontamination cycle, the load fully recovered to its original value and the cooler stroke returned to 6.4 mm. Note also that contamination after the warm-up cycle reoccurred at a much slower rate than the original contamination, thus confirming that the contamination was external to the cryocooler.
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To estimate the level of load increase that occurred, one can examine representative curves11 for the performance of the BAe 50-80K cooler as shown in Fig. 7. By plotting the MOPITT operating points on these curves it is seen that a stroke increase from 6.4 mm to 6.9 mm at the observed coldtip temperatures corresponds to a load increase of about 130 mW (8%) for cooler #1 and 230 mW (15%) for cooler #2. This is not too different from the ISAMS experience, although ISAMS chose to conduct its decontamination cycles at more frequent intervals. BAe 80 K Cooler Life Test Experience As a third example, a BAe 80 K cooler was life-tested in a vacuum bell jar at TRW in 1992 for around 160 days.12 During this lifetest the chamber was maintained with a vacuum around and the cryocooler coldfinger was wrapped with multilayer insulation (MLI). As shown in Fig. 8, the coldtip experienced a continuous temperature increase over the lifetest period at a near-constant rate of 5.48 K/year. Similar to the ISAMS and MOPITT performance, the
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performance degradation was traced to contamination of the external low-emittance thermal surfaces. From Fig. 9 it can be seen that 5.48 K/year at around 52 K equates to a radiation load increase of around 140 mW/year, again similar to the ISAMS and MOPITT experience. To evaluate the load increase in terms of emittance increase requires the surface area of the cold surfaces and the temperature of the hot surrounds. Figure 10 is a scale drawing of the BAe 80K coldfinger from which the MLI area has been estimated to be around The bell jar was at room ambient (296 K). Solving for the increased emittance gives:
If, as opposed to MLI, this system had a single low emittance surface, Fig. 1 suggests that the emittance increase would amount to a single-surface ice film thickness growth of around 0.7 µm/ Similarly, for a single surface, Eq. 4 can be used to estimate what the background water vapor pressure would have had to be for this rate of film growth. Thus
Since this partial pressure of water is well below the expected vacuum chamber water vapor pressure, it suggests that using multilayer MLI is much better than a single low-emittance surface in controlling the effects of contamination. The data also suggest that to fully stabilize low
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emissivity surfaces either requires very low partial pressures of water (say ), or more effective MLI involving many more layers that are kept contamination free. For ground life tests, one means of achieving a very low partial pressure of water is to shield the cryocooler coldend with a cryopumping shield with a temperature below around 120 K.
SUMMARY AND CONCLUSIONS A brief review of the physics of cryocondensation has been presented to provide insight into the in-space cryocontamination issue and its effects on low-emittance surfaces. However, appropriate values for key parameters such as the effective partial pressure of water in a long-term space mission and the sensitivity of MLI to moisture film buildup are not readily available. As one means of estimating these unknown parameters, the in-space performance of some previous cryocooler-based space missions has been examined. Although the mission sample is small, the data indicate that around a 10-20% load increase may be expected on-orbit due to contamination of surfaces. This is for a typical cryogenic space instrument with a total cryogenic load on the order of one watt. Historically, the established way to deal with this load increase has been to periodically boil off the contaminants by heating the cryogenic surfaces to near room temperature. This procedure has been found to work well, but the deep thermal cycling can also be very stressful to the systems involved. In addition, to provide sufficient cryocooler headroom to insure long uninterrupted operating periods between decontamination cycles, cryocoolers need to be appropriately oversized to accommodate this contamination-driven load increase. In the future, it would be desirable to incorporate unproved design features into cryogenic space applications that would greatly reduce or eliminate the degrading effects of on-orbit contamination. In the mean time, the issue of managing in-space contamination needs to be a focused part of the design and operation of any long-term cryogenic space application.
ACKNOWLEDGMENT The work described in this paper was carried out at the Jet Propulsion Laboratory, California Institute of Technology, and was sponsored by the National Aeronautics and Space Administration.
REFERENCES 1.
W. Viehmann and A.G. Eubanks, Effects of Surface Contamination on the Infrared Emissivity and Visible-Light Scattering of Highly Reflective Surfaces at Cryogenic Temperatures, NASA Technical Note TN D6585, NASA Goddard Space Flight Center, February 1972. 2. Barron, R.F., Cryogenic Systems, Oxford Univ. Press, New York, 1985, p. 225. 3. O'Hanlon, J.F., A User's Guide to Vacuum Technology, John Wiley & Sons, New York, 1989, p. 59. 4. Dushman, S. and Lafferty, J.M., Scientific Foundations of Vacuum Technique, John Wiley & Sons, New York, 1962, p. 14. 5. Barron, R.F., Cryogenic Systems, Oxford Univ. Press, New York, 1985, p. 455. 6. Boies, M. et al., “Measurement of long-term outgassing from materials used on the MSX spacecraft,” Proceedings of SPIE Vol. 4096 (2000), pp. 28-40. 7. ISAMS cooler data distributed by John Whitney, Oxford University Dept. of Physics, Oxford, UK, 19911993 (Private Communication). 8. Ross, R.G., Jr., Johnson, D.L., and Kotsubo, V.Y., BAe 80K Stirling Cooler Performance Characterization, JPL Internal Document D-9912, Jet Propulsion Laboratory, Pasadena, CA (1992). 9. Mand, G.S., Drummond, J.R., Henry, D., and Hackett, J., “MOPITT On-orbit Stirling Cycle Cooler Performance,” Cryocoolers 11, Kluwer Academic/Plenum Press, NY (2001), pp. 759-768. 10. MOPITT cryocooler year 2000 operations summary data distributed by G.S. Mand, Univ. of Toronto Dept. of Physics, Toronto, Canada, January 2001 (Private Communication). 11. Smedley, G.T., Mon, G.R., Johnson, D. L. and Ross, R.G., Jr., “Thermal Performance of StirlingCycle Cryocoolers: A Comparison of JPL-Tested Coolers,” Cryocoolers 8, Plenum Publishing Corp., New York (1995), p. 187. 12. TRW's BAe Cooler Lifetest Data, distributed by Bill Burt, TRW, Redondo Beach, CA, 1992 (Private Communication).
Performance Characteristics of the ASTER Cryocooler in Orbit M. Kawadaa, H. Akaob, M. Kobayashib, S. Akagib, T. Maekawac, O. Nishiharac, M. Kudohd and H. Fujisadae a
National Institute of Advanced Industrial Science and Technology, Tsukuba, Ibaraki, 305-8568, Japan b Mitsubishi Electric Corporation, Kamakura, Kanagawa, 274-8520 Japan c Fujitsu Limited, Kawasaki, Kanagawa, 211-8588 Japan d Japan Resources Observation System Organization, Chuo-ku, Tokyo, 104-0032 Japan e Sensor Information Laboratory Cop., Tsukuba, Ibaraki, 305-0005 Japan
ABSTRACT The Advanced Spaceborne Thermal Emission and Reflection Radiometer (ASTER) was developed by the Japanese Ministry of Economy, Trade and Industry (METI) for installation in the NASA EOS Terra spacecraft. The ASTER instrument consists of a visible and near-infrared radiometer (VNIR), a short-wave infrared radiometer (SWIR) and a thermal infrared radiometer (TIR). Two cryocoolers are required to cool the infrared detectors for the SWIR and the TIR subsystems. In the first functional checkout of the two subsystems in orbit, the SWIR cryocooler cooled the PtSi detector to an operating temperature of 77 K in a cooldown time of 22 minutes, and the TIR cryocooler cooled the MCT detector to an operating temperature of 80 K in a cooldown time of 23 minutes. The two cryocoolers have been smoothly working in orbit for over 2 years. The temperature of each detector has stabilized in the allowable temperature range. Long-term data have been acquired on the cooling performance and power consumption (or motor current) under normal operating conditions for each cryocooler, and they have also been evaluated based on the detector temperature and temperature stability in this phase. With the recognition that the coolers completed fabrication and qualification testing about four years before launch, and now have acquired over two years operation in orbit, it is clear that the performance requirements for the ASTER cryocoolers have been met. This paper summarizes both the ground test results prior to launch as well as the on-orbit performance characteristics of the ASTER cryocoolers.
INTRODUCTION The ASTER instrument, which consists of a visible and near-infrared radiometer (VNIR), a short-wave infrared radiometer (SWIR) and a thermal infrared radiometer (TIR)1,2, is required to have high sensitivity and high spatial resolution. To achieve these requirements, two cryocoolers Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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are used for cooling the PtSi detector of the SWIR and the MCT detector of the TIR. The mission life of the ASTER instrument is five years, and an operation life longer than five years is thus required for the ASTER cryocoolers.3,4,5,6 The goal in the development of the ASTER cryocooler was to achieve a durability of over 50,000 hours and mechanical vibration forces below 0.1 N in the frequency range from 40 Hz to 135 Hz in all three axial directions. A split-Stirling cycle cryocooler with clearance seals and linear electric motors was employed to realize the above goal. The twin-opposed piston configuration of the compressor unit was adopted as the driving mechanism of the compressor piston to eliminate mechanical vibration and compensate for the momentum. The mechanical vibration generated from the expander unit was reduced by an active balancer. Each cryocooler for the SWIR and TIR has a cooling capacity of 1.2 W at 70 K with power consumption lower than 55 W in normal operation. The development of two cryocoolers was completed in autumn 1995. After they were integrated in each subsystem, subsystem functional and performance tests including environmental tests such as thermal vacuum and vibration tests were carried out. Afterwards, the three subsystems of the VNIR, SWIR and TIR were integrated into the ASTER instrument system and system functional and performance tests were completed for the entire ASTER instrument. The EOS-AM1 (Terra) spacecraft was launched from Vandenberg Air Force Base, on December 18, 1999, and began collecting ASTER science data on February 24, 2000.
OUTLINE OF ASTER CRYOCOOLER Design Features and Evaluation Test Results of Key Technology Figure 1 schematically depicts the structures of the SWIR and TIR cryocoolers. Locations of sensors for measuring the temperature of the cryocoolers in orbit are shown in this figure. In addition to the temperatures shown in the figure, temperatures of the local radiator and dewar are monitored in orbit to maintain the performance of the SWIR cryocooler on the spacecraft. The temperature monitoring of the local radiator of the TIR cryocooler is also not shown in this figure. To achieve an operating lifetime of 50,000 hours, the ASTER cryocoolers utilize noncontacting clearance seals provided by a very narrow gap between the piston and the cylinder to avoid wear of seal materials. The suspension-spring flexures are made as thin metal disks that have spiral arms. Each piston of the compressor unit is supported by two sets of several suspension springs, as shown in Fig. 1, to maintain the clearance, concentricity and parallelism with the cylinder. The displacer piston and active balancer of the expander unit are also supported by suspension springs similar to those used to support the compressor pistons. Lifetime evaluation of the suspension springs is the most important issue in realizing longterm operation of the cryocooler. To meet the durability goal for the cooler, the suspension springs must be able to achieve 50,000-hour operation. The durability of the suspension spring design was examined by experiments using accelerated operation and stress analysis utilizing finite-element methods (FEM). The suspension spring design of the SWIR cryocooler was thoroughly examined repetitively of 109 times under a higher stress than the design value at an operation frequency of 45 Hz. The suspension spring of the TIR cryocooler was examined for 109 cycles for a higher frequency than the actual operation frequency of 40 Hz with constant stroke. The test results proved the suspension springs had sufficient endurance. Impurities contained in the working helium gas directly affect the cryocooler performance and reliability. The effect of contamination on the cooling performance was evaluated with a breadboard model (BBM) cryocooler. The presence of contamination was confirmed using samples of the same materials for each cryocooler. Moreover, the effect of the amount of moisture on the cooling performance was clarified experimentally. Based on the test results, contamination released from the materials used was a major limiting factor in the cooling performance. After assembly, a bake-out procedure was required to guarantee that there is no degradation due to contamination by outgases.
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Performance Characteristics of SWIR and TIR Cryocooler The cooling capacity of the SWIR cryocooler is a nominal value of 1.2 W at 70 K; the measured power consumption is 43.5 W, which satisfies the requirement that it be less than 55 W. The cooling capacity of the TIR cryocooler is a nominal value of 1.2 W at 70 K; the measured power consumption is 50 W, which satisfies the requirement that it be less than 55 W. The frequency characteristics of the induced vibration were measured using a load cell on the measurement fixture. The 45 Hz factors correspond to the operating frequency of the SWIR cryocooler, and the factors at 90 Hz and 135 Hz correspond to its upper harmonics. The 40 Hz factors correspond to the operating frequency of the TIR cryocooler, and the factors at 80 Hz and 120 Hz correspond to its upper harmonics. Figure 2 shows the test results of the vibration forces induced by a proto-flight model (PFM) cryocooler and the displacement of each unit. According to the test results, most of the force factors are less than the upper limit of 0.1 N. Compared with the peak thrust force, which is generated by the linear motors and even reaches 10 N to 100 N during normal operation, the measured vibration forces as a two-opposed configuration are sufficiently small. All the data, after conversion to displacement values, are of submicron order in the direction of the longitudinal axis. In the functional and performance test of the SWIR subsystem, the cryocooler was driven for
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1,160 hours and the number of start and stop (ON/OFF) cycles was about 210. In the functional and performance test of the TIR subsystem, the cryocooler was driven for 2,370 hours and the number of start and stop (ON/OFF) cycles was about 510. Subsequently, two cryocoolers were driven in various tests carried out before the launch. Figure 3 shows the cooling performance characteristics of the TIR cryocooler in thermal vacuum tests. The left side and the right side show the cooldown time before and after the thermal vacuum tests respectively in this figure. Numbers 1, 3, 5 and 7 show the test results of the cooldown time under low-temperature conditions below 21 °C for the expander unit in the figure, and numbers 2, 4 and 6 show the cooldown time under high-temperature conditions above 25 °C. A cooldown time of the cryocooler of 21-24 minutes was required within the
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environmental temperature range of 14 °C to 35 °C. Figure 4 shows the life test progress of the SWIR cryocooler. The transition of the power consumption of the cryocooler was observed during the life test under the condition that the cooling capacity of 1.2 W, the cooling temperature of 70 K and the mounting surface temperature of 40 °C were kept constant. As shown in Figure 4, the life test cryocooler has been running for 42,000 hours without any failure or degradation in performance, and continues to operate.
COOLING PERFORMANCE CHARACTERISTICS OF CRYOCOOLERS IN ORBIT The SWIR and TIR subsystems use cold plates supplied by NASA. The cold plate acts as a heat sink for each subsystem and the temperature is maintained in the range between 20 °C and 25 °C by a capillary-pump heat-transfer system (CPHTS). While the TIR expander unit is thermally controlled by a local radiator with a heat pipe, the SWIR expander unit employs a local radiator for thermal control. In fabricating the engineering model (EM), the configuration layout of the SWIR subsystem was changed from the vertical type to horizontal type. As a result, the thermal design for heat rejection from the expander unit was changed. The heat rejected from the expander unit would be radiated to the earth direction by the local radiator. The sensor locations for measuring the temperature of two cryocoolers are shown in Figure 1. Effect of the mechanical vibration is indirectly evaluated by analyzing image data, because it is not possible in orbit to measure mechanical vibration characteristics of cryocooler.
Cooldown Time of SWIR and TIR Cryocoolers The initial function checkout of the SWIR subsystem was carried out on January 7, 2000, and that of the TIR subsystem was carried out on January 9, 2000.7 Figure 5 shows the relationship between cooldown time and temperature of each detector. The operating temperature range of the SWIR detector is 77 ±0.3 K. The detector temperature reached 77 K in the cooldown time of 22 minutes. Based on the result of the initial function checkout of the SWIR subsystem, the cryocooler is concluded to have satisfied the required specifications from the temperature characteristic curve in the figure. In the thermal vacuum test of the SWIR subsystem, the cooldown time of the cryocooler ranged from 20 to 25 minutes. The cooldown time in orbit was almost equal to the result obtained in the thermal vacuum test. It is considered that there has been no performance degradation over the long term, because of the good coincidence between the cooldown time in orbit and the result of the thermal vacuum test. The operating temperature range of the TIR detector is 80 ± 2 K. The detector temperature reached 80 K in a cooldown time of 23 minutes. The cooldown time in orbit was almost equal to the result obtained under the low-temperature operation condition in thermal
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vacuum tests. Based on the result of the initial function checkout of the TIR subsystem, it is concluded that the cryocooler has satisfied the required specifications from the temperature characteristic curve in Figure 3. The TIR cryocooler monitors motor drive current but not power consumption. The data from the thermal vacuum test were acquired in May and June 1996. The motor drive current of the TIR cryocooler at some observation modes in the initial checkout in orbit showed a constant value, and the value was almost the same in the final test of the TIR subsystem.
Detector Temperature under Normal Operation Conditions Figure 6 shows the relationship between the operating temperatures of the SWIR and TIR detectors with elapsed time under normal operation in orbit. The specification for the operating temperature range of the SWIR detector is 77 ± 0.3 K. The measured temperature satisfies the required specification. From the result of analyzing the image acquisition data, there appears to be no effect of the mechanical vibration on the cryocooler.
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The temperature of the TIR detector was stabilized at approximately 80 K. The supplementary data contain more precise information, and it is possible to evaluate the operating temperature of the detector. In the long-term calibration performed from March 2000 to January 2002, the operating temperature and temperature fluctuation of the detector were 79.99 ± 0.02 K. From the results of analyzing the supplement data, it was confirmed that the cooling performance of the cryocooler satisfies performance requirements. There were, most likely, no mechanical vibration effects on the cryocooler, based on the analysis results of the image acquisition data.
ENVIRONMENTAL TEMPERATURE AND POWER CONSUMPTION OF CRYOCOOLERS UNDER NORMAL OPERATION CONDITIONS IN ORBIT Two cryocoolers of the SWIR and TIR subsystems monitor motor current but not power consumption. The relationship between environmental temperature of the cryocooler and the motor drive current is described.
Environmental Temperature and Power Consumption of SWIR Cryocooler The heat rejected from the compressor unit of SWIR cryocooler would be transferred to the cold plate by thermal conduction. The heat is transported from the cold plate to the CPHTS radiator, and is radiated from the radiator to space. The heat rejected from the expander unit would be radiated from the local radiator arranged in the panel that turns toward the earth. Figure 7 shows the temperatures of the expander unit, the compressor unit and the local radiator with elapsed time in orbit. The temperature of the expander unit reached 40 °C over 150 days in 2001. The temperature of the local radiator also increased similarly and the temperature difference with the expansion unit is constant. In the compressor unit, the temperature of the compressor body is measured in the neighborhood of the baseplate and the temperature of the compressor cylinder is measured in the neighborhood of the compression space of the working helium gas. Since the compressor unit is directly mounted on the baseplate, temperature of the compressor body is steady within the range of 30 °C to 32 °C. The temperature of the compressor cylinder increased linearly with time. Because the temperature of the expander unit had reached 40 °C, the temperature of the cold plate was changed at the setting point from 20 °C to 15 °C from July 24 through 26, 2001. The factor responsible for the temperature increase is described. The environmental temperature increased compared to the design value obtained by thermal design. Although the thermal design of the local radiator was performed based on the heat load in operating the SWIR subsystem at 8 percent of the duty value, it is operated at 100 percent actually due to the actual of SWIR scenario. The key factor responsible for the temperature increase is that the normal
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heat load had increased by continuous operation of a pre-amplifier. The local radiator received heat loads such as albedo and sunlight which change seasonally. The working helium gas temperature in the compression space is influenced by the temperature of the local radiator, and temperature of the compressor cylinder increases in the same manner as that of the local radiator. The SWIR cryocooler is monitored with the drive current of four linear motors. Two compressor pistons, an expander piston and an active balancer are driven by their linear motors. The drive current of each linear motor over time is shown in Figure 8 in the normal operation mode. The drive current of the compressor piston increases in the initial stage of operation in orbit. In the SWIR subsystem, prior to the start of operation, the temperature was maintained low. Then, the power consumption of the SWIR cryocooler increased, since the temperature of the local radiator increases when operation begins. It is seen that the characteristics of two motors are the same for the current measured values. Even when the SWIR cryocooler has been working at an environmental temperature of 40 °C in the normal operation mode, the detector temperature shows a constant value of 77 K as seen in Figure 6. It is confirmed that the SWIR cryocooler has a margin for the design specifications. Because neither the temperature increase of the detector nor the fluctuation of the detector temperature is observed, it is thought that there is no degradation in the cryocooler performance.
Environmental Temperature and Power Consumption of TIR Cryocooler Figure 9 shows the temperature of the TIR expander unit, baseplate and radiator in the normal operation mode. The heat rejected from the compressor unit of theTIR cryocooler would be transferred to the cold plate by thermal conduction. The environmental temperature of the compressor unit is controlled to be between 20 °C and 25 °C by the CPHTS. The feature of the TIR cryocooler is that the environmental temperature of the expander unit is in the range of 15 °C to 19 °C, and is maintained lower than that of the compressor unit. Because the radiator has a sufficient thermal radiation area and its operating temperature is maintained at about – 18 °C, the expander unit is maintained under low-temperature conditions. Figure 10 shows the drive current with time in the normal operation. For the TIR subsystem, prior to the start of operation, the temperature was kept low. Then, the power consumption of the TIR cryocooler increased, since the temperature of the radiator increases when operation begins. The drive current rose from 0.45 to 0.5 amperes with the temperature of the expander unit, and maintained a constant value afterward. The TIR cryocooler has been operating under low-temperature conditions, and the detector temperature shows a constant value of 80 K as shown in Figure 6. It is confirmed that the
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TIR cryocooler has a margin for the design specifications. Because neither the temperature increase of the detector nor the fluctuation of the detector temperature is observed, it is thought that there is no degradation in the cryocooler performance.
OPERATION RESULT OF ASTER CRYOCOOLER IN ORBIT Two cryocoolers of the SWIR and TIR subsystems are driven at different environmental temperatures as mentioned above. That is, the expander unit is driven under high-temperature conditions for the SWIR cryocooler. On the other hand, the expander unit is driven under lowtemperature conditions for the TIR cryocooler. Although the operation time record in the orbit of the two cryocoolers amounts to over 19,000 hours, no performance degradation is observed. The SWIR obtains images of six bands in the short-wave infrared region with a spatial resolution of 30 m. The TIR obtains images of five bands in the thermal infrared region with a spatial resolution of 90 m. Large amounts of clear image data were obtained in the normal operation mode in orbit of the SWIR and TIR subsystems. From the results of the image data analysis, the vibration forces induced by the ASTER cryocooler appear to be negligible. The temperature of the cold plate (CPHTS), temperature information on each subsystem and image acquisition data must be monitored in order to evaluate the cooling performance.
SUMMARY In this paper, the performance of the ASTER cryocooler in orbit was described. Two cryocoolers of small-size, lightweight and low mechanical vibration was realized for the ASTER instrument. The coefficient of performance of the ASTER cryocooler is more efficient than that of the other one. It was confirmed that there was no performance degradation by comparing the results of the initial checkout in orbit and the final test on the ground. After the
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initial checkout, the operation of each subsystem was changed to the normal observation phase. The operational result in orbit of the two cryocoolers was experienced over 19,000 hours, and no performance degradation was seen. Each cryocooler has sufficient performance levels, satisfying design specifications as seen from the results of the current data analyses. For the two cryocoolers, the long-term operation in orbit is expected, because mechanical failure did not occur. Long-term data have been acquired on cooldown performance and power consumption in the normal operation mode of each cryocooler, and the detector temperature and its temperature stability in this phase have been evaluated. The vibration forces induced by the ASTER cryocooler have been evaluated from the results of the image data analyses.
ACKNOWLEDGMENT The development of the ASTER was carried out as a project of the Ministry of Economy, Trade and Industry (METI). REFERENCES 1. A.Akasaka et al., "Short Wavelength Infrared (SWIR) Subsystem Design Status of ASTER" , Proc. of SPIE”, vol. 1490 (1991), pp.269-277. 2.
Y. Aoki et al., "Thermal Infrared Subsystem Design Status of ASTER”, Proc. of SPIE, vol. 1490 (1991), pp. 278-284.
3. M. Kawada and H. Fujisada, "Mechanical Cryocooler development Program for ASTER", Proc. of SPIE”, vol. 1490 (1991), pp. 299-308. 4.
M. Kawada et al., "Long-Life Cryocooler Development Program for ASTER”, Cryocoolers 8, Plenum Press, New York (1995), pp. 35-41.
5.
M. Kawada et al., "Verification Test and Evaluation of Vibration Characteristics for ASTER Cryocooler, " Proc. of SPIE, 2553 (1995), pp. 272-2287.
6. M. Kawada et al., "Performance of long-life Stirling cycle cryocoolers for ASTER instrument”, Proc. of SPIE, vol. 3221, (1997), pp. 220-229. 7. M. Kawada et al., “In-flight performance of the ASTER cryocooler”, Proc. of SPIE, vol. 4169 (2000), pp. 88-97.
AIRS Pulse Tube Cooler System-Level and In-Space Performance Comparison R.G. Ross, Jr. Jet Propulsion Laboratory California Institute of Technology Pasadena, CA 91109
ABSTRACT During this past year, JPL’s Atmospheric Infrared Sounder (AIRS) instrument completed thermal vacuum testing at the spacecraft level, was launched into Earth orbit on NASA’s Earth Observing System Aqua platform on May 4, 2002, and cooled down to operating temperature on June 12, 2002. The instrument, which is designed to make precision measurements of atmospheric air temperature over the surface of the Earth, uses a redundant pair of TRW pulse tube cryocoolers operating at 55 K to cool its sensitive IR focal plane. This use of redundant coolers with no thermal switches creates certain challenges when it comes to performance measurement and verification at the systems level. In support of the mission, new test and analysis techniques have been developed to allow accurate intercomparison of cooler-level, instrument-level, spacecraft-level, and in-space performance of the coolers. The driver for the development of these techniques is the strong dependency of the performance of the cooler system on gravity. Specifically, the off-state conductance of the nonoperating redundant cooler is highly dependent on convection, and thus gravity level and orientation. Because the instrument-level, spacecraft-level, and in-space environments have substantially different orientations and gravity levels, special test and analysis techniques had to be developed to allow accurate intercomparison of the results. This paper presents the derivation of the test and analysis techniques as well as the measured system-level performance of the flight AIRS coolers during instrument-level, spacecraft-level, and in-space operation.
INTRODUCTION A critical aspect of the integration of a cryocooler into a high-value application, such as a multiyear-life spacecraft, is monitoring and confirming the health of the cooler system after each important stage of integration and qualification testing: 1) after completion of the cooler itself and its qualification testing, 2) after integration of the cooler into its immediate application (instrument) and the instrument-level qualification testing, 3) after integration of the overall instrument onto the spacecraft and spacecraft-level qualification testing, and 4) after launch of the entire instrument/spacecraft system into space. Initial testing of a development model cooler is generally relatively straightforward as such coolers may be fully instrumented with few constraints on access to measurement data such as truerms input power to the compressor, or heat applied to the cold-load interface. In contrast, a flight Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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model cooler often will come with space-quality drive electronics that provide little or no ability to accurately assess the internal distribution of power such as that going to the compressor. And once the cooler is integrated into the instrument, the details of the cryogenic load generally become obscured by the dozens of unquantified load contributors and test environment variables such as MLI effectiveness, vacuum level, and background temperatures. In addition to these typical performance verification issues, the AIRS instrument has a redundant pair of cryocoolers, and the thermal conduction loss to the off cooler, which equals ~40% of the total cryogenic load, is gravity and orientation dependent. This is because the conduction load from the off cooler is dominated by convection within its pulse tube, and the gaseous convection is highly sensitive to orientation with respect to gravity, and thus to the presence of gravity. The problem of cooler performance verification at the various stages of cooler, instrument, and spacecraft integration is further compounded by the inevitable fact that the cooler orientation is likely to be, and was for AIRS, different in each test, and was of course different again in space. The focus of this paper is on how these issues of performance verification and tracking were successfully dealt with for the specific example of the AIRS instrument and its redundant pair of pulse tube cryocoolers.
AIRS INSTRUMENT CRYOGENIC AND CRYOCOOLER DESIGN The technical foundation of the AIRS instrument is a cryogenically cooled infrared spectrometer that uses a pair of TRW 55 K pulse tube cryocoolers to cool the HgCdTe focal plane to 58 K.1 The instrument also includes a 150 K-l90 K two-stage cryogenic radiator to cool the optical bench assembly to around 150 K. Configurationally, the 58K 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 190K thermal radiation shield and MLI blankets. The ambient portion of the instrument contains the high power components including the instrument electronics and the cryocoolers and their electronics. Figure 1 schematically illustrates the cryogenic elements of the AIRS instrument design. As noted, the system incorporates two independent 55 K pulse tube cryocoolers, a primary and a nonoperating backup, each connected to the 58 K 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
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loads is the BOL/EOL temperature of the optical bench and pulse tube housing—assumed to be 145 K/160 K and 309 K/314 K, respectively. To provide an initial performance benchmark, extensive characterization of the AIRS cryocooler performance was carried out during the cooler development and qualification testing phases at TRW and JPL.2 Of particular interest is the measurement of the cooler's off-state conduction as a function of orientation angle, shown in Fig. 2, and the cryocooler overall system-level refrigeration performance shown in Figs. 3a and 3b for a heatsink temperature of 25°C. Because the AIRS coolers' pulse tubes face in opposite directions, the preferred orientation for testing is with the off cooler's pulse tube facing up as noted in the drawing in Fig. 2. This orientation inhibits convection and gives the ~0.5 watt conduction load expected in space as noted in Table 1. Note in Fig. 2 that when the off pulse tube is inverted or horizontal, the measured conduction load is double or triple the 0.5 watt value due to convection effects.3
AIRS Cryocooler Instrument Level Testing During instrument-level testing the cryocoolers' performance was repeatedly characterized to provide a prediction of the expected space performance, and to confirm that no degradation had occurred to either cooler due to imposed qualification test environments.4 However, because of test geometry limitations, the pulse tube orientation was either fixed with the cooler A pulse tube vertically up, or with both pulse tubes horizontal. Thus, the data interpretation was immediately faced with the large and uncertain angle-dependent conduction load noted in Fig. 2. The means used to
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resolve the large gravity-dependent load increase was to run both coolers simultaneously, unlike their operation in space. The challenge was thus to predict the coolers' eventual in-space performance and health from a test configuration that differed significantly from the space operating conditions.
TEST METHOD OVERVIEW The test methodology developed for assessing cryocooler performance is fundamentally based on comparing the "measured cryogenic heat load" at any point in the integration and test flow against a "baseline predicted heat load" derived from analytical modeling calibrated using heritage measure-
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ments as presented in Ref. 2. The "baseline predicted heat load" is that shown in Table 1, and involves contributions from the focal plane dewar assembly, from the cold rod and its flex braid assembly, and from the coldblock region of the pulse tubes themselves. As usual, the devil is in the details, and in this case the detail is how the "measured cryogenic heat load" is measured when the second redundant pulse tube is presenting a large unknown heatload contribution. The fundamentals behind the methodology are most easily seen by appealing to Fig. 4. This figure presents an electrical circuit analog for the interconnection of the two pulse tubes and the cold link assembly in the AIRS instrument. The resistor network flowing from the primary cooler through the redundant cooler represents the 'off cooler' thermal resistance to the ~313 K pulse tube housing temperature In particular the resistor represents the highly variable and thus unknown thermal resistance of the redundant pulse tube as noted in Fig. 2. Appealing to Fig. 4, it is also seen that the affect of the redundant cooler on the heat load presented to the primary cooler is fully defined in terms of the resistors and and the pulse tube coldblock temperatures and independent of the value of The temperatures and are indeed known, and are accurately measured in flight as part of the AIRS cooler closedloop temperature control operation of the two coolers. Thus, we have a way of quantifying the load presented by the off cooler by monitoring the between the two pulse tube coldblocks. Because this is not always controllable, the performance assessment methodology uses as a characterization variable as shown in the detailed steps enumerated below.
Detailed Performance Measurement Steps The rationale and details of each of the measurement steps are described below: 1) Achieve Baseline Operating Conditions. The cryocooler drive levels and heat rejection temperatures are first adjusted to achieve the baseline flight operating conditions as defined below: Cryocooler cold plate temperature: 293 ±10K (preferably 293±1K) Spectrometer Temperature: 155 ±10 K (preferably 154±1K) Primary pulse tube temperature: 56 K±0.1K Redundant pulse tube temperature to achieve Focal plane temperature: 58K±2K Because this baseline set of operating conditions is not totally controllable, the critical parameters have been given a range and a preferred target value. The reference baseline load is then adjusted for the actual test conditions achieved, as described in step 3. 2) Heat Load Determination. No instrumentation is available to measure the total cryogenic heat load in the AIRS instrument. However, if the cooler performance has not degraded, the cooler input drive level, which is monitored in the flight system, can be used to estimate the cryogenic load. This requires use of the previously measured performance characteristics of the coolers as presented in Figs. 3a and 3b, with an appropriate correction applied if the heat rejection tempera-
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ture is different from that used in the plots. Specifically, to determine the total cryocooler heat load, a point is placed at the intersection of the cooler %Drive applied to the primary cooler and the effective coldblock temperature (T*) of the primary cooler. The effective coldblock temperature (T*) is the ~56 K coldblock test temperature adjusted to correct for compressor heat rejection temperatures that are different from the 298 K baseline for which Figs. 3a & 3b were generated. T* is defined in terms of the measured pulse tube coldblock temperature of the primary cooler and the compressor cold plate temperature by the following equation:
The total cryogenic heat load is then read from the plot abscissa directly below the plotted point. Equation (1) was derived using the measured relationship between heat rejection temperature and coldblock temperature shift as described in detail in Ref. 5 3) Baseline Load Determination. The AIRS cryocooler "baseline predicted heat load" is the reference heat load used for comparison against the "measured" heat load in any given verification test. In particular, it is the data contained in Table 1 adjusted for the effects of the actual test temperature of the AIRS optical bench (dewar flange), and the actual test temperature of the AIRS cryocooler pulse tube housing. Because the pulse tube housing temperature is not measured directly, this temperature can be written in terms of the cryocooler heat rejection (cold plate) temperature and the housing's 15°C measured temperature rise above the cold plate temperature. In equation form, the baseline heat load can thus be written as:
where:
Focal plane emittance (~0.04) Sapphire rod emittance (~0.04) Flexlink MLI emittance (~0.06) Pulse tube MLI emittance (~0.04) Dewar flange temperature (~155K) Pulse tube housing temperature (cold plate temperature + 15 K)
When the baseline parameter values shown above are used, Eq. 2 reproduces the baseline cryocooler load components presented in Table 1. The equation is used to refine the baseline cryocooler load for the actual dewar flange and cooler cold plate temperatures used in a given test. 4)
The between the "measured cryogenic heat load" and the "baseline predicted heat load" is the critical performance measure for a given test condition. This is just a subtraction of the results of step 3 from the results of step 2. The result may then be plotted on a plot of versus pulse tube temperature differential
5) Coldend Heatload Assessment. As an additional check on the system health, the temperature difference between the focal plane temperature and the average temperature of the two pulse tube cold blocks is computed. Given the fixed thermal resistance of the coldrod assembly and flexbraid, this value provides an assessment of the dewar heatload conducted down the coldrod assembly that can be compared with heritage values. For AIRS, this focal plane temperature rise above is around 1.0 K.
AIRS COOLER DATA ANALYSIS Using the above methodology, a number of AIRS cooler performance measurements have been analyzed using data acquired: 1) during AIRS instrument-level test at Lockheed Martin (LMIRIS) in Lexington, MA in 1999, 2) during instrument checkout in the Aqua spacecraft thermal vacuum testing at TRW in Redondo Beach, CA during September 2001, and 3) after AIRS cooler turn-on in space on June 12, 2002. During these tests the pulse tubes had a variety of orientations with respect to gravity, and in many cases both pulse tube coolers were run simultaneously to defeat the large convective heat load that would have occurred if the redundant cooler had been turned off.
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Figure 5 provides an intercomparison of these three data sets made by plotting the difference between the measured cryocooler heat load and the baseline heat load against the pulse tube differential temperature as described earlier. Note the strong and uniform dependence of the measured performance on pulse tube temperature differential and the ability to easily interpolate the expected in-space performance for each data set. Understanding the performance of the primary operating cooler as a function of this temperature differential was an important accomplishment that enabled comparison of these diverse test conditions. Note that the data measured during spacecraft thermal vacuum (T/V) testing are consistent with, and slightly better than the cooler performance measured at the instrument level two years
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earlier. This ~70 mW improvement in performance may be partially due to the careful effort made to outgas the Aqua S/C and the AIRS instrument prior to cooling the spectrometer and coolers in the large spacecraft T/V chamber at TRW. Additionally, these tests were conducted immediately after spectrometer cooldown, perhaps before any long-term contamination had a chance to build up on the low-emittance coldend surfaces. Of course the most important data are the post-launch performance measurements plotted at the measured differential temperature in space of 2.2 K. These data, which have also been added to the cooler performance plots in Fig. 3, confirm that the coolers are performing in-orbit as designed, with a total cryogenic heat load at beginning of life within 25 mW of the predicted baseline value. Also, the data confirm that the in-space pulse tube off-state conduction is well approximated by the off-state conduction measured on the ground with the off pulse tube oriented vertical with its hot end up. Lastly, as a final check on the system health, the temperature difference between the focal plane temperature and the average temperature of the two pulse tube cold blocks was measured in space. This too agreed with the 1.0 K measured earlier during spacecraft thermal vacuum testing and confirms the post-launch health of the AIRS instrument and coolers. As the AIRS space mission progresses, the total cryogenic load can be expected to increase somewhat due to surface contamination effects.6 Monitoring this expected load increase will be a future use of the detailed methodology presented here.
SUMMARY In summary, a test methodology has been developed that allows the thermal performance of a complex flight system to be monitored through changing environmental conditions using only the flight telemetry data. The needs and complexity of such performance monitoring are often overlooked during the initial instrument design phase, and may only surface late in the testing process. This was the case for AIRS, and it is hoped that the successful lessons learned presented here will be of value to future cryogenic applications. Most importantly, the acquired data confirm that the AIRS coolers and dewar are performing to specification and are ready to support the flight mission's science objectives.
ACKNOWLEDGMENT The work described in this paper was carried out at the Jet Propulsion Laboratory, California Institute of Technology, and was sponsored by the NASA EOS AIRS Project through an agreement with the National Aeronautics and Space Administration. Special thanks are due Dean Johnson of JPL, and Ken Green and Ken Overoye of BAE Systems (formally Lockheed Martin LMIRIS), who spent long hours helping to acquire the data described here.
REFERENCES 1 Ross, R.G., Jr. and Green K., "AIRS Cryocooler System Design and Development," Cryocoolers 9, Plenum Publishing Corp., New York, 1997, pp. 885-894.
2. Ross, R.G., Jr., Johnson, D.L., Collins, S.A., Green K. and Wickman, H. “AIRS PFM Pulse Tube
3. 4. 5. 6.
Cooler System-level Performance,” Cryocoolers 10, Plenum Publishing Corp., New York, 1999, pp. 119-128. Johnson, D.L., Collins, S.A., Heun, M.K. and Ross, R.G., Jr., "Performance Characterization of the TRW 3503 and 6020 Pulse Tube Coolers," Cryocoolers 9, Plenum Press, New York (1997), pp. 183-193. Ross, R.G., Jr., Test Results of the AIRS Cryocooler / Dewar Heat Load T/V Test at TRW, AIRS Design File Report ADF# 541 (Internal Document), Jet Propulsion Laboratory, September 13, 2001. Ross, R.G., Jr. and Johnson, D.L., “Effect of Heat Rejection Conditions on Cryocooler Operational Stability,” Advances in Cryogenic Engineering, Vol. 43B, Plenum Publishing Corp., New York, 1998, pp.1745-1752. Ross, R.G., Jr., “Cryocooler Load Increase due to External Contamination of Cryogenic Surfaces,” Cryocoolers 12, Kluwer Academic/Plenum Publishers, New York (2003).
Final Qualification and Early On-Orbit Performance of the RHESSI Cryocooler R. Boyle, S. Banks, and K. Shirey NASA/Goddard Space Flight Center Greenbelt, MD 20771, USA
ABSTRACT The Reuven Ramaty High Energy Solar Spectroscopic Imager (RHESSI) spacecraft was launched on February 5, 2002. It now observes the Sun with the finest angular and energy resolutions ever achieved from a few keV to hundreds of keV, using an array of nine germanium detectors operating at 75 K. The spacecraft was originally scheduled for launch in July 2000, but a vibration facility mishap damaged the primary structure of the spacecraft, along with the cryocooler. This paper describes issues in the qualification of a replacement for the original flight cooler, and describes early on-orbit performance.
INSTRUMENT DESCRIPTION The High Energy Solar Spectroscopic Imager (HESSI) spacecraft (Figure 1) was selected by NASA in 1997 for its Small Explorer (SMEX) program. It was built by a team led by the University of California at Berkeley (UCB). Launched in February 2002 after a number of delays, and renamed RHESSI, it should observe thousands of flare events occurring near and slightly after the peak of the eleven-year solar cycle during its planned two years of observations. The spectrometer uses an array of nine large germanium detectors, mounted in a cryostat on a common 75 K coldplate. While the detectors themselves have no measurable dissipation, each detector requires two FET amplifiers operating below 150K, each dissipating approximately 30mW. An off-the-shelf Sunpower M77B Stirling-cycle cryocooler was chosen for the mission. This cryocooler has a pneumatically driven displacer, and an integral counterbalance motor that can be used for vibration attenuation. (Figure 2). The compressor and displacer are each supported on gas bearings, enabling a very long service life. Copper fins intended for air cooling have been machined away to allow for a conductive thermal interface. Mounting tabs that were brazed to the outer housing were removed, and a support cradle was used in its place.1 The M77 cryocoolers were vibrated to 14.1 grms without any apparent ill effects. The cryocooler was designed for long life, and in fact the first flight model eventually ran for 17000 hrs. But mishaps during instrumentlevel vibration testing raised a number of questions about the coolers, and a significant effort was eventually required to qualify a second unit for flight.
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Engineering Test Unit Cooler (ETU) The engineering test unit (ETU) cooler was integrated with the ETU spectrometer in March 1999, and demonstrated that the system provided satisfactory thermal performance. The ETU spectrometer was then damaged during its first vibration qualification test due to assembly problems. Though the cryocooler seemed at first unaffected, its performance after the vibration test was qualitatively different than before. It required slightly more power to reach operating temperature, and the coldfinger temperature showed an instability which was at first attributed to electrical noise on the sensor. But the ETU cryocooler remained in operation for six months, before finally being retired, along with the ETU spectrometer assembly. At the time, the loss of performance was attributed to changes in thermal blanketing in the spectrometer, which had occurred at the same time as the vibration test. There was no investigation of any damage to the cooler at the time it was retired.
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First Flight Model Cooler (FM1) The first flight model cooler (FM1) was integrated into the FM spectrometer in September 1999, and immediately put into an instrument-level vibration test. Subsequent thermal performance was satisfactory, and the cooler continued in operation through spacecraft integration and test. On March 21 2000, the spacecraft was subjected to an accidental over test during final vibration qualification.2 The solar panels and the telescope support ring were broken, and the main deck of the spacecraft was deformed beyond allowable tolerances. The spacecraft was disassembled to the box level for damage inspection, and the spectrometer was found to require significantly higher power for the cryocooler to achieve operating temperature. In addition, the performance of the cryocooler had become much more sensitive to its orientation relative to gravity, and the coldfinger temperature oscillated between two stable operating points 3K apart. The decision was quickly made to replace the FM! Cooler, but the backup cooler, then in final qualification testing, was found to be overstroking at high power levels that had not previously been a problem. A search began for a replacement unit that would be eligible for flight.
RECOVERY EFFORT The ETU and FM1 cryocoolers had come from an initial batch of eight M77B cryocoolers manufactured in 1994. Of the other six units, two were damaged in removing the copper cooling fins, two were damaged during bakeout operations, one had a counterbalance that was damaged during development of vibration control software, and one developed an intermittent electrical problem. A second batch of eight coolers, M77C units without integral counterbalancers, was produced in 1999, but problems with the cleaning process at GSFC apparently caused other problems in the cooler, and all eight, including the first flight backup cooler, were eventually excluded from flight. At the time of the spacecraft mishap, though, all these coolers were considered as options for flight, plus a batch of four additional coolers that the Naval Research Lab had ordered for its own flight experiments. New coolers were put on order from Sunpower, and a “Tiger Team” was assembled to examine all of RHESSI’s cryocooler options. Due to the dependence of RHESSI’s science on the timing of the solar cycle, it was particularly important that the cooler situation be resolved as quickly as possible. Failure Mode Analysis In selecting a replacement for the FM1 cryocooler, the Tiger Team was concerned about possible failure modes in the M77 coolers, including: 1. Mechanical and structural failures, such as the internal support structure, compliant feed through rods, spring assemblies, or the loosening of mechanical attachment hardware (i.e. screws and nuts). 2. Contact of moving parts, either piston or motor, with stationary cryocooler structure, such as the cylinder and feed through surfaces. 3. Contamination of the internal helium working fluid either by water or particulates. 4. Electrical failure (either an electrical short or open). 5. Working fluid leakage from the cryocooler structural housing or working fluid fill tube. To address the mechanical failure modes, the ETU and FM1 coolers were inspected for mechanical and structural failures after removal from the RHESSI spectrometers. The EM cooler coldfinger was out of alignment with the rest of the assembly by about 0.25°, or 0.3mm at the cold end, and the FM1 coldfinger by about 0.09°, or 0.1mm. It was not possible to determine if the coldfinger had been deformed, or if internal components had shifted. One nut had come loose in the FM1 cooler, on a joint in the counterbalance assembly that was not critical to alignment. Analysis of the vibration loading on the coldfinger suggested that there was a large margin on the yield strength of the coldfinger, but no verification testing was performed on an actual coldfinger. The
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team felt that it was possible that the coldfinger misalignment had caused the loss of thermal performance, but was unable to positively conclude what had caused it. The Team considered a number of different ways to check the cryocoolers for contact between the moving and stationary parts, but could identify no feasible technique. Low-frequency stiction testing is not possible with these coolers, since the gas bearings are not functional at low amplitude / low frequency, and there is no position sensor to indicate piston position. X-ray examination does not give adequate resolution of the materials in the coldfinger. Vibration measurements of the cryocooler body do not show a distinct signature of touch contact among the large existing forces of normal vibration. The team did commission an analysis of the gas bearings, and found that the bearings should be able to adequately support the pistons under one-G operation, as well as in the 15rpm rotational environment of the spacecraft. Prior to disassembly the ETU and FM1 coolers were run in a variety of orientations, showing that thermal performance was not only dependent on whether the coldfinger was pointing up or down, but even on which side of the cryocooler was down when the cryocooler was run horizontally. This sensitivity to roll angle, as well a test with the ETU in which weights were hung on the coldtip during operation, strongly suggested that the components in the coolers were rubbing in certain operating orientations. This roll angle sensitivity became one test for touch contact in flight candidate cryocoolers . To address contamination, the working gas in the ETU and FM1 coolers was sampled with a residual gas analyzer. Small amounts of CO and were found, but no measurable water. While the gas contamination levels did not seem to have affected the thermal performance, there was a significant amount of dust buildup in the assemblies, probably due to touch contact during wear-in of the machines. The flow impedance of the gas bearings had changed from the initial manufacture, possibly due to accumulation of dust in the bearing ports. It seemed possible that dust deposits might accumulate benignly, only to be kicked loose into a bearing port during a vibration test. While this scenario was not eliminated as a cause of the anomalies with the ETU or FM1 coolers, it seemed unlikely to affect the FM2 cooler, which had already been vibrated four times prior to launch, with no apparent ill effect. No indications were found in the ETU or either of the FM coolers of any electrical problems, or any helium leakage from the pressure boundary of the machine.
Selection of FM2 After consideration of all the available M77 coolers, extensive testing of the primary candidates, and analysis of the sensitivity of the science mission to delays in the launch, the cooler with the damaged counterbalance was selected as the FM2 cooler. The damaged counterbalance was removed, and a replacement counterbalance was cut from the body of a cooler with damaged feedthroughs. The replacement counterbalance was carefully welded to the cooler body, and put into qualification. First, the new assembly underwent a thorough bakeout, followed by thermal vacuum, vibration qualification, and a cold-start test at –30°C. When the cooler was operated horizontally, it’s coldfinger temperature dropped by as much as 5K, relative to its operating point with the coldfinger in the vertical orientation. Its performance in horizontal orientations was never any worse than vertical, in which most of the cooler’s 10,500 hours of operating time had been accumulated. If there was touch contact inside the cooler, it seemed likely that it had always been present. The cooler was finally delivered to UCB, and installed in the FM spectrometer in September 2000.
OPERATION OF FM2 The spectrometer was quickly put into vibration testing, prior to thermal characterization. When the cryocooler was eventually cooled down, it was found to give adequate thermal performance, holding the detector temperature at 78 K for 67 W of input power. This was slightly higher than the FM1 cryocooler, but there was no way to determine if the cooler performance was a little low, or if items such as the spectrometer MLI had degraded during service and handling of the system.
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Malfunction of electronics box During spacecraft-level testing, a crash of the main computer cut off the clock that drove the output waveform of the cryocooler’s power amplifier. This should have put the full 28 VDC bus voltage across its terminals, but the cooler was inadvertently spared, as a small instability in the amplifier drove it into an un-clocked oscillation at a frequency of about 3 Hz. The cooler was not damaged, and the electronics were modified to eliminate the dependence on the spacecraft clock.
Final Instrument Warmup The spectrometer was kept cold almost continuously from the time it came out of vibration testing to the time it was mated to the launch vehicle, in an effort to keep the detectors cold enough to avoid incidental cosmic ray damage. During most of this time, liquid nitrogen was used to maintain the temperature, allowing the cryocooler to remain dormant. Due to a series of mishaps with the Pegasus family of launch vehicles, the launch date slipped from 2001 into February of 2002. The cooler was used for the last time in January 2002, to control the internal temperature distribution of the spectrometer during the final pre-launch warmup to room temperature. The cooler was working nominally, and the spacecraft was shipped to the launch site.
LAUNCH AND EARLY OPERATIONS The spacecraft was finally launched February 5, 2002, and power was applied to the cryocooler about eight hours after launch. The detectors reached an operating temperature of 65K over the next six days, with the cooler operating at about 90 W of input power (Figure 3). The power was subsequently trimmed back to about 50 W, to maintain a coldplate temperature of 72-75K. The cooler performance compares well with data from spacecraft thermal/vacuum testing, indicating that the cooler performance was not measurably affected by the actual launch environment. The system has now operated for about twenty weeks on-orbit as of June 2002, with no signs of change in its performance.
CONCLUSIONS The M77 cooler seems to have been successfully put into service on-orbit on the RHESSI spacecraft. Although the M77 was not originally intended for spaceflight use, it was possible to qualify individual units for flight, and show with some confidence that they have good potential for
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lasting through a two-year mission. Due to the low number of coolers produced, and the variability from unit to unit, it is still difficult to say what the overall reliability might be.
REFERENCES 1.
Boyle, R., Banks, S., Cleveland, P. and Turin, P., “Design and Performance of the HESSI Cryostat,” Cryogenics 39, no. 12 (1999), pp. 969-973.
2.
Report on High Energy Solar Spectroscopic Imager (HESSI) Test Mishap, ftp://apollo.ssl.berkeley.edu/ pub/hessi/JPL_Mishap/HESSI_MIB_vol1.pdf
Operation of a Sunpower M87 Cryocooler in a Magnetic Field S.R. Breon, K.A. Shirey, I.S. Banks, B.A. Warner, R.E Boyle, and S. Mustafi NASA Goddard Space Flight Center Greenbelt, MD, USA 20771
ABSTRACT The Alpha Magnetic Spectrometer–02 (AMS–02) is an experiment that will be flown as an attached payload on the International Space Station to detect dark matter and antimatter. It uses large superconducting magnets cooled with superfluid helium to bend the path of cosmic particles through a series of detectors, which then measure the mass, speed, charge, and direction of the particles. Four Sunpower M87N Stirling-cycle cryocoolers are used to extend the mission life by cooling the outer vapor-cooled shield of the dewar. The main magnet coils are separated by a distance of approximately 1 m and the coolers are located approximately 1.5 m from the center line of the magnet, where the field is nearly 1000 gauss at the cryocooler cold tip. Interactions between the applied magnetic field and the linear motor may result in additional forces and torques on the compressor piston. Motion of the compressor and displacer pistons through the magnetic field spatial gradients will generate eddy currents. Additional eddy currents are created during magnet charge, discharge, and quench by the time-varying magnetic field. The results of tests to demonstrate the performance of the cryocoolers in an external magnetic field, with and without magnetic shielding, are presented.
INTRODUCTION The Sunpower M87N cryocoolers that are used on AMS–021 are exposed to substantial steadystate magnetic fields that could interfere with the thermal performance of the cooler or degrade the cryocooler’s lifetime. Since the purpose of the cryocoolers is to reduce the parasitic heat load and extend the lifetime of the AMS–02 superfluid helium dewar, either of these effects would shorten the mission duration. It is conceivable that the coolers would not operate at all, or could be damaged by eddy current forces during a magnet quench. AMS–02 is the first space-flight mission that will use cryocoolers to help cool a system that includes a large superconducting magnet. Studies of the Astromag project2 considered the use of cryocoolers for a free-flyer version, but the technology development effort to examine this question was never funded. The XRS instrument3, which was lost due to a launch failure, is being rebuilt for launch on Astro-E2. XRS does have a superconducting magnet and the XRS2 version will use a cryocooler to extend the life of the superfluid helium dewar. Compared to AMS–02, however, the XRS2 magnet is much more modest in size and field, and the magnet is heavily shielded. The field at the XRS2 cryocooler will be < 10 gauss. Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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For ground applications, Gifford-McMahon (GM) cryocoolers have been used to cool superconducting magnets directly in magnetic resonance and energy storage applications4,5,6. Reports on these applications have focused more on how effectively the cryocooler cools the magnet and whether or not the vibrations from the cryocooler disturb the magnet or system. A small performance loss of the cryocooler would likely go unnoticed. Furthermore, in ground-based applications with a split compressor/expander configuration, the compressor can be placed further away from the magnet in a location where the magnetic field is low.
EXTERNAL MAGNETIC FIELD AMS–02 uses four cryocoolers: two are mounted on the upper support ring of the vacuum case and two are mounted on the lower support ring of the vacuum case, as shown in Fig. 1. The cryocoolers point radially inward so that the cold tip of each cryocooler is close to the outer vaporcooled shield. Each cryocooler experiences different field strengths and field gradients depending on its location relative to the magnet. Due to magnet symmetry the fields and gradients at the cryocoolers on the upper ring are a mirror image of each other. The same is true on the lower ring. The magnet is being designed and built by Space Cryomagnetics, Ltd. (SCL). Using its model of the AMS–02 magnet, SCL generated values of the magnetic field in all three directions at 50 grid points specified for each cryocooler. The values of the fields, given in AMS–02 coordinates, were translated to cryocooler coordinates. The cryocooler coordinate system and the grid points are shown in Fig. 2. The values of the field at the cold tip, motor, and balancer for one of the cryocoolers on the upper ring are given in Table 1. The cryocoolers on the upper ring are exposed to higher fields and field gradients than the cryocoolers on the lower ring. Table 2 shows the maximum field gradients along the cryocooler axis and normal to the axis.
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During most of the time on-orbit, the external magnetic field will be constant. The magnet is first charged once AMS–02 is installed on the International Space Station and a persistent switch is closed. It is not expected that the magnet will be discharged and recharged frequently, if ever, on orbit. In the event of a quench, the magnet’s protection circuit will discharge the magnet rapidly over a period of <10 seconds as shown in Fig. 3. The maximum magnetic field rate of change during a quench is 297 gauss/second.
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MAGNETIC FIELD/CRYOCOOLER INTERACTIONS The focus of our investigation is the effect that the applied magnetic field has on the cryocoolers, which have a fixed coil, moving magnet ac linear motor. Although the cryocoolers will perturb the external field locally, they are far removed from the central bore of the magnet system and the disturbance at the detectors is expected to be negligible. The magnet developer may include the cryocoolers in an update of the AMS–02 magnetic field model. For most of the time on-orbit the cryocoolers will be operating in a non-uniform steady-state magnetic field. Regardless of whether the cryocoolers are turned on or off, the external field will interact with any magnetic materials in the cryocooler. Outside of the motor, the only significant magnetic material is a steel ring that is part of the counterbalance. For reasons that are explained later in this paper, we are considering replacing this ring with a non-magnetic stainless steel ring. The motor itself is designed to have no net dipole moment, so any induced magnetic forces and torques on the cryocooler will be due to secondary effects including the motor’s quadrupole moment, non-ideal characteristics of the motor such as mechanical tolerances, and variations in the magnetization of the material. Structurally, these forces and torques are easily counteracted by the support bracket. If the field were strong enough, it could alter the magnetization of the cryocooler motor.
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When the cryocoolers are operating, forces and torques on the moving magnet could cause misalignment of the compressor piston and cause vibration and rubbing, which could shorten the lifetime of the cryocooler. If the external magnetic field saturates the motor laminates, it will reduce the motor efficiency and degrade the thermal performance. In addition, the external magnetic field will generate eddy currents in all moving metallic pieces such as the displacer and the compressor pistons. These eddy currents will generate heat that will reduce the cryocooler thermal performance. Eddy current forces will tend to inhibit the motion of the pistons, thus further degrading thermal performance. When the cryocoolers are exposed to a transient field; i.e., when the magnet is being charged or discharged or if it quenches, eddy currents will be generated in all metal components in the cryocooler. Since these events will be rare, any degradation in thermal performance during a transient will not be of concern to the project. The main concern is that the transients not cause damage to the cryocoolers.
TEST RESULTS We tested two engineering model cryocoolers in a magnetic field at a cyclotron facility located at the Massachusetts Institute of Technology (MIT), as shown in Fig. 4. At the center of the cyclotron, the field is vertical and very uniform. Near the edges there is a substantial gradient in the radial direction parallel to the cyclotron pole pieces. After mapping out the magnetic field we chose to place the cryocooler in a horizontal position in the fringing field near the edge of the cyclotron. The magnetic field at the motor is considered to be the most important, so the vertical field at the center of the motor was chosen to be the standard for the tests: the cold tip was in a larger field and the balancer was in a smaller field during the test compared to the flight configuration. Although the cyclotron does generate a small radial field (axial to the cryocooler), above the midplane the field is radially inward and below the midplane it is radially outward, so it does not mimic the AMS–02 field. As can be seen from Table 1, the field normal to the cryocooler axis at the motor is approximately 600 gauss. A fiducial was marked on the cryocooler fixture near the motor and used as the reference point for magnetic field measurements during the test. The maximum expected field at the fiducial was 610 gauss. Both cryocoolers were operated at full field with no degradation in performance. The cryocoolers continued to operate at fields up to 1,200 gauss, or twice the maximum expected field, with only a small degradation in performance. The cryocoolers were cycled on and off when exposed to a 1,200 gauss field without incident.
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The cryocoolers were tested with two cylindrical iron shields (see Fig. 5) to determine if shielding would reduce or eliminate the performance degradation. The shields extended from the cold tip to the aft end of the compressor. One shield had an outer diameter of 135 mm and a thickness of 4.6 mm. The other shield was 152 mm in diameter and 12.7 mm thick. Both shields were 279 mm long. Although the shields did screen out some of the external field, they had no effect on the performance degradation. A systematic set of load curves was taken for one of the unshielded cryocoolers at three input powers and three heat loads as a function of increasing magnetic field. The results, shown in Fig. 6, are corrected for drift in the heat sink temperature around a nominal 30°C. The minimum and maximum input powers at which the cryocoolers will be operated are 60 W and 150 W, respectively. The nominal input power is limited to 100 W by power restrictions on the AMS–02 payload. The performance degradation is most pronounced at low input power and high heat load: at 60 W input power and 6 W heat load, there is a 4% loss in cooling power at ~900 gauss, or 1.5 times the maximum expected field. At 100 W input power and 6 W heat load, the cooling loss at ~900 gauss is less than 1%. One explanation for the performance degradation which was considered was that the thermometers used to measure cold tip temperature were affected by the magnetic field. This would not have been expected because the Cernox thermometers mounted on the cold tip are supposed to be insensitive to magnetic fields. However, to verify that the problem was not associated with the thermometry, we rotated the cryocoolers to place the cold tip in a lower magnetic field. Surprisingly, the performance degradation increased. We then realized that the balancer was now in a higher field, and the interaction between the magnetic steel in the balancer and the external magnetic field was hurting the cryocooler performance. To test this hypothesis we mounted the cooler on its flexible mount and took load curves with the balancer attached and with the balancer removed (see Fig. 7.) We also rigidly mounted the cryocooler with the balancer removed. As expected, the cryocooler on a flexible mount with the balancer attached shows a larger rise in temperature as well as sensitivity at lower fields than the other two cases. When the cryocooler is mounted on a flexible mount and the balancer is removed, the greater unbalance results in a warmer cold-tip temperature. The increase in temperature is attributed to some of the energy going into the cryocooler now being dissipated in viscous damping at the mount rather than in useful work on the gas. This explains the higher cold tip temperature when the cryocooler is flexibly mounted both with no balancer as well as with the balancer interacting with the magnetic field. When the cryocooler is rigidly mounted it does not lose energy in the mount. Interestingly, there is still a small performance degradation even when the balancer is removed from the cryocooler, which might indicate a loss of motor efficiency. It is also possible that the effect is due to rubbing of the piston, although the repeatability of the phenomenon and absence of hysteresis suggests that this is not likely.
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Additional tests will be conducted at the MIT cyclotron facility to understand the performance degradation at high fields. We will also measure the acceleration of the cryocooler to determine if the vibration forces are increased by the interaction with the magnetic field. Alternatives to the magnetic steel ring in the balancer will be tested, including a nonmagnetic stainless steel ring and a laminated nonmagnetic stainless steel ring. In addition, forces on a sample cryocooler motor will be measured.
ANALYSES The complex geometries encountered in this problem do not yield to simple calculations. The need to perform finite element analyses of the interactions between the applied magnetic field and the cryocooler is being evaluated. We have done a very rough estimate trying to bound the eddy current heating in the cold finger and the compressor piston as they move through the magnetic field gradient. We have assumed that the maximum resultant field in both of the locations is coaligned with the cylindrical axis and that the magnetic field gradient equals the maximum value at both locations. The eddy current heating, P, in a cylinder of radius and thickness L moving sinusoidally in a non-uniform magnetic field with gradient dB/dx is given by:
where
is the electrical resistivity and dB/dt is:
where is the peak-to-peak stroke of the piston and where f is the operating frequency of the cryocooler. The maximum eddy current heating is equal to:
Similarly for a ring of thickness L with outer radius
and inner radius
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The maximum heating in the cold finger is approximately 1.5 mW and in the compressor and warm end of the displacer it is approximately 15 mW. This amount of heating is negligible for the cryocooler performance since the cooling power at the cold tip is on the order of 4 W and the input power to the compressor is
CONCLUSION The Sunpower M87N cryocooler is capable of operating in a magnetic field as large as 1200 gauss perpendicular to the axis of the cryocooler. The cryocooler can start and stop in a 1200 gauss field. A 1 – 4% thermal performance degradation has been seen at fields between 900 and 1200 gauss. Part of that degradation is attributed to damping of the passive balancer, resulting in additional energy loss in the flexible mounts. We are exploring the possibility of eliminating this effect, but would still expect to see a very small degradation at high fields due to magnetic field effects on the motor. These effects are not seen at the maximum expected field on-orbit. The true test of operating the cryocooler in the AMS–02 magnetic field will only occur when the flight cryocoolers are integrated and operated with the flight magnet. Any long-term degradation will not be known for certain until AMS–02 is on-orbit. The test program we have conducted has given us confidence that the short-term performance degradation will be negligible even up to 1.5 to 2 times the maximum expected field. Additional tests and analyses are planned to examine the question of long-term degradation that could lead to the failure of a cryocooler before the end of the mission.
ACKNOWLEDGMENTS The authors wish to acknowledge the support of Dr. Henning Leidecker of NASA Goddard Space Flight Center in formulating approaches to the problem. We also wish to thank Dr. Ulrich Becker of MIT for access to the cyclotron facility and Messrs. Stephen Harrison and Steve Milward of Space Cryomagnetics, Ltd. for calculations of the magnetic field parameters.
REFERENCES 1.
Shirey, K.A., Banks, I.S., Breon, S.R., and Boyle, R.F., “Space Flight Qualification Program for the AMS–02 Commercial Cryocoolers,” to be presented at the 12th International Cryocooler Conference, Cambridge, MA, 2002.
2.
Green, M.A. and Castles S., “Design Concepts for the Astromag Cryogenic System,” Advances in Cryogenic Engineering, vol. 33, Plenum Press, New York (1988) pp. 631-637.
3.
Kelley, R.L. et al., “The Microcalorimeter Spectrometer on the Astro-E X-ray Observatory,” Nucl Instrum Meth A, vol. 444, no. 1-2 (2000), pp. 170-174.
4.
Hoenig, M.O., “Design Concepts for a Mechanically Refrigerated 13 K Superconducting Magnet System,” IEEE Transactions On Magnetics, vol. MAG-19, (1983).
5.
Urata, M., Maeda, H., Aoki, N., Uchiyama, G., “Compact 17 T Epoxy-impregnated Magnet without Bore Tube,” Cryogenics, vol. 31, no. 7, (1991) pp. 570-574.
6.
Laskaris, E.T., Ackermann, R., Dorri, B., Gross, D. Herd, K., and Minas, C., “A Cryogen-Free Open Superconducting Magnet for Interventional MRI Applications,” IEEE Transactions on Applied Superconductivity, vol. 5, no. 2 (1995) pp. 163-168.
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Active Vibration Cancellation in Astrium Stirling Cycle and Pulse Tube Coolers S. Akhtar and R. Hunt Astrium Ltd Stevenage, SG1 2AS England
ABSTRACT The expanding use of cooled detectors in orbital science and defence applications has produced a natural evolution towards higher performance in the detector cooling system. The control of forces emitted by the cooling system becomes ever more significant with the demands for greater instrument pointing accuracy and, for detectors operating at sub-Kelvin temperatures, control of parasitic heat loads due to vibration of the detector cryostat. This trend in requirements towards lower vibration is clearly evident in the recent supply of flight cooler systems by Astrium to INTEGRAL, HELIOS II and ENVISAT (AATSR and MIPAS). In the instruments for these programs the Astrium coolers are operated in a back-to-back configuration for minimum exported force. The subject of this paper is the reduction of exported forces to a minimum by the active control of the cooler drive power. This paper will discuss some active steps which can be taken to reduce the forces exported from a cooler system using active control in analogue/digital electronics. It will outline a role in which embedded Digital Signal Processors (DSPs) can further improve performance.
INTRODUCTION The dominant source of emitted force is the cooler compressors. The compressors are essentially linear motors with reaction masses suspended on diaphragm springs. Due to inherent system non-linearities, e.g. in the motor drive electronics, the linear motor, the piston support springs, and the gas dynamics, the movement of the compressor piston will create a spectrum of emitted forces at the compressor fundamental drive frequency and its harmonics. A sample spectrum from a singleended compressor is shown in Figure 1. For this example the force emitted at the cooler drive frequency of 44Hz is approximately 55N. The scale has been enlarged to show the harmonic structure caused by the non-linearities of the mechanical system. The objective is thus to minimize this force spectrum. There are various strategies which can be employed to minimize these emitted forces. For example: Minimize the moving mass. Use of a balanced compressor configuration (Astrium Miniature Pulse Tube Cooler, Gibson et al. 1 ) Internal compressor geometry improvements (Vibration Reduction in Balanced Linear Compressors, Dadd et al. 2 ) Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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Measure the exported force and modify the compressor drive waveform accordingly to minimize the force. The first three strategies can be identified as removing the problem at the source. By careful design and construction the emitted forces can be minimized. However, there will always be some non-linearities present especially in the gas spring dynamics which will generate forces at the harmonic frequencies of the drive waveform. It is the roll of the vibration control strategy to minimize this force spectrum and thus minimize the vibration of the instrument of which the cooler system forms a part.
BASIC CONTROL The basic approach to control these forces is to place compressors in an opposing configuration with the two pistons sharing a common axis for their motion or use a balanced dual compressor, as in our Mini Pulse Tube Cooler, and drive the compressor linear motors in phase. In this manner, the forces produced by the compressors are in opposition and can be balanced against each other. The balancing is determined during ground testing only and the adjustments required are then left fixed for flight. The INTEGRAL program utilizes this form of force control, Figure 2. The emitted forces were measured by a Kistler™ force transducer mounted on the cooler mounting frame. This signal was conditioned and then logged on a PC-based controller. The force spectrum was analyzed by applying a Fourier Transform (FFT) to the logged data. An averaging function was used to estimate the total emitted force. The function was formed as a sum of the emitted forces at the compressor drive frequency, F0 and its first four harmonics, i.e.
The control methodology was simply to declare one of the compressors to be the master (independent) and the phase and amplitude of the drive to the second compressor was adjusted to minimize the forces measured as The INTEGRAL cooler system consists of two pairs of
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coolers (SPICO 1 and SPICO 2) mounted on a single frame (see Figure 2). Each compressor and displacer of a pair is mechanically placed such that their motion oppose each other when driven in phase. Sample data taken during INTEGRAL cooler system tests at Astrium using thus methodology are shown in Table 1. When only one cooler subsystem was operated, e.g. SPICO 1, the other cooler was unpowered. Each cooler was set for a compressor input power of exactly 44 watts or a fixed stroke of 9mm. It is interesting to note that the equal power setting, rather than equal compressor strokes, gave better low vibration figures. The SPICO 2 results showed a similar trend. In the final test, both cooler subsystems were operated simultaneously with the same power, 50W, applied to each unit. In this four-cooler mode of operation, one cooler of each pair was made the master reference and the phase of the second cooler was adjusted to give minimum force in the x-axis. All the above low vibration tests were performed steady state, with the cooler cold-tip at 90K and a temperature stability better than 0.1K over any 10 minute period. This basic control approach shows that simply balancing the drive power to a mechanically symmetric arrangement can reduce the exported forces to 1N to 2N levels. The main benefits of this approach are simplicity and lower cost than an active solution. The main limitation is that it is difficult to update the drive balance parameters once the cooler system is in orbit.
ACTIVE CONTROL With active control, the emitted forces achieved by the basic control method can be lowered, and the dependency of the basic method on fixed parameters determined on the ground can be
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removed. The basis of the active control methodology is to determine the force spectrum as before, but this time not just balance the compressor drive powers for minimum emitted force, but to also examine each harmonic in turn and modify the compressor drive waveform to minimize each harmonic individually. This methodology has been shown to provide an additional l0dB of attenuation to Typically, this reduces the emitted force to 0.1N to 0.2N rms, with good attenuation of the fundamental frequency and its first six harmonics. Determination of the force spectrum can be realized in two ways, either by predominantly analog signal processing, or by predominantly digital signal processing. These two alternate methods are described below.
Analog Signal Processing The heart of the design is an analog phase locked signal detection system which can be set to determine the magnitude of any harmonic in the emitted force spectrum. The value of the emitted force at each frequency in turn is then monitored by an FPGA. The algorithm seeks to minimize the measured force by adjusting components at that particular frequency in the compressor drive signal. A functional block diagram of the design is shown in Figure 3. The emitted force signal from the Kistler™ transducer mounted on the cooler system frame is passed through a signal conditioning unit. This conditioned signal along with the reference cooler operating frequency (e.g. 44 Hz for fundamental (f0), 88 Hz for first harmonic (f1), to f(n) for the highest harmonic, are passed to a four quadrant linear multiplier and low pass filter. This circuit functions as a tunable narrow band pass filter which can be set to the fundamental and each harmonic of the force spectrum. The force level at the selected frequency is then digitized and passed to the FPGA. The FPGA then determines the level of modification to be applied to the compressor drive signal and updates the drive signal. By sequencing through each harmonic in turn the overall emitted force from the cooler compressor can be minimized. Using this method we have demonstrated an improvement of l0dB in emissions compared to the simple balanced approach.
DSP-BASED ACTIVE CONTROL The alternative active control solution is based on using a Digital Signal Processor (DSP) as the heart of the controller, Figure 4. Here the frequency components of the emitted force spectrum are determined via a Fourier Transform. Once determined the same algorithm is followed to minimize the emitted force. The use of a DSP approach where the algorithm is executed in software gives a strong benefit in that it permits the software to simply tailored to the individual application.
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SUMMARY Astrium through their development of cryogenic cooler systems, including cooling engine design, electronic design and systems integration, have developed an effective system-level solution to the control of forces from complex multi-cooler systems. These forces can give rise to micro-vibration disturbance torques at the instrument level. The primary approach to vibration control is to minimize the forces emitted by the cooler system. The drive strategy which this paper describes, that of modifying the drive waveform to one of a pair of compressors, can reduce the emitted forces down to 0.1N levels. The methods described in this paper are applicable to all cooling systems that utilize compressors. The challenge is now to refine the control algorithm further to achieve emitted forces down to 0.01N levels. Work in this area is continuing.
REFERENCES 1. Gibson A.S.et al., “Design and Characterization of a Miniature Pulse Tube Cooler,” Cryocoolers 12, Kluwer Academic / Plenum Publishers, New York (2003).
2. Dadd, M.W. et al., “Vibration Reduction in Balanced Linear Compressors,” Cryocooler 11, Kluwer Academic / Plenum Publishers, New York (2001), pp. 175-182.
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On-Orbit Cooling Performance of a Miniature Pulse Tube Flight Cryocooler D. R. Ladner1, R. Radebaugh2, P. E. Bradley2, M. Lewis2, P. Kittel3, and J. H. Xiao4 1
Lockheed Martin Astronautics Operations Denver, Colorado 80201 2 National Institute of Standards and Technology Boulder, Colorado 80303 3 NASA Ames Research Center Moffett Field, CA 94035 4 Ethicon, Inc., Somerville, New Jersey 08876
ABSTRACT The Miniature Pulse Tube Flight Cryocooler (MPTFC) developed jointly by the National Institute of Standards and Technology (NIST) and Lockheed Martin Astronautics Operations is a twostage pulse tube (PT) cryocooler designed for flight demonstrations aboard the Space Shuttle. It was first flown on STS-90 in April 1998 and again in December 2001 on STS-108. The design, testing, and data from the first flight have been reported previously. In this paper we report on the on-orbit cooling performance of the MPTFC with and without a heat load at the second stage and compare it to post-flight cooling data. The on-orbit warm-up data with a 2nd stage heat load of 45 mW after the compressor is switched off are also compared to post-flight laboratory data. While the on-orbit and laboratory cooldown and cooling performance data are found to be nearly identical, the 90 minute orbital temperature variation on the heat rejection surface has a significant effect on the cooler component and stage temperatures. Also, while the on-orbit and laboratory steady state temperature difference data with the heat load are also nearly identical after complete warmup, the on-orbit dynamic conductance data during the warm-up exhibit anomalous behavior that is attributed to the outgassing of cryopumped contaminants.
INTRODUCTION The Miniature Pulse Tube Flight Cryocooler (MPTFC) flew as a NASA shuttle payload (GAS785) in December 2001 aboard STS-108 (Shuttle Transportation System – 108) as a technology demonstration experiment. The experiment provided ~19 hours of on-orbit cooling data and ~ 75 hrs. of additional zero-g thermal data. This project was a collaboration between Lockheed Martin Astronautics Operations and the National Institute of Standards and Technology (NIST). The STS108 re-flight was sponsored by NASA / ARC. The primary objectives of this experiment were to demonstrate pulse tube (PT) cryocooler performance and dynamic “off-state” thermal conductance in a micro-gravity environment, to obtain vibration data in zero-g, to verify launch survivability, Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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and to measure heat flows in a GAS canister with an uninsulated Experiment Mounting Plate (EMP). In this paper we report on the cooling performance and off-state conduction measurements and compare them to ground test data. The steady state conductance data are also compared to data obtained from the previous MPTFC flight (GAS-197), STS-90 in April 1998 [1]. A low budget required the extensive use of commercially available components, including a tactical compressor and drive electronics, an electromagnetic latching valve, a commercial data acquisition system, and numerous commercial electronics components. Attention to flight safety issues directly impacted the MPTFC design in terms of its flight operating pressure (2.5 MPa); its miniature sizing was dictated by limited lifetime battery-powered operation in a cold environment. As previously reported [2], the experiment design—cryocooler and support hardware — also had to address various flight hazard issues, such as mechanical and electrical integrity, EMI, redundant fusing, diode isolation of batteries and electronic subsystems, mitigation for high subsystem temperatures, launch and landing vibration, etc. The experiment design had to accommodate STS bay temperatures from -50 to +40 deg C. In addition, the experiment commands had to conform to limited opportunities for STS crew operations in the mission timeline.
DESIGN The cryocooler cold head design is a two-stage U-tube geometry orifice pulse tube (PT) based on the double inlet concept first introduced by Zhu, Wu, and Chen [3]. The cold head is schematically represented in Figure 1. This two-stage approach was selected to meet the design goal of reaching 80K and the miniaturization requirement in which the compressor and cold head are separate components. The compressor and cold head were separated to reduce vibration at the cold head and to balance the thermal operating loads at the compressor. This approach, commonly referred to as multi-inlet when two or more stages are present, reduces the regenerator loss by using a secondary orifice which diverts a small percentage of the gaseous helium working fluid directly from the compressor to the warm end of the pulse tube. This small flow produces a favorable phase shift between the pressure and mass flow waveforms and also reduces the regenerator loss. For optimal performance this configuration relies on optimized and stable flow division (provided by the three orifice impedances), minimum void volume, maximum pressure ratio, and minimization of any DC flows or turbulence. Analytical and numerical models such as REGEN 3.1 developed by NIST [4,5,6] and a thermoacoustic model developed by Xiao [7,8,9,10] were employed to design both the prototype and flight cold heads. The partially assembled MPTFC flight cold head is shown in Figure 2. It is driven by a small 0.75 cc swept volume MX-8000 compressor. Because automatic stroke control was not achievable on-orbit, the compressor was operated below full stroke to prevent excessive vibration when ini-
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tially switched on. Instrumentation on the cold head provided redundant temperatures, a fixed cold end heat load [11], system pressure, and acceleration data at the warm end. The flight cooler (MPTFC) cold head is shown in Figure 2. The figure provides an exploded view of the two stages although many components are not shown. For example, the primary and secondary orifices are integrated into the aftercooler. Also not shown are two thin plates that help establish impedances to form the secondary and intermediate orifices. Several other flight-related elements including temperature sensors and a film heater on the cold end (second stage) are also omitted. The MPTFC design features include 1) a very compact physical arrangement; 2) conduction cooling of the aftercooler and the compressor via aluminum mechanical supports that connect to the EMP; 3) a thin wall polished aluminum radiation shield that attaches to the mid-stage headers; and 4) nylon displacement stops located at both the mid-stage and cold end. Figure 3 shows the G-785 Payload (experiment and flight support hardware) minus the GAS canister enclosure during integration at NASA/KSC, while Figure 4 shows the G-785 integrated into the Space Shuttle Endeavour after delivering the primary payload (cargo module) to the International Space Station (ISS). The upper third of the G-785 consists of the MPTFC experiment itself, which is contained in the vacuum housing (small inverted canister at the top); an electromagnetic latching vacuum valve (EMV) and flex hoses to connect the experiment housing to space vacuum; and other electrical and battery venting hardware. The lower two thirds (the “battery box”) contains two redundant battery modules and the flight electronics module—the data acquisition system (DAS), computer, signal conditioners, and compressor drive electronics. The payload integrates with the 5 cubic foot NASA GAS canister that is capped at both ends and evacuated during integration. The completed GAS canister fully encloses the experiment; all electronic connections are made via feedthroughs at the canister bottom plate which holds the NASA electronics. The bottom plate is the electronic interface between G-785, the GAS relay system, and the shuttle GAS computer operated by astronauts. The flight support hardware and the cryocooler experiment/adapter plate are bolted into the bottom of the upper plate or EMP. Therefore, most of the heat generated by the G-785 is conducted directly to the EMP, but a small fraction is first radiated into the canister walls. The EMP is uninsulated and its external surface is coated with silverized teflon tape having excellent heat rejection properties (solar absorptivity = 0.12 and IR emissivity = 0.76). In Figure 4, the G-785 is the canister on the top right, horizontally mounted onto the MPESS carrier in Bay 13. Given this canister orientation with the EMP facing the shuttle port side, the radiation view factor comprises both the shuttle bay and the space and/or earth thermal environments. Most GAS payloads opt for an insulated EMP, but for MPTFC the uninsulated EMP pro-
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vides a lower rejection temperature. The disadvantage, lacking a sunshade/shielding, is that the EMP has full exposure to the orbital temperature variations and shuttle attitude changes. Other features of the G-785 payload are: 1) the GAS canister is evacuated (vs. 1 atm ) to simulate a flight instrument environment; 2) the 16 V secondary battery system consists of eight 4 V batteries in a 2-string redundant arrangement to actuate the EMV and to power the DAS, computer, and flight electronics; 3) the 24 V primary battery system consists of thirty-six 4V batteries in a 6-string redundant arrangement to power the MPTFC and its compressor drive electronics at a nominal 1.6 amps; 4) all batteries are integrated into two identical modules clocked 90 degrees apart, and the strings have different cell orientations for mitigating launch vibration. In addition to experimental data, canister and hardware temperatures are also measured for comparison to numerical models [12, 13] for powered payloads in uninsulated EMP canisters.
COOLDOWN RATE AND COOLING PERFORMANCE Figure 5 shows the on-orbit cooling data (solid symbols) and post-flight simulation test data (open symbols) plotted for comparison. Note that the EMP rejection temperature for the simulation data is relatively constant because the EMP temperature was actively controlled using an FTS RC100 recirculating cooler. In contrast, the on-orbit EMP temperature floats with the Space Shuttle environment and experiment heat load conducted into it. Therefore, after several hours of cooling, the EMP on-orbit and simulation temperatures have diverged by ~8 K. Corresponding but smaller temperature differences appear at each stage of the cold head. The lowest on-orbit temperatures were 96.5 K for stage 2 and 181.9 K for stage 1. The corresponding simulation stage temperatures were 93.1 K and 177.3 K, respectively. Also note that the cooldown rates are identical to within the experimental error associated with the rejection temperature difference. The slight temperature rise in the stage 2 simulation data at ~ 640 minutes elapsed time (ET), referenced to initial electronics “on”, corresponds to a high state of discharge on the primary battery [14]. Figure 6 is the continuation of the on-orbit cooling, which follows several hours of steady state cooling (not shown). It includes actuation of the 50 mW heater [11] on the stage 2 cold end at ET = 1310 min. During the steady state cooling the stage temperatures drifted up slightly in response to the upward drift of the EMP rejection temperature. Part of the EMP upward drift is due to the long time constants associated with thermal equilibration; however, part of it is also due to a new heat input, viz., the exothermic reaction of the batteries as they became increasingly discharged. Stage 1 has drifted from 181.9 K to ~186 K and stage 2 from 96.5 K to ~100 K when the heater on
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stage 2 is commanded “on”. The stage 2 temperature rises to 110 K in ~ one hour, while stage 1 rises to 188 K; these are equivalent to the results from previous laboratory cold functional tests [2]. Prior to commanding the experiment compressor power to “off” at ET = 1644 min. (right edge of Figure 6), the low primary battery voltage, coupled with high current demand by the compressor drive, temporarily resulted in higher input power to the compressor and increased stroke, causing the dip in stage temperatures. This was verified by noting a sharp increase in the temperature of the driver itself, followed by a temperature decrease as the battery discharge current also began to fall. This corresponds to the gradual rise in stage temperatures until stage 2 is 133 K and stage 1 is 197 K at compressor “off”. It is noteworthy for the discussion of cold head contamination below that this warm-up, duplicated in the laboratory in the post-flight simulation, is not thought to be associated with cryopumping of by-products from the SIMO dump or RCS thruster burns, both of which had occurred slightly earlier in the mission timeline. ON-ORBIT WARM-UP RATE ANOMALY One unanticipated result of the flight demonstration was the initial accelerated warm-up rate of the cold head following the compressor “off” command. Figure 7 provides a comparison between the on-orbit and post-flight simulation warm-up data with the 45 mW heat load [11] on the stage 2 cold end. (The upper set of curves are the on-orbit data in the figure.) To simulate the onorbit condition, the primary battery voltage was deliberately allowed to fall to the approximate level that occurred on-orbit, which resulted in the gradual warm-up of the stages with the compressor still operating. When stage 2 reached ~ 133 K, the on-orbit temperature at compressor “off” in Figure 6, the compressor was switched off to duplicate that initial condition. However, stage 1 was at a slightly lower temperature (~190 K) than on-orbit (197 K), consistent with the difference between the EMP lab and the space temperatures (~ 8 K). This slightly larger would accelerate the on-orbit warm-up rate in the direction observed, but it would not fully explain either the qualitative or the quantitative behavior in Figure 7. The on-orbit stage 1 temperature curve clearly exhibits an increasingly nonlinear upturn beginning at ~225 K at ET = 54 min. (first arrow) until it reaches ~ 253 K (second arrow), at which time (ET = 90 min.) it rapidly transitions to the expected behavior, The on-orbit stage 2 temperature curve also deviates from the exponential behavior; it is essentially linear over the same 36 min. period. Beyond this transition point (second arrow), a of ~ .6 K still exists between the two EMP curves, indicating that the larger alone could not have caused the accelerated warm-up. Note also that the simulation data, aside from the higher initial stage temperatures, are typical of all previous MPTFC lab tests [15] and exhibit exponential behavior over the entire warm-up period.
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The accelerated warm-up rates indicate an anomalous dynamic conductance not previously observed with the MPTFC. This is associated with a transient thermal condition (thermal short) that clearly affects stage 1 more directly than stage 2. In Figure 7 the stage 2 temperature simply tracks stage 1 within about the same for both space and lab data. If the thermal short were caused by direct mechanical contact between either stage and an ambient component (e.g., a displacement stop), then the cooling performance (Figures 5 and 6) would have been measurably degraded. However, the nominal cooldown in Figure 5 is indicative of an adequate vacuum condition in the housing and good thermal isolation. We explain the thermal short by hypothesizing that water vapor or carbon dioxide that may have been cryopumped onto one or both stages sometime during the ~19 hours of on-orbit cooling would have rapidly sublimated during the cold head warm-up. This process would produce two competing thermal effects on the cold head: 1) cooling due to evaporation from active pumping, and 2) simultaneous warming due to the gas thermal conduction between the ambient vacuum housing and the cold head for vapor pressures above These effects are in addition to the nominal MPTFC parasitic heat loads that were analyzed previously [15]. Following this hypothesis the next step is to determine whether the contaminating source was of an external (e.g., Shuttle bay) or internal (vacuum housing) origin. Figure 6 identifies several Shuttle operations that produce contaminants and which occurred during the Shuttle approach to the International Space Station (ISS). However, it is not probable that any significant cryopumping occurred as a result of either the SIMO dump (water mostly), which lasted ~ 2.5 hrs, or the Orbital Maneuvering System (OMS) burn (~1 min.), because data from previous Shuttle missions indicate little change in the bay pressure during these operations [16]. However, during RCS burns the bay pressure may rise to [17]. This pressure level would be consistent with cryopumping water vapor onto either MPTFC stage, since the corresponding dew point is ~ 187 K, above the stage 1 and stage 2 temperatures [18]. As discussed below, however, the RCS burns also turn out to be an unlikely contamination source, because the pumping line impedance limits the cryopumping rate at the cold head. We conclude that the most likely scenario is that residual water vapor trapped within the vacuum housing itself may have cryopumped onto the cold head stages. The reason that this could have happened, in spite of ~ 8 hours of on-orbit pumpout prior to cooldown, is that the pumping speed for water vapor at space vacuum pressure of torr is much slower than the pumping
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speed obtained using a laboratory pumping station equipped with a cold trap, because the base pumping pressure is reduced by an order of magnitude and backstreaming of water vapor is effectively eliminated. Also, long pumpout periods (~ 24 hours) are typically associated with our MPTFC laboratory tests. Unfortunately, the prelaunch pumpout activity at NASA/KSC was limited to about 4 hours (using a turbomolecular pump without a cold trap) before the EMV was latched closed. The tight integration schedule did not permit elevating the vacuum housing temperature during the pumpout. A manual valve was placed in series with the EMV to reduce contaminant leakage into the vacuum housing but was removed about 3 weeks before flight. Nonetheless, the EMV nominal leak rate is and this would allow only < 2E-6 g of water ingestion during the 80 days before launch. Although unlikely, we cannot rule out a higher leak rate due to a worn valve seal. Additionally, we hypothesize that the fixed cylindrical radiation shield, having an external surface area of that attaches to stage 1 comprises the primary cryopumping surface. The temperature range of stage 1 during the anomaly, 225 K < < 253 K, corresponds to saturated vapor pressures of~0.035 to 0.72 torr (4.7 to 9.6 Pa), sufficiently high pressures for gas conduction between the shield external surface and the cold head mechanical support (~1 mm gap). Although the colder stage 2 components might also be expected to cryopump some of the water vapor, stage 2 is largely baffled by this shield, which runs its entire length but is open at the bottom. Several layers of MLI between the shield and the stage 2 components might also have intercepted any vapor that diffused into the open end of the shield. An important question is whether cold head cryopumping would have caused observable cooling degradation resulting from the heat of condensation/fusion (-2830 J/g) and/or the degradation of the stage 1 shield emissivity, as recently reported for some space coolers [19]. Thus, for fixed compressor power, continuous or intermittent cryopumping of water vapor or from either the SIMO dump or from numerous RCS burns indicated in Figure 6, should measurably degrade the stage temperatures (260 mW estimated heat load), but this was not observed. However, if the cryopumping of residual trapped water vapor had occurred gradually over the entire 19 hours of onorbit cooling, it would have produced a much lower level of heating (~25 mW estimated) compared to other stage 1 parasitics [15] and would have had little effect on the stage 1 temperature. Note that during this time (Figure 6) the MPTFC was experiencing uneven cooling caused by the primary battery low voltage condition. Additionally, increased exothermic reaction of the batteries caused an increase in the EMP temperature. Therefore, a slight performance degradation might have been masked by this uneven cooling. As stated above, the warm-up of both stages for ET > 1550 min. (before compressor “off”) was duplicated in the laboratory simulation, and it is unambiguously associated with low compressor power / decreased swept volume, not with a cryopumping heat load or an increase in the radiation load. As an analytical check on our hypothesis we can estimate the mass of accumulated water ice based on the finite sublimation period (~ 36 min.) during warm-up and then compare the result to the calculated ice masses cryopumped from potential internal or external sources. It turns out that although the sublimation mass flow rate during warm-up is a very strong function of the ice temperature through its dependence on the saturation pressure [19], the limiting factor for the mass flow rate is actually the pumping speed of the ~ 100 cm long (effective) x 0.127 cm diameter pumpout flex hose and the 5/16 inch orifice of the EMV [20], At ~275 K the mass flow rate of the water vapor would be a relatively constant 3E-4 g/s for molecular flow. It should be noted that the EMV remains latched open even after the MPTFC compressor is switched off, so that the sublimating water vapor would eventually be pumped out to space, reestablishing a high vacuum condition in the housing. We also note that the rapid transition in the stage 1 curve at 253 K is consistent with the depletion of a sublimation source, at which time the pressure would fall rapidly below since the remaining vapor mass would be only To deplete the ice source so quickly (~36 min.) the accumulated mass would have to be small, and for the estimated flow rate of 3E-4 g/ s it would be ~0.65 grams. The cryopumping rate from an external source (e.g., RCS water contamination at ) through the 275 K pumpout line / EMV is only 3E-8 g/s and it would take 6000 hrs. to cryopump 0.65 g. In contrast, the cryopumping rate from an internal source is limited only by the cryopump (shield) surface area and the source vapor pressure at 275 K
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(theoretically, ~5 torr). Here, the saturation pressure for ~182 K ice would maintain the vacuum housing pressure at but the source vapor pressure at the 275 K housing surfaces would be slightly higher. Using the saturated pressure to estimate the pumping rate gives 2.2E-5 g/s and the required cryopumping time would be ~8 hours. Therefore, cryopumping would have occurred during the initial cooldown (Fig. 5) and first several hours of cooler operation. The estimated heat load would have been ~60 mW, causing an estimated increase in the stage 1 temperature of~3 K. A second important question is whether the radiation load on stage 1 would also have increased. If the ice accumulated on the shield external surface, as postulated, the average layer thickness for 0.65 g would have been ~240 microns, and the external surface nominal emissivity (~0.10) should have been adversely affected [21]. (Lacking an ice buildup inside the shield, the emissivity (~0.06) of the highly polished internal surface facing the stage 2 components would have been unaffected.) The predicted higher total heat load on stage 1 has not been rigorously calculated, but we estimate an increase of 0.3 W above the nominal parasitic load of ~0.5 W. Therefore the stage 1 temperature should have been very measurably affected throughout the experiment. Since this was not the case, further analysis of this mechanism is needed. Finally, we note that the pronounced upturn in the stage 1 temperature for Tl > 225 K in Figure 7 can be explained by 1) the transition from pressure dependent molecular conduction to viscous conduction as the mean free path of the vapor becomes less than the characteristic length of the gap (~1 mm), and 2) the reduced surface area for ice sublimation as the sublimation process progresses along the shield temperature gradient.
“OFF-STATE” THERMAL CONDUCTANCE The measured conductance of the MPTFC cold head in zero-g has been previously reported [ 1 ]. Both laboratory and space measurements of the steady state temperature differences across the stages with a fixed heat load of 45 mW at the stage 2 “cold end” are in excellent agreement, including both adverse and favorable orientations for convection in 1 g. As discussed above, the dynamic off-state conductance in zero-g during a transient period of the warm-up was found to be much higher than the conductance measured in the laboratory simulation. Because the cold end is positioned above the warmer stages in the simulation test setup (G-785 oriented as in Figure 3), convection is favored during warm-up until the stage temperature polarities reverse. For stage 2 this occurs at ET ~ 160 min (space) and 170 min (lab) and for stage 1 at ET ~ 180 min. (space) and ~ 250 min. (lab) in Figure 7. Since the zero-g space data indicate a higher stage 1 conductance (that must not be associated with convection) than the lab data prior to these reversals, any convection effects in 1 g are negligible compared to the mechanism causing the high conductance in space. We can also rule out a heater malfunction because of the qualitative difference between the space curves and those from previous measurements at higher heater powers [15]. Also, the monitored heater element voltage was nominal. Various lab warm-up data involving different initial and final steady state temperatures (rejection temperature plus ) all give about the same “system” time constant. Using the initial stage temperatures at the transition time, ET = 90 min., the time at which the thermal short apparently disappears, gives a time constant for the space data of only about half the lab simulation time constant [22]. However, the final steady state turn out to be about the same for both sets of data. The simulation and space curves are already converging at the right side of Figure 7, and by ET ~ 1200 min. (not shown) the two sets of data have reached steady state. At this time the EMP temperature is ~265 K for both sets of data. The zero-g EMP-to-stage 1 is ~ 8.2 K and the stage 1-to-stage 2 is~12.1 K. The post-flight simulation results were 7.2 K and 11.1 K, respectively. Previous measurements in zero-g on STS-90 were 7.4 K and 11.8 K at the same EMP temperature, but at ET = 900 min. Previous lab results were 6.8 K and 11.2 K at ET = 1075 min., consistent with the increased thermal conductivity at the higher temperature of 296 K on the EMP. Although convection effects for the lab data are thought to be entirely negligible for the MPTFC [1], there remains a small measurable difference between the zero-g steady state data and the lab data when corrected for the EMP reference temperature differences and the long equilibration times.
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The dynamic conductance measurements in zero-g would be consistent with a soft vacuum for the space experiment in the early part of the warm-up due to ice sublimation, with some residual vapor persisting for several hours until the high vacuum condition was fully reestablished. We therefore conclude that the anomalous dynamic conductance can be attributed to inadequate prelaunch pumpout of the vacuum housing and that otherwise the warm-up rates would have been consistent with previous laboratory data, i.e., that there was no zero-g effect on the cold head dynamic thermal conductance.
ON-ORBIT TEMPERATURE OSCILLATIONS Another unexpected result of the flight demonstration was the amplification of the 90 min. orbital temperature oscillations during the time of active cooler operation, i.e., compressor “on”. This was observed in the low heat capacity cooler and hardware components that were conductively coupled to the EMP. Prior to active cooling the EMP baseline oscillation peak-to-peak amplitude was <1 K for an experiment heat load of ~ 3 W from the flight electronics only. Other components had similar temperature amplitudes. With the compressor “on” the EMP amplitude became ~ 2.4 K for a total heat load of ~ 38 W, and this effect also occurred in all MPTFC warm end components, as shown in Figure 8. Here the ordering of these curves in terms of average temperature and phase angle indicates the direction and time lag of the heat flow through them. Temperature resolution on some RTDs was a factor of ten higher than for others (e.g., at 100 mA constant current excitation there is poorer resolution on the aftercooler 100 ohm RTD than on the experiment housing 1 kohm RTD), but the overall dynamic of an exposed orbiting heat rejection surface (the EMP) is quite obvious. The explanation for the increased amplitude of the oscillations is that the larger experiment heat input (with the compressor operating) added to a comparable solar input causes a higher temperature on the EMP during the orbital day, and this in turn results in enhanced heat rejection during the orbital night to achieve the required energy balance. A preliminary model [23] has been developed to explain the oscillations in terms of a system of four differential equations describing the energy balance on the four primary subsystems: the EMP, the experiment/compressor, the flight electronics/compressor drive, and the battery subsystem. Each of the last three subsystems generates heat and conducts it into the EMP where it is rejected. Results of a simpler three node system give good agreement with the EMP oscillation data using two variable parameters, the effective earth temperature and the EMP view factor. We further note that in Figure 8 the compressor drive module, located on standoffs within the convection-cooled flight electronics compartment, is effectively shielded from these orbital oscillations.
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One of the complications of large thermal oscillations at the heat rejection surface is their effect on the stability of the stage temperatures. In Figures 5 and 6 the effect is imperceptible against the backdrop of the larger overall cooling. Closer examination reveals the results shown in Figure 9, where the oscillation amplitudes of the two stage temperatures are compared by referencing their lowest excursions. Each stage temperature not only drifts but also oscillates with the EMP temperature oscillations. The 2.4 K amplitude of the EMP is attenuated at the stages, and phase lags of several minutes are obvious. These effects follow directly from the cold end temperature dependence on the PT cooling mechanism, which varies nonlinearly with the warm end temperature. An average peak-to-peak amplitude of ~2.4 K at 275 K causes a peak-to peak swing of ~ 0.67 K at 182 K and ~ 0.45 K at 97.5 K. These dynamic values are about half as large as what might be inferred from the steady state stage temperature dependence on the rejection temperature [2, 24]. Even such small long period oscillations at stage 2 would present an additional difficulty for good temperature regulation of an IR focal plane array or other temperature sensitive detector by requiring feedback to the compressor drive or out of phase electrical heating at the cold end. Shielding of heat rejection surfaces from orbital temperature variations is therefore desirable for flight cryocoolers.
CONCLUSIONS On-orbit cooldown and cooling performance data by a miniature PT cooler (MPTFC) flown on STS-108 are equivalent to post-flight laboratory data. Anomalous dynamic conductance data during warm-up were determined to be the result of outgassed water vapor which had been cryopumped from the vacuum housing internal surfaces during nominal on-orbit cooling. Although low rate cryopumping would have caused no measurable effect on the cold head temperatures, additional radiation heat loading from an increase in the emissivity was also not observed. Steady state thermal conductance data in zero-g and lab simulations are nearly identical to previous data, but the conductance for both stages in zero-g is slightly lower than the conductance in 1 g. Orbitinduced thermal oscillations of the heat rejection surface are amplified with higher experiment heat input to the rejection surface and can therefore affect the thermal stability of the PT cold end.
ACKNOWLEDGMENTS We would like to thank Dr. Ron Ross of JPL for providing pertinent information on the onorbit performance degradation associated with cryopumped contaminants in pulse tube cryocoolers.
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REFERENCES 1. D. R. Ladner, R. Radebaugh, and P. Bradley, “Off-state conductance measurement of the NIST/Lockheed Martin miniature pulse tube flight cryocooler: laboratory vs. space,” Adv. in Cryogenic Engineering, Vol. 47B, Amer. Institute of Physics, Melville, NY (2002), p. 1085. 2. P. Bradley, R. Radebaugh, J. H. Xiao, and D.R. Ladner, “Design and Test of the NIST/Lockheed Martin miniature pulse tube flight cryocooler,” Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2001), p. 189. 3. S. Zhu, P. Wu, and Z. Chen, “Double inlet pulse tube refrigerators: An important improvement,” Cryogenics 30, (1990), p. 514, 4. P.J. Storch, R. Radebaugh, and J.E. Zimmerman, Analytical Model for the Refrigeration Power of the Orifice Pulse Tube Refrigerator, NIST Technical Note 1343 (1990). 5. J. Gary, D.E. Daney, and R. Radebaugh, “A computational model for a regenerator,” Proc. Third Cryocooler Conference, NIST Special Publication 698, (1985), p. 199. 6. J. Gary and R. Radebaugh, “An improved numerical model for calculation of regenerator performance (REGEN3.1),” Proc. Fourth Interagency Meeting on Cryocoolers, David Taylor Research Center, Report DTRC-91/003, (1991), p. 165. 7. J. H. Xiao, “Thermoacoustic theory for regenerative cryocoolers: A case study for a pulse tube refrigerator,” Proc. 7th International Cryocooler Conference, Report PL-CP-93-1001, Air Force Research Laboratory, Albuquerque, NM (1993), p. 305. 8. J. H. Xiao, “Thermoacoustic heat transportation and energy transformation, Part 1: Formulation of the problem,” Cryogenics 35, (1995), p. 15. 9. J. H. Xiao, “Thermoacoustic heat transportation and energy transformation, Part 2: Isothermal wall thermoacoustic effects,” Cryogenics 35, (1995), p. 21. 10. J. H. Xiao, “Thermoacoustic heat transportation and energy transformation, Part 3: Adiabatic wall thermoacoustic effects,” Cryogenics 35, (1995), p. 27. 11. The heater power is 50 mW with the primary battery (compressor “on”) and 45 mW with the secondary battery (warm-up). 12. D. Butler, GAS eleven node thermal model (GEM), NASA Tech Note, (October 1987); “GAS Thermal Design Summary”, NASA/GSFC X-732-83-8 (July 1983). 13. F. Renken, “Thermal Analysis of a NASA GAS Experiment,” Masters Thesis, University of Colorado (December 1995). 14. Although the experiment can be powered by an external power supply, the experiment battery system was used to simulate the low voltage condition that occurred during the final hours of the flight demonstration. 15. D.R. Ladner, “Model for two-stage pulse tube cold head parasitic heat load,” Adv. in Cryogenic Engineering, Vol. 47B, Amer. Institute of Physics, Melville, NY (2002), p. 1149. 16. W.F. Denig and R. Viereck, GAS Pressure Measurements on Space Shuttle Mission-39, Phillips Laboratory Technical Report PL-TR-96-2101, Instrumentation Papers, No. 348 (April 9, 1996). 17. M. Taylor, Orbiter environmental contamination from water dumps, FCP and FES, Omitron, Inc. Tech Note; personal communications (2002). 18. R. F. Barron, Cryogenic Systems, Oxford University Press, New York, 1985, p. 225 and p. 560. cryogenic surfaces,” 19. R. G. Ross, Jr., “Cryocooler load increase due to external contamination of Cryocoolers 12, Kluwer Academic/Plenum Publishers, New York (2003). 20. G.K. White, Experimental Techniques in Low Temperature Physics, Clarendon Press, Oxford, 1979, p. 259. 21. W. Viehmann and A.G. Eubanks, Effects of Surface Contamination on the Infrared Emissivity and Visible-Light Scattering of Highly Reflective Surfaces at Cryogenic Temperatures, NASA Technical Note TN D6585, NASA/GSFC (February 1972). 22. A two-lump thermal model would have two time constants, but the time for equilibrium is found experimentally to be about 15 hours so the effective “system time constant” is about 3 hours. 23. D. R. Ladner, “Model for Orbit-Induced Temperature Oscillations in PT Cryocoolers” (to be published). 24. T.C. Nast et al., “Miniature Pulse Tube Cryocooler for Space Applications”, Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2001), p. 145.
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High-Tc SQUID Based Gradiometer Cooled by a Cryotiger Gas-mixture Cooler A.P. Rijpma, H.J.M. ter Brake, H.J. Holland and H. Rogalla University of Twente, Faculty of Applied Physics 7500 AE Enschede, The Netherlands
ABSTRACT We investigated the feasibility of a SQUID system for fetal magnetocardiography aiming at a system without a magnetically shielded room and cooled by a cryocooler. For demonstration purposes, we have selected the APD-Cryotiger, a gas-mixture Joule-Thomson type cooler. We use an alumina SQUID holder that contains three primary measurement SQUIDs and two reference SQUIDs. The three primary SQUIDs can be combined electronically into a second-order gradiometer with 6 cm baseline. Test experiments performed in a magnetically shielded room revealed a significant noise contribution arising from the cooler. Because this noise contribution decreases with increasing distance to the cooler cold head, we expect the source to be located in this cold head, most probably due to remanent magnetization. Because of the large field gradient in this remanent field, a second-order gradiometer configuration is dominated by the cooler noise. As an alternative, we formed a first-order gradiometer of the bottom SQUID and the middle SQUID, and corrected for the cooler noise in it by means of the first-order gradiometer output obtained from the middle SQUID and the top SQUID. In this set up, adult magnetocardiograms were successfully recorded. Outside the magnetically shielded room, however, second-order gradiometer operation is required. Therefore, we attempt to identify the source of the noise contribution in order to be able to remove it from the system. In the paper, the demonstrator set-up is described and experimental results are presented and discussed.
INTRODUCTION Over the last few years, fetal magnetocardiography (fMCG) has proven to be a useful tool for the detection and classification of arrhythmias and the study of congenital heart Clinically relevant information in fMCG may be found in the duration of certain waves or intervals. A database with data from many groups on the duration of the various waves and intervals is maintained by our biomagnetism research team.6 As the clinical relevance of fMCG was already anticipated, the feasibility of a cryocooler-cooled SQUID system for fetal magnetocardiography in an unshielded environment was investigated in the past years. Compared to conventional systems cooled with liquid helium and operated in a magnetically shielded room, such a system is expected to be inexpensive and user-friendly.
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DEMONSTRATOR SYSTEM As the aim is an easy-to-handle system, it was decided to use cryocoolers rather than cryogens. This can be implemented in several ways, as discussed in more detail in reference 7. Preferably, the cooler is directly connected to the sensing unit and run continuously. For this, ‘low-noise’ coolers are required (JT-coolers, GM-type pulse tube coolers). More noisy coolers (Stirling, GM) can be applied if time or space separation is used. Such a separation significantly adds to the complexity of the system. Therefore, low-noise coolers are the preferred option. We decided to incorporate the JT-type APD-Cryotiger with a non-magnetic high-performance cold head into the system (figure 1).8 Figure 2 shows a photograph of the demonstrator system. The system is shown in unshielded environment. On the left side of the figure, the cooler compressor can be recognized. To the right, the ‘sensing unit’ is suspended in a crane. The sensing unit is a vacuum enclosure containing the cooler cold head and the SQUID magnetometers. It is connected to the cooler compressor by 7.5 m long gas-lines. To support and position the sensing unit, an existing aluminum crane was used.
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Figure 3 shows the sensing unit in more detail. The central element is an alumina holder, which contains the SQUID magnetometers needed for the measurement of the magnetic fields. It is suspended below the cooler cold head. The whole is placed in a vacuum casing for thermal isolation. Moreover, multilayer insulation (MLI) is applied to reduce the radiative load. To mechanically connect the SQUID holder with the cooler cold tip, a copper disk is used. This disk is fastened to the cold head with brass screws. To connect to the alumina holder, a more elaborate construction is used. First, a slot is made in the holder. Next, a ring is built in this slot by gluing several halve rings of G10-material together. A number of holes in the resulting ring allow the holder to be fastened to the copper disk by brass bolts. Furthermore, to reduce the torque on the cooler cold head when the system is tilted, the copper disk is attached to the vacuum casing with Kevlar fibres. An important design aspect is the sensor geometry. A second-order gradiometer was found to be the optimum configuration. Furthermore, a baseline of 6 cm was found to be a reasonable trade-off between signal sensitivity and noise rejection. The material alumina is chosen for the holder for its thermal conductivity in combination with low electrical conductivity. The latter ensures that the thermally induced magnetic noise from the holder is not an issue.9-10 As a bonus, the material is also mechanically stable. The alumina holder, which contains the three SQUIDs needed to form a second-order gradiometer is shown in figure 3b. Reference sensors are added in order to be able to improve the balance by electronic noise cancellation."
INTRODUCTORY EXPERIMENTS Before the system was assembled, test experiments were carried out on the Cryotiger to check its magnetic, mechanical and thermal interference. As these measurements are discussed in detail elsewhere12,13, we will only describe the main results. Firstly, it was established that the cooler is easily capable of operating at 77 K with 0.5 W load. Next, it was found that the final temperature was unstable when the cooler was operated without load. However, the situation is improved when a heat load is present. This is illustrated in figure 4, which shows the temperature measured over several days with a total load of about 2 W. The long-term drift is about 2 K over a period of 20 days. The main short-term temperature fluctuation was measured to have an amplitude of about 10 mK on a time-scale of about 5-10 seconds. The mechanical interference of the Cryotiger cold head was determined to consist of translations at a frequency of 49 Hz with an amplitude of about 40 nm along the cold head axis. Rotations with an inclination of about 3 µrad were observed as well.
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The electromagnetic interference of the compressor and that of the cold head were examined separately. At about 2 m distance from the compressor, the (50 Hz power line) disturbance from the compressor is indistinguishable from the environmental noise. The remanent magnetism of the cold head appeared to be quite large: at 40 cm distance about with a gradient of
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Shielded Environment After the system was assembled, it was tested inside a magnetically shielded room (VAC, type AK3B) in order to verify that the cooler did not add to the intrinsic sensor noise. Figure 5a shows the magnetic field noise densities of the three primary magnetometers as measured with the cooler on and off. With the cooler operating, two noise contributions appear. Firstly, the spectrum shows an increased noise level below about 10 Hz. As this contribution decreases for increasing distance from the cooler cold head, the source for the increased noise is expected to be located somewhere in the cold head (top half of figure 3a),
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The second noise contribution does not stem from the cold head but from the compressor: periodic magnetic fields. These are not visible in figure 5a, as the compressor was outside the shielded room during this measurement. In contrast, the magnetometer trace in figure 5b clearly shows the 1 Hz magnetic field from the compressor, which is caused by 49 Hz mechanical modulation of a 50 Hz magnetic field source. This disturbance is suppressed by gradiometer formation. It is obvious from figure 5a that a second-order gradiometer configured as 1-2-1 does not make sense since it is limited by the low-frequency cooler noise. As an alternative, we formed a first-order gradiometer of the bottom SQUID and the middle SQUID, and corrected for the cooler noise in it by means of the first-order gradiometer output obtained from the middle SQUID and the top SQUID. The resulting corrected first-order gradiometer is roughly: bottom – 1.2×middle + 0.2×top. As an illustration, we measured the MCG of an adult inside the shielded room as depicted in figure 5b. In this experiment, the compressor was also inside the room (about 2.5 m from the cold head). Because the compressor has a 49 Hz asynchronous motor powered with 50 Hz14, a clear 1 Hz noise contribution results (top curve). By forming a firstorder gradiometer this contribution is significantly suppressed. The remaining noise is dominated by the low-frequency cold-head noise (middle curve). The corrected first-order gradiometer shows a nice MCG with clear QRS-complexes and T-waves (lower curve). Here, the lowfrequency noise is caused by breathing of the subject.
Unshielded Environment Next, the system is operated outside the shielded room (figure 6). In order to get proper SQUID operation, the background DC-field has to be suppressed. This is done with a coil, which is designed to fit around the sensor head and to create the same field at all three sensor locations. The spectral plots show that the second-order gradiometer is dominated by the same lowfrequency disturbance as seen inside the shielded room. Indeed, if the noise levels depicted in figure 5 are used to estimate the second-order gradiometer noise, the correspondence is good. As the system noise exceeds the environmental disturbances, first-order gradiometers are considered. The first-order gradiometer is indeed an improvement over the second-order gradiometer. Moreover, a substantial improvement is obtained by constructing the corrected gradiometer. This indicates that even the first-order gradiometer is dominated by the lowfrequency system noise. To obtain the corrected first-order gradiometer for figure 6, the references were applied to improve the balance (1 à 2%). Figure 6b shows the magnetic field ofa heart beat recorded in unshielded environment with the cooler operating.
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INVESTIGATION OF THE LOW-FREQUENCY EXCESS NOISE The low-frequency disturbance appears to be strongly correlated over the primary SQUID magnetometers. This correlation is confirmed by the performance of the corrected gradiometer as described above. However, in order to suppress this disturbance at its origin, we need to know what the cause is. A number of possible causes are considered and discussed in this section. Firstly, the data-acquisition may introduce additional noise. However, such a contribution is unlikely to disappear when the cooler is switched off. The same holds for background dc-fields or rf-fields. Moreover, these fields decrease the sensor performance and a resulting increase in noise will not be correlated between sensors. Considering the large copper disk present in the system, Johnson noise seems a likely candidate. The drop-off in the noise level around 5 Hz might be related to the self-shielding of the disk15. The disk is eliminated as a source for two reasons. Firstly, the noise of this disk was calculated before construction not to add to the sensor noise; i.e. smaller than Secondly, the disturbance should also be present when the cooler is switched off. The flux-modulation scheme can lead to noise when the modulation of the magnetometers is not synchronized and some cross-talk is present. But this contribution would remain as well when the cooler is switched off. Furthermore, in our system, the magnetometer electronics can be synchronized. Next, the cooler working fluid is considered. As the fluid is a mixture of propane, ethane, methane, argon, nitrogen and neon, no (para)magnetic behavior is expected. Although a contamination (with e.g. ferromagnetic particles) may exhibit such a behavior, a contamination is unlikely. After the system was put together, the gas lines have never been disconnected. Ground currents may cause noise in the same way as any current: through their magnetic field. In our case, ground currents may run between the cooler, the shielded room and the SQUID electronics. The cooler is connected to the system by gas-lines with a metallic outer shield, the system was occasionally grounded to the room with a cable, and the SQUID electronics had shielded cables. Repositioning of these cables did not significantly change the recorded noise and neither did the disturbance change when the grounding was altered (e.g. by removing the connection to the shielded room). At this point, the character of the disturbance should be looked at in more detail. Figure 7 shows the spectrum of a second recording carried out in the shielded room. As the amplitude decreases for increasing distance from the cooler cold-head, we expect the source to be located in this cold head. As additional information, figure 8 shows a time-trace of the same data but filtered to highlight the low-frequency disturbance. This figure shows the disturbance to be highly correlated over the three primary sensors and much less when the reference sensors are considered. In figure 9 this is made more quantitative by means of the coherence.
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Next, thermal fluctuations are considered. Thermal fluctuations of the sensors are relevant as these influence the parameters of the sensor such as the field sensitivity. There are three reasons why this is not a likely cause. Firstly, the references would be expected to experience a related thermal fluctuation and they should show a strong correlation as well. Secondly, from thermal capacitances and resistances it is calculated that any temperature variation due to the cooler with a frequency of a few Hertz or higher is strongly suppressed. Moreover, a phase shift should be present between the temperature fluctuation of the separate sensors. This is not the case, as could already be concluded from the fact that a simple scalar suffices to make the corrected gradiometer. The above considerations leave the combination of a magnetic field and a mechanical motion as a source for the disturbance. Considering the absence of electrical currents in the cold head, it seems likely that a remanent magnetic source moves with respect to the SQUIDs at a low frequency. This movement may be induced by pressure variations in the cold head. The limited correlation with the references indicates that there are either multiple sources or one source that has several independent modes of vibration (e.g. a translation and a rotation). If the source is assumed to be a magnetic dipole oriented along the system axis, then the field experienced by the three SQUID magnetometers is expressed as:
Here is the distance between the sensor and the source dipole with dipole moment m. If a vertical vibration of the source is considered, the field variation experienced by the sensor due to a displacement of the source over a distance L is derived as:
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Along the axis, the field variation due to the movement of a magnetic dipole scales with Using this, the position of the source is estimated by comparing the amplitudes of the disturbance as recorded by the three magnetometers (ratio 1:2½:10). This results in a position about 16 cm above the top SQUID. Figure 10 shows a photograph of that area. Investigations with a magnetoresistive sensor showed that the spacer was strongly magnetic (at ~1 mm distance the field easily exceeded the earth field). Furthermore, the return line has a joint in this area: the broader part. Although it is difficult to discriminate it’s field from that of the spacer, this joint seems magnetic as well. As these parts are within 1 centimeter from where the source of the lowfrequency noise is expected, they are the most likely cause of the excess noise.
DISCUSSION Measurements with the system cooled with the Cryotiger showed an extra low-frequency noise contribution when the cooler was switched on. This contribution seems to be related to remanent magnetism of a spacer. Although it is difficult to imagine this spacer to vibrate with the low frequencies involved, it is just at the location where a simple dipole model predicts the source of the noise. Therefore, we will attempt to demagnetize these parts and the system will be reassembled. If the demagnetization appears to be successful, we will perform biomagnetic experiments outside the shielded room.
ACKNOWLEDGEMENTS This research is supported by the Dutch Technology Foundation (STW), the Institute for Biomedical Technology (BMTI), Philips Medical Systems and Thales Cryogenics.
REFERENCES 1. Peters M.J., Crowe J.A., Piéri J.F., Quartero H.W.P., Hayes-Gill B.R., James D.K., Stinstra J.G. and Shakespeare S.A., “Monitoring the fetal heart non-invasively: A review of methods”, J Perinat Med, 29 (2001), pp. 408-416.
2.
Wakai R.T., Leuthold A.C., Cripe L. and Martin C.B., PACE, 23 (2000), pp. 1047–1050.
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3.
Menéndez T., Achenbach S., Beinder E., Hofbeck M., Schmid O., Singer H., Moshage W. and Daniel W.G., PACE, 23 (2000), pp. 1305–1307.
4.
Kähler C., Grimm B., Schleussner E., Schneider A., Schneider U. and Nowak H., Prenat. Diagn., 21 (2001), pp. 176–182.
5.
van Leeuwen P., Hailer B., Bader W., Geissler J., Trowitzsch E. and Grönemeyer D.H., Brit J Obstet Gynaecol, 106 (1999), pp. 1200–1208.
6.
Golbach E.G.M., Stinstra J.G., Grot P. and Peters M.J., In: Biomag 2000 Proceedings 12th Int Conf on Biomagnetism, Nenonen J., Ilmoniemi R.J. and Katila T. eds. (2001), pp. 495-498.
7.
Rijpma A.P., ter Brake H.J.M., Peters M.J., Bangma M.R. and Rogalla H., Adv. Cryo. Eng. 45 (2000), pp. 1621-1628.
8.
SHI APD Cryogenics Inc, 1833 Vultee Street, Allentown, PA 18103-4783, United States.
9.
Varpula T. and Poutanen T., “Magnetic field fluctuations arising from thermal motion of electric charge in conductors”, J. Appl. Phys. vol. 55, no. 11 (1984), pp. 4015-4021.
10. Kasai N., Sasaki K., Kiryu S., Suzuki Y., “Thermal magnetic noise of dewars for biomagnetic measurements”, Cryogenics, vol. 33, no. 2 (1993), pp. 175-179 11. ter Brake H.J.M., Dunajski Z., van der Mheen W.A.G. and Flokstra J., “Electronic balancing of multichannel SQUID magnetometers”, Journal of Physics E: Scientific Instruments 22 (1989), pp. 560-564. 12. Rijpma A.P., Bangma M.R. et al., “Interference Characterisation of Cryocoolers for a High-Tc SQUID-based foetal heart monitor”, Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2001), pp. 793-802. 13. Bangma M.R., Rijpma A.P., de Vries E., Reincke H.A., Holland H.J., ter Brake H.J.M. and Rogalla H., “Interference characterisation of a commercial Joule-Thomson cooler to be used in a SQUIDbased foetal heart monitor”, Cryogenics, vol. 41 (2001), pp. 657-663, 14. Hohmann R., SQUID-System mit Joule-Thomson-Kuehlung zur Wirbelstrompruefung von Flugzeugfelgen, PhD thesis, Institut fuer Angewandte Physik, Justus Lieblig Universitaet Giessen, Germany (1999). 15. Rijpma A.P., Fetal Heart Monitor: development of a cryocooler-cooled high-Tc SQUID system for fetal magnetocardiography in unshielded environment, PhD thesis, University of Twente, The Netherlands (2002).
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On the Development of a Non-metallic and Non-magnetic Miniature Pulse Tube Cooler H.Z. Dang, Y.L. Ju, J.T. Liang and Y. Zhou Cryogenic Laboratory, Technical Institute of Physics and Chemistry Chinese Academy of Sciences, Beijing 100080, China
ABSTRACT For continuous low-noise cooling of a highly magnetic flux sensitive high-Tc SQUID by using a cryocooler, the cooler systems have to meet rather stringent requirements concerning mechanical vibrations and electromagnetic interference (EMI). These cooler-generated disturbances must be below the intrinsic noise level of the sensors. The pulse tube cooler (PTC) is an attractive candidate for SQUID operation due to the absence of mechanical moving parts in the cold head. But the vibration and EMI noise originating from the metallic-made cold head of a PTC are still enormous obstacles for practical applications. In this study, we first analyzed the main sources of interference signals of a PTC system; then we designed and fabricated a non-metallic and non-magnetic miniature PTC (NNPTC) for lownoise cooling of high-Tc SQUIDs. The cooler itself, including all the components, was entirely made of non-metallic, non-magnetic and electronically insulating materials in order to reduce the local magnetic fields and eddy currents from that of a metallic-made cooler. System design of the cooler focused on the selections of non-metallic, non-magnetic materials in the NNPTC system and dimensional layout of each component. The cryogen-free operation of high-Tc SQUIDs by the NNPTC is under way in our laboratory.
INTRODUCTION The successful application of high-Tc Superconducting Quantum Interference Devices (SQUIDs) is strongly tied to the availability of satisfactory refrigeration technology. For many years liquid nitrogen (or liquid helium) was used to cool high-Tc SQUIDs. But there exists a lot of intrinsic inconvenience in transferring and supplying low-temperature liquids and operating with them. Especially in special places like in space, which demands a refrigeration system be able to operate in zero gravity, or in submarines, where the use of cryogenic liquids is restricted or forbidden. In many cases, it is the lack of a satisfactory refrigeration system that hinders the wider acceptance of high-Tc SQUIDS. In recent years, considerable effort has been made to provide mechanical cooling for high-Tc SQUIDs.3 High-Tc SQUIDs are usually operated in a temperature range between 50 and 80K and require a cooling power of several hundred milliwatts. Now single-stage cryocoolers can meet this relatively “simple” requirement. In fact, the difficulties of providing appropriate mechanical cool-
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ing for high-Tc SQUIDs other than by liquid nitrogen filled dewars lie mainly in mechanical vibrations and electromagnetic interferences (EMI) from the refrigeration systems themselves. These cooler-generated disturbances must be below the intrinsic noise level of the SQUIDs. It is an enormous challenge for a cryocooler. For example, the sensitivity of typical SQUIDs used to measure geomagnetic fields is times the local magnetic field. Normal commercial cryocoolers can’t be used to cool SQUIDs because their intrinsic noise is so high that the output of the SQUIDs becomes meaningless. Compared with other regenerative cryocoolers such as G-M and Stirling refrigerators, the pulse tube cooler (PTC) is a more attractive candidate for SQUID cooling because of the absence of mechanical moving parts in the cold head. This feature also has the potential for increased reliability and reduced vibration. Another problem for low noise cooling is the EMI noise originating from magnetic or metallic components of cryocoolers. If a PTC, including all the components, can be entirely made of non-metallic, non-magnetic and electronically insulating materials, the EMI noise and eddy currents from the refrigeration system could be reduced to negligible levels. In this paper, the development and design considerations of a new non-metallic and non-magnetic miniature PTC are presented.
MAIN SOURCES OF INTERFERENCE SIGNALS OF A PTC In general, the main interference signals of a PTC system are of mechanical vibrations and EMI (including magnetic signals and eddy currents). The EMI signals can be caused by additional moving magnetic fields in the proximity of the SQUIDs, or by movement of the SQUIDs in an inhomogeneous field; the mechanical vibrations caused mainly by the vibration of the system. In conclusion, the sources of interference signals are mainly the following: (1) The compressor system (2) The vibrations of a PTC system, especially that of the cold head (3) The magnetic materials in a PTC system; the eddy currents originating from the movement of the metallic materials of a PTC system in an inhomogeneous field Compressor System. The main interference signals originating from the compressor system are from two sources: 1) the EMI noise from the movement of the magnetic or metallic-made components of compressors in a magnetic field; and 2) mechanical vibration transmitted to the cold head from the moving components of compressor system. The above two kinds of interference signals can be reduced greatly or eliminated by three approaches as follows: A. Increasing the distance between the compressor system and the cold head (on which the SQUIDs are mounted) so that the compressor EMI signals are greatly reduced. B. Shielding the compressor system when necessary. The distance between the compressor system and the cold head can not be increased infinitely because the cooling power and the lowest temperature decrease gradually with increasing distance; thus the EMI signals from the compressor system may not be reduced sufficiently by increasing the distance. In this instance, we may supply a magnetic shield for the compressor system. C. Employing flexible plastic or polyimide line for the connection of the coolers to the linear compressor in order to reduce the mechanical vibration transmitted from the compressor system. The Vibrations of a PTC System. Pressure-wave induced elastic oscillations of the pulse tube and cold head are the main source of the periodic interference signals. This kind of “breathing” of the cold head can generate obviously observed signals. For example, in a field gradient of the order of 250 nT/m along the cold head axis (z-direction), a motion of 10 µm will lead to a field variation of about The pulse tube length change can be determined by Hook’s law 6 : Where, L is the length of the pulse tube, r is the radius of the pulse tube,
is the peak-to-peak
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pressure amplitude, E is the Young’s Modulus, and h is the wall thickness of the pulse tube. is thus a linear function of the When tube dimensions and materials are determined, reducing to some extent can attenuate the periodic interference signals without degrading the performance of the cryocooler too much. Another method to further decrease the vibration is a vibration compensation method.2 The sensor is mounted on a separate cold platform that is thermally, but not mechanically connected to the cold tip of the PTC. The cold head of the new cold platform is mechanically supported by means of a tube that are fixed at the warm end of the PTC, as described in Ref 2. Magnetic Materials and Eddy Currents. Any magnetic materials in a PTC system directly introduce an additional magnetic field to the local magnetic field. And metallic components moving in an inhomogeneous field (for example geomagnetic field) inevitably produce additional eddy currents that can bring interference to the SQUID’s operation. The best solution is to avoid using any magnetic or metallic materials in the manufacture of the PTC. At present, it is a crucial and unavoidable problem in designing and fabricating a non-magnetic PTC. Unfortunately, there is little prior work on this challenging subject. The main components of a PTC include pulse tube, regenerator tube, connecting flanges, cold head, regenerator matrix, flow straighteners, and vacuum chamber. In this paper we focus our attention on designing and fabricating all of this components using non-magnetic, non-metallic, and electrically insulating materials.
DESCRIPTIONS OF SYSTEM DESIGN Experimental Set-up We adopted a co-axial PTC configuration for our NNPTC in order to minimize the size of the cryocooler. This configuration is also very compact and convenient for connection with the cooled sensors. Figure 1 shows a schematic diagram of our specially designed PTC for cryogen-free operation of SQUIDs.
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Selection of Materials Pulse Tube and Regenerator Tube. Thin stainless steel tubes are widely used for the pulse tube and regenerator tube. However, stainless steel exhibits marked residual magnetization. For example, the magnetic susceptibilities of most stainless steels at a temperature of 77 K are greater than 1.02. Therefore, we recommended and glassfilled epoxy resins1 instead of stainless steel. Ti-A13-V2.5 exhibits very weak magnetization, but it still is a metallic material, so eddy currents remain a severe problem for highly sensitive SQUIDs. Glassfilled epoxy resin is a nonmagnetic and non-metallic material; however, due to its intrinsic molecular structure, high-pressure helium gas will diffuse through it unavoidably, which has been proven by our experiments. At one point we fabricated a regenerator tube with an inner diameter of 11mm, a wall thickness of 2 mm, and a length of 52 mm using glassfilled epoxy. When the charge pressure of the helium gas was about 1.6 MPa, obvious diffusion from the tube wall was detected. In addition, the mechanical strength of this kind of material is not satisfactory. In our experiments, the regenerator tube is made of a special machinable ceramic and the pulse tube is of Nylon 1010. As shown in Fig. 1, the pulse tube is assembled inside the annular regenerator tube. The regenerator tube has an inner diameter of 11mm, a wall thickness of 1.25mm and a length of 52mm. The pulse tube has an inner diameter of 5.5mm, a wall thickness of 0.5mm and a length of 65mm. The main properties of the machinable ceramic are shown in Table 1. It should be pointed out that the diffusion of helium through the machinable ceramic tube is almost the same as the diffusion through stainless steel. Also the axial thermal conduction losses from the hot end to the cold end are also reduced greatly due to its much lower thermal conductivity than that of stainless steel (13 W/m-K at 200K).] Cold Head. The cold head is usually directly coupled to the cooled SQUIDs, so it has the strongest impact on the sensors. Therefore, the selection of the materials of cold head is much more crucial. Pure copper is often used in a common commercial PTC to make cold head due to its high thermal conductivity. However, for high sensitively cooling SQUIDs, it is obviously unsuitable because of the eddy currents problem, which has been tested experimentally by Gerster.1 Pure titanium is often recommended due to its weaker magnetization 2, but it can’t eliminate eddy currents since titanium is still a metal. Gerster et al.1 has reported that they made a solid cold tip of polycarbonate, then glued seventeen Ml.6 threaded pins made of aluminum into the polycarbonate cold tip to guarantee sufficient heat conduction. Although it damped the eddy currents greatly compared with a traditional metallic cold head, the impact of eddy currents was not eliminated completely because of the metallic material (aluminum). In order to make an ideal cold head for a nonmagnetic PTC, the material should be completely nonmagnetic, nonmetallic, and electrically insulating. In addition, it should have relatively high thermal conductivity for the heat transfer between the cold head and the SQUIDs. After carefull considerations, a special boron nitride ceramic was adopted for the cold head. The main properties of the boron nitride ceramic are shown in Table 2. Another advantage of the boron nitride ceramic is easy fabrication. The cold head is designed with a columnar shape so that it can be conveniently glued to the outer wall of the regenerator tube. The cross section of the cold head is shown in Fig. 2.
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Regenerator Matrix. The magnetic impurities, induced eddy currents, and thermally activated currents in a metallic regenerator could produce considerable magnetic disturbance [1]. An important task in designing and fabricating a NNPTC is to find a kind of regenerator material that is non-magnetic, non-metallic, and electrically insulating, and that will not degrade the efficiency of the cryocooler with respect to the dimensions of the regenerator housing. A lot of different materials have been tested experimentally, and the screens made of PA6.6 and Teflon are found to be competent, as described in Refs. 1 and 4. In fact, these two kinds of materials are satisfactory to some extent, and considerable effort has been carried out in our laboratory to find a better one; these results will be reported in other papers. In our experiments described here, the regenerator matrix consists of a stack of 350-mesh PA6.6 screens that are annular in shape, and placed concentrically between the pulse tube and regenerator outer wall (see Figure 1). Connecting Flanges, Vacuum Chamber and Flow Straighteners. In a common pulse tube cryocooler, stainless steel is usually used to fabricate the connecting flanges and vacuum chamber. In order to replace the material exhibiting marked residual magnetization, acrylic glass was used; this is a non-metallic, non-magnetic and electronically insulating material, and its mechanical strengths can meet the requirements of vacuum chamber and connecting flanges. In the NNPTC system, we have designed four flow Straighteners. Two of them are placed at the inlets of the pulse tube and the regenerator tube, and the other two are located at the outlets of them. All of the Straighteners are made of polytetrafluoroethylene plastic, replacing the traditional copper or stainless steel to eliminate magnetic impurities and eddy currents. Flow straightener 16 is placed between the cold head and the pulse tube (as shown in Figure1). It is a vaulted shape with some serrated holes at the bottom and well-proportioned grooves in the outer wall (see Figure 3) so that the gas flows straight and evenly. Adhesive technology. As shown in Figure 1, both the cold head and the flange at the hot ends are connected to the regenerator tube by a kind of special synthetic epoxy resin adhesive named DW-3. This material can be used at temperatures between -269°C and +60°C to bond various kinds of metallic and non-metallic materials with different expansion coefficients. The lower the temperature, the higher the adhesive strength.
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Compressor System. In our experiments, we use a linear compressor with a swept volume of 2cc. To reduce EMI signals from compressor system, the connecting flexible line between the compressor system and the cold head can be as long as 1.5m. The flexible tube has an inner diameter of 2mm and a length of 2m. The compressor system is shielded by µ-metal plates to eliminate its EMI signals (see Figurel). Others. The working fluid is pure helium gas, and the operating frequency of the compressor is around 50Hz. The charge pressure is from 1.6 to 2 MPa.
CONCLUSIONS A non-metallic and non-magnetic miniature coaxial pulse tube cryocooler for cooling high-Tc SQUIDs has been designed and fabricated. All of the components of the PTC are fabricated with non-magnetic, non-metallic and electrically insulating materials to minimize EMI signals. The regenerator tube is made of a kind of special machinable ceramic without diffusion of helium into the vacuum chamber and less thermal conduction loss. The pulse tube is made of Nylon 1010 and the regenerator matrix consists of a stack of 350-mesh PA6.6 screens. The connecting flanges at the hot ends and the vacuum chamber are made of acrylic glass. The cold head is made of a special boron nitride ceramic with high thermal conductivity. Both the cold head and the flange at the hot ends are connected to the regenerator tube using a kind of special synthetic epoxy resin adhesive, and all of the flow straighteners are made of polytetrafluoroethylene plastic. Three approaches were employed to eliminate the interference signals originated from the compressor system: 1) increasing the distance between the compressor system and the cryocooler, 2) shielding the compressor system, and 3) fabricating the tube linking the compressor system to the cryocooler from flexible materials. The cryogen-free operation of high-Tc SQUIDs by the NNPTC is under way in our laboratory, and detailed experiment data will be presented later.
ACKNOWLEDGMENT This work is funded by the National Natural Science Foundation of China (Grant No. 50176052) and the Chinese Academy of Sciences.
REFERENCES 1.
Gerster, J., Kaiser, G., Reibig, L., Thurk, M., and Seidel, P., “Low Noise Cold Head of a Four-Valve Pulse Tube Refrigerator,” Advances in Cryogenic engineering, Vol. 43 (1998), pp. 2077-2084.
2.
Lienerth, C., Thummes, G., and Heiden, C., “Progress in Low Noise Cooling Performance of a PulseTube Cooler for HT-SQUID Operation,” ASC 2000, Virginia USA Paper: 3EF04 (2000), pp. 353-362.
3.
Heiden, C., “Pulse Tube Refrigerators: a Cooling Option for High-Tc SQUIDs,” SQUID Sensors: Fundamentals, Fabrication and Applications, Kluwer Academic Publishers, Netherlands (1996), pp. 289-305.
4.
Gerster, J., Krause, J., Thurk, M., and Seidel, P., “Use of Polymer Materials for the Regenerator of a Pulse Tube Regenerator,” ICEC17 (1998), pp. 97-100.
5.
Thummes, G., Landgraf, R., Giebeler, F., Muck, M., Heiden, C., “Pulse Tube Refrigerator for High-Tc SQUID Operation,” Advances in Cryogenic engineering, Vol. 41 (1996), pp. 1463-1470.
6.
Lienerth, C., Thummes, G., and Heiden, C., “Low-Noise Cooling of HTc-SQUIDs by Means of a Pulse-Tube Cooler with Additional Vibration Compression,” ICEC18 (2000), pp. 555-558.
Cryogenic Refrigerator Evaluation for Medical and Rotating Machine Applications R.A. Ackermann, D.A. Grey*, S. Funayama** and K. Ito** General Electric Global Research Center Niskayuna, NY 12309 *General Electric Medical Systems Florence, SC 29501 ** Sumitomo Heavy Industries, Ltd Tanashi-City, Tokyo 188-8585, Japan
ABSTRACT During the past two years, the General Electric Global Research Center (GE GRC) has been involved in two cryogenic projects in support of GE’s product divisions. The first project was conducted for GE Medical Systems, and involved the evaluation of Gifford McMahon (GM) cryocooler performance. The evaluation described in this paper was performed after an apparent failure of a large number of refurbished cryocoolers. The lost service productivity and increased costs resulting from the failures were significant enough to warrant a detailed examination of the problem. The root cause of the failures was evaluated using Six Sigma methodology and a corrective program developed with the supplier. The second project was the development of a superconductive generator cryogenic refrigeration system for GE Power Systems. The requirements for the superconductive application and the general approach are presented in this discussion.
INTRODUCTION Six Sigma methodology, an important failure analysis tool, was applied in this investigation to determine the root cause for the failure of refurbished cryocoolers installed in Magnetic Resonance Imaging (MRI) machines. By designing an experimental test program (DOE) to evaluate the performance of the refurbished units, we were able to determine the primary failure mechanism and propose remedial action. Through this systematic approach, the field service problem was remedied, resulting in the proper function of refurbished cryocoolers. The second major GE GRC cryogenic program, the development of a superconductive electric generator, focused on developing a high-temperature superconducting rotor coil, helium transfer coupling, and cryogenic cooling system. Prototypes have been built for each component, and tests conducted to reduce development risks. The requirements for the cryogenic cooling system and GE’s risk reduction progress are discussed in the following sections.
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CRYOREFRIGERATOR FAILURE EVALUATION The zero helium boil-off technology employed by GE Medical Systems for many of its Magnetic Resonance Imaging Systems (MRI) uses GM refrigeration systems capable of recondensing helium at 4.2 K. The GE cooling specification for these systems, at 50 Hz is 0.9 W at 4.2 K on the second stage, with 37 W capacity at 40 K on the first stage. Recently, as the MRI-installed base of zero boil-off systems increased and normal service intervals for systems occurred, it became evident that systems serviced with refurbished GM units were failing to meet the zero boil-off specification, resulting in large, unplanned liquid helium usage and increased service costs. Additionally, Field Service reported that when a refurbished unit was replaced with a new GM cooler, the systems achieved the zero boil-off specification. The first reaction to this finding was that the refurbished units failed to meet specifications, and that action by the supplier would be required to ensure the performance of the refurbished units. However, examination of supplier’s refurbished unit performance records and test procedures revealed that the required quality and performance levels had been maintained, and that supplier acceptance procedures were not the root cause of the magnet’s failure to meet the zero boil-off criteria. To determine the true cause of the failures, GE and Sumitomo Cryogenics established a program using Six Sigma methodology. The objectives of the program follow. 1. To establish the difference in performance between new and refurbished units. 2. To isolate the primary element causing the change in performance. 3. To determine the root cause of the difference in performance. 4. To review GE’s service procedures to ascertain if changes that would affect magnet thermal performance have occurred. The first step in the investigation was to review the supplier’s acceptance criteria for refurbished units. While the supplier’s acceptance criteria for refurbished units meets GE’s specifications, new units from the supplier were achieving larger first and second stage capacity. The improvement in 50 Hz performance is shown in Table 1. Over the past 3 years, the performance of new units supplied to GE has steadily improved. Over the same period, however, the performance level of refurbished units has been unchanged. The difference in performance has likely masked variances in magnet thermal performance occurring from magnet manufacturing variances, cryostat thermal degradations and servicing procedures. The emphasis of the program, therefore, was directed at determining the root cause of the performance difference. The method used to establish the principal effect influencing performance was to verify that units returned from the field failed to meet GE performance specifications, and to systematically change components and measure how performance was affected. The experiment (DOE), designed to achieve this goal, was a full factorial DOE with three variables (as shown in Table 2). Testing included eight runs per unit, in which the cold head was run with the old and new displacers, the first stage seal, and the displacer cylinder in the order shown in Table 2. Tests were repeated on several cold heads for verification.
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DOE TEST RESULTS As shown in Figure 1, the variable having the most significant effect on performance was the first stage displacer, (DOE Test 2). The following diagnostic steps were taken to determine the root cause of the first stage displacer failure. 1. An x-ray of the displacer was taken to determine if regenerator material movement occurred. 2. The displacer was cut and its lower section removed, exposing the regenerator material, which was removed and visually inspected for anomalies. 3. A microscopic evaluation was performed on the regenerator material to determine if surface contamination or deformations were present. 4. The outer cylindrical surface of the displacer was inspected for wear. The x-ray of the displacer is shown in Figure 2. The composition of the regenerator consisted of the displacer housing, wire mesh screens in the top of the displacer (warm end), and lead spheres
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located below the screens. Fiber pads were used to separate the two sections and prevent lead from moving into the screens. As shown in Figure 2, no obvious movement or mixing of regenerator materials occurred. Therefore, we concluded that a visual inspection of the regenerator internals was necessary to ascertain the failure mode. The second step in our evaluation was to cut the displacer below the lead spheres and examine the lead for solid contamination and any deformation. As shown in Figure 3, the lead adhered together to form clumps of material in the bed. The clumps were examined to determine if they were sintered together, potentially producing a channeling of the flow through the lead. Figure 4, a microscopic view of the lead surfaces, reveals the lead spheres to be deformed and adhering together to form the clumps. The result is that the lead is no longer cleanly packed with a uniformlydistributed flow, but sintered together, altering the flow pattern in these regions of the regenerator. The final step in the investigation was to examine the outer surface of the displacer for wear. The investigation revealed considerable wear on the surface of the displacer in the region of the cold end exhaust ports shown in Figure 2. In this application, the cold head is not run in the vertical
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position, but is located at -10 degrees from the horizontal. We concluded that the orientation of the cold head produced the wear, and that the wear at the exhaust holes of the displacer further contributed to the asymmetric flow through the regenerator.
CONCLUSION From evidence gathered during the physical evaluation, we concluded that the degradation in performance of the first stage was caused by the asymmetric flow in the regenerator. By installing a new first stage displacer in refurbished units, performance greatly improved, nearly matching that of new units (see Table 3). To date, this solution has successfully eliminated the field service performance degradation issue for GE.
SUPERCONDUCTIVE GENERATOR PROGRAM The Superconductive Generator Program at the GE GRC is a GE Power Systems and DOE partnership-sponsored program designed to develop and commercialize a high-temperature, superconducting rotor. The rotor will replace copper rotor windings in an electrical generator, which will use a conventional GE generator stator. The goal of the program is to develop the essential cryogenic components and superconducting rotor to enable this retrofit. The cryogenic components to be developed include the cryogenic refrigeration system, the helium transfer coupling, and the heat exchanger for the superconductive coil. The interrelationship of these components is shown in Figure 5. The cryogenic design for the cooling system was configured to have a stationary refrigeration system with a helium transfer system to transfer the heat from the generator to the cold box. A helium transfer coupling is used to convert the stationary flow from the refrigeration system to the rotating coil and the coil heat exchanger. The specifications for the cooling system, based on a GE generator retrofit, are shown in Table 4.
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In addition to cooling thermal and transfer flow requirements, the following mechanical and environmental conditions must be met. Vibration. When mounted to the base of the generator, the cryocooler components must be able to operate with no degradation of life with 60Hz and 120Hz vibration of up to 0.125 mm, peakto-peak. Humidity. The cryocooler cooling system shall be capable of operating in 100% relative humidity for extended periods of time. Temperature Variations. The cryocooler cooling system shall be capable of operating in an ambient environment of –25 C to +40 C. Turbine Oil. The cryocooler cooling system shall be capable of withstanding exposure to heavy petroleum oils. Explosion Risk. Electrical components mounted on the cold box shall be H2 explosion proof.
Reliability The superconductor must also meet the following reliability requirements. 1. Mean time between maintenance (MTBM) shall be greater than 24 months. 2. The failure rate for the cooling system must be less that 2% over the planned preventive maintenance schedule. 3. The life of the cooling system shall be greater than 30 years. The prototype generator cooling system schematic GE will investigate is presented in Figure 6. The system is comprised of a GM cryocooler that cools an external helium transfer system. The transfer system and the cold head are housed in a cold box mounted to the transfer coupling through a vacuum-insulated, flexible bellows attachment. The transfer system consists of a helium circulation pump, a counterflow heat exchanger, HX 1, a single pass heat exchanger, HX 2, and a flow ejector. To achieve the 100 W of cooling for the generator, the system is designed to have an inlet cooling temperature of 30 K with a helium flow rate of 2.0 g/s. To meet transfer coupling sealing requirements, the maximum inlet pressure to the circulator is limited to 1.7 bar absolute and the maximum pressure drop across the coil is maintained at less than 0.35 bar.
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PROTOTYPE TEST RESULTS The results of the prototype cooling system tests (Table 5) indicate that the system is capable of meeting current thermal specifications. Tests also revealed that the cooling capacity of the cooling system is reproducible, as indicated by the small standard deviations. Eight tests were run for periods of 10 hours or greater to measure the system’s variabilities.
CONCLUSION Tests conducted on the generator cooling system indicate that a regenerative cryocooler is a practical solution for the cooling capacities in this size range. The temperature stability and repeatability is good; the efficiency (150 to 1 W/W) is acceptable; and the proven reliability of GM cryocoolers is critical to the quality of the product. GE will continue to evaluate this design to ensure that the other mechanical and environmental criteria can be met. In addition, GE will evaluate whether this design approach can be effectively extended to systems requiring up to 500 W of heat removal.
REFERENCES 1.
Brue, Greg. Six Sigma for Managers, McGraw-Hill, New York (2002).
2.
Yukio, Mikami; Toshiharu, Yanada. Heat Radiation Shielding Plate Cooling Device, Patent Abstracts of Japan, Publication no. 10-311618, 11/24/1998.
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Helium Free Magnets and Research Systems J. Good, S. Hodgson, R. Mitchell, R. Hall
Cryogenic, Ltd. London W3 7QE, United Kingdom
ABSTRACT Cryocoolers operating at 4K and lower temperatures permit research systems and superconducting magnets to operate entirely without liquid helium. We describe the production of magnets that are directly conduction cooled and operate in vacuum as well as arrangements for cooling samples in a controlled environment from 300K to 1.6K using only one single cryocooler. Many designs use the 4K cryocooler to recondense liquid helium but these are different from true Cryogen Free designs. The magnet or system cryostat is relatively conventional in design with a liquid helium reservoir and the cryocooler is used to recycle and recondense helium leaving the reservoir. The systems need liquid helium and nitrogen for precooling but then can be kept cold for long periods. Our design for direct conduction cooling in vacuum has important advantages. First the design can be kept compact as there are no cold liquid vessels. All electrical feed throughs, including high current leads are at room temperature. The cooldown uses only the refrigerator. The internal design is simpler and there are fewer risks for vacuum leaks. Magnets up to 750mm in diameter providing 0.5 Tesla have been built as well as 15 Tesla small bore (50mm) coils. Samples can be cooled to low temperatures and their properties measured in a controlled manner without imposing high heat loads on the cryocooler, which is also used to cool the magnet. This has been achieved over the 4K to 300K range using a gas convection cooled sample chamber, the design of which has been patented. Lower temperatures to 1.6K can be attained by pumping on a small volume of condensed liquid helium.
INTRODUCTION The production of high field cryogen free magnets (CFMs) delivering fields up to and in excess of 15 Tesla and in a wide range of geometries has become quite routine. However, it is now possible to manufacture a measurement system that provides both a high magnetic field and a variable temperature environment for the sample covering a range from 1.5 to 300 Kelvin using a single cryocooler as the source of cooling power. The performance of the variable temperature sample space is comparable to that of a standard variable temperature insert (VTI) that is used with a liquid 4He cryostat, in terms of both the temperature stability and the speed of response of the system. It is possible to control the temperature of the sample to better than (5 milliKelvin stability and to cool the sample from room temperature to below 4 Kelvin in less than 2 hours. No cryogens (liquid 4He or N2) are required at any stage of the operation of such systems. This is a Cryocoolers 12, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2003
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great advantage in those parts of the world where these liquids are prohibitively expensive. In addition, the basic system can be tailored to a wide variety of measurement requirements. Systems can be supplied with the probes, electronics and software necessary for a complete measurement station. For example, pick up coils for AC susceptibility measurements are often incorporated into this type of system. Also under development is a 3He version of the system to produce temperatures down to 300 milliKelvin. This paper describes the design and operation of a 14T magnet system with built-in VTI. Also presented are the results of factory testing.
DESCRIPTION OF THE SYSTEM The system described here (Fig. 1) consists of a 14T magnet and a VTI with a temperature range of 1.5K 300K. A single two-stage cryocooler is used to cool down both the magnet assembly and the return gas from the VTI. On this unit the flow around the gas circuit is established using an external oil-free scroll pump that permits continuous operation at 1.5K rather than using an internal gas convection design which limits the base temperature to around 4K. The magnetic field is generated using two concentric superconducting solenoids, the outer being wound from NbTi and the inner from Nb3Sn conductor. Both coils are thermally grounded to the second cooling stage of the cryocooler coldhead. The second stage of the cryocooler provides a working temperature from 2.7 to 4K. The leads carrying current to the magnet must necessarily have a high thermal gradient i.e. from room temperature external to the cryostat to the operating temperature of the magnet (~4K and below). In order to reduce the thermal load on the magnet assembly the current leads are in part constructed from HTS material. The leads are thermally grounded to the first cooling stage of the cryocooler coldhead, which runs at 30 40K. A superconducting switch with a heater to drive it above its transition temperature is connected across the coil to allow operation in the persistent mode. The whole assembly is protected from thermal radiation by aluminium radiation shields that are thermally grounded to the first cooling stage of the cryocooler coldhead. Multiple layers of superinsulation provide additional shielding. The operation of the sample space is similar to that of other 4He cooled sample spaces in that liquid 4He is drawn through an impedance, in this case a needle valve, causing adiabatic expansion and hence further cooling. The helium is drawn from a small reservoir which is replenished by gas condensed from room temperature to 1.5K typically. This flow of cold helium gas provides the cooling source for the sample. The helium passes through a heat exchanger to which are fixed a heater and a thermometer. The heat exchanger and thermometer set the temperature of the gas passing to the sample space which can be set to any temperature from 1.5K to 300K. The flow rate is kept the same at all temperatures. The reservoir is very small (approximately l00cc)
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and is maintained full by recondensing the return gas from the mechanical dry pump at room temperature at the second stage of the cryocooler. Re-condensation occurs in two stages; initial cooling from room temperature to 30K occurs at the first stage of the cryocooler where there is a large cooling power. Provided that the flow rate is controlled the second stage is not unduly loaded. To ensure that the liquid helium for the sample space is always at the same temperature the small reservoir is not in thermal contact with the second stage of the cryocooler. Temperatures throughout the system are monitored using calibrated RhFe thermometers. In cases where the magnetic field is strong Cernox or similar thermometers are used that have low magneto-resistance.
DISCUSSION The crucial consideration when designing a system such as this is how best to balance the heat loads on the cryocooler due to the VTI and the magnet. Obviously, there is a limit to the temperature above which the magnet will operate; therefore, the magnet must be maintained at a temperature below this limit. The limiting temperature is also reduced as the field increases. To further complicate matters, the rate at which the magnet can be energized is also limited by its temperature, since thermal energy is dissipated in the coil as it is energized. This dissipation is proportional to the rate of energizing for a given field. Therefore, careful consideration must be given to the tune required for maximum field to be achieved. Similarly, successful VTI operation is dependent on the available cooling power. With insufficient cooling power, cooling a sample from room temperature is painfully slow and the base temperature that can be achieved will be higher.
TEST RESULTS The total time for the system to cooldown from room temperature to its operating temperature was 28 hours. The base temperatures at points throughout the system after initial cooldown are given in Table 1. Figure 2 shows the dependence of the magnet temperature on the VTI pressure. It was found that the temperature of the magnet was dependent on VTI pressure alone i.e. independent of the temperature at which the sample is being controlled. This has the advantage that, having defined a rate of flow through the VTI using the needle valve, the sample can be controlled to any
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temperature up to room temperature without affecting the performance of the magnet. It was also noted that the VTI pressure remains constant over the entire sample temperature range. Initial testing of the magnet was done with the VTI needle valve closed. In other words, there was no gas flow through the VTI and therefore no additional heat load on the cryocooler due to the VTI. Table 2 shows the quench history of the magnet. The coil could with further training be expected to achieve fields in excess of 15T, but no attempt was made to achieve higher fields. The magnet and variable sample space was tested further with a pressure of 10 mbar in the VTI. Persistent mode at 14 Tesla was achieved with no quenches whilst controlling the VTI at 10K and 300K.
CONCLUSION The test results show that cooling power for the sample space and magnet can be supplied by a single cryocooler. It was also demonstrated that the two can be operated independently without adversely affecting the performance of the other. The fact that only a single cryocooler is required reduces both the capital cost of the equipment and the operating cost.
REFERENCES 1.
Sumitomo Heavy Industries, Ltd., 1.5 watt/4K SRDK-415D Cryocooler, Cryogenics Department Precision Products Division, 2-1-1 Yato-Cho, Nishi Tokyo, Tokyo 188-8585, Japan
Proceedings Index This book draws upon papers presented at the 12th International Cryocooler Conference, held in Cambridge, Massachusetts on June 18-20, 2002. Although this is the twelfth meeting of the conference, which has met every two years since 1980, the authors’ works have only been available in hardcover book form since 1994; this book is thus the fifth hardcover volume. Prior to 1994, proceedings of the ICC 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 eleven previous proceedings of the International Cryocooler Conference. 1) Refrigeration for Cryogenic Sensors and Electronic Systems, Proceedings of a Conference held at the Nat'l 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 Charred 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 1719, 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. 10) Cryocoolers 10, proceedings of the 10th ICC held in Monterey, California, May 26-28, 1998, Ed. by R.G. Ross, Jr., Kluwer Academic/Plenum Publishers, New York, 1999. 11) Cryocoolers 11, proceedings of the 11th ICC held in Keystone, Colorado, June 20-22, 2000, Ed. by R.G. Ross, Jr., Kluwer Academic/Plenum Publishers, New York, 2001. 817
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Author Index Abhyankar, N., 51, 59 Ackermann, R.A., 805 Akagi, S., 737 Akao, H., 737 Akhtar, S., 771 Alexeev, A., 595 Babitsky, V.I., 717 Baik, J.H., 349 Bailey, P.B., 247, 255, 275 Banks, I.S., 37, 755, 761 Barber, D.S., 627 Bauwens, L., 389 Benschop, T., 31, 87, 165 Bhandari, P., 627, 637 Bhatia, R.S., 651 Bonney, G.E., 411 Borders, J.W., 637 Bowman, R.C., Jr., 627, 637 Boyle, R.F., 1, 37, 755, 761 Bradley, P.E., 523, 777 Bradshaw, T.W., 173 Breedlove, J.J., 563 Breon, S.R., 37, 761 Brest, M.L., 69 Brisson, J.G., 587, 675, 681 Bruins, P., 31, 87, 109, 165 Bruninghaus, C.H.Y., 51 Bugby, D., 693 Burger, J.F., 643
Cai, J.H., 115 Canavan, E.R., 661 Cao, Y., 609, 615, 621 Cauquil, J.M., 87 Champagne, P.J., 205, 225 Chan, C.K., 219 Charles, I.,131, 165, 173, 265 Chase, S.T., 651 Chen, G.B., 325, 451 Chen, G.M., 431, 615, 621 Cheuk, C.F., 247, 275 Clappier, R., 225 Colbert, R,, 191 Collaco, A., 225 Crespi, P., 31 Crook, M., 173
Crumb, D., 627 Curran, D.G.T., 225 Dadd, M.W., 247, 255 Dai, W., 331, 343 Dang, H.Z., 799 Davey, G., 255 Davis, T.M., 219, 255 de Boer, P.C.T., 265 de Koning, A., 109 de Waele, A.T.A.M., 283 Deng, X., 389 Derossett, W.R., 183 Desai, P.V., 547, 555 DiPirro, M.J., 661 Dolan, F.X., 563 Duband, L., 131, 165, 173, 265, 669 Duval, J.M., 131, 265 Elwenspoek, M., 643 Evtimov, B., 205, 213, 225, 241 Fabris, D., 499 Fang, L., 389 Flake, B.A., 9, 59 Fleming, J., 687 Foster, W.G., 225 Francis, J., 661 Fujisada, H., 737 Funayama, S., 805
Gan, Z.H., 325 Gedeon, D.R., 139 Gerstmann, J., 587 Gibson, A.S., 173 Gifford, P.E., 293 Glaister, D.S., 45 Godden, J., 191 Goloubev, D., 595 Gong, M.Q., 603, 609, 615, 621 Good, J., 813 Grey, D.A., 805 Gschneidner, K.A., Jr., 457 Gully, W.J., 45 Guo, F., 439 Häfner, H.U., 149 Hall, R., 813
Hanes, M., 75 Hannon, C.L., 587 Harvey, D., 191 Harvey, J.P., 547 He, Y.L., 325 Helvensteijn, B.P.M., 199 Hendricks, J.B., 483 Hernandez, C.G., 205 Hernandez, P., 669 Hill, N.G., 247, 275 Hodgson, S., 813 Hofman, T., 109 Holland, H.J., 643, 789 Hong, Y.J., 95, 103 Hou, Y.K., 115, 123, 361, 539 Hristov, V.V., 651 Hu, Q.G., 603 Huang, Y.H., 451 Hunt, R., 173, 771 Ichikawa, M., 309 Ikeya, Y., 403, 473 Inoue, T., 309 Ito, K., 805 Itoyama, T., 157 Izenson, M.G., 571 Jackson, M., 661 Jaco, C., 219 Jeong, S., 371, 531 Jiang, N., 325 Jiang, Y.L., 325 Jin, T., 451 Jing, W., 115, 123, 361 Ju, Y.L., 115, 123, 317, 361, 539, 799 Jung, J., 371, 531 Jungkman, D., 225 Kamoshita, T., 157 Kashani, A., 199, 349 Katsuragawa, S., 309 Kawada, M., 737 Keating, B.G., 651 Kelly, K., 489 Khrustalev, D., 709 Kikuchi, R., 157 Kim, H.B., 95, 103
819
820
King, T.T., 661 Kirkconnell, C.S., 183, 233, 379, 547 Kittel, P., 199, 777 Kobayashi, H., 331 Kobayashi, M., 737 Koh, D.Y., 95, 103 Kotsubo, V., 213 Koyama, T., 301 Kroliczek, E., 693 Kudoh, M., 737 Kuo, D.T., 79 Ladner, D.R., 777 Lange, A.E., 651 Lee, K.B., 95 Lester, J., 579 Levenduski, R., 579 Lewis, M.A., 517, 523, 777 Li, N., 325 Li, Q., 439 Li, R., 301, 403, 473 Li, X., 421, 431 Liang, J.T., 115, 123, 317, 361, 539, 799 Linder, M., 165, 173 Ling, H., 421, 431, 447 Liu, H., 447 Loc, A.S., 627, 637 Lody, T.D., 79 Luo, E.C., 421, 425, 431, 447, 603, 609, 615, 621 Maddocks, J.R., 349 Maekawa, T., 737 Mantwill, E., 595 Marland, B., 693 Marquardt E., 45, 507, 579 Martin, J.Y., 87, 131 Mason, P.V., 651 Matsubara, Y., 331, 343 Matsumoto, N., 157 McCandless, A., 489 McCormick, J.A., 563, 571 Miller, F.K., 675 Mills, G., 687 Mirbod, F.I., 69 Mitchell, A., 225 Mitchell, M.P., 499 Mitchell, R., 813 Morgante, G., 627, 637 Motakef, S., 489 Mullié, J., 31, 165 Mustafi, S., 761 Nakane, H., 467 Nam, K., 531 Nash, A., 637 Nast, T.C., 205, 213, 225, 241 Nellis, G.F., 349 Nguyen, T.,. 219
AUTHOR INDEX
Nishihara, O., 737 Nogawa, M., 309 Nozawa, H., 473 Numazawa, T., 397, 403, 467, 473 O’baid, A., 75 Ogura, T., 301 Ohshima, K., 157 Olson, J.R., 205, 213, 225, 241 Orlowska, A.H., 173 Orsini, R., 191 Park, S.J., 95, 103 Pearson, D., 637 Pecharsky, A.O., 457 Pecharsky, V.K., 457 Peskett, G.D., 27 Pfotenhauer, J.M., 349, 523 Philhour, B.J., 651 Phillips, C., 681 Price, K.D., 183, 233 Prina, M., 627, 637 Pruitt, G.R., 183 Pundak, N., 717
Qi, Y.F., 603, 609, 615, 621 Qiu, L.M., 325 Raab, J., 191, 247, 275 Radebaugh, R., 225, 507, 517, 523, 777 Ravex, A., 31, 131, 165, 669 Reed, J.S., 27 Reiter, J.W., 627 Renna, T., 205 Riabzev, S.V., 717 Rijpma, A.P., 789 Roberts, T., 9, 555 Rogalla, H., 643, 789 Rolff, N., 149 Ross, B., 69, 183 Ross, R.G., Jr., 1, 727, 747 Roth, E., 241 Sahimi, M., 379 Salerno, L.J., 199 Salazar, W.E., 17 Sarri, G., 205 Satoh, T., 397, 473 Schmauder. T., 337 Schmelzel, M.E., 627, 637 Schroth, A., 379 Seidel, P., 337 Shen, Y., 451 Shirey, K.A., 37, 755, 761 Shirron, P.J., 661 Sirbi, A., 637 Sirbi, G., 651 Sixsmith, H., 563, 571 Smith, J.L., Jr., 587 Stack, R., 45
Stouffer, C., 693 Strauch, R., 275 Sugita, H., 343 Swift, W.L., 563, 571 Takeuchi, T., 157 Tanaeva, I.A., 283 Tang, K., 451 ter Brake, H.J.M., 643, 789 Thummes, G., 149 Thürk, M., 337 Toma, G., 191 Tomlinson, B.J., 9, 51, 59, 219, 255 Tooyama, S., 343 Toyama, K., 157 Traum, M., 587 Trollier, T., 31, 131, 165 Tsukahara, Y., 157 Tu, Q., 439 Turner-Valle, J., 687 Tuttle, J.G., 661 Veprik, A.M., 717 Wade, L.A., 637 Waldauf, A., 337 Wang, C., 293 Wang, L., 539 Warner, B.A., 761 Wei, Z., 687 Will, E., 225 Wilson, K.B., 139 Wright, G.P., 45 Wu, Jihao., 439 Wu, J.F., 421, 431, 447, 603, 609, 615, 621 Xiao, J.H., 777 Xu, M.Y., 301 Yamaguchi, T., 467 Yamazaki, S., 467 Yan, L.W., 115 Yan, P.D., 301 Yanagitani, T., 473 Yang, L.W., 149 Yang, M., 421, 431 Yarbrough, S.A., 59 Yasukawa, Y., 157 Yoon, K.W., 651 Yoshizawa, S., 467 Yu, B.K., 95 Yu, Z., 439 Yuan, K., 115, 123, 317, 539 Yuan, S.W.K., 79 Zagarola, M.V., 571 Zhou, Y., 115, 123, 539, 603, 609, 799 Zhu, S.W., 309 Zhu, W.X., 115
Subject Index Accumulators (see Thermal storage) ACTDP cryocooler development, 1 ADR (see Magnetic refrigerators) Adv. Mechanical Tech., Inc., 587 Advanced Research Systems, Inc., 411 Aerospace Corporation, 225 Air Force Research Lab (AFRL): Astrium 10K cooler testing, 59 Ball 35K/60K PSC testing, 51 linearity of flexure suspension systems, 255 modeling periodic flow in regenerators, 555 research initiative overview, 9 TRW HCC two-stage PT cooler, 219 Air Liquide DTA (France): 50-80K PT cooler for space, 165 high capacity 5W-80K PT cooler, 131 high capacity Stirling cooler for ISS, 31 sub-Kelvin sorption/PT cooler, 669 AIRS instrument cryocoolers: ground and in-space performance, 747 in NASA cooler programs overview, 1 Aisin Seiki 4K pulse tube, 309 Alabama Cryogenic Engineering, 483 Ames Research Center (NASA): LMA pulse tube flight experiment, 777 TRW HEC PT cooler testing, 199 Ames Laboratory (Iowa State Univ.): low temp, regenerator mat'ls overview, 457 AMS-2 spectrometer: qualification of coolers for, 1, 37 cooler operation in high magnetic fields, 761 APD Cryogenics: Cryotiger in SQUID application, 789 Applications of cryocoolers (see Integration with cryocoolers) ARC (auto-refrigerating-cascade) cycle, 595 ARC, NASA (see Ames Research Center) ASTER flight cooler performance, 737 Astrium cryocoolers (formerly BAe and MMS): 10K 2-stage testing at AFRL, 59 50-80K MOPITT cooler performance, 727 active vibration cancellation systems, 771 miniature space pulse tube cooler, 173 Atlas Scientific, 199, 349 Ball Aerospace cooler activities: 6K ACTDP hybrid Stirling/J-T cooler, 1
35K/60K PSC testing at AFRL, 51 multistage Stirling cooler development, 45 optical cooling mirror leakage effects, 687 parallel plate heat exchangers, 507 Redstone hybrid 10K J-T/GM cooler, 579 Bearings: flexures in tactical coolers, 17, 69 flexures in space coolers, 31 linearity of flexure suspension systems, 255 Brayton cycle cryocoolers (see Reverse-Brayton cryocoolers) British Aerospace (BAe) cryocoolers (see Astrium cryocoolers) Calif. Institute of Tech. (see also Jet Prop. Lab): sub-Kelvin cooler for Polatron, 651 CEA/SBT (France): 50-80K PT cooler for space, 165 high capacity 5W-80K PT cooler, 131 miniature space pulse tube cooler, 173 piston resonance in orifice PT, 265 sub-Kelvin sorption/PT cooler, 669 Chinese Academy of Sciences, Cryogenics Lab: 40-80K linear-compressor PT cooler, 115 60K PT cooler system design, 123 capillary tube expanders for mixed gases, 609 dc flow, caused by regen nonlinearities, 425 dephlegmation separator for mixed gases, 603 GM-type coaxial PT cooler experiments, 317 nonmetallic pulse tube for SQUIDs, 799 oscillating regenerator flow at low temp, 539 role of PT orifice and secondary bypass, 361 thermoacoustic device analysis, 431 thermoacoustics, high freq. operation, 439 thermoacoustics, influence of convection, 447 thermoacoustics, onset temp gradient, 421 vortex tube applied to auto-cascade J-T, 621 vortex tube modeled as heat exchanger, 615 CMC Electronics: tactical cooler lifetest results, 79 Collins-type 10K cryocooler, 587 Compressors: hybrid sorption compressor elements, 627 linearity of linear suspension systems, 255 producibility of linear compressors, 275 reverse Brayton, 571 scaling of linear compressors, 247
821
822
Conductance, of bonded regenerator mat'ls, 517 Contamination: in LMA PT cooler flight experiment, 777 overview, for low-e surfaces in space, 727 Creare Cryocoolers: 6K ACTDP cooler development, 1, 571 6K Brayton cooler developments, 571 NICMOS rev-Brayton, 1, 563 Cryogenic Ltd (UK), 813 Cryomech, Inc: 4K and 10K pulse tubes, 293 4K precooler for sub-Kelvin sorption, 669 Cryotiger, integration of SQUIDs with, 789 DC flow: caused by natural convection, 447 caused by regenerator nonlinearities, 425 Dephlegmation separator, 603 Dilution refrigerator on 4K PT, 669 DRS tactical Stirling coolers, 17 Dynacs Engineering, 59, 51 Eindhoven Univ. of Technology, 283 Electric field emissions (see EMI/EMC measurements and suppression) Electromagnetic interference (see EMI/EMC measurements and suppression) Electronics, vibration cancellation, 771 EMI/EMC measurements and suppression: HTS SQUID applications, 789, 799 operation of coolers in high mag fields, 761 Polatron microwave receiver, 651 Erbium regenerator materials (see Regenerators) European Space (ESA/ESTEC) activities: 50-80K PT cooler for space, 165 Astrium miniature space PT cooler, 173 Lockheed 2.7W-80K PT cooler, 205 Flexible thermal links, 205, 709 Flexure bearings (see Bearings) Friedrich-Schiller-Universität, Jena, 337 Fuji Electric 70K commercial PT cooler, 157 Fujitsu, Ltd., 737 Gas-gap heat switches (see Heat switches) Gedeon Associates, 139 General Electric MRI applications, 805 Georgia Institute of Tech., 547 Gifford-McMahon Cryocoolers: improvement at 4K using GOS, 403 regenerator degradation analysis, 805 thermal hysteresis at 4K, 411 Gimbal mount: thermal transport system for, 693 vibration reduction for, 717 Goddard Space Flight Center (NASA): continuous magnetic refrigerator, 661 NASA cooler program overview, 1 operation of coolers in high mag fields, 761 qualification of coolers for AMS-02, 37
SUBJECT INDEX RHESSI on-orbit performance, 755 GOS regenerator material (see Regenerators) Heat conduction (see Conductance) Heat exchangers: compact high-effect. parallel plate, 507 Heat pipes: cryogenic, 709 Swales development and testing of, 693 Heat switches: Swales development and testing of, 693 HESSI(see RHESSI) High temperature superconductor applications (see Integration of Cryocoolers with) HIRDLS instrument cryocoolers, 1 Holmium regenerator materials (see Regenerators) HTS applications (see Integ. of cryocoolers with) Huazhong Univ. (China), 439 Hybrid multistage coolers: Ball 6K ACTDP J-T/Stirling, 1 dilution refrigerator on 4K PT, 669 Redstone 10K J-T/GM cooler, 579 sub-Kelvin sorption on 4K PT cooler, 669 TRW 6K ACTDP J-T/PT, 1 Hydrides (see Sorption cryocoolers) Hymatic Engineering Co: producibility of linear compressors, 275 scaling of linear compressors, 247 IGC-APD Cryogenics (see APD Cryogenics) INTEGRAL mission coolers, 205, 771 Integration of cryocoolers with: gimbal mounts, 693, 717 heart monitor, 789 heat pipes, 693, 709 heat switches, 693 low-emittance surfaces, 727 MRI systems, 805 Polatron GHz receiver, 651 space experiments (see Space experiments) space instruments (see Space instruments) SQUIDs, 789, 799 superconductive generators, 805 superconductive magnet systems, 813 thermal storage, 233, 693 vibration control systems, 109, 717, 771 ISAMS, in-space cooler performance, 727 J-T cryocoolers: Ball ACTDP 6K cooler development, 1 capillary tube expanders for mixed gases, 609 Cryotiger for cooling SQUIDs, 789 dephlegmation separator for mixed gases, 603 mixed-gas auto-cascade J-Ts, 595 Redstone hybrid 10K J-T/GM cooler, 579 superfluid compressor performance, 675 TRW ACTDP 6K cooler development, 1 vortex tube applied to auto-cascade J-T, 621 vortex tube modeled as heat exchanger, 615 JAMI flight coolers, 191
SUBJECT INDEX Jet Propulsion Lab: ACTDP cooler development program, 1 AIRS PT cooler in-space performance, 747 contamination of low-e surfaces, 727 NASA cooler program overview, 1 Planck hydride compressor elements, 627 Planck EBB breadboard cooler testing, 637 Joule-Thomson Cryocoolers (see J-T cryocoolers) Kleemenko cycle refrigerator, 595 Kogakuin Univ., 467 Konoshima Chemical Co., 473 Korea Adv. Inst. of Sci. and Tech.: surface heat pumping loss in PT, 371 regenerator model with oscillating flow, 531 Korea Inst. of Machinery and Materials: linear free-piston Stirling experiments, 95 linear free-piston Stirling modeling, 103 Leybold Vacuum GmbH: 5W-80K pulse tube cooler, 149 Life estimation methods, 79, 87 Life test results: Ball Aerospace 35K/60K, 51 CMC tactical coolers, 79 Thales tactical coolers, 87 LIGA-fabricated regenerators, 489 Lockheed Martin Astronautics Operations: pulse tube flight experiment, 777 Lockheed Martin Advanced Tech. Center: 6K ACTDP PT cooler development, 1 10K PT for space applications, 241 35K two-stage PT cryocooler, 213 35K/85K high capacity 2-stage PT, 225 80K 2.7W INTEGRAL PT devel., 205 Loop heat pipes, cryogenic, 709 Loughborough Univ. vibration suppression, 717 Magnets, helium free cooling of, 813 Magnetic fields (see EMI/EMC): cooler operation in high mag fields, 761 Magnetic refrigerators: continuous ADR for below 50 mK, 661 regenerator materials for (see Regenerators) refrigerant materials for (see Refrigerants) Massachusetts Institute of Technology: Collins-type 10K cryocooler, 587 superfluid He J-T refrigerator, 675 superfluid He Stirling refrigerator, 681 Materials: conductance of (see Conductance) refrigerants (see Refrigerants) regenerator (see Regenerators) Matra Marconi coolers (see Astrium coolers) Meisei Univ., 467 Messer Cryothenn GmbH, 595 Mezzo Systems, 489 Microcooler, Univ. of Twente sorption, 643 Microphonics (see vibration) Mitchell Stirling, 499
823 MIT (see Massachusetts Institute of Technology) Mitsubishi Electric, 737 Mixed refrigerants (see Refrigerants, and J-T cryocoolers) MOPITT cooler performance, 727 MRI applications (see Integ. of cryocoolers with) NASA (see individual centers): cooler program overview, 1 NASDA (Japan): pressure wave generator for PT, 343 Nat'l Inst. of Adv. Ind. Sci. & Tech. (Japan), 737 Nat'l Inst. of Standards and Tech. (see NIST) NICMOS: reverse-Brayton cooler flight data, 563 in NASA cooler programs overview, 1 Nihon University: 4K VM-type pulse tube cooler, 331 pressure wave generator for PT, 343 NIST: bonded regenerator matrix mat'ls, 517 LMA pulse tube flight experiment, 777 LM-ATC high capacity 2-stage PT, 225 parallel plate heat exchangers, 507 regenerator loss measurements, 523 Northrop Grumman: Elect. Sensors and Systems, 225 Space Technology (see TRW) Optical cooling mirror leakage effects, 687 Optical Engineering Associates, 687 Orientation, effect on performance: Sunpower pulse tube cooler, 139 pulse tube off-state conduction, 747 thermoacoustic driver, 447 Phase change materials (see Thermal Storage) Planck sorption coolers NASA cooler program overview, 1 compressor performance testing, 627 EBB breadboard cooler testing, 637 Polatron, sub-Kelvin coolers for, 651 Producibility, of linear compressors, 275 Pulse tube cryocoolers: 1.5K 3He Eindhoven Univ., 283 4K two-stage Aisin Seiki, 309 4K two-stage Sumitomo, 301 4K VM-type pulse tube cooler, 331 40-80K linear-compressor Chinese, 115 50-80K Air Liquide for space, 165 60K linear-compressor Chinese, 123 80K-5W high capacity at CEA/SBT, 131 80K-5WGiessen/Leybold, 149 Astrium miniature space cooler, 173 Cryomech4K and l0K, 293 Cryomech 4K precooler for sub-Kelvin, 669 Fuji Electric 70K commercial cooler, 157 LM-ATC 6K 4-stage ACTDP, 1 LM-ATC 10K PT for space, 241 LM-ATC 35K two-stage cooler, 213
824 LM-ATC 35K high-capacity cooler, 225 LM-ATC 80K 2.7W for INTEGRAL, 205 Raytheon low-cost space cooler, 183 Raytheon hybrid Stirling/PT for 35K, 233 Sunpower single and 2-stage, 139 Thales 80K PT w/ flexure bearings, 109 TRW ACTDP 17K J-T precooler, 1 TRW AIRS 55K, 1, 747 TRW HCC 35K/85K two-stage, 219 TRWHEC, 191, 199, 275 TRWJAMI, 191 TRW 58K TES, 1 Pulse tube theory and investigations: 1. 5K 3He Eindhoven Univ., 283 4K VM-type pulse tube cooler, 331 effect of orientation on performance, 139 GM-type coaxial PT experiments, 317 He/H2 mixed gas, use of at 20K, 325 LMA pulse tube flight experiment, 777 mass flow control, effect of, 337 model of hybrid PT/rev-Brayton cooler, 349 model using method of lines, 379 nonmetallic, nonmagnetic for SQUIDs, 799 off-state conduction vs. angle, 747 piston resonance in orifice PT, 265 pressure wave generator for PT, 343 regenerator loss at low temp and high freq, 523 role of orifice and secondary bypass, 361 surface heat pumping loss, 371 thermoacoustic PT with mixed gases, 451 Univ. of Calgary PT analysis, 389 Pusan National Univ. (Korea), 95 RAL (see Rutherford Appleton Laboratory) Rare earth compounds (see Regenerators) Raytheon (formerly Hughes Aircraft): low-cost, lightweight space coolers, 183 comparison of regenerator models, 547 PT model using method of lines, 379 RSP2 hybrid Stirling/PT for 35K, 233 Raytheon Infrared Operations: tactical cryocooler development, 69 Recuperators: for turbo-Brayton coolers, 571 parallel plate heat exchangers, 507 Redstone Engineering hybrid 10K J-T, 579 Refrigerants: 3 He for use below 4K, 283 He/H2 mixed gas, use in 20K PT, 325 noble gas mixture for thermoacoustics, 451 superfluid He, 675, 681 Regenerators: bonded regenerator matrix mat'ls, 517 comparison of numerical models, 547 dc flow caused by nonlinearities, 425 degradation in GM MRI application, 805 etched foil, improved flow in, 499 GAP material in 4K PT, 309 GdSb, properties of, 467 GOS ceramic mat'l for 4K coolers, 397, 403, 473
SUBJECT INDEX LIGA-fabricated microchannel, 489 loss meas. at low temp and high freq., 523 modeling oscillating flow and pressure, 531 modeling periodic flow in porous media, 555 nonmetallic, nonmagnetic for SQUIDs, 799 oscillating regenerator flow at low temp, 539 overview of low temperature mat'ls, 457 perforated plates, predictions for, 483
Reliability of cryocoolers: CMC tactical cooler lifetest results, 79 degradation of low-e surfaces in space, 727 GM in medical MRI application, 805 MTTF predictions on Thales coolers, 87 STI Stirling cooler reliability data, 75 Reverse-Brayton cryocoolers: 6K Creare cooler developments, 571 Creare ACTDP 6K cooler study, 1 NICMOS operation on HST, 1, 563 RHESSI in NASA cooler missions overview, 1 in-space performance, 755 Ricor vibration protective mount, 717 Rutherford Appleton Laboratory, 173 SADA linear tactical cooler overview, 17 Santa Clara University, 499 Sorption cryocoolers: microcooler, Univ. of Twente, 643 Planck hydride compressor elements, 627 Planck EBB breadboard cooler testing, 637 sub-Kelvin for Polatron, 651 sub-Kelvin on 4K Cryomech PT cooler, 669 Space experiments: Lockheed Martin Astronautics PT cooler, 777 Swales thermal integration technologies, 693 Space instrument missions: AIRS, 1, 747 ASTER, 737 HIRDLS, 1 INTEGRAL, 771, 205 ISAMS, 727 JAMI, 191 MOPITT, 727 NASA mission summary, 1 NICMOS, 1, 563 Planck, 1, 627, 637 RHESSI, 1, 755 TES, 1 SQUIDs (see Integration of cryocoolers with:) Stirling cryocoolers: Air Liquide high capacity for ISS, 31 ASTER, space performance of, 737 Astrium 10K cooler testing, 59 Astrium/MMS 50-80K MOPITT, 727 Ball multistage cooler development, 45 Ball 35K/60K PSC testing at AFRL, 51 CMC tactical cooler lifetest results, 79 linear tactical for weapon systems, 17 Oxford l00mW 80K low power, 27 Oxford/BAe 80K ISAMS, 727
SUBJECT INDEX Raytheon RS1 for Space, 183 Raytheon RSP2 Stirling/PT for 35K, 233 Raytheon Infrared Operations tactical, 69 SADA linear tactical cooler overview, 17 Sunpower M87N, qual. for AMS-02, 37, 761 Sunpower space perf. on RHESSI, 755 Superconductor Tech., Inc., reliability, 75 Thales tactical coolers, 87 Stirling cryocooler theory and investigations: dynamic analysis of free-piston linear, 103 experimental study of phase shift in linear, 95 MTTF prediction for tactical coolers, 87 operation in high magnetic fields, 761 superfluid He as working fluid, 681 Sub-Kelvin coolers: continuous ADR, 661 dilution refrigerator on 4K PT, 669 sorption cooler for Polatron, 651 sorption precooled on 4K PT, 669 Sumitomo Heavy Industries: 4K two-stage pulse tube cooler, 301 GM regenerator degradation analysis, 805 GM improvement at 4K using GOS, 397, 403 GOS ceramic regen mat'l for 4K coolers, 473 Sunpower coolers: operation in high magnetic fields, 761 qualification of M87N for AMS-02, 37 RHESSI on-orbit performance, 755 single and 2-stage pulse tube coolers, 139 Superconductor applications (see Integration of cryocoolers with) Superconductor Technologies Inc. (STI), 75 Superfluid helium as working fluid: in Joule-Thomson refrigerator, 675 in Stirling refrigerator, 681 Swales Aerospace: components for cryo integration, 693 Planck cryocooler development, 627
TES instrument cryocoolers, 1 Texas Instruments (see DRS cryocoolers) Thales Cryogenics: 50-80K PT cooler for space, 165 high capacity 5W-80KPT cooler, 131 high capacity Stirling cooler for ISS, 31 low vibration 80K PT w/ flexure bearings, 109 MTTF of Thales tactical coolers, 87 Thermacore cryogenic loop heat pipes, 709 Thermal conductivity (see Conductance) Thermal storage: helium/charcoal TSU for 15K, 579 Swales development and testing of, 693 Thermal switch (see Heat switch) Thermoacoustic generator: analysis using distributed-parameters, 431 dc flow, caused by natural convection, 447
825 dc flow, caused by regen nonlinearities, 425 driven PT with mixed noble gases, 451 high freq. operation, investigation of, 439 onset temperature gradient, study of, 421 TRW (now Northrop Grumman Space Tech.): ACTDP 6K cooler development, 1 AIRS 55K pulse tube cooler, 1 HCC 35K/85K two-stage PT cooler, 219 HEC pulse tube cooler, 191, 199, 275 JAMI pulse tube cooler, 191 producibility of linear compressors, 275 scaling of linear compressors, 247 TES 58K pulse tube cooler, 1 Tsukuba Magnet Laboratory: GM improvement at 4K using GOS, 397, 403 GOS ceramic material for 4K coolers, 473 regenerator properties of GdSb, 467 Turbo Brayton coolers (see reverse Brayton coolers) Univ. of Calgary, 389 Univ. of Dresden, 595 Univ. of Giessen, 149 Univ. of Oxford: linearity of flexure suspension systems, 255 low-power Stirling cooler, 27 producibility of linear compressors, 275 scaling of linear compressors, 247 Univ. of Southern Calif., PT modeling, 379 Univ. of Twente: sorption microcooler, 643 SQUID-based fetal heart monitor, 789 Univ. of Wisconsin: model for hybrid PT/rev Brayton, 349 regenerator loss measurements, 523 US Army Night Vision, 17 Vibration: Astrium cancellation system, 771 suppression with gimbaled instrument, 717 Thales vibration control algorithm, 109 sensitivity of Polatron instrument, 651 VM-type pulse tube cooler, 331 Vortex tube: applied to auto-cascade J-T, 621 modeled as heat exchanger, 615 Zero-boil-off cryogen storage, 199 Zhejiang Univ.: capillary tube expanders for mixed gases, 609 PT with He/H2 mixed gas at 20K, 325 thermoacoustic device analysis, 431 thermoacoustic driven PT cooler, 451 thermoacoustics, onset temp gradient, 421 vortex tube applied to auto-cascade J-T, 621 vortex tube as heat exchanger, 615
