Cryocoolers 11

CRYOCOOLERS 11

A publication of the International Cryocooler Conference

CRYOCOOLERS 11

Edited by

R. G. Ross, Jr. Jet Propulsion Laboratory California Institute of Technology Pasadena, California

KLUWER ACADEMIC PUBLISHERS NEW YORK, BOSTON, DORDRECHT, LONDON, MOSCOW

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Preface Over the last two years we 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; they 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, spearheaded in the former USSR, to achieve low-cost cryocooler systems in the 65 - 80 K temperature range. Tactical Stirling cryocoolers, the mainstay of the defense industry, continue to find application in cost-constrained commercial applications and space missions; 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 11 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 11th International Cryocooler Conference, which was held in Keystone, Colorado, in June 2000. The program of this conference consisted of 127 papers; of these, 98 are published here. Although this is the eleventh 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 fourth 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 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.

v

vi

PREFACE

The content of Cryocoolers 11 is organized into 19 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, and associated regenerator research. Next, Turbo-Brayton, JouleThomson, and sorption cryocoolers, as well as sub-Kelvin refrigerators are covered in a progression of lowering temperatures. The technology-specific chapters end with a chapter on Optical Refrigeration; this provides a glimpse into the future with miniature solid-state refrigerators using advanced optical-based refrigeration cycles. The last four chapters deal with cryocooler reliability investigations, integration technologies, and experience to date in a number of representative space and commercial applications. The articles in these last four chapters contain a wealth of information for the potential user of cryocoolers, as well as for the developer. The expanding availability of low-cost, reliable cryocoolers is making major advances in a number of fields. It is hoped that this book will serve as a valuable reference to all those faced

with the challenges of developing and using cryocoolers.

Ronald G. Ross, Jr. Jet Propulsion Laboratory

California Institute of Technology

Acknowledgments The International Cryocooler Conference Board wishes to thank Ball Aerospace & Technologies Corp., which hosted the 11th ICC, and to express its deepest appreciation to the Conference Organizing Committee, whose members dedicated many hours to organizing and managing the conduct of the Conference. Members of the Organizing Committee and Board for the 11th ICC include:

CONFERENCE CO-CHAIRS Rodney Oonk, Ball Aerospace Richard Reinker, Ball Aerospace CONFERENCE ADMINISTRATORS Margueritte Sommers, Ball Aerospace Dianne Fisher, Ball Aerospace PROGRAM CHAIRMAN Klaus Timmerhaus, Univ. of Colorado CONFERENCE SECRETARY Jill Bruning, Nichols Research Corp. PUBLICATIONS Ron Ross, Jet Propulsion Laboratory TREASURER Ray Radebaugh, NIST

PROGRAM COMMITTEE John Brisson, MIT William Burt, TRW Stephen castles, NASA/GSFC Peter Kerney, Leybold Eric Marquardt, NIST Lawrence wade, JPL ADVISORY BOARD Guobang Chen, zhejiang Univ., china Thom Davis, AFRL Dave Glaister, Ball Aerospace Geoff Green, MAPC Tom Kawecki, NRL Peter Kittel, NASA/ARC Ralph Longsworth, APD Cryogenics Yoichi Matsubara, Nihon Univ., Japan Ted Nast, Lockheed Martin ATC Martin Nisenoff, NRL Walter Swift, 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 C. Hall, P. Irwin, J.M. Lee, R. Mestas, B. Oonk, A. Ravex, B. Reinker, C. Stoyanof, M. Stoyanof, and J. Timmerhaus.

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Contents 1

Government Cryocooler Development and Test Programs Military Space Cryogenic Cooling Requirements for the 21st Century ......

1

T.M. Davis and B.J. Tomlinson, Kirtland AFB, NM; and J.D. Ledbetter, Mission Research Corp., Albuquerque, NM

Status of Programs for the DoD Family of Linear Drive Cryogenic Coolers for Weapon Systems .................... .................. 11 W.E. Salazar, US Army Night Vision, Fort Belvoir, VA

Air Force Research Laboratory Cryocooler Characterization and Endurance Update .............................................. 17 B.J. Tomlinson, C.H. Yoneshige, AFRL, Kirtland AFB, NM; and N.S. Abhyankar, Dynacs Engin., Albuquerque, NM

Air Force Research Laboratory Cryocooler Reliability Initiatives ........... 27 S. Blankenship and T.L. Fountain, Georgia Inst. of Tech., Atlanta, GA; and T.M. Davis and B.J. Tomlinson, AFRL, Kirtland AFB, NM

Space Stirling Cryocooler Developments

35

Protoflight Spacecraft Cryocooler Performance Results ................... 35 K. Price, Raytheon, El Segundo, CA; and J. Reilly, N. Abhyankar, and B.J. Tomlinson, AFRL, Kirtland AFB, NM

Characterization of Raytheon’s 60 K 2 W Protoflight Spacecraft Cryocooler .................................................... 45 N.S. Abhyankar, Dynacs Engin., Albuquerque, NM; and C.H. Yoneshige, B.J. Tomlinson, and J. Reilly, AFRL, Kirtland AFB, NM

The Development of a 10 K Closed Cycle Stirling Cooler for Space Use ...... 55 G. Baker, D. Féger, and A. Little, Astrium, Stevenage, UK; A.H. Orlowska, T. Bradshaw, and M. Crook, RAL, Chilian, UK; B.J. Tomlinson, AFRL, Kirtland AFB, NM; and A. Sargeant, Cubic Appl. Inc., Lacey, WA

Development of a 12 K Stirling Cycle Precooler for a 6 K Hybrid Cooler System ................................................. 63 W.J. Gully, D.S. Glaister, and D.W. Simmons, Ball Aerospace, Boulder, CO

Thermodynamic Optimization of Multi-Stage Cryocoolers ................ 69 C.S. Kirkconnell and K.D. Price, Raytheon, El Segundo, CA

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CONTENTS

Long-Life Tactical and Commercial Stirling Coolers

79

The Advent of Low-Cost Cryocoolers ................................. 79 R.Z. Unger, R.B. Wiseman, and M.R. Hummon, Sunpower, Inc., Athens, OH

Performance and Reliability Improvements in a Low-Cost Stirling Cycle Cryocooler ............................................... 87 M. Hanes, Superconductor Tech. Inc., Santa Barbara, CA

Development of a Long-Life Stirling Cryocooler ........................ 97 Y. Ikuta, Y. Suzuki, K. Kanao, and N. Watanabe, Sumitomo Heavy Indus., Hiratsuka, Kanagawa, Japan

Flexure Springs Applied to Low-Cost Linear Drive Cryocoolers ............ 103 R.M. Rawlings and S. Miskimins, DRS Infrared Tech., Dallas, TX

High Reliability Coolers under Development at Signaal-USFA ............. 111 M. Meijers, A.A.J. Benschop, and J.C. Mullié, Signaal-USFA, Eindhoven, The Netherlands

Long-Life Commercial Pulse Tube Coolers

119

Development of a Long-Life Stirling Pulse Tube Cryocooler for Superconducting Filter Subsystems ................................. 119 Y. Hiratsuka, Daikin Indus., Osaka, Japan; K. Murayama, Y. Maeda, F. Imai, and K.Y. Kang, Daikin Envir. Lab, Tsukuba, Japan; and Y. Matsubara, Nihon Univ., Funabashi, Japan

Development of a 5 W at 65 K Air-Cooled Pulse Tube Cryocooler ........... 125 S-Y Kim, J-J Park, S-T Kim, W-S Chung, and H-K Lee, LG Electronics Inc., Seoul, Korea

Space Pulse Tube Cryocooler Developments

131

TES FPC Flight Pulse Tube Cooler System ............................ 131 J. Raab, S. Abedzadeh, R. Colbert, J. Godden, D. Harvey, and C. Jaco, TRW, Redondo Beach, CA

The AIM Space Cryocooler Program ................................. 139 I. Rühlich, H. Korf, and Th. Wiedmann, AEG Infrarot-Module, Heilbronn, Germany

Miniature Pulse Tube Cryocooler for Space Applications ................. 145 T.C. Nast, P.J. Champagne, V. Kotsubo, J. Olson, A. Collaco, and B. Evtimov, Lockheed Martin ATC, Palo Alto, CA; T. Renna, Lockheed Martin Communications, Newtown, PA; and R. Clappier, Clappier Consulting, Discovery Bay, CA

Gamma-Ray Pulse Tube Cooler Development and Testing ................ 155 R.G. Ross, Jr., D.L. Johnson, and A. Metzger, JPL, Pasadena, CA; V. Kotsubo, B. Evtimov, J. Olson, and T. Nast, Lockheed-Martin ATC, Palo Alto, CA; and

R.M. Rawlings, DRS Infrared Tech., Dallas, TX

High Efficiency Pulse Tube Cooler ................................... 163 E. Tward, C.K. Chan, J. Raab, T. Nguyen, and R. Colbert, TRW, Redondo Beach, CA; and T. Davis, AFRL, Kirtland AFB, NM

CONTENTS

xi

High Performance Flight Cryocooler Compressor ....................... 169 P.B. Bailey and M.W. Dadd, Oxford Univ., Oxford, UK; N. Hill and C.F. Cheuk,

Hymatic Engin. Co., Redditch, UK; and J. Raab and E. Tward, TRW, Redondo Beach, CA

Vibration Reduction in Balanced Linear Compressors .................... 175 M.W. Dadd., P.B. Bailey, and G. Davey, Oxford Univ., Oxford, UK; and T. Davis and B.J. Tomlinson, AFRL, Kirtland AFB, NM

95K High Efficiency Cryocooler Program .............................183 K. Price, Raytheon, El Segundo, CA; and V. Urbancek, AFRL, Kirtland AFB, NM

Design and Test of the NIST/Lockheed Martin Miniature Pulse Tube Flight Cryocooler .............................................. 189 P.E. Bradley and R. Radebaugh, NIST, Boulder, CO; J.H. Xiao, Johnson and Johnson, Somerville, NJ; and D.R. Ladner, Lockheed Martin Astronautics, Denver, CO

Low-Cost Pulse Tube Liquefier for In-Situ Resource Utilization ............199 C.M. Martin and J.L. Martin, Mesoscopic Devices, Broomfield, CO

GM-Type Pulse Tube Coolers for Low Temperatures

205

Performance Characteristics of a 4 K Pulse Tube in Current Applications ..... 205 C. Wang and P.E. Gifford, Cryomech, Inc., Syracuse, NY

Experimental Study of a 4K Pulse Tube Cryocooler ..................... 213 S. Fujimoto, T. Kurihara, T. Oodo, Y.M. Kang, Daikin Ltd., Tsukuba, Japan; T. Numazawa, Nat. Res. Inst. for Metals, Tsukuba, Japan; and Y. Matsubara,

Nihon Univ., Chiba, Japan

GM-Type Two-Stage Pulse Tube Cooler with High Efficiency ............. 221 A. Hofmann, Karlsruhe Inst. for Tech. Physics, Karlsruhe, Germany; H. Pan and L. Oellrich, Univ. of Karlsruhe, Karlsruhe, Germany

Developments on Single and Double Stage GM Type Pulse Tube Cryorefrigerators . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 229 J.M Poncet, A. Ravex, and I. Charles, CEA/DRFMC/Service des Basses Temperatures, Grenoble, France

30 - 50 K Single Stage Pulse Tube Refrigerator for HTS Applications ........ 235 J. Yuan, J. Maguire, A. Sidi-Yekhlef, and P. Winn, American Superconductor Co., Westborough, MA

Two-Stage 4K Pulse Tube Refrigerator . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 243 S. Zhu, M. Ichikawa, M. Nogawa, and T. Inoue, Aisin Seiki Co., Ltd., Kariya, Aichi, Japan

Compressor-Specific Design of a Single Stage Pulse Tube Refrigerator ....... 249 J.M. Pfotenhauer and J.H. Baik, Univ. of Wisconsin, Madison, WI

Hybrid Cryocoolers Using Pulse Tubes

259

A Novel Multi-Stage Expander Concept ............................... 259 C.S. Kirkconnell, K.D. Price, M.C. Barr, and J.T. Russo, Raytheon, El Segundo, CA

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CONTENTS

Numerical Study of a New Type of 4K GM/PT Hybrid Refrigerator . . . . . . 265 L. Liu, L. Gong, J. Liang, and L. Zhang, Cryogenic Lab, Chinese Acad. of Sci.,

Beijing, China

Thermally Actuated 3He Pulse Tube Cooler .......................... 273 Y. Matsubara, H. Kobayashi, and S.L. Zhou, Atomic Energy Res. Inst., Nihon Univ., Chiba, Japan

Investigation of Helium and Nitrogen Mixtures in a Pulse Tube Refrigerator ................................................... 281 Z.H. Gan and G.B. Chen, Zhejiang Univ., Hangzhou, China; and G. Thummes and C. Heiden, Univ. of Giessen, Giessen, Germany

Pulse Tube Refrigeration with a Combined Cooling and Freezing Cycle for HTSC Devices . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 291 G. Chen, Z. Gan, L. Qiu and J. Yu, Zhejiang Univ., Hangzhou, China

Experimental Investigation of a Pulse Tube Refrigerator Driven by a Thermoacoustic Prime Mover ................................. 301 L.M. Qiu, G.B. Chen, N. Jiang, Y.L. Jiang, and J.P. Yu, Zhejiang Univ., Hangzhou, China

Design, Development, and Operation of a Thermo-Acoustic Refrigerator Cooling to below -60°C .............................. 309 M.E.H. Tijani, J. Zeegers, and A.T.A.M. de Waele, Eindhoven Univ. of Tech., Eindhoven, The Netherlands

Pulse Tube Analysis and Experimental Measurements

317

Design of a Miniature Pulse Tube Refrigerator . . . . . . . . . . . . . . . . . . . . . . . . . 317 A. Halouane, French Inst. of Petroleum, Rueil-Malmaison, France; and

J-C. Marechal and Y. Simon, Ecole Normale Supérieure, Paris, France

Investigation of a Single Stage Four-Valve Pulse Tube Refrigerator for High Cooling Power ......................................... 327 T. Schmauder A. Waldauf, M. Thürk, R. Wagner, and P. Seidel, Univ. of Jena, Jena, Germany

Analysis and Experimental Research of a Multi-Bypass Version Pulse Tube Refrigerator ........................................ 337 L.W. Yang, J.T. Liang and Y. Zhou, Chinese Academy of Sciences, Beijing, China

Experimental Study of the Heat Transfer in Pulse Tubes ................. 345 S. Jeong, K. Nam and M.G. Kim, Korea Adv. Inst. of Sci. and Tech., Taejon, Korea; and H.-M. Chang and E.S. Jeong, Hong Ik Univ., Seoul, Korea

Shuttle Loss in Pulse Tubes . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 353 L. W. Yang, Chinese Academy of Sciences, Beijing, China

Numerical Study of Gas Dynamics Inside of a Pulse Tube Refrigerator . . . . . . 363 Y. Hozumi, Chiyoda Corp., Yokohama, Japan; M. Murakami, Univ. of Tsukuba, Tsukuba, Japan; and M. Shiraishi, ME Lab, MITI, Tsukuba, Japan

CONTENTS

xiii

Visualization of DC Gas Flows in a Double-Inlet Pulse Tube Refrigerator with a Second Orifice Valve ............................ 371 M. Shiraishi and A. Nakano, ME Lab, MITI, Tsukuba, Japan; K. Takamatsu and M. Murakami, Univ. of Tsukuba, Tsukuba, Japan; T. Iida, NASDA, Tsukuba, Japan; and Y. Hozumi, Chiyoda Corp., Yokohama, Japan

GM Refrigerator Developments

381

A Gifford-McMahon Cycle Cryocooler below 2K ..................... 381 T. Satoh, Sumitomo Heavy Ind., Kanagawa, Japan; A. Onishi, Sumitomo Heavy Ind., Tokyo, Japan; I. Umehara, Y. Adachi, and K. Sato, Yokohama Nat ’l Univ., Yokohama, Japan; and E.J. Minehara, FEL Lab, Japan Atomic Energy Res. Inst., Naka, Japan

High Efficiency, Single-Stage GM Cryorefrigerators Optimized for 20 to 40K . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 387 C. Wang and P.E. Gifford, Cryomech, Inc., Syracuse, NY

Remote Cooling with a G-M Cryocooler by Use of Cold Electromagnetic Valves Driving an External Flow Loop ..................... 393 K.M. Ceridon and J.L. Smith, Jr., MIT, Cambridge, MA

Optimum Intermediate Temperatures of Two-Stage Gifford-McMahon

Type Coolers .................................................. 401 T.C. Chuang, Raytheon-RCSI, Philadelphia, PA; S. Yoshida, Taiyo Toyo Sanso, Co., Kawasaki, Japan; and T.H.K. Frederking, UCLA, Los Angeles, CA

Regenerator Analysis and Materials Developments

409

Regenerator Behavior with Heat Input or Removal at Intermediate Temperatures . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 409 R. Radebaugh, E.D. Marquardt, J. Gary, and A. O’Gallagher, NIST, Boulder, CO

Measurement of Heat Conduction through Metal Spheres ................ 419 M.A. Lewis and R. Radebaugh, NIST, Boulder, CO

Innovative Technology for Low Temperature Regenerators ............... 427 L. Tuchinskiy and R. Loutfy, MER Corp., Tucson, AZ; and B.J. Tomlinson, AFRL, Kirtland AFB, NM

Ductile, High Heat Capacity, Magnetic Regenerator Alloys for the 10 to 80 K Temperature Range .................................... 433 K.A. Gschneidner, Jr., A.O. Pecharsky, and V.K. Pecharsky, Ames Lab, Iowa State Univ., Ames, IA

Low Temperature Properties of HoSb, DySb, and GdSb .................. 443 H. Nakane and S. Yamazaki, Kogakuin Univ., Tokyo, Japan; H. Fujishiro, Iwate Univ., Morioka, Japan; T. Yamaguchi and S. Yoshizawa, Meisei Univ., Tokyo, Japan; T. Numazawa, Nat. Res. Inst. for Metal, Tsukuba, Japan; and M. Okamura, Toshiba Corp., Yokohama, Japan

Manufacturing Considerations for Rare Earth Powders Used in Cryocooler and Magnetic Refrigerator Applications .................... 449 S.A. Miller, J.D. Nicholson, Starmet Corp., Concord, MA; and K.A. Gschneidner, Jr., A.O. Pecharsky, and V.K. Pecharsky, Ames Laboratory, Iowa State Univ., Ames, IA

Magnetothermal Properties of Polycrystalline Gd2In . . . . . . . . . . . . . . . . . . 457 M.I. Ilyn and A.M. Tishin, Moscow State Univ., Moscow Russia; K.A. Gschneidner, Jr., V.K. Pecharsky, and A.O. Pecharsky, Ames Labs, Iowa State Univ., Ames, IA

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CONTENTS

New Regenerator Material for Sub-4 K Cryocoolers ...................... 465 T. Numazawa, O. Arai and A. Sato, Tsukuba Magnet Lab, Nat. Res. Inst. for Metals, Tsukuba, Japan; S. Fujimoto, T. Oodo, and Y.M. Rang, Daikin, Ltd., Tsukuba, Japan; and T. Yanagitani, Konoshima Chemical Co., Kagawa, Japan

New Regenerator Materials for Use in Pulse Tube Coolers ................ 475 A. Kashani and B.P.M. Helvensteijn, Atlas Scientific, Sunnyvale, CA, P. Kittel, NASA/ARC, Moffett Field, CA; and K.A. Gschneidner, Jr., V.K. Pecharsky, and A.O. Pecharsky, Ames Labs, Iowa State Univ., Ames, IA

Turbo-Brayton Cryocooler Developments

481

Advanced Developments for Low Temperature Turbo-Brayton Cryocoolers . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 481 J.A. McCormick, G.F. Nellis, H. Sixsmith, M.V. Zagarola, M.G. Izenson and W.L. Swift, Creare, Hanover, NH; and JA. Gibbon, NASA/GSFC, Greenbelt, MD

Life and Reliability Characteristics of Turbo-Brayton Coolers .............. 489 J.J. Breedlove, M. V. Zagarola, G.F. Nellis, F.X. Dolan, and W.L. Swift, Creare, Hanover, NH; and J.A. Gibbon, NASA/GSFC, Greenbelt, MD

A Flexible Turbo-Brayton Cryocooler Model . . . . . . . . . . . . . . . . . . . . . . . . . . . 499 P.L. Whitehouse, NASA/GSFC, Greenbelt, MD; and G.F. Nellis and M.V. Zagarola, Creare, Inc., Hanover, NH

J-T and Throttle-Cycle Cryocooler Developments

505

A 10 K Cryocooler for Space Applications ............................. 505 D.S. Glaister, W.J. Gully, G.P. Wright and D.W. Simmons, Ball Aerospace, Boulder, CO; and B.J. Tomlinson, AFRL, Kirtland AFB, NM

Modern Trends in Designing Small-ScaleThrottle-Cycle Coolers

Operating with Mixed Refrigerants . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 513 M. Boiarski and A. Khatri, IGC-APD Cryogenics, Allentown, PA; O. Podcherniaev, IGC-Polycold Sys., San Rafael, CA; and V. Kovalenko, Moscow Power Engin. Inst., Moscow, Russia

Thermodynamic Analysis of an Mixed-Refrigerant Auto-Cascade J-T Cryocooler with Distributed Heat Load .......................... 523 M.Q. Gong, E.C. Luo, J.T. Liang, Y. Zhou, and J.F. Wu, Chinese Academy of Sciences, Beijing,

China

Sorption Cryocooler Developments

531

PLANCK Sorption Cooler Initial Compressor Element

Performance Tests . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 531 C.G. Paine, R.C. Bowman Jr., D. Pearson, M.E. Schmelzel, P. Bhandari, and L.A. Wade, JPL, Pasadena, CA

Sizing and Dynamic Performance Prediction Tools for 20 K Hydrogen Sorption Cryocoolers . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 541 P.Bhandari, M. Prina, R.C. Bowman and L.A. Wade, JPL, Pasadena, CA; and M. Ahart, Princeton Univ., Princeton, NJ

CONTENTS

xv

165 K Microcooler Operating with a Sorption Compressor and a Micromachined Cold Stage ....................................... 551 J.F. Burger, H.J. Holland, J.H. Seppenwoolde, J.W. Berenschot, H.J.M. ter Brake, J.G.E. Gardeniers, M. Elwenspoek and H. Rogalla, Univ. of Twente, Enschede, The Netherlands

Sub-Kelvin Refrigerator Developments

561

Double Stage Helium Sorption Coolers ............................... 561 L. Duband, CEA/DRFMC/Service des Basses Températures, Grenoble, France

Sub-Kelvin Sorption Coolers for Space Application ...................... 567 L. Duband, CEA/DRFMC, Grenoble, France; B. Collaudin, ESTEC, Noordwijk, The Netherlands; and P. Jamotton, Centre Spatial de Liège, Belgium

Closed-Cycle Cooling of Infrared Detectors to 0.25 K for the Polatron ....... 577 R.S. Bhatia, J.J. Bock, V.V. Hristov, W.C. Jones, A.E. Lange, J. Leong, P.V. Mason, B.J. Philhour and G. Sirbi, Caltech, Pasadena, CA; S.E. Church and B.G. Keating, Stanford Univ., Standford, CA; J.G. Glenn, Univ. of Colorado, Boulder, CO; S.T. Chase, Chase Research Ltd., Sheffield, UK; and P.A.R. Ade and C.V. Haynes, QMW College, London, UK

A Continuous Adiabatic Demagnetization Refrigerator for Use with Mechanical Coolers . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 587 P. Shirron, E. Canavan, M. DiPirro, M. Jackson, J. Panek and J. Tuttle, NASA/GSFC, Greenbelt, MD; and N. Abbondante, M. Grabowski and M. Hirsch, Worcester Polytechnic

Institute, Worcester, MA

Reaching 96 mK by a Pulse-Tube Precooled Adiabatic Demagnetization Refrigerator ................................................... 597 G. Thummes and M. Theiß, Inst. of Applied Physics, Giessen, Germany; and M. Bühler and J. Höhne, CSP GmbH, Ismaning, Germany

Dissipation in Metal Welded Bellows and Its Consequences for Sub-Kelvin Refrigerators . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 605 C.L. Phillips and J.G. Brisson, MIT, Cambridge, MA

Optical Refrigeration Developments

613

Design and Predicted Performance of an Optical Cryocooler for a Focal Plane Application .......................................... 613 G.L. Mills, A.J. Mord and P.A. Slaymaker, Ball Aerospace, Boulder, CO

Optical Refrigeration Using Anti-Stokes Fluorescence from

Molecular Dyes ................................................ 621 G. Rumbles, B. Heeg, and J.L. Lloyd (née Clark), Imperial College, London, UK; P.A. DeBarber, MetroLaser, Inc., Irvine, CA; and B.J. Tomlinson, AFRL, Kirtland AFB, NM

Solid-State Optical Cooler Developments .............................. 631 B.C. Edwards, J.E. Anderson, and R.I. Epstein, Los Alamos Nat’l Lab, Los Alamos, NM; and C. W. Hoyt and M. Sheik-Bahae, Univ. of New Mexico, Albuquerque, NM

Cryocooler Reliability Investigations and Analyses

637

Cryocooler Reliability and Redundancy Considerations for Long-Life Space Missions ................................................. 637 R.G. Ross, Jr., JPL, Pasadena, CA

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CONTENTS

Space Cryocooler Contamination Lessons Learned and Recommended Control Procedures ............................................. 649 S. Castles, NASA/GSFC, Greenbelt, MD; K.D. Price, Raytheon, El Segundo, CA; D.S. Glaister and W.J. Gully, Ball Aerospace, Boulder, CO; J. Reilly, AFRL, Kirtland AFB, NM; and T. Nast and V. Kotsubo, Lockheed-Martin, Palo Alto, CA

Cryocooler Contamination Study: Temperature Dependence of Outgassing ................................................. 659 S.W.K. Yuan and D.T. Kuo, BAE Systems, Sylmar, CA

BAE’s Life Test Results on Various Linear Coolers and Their Correlation with a First Order Life Estimation Method .......................... 665 D.T. Kuo, T.D. Lody and S.W.K. Yuan, BAE Systems, Sylmar, CA

Initial Observations from the Disassembly and Inspection of the TRW 3503 and Creare SSRB ...................................... 673 B.J. Tomlinson and C.H. Yoneshige, AFRL, Kirtland AFB, NM; and M.L. Martin, Dynacs Engin., Albuquerque, NM

Cryocooler Integration Technologies and Materials

681

Cryogenic Material Properties Database ............................... 681 E.D. Marquardt, J.P. Le, and R. Radebaugh, NIST, Boulder, CO

Experimental Results on the Thermal Contact Conductance of

Ag-Filled Epoxied Junctions at Cryogenic Temperatures ................ 689 Z. Wang, A. Devpura, and P.E. Phelan, Arizona State Univ., Tempe, AZ

A Fail-Safe Experiment Stand for Cryocooler Characterization ............. 699 C.H. Yoneshige, J.P. Kallman, G. Lybarger, AFRL, Kirtland AFB, NM; and N.S. Abhyankar and M.L. Martin, Dynacs Engin., Albuquerque, NM

Development and Testing of a Gimbal Thermal Transport System .......... 707 D. Bugby, B. Marland, and C. Stouffer, Swales Aerospace, Beltsville, MD; and B. Tomlinson and T. Davis, AFRL, Kirtland AFB, NM

Cryocooler Interface System ........................................ 719 G.S. Willen, Tech. Applications, Inc., Boulder, CO; and B.J. Thomlinson, AFRL,

Kirtland AFB, NM

Development and Testing of a High Performance Cryogenic Thermal Switch . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 729 B. Marland, D. Bugby, and C. Stouffer, Swales Aerospace, Beltsville, MD; and B. Tomlinson and T. Davis, AFRL, Kirtland AFB, NM

Thermally Conductive Vibration Isolation System for Cryocoolers .......... 739 G.S. Willen, Tech. Appl., Inc., Boulder, CO; and E.M. Flint, CSA Engin., Mountain View, CA

Advanced Cryogenic Integration and Cooling Technology for Space-Based Long Term Cryogen Storage . . . . . . . . . . . . . . . . . . . . . . . . . . . . 749 B.J. Tomlinson and T.M. Davis, AFRL, Kirtland AFB, NM; and J.D. Ledbetter, Mission

Research Corp., Albuquerque, NM

CONTENTS

Space Cryocooler Applications

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759

MOPITT On-Orbit Stirling Cycle Cooler Performance . . . . . . . . . . . . . . . . . . . 759 G.S. Mand and J.R. Drummond, Univ. of Toronto, Toronto, Canada; and D. Henry and J. Hackett, COM DEV Inter., Cambridge, Ontario, Canada

HIRDLS Instrument Flight Cryocooler Subsystem Integration and Acceptance Testing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 769 W.C. Kiehl, D.J. Berry, D.S. Glaister, J. Richards, and R.G. Stack, Ball Aerospace, Boulder, CO

Low-Temperature, Low-Vibration Cryocooler for Next Generation Space Telescope Instruments . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 775 R.L. Oonk, D.S. Glaister, W.J. Gully and M.D. Lieber, Ball Aerospace, Boulder, CO

Commercial Cryocooler Applications

783

Considerations in Applying Open Cycle J-T Cryostats to Cryosurgery ....... 783 R.C. Longsworth, 1GC-APD Cryogenics, Allentown, PA

Interference Characterization of Cryocoolers for a High-Tc SQUIDBased Fetal Heart Monitor . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 793 A.P. Rijpma, M.R. Bangma, H.A. Reincke, E. de Vries, H.J. Holland, H.J.M. ter Brake

and H. Rogalla, Univ. of Twente, Enschede, The Netherlands

Vapor Precooling in a Pulse Tube Liquefier . . . . . . . . . . . . . . . . . . . . . . . . . . . . 803 E.D. Marquardt, R. Radebaugh, and A.P. Peskin, NIST, Boulder, CO

Terrestrial Applications of Zero-Boil-Off Cryogen Storage . . . . . . . . . . . . . . . . 809 L.J. Salerno and P. Kittel, NASA/ARC, Moffett Field, CA; J. Gaby, NASA/GRC, Cleveland, OH; R. Johnson, NASA/KSC, FL; and E.D. Marquardt, NIST, Boulder, CO

Indexes

817

Proceedings Index . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 817 Author Index . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 819 Subject Index . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 821

Military Space Cryogenic Cooling Requirements for the 21st Century Thom Davis1, B. J. Tomlinson1, and Jim Ledbetter2 1

Space Vehicles Directorate, Air Force Research Laboratory Kirtland AFB, NM 87117-5776 2 Mission Research Corporation Albuquerque, NM, USA 87106-4266

ABSTRACT

Current space cryocooler developments have achieved performance and capability that have made the use of active refrigeration in space missions feasible. Space flight demonstrations such as the Sandia National Laboratory Cobra Brass and Multispectral Thermal Imager missions, the National Aeronautics and Space Administration SABER, Hyperion, and AIRS missions baselined and implemented active refrigeration to achieve mission goals. The NASA retrofit of the NICMOS cooling system on the Hubble Space Telescope, due to be installed during a 2001 servicing mission, will use a reverse Brayton cycle cryocooler to provide cooling for the NICMOS sensor due to a prematurely depleted cryogen dewar. These applications of cryocooler technology validate the improved mission capabilities and reliability and lifetime confidence in active refrigeration in space. Past development efforts have focused primarily on reliability and the achievement of long life. However, looking ahead at 21st century military space applications, there are improvements needed in several aspects of current cooling technology including higher capacity cooling loads, mass reduction, and improvement in efficiency, low temperature performance, and lifetimes greater than 10 years. In addition, cryogenic integration technology must be developed to allow efficient cryocooler to cooled component integration. Significant improvements in cryocooler technology can easily be overshadowed by gross parasitic heat loads and unacceptable cryogenic system penalties. This paper focuses on mid-term and out-year cooling requirements for the Air Force Space Based Infrared System Low, Space Based Laser, Advanced Space Based Infrared System, and other Department of Defense space missions. INTRODUCTION With the advent of the Strategic Defense Initiative in the mid 1980s, the Department of Defense recognized the improved mission capabilities of cryogenic cooling for detectors in space applications. Cooled detectors allow the collection of photons at longer wavelengths, allowing vast improvements in identification and discrimination capability with a minimum of sensor aperture growth. Smaller aperture produces cheaper, lighter sensors, much easier to host in a space-based environment. Other space missions such as communications, remote sensing, and weather monitoring can benefit from subsystems using cryogenic technology including super Cryocoolers 11, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2001

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conducting electronics, high data rate signal processors, and high speed/low power analog to

digital converters. Priorities for the Air Force Research Laboratory cryocooler effort are to develop and demonstrate space qualifiable cryogenic technologies required to meet future requirements for Air Force and Department of Defense (DoD) missions. Other objectives are to develop state-ofthe-art cryocooler technology, characterize and evaluate the performance of development hardware, pursue advanced concepts for future spacecraft missions, and work to enhance cryocooler to spacecraft integration. Cryocooler development is tracked under the Defense Technology Objectives initiative with performance improvement objectives having been established for life, power, mass, and vibration. In addition to cryocooler development, utilization of improved integration technologies ensures an optimum cryogenic thermal management system is developed that limit or eliminates operational constraints imposed on the spacecraft platform. Progress is reviewed annually at DoD level. Collaboration with other government development activities and private industry has been a major strength of the AFRL program. This has resulted in leveraging of scarce development funding and more rapid transition of cryocooler technology to the space community. Components that significantly improve the efficiency, extend life, reduce mass, or limit induced vibration are developed and transitioned into next generation cryocooler designs. The Air Force Research Laboratory (AFRL) and its predecessors, Phillips Laboratory and the Air Force Space Technology Center, has been the primary agent of the Ballistic Missile Defense Organization (BMDO) and the Space Based Infrared Low program office for the development of low capacity cryogenic refrigerators and integration technologies for space applications since the mid 1980s. These cryocooler development programs concentrated on addressing the negative impacts of mechanical refrigerators on optical space systems: induced line of sight vibration, longevity, power consumption, and mass. Additionally, initiatives funded through the Air Force Science and Technology budget have addressed critical issues for other Department of Defense users of cryogenic technology. Early development efforts were on comparatively large capacity machines to support cooling requirements for the Space Surveillance and Tracking System (SSTS). The protoflight Cryocooler program produced two three-stage 10K cryocoolers (Contractors: Air Research and Arthur D. Little) for cooling of the long wave silicon focal plane arrays. An additional program

aimed at 10K primarily developed by NASA’s Jet Propulsion Laboratory and Aerojet using sorption, culminated with the BESTCE Shuttle flight experiment in 1995. The Standard Spacecraft Cryocooler program (SSC) initiated in 1990 marked a change in emphasis from relatively large machines to more compact and efficient cryocoolers aimed at meeting cooling needs in the range from 60K to 150K for MWIR applications. Using Oxford Stirling cycle technology developed primarily in the United Kingdom, these machines utilized linear drive motors and tight clearance seal non-contacting piston shafts. The pulse tube cryocooler, a variation of this technology, replaces the actively moving expander piston with a non-moving regenerator and pulse tube. AFRL has also pursued alternate cryocooler concepts including reverse Brayton cycle designs; and for extremely low temperature cooling (~10K), variant using Joule-Thomson combinations and improved Stirling and pulse tube designs are being considered.

As user confidence in cryocooler reliability has improved, focus is also being placed on reduced mass and improved efficiency. REQUIREMENTS IDENTIFICATION

The primary purpose of AFRL cryogenic technology development program is to support military unique mission requirements. An essential element of long-term Air Force planning is to ensure users have superior military capabilities by quickly developing and sustaining the right weapon systems. It is important to fully understand users' operational requirements, to develop a range of alternative solutions and identify the critical, enabling, and enhancing technologies. At the heart of this effort are the Air Force Technical Planning Integrated Product Teams (TPIPTs) who facilitate the planning and development of technically superior and affordable solutions to

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operational needs. TPIPTs who facilitate the planning and development of technically superior and affordable solutions to operational needs. A typical TPIPT is facilitated by Product Center development planners or program office planners and consists of members from operational

commands, laboratories, air logistic centers, test centers, program offices, system engineering, and intelligence agencies. TPIPTs do development planning for the operational mission or mission support areas to include near-term planning through long-term planning -- "now until forever." TPIPTs support the user's Mission Area Assessments, Mission Needs Analyses and development of Mission Area Plans and road maps to project future operational capabilities. The TPIPT process includes the user and Air Force team members to gather, analyze, coordinate, and disseminate information in each area. TPIPTs are also a critical component of the Air Force technology master process and serve, as the primary source of weapon system technology needs. TPIPT products document candidate systems solutions to user needs, development roadmaps, and technology investment recommendations. The primary mission area for cryogenic cooling is the Space Control TPIPT that addresses space surveillance, counter space, missile warning, and space based ballistic missile defense command, control, and missile defense tasks. Other requirements are obtained from direct coordination with various Air Force and DoD program offices assessing their technical needs against current technology and desired mission improvements. One other important source for identifying technical voids is the systems contractor who eventually develops the next generation space surveillance systems. Coordination with the contractor payload developers is essential to accurately forecast needed advancements in cryogenic cooling technology. The end product of the planning process is technology roadmaps and technology investment plans, which define specific technical objectives and expected funding. Finally, AFRL works closely with other government cryocooler developers to leverage scarce technology funding and assure research efforts are not duplicated. CURRENT DEVELOPMENT PROGRAMS

The primary emphasis of the current cryocooler development efforts is to support the Engineering Manufacturing Development (EMD) requirements for the Space Based Infrared System Low (SBIRS Low) satellite program. Several machines developed by AFRL were based

lined for the SBIRS Low Flight Demonstration System (recently terminated) and Cobra Brass flight experiment. The Cobra Brass experiment and then Brilliant Eyes program office developed requirements for the TRW 150K Protoflight Spacecraft Cryocooler (PSC). This “mini pulse tube” machine has a large capacity (>2W) and high efficiency for 150K cooling, and more limited capability at colder temperatures (down to about 65K). A TRW mini pulse tube was flown on the unsuccessful NASA SSTI satellite launched in 1997 and two other units are currently flying on the Cobra Brass payload. A larger capacity TRW pulse cryocooler originally developed with BMDO funded was launched in early 2000 on the Department of Energy’s Multi Spectral Thermal Imager Satellite. AFRL managed the development of the space qualified cryocooler and flight electronics for Sandia National Laboratories. The NASA/Langley SABER flight experiment scheduled for early 2001 launch is also using a mini pulse tube supplied by AFRL.

Figure 1. TRW 150K Miniature Pulse Tube Cryocooler.

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Figure 2. TRW 60K Pulse Tube Cryocooler Integrated into the

Multispectral Thermal Imager Payload. The SBIRS Low operational system (EMD) is currently scheduled for deployment in 2006. The major drivers for EMD cryocooler requirements are increased duty cycle, higher cooling loads for the tracking sensors, increased cooling loads for the fore optics, 10 year design life with complete mechanical and electronic redundancy, and higher tolerance to radiation environments. The cryocoolers must also have a low system mass penalty and improved efficiency. While the final EMD design is still evolving, current concepts require two focal planes in the track sensor

imager to stare at a target simultaneously in both the MWIR and LWIR bands. Dual temperature cryocoolers offer attractive system benefits over single stage cryocoolers if redundant

cryocoolers are mandated. System mass penalty is defined as the sum of the cryocooler mass, the mass of the electrical power system necessary to drive it, control electronics mass, and the radiator area mass needed

to reject the waste heat. It is used as a measure of cryocooler impact on spacecraft design. The Air Force Research Laboratory is addressing the issue of system mass penalty in several ways. Thermal Storage Units (TSU) have been developed to absorb the wide thermal load variations during peak duty cycle and allow system designers to size the cryocooler for the average load instead of the peak heat load. An Air Force funded 60K TSU was successfully flight demonstrated in October 1998 aboard the STS-95 Shuttle mission. Under the Swales Aerospace CRYOBUS program, several cryogenic integration technologies are being designed and fabricated for potential ground demonstrations. This program is developing cryogenic Thermal Switches (CTSW) with high thermal resistance in the “OFF” state and low resistance in the “ON” state to make redundant cryocooler designs feasible. Additionally, the program is addressing various integration approaches for the optics cooling and heat rejection problems. A cryogenic capillary pumped loop, flexible cryogenic link, ambient loop heat pipe, and cryogenic looped heat pipe are being developed to support EMD needs.

Figure 3. Conceptual TRW 95K High

Figure 4. Conceptual Raytheon 95K High

Efficiency Pulse Tube Cryocooler.

Efficiency Cryocooler.

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An additional requirement for SBIRS Low EMD is a cryocooler to support cooling of the increased heat loads on the fore optics. Two AFRL programs, jointly funded by BMDO and SBIRS Low, have the objective to design and develop an advanced high efficiency cryocooler to meet on-gimbal optics cooling requirements of 10 Watts @ 95K. The 24-month programs will

emphasize ease of integration, low EMI signatures, low EMC susceptibility, and the ability to survive the launch loads of any existing Air Force launch vehicle. A contract was awarded to TRW in September 1998 for a potential SBIRS Low optics cooler producing 10 Watts of cooling at 95K with the added objectives of minimal cryocooler mass and input power. TRW’s design objectives are for a pulse tube cryocooler with mechanical mass of 4 kilograms and a specific power of less than 10 W/W. The AF, BMDO, and NASA are jointly funding the program. NASA is considering the cooler as a candidate for the space transportation system on the Mars Exploration Mission. While still under development, the program has already made significant improvements of the previous state-of-the-art in reduced mass, weight, and input power. A second contract was awarded to Raytheon in June 1999. This parallel effort uses a unique two-stage hybrid Stirling expander and pulse tube cold head design to achieve the desired cooling loads. An important element of both programs is the inclusion of producibility issues as an important program objective and the development of radiation hardened flight electronics. FUTURE SPACE SURVEILLANCE NEEDS

Improved discrimination utilizing high performance multicolor and multi-spectral focal planes will provide a significant improvements in operational capability for surveillance and missile tracking and detection and is being considered as a block change to the SBIRS Low system following deployment of the initial constellation. Multiple spacecraft applications could require near 10K operation with the use of Si:As infrared sensors for missions such as midcourse missile detection 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 large system penalties with Dewars and prohibitive mass penalties for most missions. Dewars are mostly applicable for short duration (< 1 year) experiments or very small cooling loads (<10 mW). 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. Near term cooling requirements for available 128 x 128 Si:As focal plane arrays are estimated at between 50 to 150 milliWatts depending on the manufacturer and AC power dissipation. In all likelihood, future systems will require larger arrays with the resulting increase in heat loads. In addition to the cooling of the tracking sensor, VLWIR surveillance systems will cool of the aft optics at temperatures projected from 40 to 60K. While current cryocooler requirements have addressed missile defense and tracking missions, cryocoolers are also being used for hyperspectral applications such as proliferation detection and treaty verification. Near term payloads are adapting previously coolers. Future systems will require cryocoolers with much larger cooling capability to meet the expected increase in focal size with projected cooling loads as high as five Watts at 35K. Large capacity, multi-load cryocoolers capable of cooling sensors, optics, and optical benches will greatly simplify the cooling approach for next generation payloads. Two current development programs are aimed at developing a low temperature capability. A BMDO funded near term effort with Astrium (formerly MMS) and Rutherford Appleton Laboratory developed a two stage Stirling cryocooler with a capacity of between 0.045 Watts at 10.3K. Astrium delivered the cryocooler to AFRL in early May 2000 for performance characterization and endurance evaluation. An Air Force funded program with design goals of 250 milliwatts @ 10K and with a specific power of <1000W/W was initiated in May 1998 with Ball Aerospace Corporation in Boulder, Colorado. This cryocooler uses an existing Stirling precooler combined with a Joule-Tompson based Redstone Interface to achieve cooling at 10K.

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Figure 5. Ball Aerospace 10K Cryocooler.

Figure 6. Ball Aerospace 10K Cryocooler.

The current program schedule calls for delivery of an Engineering Development Model to AFRL in May 2001. Another candidate technology that could support both the 10K and high capacity 35K

cooling needs is Creare reverse Brayton cryocooler based on 65K Single Stage Reverse Brayton (SSRB) cryocooler developed with BMDO, Air Force, NASA funding. A modified version of this design was successfully flight demonstrated on STS-95 and will be installed on the Hubble

Telescope during an early 2001 flight servicing mission. Under a recently awarded Air Force

Small Business Innovative Research contract, Creare is evaluating designs for a range of cooling loads and temperatures at both 35K and 10K. Minimal total input power for the cooling and operating life in excess of 10 years are program design objectives. Low frequency vibration

(<150Hz) in this unit is negligible due to the operation of the turbine at such high rotational speeds that results only in low Q high frequency vibration that have negligible effects on the ‘jitter’ of the sensor. A radial flow heat exchanger is also being incorporated into the design will reduce volume by 75% and be 70% lighter than the existing SSRB heat exchanger. If approved for Phase II, the program will produce an engineering development model at the temperature of most interest to Air Force systems developers. SPACE APPLICATIONS OF CRYOGENIC INTEGRATION TECHNOLOGY

Cryogenic system integration is becoming more and more important to the overall use of cryocoolers in space. Depending on the application, various cryogenic and ambient thermal management technologies are needed to augment and improve the cryocooler capabilities. The realization that to ensure mission success the cryogenic application must be viewed as a system and not as a component level mix and match is pushing mission planners to consider “end-toend” issues within the cryogenic system. Issues such as cryocooler redundancy, remote and

flexible cryogenic and ambient heat transport, thermal storage, and more efficient cryogenic system integration schemes are evident in near and far term system designs and mission applications. Thermal straps are currently the most common technology for cryogenic integration, but disadvantages associated with these devices include limitations on length, temperature stability through the strap, and imposed parasitic heat loads. Although adequate for many current applications, future systems will require additional capabilities that include long transport distances (over 2 meters), thermal storage for duty cycle heat load applications, stringent temperature stability, and thermal switching. Cryocooler redundancy is a primary consideration within the cryogenic system for space applications. In order to meet the long life goals (10+ years) for mission performance, system

developers are considering redundant cryocoolers to assure mission success. For single cooling

stage coolers (and no additional coolers cooling higher temperature radiation shields) the question of redundancy is reduced to selecting one of two options. The first is to select a larger capacity cooler to meet the useful cooling load of the sensor and the parasitic heat load imposed

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by the off cooler. The second is to utilize a cryogenic thermal switch to reduce the parasitic load from the off cooler. There are several methods of achieving thermal switching and include technologies such as “gas-gap” and mechanical thermal switches, diode heat pipes, loop heat pipes (LHP), and capillary pumped loops (CPL). All of these technologies provide a high conductance when “on” and have the capability to provide a low conductance when “off” to isolate the warm cryocooler.

The reduction in parasitic loads of all these devices versus thermal straps is on the order of 3 to 6 fold, but system designers must consider the overall increase in system complexity and decrease in system reliability versus a simple strap. The need for thermal switching is diminished by the use of multistage / multiload cryocoolers. In a conceptual redundant system with these coolers, the higher temperature, more efficient upper stages are cross-strapped with the off cooler. This allows the parasitic heat load

to be intercepted at a higher temperature and minimize the power increase to the operating cooler and the parasitic heat load on the cooled sensor. Heat pipes, LHPs, and CPLs are passively pumped systems. This means that the source of pumping for the working fluid is provided by capillary action in a wick structure. These devices have the capability for long distance, efficient cryogenic heat transport. In the case of the LHP and the CPL, the cooling loop has an evaporator to absorb the cryogenic heat load and a condenser to reject that heat at the cryocooler cold tip. These components are connected via long thin walled tubes and provide excellent thermal and vibration isolation between the cooler cold tip and the sensor. Additional capabilities that are being developed are multiple evaporator and condenser interfaces for these loops. Actively pumped systems utilize an ambient temperature compressor, an interface for the cryogenic load, and a cryocooler as a precooler for the loop. This concept is very flexible for many different system concepts, however it has an added concern for system reliability due to the compressor. Technology is under development to address concerns for flexible cryogenic heat transport across a two-axis gimbal and development of a hybrid system to achieve 10 Kelvin cooling.

Figure 7. Cullimore & Ring – Swales Aerospace – NASA GSCF Cryogenic Capillary Pumped Loop.

Figure 9. Technology Applications Inc. Cryocooler Interface System.

Figure 8. Swales Aerospace 65 K Thermal Storage Unit (CRYOTSU).

Figure 10. Swales Aerospace Cryogenic Thermal Switch.

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GOVERNMENT CRYOCOOLER DEVELOPMENT AND TEST PROGRAMS

Thermal storage and temperature stability is also a consideration for future systems. Thermal storage is a consideration for systems that have a significant heat load duty cycle. The thermal storage unit (TSU) is a phase change or sensible heat device that allows the cryocooler to be sized for the average heat load of the system and not the peak. When the heat load is low, the cryocooler reverses (for solid-liquid systems), recondenses (for liquid-gaseous systems), or recharges (for variable temperature devices that are part of a cooling loop) the TSU. When the load is high the cooler remains cooling at the average level and the TSU absorbs the additional heat at constant or near constant temperature. TSUs do provide temperature stability, but future systems may require much more stringent control of the sensor temperature. Future cryogenic applications will be extremely system design dependent. However, cryogenic integration technology is currently under development to increase performance over the state of the art, significantly reduce cryogenic system penalties, and enable new system concepts. POTENTIAL SPACE-BASED LASER CRYOGENIC REQUIREMENTS

Cryocoolers potentially are enabling technology for future SBL space systems, which will have significant cooling requirements for the cryogenic fuel. 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 replenishment. Cryocooler integration for large tanks will be a significant issue. High capacity cryocoolers and long term (>20 years) on orbit propellant storage are potentially enabling technology for future High Energy Laser (HEL) space systems, orbital transfer vehicles, and on orbit propellant depots. A number of critical issues are currently being evaluated by the SBL program office and contractor teams regarding the use of elevated temperatures for cryo-gas, use of multi-load cryocoolers and reduction of storage pressure. Other significant concerns for cryogenic applications in space based systems requiring long term cryogen storage includes substantial cooling requirements for subcritical cryogens, cryocooler redundancy, on orbit cryogen transfer from vehicle to vehicle, large shield cooling, long term gas and liquid cryogen storage, large distributed cooling surfaces, cryogenic system integration, and the significant spacecraft system penalties due to mass and input power. AFRL has pursued low capacity cryocooler concepts including reverse Brayton cycle, single and multiple stage Stirling cycle, advanced Joule-Thomson cycle, and Pulse Tube (Stirling cycle variant) designs and the technology development spans a wide range of cooling temperatures (from ~10 Kelvin to 150 Kelvin) and heat loads (up to 10 Watts at 95 Kelvin). Additionally, AFRL has pursued advanced cryogenic integration technology including cryogenic thermal switches, cryogenic heat transport, thermal storage, and cryogenic integration schemes to reduce system mass and input power penalties. Current cryogenic integration and cryocooler development programs address the negative impacts of the cryogenic system on optical space systems: including induced line of sight vibration, longevity, power consumption, mass, thermal transport, thermal storage, and thermal switching. However, the cryogenic cooling requirements for future Air Force systems may require large capacity cryogenic cooling, extremely mass and power efficient mechanical refrigerators, and significant improvements in long term on orbit cryogen storage. 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. CONCLUSIONS

The AFRL leveraged approach to cryocooler development is providing technology to support near term requirements for SBIRS Low and other DoD programs. Cryocooler mass has

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been steadily reduced, cryocooler induced line of. sight vibration has fallen below the value allocated for other sources, and user confidence in cryocooler reliability has improved. The cryogenic integration effort at AFRL is developing hardware to meet out-year cryogenic integration requirements. Integration developments lead to more efficient cryocooler / cryogenic

systems, which further enhance the mission capabilities of this already enabling technology. Future systems will benefit from increased investments in multi-temperature, larger capacity cryocoolers, and more compact, radiation hardened flight electronics. Pursuit of advanced concepts such as optical cooling could produce systems with no moving parts, and use of MEMS for improved fabrication technology can reduce weight and improve performance with the resulting a much enhanced integrated sensor/cooler/processing packaging, improved reliability, and enabling surveillance at very long wavelengths.

REFERENCES 1. Swift, W.L., “Single Stage Reverse Brayton Cryocooler: Performance of the Engineering Model,” Cryocoolers 8, Plenum Press, New York (1995), pp. 499-506. 2. Burt, W.W., and Chan, C.K., “Demonstration of a High Performance 35 K Pulse Tube Cryocooler,” Cryocoolers 8, Plenum Press, New York (1995), pp. 313-319. 3. 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. 4. Curran, D.G., “Use of Two Stages of Cooling to Reduce Space Based Laser (SBL) Cooling Requirements for Both IFX and EMD Cryocooler Procurement,” Aerospace Corporation Thermal Control Department briefing, Mar 99. 5. Orlowska, A.H., Bradshaw, T.W., Scull, S., Tomlinson, B.J., “Progress Towards the Development of a 10K Closed Cycle Cooler for Space Use,” Cryocoolers 10, R.G. Ross, Jr., Ed., Plenum Press, New York (1999), pp. 67-76. 6. Bugby, D., Stouffer, C., Davis, T., Tomlinson, 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.

Status of Programs for the DoD Family of Linear Drive Cryogenic Coolers for Weapon Systems 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 second-generation thermal imaging systems. DOD has established a family of SADAs to address high performance (SADA I), mid-to-high performance (SADA II), and compact class (SADA III) systems. SADAs consist of the Infrared Focal Plane Array (IRFPA), Dewar, Command & Control Electronics (C&CE), and the cryogenic coolers. SADAs are used in weapons systems such as Comanche, the M1 Abrams tank, the M2 Bradley fighting vehicle, and the Javelin CLU. The linear drive cryocoolers maintain the Infrared Focal Plane Arrays (IRFPAs) at the desired operating temperature. Stirling linear drive cryocoolers are being used in place of Stirling rotary coolers. DOD has defined a family of tactical linear drive coolers in support of the family of SADAs. These coolers are required to have low input power, a quick cool-down tune, low vibration output, low audible noise, and higher reliability. This paper (1) outlines the characteristics of each cooler, (2) presents the status and results of qualification tests, and (3) presents the status and test results of efforts to increase cryocooler reliability. Flexure-spring designs of the 0.15 watt and 1.0 watt coolers are currently in reliability growth testing. INTRODUCTION

The US Department of Defense (DoD) has chartered a strategy to standardize second generation infrared (IR) components throughout the services. A family of second-generation (2nd Gen.) infrared-imaging critical components called the Standard Advanced Dewar Assemblies (SADA’s) has been developed to support this strategy. SADA I is designed to address requirements for high performance systems, SADA II for mid-to-high performance, and SADA III for compact class systems. SADAs consist of the Infrared Focal Plane Array (IRFPA), dewar, Command & Control Electronics (C&CE), and the cryogenic cooler. The US Army CECOM Night Vision and Electronics Sensors Directorate (NVESD) has developed a family of Stirling cycle linear drive coolers, shown in Fig. 1, in support of this standardCryocoolers 11, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2001

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Figure 1. Family of linear drive coolers.

ization effort.1 These coolers address the shortcomings of rotary coolers such as low reliability, poor shelf life, multi-axes vibration & torque, excessive acoustic noise, and poor temperature stability of the detector array. This paper highlights the latest developments and results involving US Army programs for linear drive coolers. Table 1 highlights the key parameters of the family of coolers.

PROGRAMS FOR DoD FAMILY OF LINEAR DRIVE COOLERS

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QUALIFICATION REQUIREMENTS FOR CRYOCOOLERS

SADAs and linear drive coolers are products that require qualification prior to first delivery. These components are qualified once they pass through a series of tests approved by NVESD and

the procurement activity. The government or the cryocooler manufacturer may perform these tests. The government approves all test procedures, equipment, and test facilities prior to testing. 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, Improved TOW Acquisition Sensor, and Long Range Advanced Scout Sensor Suite (LRAS3). The tests in Table 2 are part of the qualification effort for linear drive coolers. 0.15-WATT LINEAR DRIVE COOLERS The 0.15-watt linear drive cooler was developed for second-generation FLIR man-portable applications. The 0.15-watt cooler from DRS Infrared Technologies (formerly Texas Instruments) was originally qualified in 1997 for use in the Javelin Command Launch Unit (CLU). Javelin is an anti-tank missile system. This 0.15-watt cooler was re-qualified in 1999 following an Army funded Manufacturing Technology (Mantech) program with DRS.2,3 This Mantech program was performed on both the 0.15watt 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. It developed new production techniques and processes that reduced the cost to manufacture the 0.15-watt cooler by 30%. It focused on process improvements to the

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compressor clearance seals, gas decontamination process, motor manufacturing, and cooler final assembly. It also replaced 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 8000 hours MTTF. A significant cost avoidance will be realized due to lower manufacturing costs and higher reliability coolers, and the potential production requirement for more than 7,000 of these coolers in the next 15 years. The 0.15-watt Javelin cooler with flexure springs is currently in full-scale production. It passed all required Javelin detector dewar cooler assembly performance and qualification tests prior to production. Three Javelin coolers began life testing on February 1, 1999 at DRS. These life test coolers have accumulated an average of 4777 relevant hours of operation through January 28, 2000. Two

units, MSN 18 and MSN 13, have 5289 hours each and the third unit has 3754 hours since being returned to the life test. Javelin cooler numbers 18 and 13 have been failure free. The only anomaly during the period occurred when the chamber failed to hold the prescribed temperatures during the cycle. 1.0-WATT LINEAR DRIVE COOLERS

The 1.0-watt cooler is 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 the improved TOW acquisition system. Status of Qualification Efforts

The DRS Infrared Technologies 1.0-watt coolers were 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 is also a 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. Both cooler manufacturers demonstrated reliability over the 4,000-hour MTTF requirement. DRS coolers accumulated an average of 6,486 hours and the AIM coolers accumulated 4,753 hours. Qualification testing of Litton Life Support 1.0-watt coolers is near completion. These coolers have passed most formal qualification tests. Only the Electromagnetic Radiation and the reliability tests remain to be completed. The coolers have accumulated an average of 2,500 hours in reliability testing. NVESD and PM FLIR support Litton’s formal qualification testing. Manufacturing Technology (Mantech) Efforts

The 1.0-watt cooler was also the beneficiary of an Army Mantech effort that resulted in a 32% decrease in cooler manufacturing costs at DRS. 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 (12,000 coolers projected in the next 20 years). In order to maximize competition, the Mantech program included a technology transfer effort that provided DRS reports and briefings to approved cooler manufacturers. Reliability Improvements The DRS 1.0-watt cooler with flexure springs is currently in qualification and life testing. This effort is funded by the Army’s Operation and Support Cost Reduction (OSCR) Program and managed by NVESD. The goal is to increase the life of the 1.0-watt cooler from 4,000 hours to 8,000

PROGRAMS FOR DoD FAMILY OF LINEAR DRIVE COOLERS

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Figure 2. Reliability Test Profile, 1.0-watt linear cooler with external electronics.

hours Mean Time to Failure (MTTF) in order to reduce operation and support costs. The lessons learned in the development of the DRS 0.15-watt Javelin cooler with flexure springs were applied to the 1.0-watt cooler effort. Figure 2 depicts the reliability test profile for 1.0-watt coolers. Six 1.0-watt coolers with flexure springs have been built, and test results show that they match the performance of the 1.0-watt coolers (helical springs) currently qualified and in production. Several modifications have been necessary since testing started in 1999. One modification deals with the addition of more rugged

clamping mechanisms after the flexure spring stack shifted during the mechanical shock and mechanical vibration tests. Reliability testing on three 1.0-watt coolers with flexure springs began on April 16, 1999. One unit was removed from testing due to unacceptable vibration output levels. A defective magnet assembly caused the unacceptable vibration output levels. This unit was replaced with two addi-

tional coolers to total four coolers in life test. These reliability test coolers have surpassed the required 4,000 relevant hours of operation with an accumulated average count of 4,600 hours. Testing will continue until 8,000 hours MTTF are demonstrated or until failure of the coolers. Additional tests are required to demonstrate full conformance to qualification requirements. 1.75-WATT LINEAR DRIVE COOLERS The 1.75-watt cooler is designed to address the needs of the high performance second generation infrared imaging systems that will use a SADA I or other equivalent performing system whose cooling capacity requirements or faster cool-down times cannot be met with a 1.0-watt cooler.

These coolers are currently used in many DoD programs to include the Army’s Comanche and Apache helicopters.

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Several AIM 1.75-watt coolers were purchased and are undergoing evaluation and formal qualification testing as part of a Foreign Comparative Test (FCT) program with NVESD and the Program Manager for Comanche (PM Comanche). Several 1.75-watt coolers have been successfully integrated into Comanche, Apache, and several other high performance FLIR systems to include Quantum Well FLIRs. The production requirement for these coolers is estimated at over 8,000 in

the next 20 years. FUTURE COOLER EFFORTS

Army efforts will continue to focus on reducing manufacturing costs, improving reliability, and improving the performance of the family of linear drive Stirling cryocoolers. NVESD, PM NV/RSTA, and PM Comanche are supporting efforts to increase the reliability of 1.0-watt coolers to 12,000 hours MTTF and the 1.75-watt cooler to 8,000 hours MTTF. Fiscal year 2001 funds are possible for these efforts. A related effort will aim to reduce the manufacturing costs of the 1.75-watt coolers through the transfer of manufacturing processes developed under the 1.0watt and 0.15-watt cooler Mantech program. With the advent of longer life coolers (8000-12,000 hours) there is a need to develop a method to shorten the reliability testing while demonstrating cooler life. NVESD is exploring alternatives to the current reliability test method. The current test method demonstrates cooler life by actually running the coolers until failure. With the current method the coolers accumulate 450 hours of runtime per month. It will take 30 months to demonstrate a 12,0000-hour cooler introduction of a valid accelerated reliability test. The cooler is still the least reliable component in FLIR systems and there is a need to accelerate the introduction of longer life cryocoolers into DoD systems. SUMMARY A family of second-generation (2nd Gen.) infrared-imaging critical components called the Standard Advanced Dewar Assemblies (SADA’s) has been developed to support a DoD standardization strategy. A family of linear drive coolers has also been established in support of this standardization strategy. The US Army CECOM Night Vision Directorate in conjunction with several US Army Program Managers has embarked in linear drive cooler efforts aimed at qualifying coolers, reducing cooler prices, and increasing cooler reliability. Qualification test efforts for the DRS 0.15-watt flexure spring and for the DRS & AIM 1.0watt coolers were successfully completed in the last five years. Qualification testing for the 1.75watt AIM cooler and two additional 1.0-watt cooler designs (DRS flexure springs and Litton) is ongoing. Some of the weapon systems supported by these qualification efforts include the Army’s Second Generation Forward Looking Infrared (FLIR) Horizontal Technology Integration (HTI), Comanche, Apache, Javelin, Improved TOW Acquisition Sensor, and Long Range Advanced Scout Sensor Suite (LRAS3). The US Army Manufacturing Technology program and the Operation and Support Cost Reduction program were successful in reducing manufacturing costs and in implementing flexure spring designs aimed at increasing the reliability of the 0.15-watt and 1.0-watt coolers. Additional work is planned to transfer these improvements to the 1.75-watt coolers and to further increase the reliability of 1.0-watt coolers.

REFERENCES 1. J. Shaffer and H. Dunmire, “The DOD Family of Linear Drive Coolers for Weapons Systems,” Cryocoolers 9, Plenum Press, New York (1997), pp. 17-24. 2. Raytheon Texas Instruments Systems, “Linear Drive Cooler Mantech Program,” Contract DAAB0795-C-J513 Industry Review, December 1997. 3. DRS Infrared Technologies, “Linear Drive Cooler Mantech Program,” Contract DAAB07-95-CJ513 Industry Review, February 1999.

Air Force Research Laboratory Cryocooler Characterization and Endurance Update B. J. Tomlinson1, C. H. Yoneshige1, and N. S. Abhyankar2 1

Air Force Research Laboratory

Kirtland AFB, NM, USA 87117-5776 2

Dynacs Engineering

Albuquerque, NM, USA 87106-4266

ABSTRACT

The Air Force Research Laboratory (AFRL) has been instrumental in advancing space cryocooler technology through cryocooler development and characterization of the. long-term performance of different types of cryocoolers. These coolers were developed to support the long life space mission requirements of the United States Air Force SBIRS-Low Program Office, the Ballistic Missile Defense Organization (BMDO), the National Aeronautics and Space Administration and the Department of Defense. Long life cryocooler applications include cooling infrared sensors, focal planes, optics and electronic circuits for various space missions of national interest.

The main objective of this paper is to present the status of the cryocoolers currently undergoing characterization and endurance evaluation at AFRL. The information gained through these processes is shared with industry partners, cryocooler developers, technology sponsors, and technology users. This feedback is essential for cryocooler design enhancements and future cryogenic technology development efforts. There are two cryocoolers undergoing characterization at AFRL. These include the Astrium (formerly Matra Marconi Space) 10 K Cryocooler, and the TRW 150 K miniature pulse tube (150K MPT). In addition, there are five cryocoolers undergoing endurance evaluation at AFRL These include the Raytheon Protoflight Spacecraft Cryocooler, Raytheon Standard Spacecraft Cryocooler (SSC) II, TRW 3585, TRW 6020, and the Defense Evaluation and Research Agency (DERA) cryocooler. This paper includes each cryocooler’s status, and update or initial report on its performance, elapsed runtime hours, performance anomalies and updated characterization and endurance evaluation data. INTRODUCTION

The most critical characteristics of cryocoolers for strategic space applications are lifetime and reliability. This is what distinguishes them from their short-term tactical cousins. The primary purpose of the Air Force Research Laboratory Cryogenic Cooling Research Facility (CCRF) is to explore the thermodynamic performance characteristics of one-of-a-kind or firstof-a-kind engineering design model or protoflight space cryocoolers and assess their lifetime and Cryocoolers 11, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2001

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reliability to help meet technology needs for the Air Force, the Ballistic Missile Defense Organization (BMDO), and Department of Defense (DoD) strategic space cryogenic cooling applications. Cryocoolers are enabling technology for cryogenic optical systems and infrared

sensors. During characterization, each cryocooler undergoes a rigorous series of experiments, which determine its thermodynamic performance capabilities. The effects that sensor duty cycling, changes in heat rejection temperature, heat load, and input power have on the cryocooler’s performance are investigated. During the endurance evaluation, the cryocooler is constantly run at its nominal operating parameters until it meets pre-determined failure criteria. One of the major issues with establishing confidence in the reliability of cryocooler technology is the cryocoolers’ 10+ year design life and the absence of accepted accelerated testing methods. The CCRF is equipped with one 6 foot, two 36 inch, eight 24 inch and two 28 inch thermal vacuum chambers. Each of these chambers is capable of high vacuum (~10-7 torr), which helps closely simulate a space environment, and has independently controlled conductive heat rejection surfaces to simulate a potential spacecraft thermal interface. Each chamber also has multiple feed-throughs for chiller fluid lines, cryocooler control electronics, data acquisition instrumentation, and computerized environmental controls. Cryocoolers can also be characterized in a table-top configuration, especially cryocoolers that are not qualifiable for space flight, but have significant heritage to current or in development space flight cryocoolers. CHARACTERIZATION

Characterization consists of two major portions. The first part of the characterization process is the initial baseline / acceptance evaluation. This evaluation is done to ensure that the cryocooler meets the manufacturer’s performance specifications and verify the “as delivered” performance prior to continuing the experiment. Before the cooler even reaches the laboratory, AFRL engineers have spent time with the contractor to gather vital data necessary to design the experimental test stand and prepare an adequate plan for characterization. The second part of the process is the actual characterization of the cryocooler. AFRL engineers and technicians perform experiments with the cooler to fully understand and map its thermodynamic performance and examine the complex interrelationships between the different operating

parameters. During this process, it is possible to concentrate on various areas of interest where the technology under examination could potentially be applied. This is part of the development life cycle for cryocoolers and requires inputs from technology users, government, and system integrators. Initial Baseline Evaluation When a cryocooler arrives at AFRL, it normally undergoes an initial baseline/acceptance evaluation. This evaluation consists of refrigeration performance baselines, nominal load lines, low temperature stability trials, cool-down to lowest temperature trials (no heat load applied) and stiction tests. This phase of characterization also serves as a screening method to identify any manufacturing defects or shipping and handling damage. There are two cryocoolers currently in this phase of characterization. They are the TRW 150K miniature pulse tube (MPT) and the Matra Marconi Space (MMS) 10K Cryocooler. TRW 150K MPT. The TRW MPT is designed to lift 1W @ 150K at various heat rejection temperatures. The AFRL MPT #006 was provided to NASA in exchange for MPT#002. The #006 has been qualified and integrated into the NASA Sounding of the Atmosphere using Broadband Emission Radiometry (SABER) experiment. MPT #002 is currently at AFRL and is being integrated with a 6-axis dynamometer for a functional test and an induced vibration trial. The cooler will then be moved to a 24” thermal vacuum chamber for an initial baseline evaluation and characterization. Astrium 10K. The Astrium 10K cryocooler (formerly Matra Marconi Space) is designed to lift 45 mW at 10.4 K. It was delivered to AFRL in early May 00 and was set up on a table top for a functional test. This cryocooler will be integrated with a 36” thermal vacuum chamber for

AFRL COOLER CHARACTERIZATION AND ENDURANCE UPDATE

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Figure 1. Astrium 10K cryocooler.

its initial baseline evaluation and characterization. Figure 1 shows the Astrium 10K cryocooler in its table-top configuration. Performance Characterization

In order to fully understand the cryocooler technology developed for the Air Force and DoD, the CCRF conducts experiments, which are meant to explore the full thermodynamic range of performance for each cryocooler. The AFRL objectives for characterization are as follows:

1. Characterize the full thermodynamic performance envelope of emerging cryocooler technologies to establish parametric performance models.

2. Feedback performance data, models, and lessons learned to parties involved in the cryocooler development process. During the characterization phase, the cryocooler goes through a series of experiments designed to characterize its thermodynamic idiosyncrasies versus predicted performance as well as its potential to be integrated with a spacecraft cryogenic system. Load lines, refrigeration performance baselines, temperature stability experiments, transient heat rejection thermal response, cool-down curves, and off-state conduction (determination of the parasitic heat load)

are normally done during this phase of characterization. This data is fed into parametric performance models to provide spacecraft systems integrators tools to model the cryocooler performance versus specific mission requirements. AFRL is expecting the re-delivery of the Ball Aerospace 35/60K cryocooler. This cryocooler was designed to lift 0.4W @ 35K and 0.6W @ 60K. It is a split-Stirling cryocooler with 3 stages and has the ability to lift 2 heat loads at 2 different temperatures. The initial

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baseline/acceptance evaluation was performed to evaluate its nominal performance. The cryocooler was returned to Ball for retrofit of the expander cold head due to a design flaw discovered on the NASA Ball 30K cryocooler. The cooler is currently in the final stages of reassembly and should be delivered to AFRL in July 00. ENDURANCE EVALUATION

AFRL objectives for endurance evaluation have evolved to meet the following requirements for demonstration of the lifetime and reliability of cryocooler technology developed for the Air Force, BMDO and the DoD: 1. The focus of the long life endurance evaluation of cryocooler technologies is on the characterization of the long term performance degradation, system reliability, reliability contributors and detractors, an accurate assessment of the lifetime of the technology, and the development of accepted methods for the accelerated testing of long life space cryocoolers. 2. AFRL provides feedback to cryogenic technology developers, users, and spacecraft developers in the form of technical reports and conference presentations of data and lessons learned in order to aid follow-on development efforts and add to the body of reliability data on cryocooler long term performance. The endurance evaluation is normally run until the cryocooler meets predetermined failure criteria. If the cryocooler continues to perform nominally past its design lifetime requirement, it will be allowed to run until it does meet the failure requirements. Even though endurance data is valuable for understanding the life and reliability characteristics of a cryocooler, an endurance evaluation that lasts for more than 5 years will not meet technology insertion freeze dates for critical Air Force and DoD programs or provide the necessary data to impact follow-on technology development programs of similar heritage. Thus, AFRL is working on developing accepted accelerated testing methods to help increase confidence in the reliability of emerging technologies, while still meeting technology insertion freeze dates. There are five cryocoolers currently undergoing endurance evaluation at AFRL. Table 1 shows these cryocoolers and their nominal operating conditions. These operating conditions are

maintained throughout the endurance evaluation, except for periodic load line checks, which are done to track any performance drift of the cryocooler. Each of the cryocoolers’ heat rejection temperatures is cycled above and below the nominal heat rejection temperature listed in Table 1. This allows engineers to monitor steady state performance over the design thermal rejection temperature range for each cryocooler. It also allows AFRL engineers to perform baseline evaluations at different heat rejection temperatures. For comparison purposes, 300K is defined as the nominal heat rejection temperature for all cryocoolers. The rejection temperature range is defined by the cryocooler’s design margins based on an intended orbital transient temperature profile, it’s sensitivity to coefficient of thermal expansion effects and thermodynamic performance limits.

The endurance evaluation can be run in a thermal vacuum chamber or on a table-top. Coolers are run on table top if they are not space flight qualifiable, but have a significant heritage to existing or in development space cryocoolers. AFRL currently has a total of 5 cryocoolers

AFRL COOLER CHARACTERIZATION AND ENDURANCE UPDATE

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undergoing an endurance evaluation. Table 2 shows a summary of the coolers in endurance. There are two cryocoolers running in a table-top experiment set-up. They are the Defense Evaluation and Research Agency (DERA) cryocooler and the TRW 3585 pulse tube cryocooler. DERA Cryocooler. The DERA cryocooler is a miniature split-Stirling cooler designed to lift 0.25W @ 65K. It was developed under sponsorship from BMDO to examine the development of cryocoolers for Infrared Focal Plane Array (FPA) applications at wavelengths

from ~3µm to 15µm with cooling to 65K and below. The lifetime of this cryocooler was estimated to be about 20,000 hours. The cryocooler design incorporates the use of flat section springs to support the compressor pistons. The compressor unit contains two pistons set back-toback, which work on a common compression space. The free displacer is balanced using a balance mass opposed to the displacer piston. Sliding plastic seals are used inside the cryocooler. The plastic seals ensure no metal-to-metal rubbing contact along all running clearances. TRW 3585. The TRW 3585 pulse tube cryocooler has accumulated over 19,000 hours. It is designed to lift 0.85W @ 35K. This is an engineering design model single stage orifice pulse tube with a 20cc compressor. This cooler was the first of three units developed under the AFRL

TRW 35K Pulse Tube Cryocooler program. This cooler has limited follow-on applications due

Figure 2. TRW 6020 load line comparison between JPL and AFRL (Nov 1998 data, rejection

temperature 293 Kelvin).

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to evolutionary changes in the compressor and cold head design, but is being evaluated because it has significantly contributed to the family heritage of the TRW/Oxford Stirling compressor. Due to an error discovered in the calibration of the capacitance sensor electronics, this cryocooler is currently being re-baselined for comparison to data taken at the Jet Propulsion Laboratory in 1993. AFRL has three cryocoolers undergoing an endurance evaluation in a thermal vacuum chamber. They are the Raytheon (formerly Hughes) Standard Spacecraft Cryocooler (SSC) II, the Raytheon Protoflight Spacecraft Cryocooler (PSC) and the TRW 6020 pulse tube cryocooler. Raytheon SSC II. The SSC II is a split-Stirling cryocooler designed to lift 2W @ 65K with a specific power of 30 W/W. During its initial baseline/acceptance evaluation, the SSC II specific power exceeded its design specifications and the cooler developed hardware problems. The cooler was returned to Raytheon for refurbishing and sent back to AFRL. A short minicharacterization was completed for design and “as delivered performance” verification. During the mini-characterization, the cooler was not able to maintain 2W @ 65K without occasional

tripping possibly due to compressor warm-up at 300K rejection temperature. The nominal operation point was then reset to 1.5W @ 65K in safer mode. The cooler has been running in

endurance at this new operation point since January 2000. Raytheon PSC. The Raytheon PSC is a split-Stirling cryocooler designed to lift 2W @

60K. This cryocooler is the latest design from Raytheon and is directly related to the older Raytheon SSC II currently under endurance at AFRL. During its initial baseline/acceptance evaluation, the cooler demonstrated the ability to efficiently lift 1.2W @ 35K.

After characterization, the PSC entered its endurance evaluation, but was shut down temporarily due to an electronics problem in the control rack. With the problem was solved, the cooler was moved to a 24” thermal vacuum chamber and is continuing with endurance evaluation. TRW 6020. The TRW 6020 pulse tube cryocooler is a single stage orifice pulse tube with a

10cc dual opposed compressor designed to lift 2W @ 60K. It was integrated into a 24” thermal vacuum chamber for characterization and long term endurance evaluation. Its heat rejection

temperature is set by a copper block interface to a computer controlled chilled fluid loop. Apparent discrepancies in the characterization data led AFRL to conduct a detailed review of all the data on this cryocooler and also examine similar data discrepancies with the TRW 3585 and the TRW 3503 pulse tubes. The first indication of discrepancies in the data was in the form of a distinct difference in performance compared to data gathered during a characterization performed in 1994-5 at the NASA Jet Propulsion Laboratory. Additionally, data taken at AFRL indicates a possible shift in parasitic load on the cold block over time. Figure 2 shows the apparent performance shift based on data taken in November 1998. The assumptions were that

the parasitic load on the cold end was similar to the JPL set-up and that the stroke conversion values as provided by TRW were correct.

An investigation into the apparent differences in performance revealed that the conversion factors used to determine the stroke length on compressors A and B from the capacitance sensor box readings were incorrect. The conversion factors used during the November 1998 data comparison were the factors that TRW provided in the operation manual for the cryocooler. However, when the 6020 was undergoing its initial characterization at JPL, the 6020 had to be returned to TRW to stake the capacitance sensor cap internal to the cooler. During this time, TRW recalibrated the capacitance sensor electronics and provided these values to JPL. When the cooler arrived at AFRL it was assumed that the conversion values were the ones listed in the operations manual. These conversion factors for the capacitance sensor readings were put into the data acquisition software and contributed to the fact that the cryocooler appeared as if it was experiencing degradation in performance.

Stiction tests completed recently at AFRL showed that the conversion factors being used were indeed not correct and different from the conversion factors used by JPL. Figures 3 and 4 show traces from stiction tests completed in May and June 2000 at AFRL. New conversion factors were calculated using this “soft” calibration and the LabVIEW™ data acquisition software was updated.

AFRL COOLER CHARACTERIZATION AND ENDURANCE UPDATE

Figure 3. Stiction Trace from Compressor 1.

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Figure 4. Stiction Trace from Compressor 2.

The JPL load lines were reaccomplished at AFRL, and the similarity in the data shows no significant degradation in performance. Figure 5 shows the load line comparison for data taken in May 2000 and Figure 6 is a multivariable plot that includes AFRL and JPL data. Figure 7 is

a plot of AFRL data that attempted to match the temperatures and heat load of a JPL 293K, 11mm load line. The plot shows that the input power is nearly identical, but the stroke needed to achieve the JPL performance was 0.23 mm longer on average. However, this brings up another potential source of error in that the JPL stroke data on average differed 0.15 mm from compressors A to B. The AFRL data only varied 0.06 mm from side to side. Apparently, the

Figure 5. May 2000 TRW 6020 load line comparison between JPL and AFRL data (rejection temperature 293 Kelvin).

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Figure 6. Multivariable Plot for May 2000 TRW 6020 JPL and AFRL data (rejection temperature 293 Kelvin).

Figure 7. June 2000 TRW 6020 load line comparison between JPL and AFRL data.

AFRL COOLER CHARACTERIZATION AND ENDURANCE UPDATE

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output of each channel of the amplifier has a distinct impact on the actual stroke for each compressor. Differences seen in the plots are most likely accounted for by instrumentation and measurement errors, but it is apparent that the 6020 performance is very close to the data that was taken at JPL almost five years ago. A complete error analysis is still being constructed. An additional source of error that contributed to the apparent difference in cryocooler performance from JPL to AFRL in the fact that the parasitic load on the cold end was significantly larger at AFRL in November 1998. The JPL quoted measurement was acquired utilizing a cryocooler intercept to the ”off’ 6020. The method at AFRL is a parasitic warm-up test that allows the measurement of an effective warm up parasitic on the cold block. JPL had values of -0.5 W for a parasitic load compared to AFRL which had a high of 1.4 W in Jan 00 (Table 3). It was determined that the multilayer insulation (MLI) wrap, the heater, and the sensor leads were inadequate and were the source of the large discrepancy in parasitic load. To alleviate the parasitic load on the 6020, AFRL engineers and technicians designed a new MLI blanket and wrapping scheme, the instrument leads were replaced with very thin wires to minimize the parasitic heat conduction, and the oversized cryogenic heater was removed and replaced with a 5W power resistor (JPL used 0.25 W wire-wound ceramic resistors). This effort culminated in the reduction of the effective parasitic to 0.6 W. Based on these findings it is apparent that a stiction-calibration of the capacitance sensor

electronics and a parasitic warm-up test are needed components of baseline performance checks during endurance evaluation. The TRW 6020 endurance evaluation plan has been updated to include these procedures and similar calibrations and parasitic load mitigation efforts are underway for the TRW 3585 cryocooler. SUMMARY

The AFRL Cryogenic Cooling Research Facility continues to characterize and evaluate the long life performance of emerging technology for strategic space cryocoolers for the Air Force, BMDO, and the DoD. Design and execution of these experiments are critical to ensure accurate examination of the technology. Long life endurance evaluation is a crucial component for system integrators and technology users to determine life and reliability heritage. ACKNOWLKDGMENTS This work was a team effort and included contributions from Mr. John Kallman (AFRL), Mr. George Lybarger (AFRL), and Mr. Michael Martin (Dynacs Engineering).

Air Force Research Laboratory Cryocooler Reliability Initiatives S. Blankenship and T. Lynn Fountain Georgia Tech Research Institute Atlanta, GA T. M. Davis and B. J. Tomlinson Air Force Research Laboratory Kirtland AFB, NM 87117

ABSTRACT The primary concern of spacecraft developers when considering active refrigeration for space missions is lifetime and reliability. Lack of confidence in current cryocooler technology to achieve mission performance goals and achieve the necessary lifetime is a deterrent that often precludes the

consideration of cryocooler technology for many space applications. The Air Force Research Laboratory (AFRL), through the Georgia Tech Institute of Technology, has conducted two workshops with government and industry developers and technology users to address cryocooler reliability. These workshops highlighted critical issues associated with mechanical, electronic, and software reliability for cryocooler systems. Also addressed were issues on contamination control, performance testing, acceptance testing, manufacturing in-process testing, environmental qualification screening, endurance evaluation and demonstration, and producibility issues. In addition, AFRL has pursued research and development efforts to augment cryocooler technology development with the aim of increasing reliability. Efforts such as vibration mitigation through improved compressor design, improved flexure bearings, reduced gas contamination, and

potential accelerated testing approaches have been explored. AFRL in-house laboratory characterization and endurance evaluation of cryocooler technology has also contributed to the lifetime and reliability database for emerging technologies. This information is invaluable as a tool for spacecraft developers to understand the lifetime and reliability potential of a candidate cryocooler technology and establish a heritage database for cryocooler families. Additionally, AFRL has made direct contributions to improvements in follow-on designs through the feedback of performance and endurance data to industry developers. INTRODUCTION

Lifetime and reliability are driving concerns for the use of active cryogenic cooling technology in space. Military, commercial, and scientific applications have driven the requirements for the development of long life (10+ years), high reliability cryocoolers for three decades. The scope of development issues for active refrigeration includes the mechanical unit itself, the power condiCryocoolers 11, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2001

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tioning and control electronics, and the software utilized for cryocooler operation. Recent develop-

ments in the state of the art have vastly improved the current generation of cryocooler technology, but significant issues remain and chiefly center around the reliability of the devices utilized for long

life mission applications. The Air Force Research Laboratory (AFRL), Space Vehicles Directorate, Cryogenic Technology Group has pursued research and development efforts to develop cryocooler technology to meet lifetime, reliability, and thermodynamic performance to meet Air Force and Department of Defence mission requirements for over 15 years. Recently completed AFRL advanced technology

development programs such as the Raytheon Protoflight Spacecraft Cryocooler, the Ball Aerospace 35/60K Cryocooler, and the Astrium (formerly Matra Marconi Space) 10K Cryocooler have made an impact on the state of the art for mechanical refrigeration. Additionally, current ongoing

development programs such as the High Efficiency 95K Cryocooler programs with TRW and Raytheon, and the Small Business Innovative Research programs to develop cryocooler and cryogenic integration technology are making contributions to pushing the state of the art. Through such

research efforts as vibration mitigation through improved compressor design with Oxford University, improved flexure bearing design with the Aerospace Corporation, and potential accelerated testing approaches with the Ukrainian Institute for Low Temperature Physics many aspects for basic improvement of emerging cryocooler technology have been explored. Quantifying the lifetime and reliability of long life cryocooler technology is elusive. Many of

the mechanical refrigerators that have been developed or are under development are usually unique or have very low production numbers. Additionally, designs mature and evolve from cooler to

cooler to accommodate new improvements or to meet customer specifications. These changes effect the design heritage and any prediction of cryocooler reliability. A large unknown in the

useful lifetime prediction for cryocooler performance is the long-term degradation components that are observed only over thousands of hours of operation. As a necessary component to understanding the problems associated with developmental cryocooler technology, AFRL in-house laboratory characterization and endurance evaluation of engineering design model and first-of-a-kind protoflight cryocoolers has a significant role in the lifetime and reliability database for emerging technologies. This information is invaluable as a tool for

spacecraft developers to understand the lifetime and reliability potential of a candidate cryocooler technology and establish a heritage database for cryocooler families. AFRL has also made direct contributions to improvements in follow-on designs through the feedback of performance and endurance data to industry developers. THE CRYOCOOLER RELIABILITY WORKSHOPS The Air Force Research Laboratory Space Vehicles Directorate (AFRL/VS) at Kirtland Air Force Base and the Space Technology Advanced Research Center (STAR) at Georgia Institute of

Technology in Atlanta sponsored a series of workshops on the reliability of cryocoolers for space applications. The First Cryocooler Reliability Workshop was held on 17 – 18 September 1998 in Albuquerque, New Mexico, and the Second Cryocooler Reliability Workshop was held in Manhattan Beach, California on 23 - 24 August 1999. These workshops are a mechanism for involving the cryocooler community in developing

general approaches to reliability problems. In the workshops, a subset of experts met in private for two days to identify and prioritize cryocooler reliability issues and to recommend approaches to resolve the issues. By concentrating on the technical questions in a cooperative atmosphere with a breadth of viewpoints, new insights and solutions were encouraged. Participation in the workshops was determined by invitation. The two workshops drew participants with knowledge and interest in cryocooler reliability from government, industry and academia. Civilian as well as military interests were well represented. The first workshop drew 25 attendees and the second 35. (A list of the organizations represented at the two workshops is contained in Table 1.) The diversity in participants was very important to the goals of the workshops.

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The workshop process began with introductory, plenary sessions to define the scope of the workshop and to establish participant interest areas. The workshops then broke into two or three smaller parallel sessions. Each of these was charged to review a particular subset of the problem, to prioritize issues, and to propose suggested approaches to the issues. The last afternoon of the workshops was devoted debriefs by the leaders of the parallel sessions. Abbreviated workshop agendas are shown below in Tables 2 and 3.

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WORKSHOP RESULTS

The First Workshop concentrated solely on mechanical reliability of active refrigeration. As one of the recommendations from that meeting, the Second Workshop expanded the scope to include electronic and software reliability as well. A summary of each workshop including the session debriefs was produced on CD-ROM along with the available presentations. Information is also available at http://www.star.gatech.edu under Workshops.

Cryocooler Mechanical Reliability Issues The most significant cryocooler mechanical reliability issue from both workshops was concerned with the perceptions of the acquisition and user communities regarding cryocooler maturity and suitability for space flight. Several presentations in the plenary session of both meetings addressed this issue, and the parallel sessions revisited the question in several forms. The majority of system designers and users have not recognized significant improvements in cryocooler mechanical reliability. Extensive experimentation and performance evaluation is the proposed mechanism for dispelling ignorance and doubt. The workshops spent significant energy in considering how to characterize cryocoolers, how to capture lessons learned from experiments and qualification testing, and how to distribute the results widely. Much of current experimentation and testing will characterize known mechanical failure modes including failures from contamination, wear, leakage, mechanical failure, and other sources. To continue improvements in mechanical reliability, the unknown failure modes are more important to identify. While it is possible to design for and screen for known failure mechanisms, an adequate screening for the unknowns must also be devised. Both cryocooler system and component level screening are necessary. In addition to the “screening” type of testing, such as environmental qualification testing, longterm endurance evaluation plays a major role in identifying and tracking long time constant performance degradation. Endurance evaluation is difficult due to the unavailability of accepted accelerated testing methods that allow the cooler to be operated in a fashion that provides reliability data for long term operation in a greatly reduced time period. However, the nature of current cryocooler technology apparently prohibits this type of test. Cryocoolers must be run over thousands of hours, usually in thermal vacuum chambers with sophisticated instrumentation, to track performance and provide user confidence in the design. Continuing to provide for this type of performance evaluation is essential to continue to refine the reliability database for space cryocooler technology. To verify units and designs, there is a need for full space qualification units. For the low number of builds involved in most cryocooler efforts, the cost of independent qualification units is usually prohibitive. With the expected move to manufacturing lines for cryocoolers, the need to develop qualification unit test standards and procedures will become vital. Two summary recommendations concerning mechanical reliability resulted from the workshops: (1) Develop good mechanical design principles and practices Good design principles and practices should be documented, controllable, and repeatable. Some examples are the TRW High Efficiency design for manufacturing compressors and a paper on contamination lessons learned presented at the Second Workshop. Properly implemented, these standards would result in such features as larger, more predictable design margins, less unit-to-unit performance variation, and streamlined process improvement. (2) Continue testing and experimentation programs Identifying new failure modes and characterizing reliability requires continued, well constructed

experimental and qualification testing programs. A continuation of long-term endurance evaluation on coolers is essential to these goals. Cryocooler system and component level screening testing need to be expanded.

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Cryocooler Electronic Reliability Issues Most electronic reliability issues for spacecraft are not specific to cryocoolers. The major cryocooler-specific electronic reliability issue is the reliability of the controller. The specific issues that were identified include the lack of radiation-hardened parts for some functions, the reliability of sensors, the relatively large part count, the effects of electromagnetic interference, and the complexity of the control objectives. Approaches to resolving these issues were identified from several alternatives. Since the community is moving to second generation designs, the timing is right for the implementation of some of these recommendations:

(1) Develop standardization across design teams The community needs to move toward standardization and integration. The standards must maintain design flexibility to meet unique system requirements, so the question of the acceptable level of standardization must be faced. Once defined, the effectiveness of standard building blocks for circuit and package integration needs to be confirmed. Feedback with system and mechanical designers must be included in the process. The presentation by Jim Lyke at the Second Workshop included a program with many of the desired properties. (2) Define a standard digital data interface for cryocoolers By identifying potential common core standards defined by digital inputs/outputs, there should be reliability benefits for all subsystems, not just the electronic controller. (3) Develop a cryocooler-unique testing standard for controllers There are no data specifically concerned with the reliability of cryocooler controllers. A testing standard and a testing program are needed to characterize cryocooler electronics reliability.

Cryocooler Software Reliability Issues The important cryocooler software Issues sorted into two areas: organizational and technical

issues. The organizational issues are concerned with how computer scientists/software engineers are integrated into the cryocooler team. Generally, the software community is not directly a part of the mechanical or electronics communities. Consequently, the software developers usually lack enough operating experience to know the level of functionality required, they have very little experience with space cryocoolers, and builds are so infrequent that there is little continuity of experienced personnel and a loss of lessons learned. The technical issues that arise are partly the result of the lack of experience by the software developers. The requirements are often overly complex, there is not enough margin in the requirements or the design, and the differences between an initial design with commercial parts and a final

design with radiation-hardened parts are often overlooked. The inclusion of software reliability in the Second Workshop was important to initiate a discussion of this issue. The discussion was informed by personal experiences of cryocooler developers, but did not include the perspective of software developers, particularly ones who specialize in software reliability. An understanding of software failure was implicit in the comments, but a definition of software reliability did not emerge. The most detailed recommendations concerned methods to avoid complexity in software as a means of improving software reliability, without a consideration of the potential effects on overall cooler effectiveness and reliability. The discussions led to suggestions that are collected below under three major headings: (1) Integrate software developers into cryocooler development programs Integrating software workers more deeply into cryocooler development programs could resolve some of the issues. Keeping a record of lessons learned in cryocooler software development would aid in getting new workers up to speed. These would explicitly include reference to margins in computing capability. (2) Minimize the need for software Reducing the dependence on software could contribute to resolving all software problems, including reliability issues. This might not always be possible, but some activities that are implications of this approach were explored in the Second Workshop.

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(3) Establish software testing benchmarks and standards A set of cryocooler software testing methods and standards would enable the community to track progress and predict reliability. This approach will require software reliability expertise generally lacking in the cryocooler community.

General Workshop Recommendations

The perspective of the cryocooler community, as represented in the workshops, is very positive. Advances in the last decade have raised cryocooler mechanical reliability to the levels required for spacecraft application, to the point of making electronic and software reliability as much the issue as mechanical reliability. The cryocooler community has resolved a difficult mechanical

research and development issue and is eager to see the results applied on spacecraft. Several flight programs have baselined cryocooler technology and over the next several years reliability predictions, system integration issues, on orbit problems (or lack thereof), and improvements in the state of the art of technology will affect research and development and quantification of cryocooler reliability. Consequently, two major needs were identified by the workshops. The first is to demonstrate

and to advertise the improved mechanical reliability by developing and distributing supporting test data from ongoing performance characterization and endurance evaluation activities. The second is to attack the remaining reliability issues, not only in the mechanical systems, but also in electronics and software. The general recommendations of the workshops, summarized in Table 4 as action

items for specific areas of the community, were intended to address these two major needs. SUMMARY

The Cryocooler Reliability Workshops have proved to be very useful for determining critical issues in space cryocooler reliability. Expanded awareness of these issues from the technology

developer and user communities will allow more efficient development and technology investment

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strategies across industry and government. Comprehensive use of active cryogenic refrigeration in space through a growing number of operational flight programs is pushing the need for defined and quantifiable cryocooler reliability. It is recognized that reliability must be pursued as an integral part of cryocooler development. ACKNOWLEDGMENTS

The authors would like to acknowledge the significant and essential contributions of the workshop participants, the briefing presenters, and the facilitators toward making these workshops

a success.

Protoflight Spacecraft Cryocooler Performance Results Kenneth Price1, John Reilly2, Nandu Abhyankar2, Ben Tomlinson2 1

Raytheon Systems Company El Segundo, CA, 90245, USA 2

Air Force Research Laboratory Albuquerque, NM, 87117, USA

ABSTRACT

The Protoflight Spacecraft Cryocooler (PSC) is a flight-qualified Stirling cryocooler that delivers 1.2 W refrigeration at 35 K and 3 W refrigeration at 60 K. This Oxford-class unit employs three finger tangential flexures in the compressor module. These flexures have been shown to provide smoother, lower vibration piston motion than previously obtained with conventional three finger spiral flexures. Acceptance Tests validated the required performance capabilities and Qualification Tests validated the cooler for flight. Acceptance Tests included performance mapping at rejection temperatures from 275 K to 325 K, residual vibration measurements for each module, temperature stability under various conditions, cold tip motion, cold tip side load capacity. Qualification Tests included three-axis random vibration, thermal cycling, hermeticity and EMI/EMC tests. The 12.5 kg cooler was delivered with a brassboard command and control module that provides simultaneous temperature and Adaptive Feed-Forward vibration control. The PSC with brassboard electronics have been delivered to the Air Force Research Lab in Albuquerque, where additional testing has validated the results obtained at Raytheon. INTRODUCTION Air Force Research Laboratory awarded the PSC Program1 to Raytheon Company to build, test, and deliver a flight qualified Thermo-Mechanical Unit (TMU) and a brassboard Electronic Control /Power Conditioner (EC/PC). Objectives included improving thermodynamic performance, reducing weight, and upgrading hardware to flight quality in comparison to the previously built 65K Standard Spacecraft Cryocooler Engineering Model. The PSC system has been successfully assembled, tested and delivered. Acceptance Tests performed at Raytheon and subsequently repeated and expanded at the AFRL and JPL demonstrate that the PSC TMU performance requirements have been met or exceeded in all categories. In particular, outstanding performance has been achieved over the 25K to 120K operating range. Up to 3W refrigeration can be delivered at 60K and up to 1.2W at 35K. The efficiency of the cooler at 35K is a remarkable 85.9W/W, measured at the motor inputs. The large data base now established by AFRL, Raytheon, and JPL will provide a definitive benchmark against which to measure the performance of the PSC during life testing planned to start Cryocoolers 11, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2001

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Figure 1. Protoflight Spacecraft Cryocooler.

at AFRL in late 1999. Additional cooler characterization will be performed by AFRL prior to life

test. Emphasis will be placed on the PSC’s ability to continue operating efficiently over the full rejection temperature range.

CRYOCOOLER SYSTEM DESCRIPTION

The PSC TMU is a single stage cryocooler housed in two modules connected by a compliant steel transfer line. See Figure 1. Refrigeration is produced at the tip of the cold cylinder, where the user’s thermal interface is a cylindrical copper projection. Waste heat is removed from the expander module at a flange located at the base of the cold cylinder. Compressor waste heat is removed from a flange at the center of the housing. The cryocooler weight requirement is specified as the combination of physical weight plus a weight penalty† proportional to power consumption. Penalty weight is defined as 0.3 kg/W times the motor power drawn by the TMU when supporting 2W at 60K and rejecting heat to 300K. This combination is required to be less than 33.0 kg. The PSC compressor module weighs 7.0 kg and the expander, 5.5 kg for a TMU physical weight of 12.5 kg. This is a 30% reduction compared to Raytheon’s predecessor programs such as the 65K SSC.2 Since the PSC consumes 57W at the specified operating point, the weight penalty is 17.1 kg. Therefore, the combined physical and penalty weight is only 29.6 kg, which are 3.4 kg (10.3%) below the allowed maximum. Each module is internally balanced for optimal control of residual vibrations. The compressor employs two pistons working in opposition against a common compression chamber located at the center of the module. The expander employs a single displacer piston to produce refrigeration. A matching mass driven in opposition dynamically balances the displacer piston. Residual vibrations produced by each module are corrected by the electronics via an Adaptive Feed Forward (AFF) algorithm operating from feedback obtained by monitoring the three load cells located at each module’s three symmetric mounting points. The PSC TMU design retains significant design heritage to the previously built 65K Standard Spacecraft Cryocooler (SSC) and its successor, the Improved Standard Spacecraft Cryocooler (ISSC). This decision was implemented to retain legacy to the extensive life test data obtained on ISSC unit 1 and 2 (over 23,000 hours each) and flight test data from ISSC unit 3 (NASA Cryogenic System Experiment, STS-51). Design modifications from these earlier machines were made to improve †

Weight penalty is a measure of the weight impact a cryocooler’s power consumption imposes on a spacecraft. It is an estimated weight of hardware required to supply electrical power and to dissipate waste heat.

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Figure 2. Cross-sections of the PSC compressor (top) and expander (bottom).

thermodynamic performance, reduce residual vibration, reduce weight, and increase ruggedness and reliability. Figure 2 shows cross sections of the two modules. The compressor is nominally 418mm long and 103mm in diameter. The expander is nominally 384mm long and 124mm in diameter. The cold cylinder extends 102mm beyond the expander’s warm-end waste heat rejection surface. This arrangement typically enables the cold cylinder to project into a sensor cold volume and places it beneficially adjacent to the user’s load. QUALIFICATION AND ACCEPTANCE TEST PROGRAMS

Qualification and Acceptance Testing was performed as specified in the Technical Requirements Document and the cooler has been shown to meet or exceed all requirements. Qualification Tests included thermal cycling between 233K and 327K and 12.9 Grms random vibration on each of three orthogonal axes. To verify alignment stability of the pistons, sticktion hysteresis loops were recorded before and after each vibration test and before, during, and after each thermal cycle. Housing gas hermeticity was verified by helium leak testing after final assembly and again prior to delivery. Acceptance Tests performed at Raytheon included a variety of thermodynamic performance characteristics, measurement of cold tip side load capacity, hermeticity, cold tip motion, and temperature stability and repeatability. Self-generated vibrations and cooler EMI were measured at the JPL. Most of these tests have been repeated at the AFRL with virtually identical results.

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Figure 3. PSC thermodynamic performance map at 300K rejection temperature.

Figure 4. PSC load lines at three rejection temperature and at intermediate power levels are very consistent.

CRYOCOOLER THERMODYNAMIC PERFORMANCE

Figure 3 is a performance map covering the cold tip temperatures from 35K to 120K, all with rejection temperature at 300K. At the design point, 2W at 60K, the specific power is 28.5 W/W, referenced to motor power. The efficiency relative to Carnot is 14.0%. (Carnot efficiency is defined as measured COP over Carnot COP between rejection and load temperatures expressed as per cent.) In addition, the PSC can support up to 3.0W at 60K, which is a 50% performance margin over design requirement. Figure 3 also shows test data taken at the AFRL overlaid on Raytheon test data. The AFRL data is shown as large circles, Raytheon data as solid lines. The AFRL also took additional data at 90K, shown as a dotted line. The close correlation between the two data sets indicates both accuracy and repeatability of the data despite the use of different cold tip temperature sensors, load heaters, and instrumentation. Raytheon took its data in late 1997 and the AFRL data was taken in 1999. Housing hermeticity has been continuously verified throughout the 1 1/2 years since delivery.

Although the design specification did not require performance below 60K, the single stage PSC delivers up to 1.2W at 35K for only 101 W of compressor motor power. The specific power at that point is 84.2 W/W, corresponding to a robust efficiency relative to Carnot of 9%. This high level of performance is proving to be the most significant accomplishment of the cooler. Finally, Figure 4 shows three load lines for the cooler at the piston strokes corresponding to rated power (i.e., 2W at 60K.) The load lines correspond to rejection temperatures of 275K, 300K, and 325K. Significantly, the PSC delivers almost identical refrigeration at each rejection temperature and achieves no-load temperature just above 20K. CRYOCOOLER SELF-GENERATED VIBRATION

The compressor’s self generated vibration is required to be less than 0.2 Nrms at any frequency, and the expander’s, less than 0.1 Nrms. Vibrations from each module were measured independently by the JPL to gauge the vibration contribution by each. Data are shown in Figures 5 and 6 demonstrate that the cryocooler’s internally generated vibration was well under the government’s maximum limit. The test data shown below include vibrations with and without Adaptive Feed Forward (AFF) vibration control. The AFF control was only active on the drive axis in each module (the Z-axis.) Cross axis vibrations (X- and Y-axes) were not controlled. Due to processor speed constraints (the breadboard uses a Motorola 68000 operating at 8 MHz) only the first six harmonics in each module could be controlled within a reasonable computation cycle time. Although a higher speed processor would enable more harmonics to be controlled, this was found to be unnecessary to achieve performance requirement. The performance of the compressor’s three finger tangential flexures, Figure 7a, demonstrated the predicted low vibration behavior of the Aerospace Corporation patented design. The tangential

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Figure 5. PSC compressor self-generated vibration at 2W heat lift at 60K. Cross-Axis vibrations on left, drive-axis vibrations on right. White columns control OFF, dark columns, control ON.

Figure 6. PSC expander self-generated vibration at 2 W heat lift at 60 K. Cross-axis vibrations on

left, drive-axis vibrations on right. White columns control OFF, dark columns, control ON.

Figure 7a. Compressor suspension flexures.

Figure 7b. Expander suspension flexures.

flexures were expected to displace more smoothly than conventional spiral flexure designs used in most Oxford class coolers, resulting in a significant reduction in cross-axis self-generated vibration.

The expander module employed previously designed twelve finger spiral flexures, Figure 7b. This is because data collected from earlier Raytheon coolers indicated cross axis vibrations were low and did not require further reduction. This was validated, as shown in the test results. COLD TIP SIDE LOAD CAPACITY

In order to achieve reliable Stirling cryocooler operation, the oscillating displacer and cold cylinder must not contact during operation. Therefore, after assembly with proper alignment, the

cold cylinder must be stiff enough to carry incidental side loads without deflecting into the moving displacer. The PSC was required to support a side load greater than 6N in any direction. To validate this requirement, the PSC was designed with displacer electrically isolated from the cylinder so that contact resulting from side loads can be detected by a simple conductance measurement. This is a

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very sensitive measurement and can be performed during static test at ambient and during operation at any temperature. Measurements show this requirement was exceeded by more than a factor of three with a minimum side load capacity of 19.6N. CORRELATION OF TEST PROCEDURES

The close correlation between Raytheon and AFRL thermodynamic performance data was obtained because both teams agreed to the same test methodologies, clear definitions of test parameters and used virtually identical instrumentation. Three key examples include: 1. The heat rejection temperature was defined as the temperature of the liquid cooled heat sink at the cooler’s waste heat thermal interface, as measured on the user’s side of the interface. Significantly, temperature sensors were positioned along the waste heat flow path. This includes the effect of contact thermal resistance at the interface and rejection temperature is measured at a location that is of most significance to the user. 2. Raytheon and AFRL used identical cold end instrumentation and radiation shielding. Cold end instrumentation was mounted to a copper ring clamped to the cryocooler cold tip to represent the user’s side of the cold tip thermal interface. Two calibrated silicon diode temperature sensors and a resistive heater were mounted to the ring so that measurements of cooler performance included the adverse effect of the interface thermal resistance. Again, the temperature sensors were located at the point of greatest significance to the user: on the user’s side of the thermal interface contact resistance. 3. Power consumed by the motor cables was deducted from measured power. This was done because a user’s cable resistance is likely to differ from the test cables. To enable a user to account for cable power, maps of current vs. thermal load at constant load and rejection temperature were developed. Figure 8 is the current map corresponding to the performance map shown in Figure 3. CONTAMINATION CONTROL Long-term cryocooler reliability relies on execution of quantifiable charge gas contamination control procedures. For the PSC and other Raytheon coolers, a charging procedure developed with the AFRL has been implemented that consistently achieves the required low levels of contamination. To confirm adequate control, gas samples are taken periodically throughout the charging procedure and sent to Pernicka Corporation in Fort Collins, CO for mass spectroscopy analysis. By agreement with AFRL, Raytheon follows Pernicka instructions regarding sampling hardware and procedures and reported analysis results are accepted as definitive. Contamination levels of the various gases, both condensable and non-condensable, are reported in per cent by volume (which is directly convertible to parts per million.) From this data and knowledge regarding total gas mass within the cooler, the contaminant levels are calculated as a solid volume. Experience indicates that

Figure 8. The motor current map corresponding to the performance map shown in Figure 3 is developed to aid the system integrator in selecting cable design parameters.

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the PSC and similar size coolers can tolerate at least 5 cubic millimeters. At this level of contamination, no adverse effects are observed. Table 1 shows final gas charge analysis. The table combines gas specie and volume, provided by Pernicka, with calculations of specie mass and volume performed by Raytheon. The contamination level is only 2.23 cubic millimeters. No contamination effects have been observed during testing at Raytheon, JPL, or AFRL.

PERFORMANCE VERSATILITY The PSC’s wide performance versatility is one of the program’s major accomplishments. In 1992 the AF was focused on the need for a cryocooler system that could efficiently supply 2 watts of cooling at 60K. During the course of the program, the AF interest in longer wavelength focal plane arrays extended refrigerator performance requirements to the 35K range. To address this

need, the performance and acceptance test structure was expanded to map the PSC’s maximum performance envelope. It was determined that the PSC, without design modification, outperformed both in efficiency and cooling loads, existing cryocoolers specifically designed to perform in the 35K region. At 35K, the PSC was able to cool loads over one watt with thermo-mechanical efficiencies under 90w/w from a reject temperature of 300K. In addition, the system was able to deliver

up to 3W at 60K, a 50% increase over the 2W cooling load requirement. COMMENTS ON IPT PROGRAM MANAGEMENT

The PSC Integrated Product Team (IPT), including the government Project Officer and key Raytheon personnel, encountered and solved a number of problems during the course of the program. This successful IPT provides an excellent model that can be used by the Air Force to improve program management in the future. Program success can be directly linked to the restructuring of the Statement of Work (SOW) and Technical Requirements Document (TRD) in early 1995. This

was driven by both technical problems and limited funds following the completion of the Critical Design Review (CDR) in late 1994. The result of these efforts is a system that has exceeded all initial government requirements. The system was completed with additional funds that was about 50% of the funds spent up to and including the CDR. The initial funding had been almost exhausted due to two separate developmental problems. The first problem was excessive vibrations transverse to the drive axis caused by non-uniform displacement of the compressor’s three finger spiral flexure bearings. The solution to this problem was to replace the spiral flexures with the newly developed Aerospace patented three finger tangential flexures. The tangential flexure design was found to displace smoothly, resulting in far greater transverse motion stability. The changeover required significant redesign of the flexure support system and the drive motor, which was completed by late 1994. The new flexures reduced transverse axis vibrations by a factor of about 40 and reduced axial vibrations by a factor of at least 10. The second problem was the high cost and projected unsatisfactory performance expected to be obtained from axial vibration control feedback electronics. Specifically, the high-speed processor

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and related circuitry to be provided by SCI was intended to perform the numerically demanding narrow bandwidth vibration control algorithm in near real time, but the project suffered from cost and technical problems. The technical limitation was the inability to correct vibrations above the third harmonic. The cost and performance issues compromised the program. The technical problem was solved in early 1995 when the government and Raytheon agreed to replace the narrowbandwidth vibration control algorithm with a feed-forward vibration control algorithm that could be implemented in simpler, lower cost electronics. This system could separately correct excessive vibration levels in the first six harmonics, which was sufficient to achieve program requirements. The feed-forward algorithm was also found to more effectively reject electrically generated feedback noise because it sampled vibrations over several complete machines cycles before calculating a solution. The cost problem was addressed by rescoping electronics maturity from prototype to an engineering design model (EDM) implemented within a commercial laboratory grade rack. DISCUSSION OF SELECTED TECHNICAL PROBLEMS

Several problems arose during the final development of the PSC system that were resolved to meet or exceed program requirements. Each is discussed below: a. One compressor motor coil shell buckled during pressurization testing in Dec of 95. The motor coil assembly is comprised of a thin wall titanium bobbin on which the coil is wound and a thin wall titanium cover shell welded over the coil to contain potential gas contaminants. Both bobbin and shell are thin wall elements to minimize eddy current losses and to

minimize magnet gap distance. Coil lead wires pass through the bobbin via two glass sealed connector pins. In order to resist buckling in the pressurized cryocooler, the titanium coil structure must be well supported on the inside. Encapsulating the coils in sufficient filler material to make a line-to line assembly between titanium and filler provides the required support. This coil had been built up with insufficient filler, resulting in a gap between coil and titanium cover large enough to result in buckling when pressurized. A new coil with the proper assembly tolerances was obtained and the program continued. b. A piece of the tip of the expander’s titanium piston broke off during early testing in Jan 1996. The location of the fracture was at a hole near the cold end of the piston that had been cut by electrical discharge machining (EDM.) A series of these holes is used to secure a clip that retains the regenerator packing inside the piston. The primary cause of the fracture was that the EDM process melts and recasts titanium in the zone immediately adjacent to the hole, resulting in embrittlement at the edges of the hole. The recast material had not been removed after EDM. To correct the problem, a stainless steel displacer piston was substituted for the titanium piston. Stainless steel pistons had previously been successfully used in all previous Raytheon Oxford class coolers. The subsequent thermo-mechanical performance to date has been outstanding. c. The most critical problem encountered from a time and cost standpoint was loss of compressor and expander housing hermeticity, also detected in Jan of 96. Hermeticity was lost when the electrical connectors were electron beam welded to the housings. These connectors contain multiple glass-to-metal 16-guage pins that are used to transfer motor power and position sensor leads to/from the compressor and expander modules. Localized heating by the electron beam generated thermal stresses that caused the glass to separate from the pins and/or shell, resulting in leakage paths. Fortunately, the problem had already being solved for a set of coolers being produced for the SBIRS Low Program. For these coolers, an inconel connector shell was developed to replace the previously used steel shells. Inconel welded successfully because it has a lower coefficient of expansion, which reduces stress during nonuniform heating, and distributes heat more uniformly, which reduces the thermal strain potential during welding. Reworking the PSC to incorporate these connectors required complete disassembly of both modules and remanufacture of several housing components, which extended development schedule and increased costs substantially. d. The high cost to build flight-qualified electronics to drive and control the cooler led to rescoping of the contract to build laboratory grade rack electronics. The electronics were adapted from

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a previously developed design and used a combination of new hardware, spare hardware from a previous build, and some hand wired circuitry. Unfortunately, the remaining funds for electronics were barely sufficient for the task, leading to workmanship and quality control problems. Correcting the problems caused schedule slips and testing delays. e. Access to key personnel proved to be a problem during periods of SBIRS Low program activity that required experienced cryogenics engineers and technicians. SBIRS Low’s higher priority and tight schedule caused the PSC program schedule to slip 6 months while selected personnel were redirected to build and integrate cryocoolers for the SBIRS Low Flight Demonstration System. SUMMARY

Despite a significant number of problems and setbacks, the PSC delivered in 1997 has proven to exceed all cryocooler performance requirements. As described, all difficulties were resolved through the close working relationship of the entire IPT, particularly between the Raytheon Program Manager and the AF Project Officer. The excessive cross axis vibration in the compressor module caused by the original three-finger spiral flexure design was corrected by using tangential flexure bearings. Two control electronics problems were solved the first time was with the SCI processor development, which significantly impacted both cost and schedule. The second incident was the repair and corrections to the engineering model electronics at the time Raytheon was preparing for cryocooler acceptance testing. Other problems such as the motor coil buckling and the tip of the expander piston breaking off could be expected on any program; however the loss of personnel to the SBIRS Low program for the good part of a year was never anticipated. What seemed like an equally crushing occurrence was the leakage problem with the electrical connectors. This problem took over four of five months to resolve. What cannot be minimized is the outstanding product that has resulted from this endeavor. Not only does the cryocooler have the capability to cool greater than 50% of the original contracted cooling load at 60K with outstanding efficiency (under 30 w/w), but the cooler has demonstrated a great cooling load versatility over a wide cooling spectrum. This spectrum goes far beyond what was called-out in the original contract. The most noteworthy of all the cryocoolers capability is the outstanding performance at 35K. It is able to provide over a watt of cooling for less than 90w/w. ACKNOWLEDGMENTS

This work was sponsored by the Ballistic Missile Defense Organization. The Air Force Research Laboratory, Albuquerque, NM, managed the project. D. Johnson, S. Collins and P. Narvaez of the JPL and T. Pollack and M. Kieffer of Raytheon made significant contributions in acquiring the data reported herein. REFERENCES 1. Price, K.D., Barr, M.C. and Kramer, G., “Prototype Spacecraft Cryocooler Progress,” Cryocoolers 9, Plenum Press, New York (1997), pp. 29-34. 2. 65 K Standard Spacecraft Cryocooler Program Final Report, Contract #F29601-89-C-0082, Hughes Aircraft Company, Electro-Optical Systems, El Segundo, CA, November 1995.

Characterization of Raytheon´s 60 K 2W Protoflight Spacecraft Cryocooler N. S. Abhyankar2, C. H. Yoneshige1, B. J. Tomlinson1 and J. Reilly1 1

Air Force Research Laboratory Kirtland AFB, NM, USA 87117-5776 2 Dynacs Engineering Co. Albuquerque, NM, USA 87106-4266

ABSTRACT The Air Force Research Laboratory Cryogenic Cooling Research Facility is supported by the Ballistic Missile Defense Organization (BMDO) and the U.S. Air Force SBIRS Low Program Office. It was created to characterize the thermodynamic performance and long life

potential of cryogenic cooling technologies developed by various defense industry contractors. The objectives of the characterization process are to explore the cooler’s ability to perform at its design point and to map its range of thermodynamic operation. This provides a detailed performance envelope for alternative space applications and aids in providing valuable feedback to cryocooler developers. This paper provides an overview of the AFRL characterization of the Protoflight Spacecraft Cryocooler (PSC) built by Raytheon Systems Co. The PSC is a split Stirling cryocooler with dual, opposing motion compressors and a displacer, which uses a mass balancer to reduce vibration. It is designed to lift a heat load of 2W at 60K, with a nominal rejection temperature of 300K. The characterization of the PSC involved a series of experiments, including cool-down to lowest temperature, design point verification, parameter optimization, stiction tests, long term-stability, transient thermal response (due to orbital temperature variation) and a proto-qualification thermal vacuum test. The performance map is charted providing a graphical display of important parameters including specific power, calculated as input power per watt of cooling. The off-nominal performance evaluation included lifting 1.2W at 35K and 6W at 120K, while operating within recommended total input power and stroke boundaries. INTRODUCTION The Air Force Research Laboratory Cryogenic Cooling Research Facility (CCRF) is engaged in the long-term functional and thermal evaluation of various types of cryocoolers. The thrust of operations is to characterize these coolers and verify their design requirements, to provide feedback to cryocooler developers for further improvements in future designs and to allow potential users to evaluate and compare different cryocoolers for specific space flight missions. The CCRF, with assistance from the AFRL and the Ballistic Missile Defense Organization (BMDO) is equipped with vacuum chambers, instrumentation, material resources

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and personnel to carry out short and long-term evaluations of the cryocoolers. In the facility, a cryocooler can undergo numerous years of continuous operation to identify and evaluate its efficiency, reliability and performance changes over time.

Raytheon’s Prototype Spacecraft Cryocooler (60K PSC) is a split Stirling, long life (7+ years) space cryocooler designed to lift 2 W of heat at 60K. This cryocooler is a protoflight unit intended to demonstrate the advancement of Stirling cryocooler technology for use in long life strategic space applications. It is designed to operate for 7 to 10 years on orbit (continuous duty cycle) with very high reliability (>0.95). The motive for developing the 60K PSC was provided by the Space Based Infrared System Low (SBIRS Low), BMDO and the AFRL management team for cooling infrared (IR) sensors aboard surveillance spacecraft. The 60K PSC incorporated lessons learned from Raytheon’s internal research and development coolers (the Improved

Standard Spacecraft Cryocoolers) and the Air Force sponsored Standard Spacecraft Cryocoolers. The 60K PSC has other potential cooling applications including chilled electronics and superconducting devices.

DESCRIPTION OF THE CRYOCOOLER The 60K PSC incorporates the use of both spiral and tangential flexure bearings. The compressor unit is a dual dynamically balanced gas piston design, which works on a common compression space. The expander is balanced using a balance mass operating in an opposed motion to the expander piston, thereby cancelling vibration. Clearance seals with liners ensure no metal-to-metal rubbing or wearing contact along all running clearances. The motor coils are

canned to mitigate the potential for wires or potting material to out-gas and contaminate the helium working fluid. The cryocooler is operated using a single control electronics rack that is convectively cooled internally. The cryocooler has been validated for operation with a rejection temperature (TR) between 2°C (275K) and 52°C (325K). The control electronics are not qualified for space flight in their present form.

Specialized Mounting Fixtures As shown in Figure 1, the mechanical fixtures were specially designed for this experiment to allow the cryocooler to be mounted inside a 36” thermal vacuum chamber. The heat rejection

surface is provided by standard cold plates, which have chiller fluid transfer lines for cooling and electrical coils for heating. The rejection temperature at the interface is automatically maintained by temperature controllers switching between cooling and heating.

Figure 1. 60K PSC integrated with a 36” thermal vacuum chamber.

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Cold End

There are two Lakeshore silicon diodes attached to the cold end of the PSC. One of them is directly connected to the PSC controller electronics, which allows the use of an automatic temperature controller. The other diode is used for independent monitoring and performance verification through the PSC data acquisition software. The calibration curves, which range from 4K to 330K, were inserted into the temperature controller instrument to ensure accurate measurements. The diodes and the heater are wired with cryogenic grade, low conduction twisted wires. The entire cold finger assembly is shielded from the vacuum environment and from the expander base by multi-layer insulation (MLI). The MLI is isolated from the cold finger by baffles. This set-up helps reduce conductive and radiative parasitic losses and give a more accurate estimation of the cryocooler’s performance. Instrumentation for the Controller, Data Acquisition and Environmental Control The instrumentation for the PSC is separated into three stacks: one set for controlling the cryocooler itself, another for data acquisition (including power and temperature measurements) and a third for monitoring and controlling environmental parameters (such as heat rejection temperature and vacuum pressure). The instruments are controlled by LabVIEW™ via IEEEGPIB or RS232 communication interfaces. An automated data acquisition system records the experiment data and environmental conditions. The data acquisition is also provided by LabVIEW™ software.

CRYOCOOLER CHARACTERIZATION The PSC was integrated in a 36” thermal vacuum chamber and connected to its peripheral

equipment according to procedures established by AFRL and Raytheon personnel. A test readiness review was held, where laboratory personnel presented their understanding of the performance and safety requirements and showed that the experiment stand was adequate to carry out the characterization activities. A series of experiments were performed as an acceptance evaluation before the more elaborate characterization experiments were undertaken. The characterization plan was created, which established the sequence of operations according to an experiment matrix. The characterization plan included the following evaluation activities: 1. Optimization Evaluation 2. Characterization Load Lines/Performance Mapping at Different Heat Rejection

Temperatures 3. Temperature Stability at Different Heat Rejections (With and Without Heat Load) 4. Transient Thermal Response 5. Cool-down at Different Heat Rejection Temperature (With and Without Heat Load) 6. Parasitic Heat Load Determination (Off State Conduction) 7.Thermal Cycle/Thermal Vacuum Optimization Evaluation

The performance of the cooler depends on various parameters such as compressor stroke length, expander stroke length, the phase angle between the compressors and expander, operation

frequency, piston offsets, and heat rejection temperature. To ensure the most efficient cryocooler performance, these parameters must be optimized. In the laboratory, optimization is carried out

by varying one parameter at a time while all other operation parameters and conditions are kept the same. The cold end temperature and input power are used as performance evaluation parameters. These are recorded for a set of heat loads while obtaining the sensitivity of the performance to the chosen parameter. The dimensional effect of the temperature and power values is avoided by defining the efficiency factor in terms of a percent Carnot efficiency, which is given by the ratio of thermodynamic coefficient of performance (COP) to the Carnot

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Figure 2. Percent Carnot COP for expander stroke length variation.

efficiency. (The thermodynamic COP is defined as the total input power minus the I2R losses divided by the cooling power expressed as a heat load.) As an example, the variation of the expander stroke length is presented in Figure 2. The fixed parameters were as follows: Compressor A stroke length = 7.5mm, Phase Angle = 72.4°, and Operation Frequency = 35 Hz.

These values were carefully monitored by changing the command signals. Vibration control was turned off and the heat rejection temperature was 300K. Data points were obtained for discrete heat loads between 1W and 5W. As seen in the figure, the percent Carnot COP for an expander stroke length of 2.7mm is

higher than that obtained for 2.5mm. The maximum expander stroke length at a rejection temperature of 300K is about 3.2mm and decreases to 3.0mm at a rejection temperature of 325K. In order to maintain a large enough safety margin and to lower the input power, for a fixed stroke length on compressor A, 2.5mm was chosen as the nominal expander stroke length. If a lower

heat rejection temperature were used, the expander stroke length could be slightly increased

without any significant increase in input power. Characterization Load Lines/Performance Mapping

A baseline depicting input power and cold end temperature changes over a range of heat loads provides a picture of the cryocooler’s nominal performance, which can be tracked periodically during a long-term endurance evaluation. The off-nominal performance map shows the cryocooler’s entire performance with different heat loads, stroke lengths, and lowest cold end

temperature. The limits of the off-nominal performance evaluation are based on the cryocooler’s upper limit of input power (~100W for PSC) and the highest recommended operation temperature (~150K for PSC), which were established by payload considerations and the properties of the motor coils. The data is obtained as a series of points at steady state conditions. The performance map is shown as a carpet plot of input power versus heat load for constant

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Figure 3. Performance map for heat rejection temperature of 300K.

compressor stroke lengths and constant cold end temperature. It can also include specific power

lines, which are used as a tool to compare cryocooler performance. The specific power is the inverse of the thermodynamic COP. In comparing two different cryocoolers, at comparable conditions, the cooler with the lower specific power is considered as a better performer. The performance map also allows the estimation of additional information for constant temperature or constant stroke lines without actually performing the experiments along those lines. Performance maps are dependent on heat rejection temperature. The performance map for a heat rejection temperature of 300K is shown in Figure 3. The

data points are obtained at steady state for fixed stroke lengths ranging from 2 mm to 8.5 mm and fixed heat loads from 0W to 6W. The input power and resultant cold end temperature values

were analyzed, processed using curve fits, and reduced to get constant temperature lines. The dark lines show the characteristics of performance at constant cold head temperature and constant stroke length. The constant temperature lines are verified by running a set of constant

temperature load lines. These data points are shown as filled circles in Figure 3. The curve fit trend lines provide a comparable accuracy for the constant temperature lines. The specific power lines are shown as dotted lines ranging from 5W/W to 150W/W. The data point for a 2W heat load with the cold end temperature at 60K shows that the specific power is below 30W/W, as specified in the design requirements. Similar performance maps were generated for heat rejection temperature of 275K and 325K.

Cool-down Curves Cool-down curves provide important information regarding the cooling effectiveness of the cryocooler. The time to cool down to the design temperature and the lowest possible temperature can be used as one of the performance criteria for comparing two different cryocoolers or for the selection of a cryocooler for a rapid cooling mission requirement.

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Figure 4. Cool-down curves for heat rejection temperatures of 325K, 300K and 275K with no heat load and a 2W heat load.

A total of six cases were considered. For the first three cases, heat load was not applied, and the heat rejection temperature was set at 300K, 275K and 325K. For the next three cases, a 2W heat load was applied when the cold end temperature approached 225K at the same three heat rejection temperatures. The reason the 2W load was not applied until the cold end reached 225K was to protect the resistor heater, which is rated for 0.1W at room temperature, from burning up. The six curves are shown together in Figure 4. The time is scaled so that there is a common zero

where the cooling starts. During the cooldown, the input command is held constant by the LabVIEW™ controller VI, branding it as a constant input command rather than a constant stroke cool-down since the stroke length changes as the cooler cools down. Design Point Performance The operational parameters for the 60K at 2W design point are given in Table 1 for all three heat rejection temperatures. They are provided as a reference for future evaluation. Temperature Stability A space flight cryocooler must show temperature stability for a constant heat load and other fixed conditions. It should maintain a given cold end temperature for a certain length of time.

From a laboratory-experiment standpoint, the temperature should stay within a fraction of a Kelvin for a period of one hour. A larger window is given for a 24-hour period. Although this is not a pass/fail criterion, it is a desirable feature. In Figure 5, a long term (>48 hours) stability evaluation for a heat rejection temperature of 300K and a 1W heat load is shown. The stability

CHARACTERIZATION OF RAYTHEON’s 60 K 2W PSC COOLER

Figure 5. Temperature stability at heat rejection temperature of 300K with 1W heat load.

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criterion of ± 0.1K for 1 hour and ± 0.5K for 24 hours was selected. The banded profiles

superimposed on the actual cold end temperature are shifted by an hour with a value change of ± 0.1K and by 24 hours with a value change off 0.5K. For strict stability, the cold end temperature should stay within these bands. This was achieved for a heat rejection temperature of 300K and a

heat load of 1W, as shown in Figure 5. The bands take into account the fact that rack electronics are influenced by the ambient temperature in the laboratory, which fluctuates between daytime and nighttime outdoor temperatures. This made it difficult to maintain a given temperature for a fixed heat load. During the stability evaluation, the input commands for the cryocooler components were fixed through the LabVIEW™ software controller VI.

Transient Thermal Response. For a cryocooler to be considered for cooling space electronics, its sensitivity to the orbital fluctuation in heat rejection temperature must be determined. The transient thermal response (TTR) evaluation done on the PSC assumed a sinsusoidal variation in heat rejection temperature between 295K and 305K with a 90-minute period. The heat load was fixed at 2W. Figure 6 shows the variation in rejection temperature and the corresponding variation in cold end temperature. The rack electronics temperature is also included to show its variation in the same

time frame and extrapolate its effect on the cold end temperature. The cold end temperature is shown at the bottom of the plot on the primary axis and shows a cyclic variation due to the variation of the heat rejection temperature, as well as a slight creep due to the drift in rack electronics temperature. The overall temperature variation during this time period is about ± 0.5K. Thermal Cycle/Thermal Vacuum Effects (233/330K Proto-Qualification Level Cycles)

The thermal cycle evaluation is performed by putting the entire cryocooler assembly through extreme changes in temperature, as are normally encountered in typical space flight conditions. The idea behind this experiment is to pick out any failure mechanisms that arise as a result of a

Figure 6. Transient thermal response with heat rejection temperature of 300±5K, 90min period.

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Figure 7. Thermal cycle/thermal vacuum heat rejection temperature profile.

mismatch in coefficients of thermal expansion. The PSC was cycled through a predefined profile of heat rejection temperatures ranging from 230 K to 328 K as shown in Figure 7. The cryocooler was operational while the heat rejection temperature was between 275K and 325K. The dwell time on the first and last cycle lasted for 6 hours at each extreme temperature. For

shorter cycles, the dwell time was limited to 1 hour. As the experiment progressed, it became

apparent that the limited chiller capability did not allow a defined dwell time at the lowest temperature. The four cycle test was then repeated with a 3 hour dwell time to account for temperature settling. Figure 7 shows the entire process, which lasted for about 5 days. The

controlled profile was generated automatically through control software. The dwell time, rise/fall rate and start/end temperature settings were adjustable. The desired temperature change rate was 1K/min. Stiction tests were performed during the 325K dwell times of the first and last cycles to verify that no contact was being made between the piston and the cylinder.

CONCLUSION

The Raytheon 60K PSC performed well at the design point of 2W @ 60K with a heat rejection temperature of 300K. The design point settings required less than 30W/W specific power showing an improvement in performance over split-Stirling space cryocoolers of the past. The optimization of parameters was carried out to obtain the most efficient combination of operation parameters. A performance map was generated for heat rejection temperatures of 275K, 300K and 325K. A long-term stability evaluation showed that the dependence of the laboratory grade rack electronics on the ambient laboratory temperature caused the cold end temperature to fluctuate. The proto-qualification thermal vacuum/thermal cycle evaluation did not bring out any failure mechanisms with the heat rejection temperature between 230K and 328K. The PSC was moved to a 24” chamber and is undergoing an endurance evaluation. So far, the PSC has accumulated over 3600 hours of operation.

The Development of a 10 K Closed Cycle Stirling Cooler for Space Use G. Baker†, D. Féger†, A. Little†, A.H. Orlowska#, T.W. Bradshaw#, M. Crook#, B.J. Tomlinson*, and A. Sargeant+ †

Astrium (UK) Ltd, Stevenage, United Kingdom Rutherford Appleton Laboratory, Chilton, Didcot, United Kingdom *US Air Force Research Laboratory, Kirtland AFB, New Mexico, USA + Cubic Applications Inc, Lacey, Washington, USA

#

ABSTRACT

A two-stage, split, Stirling cycle 10 K cooler is being developed for space applications to achieve the requirement for cooling silicon-based IR detectors. This program has been sponsored

by the US Ballistic Missile Development Organization. It is a further extension of a Rutherford Appleton Laboratory (RAL) 20 K cooler being space qualified at Astrium (formerly Matra Marconi Space (MMS)). New features include optimized geometry and enhanced regenerator materials as well as a larger compressor system. This paper describes the experimentation performed to develop the optimized geometry for the cooler displacer. The experimentation was achieved using existing two-stage Stirling cycle 20 K cooler hardware. Operation at 10 K is a challenging concept for a Stirling cooler and necessitated the development of new heat exchanger materials and configurations optimized for this new temperature range. The proof of concept program included the detail design of the optimized displacer and compressor, and the manufacture and assembly leading to proof of concept laboratory testing. During testing, the cooler reached a base temperature of 9.4 K. INTRODUCTION

This 10 K cooler has been derived from an initial 500 mW at 35 K US Air Force Research Laboratory development effort, which was modified in 1997 to focus on a 45 mW at 10.3 K requirement to meet the operational temperature and heat lift needs of Si:As IR sensors.

The effort was conducted in collaboration with RAL, and proposed to meet the new 10.3 K requirement using a two-stage Stirling cooler based on the experience gamed with the design of the MMS 20-50 K cooler, but using rare earth type regenerator materials and an increased swept volume; the MMS 20-50 K cooler has a 12 K base temperature and was developed for the European Space Agency. DEVELOPMENT STRATEGY

The major problem that prevents a regenerative cycle from reaching very low temperatures is

the regenerator heat capacity. As the temperature falls, the density of the gas passing through the Cryocoolers 11, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2001

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regenerator increases, and hence the heat flow into the regenerator also rises. Unfortunately, the heat capacity of most conventional regenerator materials falls as the temperature decreases. Improvements in regenerators have therefore focused on those materials which undergo mag-

netic transitions, at an appropriate temperature, with a large magnetic contribution to the specific heat in order to get a large volumetric specific heat in the 9 to 20 K temperature range. The optimization of regenerator materials has to also fulfil manufacturing and cooler operational re-

quirements. The correct operational parameters are extremely important and any gains made with a better choice of regenerator material may be cancelled by a poor design. In particular, the

regenerator must be porous—with a high surface area to volume ratio—allowing the working fluid to pass though it, and facilitating good heat exchange. At the same time, the regenerator-material trade-off must take into account the thermal and fluidic performances, as well as manufacturability and compatibility with the cooler environment, such as launch vibration or extreme temperature cycling during operation. A systematic review of regenerator candidate materials was performed

to identify materials having the highest volumetric heat capacity, and, at the same time, mechanical and chemical properties compatible with the manufacturing process and launch environment. The selected materials went through a qualification process with regard to their mechanical,

thermal and chemical compatibility. The second problem of low temperature in any refrigeration systems is one of efficiency. The

maximum Carnot coefficient of performance at 10 K is 3.5%, and a small Stirling cycle cooler would typically only operate at 1 % of Carnot at 10 K. This implies that the input power required to obtain 45 mW at 10 K would be well above the capability of a pair of existing MMS cooler

compressors such as those used in the MMS 20-50 K cooler, which are limited to 50 W input power. As the current cooler is a proof-of-concept model, and the key issue is the design of a regenerator to provide enough efficiency at temperatures below 20 K, MMS proposed to focus the development effort on this part of the cooler. For the compressor, it was decided to use two pairs of MMS 20-50 K cooler compressors; they would be used as a development tool to test the most critical component of the cooler, the regenerator.

For the same reasons, as well as to constrain costs, the displacer mechanism and its momen-

tum balancer were based on the existing MMS 20-50 K design. In order to give confidence in the ability of this type of cooler to achieve the specified

performance, both laboratory test work and mathematical modelling have been utilized. The computational modelling has focused on predicting the effect on cooler performance of geometrical and material changes to the regenerator design. The test effort has been used to correlate the mathematical model and to demonstrate actual measured performances of various displacer configurations. This approach resulted in a first prototype running at RAL in 1998 on which different candidate regenerator materials were tested for performance while MMS was qualifying the corresponding manufacturing processes to insure that they would meet mechanical and thermal envi-

ronments. This first prototype reached a 9.9 K base temperature in November 1998.

Based on these results and a pressure drop analysis (see Orlowska1), a new displacer design was defined and manufactured to further improve the performance. This displacer was, in 1999, connected to the existing two pairs of compressors to make a proof of concept model of the 10 K cooler whose design and testing are described hereafter. DESCRIPTION OF THE 10 K PROOF OF CONCEPT COOLER

Figure 1 gives an overview of the overall configuration of the cooler. Although this is a

laboratory proof of concept cooler, it is made of flight-type hardware and follows a flight like architecture: - Two pairs of back-to-back MMS 20-50 K cooler compressors

- One MMS 20-50 K cooler displacer mechanism - One MMS 20-50 K cooler momentum balancer

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Figure 1. 10 K cooler overview.

Two testing configurations have been considered for this prototype: one at ambient atmosphere, with air cooling, at MMS premises, and the other in a thermal vacuum chamber, with conductive cooling, at the US Air Force Research Laboratory at Kirtland Air Force Base, To be

able to operate in vacuum the whole cooler is mounted on a single base plate, which mimics the spacecraft mechanical and thermal interfaces. The four compressors are mounted in two co-axial pairs to minimize the level of force exported on the assembly. Each pair of compressors has a transfer line and a manifold leading from the compressed volume inside the cylinder head to the displacer. The compressor pairs are run in anti-phase with respect to each other. The compressor pairs are mounted on a simple supporting structure. This structure is made from aluminium alloy, and transfers heat from the unit heads to the base plate. The compressors

are attached to the structure via adapter plates with thermal gaskets fitted to enhance the thermal conduction across the joint. An aluminium thermal housing is fitted to each compressor linking

the copper cylinder head to the support structure to increase the heat rejection capability. A thermocouple is attached to each of the compressor heads for temperature monitoring during operation.

The compressor support structure is bolted to a base plate using thermal filler at the interface to reduce the compressor head temperature in a conductively cooled configuration. A number of options were considered for the manifold and transfer line assembly. These

included: - Four compressors feeding into one manifold connected to the displacer; - Two manifolds joining the two pipes from each compressor pair and leading to a third manifold joining these two pipes, with a single pipe leading to the displacer. The design selected features two transfer lines leading to the two compressor pairs to separate inlet ports in the displacer head. The benefits of this arrangement include: - Use of existing compressor assembly without modification; - Low overall transfer line length and manifold complexity; - More even pressure distribution in the displacer. A cylindrical support tube which surrounds the momentum balancer and the displacer mechanisms provides at the same tune mechanical support and conductive heat sinkage. This is repre-

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sentative of a flight bracket, with the exception that provision of force transducer is not included. The exported vibration of this assembly is low level, but not actively controlled. The concept can be extended to include force transducers between the displacer mounting plate and the tube assembly. In that case, flexible thermal straps would be used to link the displacer mounting plate to the support tube. The cold finger itself was initially equipped with a launch support tube to provide structural support to the cold finger during launch vibration. Based on the MMS 20-50 K design, this support tube features resilient snubbers giving a small clearance at the cold end of the cold finger to limit deflection and loads during vibration with limited heat leak. Thermal blankets are mounted inside the launch support tube to insulate the cold finger. This launch support tube was not required for this proof of concept for which no launch vibration requirements were specified by the customer. For the case of the MMS 20 -50 K cooler, the European Space Agency put stringent requirements on the cooler side loads capabilities, so, as explained below, this support tube was later-on removed to improve the cooler with respect to heat leak and heat shield performance. TEST SET UP

For test purposes at ambient pressure, the cold finger was surrounded by a vacuum can, which was linked to a vacuum pump. This vacuum can was equipped with an electrical feedthrough for the thermal sensors and the heater wires. Figure 2 gives an overview of the cooler test configuration. The cooler was tested on a fill and purge test bench with which it was possible to evacuate the cooler and adjust as needed its fill pressure. Several fens blowing on the displacer and the four compressors provided forced-air cooling. Thermocouples were located on the compressor heads and on the displacer flange to monitor heatsink temperatures. The cold tip was equipped with two high precision temperature sensors. An additional sensor provided the mid-stage temperature. The cold tip was equipped as well with a heater to simulate the user’s heat lift requirement. The compressors, displacer, and momentum balancer were driven by two sets of laboratory drive electronics, each driving one of the pair of compressors, and either the displacer or the momentum balancer. One of the drive electronics was used as the “master”, to which the second one was “slaved” to provide adequate phasing and operating frequencies between all the different components.

Figure 2. 10 K test set up overview; note launch support tube around cold finger.

10 K CLOSED CYCLE STIRLING COOLER FOR SPACE USE

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TESTS RESULTS

Performance testing of the optimized cooler started in September 1999, but the cooler only achieved 40 K. MMS and RAL ran a joint investigation of the performance. Investigation focused quickly on the displacer and it was decided to disassemble it to check: 1) the regenerator material packing, 2) the regenerator geometry, and 3) the cold finger bore geometry. The regenerator geometry was found to be faulty, generating an oversized clearance between

it and the cold finger bore, leading to bypass flow of gas toward the cold tip expansion volume. Based on these findings, and in agreement with the customer, a recovery plan was set up at the beginning of October, to re manufacture a new regenerator. The displacer was rebuilt by the end of November. After rebuilding the displacer, the cooler achieved 12 K during initial testing, confirming the investigation findings. Testing was re started in January 2000 at MMS after vacuum baking, to remove any moisture in the system. Optimization of the cooler was carried out with different fill pressures. The lowest temperature achieved on the cold tip was then 11.2 K with a mid stage temperature slightly under 200 K. Review with the customer and RAL of these second performance tests results concluded that the cold stage was performing correctly, but that the temperature of the mid stage was much higher than expected. One of the main explanations identified was that there could be some extra heat leaks into the system, especially as RAL pointed out that they did not consider, either in their model or in their previous 10 K prototype, the implementation of the above-mentioned launch support tube; the tube was inherited from the MMS 20 K cooler design. As the customer had, on this proof of concept model, no requirement with regard to launch vibration, this launch support tube was removed. At the same tune, the Multi Layer Insulation (MLI) was reviewed and considered non-optimized as well, so it was decided to: 1) remove the launch support tube, 2) redesign and re manufacture the MLI around the cold finger, and 3) design and manufacture an aluminium

heat shield attached to the mid stage. Routing of the wires of the cold tip temperature sensors was

reviewed too, to reduce the corresponding heat leaks. As thermalization of the displacer was identified also as being critical, an additional copper heat sink was designed and manufactured, using the bolted interface of the former launch support tube to provide additional heat sink around the base of the cold finger. This extra heat sink was implemented later in final testing so the impact of the heat leak reduction actions could be seen initially. Testing was done on the cooler to verify the improvement on performance seen on the cold tip, due to these modifications. The heat load to be applied to the cold tip to reach the temperature achieved before these heat leak reduction actions was 67 mW. The mid stage temperature was reduced by 30 K. Further optimization, especially with regard to fill pressure, led to a base temperature of 10.4 K, and to 11.6 K with a 45 mW heat load. Following these results, the cooler was warmed up to ambient in order to implement the additional displacer heat sink, replace the cold finger open bracket with a flight-like cylindrical bracket, and to crimp the fill ports; this brought the cooler to its deliverable configuration. Further optimization tests with regard to phase, frequency, and displacer and compressor strokes were then performed in March 2000. These led to an ultimate record breaking base temperature for a Stirling space cooler of 9.4 K, and 10.4 K with 45 mW heat load as shown in Fig. 3. CONCLUSIONS

This test campaign, summarized in Fig. 4, successfully demonstrated the feasibility of reaching the 10 K temperature range with a space-rated, two-stage Stirling cooler. This achievement is mainly due to the successful use of advanced rare earth regenerator materials. With 10.4 K, the target specification of 45 mW at 10.3 K was nearly met during test with ambient-air cooling; this cooling was not able to keep the compressors heads and displacer flange below 30 C. During the foreseen thermal vacuum tests at the US Air Force Research

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Figure 3. 10 K proof of concept cooler heat load line.

Laboratory, tests at colder heat rejection temperatures should allow the cooler to overcome this present limitation and obtain the ultimate performance of the cold regenerator.

With regard to design, the regenerator was successful, especially the cold stage. The mid stage temperature being 30 K higher than foreseen, we therefore plan to review its sizing to further lower its temperature. This mid stage redesign should be performed taking into account the heat shielding and sensor interface strategies, as they are key parameters for the overall instrument cryogenic performance, and should be investigated in detail to review the cold finger design accordingly. The displacer mechanism should be reviewed to improve its heat sinking capability; the displacer waste heat is estimated at 60 W, leading to a displacer flange temperature as high as 30°C when using ambient (20°C) forced-air convection cooling.

Figure 4. 10 K proof of concept cooler temperature history.

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The four 20-50 K cooler compressors used to feed the 10 K displacer in this proof of concept cooler fulfilled their task with limited heat sink capability and input power margins. A new compressor, with a greater swept volume, should be developed to meet the 10 K cooler requirements. This should reduce the mass and cost of the cooler and increase its efficiency. This future work should lead to a qualified and optimized Stirling cooler able to provide up to 100 mW in the 10 K temperature range, either to meet the requirements of future Silicon based sensors or to act as a pre cooler for Joule Thomson coolers operating in the 3 to 4 K temperature range. ACKNOWLEDGMENT

This development work was funded by the Ballistic Missile Defence Office. REFERENCES 1.

A.H. Orlowska, T.W. Bradshaw and S. Scull, “Progress Towards the Development of a 10K Closed Cycle Cooler for Space Use,” Cryocoolers 10, Plenum Publishers, New York (1999), pp. 67-76.

Development of a 12 K Stirling Cycle Precooler for a 6 K Hybrid Cooler System W.J. Gully, D.S. Glaister, and D.W. Simmons Ball Aerospace & Technologies Corp. Boulder, CO, USA 80306

ABSTRACT There is a need for reliable, space-qualified mechanical coolers for temperatures of 10

Kelvin and below for use in both X-ray and infrared systems. Our analysis shows that hybrids consisting of a Joule-Thomson (J-T) cooler coupled with a mechanical cooler are the most efficient when used in a low Earth orbit. We are developing a hybrid cryocooler consisting of a helium J-T system coupled to a Stirling cycle mechanical cooler for these applications. We plan to use the relatively mature Stirling cycle mechanical cooler to provide all of the precooling, and

use the J-T recuperative system for refrigeration below 20 K, a region that has historically been difficult for regenerative coolers. In this paper we discuss our work in developing this precooler on our NASA Explorer 6 K program. A discussion of the J-T portion of this system will occur elsewhere in this conference.1 Our precooler work is a continuance of our previous work on linear Stirling coolers. We

have developed one-, two-, and three-stage Stirling coolers for various NASA and Department of Defense programs. For the Explorer 6 K, we have focused on adapting our three-stage Stirling cooler for operation at lower temperatures. In this program we are using breadboard tests to acquire both an empirical knowledge and an analytical understanding of the factors that affect the low-temperature performance of our cooler. This information will be used to develop a spacequalified precooler during a later phase of the system development. In our initial tests we have demonstrated improved performance when we substituted a lead shot regenerator for the previous regenerator constructed from phosphor bronze screens. Although the data suggests future directions for more cooling, at present the performance falls short of our analytical expectations.

MOTIVATION FOR A HYBRID COOLER

Our use of a hybrid system reflects the difficulties in achieving the low temperatures with just a regenerative mechanical cryocooler.2 With a hybrid, the effort below 20 K is shared between the mechanical cooler and the J-T cooler (Figure 1). Our analysis suggests that the hybrid system can be designed to run with approximately the same power efficiency with the precooler operating anywhere from 12 K to 16 K. Our task is to learn what we can actually achieve with our precooler. The applications we are looking at, a sub-Kelvin ADR for the X-ray work and a medium/far infrared wavelength detector for the infrared, require about 100 mW of cooling at this temperature. Cryocoolers 11, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2001

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SPACE STIRLING CRYOCOOLER DEVELOPMENTS

Figure 1. A schematic of a hybrid cooler, consisting of a J-T system coupled to a multistage mechanical cooler.

There are a number of intrinsic problems that regenerative coolers face at these temperatures. The density of helium gas has become significant and there is a lack of sufficient matrix heat capacity to store the internal energy of the working fluid during the cycle. In addition,

because of the high densities, miniscule dead volumes in the cold end hold substantial amounts of nonproductive helium and degrade the performance. The pressure dependence of the heat capacity in the helium working fluid leads to a phenomenon in the regenerator similar to the imbalance in a recuperative heat exchanger and causes unavoidable heat leaks. There are additional problems associated with the use of the spring suspended linear cool-

ers. These coolers were developed to eliminate the need for a wearing seal in both the compressor and in the cold head. With a stiff radial suspension, the gap clearance can be made small enough to dynamically support a pressure wave at sufficiently high frequencies. These coolers typically operate above 30 Hz to sustain the pressure and to run on resonance. Early work has been done on the extra losses in the cold head due to this effect.3 In contrast, commercial mechanical coolers with wearing seals easily provide watts of cooling at these temperatures. Because the regenerator heat current scales with operating frequency, the ability to run at low frequencies (typically a few Hz in these coolers) is a thermal advantage and results in higher effectiveness. Due to our reliability requirements, we must use our noncontacting Stirling precooler and operate at the higher frequencies.

6 K CRYOCOOLER TEST PROGRAM

We are using a three-stage Stirling cycle breadboard cooler originally developed for the Air Force 35/60 K program. At the time our goal was to provide two separate refrigeration stages on the same cold finger. In this effort we use the multiple stages to provide efficient operation at low temperatures. With three refrigeration stages, we can divide the temperature drop across more regenerators and reduce the heat load on the last stage. This compensates for the increased ineffectiveness and allows for the lowest possible temperatures at the cold tip. The multistage expander offers several advantages as a precooler for a J-T system. The intermediate stages are used to absorb the heat due to the inefficiencies of the various recuperative heat exchangers that precool the fluid for the J-T system.

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Figure 2. The Stirling cycle displacer breadboard used in the development of the low-temperature

cooler.

Figure 3. The baseline thermodynamic performance of the 35/60 K cooler at its lowest temperatures.

The baseline configuration of our breadboard cooler is shown in Figure 2. Its three expansion stages are indicated by steps in the outer tube. Both middle and cold stages have interfaces for attaching loads. In this particular case the warmest stage, which typically runs around 180 K, is used primarily to improve the thermodynamic efficiency of the unit and has no external load. This breadboard is particularly suited to regenerator studies. It has demountable interfaces at the base, midstage, and cold tip, which allow the external regenerators to be removed and replaced without much difficulty. In particular, this can be done without disturbing the displacer drive mechanism. Our goal was to understand the specific influence of a number of factors at low temperatures, so we began by making a careful measurement of the performance of the existing 35/60 K cooler. In this configuration we used typical wire screen regenerators. The results are shown in Figure 3. As an example, when the cooler was lifting 0.1 watt at the cold tip, and had no load on the midstage, the cold tip was at 20.5 K and the midstage was at 42 K. The input power in these circumstances was approximately 48 electrical watts to the compressor.

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Figure 4. The breadboard cooler cold finger, showing the larger diameter outer regenerator on our displacer drive for the lead regenerator. The original outer tube for the screen regenerator is to the left.

We performed trend tests to explore the cooler’s limitations near its lowest temperatures. The strongest indication of regenerator limitations was the lack of response to increased pressure and frequency. These trends generally agreed with the results of our analytical models,4 and could be explained by the vanishing capacity and subsequent loss in the regenerator. These trends and the analytic models suggested that the cryocooler was regenerator limited. Consequently, an improvement in the matrix capacity should produce a substantial improvement in the low-temperature performance because of improved efficiency. Improving the efficiency removes a “heat leak” between the midstage and the cold tip, and we predicted a quarter watt increase in lift at 20 K, along with a similar loss of heat lift at the midstage. TEST WITH LEAD SPHERES

Our first effort for converting to work at low temperatures was to introduce a lead (95/5 PbSb “babbit”) regenerator. This is a traditional choice for a low-temperature regenerator because of its relatively high capacity, because it is well characterized, and because it is available as fine shot, a form that can be used in a regenerator. We procured the lead from Clad Metal Industries.5 Our goal was to understand its real impact on our low-temperature performance. The regenerator influences the cryocooler performance in a number of ways, and it can be difficult to separate them out. In an ideal situation we could have obtained woven lead screen with the same porosities and flow factors as before, so the substitution would be limited to the material properties only. However, we have tried to size the lead regenerator to mimic as many of these flow properties as we could. The remaining determinant factor in the cooler performance should be the change in matrix material capacity. The cold finger, with the larger cold regenerator, is shown in Figure 4.

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Figure 5. Thermal performance of the breadboard cooler with lead ball regenerator and midstage (mid) and cold tip (ct) heat loads as indicated.

The key results of our tests are shown in Figure 5. The cooler had the same pressure ratio

and input power as before. But as a result of the substitution of the lead ball regenerator, there was a substantial shift up in no-load temperature at the midstage of approximately 5 degrees Kelvin, and a drop in temperature of approximately 2 K at the cold tip. Both shifts were qualita-

tively expected, but quantitatively the improvement at the cold tip seems to be too small. The midstage shift suggests the loss of 300 mW of indirect refrigeration, presumably by the regen-

erator loss, but the cold tip seems to have improved only by roughly 100 mW. We initially suspected that flow restrictions at the midstage may have reduced the gross capacity at both

stages and explained this change, but subsequent testing with larger channels showed that not to be the case. Our auxiliary trend studies now indicate a direction for improvement, but are also somewhat contradictory, which means more work is needed before the low-temperature performance will be understood. As before, the trends indicate that the cooler is regenerator limited, in that it does not respond to increased pressure and input power with more refrigeration. But the cold tip responds strongly to compressor stroke to the end of its range, suggesting that the next step is to increase the compressor’s displacement relative to the existing hardware. Conversely, the cooler does not respond to displacer stroke, suggesting ways to eliminate dead volume. One trend study with the lead regenerator was of special interest. As discussed in the introduction, there are solid arguments for the need to operate a cooler at low frequencies for efficient operation at low temperatures. This motivated a series of tests in which we maintained the same strokes and pressure, but dropped the frequency. The result for the cold tip temperature (at a load of 100 mW) is shown in Figure 6. There is a general improvement as we go to lower frequencies, but it is not particularly strong. UPCOMING WORK

Because of the previous results, we are currently rebuilding our compressor with increased displacement. We also are making a number of small changes to the displacer hardware to wring out some amounts of cold dead volume. The goal is to find the right proportions for best performance. Regenerator work up to this point has focused on the cold regenerator. We expect to also

modify the upper regenerators to both improve the heat lift in the vicinity of 40 K, and if pos-

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Figure 6. Thermal performance as a function of drive frequency for /the breadboard cooler with a lead ball regenerator.

sible, reduce dead volume to enhance the performance of all stages. We will explore whether this be best done with modified screen mesh regenerators, or with powders of more exotic materials. The goal is to develop the best cooler possible by modifying the regenerators on our bread-

board hardware. ACKNOWLEDGMENTS This work is supported by a grant from the Explorer Technology Initiative program at the Goddard Space Flight Center, NAS5-99239. We also would like to thank Steve Castles at GSFC for his permission to use hardware developed on the 30 K multistage cryocooler program in support of this work. REFERENCES

1. Gully, W.J., Lester, J., Levenduski, R.C., Simmons, D.W., Wright, G.P., Tomlinson, B.J., Davis, T.M., and Reilly, J., “Rotary Vane Compressor Development for a 10 K Cryogenic Cooler,” Cryocoolers 11, Plenum Press, New York (2001). 2. Bradshaw, T.W., Orlowska, A.H., Jewell, C., Jones, B.G., and Scull, S., “Improvements to the Cooling Power of a Space Qualified Two-Stage Stirling Cycle Cooler,” Cryocoolers 9, Plenum Press, New York (1997), pp. 79-88. 3. Keung, C., and Lindale, E., “Effect of Leakage Through Clearance Seals on the Performance ofa 10 K Stirling-Cycle Refrigerator,” Proc. of the Third Cryocooler Conference, N.B.S. Pub. 698 (1985), pp. 127-134.

4. Gary, John, Daney, David E., and Radebaugh, Ray, “A Computational Model for a Regenerator,” Proc. of the Third Cryocooler Conference, N.B.S. Pub. 698 (1985), pp. 199-211. 5. Clad Metal Industries, Inc., 40-T Edison Ave., Oakland, NJ 07436 USA.

Thermodynamic Optimization of Multi-Stage Cryocoolers C. S. Kirkconnell and K. D. Price

Raytheon Electronic Systems El Segundo, California, USA 90245

ABSTRACT

Active Stirling class cryocoolers, including pulse tube coolers, are complex, difficult to optimize machines. The large number of characteristics and properties associated with geometry, materials, gas properties, heat transfer devices, flow manifolds, mechanical

mechanisms, and electro-mechanical devices that determine a particular machine’s performance make quick optimization difficult. Single-stage coolers are now sufficiently well understood that design optimization is reasonably straightforward. However, multi-stage coolers compound the design problem by virtue of the dramatically enlarged number of variables, and optimization is

still a challenge. Often, multi-stage machines are “optimized” by a brute force search of the design space or design decisions are made based on overly generalized or inaccurate assumptions

about relationships between variables. The schedule-constrained time typically available to perform optimization procedures combined with the large number of variables and their complex

interaction results in sub-optimal products. This paper presents a concept for optimization that more rapidly converges on an optimal design.

INTRODUCTION

This paper presents a cryocooler design optimization method based on analysis of exergy flow. Exergy analysis follows the destruction of energy availability from the machine input to the low temperature stage(s) where refrigeration is produced.

First, the terms and approach of the exergy method are reviewed and clarified with reference to thermodynamic analysis of one-stage coolers. Then the analysis is extended to two-stage

coolers. This is followed by a discussion of the optimization technique in one- and two-stage cryocoolers and, finally, the generalization of the method to multi-stage cryocoolers.

THERMODYNAMIC ANALYSIS

Single-Stage Cryocooler Thermodynamics A careful examination of the thermodynamics of the comparatively simple single-stage cryocooler is helpful in describing the operation of the more complicated multi-stage devices. For that reason, a brief review of fundamental single-stage cryocooler thermodynamics is presented. The conventional approach to cryocooler thermodynamic design involves the variation of design parameters, such as component volumes, regenerator matrix type and dimensions, etc., so Cryocoolers 11, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2001

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SPACE STIRLING CRYOCOOLER DEVELOPMENTS

Figure 1. Energy flow schematic for single-stage refrigerator; arrow denotes energy flow direction.

Figure 2. Temperature entropy diagram for a single-stage Stirling refrigerator.

that the capacity objective is achieved with the maximum thermodynamic efficiency.

ARCOPTR is an example of a widely-used model for this type of optimization.1 Though challenging in practice because of the complicated and interdependent physical phenomena involved, this approach is conceptually simple. Consider the energy flow diagram for a singlestage cooler shown in Figure 1. The conservation of energy dictates that

and the coefficient of performance (COP) is given by

The design parameters are therefore varied such that is maximized for the desired at The effectiveness of this approach is hampered by an imperfect knowledge of the detailed internal thermodynamics and the vast trade space afforded by the multitudinous design options.

The primary shortcoming of the above approach is that it provides no figure of merit by which to evaluate the relative goodness of the achieved. The standard figure of merit, the Carnot COP is obtained by considering a reversible refrigerator operating between the temperature extremes of interest, and is shown in innumerable thermodynamics texts to be

defined as follows:

A Second Law COP can now be defined in terms of Eqns. (2) and (3):

In words, Eq. (3) states that is the maximum amount of refrigeration that can be obtained from a refrigerator operating between and given an input power of This can be clearly shown on a temperature-entropy (T-s) diagram for a refrigerator operating on a

Stirling cycle (Figure 2). The area bounded by the curves is the net refrigeration produced. The actual refrigeration is less than the maximum possible (reversible) refrigeration for a refrigerator operating between the given temperature extremes because of practical inefficiencies (irreversibilities) which can be roughly grouped into two categories, pneumatic losses and

parasitic losses. Pressure drops, rounding of the corners of the T-s area due to real-world limits

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71

over phase control, seal leakage losses, and other such losses that decrease the amount of gross refrigeration produced are termed here “pneumatic losses.” Conduction losses, radiative loads, regenerator inefficiency, and other losses that consume a portion of the gross refrigeration produced are differentiated into another category called “parasitic losses.” Multiple expressions

for the net refrigeration rate arise from these definitions:

Optimum refrigerator performance can also be sought through the method of entropy generation minimization. Referring back to Fig. 1, it is clear that the rate of entropy generation is given by

Efficiency is maximized when is minimized. Though intuitively evident from these simple expressions, it is interesting to note that the equivalence of the First Law and Second Law methods can be demonstrated by considering the implications of irreversibilities on a

refrigeration cycle from two distinct perspectives. Using the interpretation provided above that fixes the input power and attributes the difference between the reversible refrigeration rate and the actual refrigeration rate to the irreversibilities in the system, an expression can be obtained for the lost refrigeration in terms of the other system parameters:

Using an alternative interpretation in which the refrigeration rate is fixed and the irreversibilities manifest as an increase in the actual required input power above that required for a reversible system, i.e.,

a similar expression to Eq. (7) can be obtained for the lost input power:

The algebraic combination of Eqns. (7) and (9) reduces to the simple expression

Using the definition for lost work (power) provided in Bejan and elsewhere, this analysis finally yields to an expression for the lost refrigeration capacity in terms of the entropy generation:

This expression demonstrates that the First Law optimization approach of minimizing lost refrigeration capacity is functionally equivalent to minimizing the total rate of entropy generation, whatever the source. (The reference temperature can be taken as for a refrigerator rejecting waste heat to the ambient environment.) All of these concepts of energy flow and irreversibility come together in the exergy flow map provided in Figure 3, which is modeled after the techniques demonstrated by Bejan.2 The input exergy, also called availability, provided by the compressor flows down from the warm end reservoir to the refrigeration temperature where refrigeration is produced. Exergy is destroyed through the pneumatic and parasitic loss mechanisms. The exergy content of a power interaction is equal to the power, while the exergy associated with a heat transfer interaction is a function of the temperature at which the heat transfer occurs:

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Figure 3. Exergy flow diagram for a singlestage refrigerator. Exergy is destroyed through parasitic and pneumatic losses.

Figure 4. Energy flow diagram for a particular type of two-stage refrigerator. Other two-stage configurations may dictate alternative energy

flow diagrams.

Thus the rejection of heat to ambient carries with it zero exergy. A refrigerator’s minimum achievable “no load” temperature occurs at the temperature where all of the input exergy has been destroyed through irreversible losses. Of note is the fact that the distribution of losses between pneumatic and parasitic is immaterial from the standpoint of optimization through entropy generation minimization for single-stage refrigerators. As shown in the next section, this is not the case for multi-stage devices. Two-Stage Cryocooler Thermodynamics The thermodynamic analysis and optimization of a two-stage refrigerator follows directly from the analytical approach used for a single-stage device. The stages of multi-stage refrigerators can be arranged thermodynamically in many ways; an energy flow schematic for the type of staging of present interest, a single ambient compressor with the coldest stage rejecting heat to the intermediate stage, is provided in Figure 4. The conservation of energy equation is again defined at the boundaries of the refrigerator and is given by

The input power that drives both stages originates at from the ambient compressor. Therefore, the definitions for both the first stage and second stage Carnot COPs are based upon the ambient temperature

The standard energy flow diagram used in Figure 4 is somewhat misleading in that it appears to show an engine residing between

and

and this leads to the temptation to use

in Eq.

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73

(14a). This is incorrect for the single compressor problem. If one were to use it can be shown that the perfectly reversible system yields negative entropy generation, not zero, which

violates the Second Law of Thermodynamics. In contrast, if a multi-stage refrigerator were to be considered in which individual compressors operating between each temperature level were used to pump heat between those adjacent temperature reservoirs, then the Carnot COP definitions would be based upon the temperatures bounding each stage. The input power can be conceptually partitioned into the individual portions required to drive each stage. The reversible input power can thus be defined as

where

and

are the actual net refrigeration rates. The reversible input power for each

stage represents the minimum theoretical power required to deliver the prescribed refrigeration

rate between and the refrigeration temperature. The irreversible, or lost, power can be similarly partitioned, and the approach demonstrated for the single-stage cooler in developing Eq. (10) leads to similar looking expressions for the two-stage device:

The input power lost due to irreversibilities in the first stage represents an overall system loss, but some of the second stage loss, in particular that portion due to parasitic losses between and is partially recoverable. The energy transfer from to through conduction losses and regenerator enthalpy flux due to heat exchanger inefficiency decreases the capacity at the second stage, but it increases the capacity at the first stage by the same amount. Therefore, the parasitic loss portion of is partially recovered:

Note that the recoverable fraction of the second-stage power lost due to parasitics is given by the ratio of the Carnot efficiencies, The thermodynamic intricacies of the two-stage cooler analysis are captured in the exergy

flow map provided in Figure 5. The flow of input exergy from ambient and its partitioning between the refrigeration stages illustrates the proper selection of in Eq. (14a) as the warm end reference temperature. The map also shows that the second stage parasitic loss carries with it positive exergy, that a portion is destroyed in the transferring of capacity from the colder stage to the intermediate stage, and that the remainder represents the recoverable exergy. The recoverable exergy can be shown to be equal to the recoverable power defined in Eq. (17) by considering the net exergy change due to the second stage parasitic loss component:

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Figure 5. Exergy flow map for two-stage cryocooler. Parasitic heat transfer from middle stage to cold stage increases the capacity at and decreases the capacity at by the same amount. This represents a net exergy loss to the system.

The right hand side can be expressed in terms of the Carnot efficiencies:

This expression reduces to

where the second term on the right hand side represents the exergy destruction and the first term

is the recovered exergy. Note that the right hand side is always negative; although there is partial recovery, some exergy is destroyed through the second stage parasitic loss.

CHARACTERIZATION AND OPTIMIZATION TECHNIQUES

Single-Stage Cryocoolers Thermodynamic efficiency is maximized when the rate of entropy generation is minimized. The optimization of single-stage cryocoolers through numerical analysis typically involves the variation of design parameters such as component size, operating frequency, and phase angles in search of a maximum refrigeration capacity-to-input power ratio. This is simply an indirect method of minimizing entropy generation. For a typical design study, the refrigeration load, refrigeration temperature, and nominal heat rejection temperature are known, and the input power required to deliver the needed capacity is being minimized. Referring to the expression for entropy generation in Eq. (6), the only unknown on the right hand side is so entropy generation is minimized when the heat rejection rate is minimized. The reversible input power is fixed by the refrigeration load and the temperature extremes, so is known as well. Combining Eqns. (1) and (8) yields

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75

in which only is unknown. Therefore, the task reduces to the minimization of The intent of introducing these Second Law considerations is not to replace the parametric characterization of design variables in cryocooler optimization, rather the purpose is to improve the parametric studies by guiding the efforts through an improved understanding of the physics of the system. The Second Law techniques are particularly useful for characterizing the efficiency of the cryocooler using as a figure of merit. For example, if a cryocooler is sufficiently well understood to express the losses in terms of known system parameters, then determined from the entropy generation calculations, can be used to describe the efficiency of the cryocooler over a range of refrigeration temperatures. Assume the refrigeration losses can be approximated by an expression of the form

where

and and are known. (Ideally, the cryocooler is designed such that is minimized in the vicinity of the design point corresponding to From Eqns. (3) and (10), the total input power is given by

The Second Law COP follows immediately from the above using Eqns. (2) and (4). Examples of where this type of characterization might prove useful include design efforts to rescale an

existing cryocooler to increase capacity and thermal system trade studies in which various cryocoolers are being considered for applications outside their originally intended and better characterized range of operation. Two-Stage Cryocoolers

The increased thermodynamic complexity of a multi-stage cryocooler gives rise to a substantially more challenging task of parametric characterization than for a single-stage cooler,

thus making the practical application of Second Law techniques all the more valuable. Consider a two-stage cryocooler design in which the refrigeration loads and temperatures at both stages are prescribed together with the heat rejection temperature. As described above, entropy generation minimization is equivalent to reducing the total lost input power, which is given by the algebraic combination of Eqns. (16a), (16b), and (17):

Substituting and combining terms yields

Expressions are needed for the lost refrigeration in the form of the general equation

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SPACE STIRLING CRYOCOOLER DEVELOPMENTS

and for the second-stage parasitic loss

The above expressions assume the parasitic loss between the intermediate and coldest stage is driven by the temperature difference between those stages, as would be expected for a conductive or convective loss, but allows for the possibility of more complicated interactions between the stages in determining the total loss, which includes the pneumatic loss component. However, since the adjacent boundary temperatures likely dominate the total loss between

stages, Eq. (27) can be simplified:

As an example, consider an application in which a two-stage cryocooler design is being considered to meet a single prescribed refrigeration load and temperature. No intermediate refrigeration capacity is required, so the temperature at the intermediate stage is immaterial to the user. Given this flexibility, the cryocooler designer seeks the optimum intermediate temperature

such that the design capacity at the coldest stage is achieved with minimum entropy generation. Using a simplified form of the loss equation from Eq. (22) in which it is assumed any nonlinearities with respect to temperature are captured in the first coefficient, the following expressions for total lost refrigeration are obtained:

For the purposes of this example, assume the parasitic loss is a linear function of the temperature difference (conduction and convection dominated):

The optimum intermediate temperature occurs where the total lost input power (Eq. 26) is minimized, i.e., where

The partial derivative of interest is calculated through several sequential applications of the chain

rule, only a few of which are shown below for the sake of conciseness.

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The chain rule is applied again for each loss component. For example,

The functions and and the second-stage parasitic loss constant are assumed to be known, which is consistent with the premise that the performance characteristics of each stage is reasonably well understood, and it is the optimum combination of those stages into a single

device which is sought. Eventually, through substitution and further application of the chain rule, the lost power derivative in Eq. (33) is reduced to a singular expression in terms of the

system’s one unknown, The derivative can then be calculated over the temperature range of interest and the optimum value for thus identified. DISCUSSION

Multi-Stage Cryocooler Figure of Merit The application of Second Law principles to cryocooler thermodynamic design helps guide the design by providing a figure of merit that relates performance to that of an ideal cryocooler. Eq. (4) defines that figure of merit, for a single-stage cryocooler. For two-stage and other multi-stage cryocoolers, the concept of exergy flow can be used to define similar figures of merit. As shown in Eq. (12b), the exergy associated with a heat transfer interaction between a system and its surroundings is proportional to the inverse of the Carnot COP corresponding to the temperature at which the heat transfer process occurs. For a cryocooler, the refrigeration capacities occurring at various temperature levels can be normalized to a single refrigeration capacity at a single arbitrary temperature, and the resulting normalized efficiency can then be compared to that of a single-stage Carnot refrigerator. Typically, the temperature at the coldest stage is used to normalize the capacities because it is the achievement of positive refrigeration at that temperature level which stresses the design. It follows that the normalized “pseudo single stage” refrigeration for an n-stage refrigerator is given by the expression

The COP and Second Law efficiency definitions follow directly from Eqns. (2), (4), and (35):

These expressions are useful in comparing multi-stage cryocoolers to each other and to singlestage cryocoolers, and they can also be used to evaluate the relative efficiency between varied temperature distributions and heat loads for a multi-stage cryocooler operating. The latter is

essentially a generalized extension of the two-stage example used above where the net intermediate refrigeration is allowed to vary from zero.

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Extension of Exergy Flow Concepts to n-Stage Cryocoolers The thermodynamic analysis of cryocoolers with more than two cold stages follows directly from the two-stage cryocooler analysis. Assuming a single ambient compressor is driving the multi-stage expander, the Carnot efficiency for each stage is based upon the ambient temperature and the refrigeration temperature for that stage, i.e.,

The concepts of lost power at each stage, the partial recovery of exergy flow due to parasitic heat transfer losses, and the minimization of total lost power for maximum efficiency all translate directly from the two-stage cryocooler analysis. The number of terms in the total lost power

expression (Eq. 25) grows with n, so the calculation of the partial derivatives required to characterize the efficiency of the cryocooler over the range of interest becomes more complicated, but the fundamental approach is the same. The value of employing these techniques actually increases as the number of refrigeration stages increases because the

voluminous trade space associated with multi-stage cryocoolers becomes nearly impossible to adequately characterize with brute force parametric design trades. CONCLUSION The introduction of exergy flow concepts into the thermodynamic analysis of cryogenic refrigerators provides a clear illustration of how internal irreversibilities, which destroy exergy, are manifested at the system boundary as decreased performance. The decreased performance can be interpreted as either a loss in refrigeration capacity for a given input power or an increase in the required input power for a desired refrigeration capacity. The Second Law optimization approach of minimizing entropy generation (i.e., exergy destruction) is shown to yield particular advantage for multi-stage refrigerators because the trade space is broad and the thermodynamic interactions complex. Interestingly, Streich reached a similar conclusion in his exergy analysis of a quite different problem, the thermodynamic characterization of a mixing process involving two natural gas streams.3 He concluded, in part, that exergy analysis is particularly useful for “pioneering” and “systematic studies,” terms which aptly describe the exercise of developing and optimizing a multi-stage cryogenic refrigerator. Future work is planned in which quantitative loss correlations for a multi-stage cryocooler will be substituted into the analytical model described herein to demonstrate the proposed Second Law optimization approach. REFERENCES 1. Roach, Pat R. and Kashani, AH, “A simple modeling program for orifice pulse tube cryocoolers,” Cryocoolers 9, Plenum Press, New York (1997), pp. 327-334.

2. Bejan, Adrian, Advanced Engineering Thermodynamics, John Wiley and Sons, Inc., New York (1988), pp. 111-123.

3. Streich, Martin, “Opportunities and limits for exergy analysis in cryogenics,” Chem. Eng. Technol., vol. 19 (1996), pp. 498-502.

The Advent of Low Cost Cryocoolers R.Z. Unger, R.B. Wiseman and M. R. Hummon Sunpower, Inc. Athens, OH 45701 USA

ABSTRACT

A new cryocooler, the M87, has been designed. It combines the well-tested thermodynamics of the M77 cryocooler with features developed to permit the manufacture of large-volume freepiston (linear) compressors. The price of the cryocooler is projected to be $2,000 per unit in lots of 10,000, with availability in the first quarter of 2001. In comparison to the M77 cryocooler, the M87 shows enhanced rated efficiency (20%), improved cooling power, and reduced length and mass. The M87 is rated at 7.5 W lift at 77 K with 150 W(e) input. The enabling market for development of the M87 is the growing need for portable oxygen therapy for home use. The development of significant new markets and new products based on the availability of a low cost cryocooler is anticipated. For instance, in the telecommunications field alone, estimates of the world-wide market by 2005 range from 120,000 to 400,000 units/year. INTRODUCTION

Many cryocooler designs have been developed and tested over the past dozen years. In the absence of a large market for applications of high temperature super-conductivity (HTS), there has been little impetus to undertake the development work and investment required for mass manufacture of any of these designs. In turn, the lack of low-cost cryocoolers has been an impediment to the development of products that require such technology. Production of a low cost cryocooler requires a design appropriate for low cost manufacturing. Sunpower, Inc. has developed an appropriate design and is now undertaking production in a pilot manufacturing facility. The new cryocooler, the M87, was designed for high volume manufacture and is based on a proven thermodynamic design, Sunpower’s M77 cryocooler. The M77, also a free-piston Stirling cryocooler, is a low volume, relatively high cost unit widely used for aerospace testing. Extensive experience with other free-piston designs contributed to this effort, especially work on linear compressors intended for very high volume, low cost white goods. The first application of the M87 will be in a novel medical device under development by In-X, Inc. of Denver, Colorado. Their system will supply liquid oxygen to meet the needs of patients requiring home-based portable oxygen therapy. The well-established, high volume market for home oxygen therapy provides the enabling market for the initial manufacture of the M87 cryocooler. This paper describes the features of free-piston technology that support the M87 cryocooler design, giving examples based on the M77 cryocooler and on linear compressors. The description of the M87 design includes new techniques that support mass manufacturing (patents pending) and Cryocoolers 11, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2001

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are broadly applicable to free-piston technology. Both the equipment and the processes used in manufactured are outlined. Finally, we speculate on additional markets that are likely to develop with the availability of reliable, low-cost cryocoolers.

FREE-PISTON TECHNOLOGY Free-piston designs at Sunpower have a number of features in common. There is no crank

mechanism. The piston moves back and forth within a closely fitting cylinder, typically at a resonant frequency determined by the overall design. The piston and, where present, the displacer, require no oil or lubrication, but instead are supported by gas bearings using the working gas of the machine. With non-contact gas bearings and clearance seals, there are no life-limiting wear components and the mechanical efficiency is high. Controlling the amplitude of the piston stroke via input voltage can easily and continuously modulate the output of these machines.1 ,2 Frequency is constant. A compliant member, together with planar springs and gas bearings, permits the use of tight but standard manufacturing tolerances.3 An integral linear motor / linear alternator is sealed within the machine,4 with the permanent magnets attached to a cost effective drawn or rolled sheet stainless steel part.5 In a Stirling engine, piston motion is converted into AC electricity. In a Stirling cooler or in a linear (free-piston) compressor, electricity drives the piston motion to implement the Stirling cycle in the case of a cooler, or the vapor compression cycle or pumping action in the case of a compressor. The free-piston design is applicable to a wide range of designs and applications. M77 Cryocooler

The M77 cryocooler (Figure 1) is an example of a Stirling cooler using free-piston technology. This cooler was designed for low volume assembly in our research and development lab; nevertheless, its cost, $35,000, is well below competing options. The first units of this design were deliv-

ered in 1992 and used for HTSC filter applications. To date, a total of 71 units have been delivered to university laboratories, private industry, and U.S. government agencies. Test uses include nuclear detection, earth-orbital weather balloons, oxygen liquefaction, and space applications. The M77 will cool germanium detectors in NASA’s High Energy Solar Spectroscopic Imager (HESSI), a

satellite-based observatory of X-rays and gamma rays from the sun.6 Tests of the M77 by NASAGoddard resulted in the following statement: “In extensive studies over the past two years at GSFC, M77 coolers have been vibrated to the GEVS mandated 14.1 Grms, run under thermal vacuum conditions from -25C to +30C and life-tested already (continuing) for hrs. Monitoring during these tests showed no internal contamination of the working gas. Units already tested at GSFC are fully qualified for flight, and will be used for HESSI.7 ” A related higher temperature cooler produced under license by Global Cooling Manufacturing, Athens, Ohio, has been incorporated into commercial equipment to test jet fuel by ISL, Verson, France.8 An older high temperature cooler, in nearly continuous operation at Sunpower since 1995, has accumulated to date (June

Figure 1. M77 Cryocooler with Passive Balancer.

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Figure 2. Linear Compressor.

2000) just over 40,000 hours of operation without repair, adjustment, or degradation of performance. In all, this combined experience proves the general design of free-piston machines and the material used, and validates the thermodynamics of the coolers. Linear Compressor

Sunpower has developed a free-piston, fractional horsepower hermetic compressor for use in domestic refrigeration (Figure 2). The design is intended for high volume mass manufacturing and low price (approximately $40 each). The free-piston design is appropriate for the long life expected for refrigeration compressors, 10 to 20 years of service. This reliability will meet the industry need for negligible warranty rates for the industry standard long warranty period (2 to 5 years). These compressors have higher efficiency than conventional compressors9 ,10 for several reasons: the design is mechanically efficient; losses associated with conversion from rotary to linear motion are eliminated; modulated operation through stroke control eliminates frequent on-off surges, losses, and wear. Recent analysis using pre-production units demonstrates that the linear compressor is 30% more efficient than conventional units.11 The very low piston side loads permit either oil-less operation with gas bearings, or the use of low viscosity oil. For manufacturing purposes, the compliance concept2 supports ease of assembly, with no alignment required. This compressor design is entering the final stages in preparation for mass production, with production units expected in 2001.11 While improved performance is demonstrated, note that cost targets must still be met in order for the linear compressor to enter widespread use. M87 Cryocooler Design

To produce the M87 cryocooler design appropriate for manufacture, Sunpower combined its experience with the thermodynamically robust M77 cryocooler and techniques from the linear compressor design that support manufacturing. Additional intellectual property was also developed for this project, and will serve other free-piston applications. Table 1 compares aspects of the M77 and the M87. Figure 3 shows the thermal performance of the M77. Thermal performance for the M87 is shown in Figure 4. Overall, the M87 reduces mass by 23%, increases lift at 77K, and improves rated efficiency by 20%. The M77 is a general-use cryocooler rated at 4 W cooling power (heat lift) at 77 K to 40°C reject with 100 W(e) input. Heat can be rejected through the use of a cooling loop (liquid) or fins (air). Cooler orientation during operation is unrestricted. The controller is operated from a DC

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Figure 3. Thermal performance of M77 cryocooler; reject temperature 40°C (313K).

Figure 4. Thermal performance of M87 cryocooler; reject temperature 55°C (328K).

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Figure 5. A. M87 cryocooler, 3-D cutaway view (left). B. Photograph (right).

source and incorporates a closed loop temperature control. Vibration is controlled through either a passive unit (a mass-spring combination) or an active unit (a driven linear motor). The M87 is

designed specifically for liquefying oxygen (Figure 5). Cooling power is 7.5 W at 77 K and 10 W at 87 K with 150 W(e) input. Reject temperature is 55°C into ambient air through fins. The M87 is orientation dependent (vertical, cold-end down) and has no temperature control. Vibration attenuation is through a passive mass-spring combination. The driver-controller is triac based,12 using 60

Hz 120 V AC for power. The piston is centered for start-up by a DC circuit that lifts the piston to position.13 To achieve a design suitable for manufacture, the M87 part count, the number of weld/brazed joints, and the use of polymers and glue joints were each reduced by 50% or more in comparison to the M77. The M87 uses a single gas bearing system for both the piston and displacer, implemented by means of a greatly simplified design.14 A new design for the heat exchangers and heat path improves heat transfer and simplifies manufacture.15 The regenerator is constructed of a random

fiber material. M87 Manufacturing Facility To support the development of the manufacturing facility, Sunpower recognized that its expertise and practices were unlikely to support a manufacturing start-up effectively. A separate manufacturing facility (6,000 sq ft) was developed with the capacity to manufacture over 30,000 cryocoolers per year with three shifts. An experienced manufacturing engineer was hired to oversee both start-up and manufacturing, and a newly hired, entirely separate team is assigned to M87

manufacturing activities. The M87 is composed nearly entirely of vendor-supplied parts. These are assembled at our facility into five major sub-assemblies (one example in Figure 6). Sub-assemblies are then baked out at controlled temperature under vacuum to remove all contaminating gases, and assembled and then sealed by plasma welding in dry nitrogen (Figure 7). The coolers are then transferred to an evacuation / charging station (Figure 8). Table 2 lists the equipment in place to support these activities. An investment of over $3 million has been required to set up, equip, and debug the manufacturing facility and processes.

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Figure 6. M87 cryocooler sub-assembly.

Figure 7. M87 cryocoolers ready for final weld in glove box.

Figure 8. M87 cryocooler pre-production batch, evacuation and charging station.

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Manufacturing Strategy

The general manufacturing plan is to focus on assembly processes and some specialized testing and finishing processes. Purchased parts (over 90% of machined parts inventory) will be scheduled and received on a just-in-time basis. We will use a batch process. Batch size, initially, will equal one month’s production. When the monthly batch quantity exceeds the vacuum oven batch quantity, the job batch quantities will equal the oven batch capacity. The initial production cycle

for a batch will span two-and-a-half to three calendar months. As a result, there will be multiple active jobs in the system at any given time in order to maintain monthly shipping schedules. Strict document control systems and incoming inspection will be used to maximize efficiency, yield, productivity, and end product quality. Additional on-line inspection of work-in-process quality control sampling of work in progress will be used to minimize scrap and rework. Statistical process control will be implemented over time. Employees will be trained in statistical process control procedures and processes will be prioritized on the basis of potential benefit. OTHER APPLICATIONS FOR LOW COST CRYOCOOLERS

The telecommunications industry is seeking ways to enhance cell phone service while reducing cost by using cooled HTSC components. The M87 can meet some of this need without modification. The design is also well understood, and can be fairly easily modified to enhance performance at other lifts and under differing conditions.16 The manufacturing techniques and intellectual property already developed will allow further manufacturing to occur even more easily. For the telecommunications market, reasonable projections suggest a market by 2005 for at least 120,000 units per year, and perhaps over 400,000 per year. With the availability of mass-produced, low cost cryocoolers, other markets are likely to develop. CONCLUSIONS A new cryocooler, the M87, has been developed and is being placed into production at a specially built manufacturing facility. The cooler combines excellent thermodynamic performance with a volume manufacturing design. The initial and enabling market is a unique oxygen therapy device. Production units will be available in the first quarter of 2001. Price in lots of 10,000 is expected to about $2,000 per unit. With the availability of mass-produced cryocoolers, it can be predicted that other applications will develop for the M87 itself, and for related designs optimized for other specifications, including the telecommunications market.

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ACKNOWLEDGMENT

The authors and Sunpower thank Charles Henry of In-X, Inc., for his early and continuing support for manufacture of the M87 cryocooler. REFERENCES

1. U.S. Patent 5,342,176; Method and Apparatus for Measuring Piston Position in a Free Piston Compressor, 8-30-1994. 2. U.S. Patent 5,496,176; Method and Apparatus for Measuring Piston Position in a Free Piston Compressor, 3-5-1996. 3. U.S. Patent 5,525,845; Fluid Bearing with Compliant Linkage for Centering Reciprocating Bodies, 6-11-1996 4. U.S. Patent 4,602,174; Electromechanical Transducer Particularly Suitable for a Linear Alternator Driven by a Free-Piston Stirling Engine, 7-22-1986. 5. U.S. Patent 5,642,088; Magnet Ring Support, 6-24-97. 6. http://hesperia.gsfc.nasa.gov/hessi/; http://ssl.berkeley.edu/hessi/ 7. http://hessi.ssl.berkeley.edu/instrument/cryocooler.html 8. http://www.isl-france.com/fpp_more.htm 9. Van der Walt, N.R., and R.Z. Unger, “Linear Compressors—A Maturing Technology,” Proceedings 45th International Appliance Technical Conference, University of Wisconsin, Madison, Wisconsin (May 1994). 10. Unger, Reuven, “Development and Testing of a Linear Compressor Sized for the European Market,” Presented at International Appliance Technology Conference, Purdue University, West Lafayette, Indiana, May 10-12, 1999. Available at: http://www.sunpower.com/tech papers/pub74/iatc99.html 11. Lee, H.K., G. Y. Song, E.P. Hong, K.B. Park, J. Y. Yoo, and W.H. Jung, “Development of the Linear Compressor for a Household Refrigerator,” to be presented at the 2000 International Compressor Engineering Conference, Purdue University, West Lafayette, Indiana. (July 25-28, 2000). 12. U.S. Patent 5,592,073; Triac Control Circuit, 1-7-97. 13. DC Centering of Free Piston Machine, U.S. Patent application pending. 14. Gas Bearing and Method of Making a Gas Bearing for a Free Piston Machine, U.S. Patent application pending. 15. Heat Exchanger and Method of Constructing Same, U.S. Patent application pending. 16. SAUCE, Sunpower’s proprietary Stirling design software.

Performance and Reliability Improvements in a Low-Cost Stirling Cycle Cryocooler M. Hanes

Superconductor Technologies Incorporated Santa Barbara, CA 93111

ABSTRACT

The use of a free piston, gas-bearing Stirling cycle cryocooler for commercial high temperature superconductor (HTS) applications dictates the cooler must not only be low cost, but also have long life and high reliability. Over the past two years Superconductor Technologies Inc. (STI) has integrated a new cryocooler design into our commercial HTS systems. This cooler is an improvement over the previous cooler we were using in this application however, further improvements were still attained with this new design, which presently has achieved a 1/COP of <17 w/w at 77K cold end temperature and 23°C ambient temperature. This paper will describe the performance, life, reliability and manufacturing cost improvements made to the cooler design over this time period. Data will be presented detailing the performance of the cooler over an ambient temperature range of 20°C to 70°C at various operating conditions. In addition to laboratory results, a Weibull analysis of life and reliability data from over 140 units in the field with a combined runtime of over 300,000 hours will be presented. INTRODUCTION

Superconductor Technologies Inc. (STI) has as its primary product high temperature superconductors (HTS) which are utilized as filters in cellular applications to provide increased performance over conventional filters. One of the ramifications of utilizing HTS in the filters is the necessity of having to maintain the HTS at cryogenic temperatures. There are a multitude of cryogenic refrigerators which could be considered for this application; STI chose to use a Stirling cycle, free piston, linear motor design. This design provides for a compact, efficient, long life system without the need for helium lines connecting the cold end to the compressor, as with a Gifford McMahon type cooler. If the application involves mounting the HTS products on a

tower, or other remote locations, the elimination of the helium lines becomes a more significant advantage. The Stirling cycle cooler designed for this application must have a long life to allow the cryogenically cooled filters to compete effectively with the conventional, ambient temperature filters. The performance requirements for this application were determined to be 4 watts of lift at

room ambient with less than 100 watts input power, an operating range of -40°C to 60°C, a mass of less than 5 kg. The target for the cooler life is 60,000 hours with no required maintenance.

Cryocoolers 11, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2001

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DESIGN OVERVIEW

The cooler which STI is currently manufacturing and incorporating into its HTS filter systems consists of the following design features: Integral design – the compression and expansion volumes are contained within a common pressure vessel, as opposed to a split design, where the two volumes are connected by a transfer line. Motor – the cooler uses an internal coil linear motor with radially magnetized Neodymium – iron – boron magnets. Displacer – is of the “free piston” design where a pressure difference between two volumes acts on the cross sectional area of a drive piston to provide the correct displacer motion. Regenerator – a non metallic material is used for the regenerator in this design. This provides an inexpensive regenerator which performs well down to below 50 Kelvin. Gas bearing – a gas bearing is utilized to prevent the compressor piston from coming into contact with the compressor cylinder. This eliminates both the wear and debris generation which would result if there is contact between these surfaces. Variations from previous design. The most significant change was to the linear motor. The previous design utilized an external coil with Samarium cobalt magnets and the current design uses an internal coil with magnetized Neodimium – iron – boron magnets. In addition to these fundamental differences in design, the new motor has superior performance. The efficiency of the new motor is approximately 40% higher than the old one, resulting in the same cooler performance with only 100 watts input power as opposed to 140. This is achieved in part

from the internal coil, which allows smaller air gaps in the motor’s magnetic circuit. This also allows for the use of smaller magnets, which reduces cooler cost. Another factor for the increased efficiency is the use of individual laminations for the internal motor iron, as opposed to a solid iron with machined slots. The individual laminations greatly reduce the eddy current losses within the internal iron, which not only reduces input power, but decreases the internal temperature of the cooler. Both designs utilized individual laminations for the outer iron. Another advantage of the latest motor is the lower amount of magnetic side force for a given

amount of misalignment between the motor components. This is especially critical on a design which utilizes gas bearings as the means for preventing contact between the piston and the cylinder. Figure 1 shows the relative magnitudes of the side force vs. the offset. This lower side force not only increases the effectiveness of the gas bearings, but it allows use of less precision parts and for simpler and less time consuming assembly of the cooler, which results in a lower cost cooler. Lastly, the force constant of the new motor is more constant over the displacement of the motor along its axis. This results in a cooler which is easier to design driver electronics for and to use the motor’s back emf to approximate the location of the piston. The difference in the force constants are shown in Figure 2. GAS BEARING DESIGN PRINCIPLES

In order to overcome the magnetic side force and weight of the motor this cooler utilizes a gas bearing scheme to ensure long cooler life. The gas bearing eliminates virtually all contact

between the compressor piston and the compressor cylinder, hence eliminating friction and wear. The piston essentially floats on a thin layer of helium gas, which is the same gas used as the working fluid for the thermodynamic processes within the cooler. A cross section of the piston gas bearing - cylinder assembly is shown below in Figure 3. During cooler operation, the high pressure reservoir is kept at a relatively constant and high pressure by the action of the check valve. During the portion of the cycle where the working pressure in the warm end of the cooler

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Figure 1. Motor magnetic side force vs. magnet offset.

Figure 2. Motor force constant profiles.

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Figure 3. Typical piston gas bearing layout.

is higher than the pressure of the high pressure reservoir, helium flows from the warm end into the reservoir and “recharges” it. During the time when the warm end pressure is lower than the reservoir pressure, the check valve is closed, preventing helium from escaping from the reservoir. During the entire cycle, helium is flowing from the reservoir through the piston flow restrictors and into the bounce volume. The three pressures within the system are shown in Figure 4. As shown on the graph, all the pressures initially start at the same level. As the cooler begins to run, the pressure in the reservoir begins to pump up to an almost constant level. The magnitude of the fluctuation in the reservoir pressure is a function of the reservoir volume and the piston flow restrictor flow rates. Therefore, if these parameters are designed correctly, the gas bearing will operate over an almost constant pressure difference, in spite of the oscillatory nature of the pressure in the warm end, or working, volume of the cooler. Figure 5 shows an expanded piston - cylinder gap to illustrate the principles of a gas bearing supported piston. A piston which is supported by a gas bearing will have the flow resistance of the piston flow restrictor approximately equal to the flow resistance of the annular gap between the piston and the cylinder, when the piston is centered in the cylinder. This results in the

Figure 4. Pressures in various cooler volumes.

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Figure 5. Gas bearing with piston off axis.

pressure in the gas bearing pad being approximately halfway between the reservoir and the

bounce volume pressures. Also, when the piston is centered, the pressures in the pads are equal

on all sides of the piston, and there are no net bearing forces acting on the piston. However, when the piston is forced off center, as depicted in the above drawing, the resistance of “gap 2” becomes lower than that of “gap 1” and the pressure in the gas bearing pad associated with gap 1 increases (becomes more closely coupled to the higher pressure reservoir), while at the same time the pressure in the pad on the opposite side decreases (becomes more closely coupled to the lower pressure bounce volume). This results in a pressure difference between the two sides of the piston, which act upon the projected area of the piston to provide a centering force. Since the flow resistance of the gap is proportional to the inverse of the gap width cubed, large pressure differences will exist for very small piston offsets. This self centering gas bearing will have a spring constant in the range of 10,000 lb/in per set of gas bearings, which is adequate to prevent

piston to cylinder contact, and ensure the longevity of the cooler. IMPROVEMENTS TO CURRENT DESIGN

Since the incorporation of the new cryocooler design there have been several improvements to the design which are intended to improve the efficiency and/or the reliability of the cooler. Two of the most significant changes are discussed below. Piston centering port modification. During operation there is a continuous flow of helium out of the warm end volume of the cooler, into the gas bearing reservoir, out the gas bearing restrictors and into the bounce volume (see Figure 3). If this flow was not compensated for, the average pressure in the bounce volume would become higher than the average pressure in the warm end volume. The effect of this would be to bias the piston toward the warm end volume and cause it to hit the compressor cylinder at an unacceptably low piston displacement (or input power). To offset this flow of gas from the working volume to the bounce volume and “centering port” circuit is used. This circuit allows gas to flow back from the bounce volume to

the working volume. In the initial design, under certain operating conditions, this circuit was not allowing enough gas to return to the working volume. This resulted in the cooler knocking (piston hitting the compressor bore) at a low input power. The centering port circuit was modified to keep the average pressures on the two sides of the piston more closely matched. The results of this modification are shown in Figure 6. The vertical axis shows the average (or midstroke) location of the piston as a function of input power. As can be seen, with the “no mod” centering ports, the piston drifts toward the bore as soon as power is applied to the cooler.

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compressor bore at the extremity of its displacement. With the modified centering port circuit, the piston initially drifts away from the compressor bore, until at 80 watts, it has returned to the zero power location. Under the conditions which the cooler normally operates, 80 watts is typically the max power required and the redesigned centering ports have the piston centered in the motor at this point, which provides the highest operating efficiency. When higher input power is required, such as at high ambient temperatures, dewar degrade or cooler efficiency degrade, the piston will not hit the compressor bore before the motor reaches its maximum input power. The details of this design are not shown here as a patent has been applied for and not yet issued. Cold end efficiency improvement. A key component to the overall efficiency of a cryocooler is how well heat is transferred from the heat load to the helium in the cold end of the cooler. The more efficiently this heat transfer occurs, the lower the temperature difference between the heat load and the helium in the cold end will be. This results in the helium not having to be as cold,

and, hence, a more efficient cooler. An effort was undertaken to increase this cold end heat transfer efficiency. There are essentially three resistances to this flow of heat; 1) the interface resistance between the heat load and the heat acceptor portion of the cold end, 2) the conduction

resistance across the heat acceptor, and 3) the convection resistance from the heat acceptor to the helium gas inside the cold end. It is this last resistance which is the dominant term in the overall resistance, therefor, this was the area which was to be improved upon. This resistance is defined as 1/(hA), where h = convective heat transfer coefficient, and A = heat transfer area. Increasing either of these parameters will reduce the heat transfer resistance. The modifications which were

made to the cold end increased both of these parameters. Again, the details of this design are not shown here as a patent has been applied for and not yet issued.

Regenerator material change. This cooler used a non metallic material for the regenerator. Although this is an inexpensive method to achieve an effective regenerator, the variation in the characteristics, particularly the diameter, of the purchased material was causing variations in the

Figure 6. Piston drift vs. input power.

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performance from cooler to cooler. This occurred as a result of the variation of the regenerator effectiveness and from the variation in pressure drop across the regenerator, which, in a “free piston” displacer cryocooler, will have an effect on the compressor piston to displacer phasing, which affects cooler efficiency. An effort was made to find a similar material which would not only be more consistent, but also increase the regenerator effectiveness. In order to increase the regenerator effectiveness a material with a smaller diameter was desired, as this would increase the surface area for a given regenerator mass. A new material was found, and it did meet the requirements which were desired. The results of the new regenerator material, cold end efficiency improvement and the modified centering port design are shown in Figure 7. The modification to the centering ports allow the cooler “with mods” to run at the higher power and the new regenerator material and higher efficiency cold end account for the higher lift. At 100 watts input power the new regenerator material accounts for ~0.5 watt increase and the high efficiency cold end adds the other 1.5 watt increase. COOLER RELIABILITY

The goal for life is over 60,000 hours (5 years) of continuous operation with no maintenance requirements. We currently have about 140 units in the field with a combined run time of over

490,000 thousand hours with no failures. Figure 8 shows a Weibull analysis of the coolers in the field as of 5/9/00. This analysis predicts a 56,000 hour characteristic life based on the current data. The slope of the curve on the Weibull graph (beta) is based on the types of failures which have occurred. Since there are no failures to determine this value, it is based on the failures of the previous design. Although this analysis is not complete, it does show that the goal of a 60,000 hour cooler life is attainable. SUMMARY Cooler performance. Table 1 is a summary of the cooler characteristics for a range of parameters and Figure 9 shows the performance of the cooler at 23 °C and 60°C ambient temperature over a range of input power. The COP of the cooler at 100 watt input and 23° C is 0.063 watts lift / watt input power. The Carnot efficiency at these conditions is 0.35, which yields an efficiency of 18% of Carnot.

Figure 7. Cooler lift vs. input power, current and original design.

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Figure 8. Weibull analysis of current coolers in the field.

IMPROVEMENTS IN A LOW-COST STIRLING CRYOCOOLER

Figure 9. Cooler performance summary graph.

Conclusion. The cooler as originally designed provided adequate performance for the application it was intended to be used in, however the performance improvements not only reduce the power required by the HTS system, it also increases the life of the cooler in the event of failure mode which results in a slow degrade in performance. Future improvements to the cooler include incorporating a gas bearing system on the displacer. Although very little wear has been observed on the displacer in the current design, it is the only part in the cooler which has

moving contact during operation.

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Development of a Long-Life Stirling Cryocooler Y. Ikuta, Y. Suzuki, K. Kanao and N. Watanabe

Research and Development Center Sumitomo Heavy Industries, Ltd.

63-30, Yuhigaoka, Hiratsuka, Kanagawa, 254-0806, Japan

ABSTRACT

A Stirling cryocooler for high temperature superconducting devices has been developed. High efficiency and long lifetime are emphasized in the development. Its performance is 6 W at 70 K with 150 W input under 23°C ambient conditions; at 50°C ambient, it produces 5 W at 70 K with

150 W consumption. The mean time before failure for the design is over 40,000 hours. Flexure springs are used to support the compressor cylinders and the displacer in the cold head to realize

this lifetime. The flexure springs are designed using finite element modeling (FEM) to reduce the

stress to below the fatigue limit. INTRODUCTION

High Temperature Superconducting (HTS) filters (as used in ground base stations for cellular phone systems) are passive devices and must be cooled to 60-80 K. The needed cooling power is expected to be 2-5 W at 70 K. Over 5 years lifetime is required.

Sumitomo Heavy Industries is a leading Stirling cryocooler manufacturer in Japan. The coolers have been developed and manufactured since the 1980s, and are mainly used for cooling sensors, especially infrared sensors. The sensor systems require a cooling power of 0.3-1.5 W at 80 K. In 1998, development of a new cryocooler with more cooling power and high reliability was initiated for the new HTS markets. SPECIFICATION AND DESIGN

The main cooler specifications are listed in the Table 1.

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To achieve these specifications, the development process was divided into two phases. In the first phase, the proto1 cooler was designed using a combination of a conventional linear motor and a conventional free-displacer cold head, and without using long-life technologies; it was made to examine the cooling power of the design. In the second phase, the proto2 cooler was made, adding long-life technologies and a high efficiency linear motor to the protol design. Cooling Power Design

The target of the cooling power was set at 5 W at 70 K under 23°C ambient conditions. This is the expected requirement for HTS devices. Design calculations were based on Walker’s equations, Schdmit’s equations1 and our experience. The results are considered as the average cooling power at the beginning of life. The guaranteed cooling power will take into consideration the results of lifetime qualification and the variation of cooling power in mass production. The target for the input power was set at less than 150 W. The reason is the difficulty of heat rejection from the cooler surface to the atmosphere. In the test of the proto1 with 250 W input, the surface temperature of the cooler was 50°C higher than the ambient. Generally speaking, if the cooler is designed smaller, the input power becomes larger. That makes the heat rejection more difficult and the heat sinks larger. In other words, a low input power cooler makes the heat rejection of the whole system easier. Mechanical Design for Longer Lifetime

Figure 1 illustrates the overall mechanical design of the cooler. The main factors affecting cooler lifetime include: piston and displacer seal wear, gas contamination, and gas leakage.

Prevention of Seal Wear In the proto2 design, a linear motor was incorporated to reduce the load on the piston seal. The moving parts in the proto2 compressor consist of the cylinders and coils supported by flexure

springs at both ends of the moving elements.2 This structure has the advantage of keeping the gap

between the pistons and the cylinders constant, but also has the disadvantage of making the system heavier. From the viewpoint of reliability, the moving cylinder structure was adopted. The displacer in the cold head is also supported by flexure springs, which are located at two points along

the rod that extends to the room temperature side.

Figure 1. Schematic of the proto2 cooler.

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The flexure springs were designed using FEM to reduce the stress to below the fatigue limit. Ordinary fatigue characteristics of materials are generally tested to less than cycles. On the other hand, our lifetime target is 40,000 hours; that means over cycles. In the region over 107 cycles, what happens is unknown. Conventional fatigue design theory has not yet been evaluated in this region. Therefore, fatigue tests were carried out with various flexures. A variety of shapes of springs with various stresses were tested. All the spring were designed using the same method, and manufactured using photo-etching. In the flexure life test, linear motors are used to force the springs to displace at the same amplitude and speed as in the cooler. The tests are being performed in atmosphere, at room temperature, and the experimental apparatus is cooled by a small fan to avoid overheating. There is

no vibration except the basic mode. To date, all ten samples of the flexure springs have survived one year of running, and the test is still ongoing. Other flexure springs with higher stresses broke earlier as expected.

Prevention of Gas Contamination Gas contamination can result from outgassing from internal cooler components, especially from nonmetallic materials used as insulators. In our design, two directions have been tried to reduce outgassing. One is minimization of the use of nonmetallic materials; the other is avoiding

overheating. The latter direction is based on the thought that the hotter the inside of the cooler, the more gases come out of the materials. Therefore, reducing the input power to the linear motors was tried.

Prevention of Gas Leakage The problem of gas leakage has been addressed by using a double seal structure. If this is not enough, a welded structure will be adopted as an alternative. CHARACTERISTICS OF THE COOLERS

In this section, characteristics of the developed coolers are described. Measurements of the dependency of thermal performance on ambient temperature were performed after first making

sure that the surface temperature of the compressor was equal to the ambient temperature. Figure 2 is a photograph of the proto2 cooler in the thermal chamber.

Figure 2. Photograph of the proto2 cooler in the thermal chamber.

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Figure 3. Cooling power of the proto2 cooler.

Figure 4. Ambient temperature dependency of the cooling power at 60 K.

Ambient Temperature Dependency

The experimental results, measured according to the test conditions in Table 2, are summarized in Fig. 3. The cooling power is the heat dissapated by an electric resistance heater on the cold stage; measurements were made using the 4-wire method. The design cooling power is 5 W at 70 K, but the achieved performance is 6.0 W at 70 K, with a COP of 3.9% (23°C ambient). In Figs. 4 to 6, the relationship between the cooling power and the ambient operating temperature of the proto1 and proto2 coolers is shown. From these results, the proto2 cooler is seen to have

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Figure 5. Ambient temperature dependency of the cooling power at 70 K.

Figure 6. Ambient temperature dependency of the cooling power at 80 K.

1.5 times more cooling power than the proto1 cooler at 23°C. The reasons responsible for this improvement are the reduction of the dead volume in the cold head and improvements made to the linear motor. The effect of the dead volume reduction is confirmed by our design calculations. Also, the cooling power in cold ambient environments was improved; this relates to the thickness of the compressor piston clearance seal. The clearance seal of the proto2 cooler is thinner than that of the proto1 cooler.

CONCLUSIONS AND FUTURE WORK The development target for cooling power has been satisfied. The COP data for the cooler are listed in Table 3, where the COP is defined at each temperature as the ratio of the heater input power to the input power to the cooler (not to the drive electronics). From Table 3, it is seen that world class performance was achieved. The mechanical design has also been verified via the flexure spring fatigue test. Measuring the cooler's resistance to contamination remains to be done. After this, the design lifetime of 40,000 hours will have been established.

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REFERENCES 1. Walker, G., Cryocoolers, Plenum Press, New York (1983). 2. Aubon, C.R. and Peters, N.R., “Miniature Long-life Tactical Stirling Cryocoolers,” Cryocoolers 9, Plenum Press, New York (1997), pp. 109-11

Flexure Springs Applied to Low-Cost Linear Drive Cryocoolers R.M. Rawlings and S. Miskimins

DRS Infrared Technologies, L.P. Dallas, Texas USA 75243

ABSTRACT

Flexure spring suspensions have demonstrated the ability to provide long operating lifetimes for cryocoolers intended for space-based applications. Insertion of this technology into coolers intended for tactical or commercial application has been slow due to cost considerations. This paper describes the development and testing of a flexure spring system for small tactical cryocoolers that provides a doubling of operating life while costing approximately the same as the traditional helical coil spring suspension system. The flexure spring system described in this paper successfully achieves the high radial stiffness characteristic of the flexure spring design in a low-cost package. In addition, the concept has been implemented in cryocoolers weighing less than a pound and smaller than a soft drink can. This design has been qualified for use in U.S. Army applications. Qualification and life test data is presented to demonstrate the robustness of the design in tactical environments. The producibility of the design is evidenced by the on-going production of these coolers for various applications. INTRODUCTION

For a number of years the development of long-life cryocoolers for spacecraft applications has been reported in the literature.1,2 These coolers have all shared a design heritage from the Oxford Stirling cooler; namely flexure springs or bearings as they have been called. The specific implementation or configuration varied from investigator to investigator, but the use of these springs to suspend the moving masses within the compressor was universal. Also consistent from unit to unit was the high cost of producing these “space” coolers, some perhaps approaching the

$1-million figure. At the same time, tactical cooler developers were designing and producing linear drive cryocoolers for sensor systems designed for use in day-to-day battle operations. These units have shared design philosophies that emphasize low weight, low input power, and low cost. Lifetime requirements have been an order of magnitude less than the requirements imposed on the space coolers. These tactical cryocoolers have provided improvements in performance, induced vibration, audible noise, and operating life when compared to the earlier rotary cooler designs they replaced. These improvements, however, came with an increase in acquisition costs. In 1995 the U.S. Army, through the Night Vision and Electronics Sensors Directorate (NVESD) began a Linear Cooler Manufacturing Technology program with DRS Infrared TechCryocoolers 11, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2001

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nologies (a part of Texas Instruments at that time) to reduce the cost of linear drive coolers for U.S Department of Defense programs. The goal of the program was to reduce the acquisition cost of linear coolers by 30 percent. As a part of on-going development at DRS Infrared Technologies, initial prototypes of flexure spring compressors with the same form factors as standard tactical coolers had been designed and built. It was proposed that a flexure spring compressor for tactical cryocoolers would offer a life cycle cost reduction due to the increase in life that it would provide. The potential for maintaining or even reducing the acquisition cost was also recognized, if the design were approached accordingly. The flexure spring compressor initiatives were added to the Manufacturing Technology program with the specific objectives of developing flexure spring compressors for the Javelin system’s Command and Launch Unit (CLU) and the U.S. Army’s Standard Advanced Dewar Assembly (SADA)-II The balance of this paper describes the design considerations, the resulting designs, and test results for those designs. These results, coupled with the on-going production for these programs, illustrate the success of the effort. DESIGN CONSIDERATIONS The cooler development objectives required that a number of factors be considered during the

design phase. As is the case in any design, some of these factors tend to be mutually exclusive. Some carry more weight than others, and the importance of each factor may vary as the design progresses and additional issues arise and are addressed. The primary considerations in undertaking the designs for these two coolers included: •

Ease of Assembly – During the design phase an overriding consideration was the ease of assembly. These cooler designs have been developed for production programs, not prototype

or single unit installations. As an example, the Javelin program will build and deliver nearly 5,000 nightsight systems in a six-year period. Cooler deliveries to support this program will

•

require 70 coolers per month- nearly 20 per week. Such rates cannot be attained or sustained if the subassemblies and final assembly require significant effort to assemble, align, or adjust. Compatibility with existing designs - DRS Infrared Technology has been supplying linear drive cryocoolers since the mid-1980s to a variety programs. There is now a significant installed base of these coolers. The using programs have developed systems around these designs; allowing a specific space envelope and weight budget, and providing well-defined

electrical and thermal interfaces to the cooler. Design modifications to these coolers must maintain a backwards compatibility with these fielded systems. New cryocooler designs cannot require more space, different electrical characteristics, or more thermal management. This forces the flexure spring cooler design to be form, fit, and function interchangeable with the previously fielded helical spring designs. • Less than benign installations – Tactical cryocoolers are not destined to spend their lifetimes quietly producing refrigeration in the corner of some laboratory. They are developed to provide cooling in systems that will be subject to wind buffet at Mach 1, tracked vehicle vibration, roof mounted on an off-road personnel carrier, or slung over the shoulder of infantry personnel as they work their way through rugged terrain. Above all else, the flexure spring design for tactical cryocoolers had to consider the handling and operational scenarios to which the cooler would eventually be subject. The design had to take into account mechanical shocks, operation in extreme ambient temperatures, exposure to long-term external vibration sources, and handling in battlefield conditions. The design must be mechanically robust. •

Cost Impact – As second-generation night vision systems have matured, investments have been made by the using community, most notably the U.S. Army Night Vision Laboratory and Defense Advanced Research Projects Agency (DARPA), along with the component manufacturers, to drive down life cycle costs, including cost of acquisition. Design changes to the cryocooler are improvements only if the net result is a reduction in total life cost of ownership

for the user. The implementation of flexure springs into the tactical cooler designs had to

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result in a reduction in the total cost of ownership. Some increase in the cost of acquisition could be offset by the increase in operating life; however, DRS’ goal was to implement flexure springs into the tactical designs with no increase in cost of acquisition; resulting in a reduction in the total cost of ownership for the using programs. • Operating Lifetime Improvement – DRS’ standard helical coil spring cooler designs have repeatedly demonstrated lifetimes in excess of the 4,000-hour operational requirement.3 As a goal, the incorporation of flexure supports into the compressor designs was designed to produce a doubling of this lifetime. Throughout the design process, this goal was employed to drive design decisions.

FLEXURE SPRING DESIGN AND COMPRESSOR CONSTRUCTION

Compressor Design

Given the constraints identified in the previous section, the task remained to develop a flexure design that would provide the required axial spring rates, provide the radial stiffness characteristics that are the primary benefit of the flexure spring, and fit within the space envelope of the existing compressors. The design process that was implemented was iterative in nature. The first step was to identify a suspension configuration that would fit within the length constraints. Since the requirements for the design included working with the installed base of cooler drive electronics, the electrical characteristics of the compressor could not be significantly altered. The factors such as induction, coil resistance, back EMF, etc. had to be unchanged. Traditional flexure suspensions are fairly long in nature, to provide maximum radial stiffness. The overall length of the tactical cooler compressor (less than 5 inches for a dual-opposed piston design) necessitated a short spring system. After a basic design philosophy for the suspension system (hereafter referred to as the spring stack) was determined, the actual design of the flexure was undertaken. This again was an iterative process, attempting to balance out the conflicting design parameters of spring rate, radial stiffness, maximum allowed piston stroke, and induced stress in the material. The balancing act had to arrive at a stack design that would produce the same axial spring force as the helical coil spring suspension system (to maintain the system’s resonant frequency) while not overstressing the spring material. In addition, for cost considerations the number of flexures in each stack was desired to be kept to a minimum. Each additional spring would add cost in both material cost and increase the time required for assembly and complexity of the alignment process. The detail design of the flexure was also driven by cost considerations:

• Standard design rules for photo-etch processing were followed to ensure that the flexures could be produced by a variety of photo-etch vendors using standard industry processes. • The material chosen for the flexures was readily available blued and tempered spring steel in stock thicknesses and widths, again to insure availability and lower cost. When building a limited number of coolers, such cost considerations might be secondary. With the production rates required for these coolers, piece part cost and availability becomes a major factor. As a reference point, to support the Javelin cooler build requirements, over 15,000

individual flexure springs need to be procured each year; they must be mass-producible. The final design of the flexure for the 1/5-watt compressor is shown in Figure 1. Comparison with the adjacent coin illustrates the small geometries that were required for this design. The compressor housing is only 1.5 inches in diameter. The final stack design then used the flexure design and addressed the remaining issues of retaining the same spring-mass-damper characteristics of the helical coil design. Overall mass was adjusted and various design modifications were made to adjust the center of gravity of the moving components (to minimized the potential for side-loading of the clearance seals and/or orientation sensitivity of the compressor). The completed stack assembly is shown in Figure 2.

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Figure 1. Javelin compressor flexure spring.

Assembly Processes

The compressor assembly process begins with the spring stacks. These are assembled in fixtures that locate the various components in relationship to one another as each component is

added to the assembly. Once all the comonents have been assembled they are clamped in place. Completed stacks are then bonded to the magnet assemblies to complete the mechanism assembly. Mechanisms are subjected to testing for stiction, to verify proper alignment of the stack, and basic compressor output functions, to verify proper operation when mated to the expander assembly. Following this testing, the compressor are welded into their housings, mated with an expander assembly, and eventually delivered to the using programs.

Figure 2. Javelin compressor flexure spring stack assembly.

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ENVIRONMENTAL AND LIFE TESTING

The flexure spring compressors discussed in the previous sections were designed as replacements for cryocoolers already in production and fielded in various systems. The currently fielded designs have been proven through numerous qualification and life test series, in addition to successful operation in the field. Prior to switching to the flexure design, some level of testing, validation, and evaluation is required to verify that the new designs are suitable for the intended applications. Once the designs were complete, prototype units were assembled for engineering characterization and evaluation. After it was determined that the designs were performing as intended, additional compressors were assembled for environmental qualification and life testing. The first design to be tested was the 1/5-watt compressor, specifically in the Javelin configuration. The series of environmental tests performed on the three test units is shown in Table 1.

At the same time, three additional units were started into a life test to verify reliability improvements. These units were mated to the Javelin detector-Dewar assemblies to provide the actual refrigeration demand seen by the cooler in its application. The test cycle for the life test is illustrated in Figure 3. As can be seen, the cycle includes thermal soaks and operation at temperatures ranging from –32C to +52C. These coolers are still running in test and have, at this time, accumulated over 21,000 hours of operation for a current mean-time-to-failure (MTTF) in excess of 7,000 hours. As the testing continues the MTTF grows at the rate of about 600 hours per month. Figure 4 shows the performance of one of the life test coolers since the inception of the test. The data clearly shows that the cooler performance has been extremely stable and no degradation has occurred. The one-watt design was undertaken after the 1/5-watt design and has not progressed as far through the qualification and life test process to date. The qualification testing process is very similar to that of the 1/5-watt design. The test sequence is listed in Table 2. One of the primary concerns during the design of the 1-watt flexure compressor was the extreme magnitude (250 g’s over 10 milliseconds) of the mechanical shock that the unit must survive as a part of the qualification procedure. To mitigate the risk, several early prototypes were assembled and subjected to the shock requirement. The prototype units met the requirement and the design moved forward. During the qualification testing, however, externally applied mechanical shock and vibration proved to be a stumbling block for the flexure design. Analysis of the problem led to a modification of the flexure stack assembly, which is now being prototyped for re-testing.

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Figure 3. Reliability Test Cycle.

Figure 4. Javelin flexure cooler reliability test data.

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Figure 5.1-watt SADA-II cooler reliability test data.

As is the case with the 1/5-watt flexure design, 1-watt flexure coolers are also in a life test. In

this instance, four units are running in test. The test profile is the same (Figure 3). These coolers have not been in test as long as the 1/5-watt units but have accumulated over 19,000 total operational hours for a current MTTF approaching 5,000 hours. These coolers are also accumulating additional MTTF at the rate of approximately 600 hours per month. Figure 5 shows the performance of one of the test units since the start of the test. The anomalies in performance that are seen at various points in the graph are caused by test equipment problems, demonstrating that the coolers are more reliable than the environmental test chamber and computer-controlled test equipment.

SUMMARY AND CONCLUSIONS This paper has detailed the development, test, and performance of flexure spring designs

adapted to low-cost tactical cryocoolers. The adaptation has successfully combined the flexure spring technology utilized to provide long-life for space coolers with the small-sized, low-cost compressors used in tactical thermal sensor systems. The compressors incorporate flexure springs for long-life but have been designed such that the cost of implementation is on the same order as that for the standard helical coil spring designs. These new designs are being proven through qualification testing. The true measure of the success of the program has been the introduction of the 1/5-watt flexure design into the current Javelin production program. These coolers are being built at a rate of nearly 70 units per month, demonstrating that the goals for ease of assembly have been achieved. The continuing life tests have demonstrated reliability improvements under demanding operational scenarios and environments. ACKNOWLEDGMENT Much of the work described in this paper was performed by DRS Infrared Technologies under contract to the U.S. Army Night Vision and Electronic Sensors Directorate , Contract DAAB07-95-C-J513.

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REFERENCES 1

Jones, B.G., Development for Space Use of BAe’s Improved Single-Stage Stirling Cycle Cooler for Applications in the Range 50-80K,” Cryocoolers 8, Plenum Publishing Corp., New York, 1995, pp. 1-11.

2

Nast, T.C., et.al., “Design, Performance, and Testing of the Lockheed-Developed Mechanical Cryocooler,” Cryocoolers 8, Plenum Publishing Corp., New York, 1995, pp. 55-67.

3

Rawlings, R.M., Granger, C.E., and Hinrichs, G.W., “Linear Drive Stirling Cryocoolers: Qualification and Life Testing Results”, Cryocoolers 8, Plenum Publishing Corp., New York, 1995, pp. 121127.

High Reliability Coolers under Development at Signaal-USFA M. Meijers, A. A. J. Benschop and J.C. Mullié

Signaal-USFA Eindhoven, The Netherlands

ABSTRACT

Since 1997 Signaal Usfa has been working on the development of high reliability cryocoolers. These coolers have been developed with the objective of eliminating the lifetime determining factors of conventional cryogenic tactical coolers. The intention of this study was the development of a family of cryocoolers that could be used to cover a large range of cooling powers. Today, these developments have resulted in a new range of flexure-bearing cryocoolers currently

available at Signaal Usfa, with cooling performances between 0.5 and 3 W at 80K and estimated lifetimes of more than 20,000 hours MTTF.

The basis for the extended lifetime of these coolers is our unique flexure-bearing compressor with moving-magnet technology. Inside the linear dual-opposed-piston compressor both moving pistons are fully supported at the back and front of the piston by optimized flexure bearings. With this flexure-bearing suspension, side loads on the piston seals are avoided. A proper alignment procedure ensures no contact between piston and cylinder during operation, resulting in the absence of wear of the piston coating, which normally determines the lifetime of a cooler.

The moving-magnet technology, as applied in our flexure-bearing compressor, has several major advantages over moving-coil linear motors as applied in most conventional linear compressors. First of all, the coils, known to be a possible source for gas contamination, can be placed outside the hermetically-sealed compressor housing containing the working gas. Avoiding any synthetic materials inside the cooler reduces the risk of gas contamination during the life of the cooler. The fact that the coils can be placed outside the hermetically-sealed compressor also means that no glass feedthroughs are required. In this way, risks of glass feedthrough leakage due to extreme temperature shocks or mechanical shocks are no longer present. Finally, the absence of moving coils in the compressor design also means that flying leads to supply power to the

coils are no longer present. Several qualification tests have been performed on flexure-bearing cryocoolers with different sizes of coldfingers, resulting in technical specifications currently available at Signaal Usfa

on all presented cooler types in the above-mentioned range. A flexure-bearing cooler for applications requiring over 6 watts cooling power at 80K is currently under development as well. Tests performed on a first prototype of this cooler have been successful, with a measured cooling performance of more than 8W at 80K for a 23°C ambient temperature, with 150 Wac of input power.

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LIFETIME LIMITATIONS OF A LINEAR STIRLING CRYOCOOLER

In conventional linear Stirling cryocoolers the lifetime of the cooler is limited by the compressor. Inside a linear compressor usually two pistons, driven by moving coil linear motors, are translating in opposite phase, generating a pressure wave in the compression space between the pistons, which is connected to the warm end of the coldfinger. The magnitude of the generated pressure wave directly determines the cooling performance obtained with a certain Stirling coldfinger or pulse tube. To prevent gas leakage along the pistons inside the compressor, which reduces the generated pressure wave and thus the cooling performance, close tolerance seals are often applied. The principle of close tolerance seals is a very small annular gap between the piston and the cylinder, typically with a length of a few centimeters, which prevents gas from flowing from the compression space to the larger (buffer) space behind the pistons. To limit the gas flow along the pistons to an acceptable level, the gap between the piston and the cylinder should be as small as possible, but still allow piston movement at different ambient temperatures between –52°C and +71°C. In practice, the initial gap between the piston and cyl-

inder will be around 10 microns. Increase of this gap due to wear of the coating applied on the piston increases the flow along the piston. In fact, the gas flow through the annular gap between the piston and cylinder is dependent on the gap height to the third power. To determine the impact of an increase of the gap height between piston and cylinder on the efficiency of the cooler, several measurements were performed on a Stirling cooler at Signaal Usfa. In Fig. 1, the results of these measurements are depicted for a Stirling cooler with a 5 mm cold finger. In this figure a graph is shown in which the measured cooling performance is plotted against the input power to the cooler. Data are presented for the standard initial gap of 10 microns between piston and cylinder, as well as for larger gaps of 17 microns and 20 microns. One can clearly see the large impact of the flow losses along the piston on the efficiency of the cooler. From these measurements it was concluded that the specified cooling performance is no longer met for piston diameter reductions of more than about 10 microns.

Figure 1. Results of measurements to investigate the impact of the gap between piston and cylinder on the efficiency of the cooler.

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HIGH RELIABILITY COOLERS AT SIGNAAL-USFA: A HISTORICAL OVERVIEW

Lifetime tests on various cooler types at Signaal Usfa have shown that, with conventional compressor wear of a PTFE-based coating due to contact between piston and cylinder, the life is limited to about 4,000 to 10,000 hours MTTF, depending of the cooler type. In order to ensure

higher MTTF values of more than 20,000 hours, wear of the pistons should be completely eliminated. The only way to ensure this is by avoiding contact between piston and cylinder. Study of possibilities to avoid contact between piston and cylinder (1997)

In 1997 Signaal Usfa initiated an internal development program investigating different possibilities to increase the lifetime of linear Stirling cryocoolers. The key goal in this development was to avoid contact between piston and cylinder, as this was considered the only possibility to obtain lifetimes of more than 20,000 hours. Two different principles with which this could be achieved were studied and worked out in different compressor designs. In the end, these compressor designs were compared on a number of criteria like: compressor efficiency, design complexity, compressor size and weight, and cost/price of the cooler. The first principle studied was the application of gas bearings. With gas bearings, the piston and cylinder are separated by a thin layer of gas under pressure. Two different ways in which the build-up of gas pressure can be achieved, often referred to as dynamic gas bearings and static gas bearings, were studied and worked out in detail. Although gas bearings are applied by several manufacturers of linear compressors, we concluded in a final comparison that they scored worst on almost all the criteria mentioned above. The main disadvantages with dynamic gas bearings were the presence of an extra motor to rotate the piston, and the connection of the rotating piston to the compressor housing. With the static gas principle, the reduction of efficiency due to the extra flow losses and the unavoidable pumpup effect in the compressor led to the decision not to continue with this design principle either.

Contact between piston and cylinder can also be avoided when the moving piston is suspended by mechanical springs that offer a high radial stiffness and allow easy movement in the axial direction. This can be achieved with several types of springs, but we have studied a suspension in three folded leaf-springs at both piston ends and a suspension with flexure bearings at both piston ends. Again, both possibilities were worked out in detail and, after several tests were

performed on both the folded leaf-springs and the flexure bearings, the conclusion was drawn that, on all criteria mentioned above, flexure bearings are the most interesting solution. First prototype flexure-bearing cryocooler (1998)

Based on studies performed in 1997 a first prototype flexure-bearing cooler was build and tested in 1998. Figure 2 shows a picture of this cooler.

Figure 2. First prototype flexure-bearing cryocooler (1998).

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With this first prototype flexure-bearing compressor, which was connected to a 7mm Stirling coldfinger, several tests have been performed that prove that the flexure-bearing principle

and the specially developed alignment procedure are working correctly. As can be seen in Fig. 2, the compressor itself is rather bulky due to the presence of flanges and bolt connections incorporated to allow the compressor to be sealed with O-rings. However, this enabled us to dismount the compressor, to inspect components after several tests, and to verify that the flexure-bearing suspension was working as expected. Although several imperfections in the compressor design were found, the fact that the compressor worked correctly after the first assembly, and the working principle of the flexure-bearing suspension was verified, was very promising. At the end of 1998 the cooler was put in life test;

up to now (May 2000) it has accumulated over 12,000 hours of operation without degradation of performance. The cooler has always been running at 80% of its maximum power generating a cooling power of approximately 750 mW at 80K, with a 23°C ambient temperature. Redesign of the first prototype flexure-bearing compressor (1999)

In the beginning of 1999 the flexure-bearing compressor was completely redesigned. Based on the experience obtained from tests on the first prototype several small design changes were introduced to simplify the compressor assembly. Also the design was made more compact with the objective of hermetically sealing the compressor by laserwelding to allow a complete qualification test program to be performed. Beside this, the flexures were also redesigned by increasing the axial spring stiffness to limit piston drift under all circumstances. To supply current to the moving coils inside a linear compressor, flying leads are commonly

applied; these are electrical wires connected to the glass feedthroughs in the compressor housing at one end, and the coils of the linear motor at the other end. However, the flying-lead connections are critical with respect to reliability and require high precision in mounting during produc-

tion. In the redesign of the flexure-bearing compressor, an innovative solution was introduced to replace the flying leads.

Based on the experience obtained with flexure-bearing and Finite Element Modeling (FEM) calculations, special spiral arms were designed to supply current to the moving coils in the redesigned flexure-bearing compressor. These “flexible leads” are very easy to mount during the compressor assembly, and can be designed with FEM tools to ensure the absence of failures. To illustrate the principle, Figs. 3 and 4 show the FEM model of a flexible lead and a photograph of a flexible lead mounted on a coil assembly.

Figure 3. Photograph of a flexible lead

mounted on a coil assembly (left).

Figure 4. FEM model of a flexible lead (right).

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Several flexure-bearing coolers have been built with 10mm Stirling coldfingers and submitted to various qualification tests. Cooling performance levels measured at ambient temperatures ranging from -52°C to +71°C were excellent and showed that the efficiency of the cooler was according to expectations. The length of the compressor is 180 mm with a diameter of 75 mm; the resulting total cooler weight is 3.5 kg. Two coolers were put in a lifetime test at the beginning of 2000 and are still running without degradation. Development of a compressor with moving magnet motor technology (1999) Parallel with the redesign of the first prototype flexure-bearing compressor in 1999, the development of a compressor with moving-magnet technology was also initiated at Signaal Usfa. With a moving-magnet linear motor, the magnets are connected directly to the pistons and the coilholders are part of the compressor housing. This offers a number of advantages that increase reliability of the cooler and allow the compressor design to be more compact. First of all, the fact that the coils are no longer moving means that flying leads or flexible leads are no longer necessary to supply current to the coils. The absence of these components

simplifies the compressor design and assembly, and reduces the compressor length. In the final flexure-bearing compressor design, the coil-holder on which the coils are wound is pan of the housing of the compressor. This means that the coils are outside this housing and outside of the working gas. As the coil insulation consists of a synthetic material that absorbs moisture, bake out of components and curing of the cooler under high vacuum at elevated temperatures is probably the most critical process in the cooler production. With the removal of the coils outside the working gas, no more outgassing-components are present inside the cooler. This reduces the risk of gas contamination during the life of the cooler. Finally, the fact that the coils are located outside the hermetically-sealed compressor means that also glass feedthroughs are no longer required. Glass feedthroughs are known to be critical components that can crack under extreme temperature shocks and severe mechanical stresses, resulting in gas leakage. All these advantages have lead to the conclusion that, for a high reliability cryocooler, moving magnet technology in combination with a flexure-bearing suspension is the best solution to guarantee long lifetimes. Final design of the flexure-bearing cryocooler with moving magnets (1999/2000) In the final design of the flexure-bearing compressor, we have been able to reduce the compressor diameter to 60 mm and the length to 165 mm, resulting in a cooler mass of 2.4 kg. FEM calculations performed on the flexures have lead to an optimized design with a high radial stiffness and low peak stress levels. FEM calculations performed on the moving magnet linear motor have lead to an optimized motor with an effective motor efficiency of more than 70%. Figures 5 and 6 show photographs of a cooler with a 10 mm coldfinger and a flexure-bearing subassembly.

Figure 5. Photograph of the final flexure-bearing cooler (left).

Figure 6. Photograph of a flexure-bearing assembly (right).

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Close to the end of 1999 several coolers were built and tested. In the beginning of 2000 two of these coolers have successfully undergone a complete qualification test program including

shock and vibration tests. Furthermore, we have put four coolers in lifetime test. One of these four coolers has also been tested in a centrifuge in which we have applied radial accelerations on the compressor for 500 hours. We have increased the radial accelerations starting with 2g up to a maximum of 10g. These tests have not resulted in degradation of the cooler. In March 2000, three coolers with 5mm Stirling coldfingers were put in lifetime test. NEW SIGNAAL USFA FLEXURE-BEARING CRYOCOOLER FAMILY Based on the final design of the flexure-bearing compressor with moving-magnet linear

motor, a complete new range of Stirling cryocoolers referred to as the LSF cryocooler family has been developed. This family consists of four main types of slip-on cryocoolers with 5mm, 7mm, 10mm and 13mm Stirling coldfingers. Qualification tests performed on these four coolers have resulted in technical specifications that are currently available at Signaal Usfa. In Fig. 7, four graphs with specified performances of these coolers are depicted.

Figure 7. Specified cooling performance at different ambient temperature levels

for all four coolers of the LSF cryocooler family.

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FUTURE DEVELOPMENTS

Besides the development of the new flexure-bearing cryocooler family, Signaal Usfa has several other development programs running. Two of these new developments are the development of a miniature pulse tube cryocooler driven by a flexure-bearing compressor, and development of a high capacity Stirling cryocooler with a flexure-bearing compressor.

Flexure-bearing compressor with pulse tube coldfinger

During the past years, Signaal Usfa has been working on an analytic model of pulse tube refrigerators based on harmonic approximations. This tool has made it possible to perform optimization of different design parameters determining pulse tube efficiency and helped us in understanding pulse tube behavior. During the past months we have been performing tests on a prototype pulse tube coldfinger which has a tube diameter of 5mm. The pulse tube design is a U-shape configuration with inertance tube and single bypass. Up to this moment we have measured a cooling performance of 500 mW at 80K with this pulse tube at room temperature conditions. To drive the pulse tube, a flexure-bearing compressor was adapted to match the volume and drive frequency of the pulse tube. Our goal is to increase the cooling performance up to 1 W at 80K and further industrialize the combination of pulse tube and flexure-bearing compressor. High capacity Stirling cryocooler with flexure bearings During the past year we have also been working on the design of a high capacity Stirling cooler. The initial goal with this cooler was to obtain a cooling power of more than 6 W at 80K with a lifetime of more than 20,000 hours MTTF. To achieve this we have designed a cooler based on a Stirling coldfinger with a diameter of 20 mm. During the beginning of this year we have build a first prototype of this cooler and we are

currently running an extensive test program on this cooler. With this first prototype we have measured a cooling performances of more than 8 W at 80K in room temperature conditions with

an electrical input power of only 150 Wac to the cooler. We are currently optimizing different parameters of the coldfinger that determine the efficiency. In the meantime we are designing a flexure-bearing compressor with moving magnet technology to drive the coldfinger in a final design that will be tested before the end of this year.

The diameter of the flexure-bearing compressor will be less than 90 mm and the length of the compressor will be less than 200 mm, resulting in a relatively compact cooler considering the high cooling performance. CONCLUSIONS From the presented work, it may be concluded that:

•

After three years of work we have been able to develop a range of affordable and compact flexure-bearing cryocoolers. These coolers are currently available for prices that are of the same order of magnitude as conventional tactical coolers currently available in the market. • The design of the flexure-bearing compressor, which has a diameter of only 60 mm and a length of 165 mm, is such that it can be matched to different Stirling coldfingers and pulse tubes with comparable void volumes. • Inside the compressor design, full support of the pistons via flexure bearings ensures the absence of piston wear. The application of moving-magnet technology has resulted in the coils being outside of the working gas, thus reducing the risk of gas contamination and eliminating

critical components such as flying leads and glass feedthroughs. •

Many qualification tests performed on these coolers have shown that the compressor design

can withstand the severe environmental conditions required for tactical applications.

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Lifetime tests that are running at Signaal Usfa indicate that the expected lifetime of more than 20,000 hours will be achieved. It’s just a matter of time to verify lifetime results in practice. • New developments currently running at Signaal Usfa offer opportunities for future markets requiring ultra-low vibration levels or cooling powers of more than 6 W at 80K combined with long lifetimes exceeding 20,000 hours MTTF. With our first prototype high capacity cooler, a cooling performance of more than 8 W at 80K with only 150 Wac input power has been measured.

Development of a Long-Life Stirling Pulse Tube Cryocooler for a Superconducting Filter Subsystem Y. Hiratsuka1, K. Murayama2, Y. Maeda2, F. Imai2, K. Y. Kang2 and Y. Matsubara3 1 DAIKIN Industries, Ltd. Semiconductor Equipment Department Osaka 592-8331, Japan 2 DAIKIN Environmental Laboratory, Ltd. Tsukuba, Ibaraki 305-0841, Japan 3 Atomic Energy Research Institute, Ninon University Funabashi, Chiba 274-0063, Japan

ABSTRACT

We have developed pulse tube cryocoolers for high temperature superconducting (HTS) filter subsystems used in the base stations of mobile telecommunication systems. In July 1999, we reported on our development of a 5 W Stirling pulse tube cryocooler with a contact-type compressor,1 with a cooling capacity of 5.5 W at 80 K for 200 W of input power. However, demands for a smaller-sized cryocooler with higher efficiency and with 5-year reliability prompted us to develop such a cryocooler with a U-type expander and a flexure-bearing-supported linear compressor with opposed pistons. We have developed an HTS filter and a long-life Stirling pulse tube cryocooler to cool the filter whose cooling capacity is around 1W at 80 K, as previously discussed in a progress report.2 For a compressor input power of 60 W at an operating frequency of 52 Hz and a pressure-volume (P-V) work of 26 W, and for a compressor efficiency of 45%, this cryocooler achieved a cooling capacity of 1.05 W at 80 K (0.63 W at 70 K), a specific power of 92 W/W, 5.5% Carnot (3.9% Carnot at 70 K), and a specific P-V work of 40 W/W, with a minimum temperature of 57 K in an ambient of 23°C. The key devices of this filter subsystem are an HTS filter and a low noise amplifier (LNA). The HTS filter is made from a YBCO HTS thin film and has a fractional bandwidth below 1.2% at 2 GHz and has a minimum insertion loss at 0.3dB. The HTS filter and the LNA are operated at a constant temperature of 70 K and the cooling capacity needed by them is 0.6 W. We integrated them with the cryocooler into a subsystem, and the external dimensions of this system are 194 mm high, 180 mm wide, 250 mm deep, with a total volume of 8.7 L.

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INTRODUCTION The development of high temperature superconducting (HTS) devices has increased the demand for small-sized cryocoolers. In 1992, E. Tward et al.3 developed a Stirling-type pulse tube cryocooler with a cooling capacity of 1 W at 80 K, similar to the power required for a conventional Stirling cryocooler. In 1998, J.L. Martin et al.4 designed a low-cost prototype Stirling-type pulse tube cryocooler for civilian electronics applications, especially for cooling

superconducting filter subsystems for base stations of mobile telecommunication systems. Furthermore, in 1999, S-Y. Kim et al.5 addressed HTS cryocoolers with a target cost of $1,000 for quantities of 10,000 cryocooler systems per year. They designed a cryocooler to demonstrate the feasibility of making a long-life, low-noise cryocooler for cooling HTS devices. In 1999, we reported on the development of a 5 W at 80 K class pulse tube cryocooler for

HTS filter subsystems. Miniaturization of subsystems along with the demand for reliable cryocoolers prompted us to further develop a non-contact type compressor whose cooling capacity is 1 W at 80 K. We next developed an HTS filter and then combined it with the pulse tube cryocooler into an integrated subsystem. This paper describes the development of our pulse tube cryocooler with a target cooling capacity of 1 W at 80 K (0.6 W at 70 K), and its associated HTS filter subsystem. GENERAL DESIGN The specifications for the cooler are shown in Table 1, while Figs. 1 and 2 show a schematic and a photograph of the unit. The cryocooler is a split Stirling-type pulse tube with opposed pistons that are driven by a linear motor. The compressor is connected to the expander

by tubes that are between 20 mm and 100 mm long. The cooling method is forced air-cooling. The system is filled with helium gas up to 3.1 MPa. During thermal testing the unit is mounted

in a vacuum chamber evacuated to about 0.13 MPa (

Figure 1. Schematic of the pulse tube cryocooler.

torr).

Figure 2. Photograph of cryocooler.

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Expander The cooling capacity of the U-type expander is 10% inferior to that of an in-line type. However, we adopted the U-type expander in which the regenerator and pulse tube are arranged in parallel because it decreases the size. The phase control system uses an inertance tube to prevent instability of the DC flow.

Compressor To develop a cryocooler that has a long-life and requires no maintenance for at least five years, we developed a non-contact type compressor in which the piston is arranged with bearings to decrease the size of the cryocooler, and with 'tangential flexures' to significantly im-

prove the radial stiffness. The clearance of the piston and cylinder can thus be kept to less than 15 µm and the piston has no seal. This non-contact type compressor has an outer diameter of 76 mm and a length of 169 mm. We compared this non-contact type compressor with a contact-type compressor developed for Stirling cryocoolers. In the comparison tests, the piston position was monitored by using a laser vibrometer, and the mass flow rate through the phase shifter was measured using pressure transducers mounted near the compressor discharge head, near the hot side of the regenerator, and near the pulse tube. These measurements were used to calculate both the pressure-volume (P-V) work of the compressor and the equivalent P-V work of the expander.

RESULTS AND DISCUSSION Cryocooler Performance The diameter and length of the regenerator, pulse tube, and inertance tubes were optimized to achieve the maximum cooling capacity. A comparison of the cooling capacity of the

non-contact and contact-type compressors is shown in Figure 3. For the contact-type compressor, for 60 W input power, the no-load temperature was 56 K, the cooling capacity was 0.65 W

at 70 K (1.1 W at 80K), and the P-V work was about 27 W. In contrast, for the non-contact type compressor, the no-load temperature was 57 K, the cooling capacity was 0.63 W at 70 K (1.05 W at 80 K), which is slightly lower than that for the contact type compressor, and the P-V work was about 26 W with a 23°C ambient. The efficiency of the non-contact type compressor was 45%; however, if a higher efficiency compressor is used, the cooling capacity at a given temperature should increase. Figure 4 shows the efficiency and cooling capacity at 70 K as a function of compressor input power. The efficiency is 3.9% Carnot at 70 K for 60 W of input power and the cooling capacity is relatively independent of input power. Figure 5 shows the cooling capacity and

Figure 3. Measured cryocooler cooling capacity vs. cold head temperature.

Figure 4. Measured cryocooler cooling capacity vs. compressor power.

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Figure 5. Measured cryocooler cooling capacity vs. cold head temperature.

Figure 6. Measured cryocooler cooling capacity vs. inclination angle.

%Carnot efficiency as a function of cold head temperature. When the cold head temperature is higher than the pulse tube cryocooler shows superior and more linear temperature characteristics compared with those of the Stirling cryocooler.

Influence of Pulse Tube Inclination The performance of pulse tube cryocoolers can be affected by natural convection, which is affected by the inclination of the pulse tube. In July 1999, we reported that cryocooler performance was unaffected by the pulse tube inclination. In general, high operating frequency and

small diameter pulse tubes reduce the convective heat loss. For confirmation, we measured the effect of pulse tube inclination on the performance of the cryocooler. Figure 6 shows the cooling capacity at 70 K and 80 K for cold head orientations of 0° to 180° downward from vertical. The data show that the cooling capacity is significantly reduced when the inclination is 135°, which differs from previous results in which the cooling capacity was independent of inclination.2 This difference might be due to the difference in shape between the in-line type and U-type expanders.

Influence of Environmental Temperature It was necessary to confirm the effect of environment temperature on cryocooler performance because the device is used outside. To operate the HTS filter for this system, the cooling capacity necessary is 0.6 W at 70 K. We tested the system by first setting the environment temperature to 60°C, 40°C, 25°C and 23°C, respectively, via a thermostat, and then changing the operating frequency and compressor input power to optimize the cooling capacity. Our results show that the cooling capacity at an environment temperature of 60°C was significantly lower than that at an ambient of 23°C (Figure 7). However, this system did not reach the target value, and at about 40°C the system becomes stroke limited. The effect of operating frequency on the cooling capacity at environment temperatures of 25°C and 60°C is shown in Figure 8. The optimal operating frequency is seen to depend on the

environment temperature. This is because the temperature and internal pressure of the cryocooler gas increase with increases in the environment temperature, and the increased pressure causes the optimum operating conditions for the compressor and inertance tube to change.

Vibration Pulse tube cryocoolers, which have no moving parts in the cold section, are more attractive than other small-sized cryocoolers because of their high reliability, simpler construction, and lower vibration levels. Vibration is not a significant problem in the filter subsystem, but,

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Figure 7. Measured cryocooler cooling capacity vs. environmental temperature.

Figure 9. Vibration measurement method.

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Figure 8. Measured cryocooler cooling capacity vs. operating frequency.

Figure 10. Measured cryocooler vibration.

for example, there are applications like an electron microscope that require almost no vibration. We therefore used a laser vibrometer to measure the vibration of a 70 K cold head (Fig. 9). To extract only the vibration at the cold head, we simultaneously measured the vibration at the flange of the expander, and then subtracted it from the cold head vibration measurement. Figure 10 shows the pressure of the pulse tube hot end (solid line) and vibration at the cold head (dotted line). The 0-peak vibration amplitude is around +0.6 µm , with a total peak-peak amplitude of 1.2 µm, and has a period similar to that of the pressure amplitude. We calculated the axial extension of the expander cylinder from this pressure amplitude, and found that the value calculated was approximately +0.94 mm, which is similar to the measured value. For comparison, we also measured the vibration of a Stirling cryocooler, and found that it was similar to that of a pulse tube cryocooler.

Filter subsystem

The key devices of this filter subsystem are an HTS filter and an LNA. Figure 11 shows the frequency response of an HTS filter. This filter is made by using a YBCO HTS thin film, and has a fractional bandwidth below 1.2% at 2 GHz and has a minimum insertion loss at 0.3 dB. Figure 12 shows a photograph of the subsystem. The external dimensions of this system are 194 mm high, 180 mm wide, 250 mm deep, with a total volume of 8.7 L. The cooling capacity needed by the HTS filter and LNA is 0.6 W at 70 K. The caloric output for the amplifier is 0.4 W, the quantity of radiation parasitics is 0.1W, and the parasitic heat conduction down the cable is 0.1W. The vacuum container theoretically lasts for 10 years.

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Figure 11. Frequency response of HTS filter.

Figure 12. Photograph of subsystem.

CONCLUSIONS We developed an HTS filter and a long-life Stirling pulse tube cryocooler to cool the filter. We then integrated them into a subsystem. The significant results of our research are as follows: 1. For an input power of 60 W, the no-load temperature of the cryocooler was 57 K and the cooling capacity was 0.63 W at 70 K, and 1.05 W at 80 K. 2. Vibration in the cold head of this pulse tube cryocooler is primarily caused by the cold-head elongation due to the cycling pressure. The vibration amplitude is about 1.2 mm peak to peak. 3. Environmental temperature dependence of this cryocooler is 15 mW/°C. 4. The cooling capacity decreases by 35% when the pulse tube is inclined 135° from the vertical. 5. The HTS filter has a fractional bandwidth below 1.2% at 2 GHz and has a minimum insertion loss of 0.3 dB. 6. The integrated subsystem is small and light weight, having a total volume of only 8.7 L.

REFERENCES 1. Y. Hiratsuka et al., “Development of a 1 to 5 W at 80 K Stirling Pulse Tube Cryocooler,” Cryocoolers 10, Plenum Press, New York (1999), pp. 149-155.

2. Y. Hiratsuka et al., “Development of a 5 W at 80 K Stirling Pulse Tube Cryocooler,” Advances in Cryogenic Engin., Plenum Press, New York (2000).

3. E. Tward et al.,“Miniature Pulse Tube Cooler,” 7th International Cryocooler Conference Proceedings, Air Force Phillips Laboratory Report PL-CP--93-1001, Kirtland Air Force Base, NM, April 1993, p. 113. 4. J.L. Martin et al.,“Design Consideration for Industrial Cryocoolers,” Cryocoolers 10, Plenum Press,

New York (1999), pp. 181-189.

5. S-Y Kim et al.,“Development of Low-Cost Pulse Tube Cryocooler for HTS Applications,” Advances in Cryogenic Engin., Plenum Press, New York (2000).

Development of a 5W at 65 K Air-Cooled Pulse Tube Cryocooler S-Y Kim, J-J Park, S-T Kim, W-S Chung, H-K Lee

LG Electronics Inc. Digital Appliance Lab. COMP Team Seoul 153-023, Korea

ABSTRACT

LG Electronics (LGE) has developed an air-cooled Pulse Tube Cryocooler (PTC) for HTS applications. The air-cooled PTC provides 5.5 W of cooling at 65 K and 25°C ambient with 280 W of input power with a single-acting linear compressor. Its performance is greater than that of an LGE water-cooled PTC of last year, 4.9 W at 65 K and 20°C, with 270 W of input power. The aircooled PTC, compared to the water-cooled one, has a lower-efficiency linear motor, a 3% decrease due to cutting down the cost of the linear motor, and a higher temperature at the surface of the

aftercooler, 30°C higher than the water-cooled one. Thus, the performance has improved significantly. We have optimized the geometry of the regenerator, the pulse tube, and the inertance tube using both simulation and test; a significant increase of the performance is noted. Also, to maintain the low cost target, we have decreased the cost of the linear motor and used a common heat exchanger for air-cooling. INTRODUCTION A pulse tube Cryocooler has a lot of advantages over a Stirling Cryocooler because it has no moving parts in the low-temperature region. This means much more reliable operation and much lower vibration in the cold region. Ideally, Stirling cryocoolers have better thermal efficiency than PTC. However, some recent studies2,3 have suggested that PTC using an ‘inertance’ tube can

generate the phase shift needed to make the PTC operate as well as a Stirling cryocooler. The inertance tube is a long, thin tube that allows the phase between the pressure and mass flow in the pulse tube to be adjusted to an extent that was not previously possible. The air-cooled pulse tube cryocooler developed by LGE aims at HTS applications such as cooling HTS RF filters in wireless communication systems. The compressor that drives the in-line pulse tube is based on a moving-magnet design combined with flexure supports. This PTC design approach strives for a better price-to-performance ratio, while maintaining a long life. Our target cost is $1,000 for quantities of 10,000 units per year. The PTC has a high potential for very low cost and demonstrates high cooling power. The design of the PTC was driven by the challenging low cost and reliability targets associated with the requirements of wireless applications. The M-CALC II report3 suggested a goal of life and cost for low-cost commercial cryocoolers. The 40,000-hour continuous duty life limits the design options available, and the cost goal of $1,000 for quantities of 10,000 a year for the complete Cryocoolers 11, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2001

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cryocooler means that this cooler must achieve goals that no other cryocooler has yet met. We continue to cut down the cost of the PTC such as by reducing the size of the linear motor magnet

and decreasing the high-precision machining. This is because the linear motor is the key component and the compressor cost reported5,6 is over 50% of the total system cost. In the air-cooled PTC, we are also using a radial copper fin type of heat exchanger and a common fan at the aftercooler to achieve a lower level of cost. We have developed a performance prediction program based on the ‘thermodynamic nonsymmetry’ effect1 and tried to find the optimum design point for the main components such as the regenerator, pulse tube and inertance tube. The performance must be improved significantly because the efficiency of the linear motor was decreased, 3% lower due to the low cost design, and the heat rejection temperature was increased, 30°C higher than the water-cooled one.7 Currently, the performance tests show very good results, 60% higher than what the water-cooled one achieved last year. COMPRESSOR The low-cost compressor consists of three main parts: a clearance-sealed piston/cylinder, spiral flexure linear bearings, and a moving magnet type of linear motor. These are enclosed in a common pressure vessel with an aftercooler interface to the pulse tube. The layout of the compressor in combination with the in-line pulse tube is shown in Figure 1. This unit has a motor diameter of 130 mm and a total prototype vessel dimension of 170 mm diameter, including vessel flange,

and 245 mm long, including aftercooler. The compressor mass is 10 kg, and the total mass is 12 kg. Production vessels will be welded shut, not bolted. This will save some mass and remove the flanges. The moving magnet motor consists of an outer stator with coil, an inner stator, and a magnet assembly on which the magnets are attached. The magnet assembly is connected to a reciprocating shaft supported on each end by flexures and connected to the piston. Moving magnet motors remove the need for problem-prone flexing leads, so they simplify the structure greatly. The motor is designed for a maximum of 400 W of mechanical output, and a maximum of 20 mm of stroke. It is

tuned by adjusting the moving mass and springs so that it works at 60 Hz from 220 VAC. The linear motor, which includes radial-laminations stacked automatically, can be produced as easily as a traditional motor in the mass production line, and uses a very small Nd-Fe-B magnet. The piston moves without rubbing contact in the cylinder to produce the pressure wave that drives the pulse tube. The clearance seal is maintained in a centered radial position by flexures, and the radial clearance is 25~35 mm. The clearance seal requires very tight tolerances and high-precision alignment. Because the clearance seals may be too expensive for mass production, one of the

Figure 1. Layout of the air-cooled pulse tube cryocooler.

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primary concerns is to considerably reduce the cost of their alignment. This is accomplished by reducing the high-precision machining, devising a simplified mechanical architecture with fewer parts, and simplifying the alignment of the piston supports. The front flexures, combined with the piston and the shaft, are bolted to the cylinder assembly maintaining the piston in the center of the cylinder; the backside flexures, bolted to a backside frame, are combined with the shaft and the magnet assembly in a way that maintains the pre-alignment. This robust and inexpensive design shows only a small amount of loss of delivered PV power. OPTIMIZATION

The LGE pulse tube has an in-line shape with an inertance tube to provide the proper phase shift. Stacked copper screens are used in the cold end, the warm end, and in the aftercooler; stainless steel 400 mesh screens are used in the regenerator. The PTC uses a 1 inch inner diameter thin

stainless steel tube for the regenerator, and a 1/2 inch tube for the pulse tube. The charge pressure is 2.7 MPa and the amplitude is 0.5 MPa. The pulse tube consists of many components that have complex thermal interactions when a part is changed, so it is very hard to optimize each component. Therefore, we decided to first increase the inner aftercooler volume and the outer surface of the aircooled PTC to provide a higher overall heat transfer coefficient. Then, the regenerator, pulse tube and inertance tube, which were expected to be the main parameters with the largest effect on the performance, were optimized using simulations and tests. Figure 2 shows the predicted performance from the simulations as a function of regenerator volume and pulse tube volume. The swept volume and the aftercooler volume are fixed and the

inertance tube is tuned to achieve maximum COP in every case. From the figure, it is clear that there is a strong trade-off between the regenerator loss and the pressure drop inside the regenerator. The optimum design point of the regenerator is easily found, although the optimum pulse tube volume varies slightly when the regenerator volume is changed. In the small regenerator volume region, in which the regenerator loss is dominant, we can see the performance rises steeply when the regenerator volume is increased. However, in the pressure drop dominant region, the curves show a gentle slope. Figure 3 shows a comparison between the simulation and test results. The test results have lower performance than the simulations due to some assumptions in the simulation—adiabatic compressor, perfect heat exchanger and so on—and some heat losses. However, the trends are very similar. The optimum point of the regenerator and the pulse tube volume is not changed with the various inertance tubes, but the simulation and test results show the performance of the PTC is changed

Figure 2. COP Predictions as a function of regenerator and pulse tube volume: Vre regenerator volume, Vpt pulse tube volume, Vs swept volume; the inertance tube is tuned in every each case.

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Figure 3. COP as a function of regenerator and pulse tube volume (25°C ambient temperature, vacuum torr): Vre regenerator volume, Vpt pulse tube volume, Vs swept volume; the inertance tube is tuned in each case. significantly when the length to diameter ratio of the inertance tube, L/D, is increased (Figure 4). In each case, the length and diameter of the inertance tube are tuned to maximize the COP of the PTC. Thus, the length to diameter ratio is increased when the diameter is increased. The test results show a different tendency beyond the particular L/D of the inertance tube, in which the performance goes down steeply. Therefore, we can conclude that the length to diameter parameter of the inertance tube has a significant effect on the performance and there is a critical point, which means the phenomena inside the inertance tube seems to be changed suddenly beyond that point. Figure 5 shows the performance test results of the optimized PTC in the torr vacuum chamber at 25 °C ambient temperature. The curve shows the 5.5 W at 65 K, and the 48 K no-heatload temperature. The cooling capacity of the optimized cooler is slightly improved over what the water-cooled one achieved last year; for the same conditions, the optimized one shows 60% higher cooling capacity. The total input power is 280 W, and the PV work is 240 W, 86% of indicated efficiency and 87% of estimated motor efficiency. SUMMARY AND CONCLUSIONS

The air-cooled optimized PTC developed by LGE has a better price-to-performance ratio than that of our water-cooled PTC of last year, even with the lower efficiency linear motor and the

Figure 4. COP as a function of inertance tube L/D ratio (25°C ambient temp., vacuum

torr).

DEVELOPMENT OF A 5W AT 65 K AIR-COOLED PT COOLER

Figure 5. Cooling capacity at the cold end (25°C ambient temp., vacuum

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torr).

higher heat rejection temperature of the aftercooler. These changes were the result of cutting down the cost of the linear motor and air-cooling. In addition, these results indicate that the PTC will be able to meet the challenging cost goals and the performance suggested by the HTS application companies.

REFERENCES 1. J. Liang, A. Ravex and P. Rolland, “Study on pulse tube refrigeration”, Cryogenics, Vol. 36 (1996), pp. 87-106. 2. D.L. Gardner, and G.W. Swift, “Use of Inertance in Orifice Pulse Tube Refrigerator,” Cryogenics, Vol. 37 (1997), pp. 117-121. 3. Pat R. Roach and Ali Kashani, “Pulse Tube Coolers with an Inertance Tube: Theory, Modeling and Practice,” Advances in Cryogenic Engineering, Vol. 43, Plenum Press, New York (1998), pp. 18951902. 4. M. Nisenoff, Cryocoolers for Electronic Technologies, M-CALC II Workshop Report, San Diego, 1998.

5. T. Nast, P. Champagne, and V. Kotsubo, “Development of a Low-Cost Unlimited-Life Pulse-Tube Cryocooler for Commercial Applications,” Advances in Cryogenic Engineering, Vol. 43, Plenum Press, New York (1998), pp. 2047-2053. 6. J.L. Martin, J.A. Corey, and C.M. Martin, “A Pulse Tube Cryocooler for Telecommunications Applications,” Cryocoolers 10, Plenum Press, New York (1999), pp. 181-189. 7. S-Y Kim et al., “Development of low-cost Pulse Tube Cryocooler for HTS Applications,” Advances in Cryogenic Engineering, Vol. 45, Plenum Press, New York (2000), pp. 19-24.

TES FPC Flight Pulse Tube Cooler System J. Raab, S. Abedzadeh, R. Colbert, J. Godden, D. Harvey, C. Jaco TRW One Space Park Redondo Beach, CA 90278 USA

ABSTRACT

The TRW Tropospheric Emission Spectrometer (TES) Focal Plane Cooler (FPC) features two integral pulse tube cryocoolers that independently control the temperature of the two instrument focal planes. The TES mission acquires high-resolution ozone concentration data in the earth’s troposphere in order to better understand the ozone: where it comes from and its interaction with other chemicals in the atmosphere. TES is scheduled to fly on the EOS-Aura platform in 2002. The TES FPC program delivered two flight coolers and electronics, and one flight spare cooler and electronics in November 1999. This paper presents data collected on the flight coolers during acceptance testing. Tests included thermal performance mapping at various reject temperatures and power levels, launch vibration testing, EMC/EMI testing, and self-induced vibration testing. Designed conservatively for a six-year life, the coolers are required to provide 1W cooling at 57K while rejecting to 35°C with less than 63W input power to the electronics. The system (cooler and electronics) required mass is less than 17.1 kg. The system also includes radiation-hardened control electronics and provides cooler control functions with a software-controlled microprocessor. INTRODUCTION The TES FPC program delivered two flight cooler systems and one flight spare cooler system, plus ground support electronics (GSE) to interface with the cooler system. The program was performed for the Jet Propulsion Laboratory (JPL) over a 31-month period. The TES mechanical cooler is shown in Figures 1 and 2. The cooler provides focal plane array (FPA) cooling via a thermal strap and rejects heat to a loop heat pipe attached to a radiator. The TES FPC system, which is the latest version of the TRW 100 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 derived from the AIRS cooler which was a split pulse tube cooler. The TES cooler has been reconfigured into an integral configuration with the same cold head and the same compressor as the AIRS cooler. The electronics is the same basic design as the AIRS and MTI (currently in orbit) electronics except that the producibility was upgraded and the software was made more user friendly. 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, EMI/EMC testing, Cryocoolers 11, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2001

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Figure 1. Integral vibrationally-balanced pulse tube cooler.

Figure 2. TES FPC envelop for the mechanical cooler.

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and burn-in. These tests, which are typical for space instruments, are performed to ensure reliability. The cooler performance, including load lines, temperature stability, self-induced vibrational force, and EMI/EMC properties, was measured. This paper reports the test data for one of the new flight coolers. There was less than 6% input power difference from unit to unit at the nominal operating condition of 1W at 57 K. 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 the resonant frequency of 44.6 Hz. Capacitive sensors are used to measure the position of both pistons. The output is used to measure and control dc offset and to provide overstroke protection. 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 and incorporates the reservoir tank. 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 and bypass line connect the gas from the orifice block to the reservoir tank and to the aftercooler, respectively. The internal wiring in the compressor is stranded, PTFE-insulated (cross-linked Teflon) wiring, or Kapton flexible cable. All wiring exits the centerplate through ceramic-insulated pins in feedthroughs attached to D-shell connectors for the cooler drive power and to the capacitive sensors and thermistor. A separate connector is used for the redundant platinum resistance thermom-

eters (PRTs) on the cold block. Redundant accelerometers are mounted on the compressor centerplate. Together with the signal conditioning electronics, the accelerometer provides a feedback signal to the vibration control algorithm in the cooler control electronics (CCE).

The CCE (Figure 3) is based on our high-reliability AIRS flight design1 modified forproducibility. New features include the horizontal slice design as shown in Figure 3 and additional internal connectors to allow for slice-level testing. There are three slice subassemblies, one for control (control

slice), one for power amplifiers (power slice), and one for power conversion (converter slice). The

Figure 3. Cooler Control electronics (CCE).

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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 28 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 AC drive power at 44.6 Hz to the motors in the compressor. The compressor moving coil and piston assemblies are designed to resonate on their gas and mechanical springs at this drive frequency, and thus

generate a 44.6 Hz pressure wave and mass flow to the cold head. The software adjusts the stroke 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. Figures 4 and 5 show the cooling load as a function of cooling temperature for different reject temperatures and input powers. For the TES FPC nominal cooling load of 1.0 W at 57 K, the cooler system requires 58.7 W of input power and the compressor operates at 45.5% stroke. For a TES FPC cooling load of 0.5 W at 57 K, the cooler system requires 34.5 W of input power and the compressor operates at 35.8% stroke. The CCE (Figure 3) plays a critical role in the overall cooler performance. When the input bus power was measured as a function of the output power to the compressor line correlation: where the efficiency

it fit the straight-

is 0.825, and the extrapolated tare power at zero compressor power is

TES FPC FLIGHT PULSE TUBE COOLER SYSTEM

Figure 4. Cryocooler performance for variable reject temperatures: 1W at 57K.

Figure 5. Cryocooler performance for variable reject temperatures: 0.5 W at 62K.

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Figure 6. Temperature stability maintained by control loop during simulated orbital temperature change.

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 6 shows that the temperature stability of the cooler is within a 22 mK band when operating with a 1-W load at 57 K and a baseplate temperature change of 0.21°C/min. The resolution of one bit in the temperature measurement electronics is 12 mK.

The vibration control algorithm continuously updates the compressor waveform to minimize cooler vibration. TRW’s special purpose dynamometer measures the three axes of the self-induced

vibration of the cooler. Figure 7 shows the force in the direction of piston motion (cooler axis) as well as the two cross axes when the cooler is mounted on a rigid structure. ENVIRONMENTAL TESTS AND COOLER ACCEPTANCE

The TES FPC acceptance testing included launch random vibration, a thermal vacuum test 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 two orders of magnitude less than the 5-year-life criterion and satisfied a 10-year-life requirement.

EMI/EMC TEST The cooler system must meet stringent requirements for radiated electric and magnetic fields, conducted emissions on the input bus power lines, and electromagnetic susceptibility. Excessive magnetic fields are a generic issue with linear-motor cryocoolers, as are excessive levels of input ripple current. The TES FPC is an integral version of the TRW AIRS design, which required magnetic shielding to pass the radiated magnetic emission requirement.2 The TES FPC will be fitted with this shield design. TRW’s newer cooler designs meet the radiated magnetic emission requirements without the need for shielding.3 The in-rush currents and the ripple currents for the nominal 29 V bus voltage requirement were

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Figure 7. Measured self-induced force of compressor in three axes at 46 compressor stroke level and 42.6W into compressor.

recorded as 4.2 amps and 123 dB micro amps, respectively (Table 1). TRW’s latest electronics have modified the TES FPC design to considerably lower the ripple current.4 The EMI and compatibility EMC qualification tests were performed at the TRW EMI test facilities to determine the degree of compliance to Mil-Std 461C requirements, as modified in TRW BDA-14A-001, EMC Test Procedure for the TES FPC program. Table 2 summarizes the test matrix. During the test series, two tests failed to meet the requirement. For the conducted emissions (CE03), an overlimit condition was observed at 200 and 100 kHz in the narrowband mode (Table 2). An external filter at the input of the CCE will enable the cooler system to meet CE03.

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CONCLUSIONS

The TES FPC performance met the program goals. The coolers were delivered in October 1999 and are awaiting integration with the payload. ACKNOWLEDGMENT

The work described in this paper was carried out by TRW and sponsored by the JPL TES Project. This report was prepared for the Jet Propulsion Laboratory, California Institute of Technology, sponsored by the National Aeronautics and Space Administration. REFERENCES

1. Chan, C.K., J. Raab, A. Eskovitz, A.R. Carden III, R. Orsini, “AIRS Flight Qualified 55K Pulse Tube Cooler,” Cryocooler 9, Plenum Press, NY (1997), pp. 895-904. 2. Johnson, D.L., S.A. Collins, and R.G. Ross, Jr., “EMI Performance of the AIRS Cooler and Electronics,” Cryocooler 10, Kluwer Academic/Plenum Publishers (1999), pp. 771-775. 3. Chan, C.K., T. Nguyen, R. Colbert, J. Raab, R.G. Ross, Jr., and D.L. Johnson, “IMAS Pulse Tube Cooler Development and Testing,” Cryocooler 10, Kluwer Academic/Plenum Publishers (1999), pp. 139-147. 4. Chan, C.K., Pamela Clancy, and John Godden, “Pulse Tube Cooler for Flight Hyperspectral Imaging,” Cryogenics, 39, Elsevier Ltd. (1999), pp. 1007-1014.

The AIM Space Cryocooler Program I. Rühlich, H. Korf and Th. Wiedmann AEG Infrarot-Module GmbH Theresienstr. 2 74072 Heilbronn, Germany

ABSTRACT

AIM is developing space cryocoolers for superconducting telecommunication components. The equipment should be available in 2001, and a mission to the International Space Station is scheduled for 2003. The basis for the cooler development is the AIM model SL200 cooler; it has a nominal cooling capacity of 3.5 W at 77 K, and is used for cooling high performance IR detectors. The space mission will have a duration of three years with two years in operation. The development was structured into the following phases: spin-off of classic design, implementation of flexure bearings, introduction of pulse tube cold head. Results are improved COP and lifetime. The current cooling capacity is about 4.3 W at 77 K with 96 W of input power. The expected lifetime exceeds 30,000 h. The qualification testing will start in July 2000. INTRODUCTION

The cooler AIM SL200, which is currently mainly being used for military IR equipment and for HTSC components, has to be improved to serve as a cooler for space missions in the 5 W-class. The design and performance of the SL200 are an ideal basis to be improved for enhancement of performance and reliability for mid term space missions (SL400). The cooler development at AIM is embedded in a government program “Superconductor and novel ceramics for communication technology of the future.” 1 A major milestone in the program is a mission to the International Space Station (ISS) in 2003 for testing HTSC filters in a cryogenic platform. The program is coordinated by BOSCH Telecom. To provide redundancy, two coolers are planned for the platform. DESIGN DESCRIPTION The AIM SL400 has a classic linear-cooler design with a double-acting compressor and a proven 12 mm cold head. The design has demonstrated sufficient life performance to serve for medium term space missions, even without flexure bearings. The SL200 is a design equivalent to the SL100, which consistently exceeds 4500 h MTTF in accelerated life testing. Rather than its reduced cooling performance, the limiting factor preventing higher MTTF during life qualification is the increased vibration output at end of life. Design features of all SL coolers at AIM are equivalent, providing a profound basis for a thorough redesign for the space demands.

Cryocoolers 11, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2001

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In order to improve the performance, the following components of the cooler have been redesigned:

• Replacement of Samarium Cobalt magnet material with high density Neodymium Iron Boron magnets • Installation of magnetic field liner with higher magnetic field saturation and also smaller hysteresis losses •

Optimization of the resonance conditions and of the regenerator in the cold head

The increase of life time was accomplished by: • Special coating of piston in compressor and of displacer providing practically friction free and wear tree operation of the clearance seal • Current leads to moving coil like professional loudspeaker design •

Compressor spring The current design has the potential to exceed 10,000 h MTTF. Verification is in progress

with three Engineering Models (EM). According to step 2 of the program plan, and in parallel with the qualification of the current configuration, a flexure bearing compressor is also in preparation together with a pulse tube cold head (in collaboration with Gießen University). These improvements will push the MTTF to exceed 30,000 h. Outline dimensions of the cooler are given in Fig. 1 (the length of transfer line for the space mission is about 180 mm); photographs of the cooler and electronics are shown in Figs. 2 and 3. The envelope dimensions of the electronics are 180 x 183 x 63 mm. Its weight is about 1.9 kg. The electronics provides an adjustable temperature control algorithm and the ability to switch between the regulated mode and an override fixed-output mode. The latter is used as an emergency mode and to provide a base load in cases where the other cooler is operating in the regulated mode. Furthermore, the input power into the electronics and the current operational mode can be monitored by telemetry signals. The internal PID algorithm works with a real power control algorithm

for the motor current. Figure 4 illustrates the configuration of the cooler in the breadboard platform. Besides the AIM cooler (left hand side), another cooler in the sub-10W class, fabricated by LEYBOLD, is foreseen for the experiment. Each of the coolers is equipped with a single separate electronics. Both

coolers can operate either separately or jointly, one providing the base load while the other operates in the temperature control mode.

Figure 1. Outline dimension of cooler.

THE AIM SPACE CRYOCOOLER PROGRAM

Figure 2. SL400 cryocooler.

Figure 3. Space Electronics, developed by Astrium.

Figure 4. Breadboard platform with cryocoolers (courtesy BOSCH Telecom).

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Figure 5. Cooling capacity of the SL400.

Figure 6. Comparison of SL400 Carnot fraction with other coolers.2,3

PERFORMANCE DATA Typical heat load requirements for HTSC applications are in the range of about 4 W at 77 K. As the total heat load of the cryogenic platform being developed by BOSCH Telecom is also approximately 4 W, the performance of the coolers have to exceed 4 W to meet the cooling requirements. Beside that, high efficiency was a major requirement for the development. Cooler

So far 4.3 W at 77 K with 96 W of input power has been achieved. The cooling capacity vs. input power for different ambient conditions is shown in Fig. 5. The Carnot fraction of the SL400

and other cryocoolers in the range of cooling capacities between 1 W and 10 W is shown in Fig. 6

for comparison.2,3

The specific cooling capacity is another important criterion, especially for space applications. The cooling capacity vs. cooler mass is shown in Fig. 7. Three curves for specific cooling capacity are given.

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Figure 7. Comparison of cooling capacity vs. mass for SL400 and others.2,3

Electronics The characteristics of the electronics are as follows: Control temperature setting 47 K - ambient

Temperature stability Temperature sensor

PT 100

Control algorithm

PID

Operation voltage

Efficiency STATUS OF QUALIFICATION The assembly of Engineering Models for the space qualification is completed. The general qualification at the system level will be performed at BOSCH Telecom. After the final test at AIM the delivery is scheduled for the end of June 2000. Vibration tests of the bare cooler against the levels given in Tables 1 and 2 have been performed successfully. The limitation for the elastic deflection of the cold head during launch vibration has required some special solutions. The key

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issue is the high mass of the thermal interface from the cold tip to the flexible thermal connection to the cryogenic platform; this mass leads to low frequency resonances, and thus to high deflections. BOSCH therefore has developed a special Kevlar support with high stiffness and low axial conduction. The assembly of the cold finger and support has successfully been tested at qualification levels. For verification against shock loads, 10 shocks per axis will be tested. The levels are given in Table 3. CONCLUSIONS AIM is performing a program for the development of space cryocoolers in the 5W class. Engineering models for the space qualification at BOSCH telecom are about to be delivered. The cooler SL400 has a cooling capacity of 4.3 W at 77 K with 96 W of input power, which is a 12.7% Carnot fraction. The qualification will be finished in fall 2000. A mission to ISS is scheduled for 2003.

ACKNOWLEDGMENT Financial support by a BMBF grant (13 N 7391) is gratefully acknowledged. The authors would like to acknowledge the good cooperation with BOSCH Telecom, LEYBOLD VAKUUM and Astrium. NOMENCLATURE EM g MTTF RMS

Engineering Model Constant of gravitation Mean time to failure Root mean square

REFERENCES 1. Schrempp, Ch., Klauda, M., Neumann, Ch., “Design of a Cryogenic Platform for New Communication Payload Technologies,” SAE Paper 1999-01-2086, 29th International Conference on Environmental Systems, Denver, 1999. 2. “Cryocooler Survey 1998,” CD-ROM by Nichols Research Corp., Albuquerque, NM. 3. Glaister, D.S. et al., “An Overview of Performance and Maturity of Long Life Cryocoolers for Space Applications,” Cryocoolers 10, Plenum Press, NY (1999), p. 1.

Miniature Pulse Tube Cryocooler for Space Applications T. C. Nast, P. J. Champagne, V. Kotsubo, J. Olson, A. Collaco and B. Evtimov Lockheed Martin Advanced Technology Center PaloAlto, CA 94304-1191 T. Renna Lockheed Martin Communications and Power Center Newton, PA R. Clappier Clappier Consulting Discovery Bay, CA 94514

ABSTRACT

Lockheed Martin’s Advanced Technology Center (LM ATC) has developed a miniature, lightweight pulse tube cryocooler system for space operation under funding from NASA/GSFC. The cold end is a U-tube configuration, and is driven by a dual opposed piston flexure bearing compressor. The compressor utilizes a moving magnet linear motor and incorporates a number of features that simplify assembly and enhance reliability. This cooler is designed for 0.3 W of cooling at 65 K with a 310 K rejection temperature, 15 W of compressor power, and a mass of

less than 1.25 kg. Three engineering model cryocoolers are to be delivered to NASA/GSFC. Two have been completed and test data is presented here on the EM performance. In a parallel effort, we have developed a lightweight, power efficient electronic controller funded by LM ATC cost sharing funds for a DARPA contract. The first Engineering Model of this controller has been completed and functional testing has verified successful operation. All

of the design goals have been met or exceeded. INTRODUCTION Numerous future spaceflight missions require a very light weight, compact cryocooler system with lifetimes of 10 years or more. Cryocooler customers also want lower costs and shorter delivery times. NASA-GSFC awarded a development contract to Lockheed Martin’s Advanced Technology Center (LM ATC) for the thermo-mechanical system in Sept. 1997. The cooling requirements are 0.3 W at 65 K with 15 W of compressor power with a rejection temperature of 310 K. Qualification and cooling performance is performed from 250 to 310 K. With the requirement for lifetimes in excess of 10 years, LM ATC selected the no-moving-parts pulse tube coldhead driven by a flexure-bearing clearance-seal compressor. At LM Cryocoolers 11, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2001

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ATC we have achieved pulse tube efficiencies comparable to Stirlings,1-4 such that the cooling requirements of the NASA-GSFC contract could be met with a pulse tube. The flexure-bearing compressor has now been demonstrated to be a reliable, robust technology. For this program, LM ATC is using a moving magnet compressor that has simplified assembly, reduced cost and

enhanced reliability over the standard Oxford-heritage compressor with moving coil. The NASA/GSFC contract calls for the delivery of 3 EMs. The first has completed testing, the second is in early stages of testing, and the third is undergoing final assembly. The load lines from EM #1 and early test results from EM #2 are presented here. LM ATC also recognized the need for a smaller, lighter, lower cost electronic controller with substantially improved reliability over previous versions. LM ATC initiated a program early in 1998 to develop a second-generation electronic controller. The circuits and control logic were designed by a consultant and LM ATC, with the development and the manufacture of the EMs and FMs by Lockheed Martin’s Communication and Power Center (LM CPC). An Engineering Model of this controller has recently completed functional testing and all goals of the design have been achieved.

SYSTEM DESCRIPTION Cryocooler Thermo-Mechanical Unit (TMU) The thermo-mechanical unit is an integral configuration with a U-tube coldhead directly mounted to the compressor. This arrangement requires only a single spacecraft-mounting interface for both structural support and heat rejection. Mounting options are available as re-

quired by the customer. 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 cyrocooler are packaged within the configuration shown. No additional hardware is required. 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

Figure 1. Cryocooler and electronic controller.

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Figure 2. Static load test of cold head.

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 vibration 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 the cold end under a static load test supporting 5 kg. 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 feedthrough 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 piecepart count which 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 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.

Cooler Drive Electronics (CDE) The electronic controller being developed has full capability for operating the cryocooler in space environments. It has a PWM amplifier for driving the compressor, feedback control of

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Figure 3. Test setup for thermodynamic testing.

piston amplitudes for temperature control, and feedback control of one of the pistons for reduction of exported axial vibration. An RS-422 port provides interfacing to the spacecraft for command and telemetry. This controller is substantially lighter, more power efficient, more reliable, and consumes less overhead power than LM ATC’s existing Stirling cooler flight electronic controller. The production cost will also be substantially lower than the existing controller. These advances were achieved by simplifying all aspects of the controller, including the control algorithms and the circuitry. Elimination of the displacer drive and control circuitry associated with the Stirling cycle also results in further simplification. SYSTEM PERFORMANCE Thermo-mechanical Unit Figure 3 shows the test setup for thermodynamic testing.

Thermodynamic Performance. Performance tests have been conducted on the EM unit at various power inputs and heat rejection temperatures. Figure 4 presents the load lines at several power inputs for heat rejection temperatures of 280 K for EM #1. Figure 5 presents the effect of heat rejection temperature on cooling loads. The requirement of 0.3 W of cooling at 65 K with a

Figure 4. Load curves for EM #1.

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Figure 5. Effect of heat rejection temperature on cooling capability with 15 W compressor power (EM#1).

310 K reject temperature was met with 18 W of compressor power. At 80 K, the 0.5 W cooling power requirement with a reject temperature of 310 K was met with 16.4 W of compressor power for EM #2, and reflects improved cooling over EM #1 resulting from improvements. Launch Vibration. Preliminary launch vibration tests have been conducted to verify integrity of the hardware. Figure 6 shows the cooler mounted on the vibration fixture and summarizes the launch load vibration environment. Life Testing. Life tests have been conducted on a version of the mini cryocooler which

was utilized for cooling of a high temperature superconducting filter package. In this test, the compressor and cold head were identical to the mini described in this paper except it was a “split” version in which the compressor and cold head were separated by 6 cm to facilitate

packaging. In this test the unit ran continuously for 6,000 hours, except for infrequent power outages.

Figure 6. Launch load environment and test setup.

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Figure 7. Mini pulse tube cryocooler mounted on 6 axis dynamometer.

Figure 8. Induced vibration along drive axis; 60 Hz drive frequency.

Induced Vibration Induced vibration tests were conducted on the cryocooler mounted on the LM ATC six axis

dynamometer. The test setup is shown in Fig. 7 and the measured forces are presented in Fig. 8 in the drive axis. In these tests the units were run at full stroke, with no closed-loop feedback control. Forces in the two lateral axes were below the peak value of 0.22 N. Induced vibration forces with the closed loop of the electronic controller are in evaluation, but are expected to be lower. Table 2 summarizes the EM parameters against the NASA/ GSFC contract specification.

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Figure 9. View of top-assembly and two boards.

Electronic Controller The cooler drive electronics (CDE) is comprised of two major subassemblies: (1) the Control Board Assembly, and (2) the Power Board Assembly. Major features of the design are described below for each subassembly. Photographs of the control board and power board are included in Figure 9. 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 amplifi-

ers. 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. 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 section contains a PWM DC-DC converter to provide the necessary bias voltages to the entire CDE. The converter design is optimized for reduced

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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. A view of the top-assembly was previously shown in Fig. 9. 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 PWMs 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 through-hole and surface mount components. The entire assembly was designed to optimize use of board area leading to a minimum overall weight.

Qualification Status CDE Functional Testing – An EM of the CDE has been assembled and tested. A brief description of the testing follows for each level of assembly.

Control Board Assembly – The control board EM has been functionally tested at room ambient temperature. Simulated acceleration feedback, position feedback and temperature feedback loops were used to verify acceptable performance. Interface through the RS-422 port was

exercised to verify proper commanding and telemetry read-back. All functions operated as expected during this level of test. Power Board Assembly – The power board was functionally tested at room ambient temperatures to verify performance prior to integration with the control board. Each PWM amplifier was exercised and performance data gathered for efficiency, output harmonic distortion and proper drive amplitude control levels. The efficiency exceeded the 90% goals and typically measured 93% at maximum output power. The inrush limiter and power supply were also tested to verify their expected operation. CDE Assembly – The control board and power board were integrated and functionally tested at room ambient temperature to ensure proper operation prior to integration with the compressors. The current control loop was tested in this configuration. Integrated CDE and Cryocooler Testing – Once the CDE was functionally tested, it was integrated with the cryocooler. Testing included verification of the temperature control loop, AFFECS vibration cancellation, and position loop. A summary of the system characteristics and the original design goals is shown in Table 3.

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Figure 10. Load line for high capacity mini version.

ADAPTABILITY AND ENHANCED COOLING CAPABILITY A unique advantage of pulse tube coldheads is that they have very short development time when compared with compressors and Stirling displacers. Thus, with a given compressor, coldheads can be developed for specific customer needs to provide the optimum performance at specific temperatures and cooling loads, or to provide specialized cooling configurations, such as multi-staging , multiple coldheads6 or split systems. We have performed numerous design studies exploring the range of cooling capacities and configurations consistent with our compressor capability in response to specific customer requirements. At the high capacity end, we have achieved 3 W of cooling at 80 K with a redesigned coldhead and 69 W of compressor power. The load lines for this configuration are presented in Figure 10. We have also operated a modified version of the prototype cooler with over 40 W of compressor power and produced 1.25 W at 77 K. This cooler, in a split configuration with a U-tube coldhead, was used to cool an engineering model of a High-Temperature-Superconducting 4 GHz Input Multiplexer for satellite communications systems,7 and accumulated 6,000 hours of semicontinuous running without problems. SUMMARY Lockheed Martin’s Advanced Technology Center, under support from NASA-GSFC has developed a miniature, lightweight pulse tube cryocooler for space applications. The contract calls for delivery of three EM units. EM #1 and #2 have been completed and are under test, and #3 is in final assembly. They produce 0.3 W at 65 K and 0.5 W at 80 K at a reject temperature of 310 K with 18 W and 16.4 W of compressor power, respectively, and weigh only 1.32 kg. The design approach of these units leads to reduced cost and improved reliability. Modified versions of the same size and weight have shown cooling capabilities of 3 W at 80 K and 2 W at 65 K. A simpler, lighter weight controller has been developed to drive the cryocooler. The controller has a mass of 1.6 kg and greater than 90% conversion efficiency with 3 W of overhead power. It has full capability, including temperature control and vibration control and full telemetry to the spacecraft. It has a reduced number of parts contained in two boards which result in lower cost and higher reliability than prior versions.

ACKNOWLEDGMENT This work was supported by NASA/GSFC (the thermal mechanical unit) and by DARPA and Lockheed Martin internal funding (electronic controller).

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REFERENCES 1. D.L. Glaister, M. Donabedian, and D. Curran, An Overview of the Performance and Maturity of Long Life Cryocoolersfor 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 Press, 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 Press, 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.

6. J.R. Olson, V. Kotsubo, and T.C. Nast, “Multiple Pulse Tube Coldheads Driven by a Single Compressor,” Adv. Cryo. Eng., 45, Plenum Press, New York (2000). 7. T.C. Nast, B.G. Williams, V.Y. Kotsubo, J. R. Olson, and D. J. Frank, “Development of a Cryogenic

60 Channel, HTS Multiplexer,” Adv. Cryo. Eng. 45, Plenum Press, New York (2000).

Gamma-Ray Pulse Tube Cooler Development and Testing R.G. Ross, Jr., D.L. Johnson, A. Metzger Jet Propulsion Laboratory, California Institute of Technology Pasadena, California 91109

V. Kotsubo, B. Evtimov, J. Olson and T. Nast Lockheed Martin ATC, Palo Alto. CA 94304 R.M. Rawlings DRS Infrared Technologies, Dallas, TX 75243

ABSTRACT For a variety of space-science applications, such as gamma-ray spectroscopy, the introduction

of cryogenic cooling via a cryocooler can greatly increase the potential science return by allowing the use of more sensitive and lower noise detectors. At the same tune, the performance benefits must be carefully weighed against the implementation cost, any possibility of degraded detector performance associated with the operation of the cryocooler, and the requirement to achieve long life. This paper describes the development, test, and performance of a novel new low-cost, lownoise, high-reliability pulse tube cooler, designed specifically for highly cost-constrained, longlife space missions. The developed cooler marries two technologies: a low-cost high-reliability linear compressor and drive electronics from the 1.75 W tactical Stirling cryocooler of DRS Infrared Technologies (formerly Texas Instruments), and an 80 K pulse tube developed specifically for the compressor by Lockheed Martin ATC. The successful new cooler achieves over 1.6 watts of cooling at 80 K at 23 W/W, and has the advantages of greatly reduced vibration at the coldtip and no life-limiting moving cold elements.

To achieve maximum life and low vibration, the compressor incorporates flat flexure springs for piston support and uses two opposing pistons in a head-to-head configuration with linear drive motors. The pulse tube is a compact U-tube configuration for unproved integration and is mounted to the compressor in a split configuration with a transfer line. INTRODUCTION

The object of this cooler development program was to make it possible to utilize high-resolution germanium (Ge) detectors for planetary gamma-ray spectroscopy on relatively low-cost space missions involving one- to two-year operational lifetimes. Use of a germanium detector cooled to around 80 K provides measurement sensitivities that are on average seven tunes greater than commonly used uncooled scintillation gamma-ray detectors. Until now, radiative cooling to space has been the best method for weight-limited, longduration planetary missions. Now, with reductions in size and power consumption, and improveCryocoolers 11, edited by R.G. Ross, Jr.

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ments in reliability, miniature mechanical coolers offer an increasingly attractive alternative. Their use would eliminate the interface requirement for a three-axis stabilized spacecraft and a broad unobstructed view to space, and would allow experiments in environments where radiative coolers cannot function satisfactorily, notably in orbit around warm planets and on the surface of planets and moons. Comparisons of size, mass, duty cycle and operating temperature make mechanical cooling more attractive than radiative cooling for many planetary missions. Developing a mechanically-cooled gamma ray spectrometer (GRS) for small, low-cost, planetary missions requires a small, long-life, low-cost cooler from which vibration, capable of inducing microphonics effects at the detector, has been eliminated. The cooler solution described here is to combine the compressor from a low-cost, miniature, high-reliability, commercially-available Stirling cycle tactical cooler, with a matched pulse tube expander made specifically for the compressor. The intended result is to produce a cooler that has minimal mechanical motion at the detector, retains the small mass and volume characteristics of the tactical cooler, and, thanks to recent improvements in pulse tube efficiency, requires a relatively low level of spacecraft power. The elimination of the tactical cooler’s Stirling displacer is expected to also add to the reliability and lifetime of the cooler, and substantially reduce the vibration environment at the cold-load interface. The use of a commercially available compressor and the simplicity of the pulse tube design is expected to preserve most of the cost advantage of the tactical cooler relative to the sophisticated long-life space coolers. The same relative simplicity is also expected to translate into additional cost savings by allowing inexpensive tactical-cooler drive electronics to be used. COMPRESSOR SELECTION AND DESIGN FEATURES

Central to achieving the cooler development objectives was the need to acquire a tactical cryocooler compressor with proven long-life potential, low vibration, and compatibility with the needed pulse tube expander in terms of swept volume and operating pressure. The specific design goal was to achieve > 1.1 watt of cooling at 80 K with a compressor specific power of less than 25 watts/watt. An analysis of available tactical cooler compressors led to the selection of an advanced 1.75 W tactical Stirling cryocooler manufactured by DRS Infrared Technologies (formerly Texas Instruments). The particular model, shown in Fig. 1, is based on an advanced linear compressor with its two pistons operated head-to-head and supported on flexure springs to achieve long life and good vibration suppression. This new flexure-supported compressor is one of a family of advanced flexure-equipped compressors being developed to achieve extended-life tactical coolers.1

Figure 1. DRS 1.75 W tactical cooler with drive electronics.

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Figure 2. Refrigeration performance of the DRS 1.75 W Stirling cooler as a function of input drive voltage, coldtip temperature, and coldtip load.

The thermal performance of this cooler, using the Stirling expander shipped with the cooler, is shown in Fig. 2. The more conventional DRS 1.75 W cooler has similar performance, but uses helical coil springs to support the pistons and has a predicted life greater than 5000 hours. JPL has had good success using the conventional (non-flexure-spring) DRS 0.2-watt, 1-watt, and 1.75watt Stirling coolers for a variety of low-cost, intermediate-life space missions.2,3,4 PULSE TUBE DESIGN AND CONSTRUCTION

The second task critical to achieving the required cooler performance was the development of a high efficiency pulse tube expander carefully matched to the compression attributes of the DRS

compressor and the interface requirements of the JPL gamma-ray detector mounting system, shown schematically in Fig. 3. This task, carried out by Lockheed Martin Advanced Technology Center, involved first thoroughly characterizing the DRS compressor, then designing and fabricating a pulse tube consistent with the compressor and the gamma-ray detector cooling load and mounting interfaces. The chosen concept was the U-shaped pulse tube shown in Fig. 3.

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Compressor Characterization To achieve an efficient design for the proposed pulse tube it was necessary to include an accurate model of the compressor's performance in the overall pulse tube design analysis. To acquire the needed data, the DRS flexure-bearing compressor was tested at Lockheed Martin ATC with dead volumes to determine its characteristics. Tests were performed using three different dead volumes (12. 1cc, 22.8cc, and 28.9cc), and five different charge pressures (300psia, 200psia, 100psia, 50psia, and 5psia). The compressor was driven with a current controlled amplifier with an input signal given by an HP signal generator. The power to the compressor was monitored with a Valhalla power meter, and a calibrated pressure transducer, mounted in the compression space, monitored the pressure. The resonant frequency at each charge pressure was determined by a frequency sweep searching for the maximum voltage for a fixed drive current, for drive currents ranging from 0.1 A up to 1.3 A. At the resonant frequency, the current, voltage, power, and pressure amplitudes were recorded. Table 1 presents a summary of the compressor parameters (for each compressor half); most were determined from the measurements, while some were provided by DRS. Compressor internal losses were also characterized to allow estimation of the expected efficiency of the overall pulse tube cryocooler. Because the compressor exit-passage parameters were designed for the standard DRS split-Stirling expander that has a small-diameter transfer line, somewhat higher losses were predicted when used with the pulse tube, which requires a larger transfer line. In the future, if more optimum performance from the compressor is desired, one should consider enlarging the internal flow passages to tailor the compressor for improved operation with a pulse tube.

Pulse Tube Design To achieve an efficient design for the pulse tube, detailed thermodynamic simulations were conducted by Lockheed Martin of the entire cooler system. Key parameters included pulse tube geometries, transfer line diameter, fill pressure, operating frequency, piston stroke, and pulse tube reservoir-line tuning.

The resulting design was predicted to provide 1.2 W of cooling at 80 K with 30 W of total compressor power and a piston amplitude of 2.9 mm. Note that the piston amplitude is well below the maximum of 5 mm. The predicted cooling capacity is slightly higher than the required 1.1 W, and the predicted specific power of 25 W/W matches the design goal. The largest uncertainty in the prediction was the internal losses within the compressor, which, in the dead volume tests, were particularly significant at high piston amplitudes. A conservative empirical model was used to represent the compressor flow losses in the analysis, which tended to reduce the piston amplitude in order to reduce the losses. A series of parametric studies was performed to predict the sensitivity of the coldhead to operating conditions. The efficiency of the coldhead was found to be relatively insensitive to mass flow rates and frequencies, typical of other L-M pulse tubes. This indicates that the coldhead design was not significantly influenced by the particular model used for the compressor losses. Figure 4 shows the (pressure-volume) PV specific power as a function of input power. As shown, the coldhead itself is predicted to have a PV specific power of 14 WAV at 80 K, comparable to other coldheads developed at Lockheed Martin ATC.5,6 Lockheed's best in-line, high-

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Figure 4. Predicted PV specific power as a

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Figure 5. Predicted cooling power dependence

function of input power.

on charge pressure.

capacity designs are typically 11-12 W/W. The slightly lower efficiency of this pulse tube is due in part to the U-tube configuration, the small diameter of the transferline internal to the compressor, and to the smaller cooling capacity. At the lower power levels, the efficiencies are still good,

although slightly decreased from the higher power levels. At 10 W of PV power, the PV specific power is predicted to be around 17.5 W/W. Figure 5 shows the predicted cooling power as a function of charge pressure at 30 W of compressor power. The frequency was varied for optimum performance, along with retiming of

the impedances. The regenerator and pulse tube remained fixed in the analysis. This plot suggests that 400 psia would be a good working charge pressure, and shows that the cooler can tolerate a slight reduction in charge pressure down to 300 psia or so without serious reduction in performance. However, if the charge pressure is decreased down to 200 psia, the performance begins to seriously degrade. If the regenerator and pulse tube were to be redesigned for lower

charge pressures, then the performance degradation would not be as severe as shown in Fig. 5. The results of the overall simulation analyses indicated that the proposed coldhead should perform well over a range of conditions. This is significant in that the compressor loss mechanisms were not known in detail. Once the compressor and coldhead were integrated, it was expected that minor tuning of the overall system would be able to achieve a good match between coldhead and compressor. Based on the modeling it was considered likely that the pulse tube would exceed the predicted

efficiency, since a conservative model was used for the compressor losses, a conservative model was used for the motor force constant, and many Lockheed coldheads outperform their predictions. Thus, it was expected that the coldhead would provide in excess of 1.1 W of cooling at better than 25 W/W. In addition, the low design stroke would allow the cooler to be driven to substantially higher strokes and power levels, although at a somewhat higher specific power.

Pulse Tube Fabrication Once the analyses and component designs were complete, the pulse tube cooler components were fabricated and assembled into a completed pulse tube expander. Figure 6 shows the piece parts ready for assembly, together with a completed pulse tube. Figure 7 shows the complete

cooler setup during verification testing at Lockheed. PULSE TUBE SYSTEM-LEVEL TESTING

After initial checkout and performance verification of the completed cooler at Lockheed Martin ATC, extensive performance characterization testing was carried out at JPL in preparation

for planned tests to validate the vibration and EMI compatibility with an actual gamma-ray detector using the setup illustrated in Fig. 3. Figure 8 presents the overall thermal performance measured at JPL as a function of coldend

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Figure 6. Pulse tube expander piece parts and final assembly.

Figure 7. Completed pulse tube cooler with DRS compressor on the left, reservoir volume on the right, and pulse tube with vacuum bonnet assembly in the center.

Figure 8. Refrigeration performance of the completed gamma-ray pulse tube cooler as a function of input drive voltage, coldtip temperature, and coldtip load.

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161

Figure 9. Refrigeration performance of the completed gamma-ray pulse tube cooler as a function of helium fill pressure.

temperature, coldend load, and input voltage (which is roughly proportional to stroke). Note that the specific-power performance at 80 K is around 22 W/W, which is better than the design goal of 25 W/W, and that the overall cooler capacity is also better than the design requirement, reaching over 1.6 watts at 80 K near full stroke (9 volts), in contrast to a requirement of 1.1 watt. To confirm the cooler's predicted sensitivity to fill pressure and drive frequency, additional parametric testing was conducted with these parameters as variables. The measured performance, displayed in Figs. 9 and 10, confirm that fill pressure increases cooling capacity with minimal effect on efficiency, while drive frequency, once the pulse tube volumes are fixed, is a relatively

sensitive parameter. For the as-fabricated pulse tube cooler, the best specific power is seen to occur at a frequency of around 42 Hz.

Figure 10. Refrigeration performance of the completed gamma-ray pulse tube cooler as a function of drive frequency.

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SUMMARY AND CONCLUSIONS This paper has described the development, test, and performance of a novel new low-cost, low-noise, high-reliability pulse tube cooler, designed specifically for highly cost-constrained long-life space missions such as planetary gamma-ray spectroscopy. The developed cooler marries two technologies: a low-cost, high-reliability linear compressor and drive electronics from the 1.75 W tactical Stirling cryocooler of DRS Infrared Technologies, and an 80 K pulse tube developed specifically for the compressor by Lockheed Martin ATC. To achieve maximum life and low vibration, the compressor incorporates flat flexure springs for piston support and uses two opposing pistons in a head-to-head configuration with linear drive motors. The pulse tube is a compact U-tube configuration for improved integration and is mounted to the compressor in a split configuration with a transfer line. The successful new cooler achieves over 1.6 watts of cooling at 80 K at 23 W/W, and has the advantage of greatly reduced vibration at the coldtip and no life-limiting moving cold elements.

ACKNOWLEDGMENT

The work described in this paper was carried out by the Jet Propulsion Laboratory, California Institute of Technology, and by Lockheed Martin ATC under contract with JPL; it was sponsored via the Planetary Instrument Definition and Development Program (PIDDP) through an agreement with the National Aeronautics and Space Administration. REFERENCES 1. Rawlings, R.M. and Miskimins, S.M., “Flexure Springs Applied to Low Cost Linear Drive Cryocoolers,” Cryocoolers 11, Kluwer Academic/Plenum Publishers, NY, 2001. 2. Glaser, R.J., Ross, R.G., Jr. and Johnson, D.L., “STRV Cryocooler Tip Motion Suppression”, Cryocoolers 8, Plenum Publishing Corp., New York, 1995, pp. 455-463. 3. 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 07803-5427-3, 1999, p. 5. 4. Johnson D.L., “Thermal Performance of the Texas Instruments 1-W Linear Drive Cryocooler,” Cryocoolers 10, Kluwer Academic/Plenum Publishers, NY, 1999, pp. 95-104. 5. Kotsubo, V.,Olson, J.R., andNast.T.C., “Development of a 2W at 60K Pulse Tube Cryocooler for

Spaceborne Operation,” Cryocoolers 10, Kluwer Academic/Plenum Publishers, NY, 1999, pp. 157170. 6. 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 Publishing Corp., NY, 1999, pp. 171-179.

High Efficiency Pulse Tube Cooler E. Tward, C. K. Chan, J. Raab, T. Nguyen, R. Colbert and T. Davis†

TRW, Redondo Beach, CA 90278 † Air Force Research Laboratory Albuquerque, NM 87117

ABSTRACT

The High Efficiency Cooler (HEC) is being developed in order to provide a long life, low mass, high efficiency space cryocooler suitable for use on lightweight gimbaled optics on surveillance missions such as SBIRS Low. This paper reports on the development and testing of this next generation family of space pulse tube cryocoolers which feature high cooling capacity, lower mass,

lower EMI and lower self induced vibration than the current state of the art. The HEC achieves low input power and large cooling power because of the efficiency of its pulse tube cold head and highly efficient compressor. The low mass (<4.3 kg) results chiefly from its next generation Oxford flexure compressor technology reported in a companion paper. The projected long lifetime and high reliability results from use of the proven low complexity flexure compressor and pulse tube cold head. Its low EMI is due to its self-shielding motor. The low self induced vibration results from its internal dynamic balancing. It features the ease of integration into an instrument of a small pulse tube cooler. The cooler achieves its 10 W at 95K cooling requirement with substantial margin while rejecting heat to 300K.

INTRODUCTION The High Efficiency Cryocooler achieves low input power and large cooling power because of the efficiency of its pulse tube cold head and efficient compressor. Its low mass results from the use of second-generation flexure compressor technology developed with Oxford University and productionized by Hymatic Engineering. It achieves long lifetime and high reliability by using a proven low complexity flexure compressor and pulse tube cold head. Its low EMI is due to its selfshielding motor. It achieves low vibration through its internal dynamic balancing, and it features the ease of integration of a small pulse tube cooler. The cooler is being developed in order to provide a low mass, high efficiency cryocooler suitable for use on lightweight gimbaled optics on surveillance missions such as SBIRS Low. Surveillance systems incorporating LWIR focal planes require cooling of focal planes and optics. This capability has long been sought as the solution to the midcourse missile flight detection problem. During midcourse missile flight, the trajectory phase between burnout and re-entry, tracking and discrimination of ballistic missiles, reentry vehicles, and deployed penaids (decoys) are difficult without the use of sensitive LWIR focal planes. In addition, space surveillance of resident space objects for tracking and identification is needed. By collecting photons at the longer waveCryocoolers 11, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers. 2001

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lengths emitted by these cold, dim objects, a vast improvement in identification and discrimination capability with a minimum of sensor aperture growth can be realized. Multiple spectral bands can be utilized to greatly improve the sensor’s ability to determine the temperature and cooldown rates,

which will aid in the discrimination of lethal objects versus decoys. Smaller apertures produce cheaper, lighter more agile sensors, much easier to host in a space-based system. To achieve the sensitivity at the long wavelength requires cooling of focal planes to temperatures less than 40K. Optics cooling is required to minimize background noise and enhance the detection sensitivity of these long wavelength, low temperature focal planes. Such missions cannot contemplate use of this detection capability unless efficient, lightweight cryocoolers are available for cooling of optics. Below the Horizon (BTH) imaging, requires the development of efficient cryocoolers as mission-enabling technology. To achieve the rapid scanning of an agile sensor either for missile tracking Above the Horizon (ATH) or for BTH tracking requires lightweight on-gimbal components. The high heat loads due to BTH imaging, along with the requirement for rapid scanning, drive telescope designs to use cooled on-gimbal foreoptics. For on-gimbal cooling the mis-

sion and careful thermal design determine the heat loads that the cooler must lift to be rejected by a radiator an-gimbal. This radiator capacity is limited by the physical area that can be accommodated on-gimbal, and is the mission limiting component. Given this radiator limit, all possible missions can be accommodated only by improving the cooler efficiencies to the point that radiator heat rejection is no longer the limit. If the cooler is inefficient, heat rejection of on gimbal radiators becomes the factor limiting the system capability. For this reason, cryocooler efficiency is critically important to the system design. In the required 95 K temperature range the current state of the art in specific power (ratio of input power to cooling power at 95 K) for flight qualified cryocoolers is in the range of 12 WAV. Previous flight qualified coolers did not possess the capability to produce the required >10 watt

cooling capacity. The best previous specific mass (ratio of mechanical cooler mass to cooling power at 95 K) was in the range of 1.5 kg/W. This project seeks to make a major improvement to the specific power while at the same time reducing the cooler specific mass by 350% to <0.43 kg/W. The specific power improvement goal will be achieved by developing an exceptionally efficient pulse tube cold head whose performance is optimized at 95K. The specific mass improvement goal will be achieved by using our next generation very low mass flexure bearing compressor technology in the mechanical compressor. Engineering models of these compressors were demonstrated on the IMAS project and to date have reached 3480 hours in life test with no detectable performance change. At the time of this writing the two protoflight compressors are complete and are reported in a

companion paper1 at this conference. Development has succeeded in producing exceptionally efficient development cold heads with a Carnot efficiency of 25% at the 95K, 10W operating point with 300K reject temperature. Flight cold heads are now in fabrication and will be integrated with

the compressors in June 2000. Delivery is scheduled for October 2000. HIGH EFFICIENCY CRYOCOOLER The High Efficiency Cryocooler (HEC) key requirements are given in Table 1. The input power goal is quite ambitious requiring major strides in the cold head design. Despite the very low specific mass of this machine achieving the mass goal is low risk because of the existence of the demonstrated compressor technology.

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Figure 1. High Efficiency Cryocooler.

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Figure 2. IMAS cooler.

The reliability and lifetime goal is representative of pulse tube cryoeoolers of this class which

has been partially verified by life testing and in-orbit operation of similar earlier-generation machines. The HEC conceptual mechanical design is shown in Figure 1. The cooler is an integral configuration pulse tube cryocooler. It incorporates a back to back flexure bearing compressor for vibration balance and the passive in-line pulse tube cold head. The pulse tube cold head is a passive Stirling cycle cooler in which the moving Stirling displacer is replaced with a passive expander. This greatly increases the reliability and producibility of the cryocooler by eliminating cold moving

parts with close tolerances. Otherwise, it follows the same thermodynamic principles and uses the same Oxford-style compressor. Because it is a much newer technology than Stirling cryoeoolers, movement up the learning curve in recent years has been very rapid. It now has the same tempera-

ture production capability and efficiency as Stirling coolers. Ten-year lifetime is achieved by eliminating all wear mechanisms. In the pulse tube cold head this is automatic since it is an all-metal plumbing system with no moving parts. As a result, it is simple to build and has no significant disadvantages. In the Oxford style compressors, long life is achieved via the flexures which eliminate piston wear. The flexure springs are very stiff in the direction perpendicular to the driven motion (much stiffer than gas or magnetic bearings) so that close-tolerance gas-gap seals can be maintained and wearing seals can be eliminated. The flexures themselves are designed for maximum stress levels well below the material endurance limits. Non-fatiguing performance is readily validated in any machine since over 10 cycles are accumulated in 4 to 5 days with these compressors. The working

fluid is inert dry helium with no lubricants. The drive is a direct voice coil motor similar to a loudspeaker driver, thereby eliminating linkages. Oven baking prior to closure reduces volatile condensables and water in the machines to negligible levels. All of these types of coolers are hermetically metal sealed to have effectively zero detectable leakage rates of helium fluid. The processes have been verified by life tests of similar pulse tube coolers including TRW units currently

in orbit. The HEC cryocooler is derived from its predecessor IMAS engineering model coolers (Fig. 2) developed for and delivered to the Jet Propulsion Laboratory.2

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Figure 3. HEC Compressor.

The HEC compressor shown in Fig. 3, incorporates simple and effective mechanical and thermal interfaces with the payload. The single mechanical mounting interface at the compressor centerplate also serves as a conductive interface for removing all the heat from the cooler. To enhance the efficiency and minimize its system impact on the payload to which it is connected, the

centerplate is designed to provide heat spreading, allowing the cooler to be mounted directly to a radiator. A secondary alternative warm mechanical/thermal interface is provided at the centerplate where the cold head attaches to the compressor. This provides the flexibility of changing cooler orientation as well as the option of a vacuum interface for ground testing. The larger diameter end cap incorporates the pulse tube cold head reservoir tank. The 10W at 95K requirement has been met in development testing with an engineering model compressor and development cold heads. At its present development stage the cold head is already more efficient at this temperature than other flight units. Figure 4 gives the leadline through the 10W at 95K operating point for the most efficient of the development cold heads. Figure 5 gives the efficiency of this cold head as a function of temperature while rejecting to 300K. The low radiated magnetic field (RE01) of the IMAS mechanical cooler at 75 watts input power shown in Figure 6 was measured at JPL. This excellent performance at the cooler operating frequency and harmonics results from the patented self-shielding voice coil motors. This basic design has been carried over to the HEC and therefore identical performance should be obtained at the same power level. A second major benefit of this compressor design is the very low self-induced vibration over a wide frequency range that is achievable with the cooler. This results from the very rigid motor design and the degree of balance of the two compressor halves. Other work is ongoing to better

match the compressor halves to further reduce the need for active vibration cancellation for many applications. The output force of the IMAS cooler shown in Figure 7 was measured on a Kistler

Figure 4. Cooler load line through 10W at 95K.

Figure 5. Cold Head Efficiency.

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Figure 6. Radiated Magnetic Field (RE01) for the IMAS Cooler.

Figure 7. Self-Induced vibration on drive axis with 80 W input

power to cooler and tailored waveform.

dynamometer using a fixed predetermined tailored drive waveform to minimize the harmonics. The force shown for 80 watts of input power is along the axis of motion of the compressor pistons. In orbit the tailored waveform can typically be determined autonomously and updated in real time, if necessary, by the existing flight electronics using feedback from an accelerometer mounted on the cooler. ACKNOWLEDGMENT This project was supported by the Air Force Research Laboratory under contract F29601-98C-0179.

REFERENCES 1. High Performance Flight Cryocooler Compressor, P.B. Bailey, M.W. Dadd,, N. Hill, C.F. Cheuk, J. Raab and E. Tward, Cryocoolers 11, Plenum Publishing Corp., New York (2001).

2. 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.

High Performance Flight Cryocooler Compressor P.B. Bailey and M.W. Dadd

Oxford University, Oxford, UK

N. Hill and C. F. Cheuk The Hymatic Engineering Company, Ltd. Redditch, UK J. Raab and E. Tward TRW, Redondo Beach, CA, USA

ABSTRACT

In this paper we report on the development of a next generation flexure bearing compressor which features high efficiency, high capacity per unit mass, enhanced producibility and ease of integration into payloads. The compressor was developed for the 95K High Efficiency Cryocooler programme.

The compressor achieves low mass by using small diameter flexure springs and having a new compact design of magnetic circuit which also has the advantage of being self shielding,

thus reducing the radiated magnetic field. A pair of compressors mounted back to back and driven in anti-phase provides low levels of self-induced vibration, which is further improved by the rigidity of the motor and the characteristics of the new motor and spring designs. Its ease of integration results from its compact size and the incorporation of a single thermal

and mechanical mounting interface in its centreplate. The centreplate incorporates heat spreading both internally for removing compressor heat as well as for spreading the heat to the radiator to which it can be attached. Producibility has been achieved by transferring the processes developed for manufacturing a similar Oxford designed long life tactical cryocooler. The compressors are being manufactured by Hymatic to a design which has evolved from earlier machines made by Oxford University. TRW will integrate the compressors into the flight qualified 95K High Efficiency Cryocooler which will be delivered to AFRL in October 2000. INTRODUCTION

A new type of compact linear motor and a new flexure spring design have been developed by Oxford University for linear compressors with the aim of meeting stringent requirements for high efficiency and low mass. The compressors are a key part of the High Efficiency Cryocooler

(HEC) which is being developed by TRW. Cryocoolers 11, edited by R.G. Ross. Jr. Kluwer Academic/Plenum Publishers, 2001

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Prior to this programme 3 single compressors (all with active balancers) and 4 balanced pair compressors were made and assembled at Oxford for TRW, together with a fifth balanced machine assembled by TRW. Two of the balanced pair compressors were delivered by TRW to NASA/JPL for the New Millenium IMAS project1. These compressors, which have an identical motor design to the High Efficiency Compressor discussed here, have been subject to extensive thermal, vibration and EMI testing. The original design has now been taken a step further in a three way 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 production. The assembly and test processes have drawn on the manufacturing and process technology from the Hymatic tactical Stirling cooler2. In the first half of 2000 two HEC

compressors have been built at Hymatic, using the new assembly and test processes and have been delivered to TRW. DESIGN PHILOSOPHY

The key feature of the Oxford design philosophy is the holistic approach to design taken too many cryocoolers have been designed that are almost impossible to assemble. From the outset, details are incorporated into the design to aid the assembly and testing of the machine. Another important feature is the elimination of many delicate and precise components, which are replaced by simpler parts, combined with extensive use of jigs and fixtures. This approach lends

itself to larger production quantities, rather than the 'one-off' approach to earlier builds.

COMPRESSOR DESIGN

The compressor is based on the well-proven 'Oxford' principles of spiral flexure springs and non-contacting clearance seals. The machine is a compact moving piston design, with the piston and cylinder located within the core of the magnetic circuit of a moving coil motor. The springs are the only component subjected to significant fatigue loading, and these are routinely batch tested at a minimum 25% overstroke to in excess of 108 cycles. The springs have been qualification tested, and the results from this predict a single spring arm reliability of 0.999998 and a reliability for the 96 spring arms used in each compressor of 0.9998. The linear motor powering the compressor is a new moving-coil design that features a very compact magnet circuit with low flux leakage and consequent high motor efficiency for the size and power. The coil is fully supported on a former, and special attention was paid to maximize the fill factor and increase motor efficiency. The structural integrity of the coil former facilitates transmission of driving forces without relying on the variable strength of the coil potting adhesive, thereby eliminating a common source of compressor failure. The motor design is self-shielding and features extremely low levels of radiated magnetic field. Test on the similar IMAS compressor showed that the compressor essentially met the requirements of the MIL-STD-461C RE01 test specification measured at 7cm distance3. The design of the compressor is such that it is inherently well balanced with the two 'compressor halves' mounted in line and operated in anti-phase. Tests on the MAS cooler indicated very low levels of self-induced vibration with 30 W of sine wave input power. The only harmonic above 40mN rms was the second harmonic, and this .probably arises from a mismatch in the 'mechanical zero' position between the two compressor halves3. The two identical compressor halves are mounted on an aluminium alloy 'Centre Plate' that contains all of the cryocooler interfaces. Electrical power is supplied by means of a hermetic feedthrough which is Electron Beam welded to the centreplate. One face of the centre plate provides 45cm2 of thermal interface for heat rejection. A large flange is provided around the connection to the pulse tube for a vacuum tank to be fitted during testing. Gas containment is by means of metal 'O' rings between the end caps and centre plate and an aluminium gasket to seal the fill port. A leakage rate ofbetter than 10-7 mbar litre/sec has been achieved consistently.

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Figure 1. Completed compressor.

Particular attention is given to the locking of fasteners and component stacks to preserve the alignment against vibration and redundant locking devices are used to enhance reliability.

The compressor has been designed to operate with a nominal 100 Watts of input power with an additional 50% input power margin. The compressor has an overall length of 226 mm, with end caps 57 nun in diameter and a mass of 2.45 kg. PRODUCABILITY

Many of the assembly processes involve bonding, and these processes are usually irreversible. Hence the key to quantity production of cryocoolers is the verification (where possible) of each and every stage of the production process, from component manufacture to final assembly.

Material Selection

Materials used in the assembly such as plastics, adhesives and primers are selected from an existing knowledge base of materials with a low out-gassing rate. New materials are extensively tested for their margin of being rendered clean by vacuum bake-out processes. A quadrupole Mass Spectrometer is used extensively for this purpose. Only traceable materials are used in the manufacture of components. Component Stage Geometric tolerances commensurate with the functional requirements of the components are specified. During component manufacture, the geometric tolerances form the basis of the method of work holding, while the surface finish requirement defines the manufacturing process. The success of the component manufacturing processes stems from the realisation of the differences between metal and plastic components when they are machined to tight tolerances. Non-contact measuring methods including laser, optical and air gauging are used extensively to

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verify the success of the manufacturing processes. Plans for transit and storage protection for each component are designed into the manufacturing processes from the raw materials to the

finished products. Verification and Development of Assembly Processes and Tooling

Assembly processes are verified by testing as defined in process development plans. Sufficient quantities of test pieces are manufactured to render these tests statistically significant. The objective of these tests is to determine the capability of the processes. Data derived from these tests is then used as the standard to which subsequent assembly operations are controlled, rather than the less stringent functional requirement of the assembly itself. If a process is capable of performing to a standard, it is considered to be under control only if it consistently attains the same standard within the natural variation of the process. The design of the Compressor relies heavily on the use of tooling. The dimensions and geometric tolerances of tooling is inspected and verified before being released for production. Component Inspection

For the initial builds, 100% inspection of all dimensions is being implemented, but as the component manufacturing processes become fully defined and robust, this will be gradually replaced by 100% inspection of 'critical dimensions only' in future builds. 'Goods Inward' inspection of components in itself is not sufficient. It is a truism that

components are at their best immediately after manufacture - from then on it is downhill - every operation, from finishing (deburring/frazing), inspection, cleaning and transport has the potential to damage components. For this reason it is vital with critical components to have a functional inspection of the part immediately prior to assembly. A detailed inspection plan has been compiled specifying the functional inspection requirement of critical parts and the method of inspection, which mimics the assembled conditions of the components. During the initial builds, this inspection procedure will detect errors on parts indicating that the processes need refining. This gateway enables improvements on the quality of components to be made before they are assembled beyond the 'point of no return'. In-Process Testing

Where practical every stage of the assembly process is verified by some form of in-process testing, both to test the validity of the actual process itself and also to ensure that the process has not had any secondary deleterious effects on the assembly. Experience has shown that one of the main problem areas is the clearance seal – it is difficult to achieve the correct clearance and easy to lose it. Hence many of the most important tests are those which check the alignment of the assembly and show that there is no friction between piston and cylinder. Placing the friction and alignment tests strategically in the assembly process, the consistency of the free frictionless movement of the finished compressor

can be assured. Test Facility. Many of the tests are carried out on a computer-controlled test rig which also functions as a data logger. Using digital-to-analogue converters the computer controls both DC and AC amplifiers for powering the compressor, together with signal conditioning and data logging functions. Among the functions available is the facility to continuously monitor the drive coil temperature and to shut down any test should this temperature become unacceptably high. This facility is essential in some of the DC tests, which are slow and could easily lead to a

coil burn-out if the tests were carried out manually. Alignment Test. This test verifies the capability of the spring suspension system to effect a linear motion of the piston within the cylinder such that the clearance between them is maintained. The test is carried out immediately after the spring stacks are assembled and aligned. The assembly is then locked to prevent movement, and the test is repeated before the piston is fitted.

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Figure 2. Alignment Test – typical result.

To perform the test, power is applied to the motor, which is taken through three complete stroke cycles and the run-out error measured. A least-squares polynomial is calculated through the data points; figure 2 shows a 'screen dump' of a typical test result. Note that the sensor used exhibits some backlash, but this is annulled when the polynomial is calculated (visible in the centre of the trace). The repeatability of the measurements is excellent – apart from a

certain amount of 'bedding-in' during the first cycle, the three curves follow each other within 0.1 µm. The repeatability shown here is an excellent demonstration of the flexure spring suspension system. The results of this particular test must be treated with a certain amount of caution, as the test is recording not only the linearity of the motion of the cylinder, but also measuring the straightness of the cylinder itself. Thus to make any sense of the test, the 'cylindricity' of the cylinder must be good, and is typically less than l.5µm. Much of the 'noise' apparent on the

trace is repeated on successive test cycles and is due to the surface finish of the cylinder. ' From such readings a complete picture of the combined linearity and cylindricity can be built up, and this is then displayed as a three-dimensional plot (figure 3). From this plot the effect of form error and alignment error can easily be separated and the true alignment error evaluated. Friction Tests. Tests are used to evaluate the friction between piston and cylinder - one for dynamic friction, and one for static friction. The tests are computer-controlled and are performed at several stages throughout the build, both before and after the piston is fitted. The dynamic test (low frequency sweep) involves driving the compressor through one complete cycle using a 0.01 Hz triangle wave. A curve of current against displacement is plotted and studied for discontinuities and to observe the size of the hysteresis loop. Process Capabilities. Using the in-process tests outlined above the capability of the

alignment and piston assembly processes can be quantified. The alignment process is expected to produce a mean error of 2.6µm with a standard deviation of 1.67. 99% of the spring suspensions systems that have been aligned by these processes are expected to have an alignment error of less than 6.5µm. The smallest diametrical clearance between the piston and cylinder should then be larger than 13µm for a true frictionless clearance seal design to be realistic.

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Figure 3. Alignment test - typical result in 3D plot.

CONCLUSION A compact, high power and producible compressor design has been achieved.

A build

procedure has been formulated and line qualified to ensure the consistency of the quality of the compressor. Alignment accuracy of less than 6.5µm (peak-to peak) has been achieved with no evidence of friction in the clearance seal. Innovative in-process test procedures have been designed and are expected to benefit the future manufacture of cryocooler compressors. ACKNOWLEDGEMENT

We acknowledge the strong support of Thom Davis of AFRL for this project. REFERENCES 1. Chan, C.K., Nguyen, T., Colbert, R., Raab, J., Ross, R.G. Jr., Johnson, D.L., "IMAS Pulse Tube Cooler Development and Testing", Cryocoolers 10, Plenum Press, New York (1999), pp 139-147. 2. Aubon, C.R., Peters, N. R., "Miniature Long Life Tactical Stirling Cryocoolers", Cryocoolers 9, Plenum Press, New York (1997), pp 109-118. 3. Tward, E., Davis, T., "High Efficiency Cryocooler", Proc. AIAA Paper No. 99-4564,1999.

Vibration Reduction In Balanced Linear Compressors M.W.Dadd, P.B.Bailey and G.Davey Oxford University, Oxford, UK T.Davis and B.J. Thomlinson Air Force Research Laboratory, Albuquerque, NM, USA

ABSTRACT

Coolers for Space applications are often powered by reciprocating compressors that use a linear compressor technology. These can deliver the requirements for long life and high reliability but have not yet produced acceptable uncompensated vibration levels at a reasonable cost. If two nominally identical compressors are mounted back to back the vibration level is reduced, but may still be too high for many applications. Further reduction of vibration is achieved through the use of Adaptive Control systems, which are expensive and reduce the reliability of the system. If the residual vibration can be reduced by better matching of the two compressors, then cheaper, more reliable electronics can be used to achieve the desired vibration level. Under a research and development effort with the Air Force Research Laboratory, all sources of vibration were considered but effort was concentrated on improving the matching of a compressor pair to limit the main causes of vibration. The dynamics of a compressor were modelled. The force generated by the coil was calculated from flux densities determined by a finite element analysis of the magnetic circuit. The rest of the system was modelled as a damped harmonic oscillator. An attempt was made to reduce the residual vibration in an opposed pair of compressors by duplicating the model to simulate a compressor pair and investigating the effect of small variations of a number of parameters. These results were used to estimate the accuracy with which various parameters must be matched to achieve a certain residual vibration and this information was then used to improve the assembly of a compressor pair. Most of the components for this compressor were already available so it was not possible to make major changes. Detailed measurements were made on all components and assemblies so that the vibration spectrum could be related to a compressor with documented manufacturing and building standards. INTRODUCTION

The Oxford University cryogenics group has been involved in the development of space cryocoolers for many years and has produced a number of prototype coolers/compressors. A prototype balanced compressor pair (designated the Capital Compressor) that had been designed Cryocoolers 11, edited by R.G. Ross, Jr.

Kluwer Academic/Plenum Publishers, 2001

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and manufactured at Oxford for a recent TRW space project, was found to have an inherently low vibration level The compressor halves had been assembled carefully to achieve true

clearance seals but there had been no conscious effort to produce a low vibration machine. The very low vibration levels required by space programmes are currently achieved by complex drive electronics that use adaptive control algorithms. Whilst effective, this approach has significant drawbacks: • The complexity of the electronics hinders the achievement of high reliability • The additional electronics increases cost and payload. • A basic adaptive control approach is only able to improve on-axis vibration levels. For the reduction of off-axis levels, additional force transducers and control electronics would be required with a further impact on reliability etc. An approach that suggested itself is the further reduction of the uncompensated vibration by improved matching of the two compressors through better matching of components and closer control of build processes. To achieve this goal it was recognised that a better understanding of the forces operating would be desirable and that methods should be found to allow build quality and component matching to be evaluated and documented. Some work has already been done on the development of a compressor model using a dynamic simulation package called Vissim. This was used to model an existing unbalanced compressor and was found to give good agreement with measured operating parameters that included out of balance forces. This work was reported in ref(l). The project that evolved had the aim of • Looking more closely at the forces acting in the compressors to see whether the existing model is adequate. • Looking at what dimensional and alignment limits are required for good balance, and what

the practical limitations might be. • Documenting the assembly of a balanced compressor pair. It was hoped that some improvement might be made over previous assemblies but this might not be possible given the restriction of using largely existing components.

• Comparing the measured vibration of the assembled compressor with the values generated by the model At the time of writing, the compressor build had not been completed so this paper will concentrate on describing the approach to modelling and build evaluation. FORCES ACTING IN A LINEAR COMPRESSOR

The types of linear compressors used in space cryocoolers have the virtue that the systems offerees operating in them are amenable to relatively simple descriptions. Figure 1. shows the main components of a typical linear compressor. These are: • A piston/cylinder assembly utilising a no-contact clearance seal • A linear motor – in this case a moving coil “loudspeaker” type motor • A suspension system comprising of two aligned sets of flexure bearings – these are usually springs with a flat spiral geometry and are known as spiral springs or flexures. The characteristics of the suspension system are extremely important to the operation of the compressor. The geometry and alignment of the springs accurately define an axis along which the spring stiffness is low and a large movement (e.g. > 10 mm) is possible without exceeding the fatigue limit of the spring material. Perpendicular to this axis the stiffness is extremely high. The obvious axial symmetry of this type of compressor and the distinct characteristics of the suspension spring assembly suggest the division offorces acting into three types: • Forces acting along the compressor axis where the spring stiffness is low– these will be termed “On-axis” forces. • Forces acting perpendicular to the compressor axis where the spring stiffness is very high – these will be termed “Off-axis forces.

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Figure 1. Schematic diagram of typical compressor.

• Forces associated with the rotation of the spiral suspension springs. The operation of these springs generally gives rise to a small rotation of the moving assembly. The forces exerted on the moving assembly, are also off-axis forces but they resolve to a zero net force and a

well-defined moment acting about the compressor axis. ON-AXIS FORCES

The On-axis forces are the dominant forces acting in the compressor. The principal forces with large magnitudes are those intrinsic to the compressors operation: • The force generated by the coil • The force produced by the suspension springs as they are deflected. • The force produced by gas pressure differences acting on the piston area. These are balanced by the inertia force acting through the centre of gravity as the moving assembly accelerates. Ideally the axis defined by the movement of centre of gravity and axes along which these forces act should coincide. The result would then be a single force acting along a single well defined axis and there would be no resulting moments. In practice there will be offsets and angular misalignments between the axes and these will result in off axis forces and moments perpendicular to the compressor axis. The radial stiffness and separation of the two sets of suspension springs determine the actual movements resulting from such misalignments. If the magnitudes of the axial forces are known then the level of misalignment that can be tolerated is readily calculated. There are other effects that will give rise to axial forces such as: • Windage • Eddy current effects due to conductive components moving in the motors magnetic field • Shear forces acting on the piston due to gas flow in the clearance seals • Forces generated by sections of the coil leads moving in areas of stray flux. These are much smaller in magnitude and would only have any significance if: • they were so uncontrollable that they resulted in big differences between the two compressor assemblies.

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•

They were large enough and acted far enough off the compressor axis that they produced sufficient moments to cause radial deflection of the springs. It is clear that windage and eddy currents could only have any net effect if there was fairly gross mismatching of the compressors. The forces generated by coil leads will be very closely matched but will not be acting along the compressor axis. The shear forces acting on the piston are sensitive to the dimensions and geometry of the clearance seals, which may not be closely matched. Frictional forces should not be present because in principal there are no contacting surfaces. The detection of vibration caused by friction would indicate a fault condition. The forces generated by friction will be very variable depending on small details of surface form etc so that it is not possible to be too specific about its form. OFF-AXIS FORCES

Off axis forces can only register a net effect because of some deviation from symmetry either required in the design or as a consequence of imperfections in the compressor build. Because of the high radial stiffness only large forces will produce significant displacements. There is also the possibility that smaller forces may have enough leverage to cause a significant angular deflection. Some possible forces that were considered are: • Unbalanced gas forces acting on the sides of the piston •

Forces generated by current carrying leads interacting with the motor flux

•

Forces generated by deviations from linear movement of the suspension system The forces acting on the sides of the piston are large and are determined by the detailed geometry of the clearance seal. Defects in the clearance seal geometry can lead to significant

imbalance and misalignment of these forces. The radial gap in a clearance seal is typically around 10 microns for reasonable seal efficiency. With machining tolerances typically around +/1 micron and a similar tolerance on alignment it is clear that this effect merits careful consideration. The forces generated by the current leads are an inevitable part of a practical design. Their magnitudes are readily estimated, and are likely to be insignificant. Off–axis Forces generated by inaccuracies in the linear movement of the suspension system need to be considered but it is likely that the linearity required for the clearance seal will automatically ensure that they are not significant. MOMENTS ABOUT COMPRESSOR AXIS

The magnitude of the rotation produced by the springs will be determined by certain aspects of the spring geometry, principally the spiral arm length and curvature. The manufacturing

processes used allow these parameters to be controlled to tight limits, typically better than +/0.1%. It is clear that provided that the axes of the compressor halves are well aligned and that the compressor strokes are well matched, any residual moment is likely to be insignificant. RESONANCES

In addition to the forces already described there remains the possibility that mechanical assemblies have resonant modes that may be excited at particular frequencies. Small out of balance forces, whether on-axis or off-axis, could be amplified to the point that they become a problem. The general approach to this problem is to keep the resonant frequencies of components as high as possible by designing the moving assembly for high stiffness. The use of materials that have some intrinsic damping can also be considered although the opportunities to do this are limited by the need for other mechanical properties. The moving assembly/suspension

spring system does present a particular instance where resonances may be a problem. The stroke required of the springs limits their stiffness to relatively low values.

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There are a number of vibration modes that should be considered: • Resonance of the spring arms – the inner and outer ends being nodes and the centre being an anti-node. • Torsional resonance of the moving assembly about the compressor axis • Torsional resonance of the moving assembly perpendicular to the compressor axis • Radial resonance of the moving assembly with respect to the compressor axis. These modes of resonance were investigated in the Capital compressor being built and the resonant frequencies were determined to be at least 400 Hz. A STRATEGY FOR GOOD BALANCE

The aim for a well-balanced compressor is that each compressor half should only produce an On-axis force and that these forces should equal so that when they are aligned back to back they cancel out To approach this goal the design has to be effective, the compressor components need to be well matched and the build quality has to be adequately controlled. The overall strategy that has been adopted is: • Close matching of the amplitude of On-axis forces. • Alignment of On-axis forces to avoid generating unnecessary couples. • Minimising of Off-axis forces • Design of suspension system with high radial stiffness to minimise deflections. • Mechanical design that avoids assembly resonances in the operating frequency range • Design which includes damping (where possible) to minimise amplitudes of resonances IMPROVING THE BALANCE OF THE CAPITAL COMPRESSOR

The above has described in general terms the effect of different forces and their possible sources without detailed reference to their magnitudes and real significance. For the Capital compressor the Vissim based model described below was used to calculate the principal On-axis forces for typical operating conditions. The effects of mismatching certain parameters were investigated. Also measurements were made of the torsional and radial stiffness of the compressor’s suspension system. This information was used as the basis for deciding which of the effects described were likely to have any real impact on the overall balance. For the Capital compressor, areas open to some improvement were identified as: • Better matching of principal On axis forces – reducing differences in axial offsets of coils, better matching of moving masses, motor parameters i.e. no of turns in coil, field in air gap • Reducing radial offsets between principal On axis forces • Improving geometry of clearance seals by closer control of component manufacture and alignment • Improving angular alignment of compressor halves with respect to each other ON-AXIS FORCES AND SPRING STIFFNESS FOR CAPITAL COMPRESSOR

Maximum values of principal On-axis forces 100 N Minimum radial stiffness of suspension system 760,000 N/m Torsional stiffness perpendicular to compressor axis (about C.O.G.) 780 Km/radian If a maximum of 1 micron radial movement of any of point on the piston is set as a criteria then maximum allowable values can be set for Off-axis forces and moments perpendicular to compressor axis. With the peak value of the principal forces, maximum angular misalignments and offsets can also be defined: Maximum Off-axis force 0.76 N Maximum moment perpendicular to compressor axis 0.023 N.m Maximum angular misalignment for On axis forces 0.0076 radians Maximum offset for On axis forces. 0.23 mm

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MODEL USED TO INVESTIGATE PRINCIPAL ON-AXIS FORCES

The model used to evaluate the effects of build mismatches is briefly described below. The differential equation defining the motion for the On-axis forces described above is: x is the displacement of the piston assembly from its rest position m is the moving mass, is the spring rate of the suspension springs is the pressure difference acting across the piston over an area A.

where P is cycle pressure, is pressure in compressor body F(x,t) is the force generated by the driving coil. The variation in is primarily generated by the cycle pressure. This is determined by the thermodynamic and pressure drop processes occurring in the refrigeration cycle and cannot be simply described. An approximate model, that is simple and useful, can be developed by treating the gas force as a spring/ damper combination The resulting differential equation is that for a damped harmonic oscillator:

c is effective damping constant for the gas, k is total spring rate, effective spring rate of the gas. The force generated by the coil is given by

is the

is the conductor length per axial length of coil, i(t) is the current through the coil The differential equation defining the behaviour of the moving coil motor as an electrical system is:

V(t) is the applied voltage, E(t) is the back emf generated by the coil, R is the coil resistance, L is the coil inductance. E(t) is given by:

The values of integral

that are required for Eqs. (2) and (4) are accessed in

Vissim as a “Look up” table. The look up table values were calculated in a spread sheet using flux distribution defined by:

These equations were derived empirically using values generated by a finite element analysis of the magnetic circuit. SOME RESULTS OBTAINED FROM THE MODEL

The work described in Dadd et al.1 showed that model values and values measured on a particular unbalanced compressor were close for all the main parameters i.e. instantaneous values of forces, currents and voltage inputs. The only value that had to be adjusted to obtain good agreement was the coil inductance. A similar comparison will be made with the Capital compressor when its build has been completed. The model was used to investigate how the mismatch of particular build parameters would

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Figure 2. Variation of net force with axial offset of coil on one side.

Figure 3. Vibration Spectrum for coil offset of 0.2 mm.

effect the balance such parameters could be moving mass, inductance, magnetic flux distribution etc. As an example, one finding is the dependence on the mean position of the coil in the magnetic circuit. This can vary because of the build up of tolerances in the compressor build. Figure 2. shows how the residual force varies with the changes in offset for one of the compressors. Figure 3. shows the frequency spectrum for the case where one coil is centred and the other is offset by 0.2mm. It will be seen that the residual force is becoming quite significant i.e. 1% of value for each half and that it is dominated by the second harmonic. MEASUREMENTS AND DOCUMENTATION OF COMPRESSOR BUILD

An important and time-consuming part of the work described in this paper was the measurement and documentation of various parameters during the build. These measurements included: • Component masses • Coil characteristics – resistance, No of turns and Inductance • Magnetic flux distribution in gap • Geometrical and dimensional features of individual components e.g. concentricity • The geometrical features e.g. alignments of the compressor assemblies as they are built up.

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The purpose of these measurements was twofold. Firstly to establish a build specification on which to base any modelling or interpretation of measurements. Secondly to actually identify areas where mismatches were significant or where components were not manufactured to adequate standards. Many of these measurements e.g. mass, coil resistance etc required only ordinary laboratory equipment and will not be discussed further. However the measurement of geometrical features and alignment required specialised metrology equipment. Many of these measurements also needed to be taken during the compressor build when cleanliness is very important. To enable inspection measurements to be made in a clean environment a TalyRond 100 was installed in the clean room. The essential features of this type of instrument are: • An accurate turntable incorporating centering and levelling adjustments • Vertical and horizontal columns allowing accurate positioning of probes • A number of sensitive displacement probes capable of determining surface forms e.g. roundness, concentricity, surface roughness etc. Whilst taking these alignment measurements it became clear that the process could be helped by having specific measurement features e.g. easily accessible reference surfaces. This was not possible with the capital compressor but may be considered in future designs. CONCLUSIONS

Although it is not possible to comment with the advantage of actual vibration measurements, some conclusions can be made: • The existing model appears to be adequate for describing the principal On-axis forces • The magnetic properties of the magnets and magnetic circuit components are variable. Good matching of these components is helped if they are chosen from within a single batch. • The matching of the on axis forces requires matching of coil offset as well as other parameters • The offsets between the axes of the principal forces can cause significant moments if their alignment is not adequately controlled • The clearance seal geometry needs to be closely controlled to avoid significant unbalanced forces acting perpendicular to the compressor's axis. • Future compressor designs could usefully incorporate specific features to facilitate measurements on build alignment In many ways good balance is synonymous with high build quality – i.e. both are concerned with good alignments, absence of friction and the minimising of resonance. Low vibration compressors require a high build standard: a high build standard may be demonstrated by low vibration measurements. Assessing build quality is an important issue because of its effect on reliability. The pursuit of low vibration may therefore have added benefits.

ACKNOWLEDGEMENTS The Capital compressor was designed and built under a contract with TRW, who also made compressor components available for this work. REFERENCES

1. Dadd, M. W., Davey, G., Lion Stoppato, P.F., Bailey, P.B., “Vibration Reduction in Balanced Linear Compressors in the 17th International Cryogenic Engineering Conference”, ICEC 17, Institute of Physics Publishing, Bristol (1998), pp.127-131.

95 K High Efficiency Cryocooler Program Kenneth Price1 and Capt. Vladimir Urbancek2 1

Raytheon Systems Company El Segundo, CA, 90245, USA 2

Air Force Research Laboratory Kirtland AFB, NM, 87117, USA

ABSTRACT

The Air Force / Raytheon 95K High Efficiency Cryocooler (95K HEC) Program is developing a new two-stage hybrid Stirling-pulse tube space qualified refrigerator with high heat lift capacity, high efficiency, low weight and size, and low production costs relative to the current state-of-the-art. The basic program will deliver a protoflight Stirling-class Thermo Mechanical Unit (TMU) with protoflight radiation hard electronics. The cooler is designed to support 10W heat lift from a 95K source to a 300K sink. Motor power consumption is to be less than 100W and system power (including electronics) is to be less than 137W. The cooler is to weigh no more than 6Kg. The TMU cold head and compressor designs are highly versatile to enable low cost tailoring to meet the needs of a wide variety of applications. The first demonstration of this versatility is a program option to deliver a companion high-capacity 35K cryocooler. This cooler will also have an aggressive efficiency requirement. The 95K and 35K TMU will share over 95% of components, resulting in significant production efficiencies. Another result of this high degree of commonality is that each cooler can be powered and controlled by standardized Command

and Control Electronics (CCE). The only adjustments needed to match the CCE to a TMU design are in selected logic parameters stored in ROM and in minor changes to winding ratios in two transformers. The CCE is designed with radiation hard components, but the initial protoflight units will be delivered with lower cost commercial substitutes, where available with the same form, fit, and function. The 25 month program will deliver a fully flight qualified 95K system

including both TMU and CCE and, if the option is exercised, a similar flight qualified 35K system. INTRODUCTION

The Space Based Infrared System (LEO), or SBIRS Low, is part of the our nation’s Ballistic Missile Defense program and it’s proposed function is to detect and track ballistic missiles during mid course flight, which is the trajectory phase between burnout and reentry. Discrimination of ballistic missiles, reentry vehicles, and deployed decoys during this phase is difficult, if not impossible, without the use of long wave infrared focal planes. Achieving the sensitivity necessary to identify and discriminate among these cold dim objects requires cooling of the focal plane to temperatures less that 40K. Optics cooling is also required to reduce Cryocoolers 11, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2001

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background thermal noise and improve detection sensitivity at these long wavelengths and low temperatures. An advanced study of a potential SBIRS Low architecture called for gimbaled fore optics at 100-110 Kelvin that required cryogenic cooling at 95K from the fore optics cooler. The study also called for tracking sensor/shielding cooling in two stages at 35K/60K. It was quickly determined that state-of-the-art cryocoolers and/or radiators would not be able to achieve the

desired temperatures and heat loads in a reasonably sized or efficient package. Thus, the government embarked on the High Efficiency Cryocooler program. The United States Air Force is sponsoring a program to design, develop, fabricate and deliver space qualifiable cryocoolers to meet SBIRS Low Technology needs. These coolers will support the SBIRS Low cooling performance objectives for the potential on-gimbal optics design at an initial design point of 95K with 10 Watts of cooling. The contract includes an option to modify the 95K design and then deliver a cooler that produces two-stage cooling at 35K and 60K for tracking sensor cooling. An added benefit of exercising the option task is that the two versions of the cooler (95K and 35K/60K) may be close enough to allow endurance testing on one unit and qualification testing on the other thereby producing a fully space qualified design. Raytheon’s High Efficiency Cooler program kicked off in August 1999. The contract is an innovative cost and risk sharing agreement between the government and Raytheon. Under the arrangement Raytheon has assumed the risk and cost associated with the design of the expander module and electronics in exchange for full data rights to these designs. The Air Force, in the spirit of acquisition reform, is funding the lower risk design of the compressor and the complete

system fabrication costs. This approach allows the Air Force to develop a new cooler for a third to a half of the cost of going it alone. Though the Air Force has limited data rights, it will get full benefit of the design because the contract assures that the cooler will be made available to any DoD contractor at a reasonable cost. On the technical side, the program was selected because it proposed developing a unique and innovative two-stage hybrid Stirling-pulse tube design which promises to be easily modified for split heat loads and potentially used in tracking sensor cooling at 35K and 60K. This is a true two-stage design, not a one-stage cooler with regenerator heat intercept that has in the past been mistakenly called two-stage. The design also allows angling of the second stage from 0 to 90 degrees relative to the first for unique ease of integration of the cold head into confined cryogenic spaces.

The Raytheon 95K High Efficiency Cooler (95K HEC) program will deliver a complete protoflight quality cryocooler including compressor, expander, and flight like electronics in September 2001. The system will be fully flight qualifiable except that contract allows the use of non-radiation hardened parts as direct substitutes, if available, in the cryocooler control electronics when radiation hardened parts are cost prohibitive.

THERMO MECHANICAL UNIT DESCRIPTION

The 95K HEC system is comprised of a Thermo-Mechanical Unit (TMU) and Command and Control Electronics (CCE). Key performance requirements are listed in Table 1. The TMU is a novel two-stage hybrid Stirling-pulse tube cryocooler comprised of separate compressor and expander modules connected by a transfer line. See Figure 1. The TMU retains significant legacy to Raytheon’s previously developed “Oxford” class machines. Compressor swept volume is 6cc produced by a pair of pistons working in opposition against a common compression volume.

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Figure 1. Protoflight 95K High Efficiency Cryocooler. The hybrid Stirling-pulse tube two-stage

expander is more efficient than single stage coolers and the unusual angled cold head offers unique integration advantages.

The two-stage expander employs a Stirling first stage and a 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 and regenerator tube and the pulse tube orifice and surge volume are thermally anchored to the first stage. The hybrid expander has several useful characteristics compared to two-stage pulse tubes and

two-stage Stirlings. Compared to a two-stage pulse tube expander with similar heat lift, the working gas volumes of the hybrid expander are significantly smaller and require lower gas flow

rate. This reduces the compressor’s swept volume and increases its pressure ratio. Compared to a two-stage Stirling expander with similar heat lift, the structural elements are much easier to construct. For example, the tight-tolerance second stage clearance seal is eliminated. Manufacturing and alignment are no more difficult than for ordinary single stage Stirling expanders.

The hybrid expander also offers noteworthy versatility. For example, the expander can be configured to angle the second stage from 0 to 90 degrees relative to the first. The unit shown in Figure 1 has a 45-degree angle between stages, which enables a redundant pair of expander cold heads to be compactly configured as shown in Figure 2. The aluminum compressor housing is nominally 338mm long and 97mm in diameter. The center section of the housing is a box-like structure that has a large area heat rejection surface on one side and an easily accessed three-point mount on the other. See Figure 1. The heat rejection area is 61.8 sq. cm (9.6 sq. inch.) This large area, combined with aluminum’s high thermal conductivity, can efficiently transport over 150W of waste heat from the module. The three-point mount on the opposite surface simplifies integration into most systems. Each of the two pistons

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Figure 2. Two hybrid two-stage cold heads angled at 45 degrees can be compactly integrated into a small sensor volume. Note that the first and second stage thermal interfaces are immediately adjacent to each other a feature that can be used to advantage in shortening thermal straps.

has a Linear Variable Differential Transducer (LVDT) position sensor. The LVDT output is fed into the electronics to prevent motion beyond a programmed “soft stop” stored in the electronics. The expander shown in Figure 1 is nominally 309mm long and 94mm in diameter. The twostage cold cylinder extends 128mm in the axial direction beyond the expander’s warm-end waste heat rejection surface. Thermal interfaces at both cold stages and at the warm end are large flat surfaces. Combined side load capacity of the two stages is a minimum of 100N in any direction. As in the compressor module, the housing is aluminum, position sensors are used to limit motion, and waste heat is efficiently removed from 21.4 sq. cm (1.66 sq. inch) of heat transfer

interface area. The protoflight compressor module is projected to weigh 4Kg and the protoflight expander, 2.5Kg for a TMU physical weight of 6.5Kg. Each module is internally balanced for optimal control of residual vibrations. The compressor employs two pistons working in opposition against a common compression chamber located at the center of the module. The expander’s first stage piston is balanced by a matching mass driven in opposition. The pulse tube second stage does not require dynamic balancing because the moving gas mass generates insignificant force. Residual vibrations produced by each module are corrected by the electronics via a novel algorithm executed by the electronics described below. The algorithm is driven by feedback from piezoelectric load washers at each

module’s mounting points. ELECTRONICS DESCRIPTION

The Command and Control Electronics (CCE) is a two-board system packaged in an aluminum “slice” housing. See Figure 3. The two boards are designated as the Power Board and the Logic Board. The Power Board includes: 1. A Low Voltage Power Supply (LVPS) with multiple DC voltage output forms 2. Spacecraft power bus voltage step-up conversion 3. Two pulse width modulated (PWM) switched compressor motor amplifiers

4. EMI filters at the board inputs and outputs

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Figure 3. Two board Command and Control Electronics. The board on the left is contains the digital elements and is called the Logic Board. The other board contains Low Voltage Power Supplies and compressor PWM amplifiers and is called the Power Board.

The LVPS generates a set of low power DC voltage forms that energize circuits within the CCE and the expander motor amplifiers. The Logic Board includes: 1. An RS422 serial port using RS232 protocols to manage spacecraft telemetry 2. A parallel RS232 serial port to enable use of a PC user interface in ground test 3. A 10KHz drive circuit to excite the LVDT piston position sensors 4. Analog signal converter circuits (e.g., a multiplexer and an ADC) 5. A set of gate arrays containing all logic needed for communication and control 6. Various digital system elements such as a ROM chip and a RAM chip. 7. Two expander motor amplifiers mounted to the Logic Board for packaging convenience 8. EMI filters on the board inputs and outputs where required The CCE is a digitally based control system that eliminates most of the conventional analog circuitry previously used in cooler electronics. This resulted in a significant reduction in parts count that increased reliability and reduced the number of required radiation hard Integrated Circuits. Total parts count is less than 550 including all resistors, capacitors, diodes, and magnetics as well as integrated circuits and FETs. Reliability at ten years is calculated to be over 97% and radiation hardness at the component level is a minimum of 200Krad. The CCE uses gate logic to implement key functions, including: 1. Telemetry uplinks (commands) and downlinks (data and status reports) 2. Cold tip temperature control on the second stage 3. Vibration control in both compressor and expander modules 4. Piston position control over the displacer and two compressor pistons 5. Launch lock enable/disable Electronics also include two separate smaller modules: an Input Ripple Filter (ERF) and a preamplifier module. The ERF attenuates reflected AC current drawn from the spacecraft power bus. The preamplifier module resides near the cryocooler to amplify low level signals from the cold tip temperature sensor and the compressor and expander vibration sensors before transmission to the CCE through potentially lengthy cables. Signal amplification reduces noise susceptibility of the signals. Two significant accomplishments demonstrated by the electronics are high power throughput efficiency to the compressor motors and a novel and effective vibration control algorithm. Power throughput efficiency is defined as the ratio of the power delivered to the cryocooler motors divided by power supplied to the electronics minus power to the LVPS. Throughput power includes power consumed by the IRF, bus voltage stepup conversion, two compressor PWM motor amplifiers, and EMI filters. This factor is closely correlated to power drawn by the compressor module, which in turn scales with amount of refrigeration produced by the cooler. Power throughput efficiency is over 90% when delivering at least 25W to the cooler motors. Power to the LVPS is approximately constant under all conditions and is designed to be about 15W. Power drawn by the LVPS includes all power required to drive circuitry within the CCE.

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Figure 4. Performance of the new vibration control algorithm. The time domain based algorithm suceeds in reducing vibrations to the noise floor.

A new vibration control algorithm was developed to work within the limited capacity of Field Programmable Gate Arrays (FPGA). This algorithm operates entirely within the time domain, which eliminates computationally intensive Fourier Transforms used by adaptive feed forward algorithms. The effectiveness of the new algorithm was tested by artificially inducing a large imbalance in a compressor module to show the algorithm converges on a low vibration

solution. The imbalance was implemented by dropping out every fifth compressor PWM power pulse in one piston assembly. This induced the large 5.8Nrms imbalance shown at 44 Hz drive

frequency in the vibration spectrum shown in Figure 4a. Vibration levels at the next several harmonics were on the order of 0.3Nrms.

Despite the induced fault, the algorithm successfully brought the compressor into balance and drove the force levels down to .030Nrms or lower over the 0-500 Hz band. See the spectrum in Figure 4b. All disturbances produced by the compressor at its drive frequency and next ten harmonics were lowered to the noise level present in the vibration feedback signal. The broadband noise is due to background environmental vibrations plus inherent circuit noise.

Correction below the background level is is believed to be impossible. Temperature control has also been demonstrated and integration of the temperature control algorithm with vibration control is currently in progress. PROGRESS SUMMARY

A breadboard TMU is in assembly as of June 2000 for test by July. Data will be used to refine the calibration of the hybrid cryocooler math model enabling optimization of the design.

The breadboard is close to a flight qualifiable unit. Once the performance of the breadboard is demonstrated the design can be readily upgraded to flight quality. A new protoflight TMU will

then be built and tested for performance and flight qualification. Most elements of the CCE circuits and logic have been demonstrated. A second set of circuit boards is currently in assembly incorporating refinements. The logic is being integrated as a complete package and will be burned into a set of FPGAs. These chips will be mounted to the new Logic Board and integrated with the protoflight TMU for system test. ACKNOWLEDGEMENTS The Air Force Space and Missile Command and BMDO sponsored this work. The Air Force Research Laboratory, Albuquerque, NM, managed the project.

Design and Test of the NIST/Lockheed Martin Miniature Pulse Tube Flight Cryocooler P. E. Bradley, R. Radebaugh

National Institute of Standards and Technology Boulder, Colorado, USA 80303 J. H. Xiao Johnson & Johnson Co. Somerville, New Jersey, USA D. R. Ladner Lockheed Martin Astronautics Co. Denver, Colorado, USA

ABSTRACT

A two-stage miniature pulse tube (PT) cryocooler, designed for a Space Shuttle flight demonstration, was built and tested at Lockheed Martin Astronautics (LMA) at Denver, CO and the NIST Boulder Laboratory. The Miniature PT Flight Cryocooler (MPTFC) was designed to provide 0.15 W of cooling at 80 K with heat rejection at 275 K. It was developed as the smallest cryocooler of its kind for the purpose of demonstrating launch survivability and thermal performance in a zero-g environment. A prototype laboratory version was first built and tested to provide information on component sizing and flow rates for comparison to numerical models. The flight version was then fabricated as a Getaway Special (GAS) Payload. Cost containment and manned flight safety constraints limited the extent of the MPTFC development to achieve performance optimization. Nonetheless, it reached 87 K driven by a commercially available tactical compressor with a swept volume of 0.75 cc. The on-orbit cooling performance was not demonstrated because of low battery voltage resulting from failed primary batteries. The first off-state PT thermal conductance measurements were successful, however, and the MPTFC also demonstrated the robustness of PT cryocoolers by surviving pro-launch vibration testing, shipping, and the launch and landing of STS-90 with no measurable performance degradation. The design and performance optimization approach for miniature two-stage PT coolers are discussed. Some factors that may limit performance in small-scale PT coolers are identified also. Laboratory pre-launch and post flight performance data of the MPTFC are presented, including cooling performance as a function of heat load and rejection temperature. Off-state conductance results are discussed in a related but separate presentation.

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The Miniature Pulse Tube Flight Cryocooler (MPTFC) flew as a NASA shuttle payload (GAS-197) in April 1998 aboard STS-90 (Shuttle Transportation System-90) as a technology demonstration experiment. The primary objectives of this experiment were to demonstrate pulse tube (PT) cryocooler performance and “off-state” thermal conductance in a micro-gravity environment and to verify launch survivability of miniature coolers having limited vibration mitigation features. This project was a collaboration between Lockheed Martin Astronautics (LMA) and the National Institute of Standards and Technology (NIST). Because of the STS schedule and the safety issues associated with manned space missions, the experiment was subject to a number of design constraints. The tight development schedule was based on limited flight opportunities preceding the construction of the International Space Station (ISS) and on a low project budget. These two constraints dictated the extensive use of commercially available components, including a tactical compressor and drive electronics (to obviate a long-life flight compressor development effort), an inexpensive 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 operating pressure, sizing for limited battery-powered operation in a cold environment, and limited design opportunity for performance optimization. The overall flight experiment design also had to address various flight hazard issues, such as mechanical and electrical integrity, EMI, redundant fusing, diode isolation, mitigation for high temperatures, etc. The experiment design had to accommodate operation over a range of STS bay temperatures from -50 to +40°C. In addition, the experiment timeline had to conform to limited STS crew operations.

The approach for completing the project on schedule was to design and test a prototype cryocooler in parallel with the overall flight hardware system definition and parts procurement. Subsequently, the flight hardware and flight cryocooler development and testing were also accomplished as parallel efforts. Lockheed Martin had primary responsibility for the flight and GSE hardware and electronics, systems engineering, and for payload management, while NIST had primary responsibility for the cryocooler development, assembly, and testing. In practice each organization

contributed to all of these tasks. DESIGN Coldheads The cryocooler coldhead design selected was a two-stage U-tube geometry orifice pulse tube

(PT) system based upon the double inlet concept first introduced by Zhu, Wu, and Chen.1 The system is schematically represented in Figure 1. This two-stage approach was arrived at based on the design goal of reaching 80K and the miniaturization requirement in which the compressor and coldhead are separate components. The compressor and coldhead were separated to reduce vibration at the coldhead 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 (approximately 10%) of the gas to travel directly from the compressor to the warm end of the pulse tube. This small flow bypasses the regenerator and then compresses and expands the gas that remains at the warm end of the pulse tube. This reduces the flow through the regenerator thus reducing the regenerator loss accordingly. 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 REGEN3.1 developed by NIST2,3,4 and a thermoacoustic model developed by Xiao5,6,7,8 were employed to design both the prototype and flight coldheads. The prototype cooler is shown in Figure 2. The test fixture for the prototype coldhead allowed the flow division between the primary and secondary orifices to be adjusted during operation using external metering valves, which were modified to minimize void volume. The approximate flow rates were easily determined for the primary and secondary orifices based on the metering valve settings. However, the intermediate flow path distribution between the first and second stages

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Figure 1. Schematic of double inlet configuration. Figure 2. NIST/LMA pulse tube prototype.

required manual adjustment. Therefore partial disassembly of the coldhead was necessary during several iterations to optimize the intermediate orifice. Calculations provided the approximate final flow impedance for this orifice. The fabrication, characterization, and best thermal performance were accomplished within only 3 months of the initial design phase. The prototype unit used a larger 2.5 cc laboratory compressor operated at reduced power to simulate the expected power of the smaller 0.75 cc flight unit. Instrumentation of the prototype coldhead provided temperatures, heat loading, and pressure data for evaluating pressure ratios and phase angles. The relatively long transfer line and small but unavoidable void volumes in the valves limited the efficiency of the prototype coldhead, but its low temperature performance proved the design feasibility of such a small system. A low temperature of 84 K with a pressure ratio of about 1.23 was achieved. For a pressure ratio of 1.26 however, 76 K was achieved. When the 0.75 cc flight compressor was attached to the optimized prototype coldhead with its attendant void volume a temperature of only 127 K was achieved. This resulted from the much lower pressure ratio of 1.13, indicating that the PV work was lower with the flight compressor. A pressure ratio of 1.2 to 1.25 was the design value for the MPTFC. Steps were taken in the fabrication of the flight coldhead to minimize any void volume in the system in order to deliver the PV work associated with the design pressure ratio. The flight cooler (MPTFC) coldhead shown in Figure 3 has a PT volume of nearly 0.54 cc (see Table 1 for other important coldhead dimensions). The figure provides an exploded view of the two stages but several components are not shown. The cold end and aftercooler are made using

OFHC copper; the regenerator and pulse tubes are thin wall 304 stainless steel. The reservoir, which is also made from 304 stainless steel, has a bracket for attaching the flight pressure transducer. Two smaller diagnostic pressure transducers communicate with the compressor and primary orifice spaces of the coldhead. The transfer line is shown prior to final flight modification,

Figure 3. NIST/LMA MPTFC Coldhead.

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and the pressure relief capillary is also visible. The primary and secondary orifices (not shown) are integrated into the aftercooler. Clamps and seal rings to connect the two stages and the fixed

radiation shield / MLI are not shown. Also not shown are two thin plates that provide impedances to form the secondary and intermediate orifices. Many other flight-related elements including other instrumentation such as temperature sensors and a film heater on the cold end are also omitted in this figure. The MPTFC schematic is shown in Figure 4. Some important features of the MPTFC design are 1) a very compact physical arrangement; 2) the aftercooler and the compressor are conductively cooled through contact to the experiment mounting plate (EMP); 3) nylon displacement stops are located at both the coldhead inter-stage and cold end. The development and test phases for the flight coldhead required about 11 months to complete concluding with final preparations for vibration testing just prior to integration.

Coldhead Instrumentation The MPTFC was extensively instrumented for temperature, pressure, and vibration measurements. In fact there were double and triple redundancies built into the system. All of the sensors employed were commercial off the shelf (COTS) items. Specifically, thin film platinum RTDs, piezoresistive and piezoelectric pressure transducers, thin film heaters, and tri-axial accelerometers were used. Schedule constraints required two diagnostic piezoresistive-pressure transducers to be epoxied in place for flight to mitigate the risk of GHe leakage. The electronics system was designed and built in-house and made extensive use of COTS hardware. The DAS was configured at a 5 minute scan rate in a “fill and hold” mode.

G-197 Design Features Figure 5 shows the G-197 Payload minus the GAS canister enclosure. This assembly can be referred to as a cantilevered frame support. The upper third of G-197 consists of the MPTFC experiment itself, a commercial latching vacuum valve to expose the experiment to space vacuum,

Figure 4. Exploded schematic of the flight cooler arrangement.

Figure 5. TheG-197 payload minus GAS canister.

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and other electrical and battery venting connecting hardware. The lower two thirds (the “battery box”) contains the battery modules and the flight electronics module — the DAS, computer, signal conditioners, and drive electronics for the experiment. This entire assembly, i.e., G-197, integrates with the NASA GAS canister, which is a 5 cubic ft. cylindrical containment vessel that is capped at both ends and evacuated during integration. There are four supports at the cantilevered end, known as bumpers, that lock the bottom of G-197 radially into the canister by preloading against the canister walls. This must be done in order to restrict the movement of the lower end of the battery box during launch and landing, thereby reducing the bending loads applied at the opposite end where the experiment attaches to the EMP with only 12 Ti alloy screws. The completed GAS canister then encloses the experiment with all electronic connections accomplished via feedthroughs at the canister bottom plate which holds the NASA GAS electronics. The bottom plate is the electronic interface between G-197, the GAS relay system, and the shuttle GAS computer which is operated by astronauts. Special features of the G-197 GAS payload are: 1) the EMP is uninsulated and coated with silverized Teflon tape for maximum heat rejection to a space environment, 2) the GAS canister is evacuated to simulate a flight instrument environment, 3) The primary battery system consisted of thirty-three 2V batteries in a 3 string redundant arrangement to power the MPTFC and its compressor drive electronics at a nominal 22V. The secondary battery system consisted of four 4V batteries in series to maintain a nominal 16V supply to power the EM valve, DAS, and computer electronics. All batteries are polyurethane foamed into the BB and are vented to EMP relief valves, 4) “low voltage” and “high temperature” cutoff circuits are employed (no longer required by NASA), and 5) battery voltages and GAS canister temperatures are measured for the uninsulated EMP for comparison to NASA numerical models developed in 1987. Flight Safety Features The MPTFC design employed several voluntary safety features as well as those required by NASA to safeguard the STS and astronauts during flight. Although not all of the required safety features were practical for mission success, they had to be accommodated in order to fly MPTFC on the shuttle. NASA vacillated on certain requirements, but the following were final. Although the sealed lead acid primary batteries were vented external to the GAS canister to safeguard against the buildup of explosive gases, both low voltage and high temperature cutoff circuitry were also required. Proof pressure testing was required for the MPTFC vacuum housing to more than ten times the maximum pressure that would exist if the cooler developed a leak or the compressor pressure exceeded the limits of its housing. Thermal cutoff switches were employed in the drive electronics to prevent an overheat condition within the compressor. Furthermore, a pressure relief mode, consisting of a capillary that was sealed using an indium soldered cap connected to the warmest location on the MPTFC, was required in case the MPTFC overheated. The compressor drive power was set conservatively to eliminate excessive initial vibration. Polyswitch fusing and redundant wiring were also required to complete the electronics package. The 3.6 V Li cell used to retain memory was double-diode isolated and fused. PREFLIGHT TESTS All optimizations of both the prototype and the flight cooler (MPTFC) systems were conducted in the laboratory in a bench test environment. All instrumentation and drive electronics were installed in the best configuration for optimization and therefore were not configured (i.e., wired or attached) for flight. This meant that upon completion of optimization the sensors were removed and reworked for the flight configuration, including rewiring, reattachment, and functional tests. Experiment Functional Testing After reworking the MPTFC for the flight environment and integration with the flight electronics and the flight support structure, a complete functional verification of all sensors and a performance verification of the G-197 system were made. This of course included a final pump-out and

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Figure 6. Cool down for flight system rejecting to 258 K. gas helium pressure charge of the cooler system to 2.6 MPa. These tests included verifying that

minimum temperatures and maximum refrigeration loads were maintained, as well as the cooldown performance. Figure 6 shows a representative cool-down for the flight system as measured during thermal vacuum testing conducted at the LMA facility. Figure 7 shows the MPTFC performance for a nominal rejection temperature of 273 K. This data was consistent with data measured

before the flight configuration and integration with the DAS and flight electronics rework. EMI Testing Although G-197 was a GAS payload, which is considered to be a very low risk to the shuttle operations / communications when fully sealed, NASA required that the radiated EMI of the payload be measured. This requirement for G-197 was due to the vacuum line, which runs from the

cooler housing via the EM valve to an EMP port for the on-orbit evacuation of the housing. The

Figure 7. Nominal cooler performance for a 273 K rejection temperature.

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MPTFC and flight electronics subsystem were tested at the LMA facilities. Testing included low

frequency to GHz radiated emissions for both narrow and broad band. The G-197 radiated EMI easily passed within the acceptable limits imposed by NASA.

Vibration Testing Complete vibration testing was required for the G-197 payload because it slightly exceeded the nominal limit of 91 kg (200 lb.). Both a low-level sine sweep and random vibration at qualification

and protoflight acceptance levels were performed in lateral and axial modes. These tests were accomplished at the NASA/ARC. It is of note to the reader that these vibration levels are rather severe and often are conservative compared to the actual launch environment. They represent a safety concern by NASA for the structural integrity of the system. However, NASA requires this conservative level of testing because of the uncertainty of both location of the payload and flight vibration loads within the STS bay. G-197 was calculated to have a lowest lateral mode resonance of 58 Hz when installed in the GAS flight canister. A measured resonant frequency of 40 Hz was obtained using the shipping canister, still above the required 35 Hz minimum. An unexpected result of this test was that some of the Pb-acid batteries failed in the axial mode. This was recognized by a decrease in the primary string voltage from a nominal 22 V to ~20 V. However, due to schedule constraints, coupled with the knowledge that the MPTFC experiment can be operated effectively at 16 V, no changes or modification to the batteries were undertaken. It was believed that there was sufficient margin to continue with the scheduled integration and launch, anticipating that there would be no further degradation. INTEGRATION AT NASA/KSC

Upon completion of the vibration testing and subsequent pre-integration functional checkouts, the G-197 payload was packaged and shipped to NASA Kennedy Space Center (KSC) for integration with the GAS flight electronics and flight canister. The integration is typically a threeday process that takes place about 3 months prior to launch. Upon completion and sign-off of the integration, there is no opportunity for further contact with the payload by the experiment investigators. Therefore the MPTFC experiment had to be able to retain its 2.6 MPa gas helium charge and an adequate battery charge over a three-month period. This requirement and STS safety considerations effectively eliminate many types of batteries used for unmanned missions.

The integration process is a very challenging exercise for a complicated powered experiment such as MPTFC. First the payload is unloaded from the shipping canister, which is quite similar to

the GAS flight canister except that it has no NASA electronics and acts only as a protection vessel for shipping. However, because the support bumpers must be positioned to safeguard the experiment during transit, disassembly is required at NASA/KSC. Next, a very thorough visual safety inspection of the battery box (BB), electronics / fuse box, and experiment itself is conducted by NASA/KSC personnel to ensure that the payload is in full compliance with all safety paperwork. After the safety inspection is signed off, the real work to prepare the experiment for flight begins. The battery systems must be top-charged and the BB and vent plumbing subjected to a pressure proof test prior to a GN2 purge. The experiment housing must be evacuated and valved off before a preliminary system functional check is made. After a final visual inspection, all fasteners are secured for flight which involves epoxying and/or lock-wiring external screws and securing wiring / cables. A weigh-in of the completed experiment verifies that it is within approved limits before NASA will proceed. Finally, the payload is installed into the flight canister, the bumpers are preloaded for flight, and the end plate with NASA electronics / interface cable is integrated to the payload. A final functional test is then made to ensure that the NASA electronics and payload are compatible. This test also serves to verify EM valve operation, MPTFC cooling, and adequate battery margin. The EM valve is reset to a closed position and a removable manual valve (which maintains a guard vacuum in the line connecting the EM valve to the EMP port) is closed and tagged for removal prior to flight. The GAS canister is evacuated and the payload is officially handed off to NASA/KSC for STS integration.

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FLIGHT DATA Minimal Cooling

The cooling data collected during the STS-90 mission was unfortunately of minimal content. After initial confirmation of primary power, the first data set consisted of 20 hrs, powered mostly by the secondary system, which operates the coldhead heater and the DAS in the absence of primary power. Attempts to operate the cooler on-orbit were apparently foiled by a fluctuating primary voltage, which dipped below 16 V causing a designed malfunction indication at the GAS control computer. Specifically, this led to periodic cutouts of relay ‘A’, eliminating power to the MPTFC compressor drive electronics. This relay supposedly resets automatically at 40 minute intervals by a NASA timer, but there was no indication of this operation in the data set. The relay was switched off and reset manually later in the mission, again with positive initial indication but little evidence of cooling in the data. This was unfortunate since the MPTFC can operate safely

below 16 V, thereby making the NASA-mandated cutout level for battery protection unnecessary. It should be noted that NASA has recently eliminated malfunction circuitry requirements for GAS

payloads. Measures to perform a re-flight of this experiment (as G-785) to demonstrate the on-orbit cooling have already been performed. The robustness of the MPTFC and its survivability without performance degradation have already been proven. “Off-state” Thermal Conductance Test The on-orbit data collected from the second data set of 15 hrs included continuous cold stage heater operation. This data comprises the on-orbit conductanc test for a pulse tube cooler in an offstate passive mode. To date there has been little if any published data of this type. In fact, less than a handful of PT coolers have been operated in space. This conductance data and its analysis are presented in a related paper at this conference.9 POST-FLIGHT TESTS

Battery testing One month after STS-90 touched down at KSC the G-197 payload was de-integrated and delivered to LMA personnel. A post-flight functional test determined that the primary battery system exhibited random voltage fluctuations that varied from ~15 to 21V. Later it was determined that at least 7 of the 33 cells were either intermittent or completely defective. At least two cells on each redundant string were affected. This condition explained the relay cutouts during flight. The low voltage condition was attributable to the vibration tests, the launch, and the -6°C STS bay temperature during initial MPTFC activation.

Accelerometer data and model results Analysis of post-flight data from a miniature tri-axial accelerometer mounted at the reservoir of the cooler during operation was evaluated using a finite element model. It predicts a lateral displacement of 1.14 at 164 Hz for the as-built u-tube configuration. However, if a third thin wall support member were used to stiffen each stage, the model predicts the displacement would decrease to 0.21 at a frequency of 391 Hz. This result indicates improvement may be made to further reduce induced vibration at the cold tip for vibration sensitive sensor packages. Cold environment testing Subsequent to post-flight battery refurbishment (see below), cold environment testing was performed at rejection temperatures below the nominal 273 K. The performance as a function of the heat rejection temperature was conducted and the results are shown in Figure 8. The heat

rejection temperature affects both stages consistently. The first stage temperature ranges from 161 K to 204 K for both a no load and a 45 mW load at the second stage based upon a rejection temperature ranging from 244 K to 300 K. For the same rejection temperature range the second

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Figure 8. MPTFC cooler performance as a function of heat rejection temperature.

stage temperature ranges from 82 K to 105 K for no load, while it ranges from 90 K to 121 K with a 45 mW load. The cold testing also served to confirm the nominal operation of heaters located on the flight electronics and compressor. PERFORMANCE DATA

This system as currently configured is certainly not very efficient, having a Carnot efficiency of only 2%. However, because of its very small size it represents an important step toward miniaturizing PT systems. As most readers of this paper are aware, extensive experimental optimization is required to achieve the best performance for PT cryocoolers after thorough modeling has been

completed. Often the first iteration in the system design based on the numerical model falls short of the intended goal. This is especially true when pushing the predicting capabilities of the model into a new scaling arena, as for the MPTFC. Although the schedule allowed for only one design iteration, much has been learned about scaling miniature PT coolers, e.g., the complexities associated with the flow distribution and optimization for a two-stage multi-inlet design. For best efficiency the importance of minimizing parasitic heat leaks can not be overlooked. Immediately after flight the MPTFC demonstrated no reduction in performance and there was no detectable change in the system pressure. In fact, it has held pressure for 2.5 yrs with only a 5% loss, but even this loss has degraded thermal performance from a no-load temperature of 111 K to 123 K at ambient rejection. A loss of only 5 % in the pressure nonetheless represents significant seal performance for a system designed for modification flexibility. RE-FLIGHT

At present a re-flight of the MPTFC experiment is in progress as G-785. A re-flight provides a potential opportunity for MPTFC improvements, depending on the manifest date set by NASA. A new battery system is always required for a GAS re-flight. Failure of the NASA-recommended primary system batteries used in G-197 resulted in the selection of new batteries of the successful secondary system type, but even these batteries revealed a sourcing issue. Specifically, while the original 4V batteries and their new replacements were of the same model from the same vendor, they were not from the same factory. A different internal design in the newer version resulted in failure during sample vibration tests; G-785 uses older version cells. Other improvements include a high resolution voltmeter, improved frequency diagnostics, and an increased DAS scan rate. All paperwork and refurbishment are complete for a re-flight and we are anticipating an opportunity in early 2001. If it becomes necessary to re-pressurize the MPTFC to 2.6 MPa, more diagnostic sensors and some radiation baffles will be installed.

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FUTURE MPTFC COLDHEAD DESIGN AND TEST

Future efforts in MPTFC research will concentrate on improving the current 2-stage coldhead, including evaluation of DC flow, further reduction of the parasitic heat loads9, modification of

stage geometry, and performance evaluation at higher operating pressures. Furthermore, a new one-stage design effort will be undertaken for comparison with the two-stage approach. Extending existing numerical models with innovations will streamline the design and test of miniature systems, thereby advancing their predicting capabilities and accuracy.

ACKNOWLEDGMENTS

We especially wish to thank Dr. Joe Martin for financial and program support, Mr. Don Hirschfield for assistance during integration at NASA/KSC, Mr. Ken Sell for intricate machining, and Mr. Aleks Bakman for vibration analysis, all of LMA; Dr. Peter Kittel of NASA/ARC for vibration testing; Mr. Michael Lewis of NIST for technical assistance during MPTFC testing; and

Mr. Mark Wallace of Campbell Scientific for software assistance. REFERENCES

1. S. Zhu, P. Wu, and Z. Chen, “Double inlet pulse tube refrigerators: An important improvement,” Cryogenics 30, 514 (1990). 2. 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). 3. Gary, J., Daney, D.E., and Radebaugh, R., “A computational model for a regenerator,” Proc. Third Cryocooler Conference, NIST Special Publication 698, (1985), p. 199.

4. Gary, J. and Radebaugh, R., “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. 5. J. H. Xiao, “Thermoacoustic theory for regenerative cryocoolers: A case study for a pulse tube refrigerator,” Proc. 7th International Cryocooler Conference, Air Force Report PL-CP-93-1001, Kirtland

6.

7. 8.

9.

Air Force Base. NM (1993), p. 305. J. H. Xiao, “Thermoacoustic heat transportation and energy transformation, Part 1: Formulation of the problem,” Cryogenics 35, 15 (1995). J. H. Xiao, “Thermoacoustic heat transportation and energy transformation. Part 2: Isothermal wall thermoacoustic effects,” Cryogenics 35, 21 (1995). J. H. Xiao, “Thermoacoustic heat transportation and energy transformation, Part 3: Adiabatic wall thermoacoustic effects,” Cryogenics 35, 27 (1995). D. R. Ladner, P. Bradley, and R. Radebaugh, “Offstate Conductance Measurements of the NIST/ Lockheed Martin Miniature Pulse Tube Flight Cryocooler: Laboratory vs. Space,” Cryocoolers 11, Plenum Publishers, NY (2001).

Low-Cost Pulse Tube Liquefier for In-Situ Resource Utilization C.M. Martin and J.L. Martin Mesoscopic Devices, LLC Broomfield, Colorado, USA 80020 ABSTRACT

NASA’s strategy for continued exploration of Mars is based on the concept of using Martian resources to supplement materials brought from earth. This in-situ resource utilization (ISRU) program allows dramatic reduction on the mass of materials that must be transported from the Earth, and is an enabling technology for a manned mission to Mars. A key part of the ISRU strategy is to use the Martian atmosphere along with hydrogen feedstock and chemical reactors brought from earth to produce oxygen and a hydrocarbon as rocket fuel for the return trip to Earth. Any oxygen produced on Mars will need to be stored for months to years as sufficient reserve is built up. The overall weight of the storage system is lower for liquid oxygen than pressurized gas, so a liquefier is required. We are developing a low-cost oxygen liquefier for insitu resource utilization. The design point for this cooler is 20 W at 89 K rejecting to 245 K. The liquefier uses a pulse tube cryocooler, with a linear, opposed pressure wave generator. The pulse tube is being designed for compactness and ease of integration with the balance of the system. The first generation cryocooler has been built and is currently being tested. Performance predictions for this cryocooler are presented. INTRODUCTION

Exploration of Mars will require utilization of indigenous resources to support human life and operations.1 One component of ISRU is the manufacture and storage of rocket propellants. Storage of these propellants can be accomplished for the lowest mass using cryogenic liquefaction and storage. Under support from NASA/Johnson Space Center through the Small Business Innovative Research program, Mesoscopic Devices is developing a pulse tube cryocooler to liquefy oxygen

and propellant in support of ISRU. This cryocooler must be low-weight, capable of operating over a large range of ambient temperatures, and scalable to large powers to support human missions. We are one year into a two year effort and have completed the fabrication of the first

generation pulse tube liquefier. The design of this cryocooler is described below. IN-SITU RESOURCE UTILIZATION In-situ resource utilization involves using the indigenous resources on other solar bodies to support their exploration. In the case of Mars, in-situ consumable production (ISCP) has the Cryocoolers 11, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2001

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potential to significantly reduce the cost and risk of robotic and human exploration.1 This would involve sending feedstock, such as liquid hydrogen, to Mars along with a chemical processing plant. The chemical processing plant would use the carbon dioxide from the Martian atmosphere

and the hydrogen brought from earth to generate rocket propellants, such as oxygen, methane, methanol or other hydrocarbons. In addition, oxygen can be produced for crew life support. The initial mission mass required in low-Earth orbit can be reduced by 20% by using ISCP. ISCP requires three components: resource collection, chemical processing, and liquefaction and storage. NASA is pursuing all three areas concurrently. The work presented in this paper addresses the liquefaction issue. A liquefier is required to cryogenically liquefy the oxygen, and probably the hydrocarbon, for storage. The requirements for the liquefier are still evolving, but there are several key features that need to be addressed in all potential designs. First, minimum weight is essential. This weight must account for the cryocooler, as well as the storage dewar, power source and heat rejection

system.2 Unlike terrestrial applications, heat rejection will have a significant impact on the overall weight and must be considered at the earliest stages of the design. Improved cryocooler

efficiency may increase the weight of the cooler slightly, but it may be more than compensated for by reducing the weight of the heat rejection system. Other key factors include the operating duty cycle and the cooling loads and temperatures required. These have been discussed previously and are an area that continues to evolve.3 The one thing that is certain is that initial missions, such as rock sample return missions, will have much lower cooling requirements than later, manned missions. For this reason, a cryocooler design that can be effectively scaled in size will provide a reduction in the risk for human missions by demonstrating key technology in smaller, earlier missions.

CRYOCOOLER DESIGN

We selected a coaxial pulse tube cryocooler driven by a linear, opposed pressure wave generator. The use of a coaxial cold head simplifies integration with the storage dewar, as the cold head can be inserted directly into the dewar neck. An annular regenerator surrounds the pulse tube. A condenser, to provide additional condensation surface area, can be bolted to the cold tip before insertion into the dewar. Figure 1 shows the cryocooler and Figure 2 shows the coaxial pulse tube. Table 1 shows the

key design parameters and overall dimensions. The liquefier uses an inertance tube to generate the required phase shift. Due to the relatively high operating frequency (60Hz), phase differences of up to 52 degrees between the volumetric flow rate and the pressure can be generated at the entrance to the inertance tube. This allows for near-ideal phase relationships at the cold end and regenerator. Another key advantage of the inertance tube design is the elimination of DC flows through the pulse tube. The baseline inertance tube is 4 mm × 2.4 m long. For testing, this tube is routed out the side of the inlet cap to an external reservoir. This arrangement facilitates optimization of the inertance tube

parameters. In the final configuration, the inertance tube is coiled in an annular volume inside the compressor surrounding the pistons. A 0.5 liter reservoir volume is provided inside the

pressure vessel, separate from the compressor back side. The pulse tube is tapered to suppress acoustic streaming. We used the methodology of Swift to calculate the optimal taper angle.4 The required angle is very small, only 1.6 degrees (full

cone angle) for this design. The aftercooler is liquid-cooled, with 100 mesh copper screens on the helium side and circumferential fins and grooves on the coolant side. For laboratory testing, we circulate a mixture of ethylene glycol and water through the 6 mm high by 2 mm wide passages to reject the heat from the cycle. For a Mars application, the aftercooler would be coupled to a radiator using

either a pumped single-phase loop or a two-phase thermosyphon or heat pipe. Design of the Mars heat rejection system will be a significant task in the second year of our program.

Mesoscopic Devices worked closely with CFIC of Troy, NY to develop the compressor for the liquefier. The compressor uses patented STAR motors and bent strap flexure bearings.5

LOW-COST PT LIQUEFIER FOR IN-SITU RESOURCE UTILIZATION

Figure 1. Oxygen liquefier.

Figure 2. Cold head, shown 180° from normal operating orientation (cold tip is up here but would be down in operation).

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STAR motors are a type of moving magnet design that uses an interdigitated stator-armature combination and separate bobbin-wound coils. The 100 W STAR motors used for this compressor represent the smallest STAR motors built to date. Motors of this type have already

been built at 100 W, 300 W, 2 kW and 10 kW. Linear compressors using this motor and flexure design can be built from 200 W to 20 kW of input power, spanning two orders of magnitude in cooling power. This provides a clear growth path from cryocoolers sized for robotic sample return missions, which will require a few kg/day of propellant, to manned missions that might require several hundred kg/day for breathing, exploration and propellant uses. The compressor design parameters are also included in Table 1. Performance Predictions

Figure 3 shows the predicted load curves for this machine. We used, DELTAE, a thermoacoustic code, to generate the predicted load curves. The two curves are for reject temperatures of 245 and 293 K, representing the design operating conditions in a Martian environment, and typical laboratory conditions. The curves indicate the maximum load, limited

by either the compressor input power (200 W maximum) or the compressor stroke limit (12 mm).

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Figure 3. Predicted load curves for reject temperatures of 245 and 293 K. SUMMARY

The Mars ISRU application requires a cryocooler that is efficient, low weight and can be scaled to larger cooling loads to support manned missions. The pulse tube cryocooler described here will meet these requirements. The same basic design can be scaled up to 1 kW of input power, and an in-line version can be scaled up to 20 kW (we are currently building 4 and 20 kW liquefiers using this same design for terrestrial applications). ACKNOWLEDGEMENTS

This work was supported by NASA/JSC through the Small Business Innovative Research program under contract NAS9 99081. REFERENCES 1. NASA Technology Plan, http://actuva-www.larc.nasa.gov/techplan (1999). 2.

J.L. Martin, et. al., “Low-Cost, High-Performance Cryocoolers for In-Situ Propellant Production”, In-Situ Resource Utilization (ISRU-III) Technical Interchange Meeting, Denver, CO, February 1999.

3. P. Kittel, L.J. Salerno, and D.W. Plachta, “Cryocoolers for Human and Robotic Missions to Mars”, Cryocoolers 10, Kluwer Academic/Plenum Publishers, New York (1999), pp. 815-821. 4.

G.W. Swift, M.S. Alien, and J.J. Wollan, “Performance of a Tapered Pulse Tube”, Cryocoolers 10, Kluwer Academic/Plenum Publishers, New York (1999), pp. 315-320.

5.

US Patent 5389844, Linear Electrodynamic Machine, February 14, 1995.

Performance Characteristics of a 4 K Pulse Tube in Current Applications C. Wang and P. E. Gifford Cryomech, Inc. Syracuse, NY 13211, USA

ABSTRACT The Cryomech Model PT405 has cryo-cooled several applications during the past year where both its 4 K temperature and pulse tube characteristics were required. The specifications for the PT405 have improved since it was introduced. Presently the typical cooling capacity is 0.6 W at 4.2 K on the second stage with simultaneously 30 W at 65 K on the first stage with an input power of 4.9 kW. With the new CP900 Series compressor packages, the PT405 maintains the same performance for both 60 Hz and 50 Hz power supplies. The force transmitted into a cryostat by the PT405 base plate is less than 3.6 N. The same test registers a force of 178 N for a 4 K GM Cryorefrigerator. The maximum displacement (elongation of the base tube assembly due to the cycling of the high and low pressures) at the 2nd stage heat exchanger is Some applications of the PT405 are presented in the paper; such as, operating a low vibration cryostat, conductively cooling a superconducting magnet, pre-cooling an ADR X-ray detector and condensing helium. INTRODUCTION Since the discovery of the pulse tube cryorefrigerator, the cryogenic community has been looking forward to benefiting from its no-displacer design. The pulse tube cryorefrigerator promises to improve reliability, to increase the meantime between maintenance, to extend system lifetime, and to decrease the cryorefrigerator cost in comparison with Gifford-McMahon (GM) and Stirling Cycles. Yet, to many, the most exciting benefit of the absence of the displacers in the PTR was the reduction of “g” forces exerted on the cryostat and lack of magnet gradient disturbance caused as the displacer moved up and down in the GMs. Over the past few years, significant progress in the design of pulse tube cryorefrigerators has led to an increase in its efficiency in comparison with conventional cryorefrigerators. There are several types of pulse tubes commercially available, such as the TRW “Stirling type” space pulse tube1 and IWATANI “GM type” pulse tube, etc. In the summer of 1999, Cryomech introduced a 4 K Pulse Tube, the Model PT405. Although others had introduced single stage PTRs, the PT405’s 4 K performance provided the first opportunity to test the PTR in several of the most challenging cryo-cooling applications. The performance of PT405 has been improved by recent modifications and was characterized for applications. Some of the current applications will be discussed in this paper. Cryocoolers 11, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers. 2001

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PT405 COLD HEAD

PT405 is a two-stage, 4 K pulse tube cryorefrigerator that includes a pulse tube cold head, connecting flexible lines and a helium compressor package. The pulse tube cold head employs a doubleinlet configuration. Figure 1 shows a photo of the PT405 cold head. A new compressor, Cryomech Model CP950, was specially designed for the pulse tube with regard to a long meantime between maintenance (MTBM). The compressor package supplies the cold head with pressurized helium through flexible metal hoses. A rotary valve in the cold head, similar to that in a GM cryorefrigerator, directs the helium gas in and out of the pulse tube system. The 1st cooling stage of the refrigerator can provide cooling power between 35 K to 80 K, and the nd 2 cooling stage below 4 K. Owing to some advanced designs in the whole system, the MTBM of the PT405 is expected to be > 20,000 hours. PERFORMANCE CHARACTERISTICS OF PT405 Cool down performance

To maintain the standard specifications for both the 50 Hz and 60 Hz electrical power supplies, the PT405 is supplied with different compressor packages: the CP950 for 60 Hz, and the CP970 for 50 Hz. The input power in both cases is 4.9 kW with the 1st stage operating at 65 K and the 2nd stage at 4.2 K. A typical cooling-load map for the PT405 is given in Figure 2; for example, it shows that the

cooler provides 0.6 W at 4.2 K on the second stage, simultaneously with 30 W at 65 K on the first stage. The performance of the PT405 has been improved upward from the initial announced capacities of 0.57 W at 4.2 K and 18 W at 65 K.2

Figure 2. Typical cooling-load map of PT405.

Figure 1. Photo of PT405 cold head.

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Figure 3. Cool down curves for PT405 with and without mass attached.

The cool down time with a thermal mass attached for the PT405 was tested with 6.4 kg of OFHC copper on the 1st stage and 0.9 kg of OFHC copper on the 2nd stage. Figure 3 shows the

cool down times of the PT405 with and without the added mass. It takes 60 minutes for a unloaded PT405 to reach 4 K and 110 minutes for both stages to reach their minimum

temperatures of 2.6 K and 35 K. With the above thermal loads attached to both stages, the 2nd stage reaches 4 K in 100 minutes, and the two stages reach their minimum temperatures in 170 minutes.

Vibration of PT405

Cryomech, Inc. has collaborated with several different groups to analyze the vibrations generated by the PT405. Figure 4 shows the test rig used by GE R&D Center to analyze the mounting forces exerted on the cryostat at the PT405 base plate. A load cell was mounted under the base plate of cold head. The displacement of the 2nd stage cold heat exchanger was measured with both an optical comparator and an accelerometer. The accelerometer was mounted on the bottom of the 2nd stage cold heat exchanger as shown in Figure 4. The optical measurement was made with a ST405 Cryostat (Figure 5). A laser beam penetrated through the optical windows in the cryostat and focused on the second stage heat exchanger, measuring the movement. The mounting force of PT405 compared with that of a mechanical driven GM

cryorefrigerator is shown in Figure 6. The mounting force of PT405 is less than 3.6 N (0.8 lb) and for the GM 178 N (40 lb). The mounting force of pulse tube is only generated by the pressurization and depressurization of helium in the pulse tube assembly. The curve is similar in its wave shape to the dynamic gas pressures. Table 1 gives the displacements of the 2nd stage cold head measured by both the optical and the accelerometer methods. The vectors of the three-axis are given in Figure 4. The two tests measured similar displacements. The maximum displacement around is in the vertical direction.

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Figure 4. Vibration test rig .

Figure 5. Photo of ST405 cryostat.

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Figure 6. Mounting forces of pulse tube and GM cryorefrigerators.

Effects of a Magnetic Field or Orientation on the Cooling Performance

The effects of a magnetic field on the operation of the PT405 have been tested. Due to the limitation of resources, we have not tested the PT405 in a large enough magnetic field that has degraded its performance. We have designed the second stage regenerator with antiferromagnetic materials to minimize the loss of heat capacity in field. In one test we moved a PT405 Cold Head while operating in a cryostat at minimum temperature toward a superconducting magnet. In this test, a field strength of 900 gauss had no affect on the rotary valve motor or bottom temperature. Higher magnetic field will stop the motor. According to the published information of the rare earth materials, it is expected that the low temperature

performance of the 2nd stage will be slightly decreased in the magnetic field higher than 1 Tesla. It is suggested for maximum performance that the pulse tube cold head operate as close to vertical as possible. “Gas mixing losses” in the pulse tubes caused by gravity decrease the performance of both stages when the pulse tube cold head is tilted off the vertical position2. However, mounting angels up to 30° off vertical will not degrade the performance greatly.

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SOME APPLICATIONS OF PT405

Low vibration cryostat

A 3 K cryogen-free cryostat with optical access was built at Cryomech, Figure 5. The Model ST405 has very low vibration, as can be seen in Table 1. A radiation shield is fixed on the 1st cooling stage to reduce the radiation loss on the 2nd cooling stage and sample holder. It is designed for two and three axis optical measurements. There are 5 holes with diameter of 12.7 mm through the radiation shield and 5 optical windows of diameter of 25.4 mm on the optical cube. The radiation losses are crucial for the sample temperatures below 4 K. The bottom temperature on the sample holder in the cryostat is 2.6 K, when all holes through the radiation shield are covered with aluminum tape; 3.0 K with one hole uncovered; and 3.5 K with two holes uncovered. Small helium-4 liquefier or re-condenser

A small helium liquefier was developed at Cryomech using the PT405. In the system room temperature helium gas is pre-cooled by the 1st stage and 2nd stage regenerator. Figure 7 shows a schematic of the liquefier. The pre-cooling of the helium from the first stage temperature down to approximately 6.5 K using the inefficiency of the 2nd stage regenerator is critical for increasing the liquefaction rate. Numerical analysis made by the present authors calculates that the precooling heat load on the 2nd stage regenerator, decreases the PT405 2nd stage cooling capacity by only 10% of the heat actually absorbed into the regenerator. Figure 8 is a curve of the rate of liquefaction based on the liquid helium level in our condensation chamber. The accuracy of the level instrument was mm, and the liquid level sensor indicated that liquid helium level increased steadily. The pulse tube liquefier condensed 1.0 liters of liquid helium at 4.2 K in five hours, for a corresponding liquefaction rate of 4.8 liters per day.

Cryogen-free operation of superconducting magnet

The first commercial cryogen-free superconducting magnet system cooled by the PT405 was developed by Cryomagnetics, Inc. Figure 9 is a photo of the magnet system. The system has a horizontal room temperature bore diameter of 32 mm. The first stage of PT405 is used for cooling intermediate thermal shield and Bi-2223 based HTS current leads. The 2nd stage conductively cooled NbTi-based superconducting magnet below 4 K. The complete system takes approximately 14 hours to cool down the system to the operating temperatures: 50 K for

the thermal shield and <4 K for the magnet. The magnet generates a magnetic field up to 9 T with 0.1% homogeneity over a 10mm diameter volume. After an intentional quench at 9.2 T, the system recovered in temperature in 3 hours.

Figure 7. Schematic of pulse tube helium liquefier.

Figure 8. Liquefaction rate of the PT Liquefier.

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Figure 9. Photo of conductively cooling a

Figure 10. Photo of precooling an ADR for SJT

NbTi-base SC magnet using PT405

Detector

Precooling an ADR for STJ detector There has been considerable interest in attaching an adiabatic demagnetization refrigerator

(ADR) to the PT405 to cool superconducting tunnel junction (STJ) detectors to 100 mK. The STJ is a powerful tool for materials microanalysis because they can accurately measure X-ray energy. The PT405 will maintain the warm end of the two-stage ADR at 4 K, as well as cool the shield and current leads with the first stage. The first integration of PT405 into an ADR was made at Lawrence Livermore National Lab, Figure 10. A rigid cold finger joined the second stage of PT405 to the 4.2 K stage of the ADR. The purpose of this rigid connection in the preliminary test is to investigate the effects of pulse tube vibration on the detector measurement. The energy resolution of the measurements taken of the X-rays using STJ degraded from 26 eV, using liquid helium to 30 eV when attached to the PT405. Since no effort was made to mechanically or electrically isolate the ADR from the pulse tube, this result is very encouraging.

CONCLUSION

During the past year, the performance characteristics of Cryomech’s 2-stage, 4 K pulse tube cryorefrigerator have been measured in several applications. In each application, the pulse tube improved the system performance over what was previously possible with a GM. The benefits of the Pulse Tubes “no displacer” design are real. The PT405 has been improved to offer the typical cooling power of 0.6 W at 4.2 K and simultaneously 30 W at 65 K for both 60 Hz and 50 Hz power supplies. The PT405 is opening up attractive applications for cryogenic cooling around liquid helium temperatures.

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ACKNOWLEDGMENT

We would like to thank R. Ackermann at GE R&D Center for proving mounting force data; Cryomagnetics, Inc. for the data of conductively cooling a SC magnet; and J. Ullom of Lawrence Livermore National Lab for the data of precooling ADR for SJT detectors. REFERENCES 1. Tward, E., Chan, C.K., Jaco, C. et al, “Miniature Space Pulse Tube Cryocoolers”, Cryogenics, vol.31, no.8(1999), pp.717-720 2. Wang, C. and Gifford, P.E., “Performance study on a two-stage 4 K pulse tube cooler”, to be published in: Advances in Cryogenic Engineering, Vol. 45B, Plenum Press, New York (2000).

Experimental Study of a 4 K Pulse Tube Cryocooler S. Fujimoto1, T. Kurihara1, T. Oodo1, Y.M. Kang2, T. Numazawa3, and Y. Matsubara4 1

DAIKIN Air-Conditioning R&D Laboratory, LTD. 3 Miyukigaoka, Tsukuba 305-0841, Japan 2 DAIKIN Environmental Laboratory, LTD. 3 Miyukigaoka, Tsukuba 305-0841, Japan 3 Tsukuba Magnet Laboratory, National Research Institute for Metals 3-13 Sakura, Tsukuba 305-0003, Japan 4 Atomic Energy Research Institute, Nihon University 7-24-1, Narashinodai, Funabashi, Chiba 274-0063, Japan

ABSTRACT

A prototype two-stage pulse tube cryocooler was previously developed that uses a doubleinlet configuration. When the second-stage regenerator was filled with conventional magnetic regenerator materials, and the cryocooler achieved a minimum temperature of 2.89 K on the second stage and a maximum cooling capacity of 170 mW at 4.2 K. The rated input power of the compressor unit is 3.3 kW at 50 Hz. In this study, to improve the cooling capacity at 4.2 K for this cryocooler, we used a new oxide regenerator material, in the second-stage regenerator. This material has a magnetic transition temperature of about 3.8 K, and has a considerably larger heat capacity compared with that of and below 4 K. When was placed in the lowest-temperature part of the second-stage regenerator, the cryocooler achieved a no-load temperature of 2.51 K at the second stage, and a cooling capacity at 4.2 K of 250 mW. By adjusting the pulse tube size for further optimization, a cooling capacity at 4.2 K of 288 mW and a no-load temperature of 2.42 K was achieved. To further evaluate the effect of we placed in the second-stage regenerator of a 4 K Gifford-McMahon (GM) cryocooler. The cooling capacity below 4 K was improved, but that at 4.2 K was degraded. By numerical simulation, we determined the effect of on cooling performance both for the 4 K pulse tube cryocooler and the 4 K GM cryocooler. For the 4 K GM cooler, the numerical results roughly agreed with the experimental results. For the 4 K pulse tube cooler, the numerical results showed that, again, the cooling performance at 4.2 K was degraded by using By numerical simulation, we also examined the temperature distribution in the secondstage regenerator and the temperature oscillation in the second-stage expansion space. The mechanism that makes effective for improving the cooling performance at 4.2 K in the 4 K pulse tube cooler remains unclear. More detailed analysis is needed to clarify this mechanism. Cryocoolers 11, edited by R.G. Ross, Jr.

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INTRODUCTION We previously developed a single-stage Gifford-McMahon (GM) pulse tube cryocooler that uses a double-inlet configuration.1 For a 1.5 kW compressor unit, the cryocooler achieves a

minimum temperature of 25.7 K and a maximum cooling capacity of 20.0 W at 80 K. We also evaluated the dependency of the operating frequency on the inclination of the pulse tube. We measured the cooling performance by turning the cooler upside down, and found that the effect of the pulse tube inclination on the cooling performance at 80 K is negligible when the operating frequency is higher than 6 Hz. We also previously developed a two-stage pulse tube cryocooler that also uses the double-inlet configuration.2 The rated input power of the compressor unit is 3.3 kW at 50 Hz. When the second-stage regenerator is filled with lead spheres, the cryocooler achieves a minimum temperatures of 33 K at the first stage and 5 K at the second stage, and a maximum cooling capacity of 22.9 W at 80 K at the first stage and 5.9 W at 20 K at the second stage. Preliminary measurements using the same compressor unit showed that the two-stage cooler reached a minimum temperature of 2.89 K and provides 170 mW at 4.2 K when conventional magnetic regenerator materials, and are used in the second-stage regenerator. In this study, to increase the efficiency of this two-stage cooler, we used a new oxide magnetic regenerator material, in the lowest temperature part of the second-stage regenerator. The significantly improved the cooling capacity both at 4.2 K and below 4.2 K. We also placed in a second-stage regenerator of a 4 K GM cooler and determined its effect on cooling performance. Furthermore, by using numerical simulation, we compared the effect of in the 4 K pulse tube cooler with that in the 4 K GM cooler. EXPERIMENTAL APPARATUS AND PROCEDURE Two-stage 4K Pulse Tube Cryocooler

Figure 1 shows a schematic of our two-stage pulse tube cryocooler. The phase shifters for first and second stages are double-inlet configurations. Two reservoir tanks, each with a volume of 500

are connected to the hot ends of the first and the second stage pulse tubes through needle valves, and connected to the suction line of the compressor unit through other needle valves. The DC fluid current, flowing from the cold end of the pulse tube to the hot end, is controlled by using these needle valves. They also permitted control of the cooling performance. The rated input power of the compressor unit is 3.3 kW at 50 Hz. The first-stage regenerator was filled with about 1100 disks of 200-mesh phosphor bronze screen. The volume of the second-stage regenerator was about The temperatures of the first and second stage cold heads were measured by using a calibrated Pt-Co resistance thermometer and a calibrated Germanium resistance thermometer, respectively. The cycle frequency was 2.0 Hz and the initial pressure was 2.1 MPa. The regenerator cylinder can be separated into first-stage and second-stage cylinders, which are connected by their flanges and sealed by a metallic O-ring. Thus, the regenerator materials in the second-stage regenerator can be easily changed. First, we used a combination of lead spheres and conventional magnetic regenerator material, and in the second-stage regenerator. The materials were separated from each other by some phosphor bronze screens in the regenerator cylinder; lead spheres were placed in the upper half, and and were placed each in one-fourth of the lower half. To improve the cooling capacity at 4.2 K and the terminal temperature, we replaced pan of with a new oxide magnetic regenerator material, in the lowest temperature part (Figure 2a). The was developed at the National Research Institute for Metals to be used in cryocoolers that are designed for 4.2 K or less.3 can hold the high entropy below 4 K, has a magnetic transition temperature of about 3.8 K, and has a heat capacity considerably higher than that of and below 4 K (Figure 3).

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Two-stage 4K GM Cooler

To evaluate the effect of in a 4K GM cryocooler, we used a two-stage 4K GM cryocooler (DAIKIN prototype cooler). is used in the lower half part of the second-stage regenerator (Figure 2b). Using the same procedure as in the 4K pulse tube cooler, we varied the volumetric ratio of by replacing part of with The rated input power of the compressor unit is 6.5kW.

Figure 1. Schematic of two-stage cooler.

Figure 2. Schematic of second-stage regenerator.

Figure 3. Specific heat of

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NUMERICAL ANALYSIS METHOD

To evaluate our experimental results and to improve the cryocooler performance, based on a recent numerical simulation of a pneumatic-driven GM cryocooler4,5 we used numerical simulation of a two-stage 4K pulse tube cooler. Complete one-dimensional governing equations are solved by using finite difference method. Conservation of mass, momentum and energy of helium gas are written as follows;

Conservation of energy of helium gas and the regenerator matrix are written as follows;

Temperature of helium gas can be determined from the internal energy and density as follows; The equation of state of real gas can be written as;

In this simulation, variable physical properties of helium and regenerator materials were considered. The mass flow rate through the orifice valves was calculated by using the nozzle equation. The numerical model also includes the compressor unit to estimate the influence of its pressure oscillation as well as its performance (Figure 4).

Figure 4. Numerical simulation model.

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EXPERIMENTAL RESULTS Figure 5(a) shows the measured cooling performance of the 4K pulse tube cooler as a function of volumetric ratio of When was not used in the second-stage regenerator, the terminal temperature was 2.89K and the cooling capacity at 4.2K was 170mW. The cooling performance was unproved by using When the volumetric ratio of to that of the total and was 0.25, the terminal temperature was 2.51K and the cooling capacity at 4.2K was 250mW. The experimental results show that the new oxide regenerator material, significantly increased the cooling performance both at 4.2K and below 4.2K in the 4K pulse tube cryocooler. By adjusting the pulse tube size for further optimization, a terminal temperature of 2.42K and a cooling capacity at 4.2K of 288mW was achieved. Figure 5(b) shows the cooling performance of the 4K GM cooler as a function of the volumetric ratio of The figure shows terminal temperatures both with heat load of 0.4

W and 0.8W on the second stage. The experimental results show that was not effective in increasing the cooling performance at 4.2K, but was considerably effective at about 3.5K. COMPARISON BETWEEN EXPERIMENTAL AND NUMERICAL RESULTS

The experimental results show that the effectiveness of the new oxide regenerator material, in the 4K pulse tube differs from that in the 4K GM cryocooler. Using numerical simulation, we determined the cooling capacity both for the 4K pulse tube cooler and the 4K GM

Figure 5. Measured Cooling Performance.

Figure 6. Calculated Cooling Performance.

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cryocooler. Figure 6 show that

is not effective for improving the cooling performance

of either cryocooler at 4.2K but is effective for improving the cooling performance of both cryocoolers below 4K. For the 4K GM cryocooler, the numerical results agree with the experimental results. If the temperature oscillation in the expansion space is considerably large, the large specific heat of can improve the cooling performance at 4.2K. Figure 7 shows that the temperature oscillation is too small to improve the cooling performance by using the high specific heat below 4K. Figure 8 shows the temperature distribution of the second-stage regenerator both for the pulse tube cooler and the GM cryocooler. No remarkable difference is evident in the temperature distribution between the 4K pulse tube cooler and the 4K GM cooler. CONCLUSIONS We placed a new oxide regenerator material, in the second-stage regenerator in a two-stage 4K pulse tube cryocooler and in a 4K GM cryocooler. For the pulse tube cryocooler, significantly improved the cooling performance at 4.2K. The cooling capacity was increased from 170mW to 250mW and the terminal temperature was degraded from 2.89K to 2.5 1K. By adjusting the pulse tube size for further optimization, the cryocooler achieved a cooling capacity at 4.2K of 288mW and a terminal temperature of 2.42K. For the 4K GM cryocooler, was not effective in improving the cooling performance at 4.2K, but was considerably effective at about 3.5K. By using numerical simulation, we determined the temperature distribution in the second-

Figure 7. Calculated Temperature Oscillation of helium gas in 2nd-stage expansion space.

Figure 8. Calculated Temperature Distribution of 2nd-stage regenerator.

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stage regenerator and the temperature oscillation in the second-stage expansion space. The mechanism that makes effective for improving the cooling performance at 4.2 K in the 4K pulse tube cryocooler remains unclear. More detailed analysis is needed to clarify this mechanism.

REFERENCES 1.

S. Fujimoto, Y.M. Kang, and Y. Matsubara, "Development of a 5 to 20W at 80K GM Pulse Tube

2.

Cryocooler," Cryocoolers 10, R.G. Ross, Jr., ed., Kluwer Academic/Plenum Publishers, New York (1999), p. 213. S. Fujimoto, Y.M. Kang, T. Kanyama, and Y. Matsubara, "Experimental Investigation of Some Phase Shifting Types on Two-stage GM Pulse Tube Cryocooler," CEC/ICMC' 99, to be published T. Numazawa, O. Arai, A.Sato, S. Fujimoto, T. Oodo, and Y.M. Kang, T. Yanagitani, "New

3.

Regenerator Material for Sub-4K Cryocoolers," Cryocoolers 11, R.G. Ross, Jr., ed., Kluwer Academic/Plenum Publishers, New York (2001). 4. 5.

T. Kurihara and S. Fujimoto, "Numerical Analysis of the Performance of Pneumatic Driven 4K-GM Refrigerator," Cryogenic Engineering (in Japanese), Vol.31 (1996), p. 197. T. Kurihara, M. Okamoto, K. Sakitani, H. Torii, and H. Morishita, "Numerical and Experimental Study of a 4K Modified-Solvay Cycle Cryocooler," Adv. Cryog. Eng. 43 (1998), p. 1791.

Nomenclature specific heat of matrix temperature equivalent diameter internal energy

T

u

e v f

total energy velocity coefficient of friction external PV work

w

acceleration of gravity density P

q t

pressure equivalent thermal conductivity of regenerator heat exchange between regenerator and cylinder wall time

Subscripts g

gas

m

matrix

GM-Type Two-Stage Pulse Tube Cooler with High Efficiency A. Hofmann*, H. Pan**, L. Oellrich** *Forschungszentrum Karlsruhe Institut für Technische Physik D-76021 Karlsruhe, Germany **Univ. Karlsruhe, Institut für Technische Thermodynamik und Kältetechnik D-76128 Karlsruhe, Germany

ABSTRACT

A two-stage pulse tube refrigerator has been designed for maximum refrigeration powers at 20 K and 50 K, when powered by a 6.5 kW of electric compressor. The modular setup of the cold head

enables easy access to all components to be modified for the optimization of the system. All gas flows at the regenerator and at both pulse tubes are controlled by solenoid valves. Additional adjustment of the flow is done by throttling valves at the pulse tubes. Two arrangements with different sizes of the second stage have been tested. With the small size second stage, the typical cooling power achieved was 55 W at 50 K for the first stage together with 3.5 W at 20 K. This corresponds to about 5.0 % of summarized Carnot efficiency. Even higher efficiency of 5.8 % Carnot was obtained for the system with enlarged second-stage components operated with 40 W at 46 K plus 10.5 W at 20 K. No-load temperatures down to 8.5 K were achieved with lead spheres for the second-stage regenerator. Some details on the design of the test rig and on operational parameters are given. In addition, the results are compared with numeric predictions based on a small amplitude thermoacoustic model. INTRODUCTION

In the early days, pulse tube coolers were considered most attractive because of their simple design, low noise, and projected high reliability; it was believed that lower efficiency than that of a comparable Stirling or GM cooler had to be accepted. However, now the pulse tube process is understood as a Stirling process. With this understanding, there are no physical reasons for having lower efficiency with a properly designed pulse tube cooler. In the past few years this has been verified for compact single-stage Stirling-type coolers1 and also for GM-type pulse tube systems.2 With multi-stage GM-type PTRs even temperatures in the range of 2 K are being achieved.3,4 Such temperatures have not yet been realised with conventional GM coolers. These examples show that pulse tube coolers may be substituted for many types of conventional Stirling and GM coolers. In the present paper, research is focused on a two-stage GM-type pulse tube cooler for operation with a 6.5 kW input power compressor.

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EXPERIMENTAL SETUP

The experimental set-up, shown schematically in Fig. 1, was designed to enable manifold modifications. A commercial product would have required a lot of simplifications. The cooler is operated with a conventional compressor (Coolpak 6000®) equipped with additional buffer volumes in the suction and exhaust lines. The time-averaged flow rate is measured with a mass flow meter (FM). The cold head has been designed for having refrigeration power at two levels, around 50 K and around 20 K. All components are flanged. Pressures are measured at the warm end of both the regenerator and the pulse tubes. Temperature sensors are attached on the outer surfaces of regenerators and pulse tubes, and in addition, at both cooling stages (copper blocks with ohmic heaters). The first-stage regenerator is filled with stainless steel mesh. Easy variation of the active length is made possible by use of a solid plug at the upper end. The second-stage regenerator is filled with Pb shot. Different tubes with diameters up to about 25 mm and lengths up to 200 mm can be connected. The uncommon configuration with an inline arrangement of the first-stage pulse tube and the second-stage regenerator has no physical reason. It was chosen to allow better adjustment of the components. Water-cooled heat exchangers are located on the warm ends of both pulse

tubes. All gas flow in the regenerator and both pulse tubes is controlled by solenoid valves. Additional adjustment of the flow is done by two pairs of throttling valves at the pulse tubes. The solenoid valves can be actuated individually. The typical timing for 2 Hz operational frequency is shown in the insert in Fig. 1. However, each configuration requires some re-adjustment. RESULTS Continuous improvement of the system has been achieved through manifold modification of the hardware components together with tuning of the valve timing and needle valve settings, accompanied by recalculations. The first step was aimed at getting the most power at the 50 K level, together

Figure 1. Schematic of the 2-stage PTR.

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Figure 2. Refrigeration power and temperatures achieved with Coolpak® compressor (6.5kW) and different second stage configurations.

with rather modest power at 20 K. This result is shown in Fig. 2a), where the temperatures and are plotted together with the heater powers applied to those stages. The valves where adjusted for getting optimum conditions for a heat load of 3 W at the second stage together with 50 W at the first stage (circled points). The respective temperatures are 18 K and 48 K. No further adjustment was done during the variation of the heater powers. In a second step, the size of the second-stage components was increased to achieve more cooling power at the second stage. Some degradation of the first-stage refrigeration power was expected. The result is shown in Fig. 2b. Here, the system was adjusted with heat loads of 40 W and 10 W at the first and second stages, respectively. The no-load temperatures went down to 26 K and 8.7 K. Somewhat lower temperatures could have been achieved by readjustment at no-load. This refrigeration chart is very close to that of commercial GM coolers designed for the same operational range. For having a better defined characterisation of the efficiency, we calculate the overall Carnot efficiency according to

Our respective results together with some literature data of other systems are compiled in Table 1. In the best case, the reference point in Fig. 2b, we get an efficiency of 5.8 % Carnot. This is close to data evaluated for commercial GM refrigerators7 operated with the same type of compressor; also

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Figure 3. a) Cool-down with nominal load and b) temperature stability.

the more advanced laboratory-type 4K GM coolers5,6 are in the same range. Appreciably higher

efficiency values have been obtained for single-stage pulse tube coolers of the Stirling-type1 and of a GM-type with active buffer phase shifter.9 Our pulse tube cooler operated with six solenoid valves and four additional needle valves has a great number of parameters to be adjusted for tuning. One important criterion is the compensation of the dc flow which can exist in different loops. Direct observation of the dc flow is not possible in our system. Instead, the change in temperature distribution detected at the different components gives valuable advice for handling. If, for instance after modification of one needle valve, the temperature in the regenerator is increased and decreased in one pulse tube, it may be argued that a dc flow from the regenerator to the pulse tube has been increased. Much experience is required for the tuning. But after documentation of all adjustment parameters, excellent reproducibility of the system is obtained even after dismantling and re-assembling of components. A typical cool-down curve with heat load applied from the beginning and the temperature stability during a 8-h run is shown in Fig. 3. Steady state operation is established after some 90 min, thereafter temperature fluctuations are less than 1 K at both stages. It has also been shown, that such pulse tube refrigerators can rather easily be adapted to specific requirements with different distributions of the thermal loads at both stages. Moreover, a powerful single-stage refrigerator is obtained just by disconnection of the second stage. The respective result is shown in Fig. 4. Typically, a refrigeration power of 120 W is obtained at 70 K. But it should be mentioned, that the same first-stage components as they have been optimised for the two-stage operation have been used. Higher efficiency might be obtained by additional modification of those components.

Figure 4.

Refrigeration power of the first stage (second stage disconnected).

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NUMERIC MODELING

Although the performance of pulse tube systems is such that they can compete with conventional GM refrigerators, their absolute efficiency is still rather small. There is certainly some margin for further improvement, but the question is where to concentrate the needed effort. With this objective, we have tried to analyse pulse tube systems using numeric models. Our code8 is based on a small

amplitude thermoacoustic approximation, on the assumption of laminar flow in all components, and on the approximation of the regenerators by equivalent parallel channel structures. This rather crude model yields quite a good prediction of the refrigeration powers and of the temperature distribution in both regenerators and in the pulse tubes. On that basis, the model is used for calculating the work flow and the losses at different positions of the system. For evaluating the different contributions in the efficiency term, it is more appropriate to consider the term

where the efficiency is not related to Carnot, with as it is for Eq.(1), but to which is the ideal pulse tube efficiency for all single-stage configurations where no expansion work rejected from the pulse tube is recovered.10 With the work flow scaling and with the flow rate fraction

this can be transformed into

Where

are the effective refrigeration powers of the first and second stage, respectively, and are the work flows (expansion powers) at the cold end of the pulse tubes. The mass flow fraction in PT 1 is 0.6 for the present simulation, and is the work flow at the hot end of the first regenerator. Those results are plotted in Fig. 5. The work flow (“pV-power”) at the inlet of the first regenerator is about 1 kW. This is supplied by the compressor with 6.5 kW of electric input power measured for typical operational conditions. Hence, the efficiency of such a valved pressure wave generator is in the range of This low compressor efficiency, a drawback ofthe GM operation, is certainly the greatest loss term, but it is about the same for conventional GM coolers.

Next, we consider the first-stage regenerator. The expansion power (pV-work flow) in the first pulse tube is 91 W. It is balanced by 44 W of enthalpy flow in the first-stage regenerator (regenerator loss) and by 12 W of heat flow coming down through the first pulse tube (“shuttle loss”). Hence the residual effective calculated refrigeration power is 35 W at 45 K, a value close to the measurement. This corresponds to a first-stage efficiency with In the second stage, the expansion power is 25 W, the regenerator loss is 5 W, and the pulse tube loss amounts to 8 W. This gives a residual refrigeration power of 12 W at 20 K, and the respective efficiency becomes The calculated mass flow is 60 % in the first and 40 % in the second pulse tube. Hence, the overall scaled efficiency is

a value

close to the measured Carnot efficiency. The term within the brackets may be considered as the overall efficiency of the cold head. It amounts to 0.42. An overall efficiency in this range can only be approached by use of a very efficient linear drive compressor, which, however, may be too bulky

and expensive for low-frequency operation, which is 2 Hz in the present case. For a more detailed evaluation, we consider the ratio of losses to the work flow. This results in 0.29 for the first regenerator, 0.13 for the first pulse tube, 0.2 for the second regenerator, and the greatest value of 0.32 for the second pulse tube. Hence, apart from the required improvement of the compressor, the most effort should be spent to improve the first regenerator and the second pulse

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Figure 5. Calculated work and heat flows.

tube. Those calculations, however, are based on some arbitrary assumptions such as longitudinal

thermal conductivity being the same as for the bulk material, and laminar flow in the pulse tubes. Hence, the regenerator loss might be overestimated. But the pulse tube loss is more likely underestimated. The reduction of those losses is no trivial problem. Another interesting term is the expansion power removed from the pulse tubes. This is 87 W at pulse tube 1 and 22 W for the second tube. Both terms are much smaller than the power fed into the regenerator. Recovering the expansion power by use of piston expanders would only yield a small improvement (in the 10 percent range) of the overall efficiency. No expansion work is recovered in the present valved system. CONCLUSIONS

Stable and reproducible operation is obtained with a GM-type two-stage pulse tube cooler operated with 3 pairs of solenoid valves for controlling the flow at the warm ends of the regenerator and of both pulse tubes. The refrigeration powers at the 50 K and at the 20 K stages can be redistributed just by modification of the second-stage components. The overall efficiency of two-stage pulse tube coolers is shown to be very close to that of conventional GM-coolers. But the absolute value is

still much smaller than the ideal Carnot efficiency. Most of this is caused by the valved pressure wave generator, which is estimated to have only 15 % efficiency. The cold head itself is analysed to

have a rather good performance with more than 40% of Carnot efficiency. The computer model, based on small amplitude approximation and on the assumption of having laminar harmonic flow

in all components, gives some hints for localising the predominant loss terms. But this model should not yet be stressed for quantitative evaluations.

REFERENCES 1.

W.W. Burt, and C.K. Chan, “New Mid-Size High Efficiency Pulse Tube Coolers,” Cryocoolers 9, Plenum Publishing Corp., New York, 1997, pp. 173-182.

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2. A. Ravex, I. Charles, L. Duband and J.M. Poncet, “Pulse Tube Development at CEA/SBT,” Proc. of 3.

4. 5. 6.

the IIR Conf., Sidney, Sept. 1999. C. Wang, G. Thummes, and C. Heiden, “Experimental Study of Staging Method for Two-Stage Pulse Tube Refrigerators for Liquid Helium Temperatures,” Cryogenics, 37 (1997), pp. 159-164. M.Y. Xu, A.T.A.M. De Waele, Y.L. Ju, “A pulse tube refrigerator below 2 K,” Cryogenics, 39 (1999), pp. 865-869. A. Onishi, “4K-GM Cryocoolers having little orientation dependency and small influence from mag-

netic field,” Cryogenic Engineering (J. Cryog. Eng. Soc. Japan, Tokyo), Vol. 34 (1999), pp. 233-235. J.N. Chafe and G,F. Green:, “Performance of a Low Temperature Giffbrd McMahon Refrigerator Utilizing Neodymium Disk Regenerator,” Advances in Cryogenic Engineering, Vol. 43 (1998),

pp. 1783-1790. Leybold-Katalog Vakuumtechnik (1998). A. Hofmann and S. Wild, “Analysis of a Two-Stage Pulse Tube Cooler by Modeling with Thermoacoustic Theory,” Cryocoolers 10, Plenum Publishing Corp., New York, 1999, pp. 369-377. 9. S. Zhu, Y. Kakimi and Y. Matsubara, “Investigation of active-buffer pulse tube refrigerator,” Cryogenics, 37(1997), p. 461. 10. P. Kittel, A. Kashani, J.M. Lee and P.R. Roach, “General pulse tube theory,” Cryogenics, 36 (1996) pp. 849-857.

7. 8.

Developments on Single and Double Stage GM Type Pulse Tube Cryorefrigerators J.M. Poncet, A. Ravex* and I. Charles Service Basses Températures, CEA-DRFMC, 17, rue des Martyrs 38054 Grenoble Cédex 9, France *present address : Air Liquide, DTA, BP 15 38360 Sassenage, France

ABSTRACT Pulse tube (PT) cryocoolers with no moving parts in their cold ends are potentially able to replace conventional Gifford MacMahon (GM) cryocoolers for most applications. Over the past several years, CEA/SBT has undertaken basic research and prototype development of GM driven PT cryocoolers. Systematic studies have been carried out to characterize flow distribution and

d.c. flow effects. Based on a good comprehension of these phenomena, several prototypes have been built and optimized for various customised applications. Their performance is presented in this paper.

1. INTRODUCTION Several years ago, CEA/SBT initiated basic research on PT cryocoolers to understand and analyse their operation. A model has been proposed1 and a numerical tool has been developed for the design and optimisation of prototypes. As suggested by Zhou2, a double inlet configuration has been systematically used for performance improvement. It is generally stated that this configuration with a gas supply to the tube at both the cold and warm ends reduces the mass mass flow rate through the regenerator, thus increasing its efficiency. In fact, the major interest of the double inlet configuration is the achievement of a better phase shift between the pressure versus mass flow oscillations at the cold end of the tube. Gedeon3, through theoretical calculations, has predicted that the double inlet configuration may allow for d.c. flow through the tube and regenerator that is able to produce a parasitic heat load of the order of the PT gross cooling power. This effect has been observed4 at CEA/SBT during Stirling type PT developments. Experimental evidence was also observed by Chen5 during the development of a double stage 4 K GM type PT. He observed a large improvement in the performance of their prototype by fine tuning of a so-called minor orifice (needle valve) set between the buffer and the

compressor suction line. In fact, a controlled d.c. flow is created at room temperature through this connection which cancels the parasitic d.c. flow in the cold region. This artifact is now well understood and commonly used.6,7 Based on a good comprehension and control of this d.c. flow, and on extensive studies on distribution valve design and phasing control, several prototypes have been designed, optimised and characterized for various applications. Cryocoolers 11, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2001

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Figure 1. SingleStage Pulse Tube (100 W at 80 K).

2. SINGLE STAGE PROTOTYPES

At the 1997 CEC in Portland, we have reported 8 results on a single stage GM type prototype with the following characteristics :

Compressor : CTI8500 / 5 kW Operation frequency : 2 Hz Ultimate temperature : 26 K Cooling power : 80 K /100 W Configuration : U shaped A picture of this prototype is given on Figure 1. Based on this first development, several prototypes have been extrapolated for various applications. Performances and main features of these coolers are reported in Table 1. The main evolutionary developements from the initial cooler have been required by specific applications. For example, to reduce the vibration exported to the devices to be cooled, the distribution valve has been systematically disconnected from the cold head with a flexible connection line up to 1 meter long.

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A well known drawback of the low frequency operated GM type PT is its performance sensitivity to the orientation of the cold finger.8 To avoid this inconvenience, operation at 5 Hz has been demonstrated with a large reduction in the orientation effect. For HTS applications or single thermal shield helium cryostats, large cooling power in the 30K/50K temperature range is required. To achieve this goal, the initial prototype has been modified by introducing lead shot in the regenerator, with the following characteristics : Compressor : Leybold Coolpak / 6 kW

Operation frequency : 2 Hz Ultimate temperature : 17 K Cooling power : 30 W at 30 K or 70 W at 50 K Configuration : U shaped Further developments are in progress on single stage PTs to develop a full range of coaxial type cold heads – which may allow for an easier integration – to be associated with 1.5 kW, 3 kW and 6 kW standard GM compressor units. 3. DOUBLE STAGE PROTOTYPES

The common applications for double stage GM coolers are cryopumping and MRI cryostat thermal shield cooling. For these applications, large cooling power at bom 80 K and 15 K temperature

levels is required. With the goal to demonstrate that PT coolers are able to achieve these requirements, SBT/CEA has developed two GM type prototypes with overall sizes (length from 300 K flange to second stage heat station : 325 mm, overall diameter : 95 mm for the largest prototype) equivalent to existing GM coolers. A picture of such a double stage GM type PT prototype on a test bench is presented in Figure 2. The performance of both prototypes, established with two types of compressors, is reported in Table 2 and in a traditional map presentation in Figures 3 and 4. It can be observed that the obtained performance is comparable to that of the corresponding GM coolers.

Figure 2. Double stage PT prototype on a test bench.

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Figure 3

Figure 4

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The prototypes were optimized for operation at respectively 80 K and 15 K on the two stages, thus using stainless-steel gauze discs (210 mesh) in the first stage regenerator and lead shot (200 in the second stage regenerator. Without any geometrical modification for optimisation, these prototypes have been also operated in the liquid helium temperature range, introducing magnetic material in the second stage regenerator. The results obtained are summarised in Table 3. These results are encouraging for operation at LHe temperatures and may be further improved by optimisation of the second stage regenerator and tube geometries.

4. CONCLUSION The performance obtained with the various single and double stage GM-type PT coolers developed at CEA/SBT demonstrates that the new PT technology, with its potential advantages of reliability and ease of integration, is mature for most applications. A collaboration with industrial partners is now necessary to introduce commercial PT coolers into the traditionnal markets for GM coolers.

REFERENCES 1.

Liang, J., Ravex, A. and Rolland, P., “Study on pulse tube refrigerator Part 1 : Thermodynamic non

symmetry effect Part 2 : Theoretical modelling Part 3 : Experimental verification”, Cryogenics, vol 36 (1996), pp. 87-106.

2. 3.

4.

5. 6.

7. 8.

S. Zhu, P. Wu and Z. Chen, “Double inlet pulse tube refrigerators: an important improvement,” Cryogenics, vol. 30 (1990), pp. 514. D. Gedeon, “DC gas flows in Stirling and pulse tube cryocoolers,” Cryocoolers 9, R.G. Ross Jr., ed. Plenum Press, New York (1997), pp. 385. L. Duband, I. Charles, A. Ravex, L. Miquet and C. Jewell, “Experimental results on inertance and permanent flow in pulse tube coolers,” Cryocoolers 10, R.G. Rosw Jr., ed., Plenum Press, New York (1999), pp. 281-290. G. Chen, J. Zheng, L. Qiu, X. Bai, Z. Gan, P. Yan, J. Yu, T. Jin and Z. Hang, “Modification test of staged pulse tube refrigerator for temperatures below 4 K,” Cryogenics, vol. 37 (1997), pp. 529. C. Wang, G. Thummus and C. Heiden, “Control of DC gas flow in a single stage double inlet pulse tube cooler,” Cryogenics, vol. 38 (1998), pp. 843. I. Charles, L. Duband and A. Ravex, “Permanent flow in low and high frequency pulse tube coolers – experimental results,” Cryogenics, vol. 39 (1999), pp. 777. A. Ravex, J.M. Poncet, I. Charles and P. Bleuzé, “Development of low frequency pulse tube refrigerators,” Advances in Cryogenic Engineering, Plenum Press, New York, vol. 43A (1998), pp. 1957.

30 - 50 K Single Stage Pulse Tube Refrigerator for HTS Applications J. Yuan, J. Maguire, A. Sidi-Yekhlef, and P. Winn

American Superconductor Co. Westborough, MA 01581 U.S.A.

ABSTRACT

The need for reliable cryocoolers for High Temperature Superconducting (HTS) applications ranging from 30 to 50 K has become apparent in the past several years. Many cryocooler designs are under development, including GM, Stirling, and pulse tube. Pulse tube refrigerators have attracted extensive interest in recent years due to their high potential for reliability and simplicity. This paper describes the development program of a 30 – 50 K single

stage pulse tube refrigerator for use in a HTS system at American Superconductor Co. Initial tests indicate that the cooler can reach a minimum temperature below 20 K and has a cooling capacity of 20 watts at 30 K, and 60 watts at 50 K, with an input power of 6 kW. Systematic investigation of the effects of valve timing, operating frequency and compressor input power on the cooler performance is presented in the paper. Additionally, the stability issue of the pulse tube refrigerator is addressed. The progress towards development of a successful 30 – 50 K pulse tube and its potential and limitations are also discussed. INTRODUCTION

In the last decade, the significant advance of the performance levels of high temperature superconducting (HTS) wire has made it suitable for commercially viable applications such as synchronous motors, generators, transformers and electric power cables. Currently, to provide

the requisite current density, the HTS wire formed from BSCCO and YBCO must be operated at temperatures around 30 K and 50 K, respectively.1 Cryogenic refrigeration systems are one of the most crucial components for the successful commercialization of HTS applications. In most small scale HTS applications, such as motors and generators, cryocoolers will be the most attractive candidates due to their size , cost and reliability. Over the past two decades, since the discovery of HTS materials, the need for reliable cryocoolers for HTS applications ranging from 30 to 50 K has become apparent. Many cryocooler designs are under development, including GM, Stirling and pulse tube. Pulse tube refrigerators have attracted extensive interest in recent years due to their potential for reliability and simplicity. However, few attempts have been made to develop a single stage pulse tube refrigerator with relatively large cooling capacity at temperatures ranging from 30 K to 50 K. The present paper describes efforts that have been made to develop a high power 30 to 50 K

pulse tube refrigerator for potential HTS applications. Initial tests indicate that the cooler can reach a minimum temperature below 20 K and has a cooling capacity of 20 watts at 30 K, and 60 Cryocoolers 11, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2001

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watts at 50 K, with an input power of 6 kW. Systematic investigation of the effects of valve timing, operating frequency, and compressor input power on the system performance are presented in the paper. Additionally, the stability issue of the pulse tube refrigerator is addressed. The progress towards development of a successful 30 – 50 K pulse tube, its potential and

limitations, are discussed. DESCRIPTION OF COOLER AND TEST APPARATUS

The pulse tube cooler described in this paper is designed for temperatures ranging from 30 K to 50 K applications. A thermodynamic model2 was used to size the pulse tube, buffer and regenerator. The U-shape configuration is utilized for the design. The cooler flow diagram and complete package are illustrated in Figure 1. As shown in Figure 1, the cooler utilizes a modified active valve configuration3 that allows control of the phase shift at the warm end of the pulse tube to optimize the cooler performance. Furthermore, this configuration allows control of the DC flow by adjusting the opening and timing of valves 3 and 4. Finally, from a practical

point of view, this configuration makes it possible to design using a rotary valve. The pulse tube is fabricated from a thin-wall stainless steel tube with a 45 mm outside diameter and 0.5 mm wall thickness. The tube is 270 mm long which yields a total pulse tube volume about 0.4 liter. To optimize the performance of the cooler at temperature ranging from 30 K to 50 K, a three layer regenerator containing different bronze mesh sizes at the warm end and fine lead shots at cold end has been designed and constructed. In the design process, particular attention has been paid to the cold end heat exchanger. Optimizing the design of the cold end heat exchanger requires balance between the dead volume, total heat transfer area, and pressure drop. Detailed calculations indicate that the heat exchanger

efficiency can be dramatically increased by carefully arranging the gas flow channels. The cold end heat exchanger has been optimized to have a high heat transfer coefficient, large heat transfer area and a small pressure drop and dead volume. Although a water-cooled warm end heat exchanger is used for the current testing, the design of a helium-cooled warm end heat exchanger is underway; this will be a part of the complete package. For test purposes, a valve system comprising solenoid and throttle valves was used to

control the gas flow in and out of the refrigeration system. The flow coefficients of the two throttle valves connected in series with valves 3 and 4 can be adjusted from zero to 0.5 while the flow coefficient of throttle valve 5 can vary from zero to 0.15. Timing of the solenoid valves has

Figure 1. Pulse tube cryocooler: (a) complete packge (b) flow diagram.

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been accomplished through the use of a controller. Different valve timings are achieved easily by programming. The design of the rotary valve similar to the one used in a GM cryocooler is underway which will replace the solenoid valve system. The valve timing of the rotary valve is fixed by the hole configuration machined into the valve plates. Nevertheless, the following test

results were obtained utilizing the solenoid valve system. Five separate piezoresistive pressure transducers, which allow measurement of pressure oscillations, are mounted on the warm ends of the pulse tube and regenerator, on the discharge and suction lines of the compressor, and on the buffer, respectively. The cold head temperature is measured by two silicon diodes. Four cartridge heaters which can supply more than 150 watts are placed at the cold end heat block to measure the cooling capacity. In order to reduce the thermal radiation loss at low temperatures, more than 15 layers of aluminized Mylar are wrapped around the cold parts. A Leybold RW6000 compressor with an input power of 6 kW was used in

most of the experiments. For comparison purpose, a Leybold CoolPak 6000 with an input power of 7 kW was also used in some of the tests. The working medium used in all tests was helium

gas.

WORKING PROCESS Theoretical understanding of the pulse tube refrigerator is based on the analysis of the thermodynamic cycle of the gas column within the pulse tube2. The refrigeration mechanism of a pulse tube refrigerator basically results from gas expansion at the cold end of the pulse tube, which is similar to that in a GM refrigerator. However, in the case of the pulse tube refrigerator, the expansion work is transferred by a “gas piston” from the cold end to the warm end of the pulse tube instead of a solid displacer as is in the case of GM cryocooler. The system cooling capacity is determined by the expansion work at the cold end of the pulse tube. To maximize the expansion work, careful control of the valve timing is necessary. A typical valve timing for the cooler is illustrated in Figure 2. The solid lines represent the time that valve is open. To better describe the thermodynamic cycle of the cooler, the time dependent pressure waves within the pulse tube and the buffer obtained by the theoretical simulation are presented in Figure 3 (a). The corresponding start points are also shown in the same graph. Both the compression and expansion processes can be divided into three steps as can be seen in Figure 2 and Figure 3 (a). The entire working process includes a) (1-2 )gas is admitted to the pulse tube from buffer, b) (2-3) gas flows into the pulse tube from compressor, c) (3-4) gas flows from compressor into buffer through the regenerator and the pulse tube, d) (4-5) gas is exhausted from the pulse tube to the buffer, e) (5-6) gas flows out from the pulse tube to the compressor and f) (6-1) gas flows from the buffer to the compressor through the pulse tube and

regenerator.

Figure 2. A typical valve timing.

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Figure 3. Pressure oscillation obtained by (a) theoretical simulation (b) experiments.

PERFORMANCES AND DISCUSSIONS Cooldown

System cooldown characteristics have been first studied based on the optimized valve timing. Figure 4 displays a typical cooldown curve with operating frequency of 1.5 Hz. As can be seen, the cooldown process takes about 40 min. The minimum no load temperature achieved to date is 19.8 K with an operating frequency of 1.5 Hz and a pressure ratio of 1.9. The pressure oscillations within the pulse tube, regenerator, buffer and compressor gathered during the experiments are displayed in Figure 3 (b). The experimental results show that the wave patterns are very similar to those obtained by the theoretical simulation as seen in Figure 3 (a). The pressure difference between the pulse tube and regenerator is due to the combined effects of the packed bed and throttle valves. The effect of valve timing can be easily observed from the slope changes of the pressure waves. One should notice that, in the theoretical simulation, the pressure drop across the regenerator has been ignored.

Figure 4. Cooldown characteristics of the refrigeration system.

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Figure 5. Influence of valve timing on cooling capacity

Influence of Valve Timing

To study the influence of valve timing on the system performance, different valve timings have been tested with same compressor discharge and suction pressures. An operating frequency of 1.67 Hz and fixed openings of throttle valves were employed for all the tests. As an example, the cooling capacities as a function of the timing ratio of compression to expansion are displayed in Figure 5. For all tests, the time ratios of each individual valve and remained the same while the time ratio of compression to expansion varies. The time ratios of compression to expansion utilized in the experiments were 270: 330, 300:300 and 330: 270. Figure 5 illustrates the test results with three different compression to expansion time ratios. As can be seen, the cooler performance strongly depends on the valve timing. The minimum temperature achieved with different time ratio 270:330, 300:300 and 330:270 are 22.5 K, 24.4 K, and 20.5K, respectively. The best cooling capacity at 30 K was obtained with the time ratio 270:300 while the minimum temperature was obtained with a time ratio of 330:270.

Figure 6. The influence of the operating frequency on cooling capacity.

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Figure 7. Cooler stability testing data over 5 days

Influence of Operating Frequency The cooling capacity of the cooler as a function of frequency is plotted in Figure 6. All the data in this figure results from tests run with the same valve timing and throttle valve opening. The time ratio of compression to expansion used in experiments was 300:270. As can be seen in Figure 6, the operating frequency of the valves displays a noticeable influence on the refrigeration capacity. In general, there is best operating frequency for a given pulse tube system at a specific operating temperature. As shown in Figure 6, the best performance at 50 K was obtained with an operating frequency of 1.5 Hz while the best performance at 77 K was obtained with an operating frequency of 1.67 Hz. Stability A stability test was carried out over a period of about five days. During this period, a variety of heat loads have been applied to the cold head. As shown in Figure 7, the temperature is quite stable and remains constant for a given load. In general, the temperature variation is less than +/0.5 K. The experimental results revealed that the higher the operating frequency, the larger the temperature variations. The reason for this may be partly related to the performance of the solenoid valves since the higher the operating frequency, the shorter the valve operating time.

Figure 8. Effects of compressor input power on cooling capacity

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Influence of Compressor Input Power

To study the influence of the compressor input power on system performance, a Leybold CoolPak 6000 compressor with input power of 7 kW was used to drive the pulse tube cooler. Figure 8 shows the test results. In both tests, operating parameters were same except for the input power. As can be seen, even though the minimum temperature for both cases are almost same, the cooling capacity with CoolPak 6000 is about 16% larger than that with RW6000 for a given temperature. CONCLUSIONS

A single stage 30 K – 50 K pulse tube refrigerator has been designed, built and tested based on a thermodynamic model. It has been found that the single stage pulse tube has the potential to supply high cooling power at temperatures ranging from 30K to 50 K. Initial test results are

encouraging for future development. In the initial tests, minimum temperatures below 20 K and 60 watts at 50 K were reached with an input power of 6 kW. Typical cooler performance demonstrated a percentage Carnot efficiency of about 5% at 50 K. Systematic investigation indicates that valve timing has significant effect on the system performance. Tests also show

that the cooler is relatively stable during the test period. Further work is needed to design the matching compressor and improve the overall system efficiency. ACKNOWLEDGEMENT

The work described here was supported by American Superconductor Co.. The authors would like to thank Don Hannus, Jay Taber and other coworkers at American Superconductor for their assistance and support with the many test setups. REFERENCES 1. Malozemoff, A. P., Carter, W., Fleshler, S., Fritzemeier, L., Li, Q., Masur, L., Miles, P., Parker, D., Parrella, R., Podtburg, E., Riley Jr. G.N., Rupich, M., Scudiere, J., Zhang, W., “HTS Wire at

Commercial Performance Levels” IEEE Transactions on Applied Superconductivity 9, 1999, pp.2469 – 2473. 2. Yuan, J and Pfotenhauer, J. M. “Thermodynamic Analysis of Active Valve Pulse Tube Refrigerator”. Cryogenics, Vol. 39. No. 4, pp. 283 – 292.

3. Yuan, J and Pfotenhauer, J. M. “A Single Stage Five Valve Pulse Tube Refrigerator Reaching 32 K”. Advances in Cryogenic Engineering Vol. 43, pp. 1983 - 1989

Two-Stage 4 K Pulse Tube Refrigerator Shaowei Zhu, Masahiro Ichikawa, Masafumi Nogawa, and Tatsuo Inoue Second Development Department Aisin Seiki Co., Ltd., Kariya, Aichi, 448-8650 JAPAN ABSTRACT This paper describes the manufacture and test of a thermally connected two-stage 4 K pulse tube refrigerator. The second stage of the refrigerator involved the use of either a “two middlebuffer with double inlet” type phase shifter or a “two middle-buffer with displacer.” The phase shifter of the first stage was a two middle-buffer type. The connecting tubes between the cold head and the pressure switching valves were 1.5 meter long in order to reduce the vibration and the electromagnetic noise near the cold head. The best cooling capacity, 0.58 W at 4.2 K, was achieved with 5.5 kW of input power and the two middle-buffer with double inlet type phase shifter. INTRODUCTION

A 4 K pulse tube refrigerator is expected to be one of the most important refrigerators for either laboratory use or for medical instruments such as MRIs and SQUID systems; this is because of its low noise, low vibration; and no need to remove the cold head during maintenance. In most papers,1,2 a gas-connected structure is adopted for the double staging of the pulse tube refrigerator. In this paper, a thermally connected two-stage pulse tube refrigerator that consists of two independent refrigerators for each stage is manufactured and tested. An advantage of the thermally connected structure is that each stage can be driven at different conditions to achieve optimization. In order to decrease the noise level, 1.5-meter-long connecting tubes between the pressure switching valves and the cold head were used. One problem caused by such long connecting tubes was a rather large amount of work loss from the compressor due to the pressure drop and void volume in the tubes. To achieve a higher efficiency, a “two middle-buffer” type phase shifter was used for the first stage refrigerator. For the second stage, two types of the phase snifters were

tested: 1) a “two middle-buffer combined with double inlet” type, and 2) a “two middle-buffer combined with displacer” type. STRUCTURE OF THE TEST REFRIGERATOR Figure 1 shows a schematic of the test machine. R11 and PT1 are considered a single refrigerator, referred to as the first stage refrigerator. R21, R22, and PT2 are also considered as another single refrigerator, called the second stage refrigerator. The cold ends of the regenerators R11 and R21 are thermally connected to each other. In the regenerators R11 and R21, screen meshes of copper and stainless steel are stacked. Rare earth materials of or shots are packed in the half volume of R22 at the cold side. The Cryocoolers 11, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2001

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Figure 1. Schematic of 4 K pulse tube refrigerator.

other half side of R22 is Pb shots. The connecting tube between the cold head and valves is 1.5 m. Several separate electromagnetic valves are used for V11-V26 to allow easy changing of the on/off timing for experiments. The compressor is a conventional G-M refrigerator compressor with an input power capacity of 6 kW.

The phase shifter of the first stage is of the “two middle-buffer” type. A set of two buffers B11 and B12, of different intermediate pressure levels, are connected at the hot end of the first pulse tube through on/off valves V13 and V14. V11 and V12 are the high and low pressure valves, respectively, connected to the compressor. The phase shifter of the second stage is of the “middle-buffer with double inlet” type. Another

set of two buffers of different intermediate pressure levels B21 and B22 are connected to the bypass line through on/off valves V23 and V24. Valves V21 and V22 are the high and low pressure valves, respectively, connecting the second stage to the compressor. The bypass line consists of two needle valves N1 and N2 that adjust the direction and rate of DC gas flow in the bypass line. Needle valve N3 with buffer B23 is mainly used to adjust the displacement of the gas at the cold end of the pulse tube PT2. The tube between N1 and V21 or V22 has an orifice-like function. The temperature of the first stage is measured by a PtCo temperature sensor that is mounted at the cold end of the pulse tube PT1. The second stage temperature is measured by a silicon diode that is mounted at the cold end of the second pulse tube PT2. Several pressure gauges are installed on each buffer and the valve side of each connecting tube. Compared to an ordinary two-stage pulse tube refrigerator, there are some advantages for the thermally connected type pulse tube refrigerator. These include: 1) the influence between the first and second stage is small, 2) each stage can be operated at a different pressure or frequency with a different compressor, and 3) the mass flow rate to each of the two refrigerator stages can be considered as half that of an ordinary two-stage pulse tube refrigerator; thus the required valve opening area is smaller. In order to get good performance, a compact and good heat exchanger between the first stage and second stage refrigerator is needed; this is one of the disadvantages. Another concept is to use a gas controlled displacer, as shown in Fig. 2, to replace the double inlet. The movement of the displacer is controlled by valve V25, which is connected to the highpressure line of the compressor, and valve V26, which is connected to the low-pressure line of the compressor. Because the displacer stops the DC gas flow, a needle valve N4 is connected between buffer B22 and the low pressure line of the compressor to introduce a suitable rate of DC gas flow.

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Figure 2. Schematic of 4 K pulse tube refrigerator with displacer.

The function of the displacer is the same as that of the bypass. It supplies gas to the hot end of the pulse tube during the pressure-increasing process, and removes gas from there during the pressuredecreasing process. The displacer can make it easier to control the shape of the PV diagram than with the double inlet. This method is similar to that the displacer is on bypass.3 Figure 3 shows typical valve timing for the first stage refrigerator in Fig. 1. The valve timing of the second stage in Fig. 1 is the same as that for the first stage. The dark bold lines represent the period of time when the valve is open; the thin lines represent when the valve is closed.

Figure 4 shows typical valve timing for the second stage refrigerator in Fig. 2. The meaning of the line weights is the same as for Fig. 3.

Figure 3. Valve timing of two middle-buffer pulse tube refrigerator.

Figure 4. Valve timing of two middle-buffer with displacer pulse tube refrigerator.

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MIDDLE-BUFFER TYPE PULSE TUBE REFRIGERATOR

One or more buffers are connected to the hot end of the pulse tube through on/off valves. The pressure in the buffers is at an intermediate or “middle” level, and it theoretically can not be the high pressure or the low pressure of the compressor. Because the pressure in the buffers is at a middle level, we call this type of pulse tube refrigerator a “middle-buffer pulse tube refrigerator.” The middle-buffer pulse tube refrigerator can be considered as a modification of the active-buffer pulse tube refrigerator,4 if the high pressure buffer and the low pressure buffer in the active-buffer pulse tube refrigerator are eliminated. The advantage of the middle-buffer type is that the theoretical mass flow rate is small compared to the active-buffer type, if the same numbers of buffers are used. In the case of two buffers, the theoretical mass flow rate of the middle-buffer type is about one third of that of the active-buffer type. This is very important for high efficiency at low temperature. In order to explain the working process of the middle-buffer pulse tube refrigerator, a numerical simulation was conducted of the first stage refrigerator of Fig. 1. Here the assumed pressure ratio of the compressor was two. This is the minimum pressure ratio of our conventional GM compressor. Figure 5 is a typical calculated pressure waveform, whereas Fig. 6 is a typical PV diagram. Figure 5 shows that the pressures in buffers B11 and B12 are at a middle level. For comparison, the PV diagrams of the active buffer with two buffers are shown in Fig. 6. It is clearly visible that the displacement of the gas in the middle-buffer type is reduced to nearly one third of the active buffer case. This means the mass flow rate required from the compressor is also de-

creased to one third. Another important point is that the PV diagrams of the middle buffer type at the cold end and at the hot end of the pulse tube are completely separated. But in the case of the

active buffer, they are not separated. This means the shuttle loss in the middle-buffer pulse tube is significantly reduced. Though the mass flow rate is significantly reduced, the calculated efficiency is lower than that of the active-buffer type; this is because the upper left corner and right lower corner of the PV diagram at the cold end of the pulse tube are cut off. Therefore the area of the PV work is smaller.

In the middle buffer pulse tube refrigerator, the swept volume of the gas from the regenerator to the cold end of the pulse tube depends on the void volume of the regenerator, the volume of the pulse tube, the refrigeration temperature, and the pressure-ratio of the compressor. In the temperature range of lower than 10 K, the mass flow rate, which is proportional to the density and swept volume of the gas through the regenerator, becomes too large due to the high density of the helium gas. Therefore, a bypass or a displacer is necessary to decrease the mass flow rate. In this experiment, a two-middle-buffer configuration, combined with either a double inlet or a displacer, was adopted as the phase shifter for the second stage.

Figure 5. Pressure wave.

Figure 6. PV diagram.

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Figure 7. Pressure of the second stage with two middle-buffer and double inlet.

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Figure 8. Cooling capacity of the second 1. Double inlet with HoCu2 2. Double inlet with 3. Double inlet with HoCu2

EXPERIMENT RESULTS In the experiment, needle valves N1, N2, N3, and valve timing of V11-V26 were adjusted to get the highest cooling power.

Figure 7 is a typical pressure waveform of the second stage when the double inlet was used. It is clearly shown that the pressures in the buffers are near a middle level. The pressure waveform of the first stage is the same as in Fig. 7. Figure 8 shows the cooling power versus the temperature of the second stage for the case of or for regenerator R22, respectively. The operating frequency is 2 Hz. When was used, the cooling power was 0.58 W at 4.2 K with an input power of about 5.5 kW. When was used, only 0.4 W at 4.2 K was achieved with 6 kW of input power. When the displacer was used, 0.45 W at 4.2 K was achieved with 5.5 kW of input power. In the case where was used, the achieved temperature decreased to 2.35 K with a smaller opening of needle valve N3. CONCLUSIONS

A thermally connected 4 K pulse tube refrigerator has been manufactured and tested. A “two middle-buffer” type phase shifter was used for the first stage, and a “two middle-buffer with double inlet” or displacer was used for the second stage. A cooling capacity of 0.58 W at 4.2 K was achieved using a 1.5-meter-long connecting tube, 5.5 kW of input power, and the two middlebuffer with double inlet type phase shifter.

REFERENCES 1. Wang, C., Thummes, G., and Heiden, C., “Performance Study on a Two-Stage 4 K Pulse Tube Refrigerator,” Advances in Cryogenic Engineering, Vol. 43 (1998), pp. 2055-2062. 2.

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.

3. Zhou, S.L., Thummes, G., and Matsubara, Y., “Experimental Investigation of Loss Mechanism in a 4 K Pulse Tube,” to be published in Advances in Cryogenic Engineering. 4.

Zhu, S.W., Kakimi, Y., and Matsubara, Y., “Investigation of Active-buffer Pulse Tube Refrigerator,” Cryogenics, Vol. 37, No. 8 (1997), pp. 461-471.

Compressor-Specific Design of a Single Stage Pulse Tube Refrigerator J. M. Pfotenhauer and J.H. Baik

University of Wisconsin - Madison Madison, WI USA 53706

ABSTRACT A single stage active valve pulse tube refrigerator has been designed to operate at 30 K and provide a nominal cooling power in excess of 30 watts. This report details the various

considerations comprising the system design, focusing on the limitations imposed by the reciprocating-type compressor commonly used for GM-style pulse tube refrigerators. We describe a design method for GM-style pulse tubes to maximize the pulse tube cooling power that can be produced from a compressor of fixed capacity. The method provides a physical understanding of the various influencing factors, and is illustrated using the specifications for the Cryomech CP640 compressor, which draws a maximum electrical power of 5.5 kW. The design process begins by defining mass flow and compressor work as a function of the discharge and suction pressures, thereby producing a compressor performance map. The compressor map in turn provides a framework from which the pulse tube system geometry can be optimized for

maximum cooling power. Various real constraints, such as pressure drop through the valves and regenerator, laminar boundary layer along the pulse tube walls, and conduction losses are included in the design process and are shown to significantly impact the optimized result. INTRODUCTION

In the process of designing a GM-type pulse tube refrigerator, one must select a compressor from a finite set of fixed compressor capacities - deliverable work is not available as

a continuum. In view of this constraint it is relevant for a designer to ask, "what pulse tube geometry will best utilize the available power from a specific compressor?" - that is, which geometry will produce the highest cooling power for the fixed limitations of the compressor? To date, this question has been answered largely by imperical means. For example, Fujimato et.al.1 varied the length of their pulse tube, finding that the cooling power increased with the length, but that "the compressor power was too large for the pulse tubes used in [their] tests." Wang, Thummes & Heiden2 found that the cooling capacity of their pulse tubes decreased as the diameters (and volumes) were increased. An exception to the imperical approach is presented by Ravex et.al.3 who confirm the ability of their numerical model to optimize their pulse tube geometry for a specific compressor. The present report follows in this non-imperical direction, Cryocoolers 11, edited by R.G. Ross, Jr.

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but provides a general procedure that enhances the physical understanding of various constraints

along the way. The following paragraphs outline a procedure for designing a GM-type pulse tube cryocooler that will provide the maximum possible cooling capacity for a fixed compressor. The procedure identifies the constraints imposed by the compressor, as well as those associated with the pulse tube and regenerator geometries. The process begins by defining a compressor map in terms of mass flow, compressor work, and the inlet and outlet pressures. Practical considerations regarding the pulse tube volume and valve system determine the attainable range on the compressor map, and an iterative method is presented for optimizing the geometry of the pulse tube refrigerator. The specific characteristics of a Cryomech model CP640 compressor are used to illustrate the compressor constraints. In the iterative process, we use the thermodynamic model of Yuan & Pfotenhauer4 to calculate the geometry dependent cooling capacity of a 5-valve pulse tube. However, in the case of other pulse tube configurations, alternate methods for calculating the cooling power as a function of the pulse tube volume could be used instead. COMPRESSOR CHARACTERISTICS

The mass flow delivered by a reciprocating compressor can be expressed (see for example references 5-7) in terms of the displaced volume, compressor speed S, clearance volume ratio the inlet and discharge pressures compression exponent n according to the Eq.:

Here

and the polytropic

is the clearance volume, and the polytropic exponent n is associated with the expression

that describes the relationship between pressure and volume during the compression and

expansion processes. The possible values of n are bounded by (minimum) for an isothermal compression process, and (maximum) for an adiabatic compression process. Here and are the specific heat at constant pressure and constant volume respectively. The expression given in Eq. (1) results from considerations regarding the volumetric efficiency of a reciprocating compressor, and has been extensively verified in a recent investigation6 of commercial compressors used in the refrigeration industry. An expression for the electric power consumed by the compressors is also provided in the same study, and is given as:

Here

is the combined efficiency of converting electrical to mechanical and mechanical to PV work, which may be roughly given as for a wide variety of compressors. Although the values for the clearance volume and the polytropic exponent that are required in Eq.s (1) and (3) are not readily available from the commercial vendors, these can be obtained from a few performance characterization measurements for a compressor of interest. Thus it is possible to accurately define the mass flow that will be delivered by a compressor for a given set of inlet and outlet conditions.

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251

The compressor map defined by Eq.s (1) and (3) for the Cryomech CP640 is shown in Fig. 1. Using our measurements of mass flow for various inlet and outlet pressures, and an inlet temperature of 315 K, we determined values of and for the clearance volume ratio and polytropic exponent respectively. The sensitivity of the mass flow rate to and its insensitivity to can be understood in terms of the expression

Except through the relatively weak dependence of the volumetric efficiency on the outlet pressure, the mass flow is entirely determined by the inlet conditions. The compressor work, on the other hand is dependent on both and and is maximized at the higher values of Constraints imposed by the compressor define three limitations to mass flow that can be identified on the compressor map. A minimum inlet (suction) pressure and a maximum outlet (discharge) pressure define the first two of these. For the Cryomech CP640 these are respectively defined by and The third constraint is unlikely to be encountered for a pulse tube system, but is defined by the condition OPTIMIZATION DESIGN PROCESS

Within the constraints imposed by the compressor, Fig. 1 displays a wide range of mass

flow rates available through many combinations of inlet and outlet pressures. The cooling power of pulse tube refrigerators is known to depend on all three of the parameters and mass flow rate. It is therefore of interest to determine both the region of the compressor map that affords the largest pulse tube cooling capacities, and which regions are accessible in real systems. The first of these questions is answered for the case of a 5-valve pulse tube using an iterative process described by the flow diagram in Fig. 2.

Figure 1. Compressor map for Cryomech CP640 reciprocating compressor.

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GM-TYPE PULSE TUBE COOLERS FOR LOW TEMPERATURES

Figure 2. Flow diagram of iterative design process to optimize a pulse tube system geometry for maximum cooling power given specific values of and mass flow rate.

The iterative design process begins by considering pressure losses across the inlet and outlet valves connecting the pulse tube system to the compressor. In view of the energy losses

directly associated with such a pressure drop, it is important to minimize the flow resistance through these valves. For all valved (GM-type) pulse tube systems, the pressure swings experienced in the pulse tube system will be less than those produced by the compressor. In Fig. 2 above, and represent the high and low pressures realized in the pulse tube system respectively. The next step requires a selection of the pulse tube operating, or cold end, temperature. This selection is of course not arbitrary and must be based on reasonable limits for a single stage pulse tube cooler appropriate selection of regenerator matrix material, and admittedly some previous experience. In the examples to follow, we have used as a cold end temperature. With the values of and mass flow rate through the pulse tube fixed, one might expect that the pulse tube volume would be fixed. In fact, an iteration process is required to determine the combination of pulse tube and regenerator gas volumes associated with the fixed

conditions that maximize the resulting cooling power. The ratio of the pulse tube volume to the gas (or non-solid) volume in the regenerator characterizes the competing effects of expansion

COMPRESSOR-SPECIFIC DESIGN OF A 1-STAGE PT COOLER

volume and pressure loss: Large values of

253

maximize the expansion volume in the pulse tube,

but produce large pressure drops through the regenerator. Small values of minimize the pressure loss through the regenerator but sacrifice expansion cooling in the pulse tube. In addition to optimizing the cooling power as a function of the length-to-diameter aspect ratio of both the pulse tube and regenerator can be optimized for each value of and mass

flow rate. For the case of the pulse tube, the aspect ratio is chosen as large as possible to minimize conduction losses through the walls, but not so large that the boundary layer velocity becomes turbulent. The laminar to turbulent transition is defined8 by the condition that the Reynolds number be less than 280. The characteristic dimension in the Reynolds number is defined by the Stokes boundary layer thickness

where is the kinematic viscosity and is times the frequency. The laminar condition constrains the cross sectional area of the pulse tube to be larger than a minimum defined by

For the regenerator, REGEN29 is used to balance the conduction and regenerator ineffectiveness against the pressure drop in through the regenerator. In our designs, we have selected a maximum allowable pressure drop of 50 kPa. Fig. 3 displays the cooling power, pulse tube volume, and regenerator gas volume as a function of for one set of and mass flow rate values. In this case, a decreasingly significant benefit is realized by increasing beyond 8. In the interest of a compact design, this value of is chosen for the optimized design.

Figure 3. Geometry optimization of pulse tube and regenerator volumes for single set of and mass flow rate values.

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GM-TYPE PULSE TUBE COOLERS FOR LOW TEMPERATURES

Figure 4. Pulse tube cooling capacities as a function of compressor characteristics.

The same process described in the previous paragraph, carried out for a variety of inlet

and outlet compressor pressures, results in the cooling powers shown in Fig. 4. At low values of the outlet pressure, a weak pressure wave results in the pulse tube and increased values of inlet pressure reduce the pressure oscillations and therefore the cooling power. At high values of outlet pressure, the dependence of the cooling capacity Q on the inlet pressure is reversed. In this case, increased inlet pressures produce an increased density at the compressor inlet, a larger mass flow rate, and a larger cooling power. It is of obvious interest to know which parts of this map are accessible for a real pulse tube system, and what steps can be taken to reach the region of both high outlet and high inlet pressures.

Pulse Tube System Options From the results displayed in Fig. 4 it is clear that the maximum outlet pressure permitted by the compressor corresponds closely with its rated power capacity. Operating at is a necessary, but not sufficient condition for maximizing the cooling power of a pulse tube system

attached to the compressor. It is also desirable to maximize

as well. How can this be

achieved? Beginning with the case of an ideal pulse tube system - that is, one with no losses - one can consider the influence of pulse tube size on the pressures in the pulse tube system. Fig. 5 depicts the high and low pressures that will result in a pulse tube system as a function of the charging pressure, for a small volume and a large volume pulse tube system. The maximum and minimum charging pressures in the pulse tube system are constrained by the compressor's maximum discharge and minimum suction pressures respectively. A small volume pulse tube

system will realize large pressure oscillations, while the large volume will realize small pressure oscillations. The arrows, representing the pressure swing realized in each pulse tube system when it is charged to its maximum allowable pressure, reveal that although the pressure oscillations will decrease as the pulse tube volume is increased, for a constant

the inlet

pressure will increase with pulse tube volume. This intuitive consideration is confirmed by the

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255

Figure 5. Depiction of pressures realized in pulse tubes as a function of the charging (average) pressure, and as limited by the minimum and maximum pressures provided by the compressor.

Figure 6. Ideal cooling power vs. the optimized pulse tube volume for the various combinations of high and low pressures produced by the Cryomech CP640 compressor.

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GM-TYPE PULSE TUBE COOLERS FOR LOW TEMPERATURES

calculations displayed in Fig. 6. Here the same cooling powers as calculated in Fig. 4 are shown Q clearly

with their corresponding optimized pulse tube volumes. For the highest values of increases as the pulse tube volume is increased. Real System Constraints

Based on the calculations displayed in Fig. 6, one may wonder whether an upper limit to the pulse tube system volume exists. In a real pulse tube system, losses associated with the pressure drop through the valves and regenerator, and conduction losses (as the pulse tube volume increases, the optimized aspect ratio decreases) impose practical limits and define a different optimum pulse tube volume for maximum cooling power. Repeating the same procedure that resulted in Figs. 4 and 6, but including the losses realized in our valve system, the 50 kPa pressure drop through the regenerator, and conduction losses through the pulse tube and

regenerator structure, provides the results shown in Fig. 7. The optimum pulse tube volume falls between 200 and 300 with a associated values of and pulse tube aspect ratio of 8. For this design, a cooling power of 60 watts is expected at 30 K. Two additional losses have yet to be included in our model; those associated with shuttle heat loss, and DC flows. We expect therefore that the actual cooling power will be less than 60 watts. The significance of valve losses are clearly evident through this design process. We are presently pursuing an improved

valve arrangement for our own pulse tube system.

The results provided in Figs. 4, 6, and 7 also permit an estimate of efficiency for GM style pulse tube refrigerators. For the ideal case (ignoring losses) depicted in Fig. 4, one finds

that the efficiency associated with the maximum possible cooling power of a pulse tube operating at 30 K and driven by the CP640 compressor is 0.03, or 28% of Carnot. A more realistic value, including conduction losses and pressure drop through the regenerator is 0.015, or 14% of Carnot.

Figure 7. Cooling power vs. optimized pulse tube volume. The optimization process accounts for pressure losses through valves and regenerator, and conduction losses.

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Applying the above procedure for the CTI 8500 compressor used by Ravex et.al.3 we find that an optimum pulse tube volume for operation at 30 K would be approximately three times larger than what they used but that the aspect ratios of the pulse tube and regenerator would be similar. Furthermore an optimum value of and a cooling power of 82 watts is found for that case. CONCLUSION A design method has been described that allows one to optimize the cooling power of a GM-style pulse tube refrigerator for a given fixed compressor. The method produces a compressor performance map and defines an iterative procedure for optimizing the cooling power of the pulse tube refrigerator as a function of its volume and geometry. Considerations regarding the volume of an ideal pulse tube suggest that the volume should be made as large a possible. However, real losses such as pressure drop through the pulse tube and conduction losses significantly influence the process and allow one to define an optimum volume and geometry. The efficiency of a GM-style pulse tube operating at 30 K is limited to less than 15% of the Carnot efficiency. ACKNOWLEDGEMENT

This work is supported by DOE through Lockheed Martin subcontract DE-AC05-960 OR22464. REFERENCES 1. S. Fujimoto, Y.M. Kang, Y. Matsubara, "Development of a 5 to 20 W at 80 K GM Pulse Tube Cooler," Cryocoolers 10, Plenum Press, New York (1999), pp. 213-220. 2. C. Wang, G. Thummes, C. Heiden, "Performance Study on a Two-Stage 4 K Pulse Tube Cooler," Advances in Cryogenic Engineering vol. 43, Plenum Press, New York (1998), pp. 2055-2062.

3. A. Ravex, J.M. Poncet, I. Charles, and P. Bleuze, "Development of Low Frequency Pulse Tube Refrigerators," Advances in Cryogenic Engineering vol.43, Plenum Press, New York (1990), pp. 1957-1964.

4. J. Yuan and J.M. Pfotenhauer, “Thermodynamic Analysis of Active Valve Pulse Tube Refrigerators,” Cryogenics vol. 39, (1999), pp. 283-292.

5. P. Popovic and H.N. Shapiro, "A Semi-empirical Method for Modeling a Reciprocating Compressor in Refrigeration Systems," ASHRAE Transactions vol. 101 (2), (1995), pp. 367-382. 6. D. Jaehnig, "A Semi-Empirical Method for Modeling Reciprocating Compressors in Residential

Refrigerators and Freezers," MS Thesis, University of Wisconsin - Madison, (1999). 7. D. Jaehnig, S.A. Klein, and D.T. Reindl, "A Semi-Empirical Method for Representing Domestic Refrigerator/Freezer Compressor Calorimeter Test Data.", ASHRAE Transactions, (June, 2000), Minneapolis, MN

8. R. Akhavan, R.D. Kamm, and A.H. Shapiro, "An Investigation of Transition to Turbulence in Bounded Oscillatory Stokes Flows - Part 1. Experiments," J. Fluid Mech. vol. 225, (1991), pp. 395422.

9. V. Arp and R. Radebaugh, "Interactive Program for Microcomputers to Calculate the Optimum Regenerator Geometry for Cryocoolers (REGEN and REGEN2)," AFWAL-TR-87-3040 Wright-

Patterson Air Force Base, (1987).

A Novel Multi-Stage Expander Concept C. S. Kirkconnell, K. D. Price, M. C. Barr, and J. T. Russo Raytheon Electronic Systems El Segundo, California, USA 90245

ABSTRACT Raytheon has developed a novel two-stage expander for use in long life, high reliability cryocoolers for space and commercial applications. The expander is classified as a Stirling machine and requires a conventional reciprocating piston compressor to drive it. The key feature is a new method for obtaining and controlling expansion at the two stages. Thermodynamic efficiency is higher than existing one and two stage coolers and the mechanical implementation is as simple or simpler. The expander device is described in both thermodynamic and mechanical terms and performance predictions given. INTRODUCTION Raytheon has developed a hybrid Stirling-pulse tube two-stage expander module for use in long life, high reliability space applications. In this design, the first stage is a traditional “Oxford” Stirling flexure suspended expander and the second stage is a U-tube pulse tube. Refrigerant gas flows through the Stirling first stage to the pulse tube second stage. The warm ends of the pulse tube, its regenerator, and the orifice/surge volume are thermally anchored to the Stirling stage. The same compressor powers both stages. The hybrid design offers significant advantages in thermodynamic performance, power efficiency, ease of construction, reliability, cost, size, and weight compared to existing cryocooler technologies. In the following sections, the thermodynamic benefits of the new expander are discussed in comparison to existing technologies, the mechanical design is described, and performance predictions are given. THERMODYNAMIC BENEFITS OF NEW APPROACH The hybrid expander achieves high efficiency by combining the best features of Stirling and pulse tube coolers while eliminating or attenuating their inefficient features. The following briefly describes the key advantages and disadvantages of Stirling and pulse tube one-stage expanders. This will be followed by an explanation of how the hybrid two-stage design advantageously combines these features to produce a high capacity, high efficiency cryocooler. Stirling expanders achieve high efficiency in part because of the relatively low gas volume that must be cyclically pressurized and depressurized. This minimizes the mass flow rate of gas into and out of the expander relative to a comparable pulse tube cooler. The low flow rate enables better optimization of regenerator parameters, reduces pressure drop losses through the machine, and reduces compressor swept volume compared to a pulse tube. A limiting factor in Stirling performance is the complexity required to obtain high heat transfer efficiency at the cold end. A Stirling Cryocoolers 11, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2001

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Figure 1. Stirling-Pulse Tube Hybrid Expander Concept. The two-stage hybrid expander simply attaches a pulse tube expander stage to the cold end of a Stirling expander. The schematic shows the basic configurations of one-stage coolers and the two-stage hybrid cooler. In the two-stage cooler, gas flows through the Stirling stage to the pulse tube stage as in conventional Stirling class multi-stage cryocoolers.

The schematic shows a linear pulse tube, but the Raytheon hybrid configures the pulse tube stage as a U-tube for compactness and ease of integration.

with regenerator integrated within the piston has relatively poor heat transfer at the cold end because there is no flow-through heat exchanger from the regenerator to the cold thermal interface. The gas simply impinges at relatively low speed on the internal surfaces. To achieve high heat transfer, the regenerator must be external to the piston (e.g., fixed to non-moving structure) so that gas flowing into the cold expansion volume can pass through an efficient heat exchanger connected to the thermal load. Mechanical arrangements to implement this are relatively complex and the structures tend to have higher conductive parasitic loads that offset the benefit of improved heat transfer. Pulse tubes achieve good performance partially because of the high-efficiency flow-through heat exchangers present throughout the cooler, but suffer because the relatively large working volume draws high gas mass flow rates. The high flow rates result in suboptimal regenerator parameters that trade heat transfer efficiency against pressure drop losses, and the larger working volume increases the size of the compressor required to drive the expander. When a Stirling first stage is combined with a pulse tube second stage, the gas flow rate through the expander is reduced relative to an all pulse tube design because the Stirling stage has a smaller gas volume compared to a pulse tube. This reduces the size of the pressure drop losses in the expander. Furthermore, the gas volume is greatly reduced compared to an all pulse tube machine, reducing the size of the requisite compressor. The Stirling stage receives a high efficiency, flow-through heat exchanger because one can be inserted in the gas flow path from the Stirling expansion volume into the pulse tube regenerator.

Furthermore, the Stirling stage can retain the regenerator integrated within the piston, which is beneficial for manufacturing and reduces parasitic loads relative to an external regenerator design.

Finally, the complexity of the Stirling stage is no greater than the complexity of well-developed single-stage Stirling coolers. For example, a tight tolerance, cryogenic clearance seal is not required to separate the two stages. Since the pulse tube stage is entirely at cryogenic temperature, pressure drop losses within the

stage are naturally low. Reduced pressure drop improves regenerator design parameters resulting in more efficient regenerator heat transfer, maximum pulse tube work-of-expansion, and low thermal conduction losses.

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The hybridized expander module is no more difficult to manufacture than a one-stage Stirling yet delivers true two-stage cryocooler performance. The second stage is entirely passive, yielding a mechanically simple design with very few tight tolerance features. The angular orientation of the second stage with respect to the first stage is arbitrary. Like a two-stage pulse tube but distinct from a two-stage Stirling, the absence of a moving piston in the second stage allows the user to orient the stages for optimum integration at the system level. In-line (0°), a 90° bend, or any intermediate

angle between 0° and 90° are all possible. The required compressor swept volume is significantly lower than would be required for either a one or two-stage pulse tube of comparable heat lift and is

not significantly larger than for a one or two-stage Stirling expander. The expansion process in each stage is controlled by a separate mechanism that gives the hybrid an unusually useful feature: refrigeration power can be allocated and reallocated between the two

stages on command and continuously in real time. The Stirling first stage piston is motor controlled, so its motion can be altered on command. A fixed orifice and surge volume controls the pulse tube expansion. Analysis shows that modifying the first stage piston motion will alter gas flow rates, phase angles, and pressure ratios in both stages resulting in the reallocation of refrigerating power between the stages on command. Thus, refrigerating power can be biased to increase heat lift at the second stage at the expense of first stage heat lift and vice versa. This functionality is not available from typical single-piston, two-stage Stirling designs because the expansion phase angles in both cryogenic stages are directly, mechanically linked. Two-stage pulse tube cryocoolers have no active phase control, so dynamic load shifting is not possible. The partial coupling of the first and second stage expansion phase angles and the capability to shift the refrigeration allocation on the fly are by-products of the unique thermodynamic performance characteristics provided by the hybrid expander design. This capability can be used to extend the performance range of a single cryocooler to different combinations of first and second stage temperature and heat load requirements or the heat lift distribution between stages can be modified during operation to accommodate

fluctuating heat loads. STIRLING-PULSE TUBE EXPANDER DESIGN

The Hybrid Expander, shown in Figure 2, is composed of a small U-tube pulse tube second stage with surge tank attached onto a first stage consisting of a moving Stirling piston assembly,

which is essentially a resized version of the Raytheon PSC piston assembly.1 The first-stage Stirling piston is supported on a suspension of 12-finger spiral flexures to provide long-life, non-contacting operation. The piston is driven by a compact linear motor. A like suspension / motor arrangement

Figure 2. Raytheon Protoflight Hybrid Expander. Second stage oriented at 45° with respect to the first stage for protoflight. Design utilizes standard Raytheon moving displacer in first stage and U-tube

pulse tube for second stage.

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is packaged to drive a countermass to null the vibration output from the expander. As in all Raytheon Stirling cryocoolers, two close-fit, non-contacting clearance seals are incorporated at the warm end of the moving first stage piston: the “plenum” seal and the “regenerator” seal. The plenum seal isolates the working gas volume from the non-working plenum volume where the flexure suspension system and control motor and balance mass reside. The regenerator seal mini-

mizes leakage of gas around the regenerator to port working gas through the regenerator. From the Stirling expansion space, the first-stage gas is ported into a copper mesh heat exchanger housed within the first-stage heat interface block, and this same block houses the pulse tube surge volume. The gas ports through a 90° turn as it enters the second-stage regenerator, the purpose being to demonstrate the hybrid’s insensitivity to relative angle between the stages. (A 45 ° orientation has been selected for protoflight, which is the configuration depicted in Figure 2.) The gas continues through the pulse tube regenerator out to the second stage heat interface block where the second-stage expansion occurs. At this block the gas is turned 180° and sent through another copper mesh heat exchanger, which also serves as a flow straightener for the cold end of the pulse tube. The gas enters the pulse tube from the heat exchanger on its way back to the first stage heat interface block where it passes through yet another copper mesh heat exchanger / flow straightener. Here it passes through a flow restrictor and is ported into the surge volume. The containment housing at the warm end is made from aluminum. The first-stage thin-walled cold cylinder is stainless steel to minimize parasitic heat loss. The first- and second-stage heat interface blocks are made from copper and the second stage regenerator and pulse tubes are made from inconel, also selected to minimize conduction parasitics. A variety of welds and braze joints are used to join together these disparate materials. The complete cryocooler (compressor + expander) can be made with all welded or brazed joints except for one mechanical joint at the transfer line. Warm end heat rejection, structural mounting and vacuum sealing are interfaced to the aluminum warm end housing. Electrical feedthrough headers are welded onto steel bosses on the aluminum housing. In total, the expander weighs about 3.2 kilograms. DISCUSSION Performance Predictions Thermodynamic models have been well correlated at Raytheon for both Stirling cryocoolers (SSC2, ISSC, PSC1, SBIRS Low) and Pulse Tube cryocoolers (35K IR&D3, Mini IR&D4, 4-Tube5, LCC6). A numerical model of the Hybrid Expander was developed based upon these building blocks, a fundamental purpose of the breadboard expander presently in assembly being to correlate

the new math model. Of particular interest are those portions of the model that represent features unique to the Hybrid Expander, not relevant to either single-stage model. Examples are the presence of a flow-through heat exchanger at the cold end of the Stirling and the use of a cryogenic surge tank for the pulse tube stage. The maturity of the models from which the hybrid model was constructed provides confidence that reasonably accurate performance predictions can be obtained for the breadboard unit. In the areas of modeling uncertainty such as those noted above, conservative assumptions were used. The nominal design point for the breadboard expander based upon the existing model is 5.6 watts at 95 K and 5.2 watts at 145 K for 75 watts input P-V power, or about 100 W input motor power for a typical Stirling-class reciprocating compressor, at a heat rejection temperature of 305 K. Extensive characterization of the breadboard expander is planned over a wide trade space of temperatures, loads, frequencies, charge pressures, and phase angles. Experimental data will be provided at such time that it becomes available. Thermodynamic Efficiency Benefits of Hybrid Expander

The concept of “specific power,” the ratio of input power to net refrigeration capacity, has become well established in the characterization of single-stage cryocoolers as providing an informative representation of a cooler’s efficiency. In a paper presently in work by the authors7, a methodology for normalizing the performance of multi-stage refrigerators in terms of a common

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“pseudo single stage” efficiency is presented. The total capacity at the various stages is normalized to a single capacity at the coldest stage by use of the following equation:

In the above expression, is the Carnot COP based upon the refrigeration temperature at each stage and the cryocooler’s warm rejection temperature, typically at or near ambient. From Eq. (1), the normalized refrigeration capacity for the breadboard hybrid expander, assuming a rejection temperature of 305 K, is 8.2 watts at 95 K. Continuing with the single-stage analogy, the specific power for the breadboard hybrid cooler, based on 100 watts motor power, is 12.2 W/W at 95 K. Preliminary predictions indicate that hybrid designs tuned for operation in this temperature range should be readily capable of specific power efficiencies of 10 W/W at 95 K, with the refrigeration capacities split between the first and second stages as desired by the user. CONCLUSION

A novel two-stage Stirling-Pulse Tube hybrid expander has been developed by Raytheon and is in final assembly as of the writing of this paper. The hybrid expander takes advantage of the strengths of the individual expander designs upon which it is based. Using a Stirling expander first stage reduces the large pressure drop loss characteristic of both single-stage and multi-stage pulse

tube expanders. The mechanical complexity of a cryogenic clearance seal is avoided entirely by using a pulse tube second stage, and this enhances the reliability of the expander by eliminating a potential single point failure mode. The actively controlled Stirling displacer in the hybrid provides the unique capability to shift refrigeration capacity between the stages, a feature not available to multi-stage pulse tubes or multi-stage Stirlings with single displacer pistons. The projected performance for the breadboard hybrid expander unit presently being assembled is 5.6 W at 95 K and 5.2 W at 145 K for 100 W input power. REFERENCES 1. Price, K.D., Barr, M.C. and Kramer, G., “Prototype Spacecraft Cryocooler Progress,” Cryocoolers 9, Plenum Press, New York (1997), pp. 29-34. 2. 65 K Standard Spacecraft Cryocooler Program Final Report, Contract #F29601-89-C-0082, Hughes Aircraft Company, Electro-Optical Systems, El Segundo, CA; November 1995. 3. Soloski, S.C. and Mastrup, F.N., “Experimental Investigation of a Linear Orifice Pulse Tube Expander,” Cryocoolers 8, Plenum Press, New York (1995), pp. 321-328. 4. Kirkconnell, C.S., Soloski, S.C. and Price, K.D., “Experiments on the Effects of Pulse Tube Geometry on PTR Performance,” Cryocoolers 9, Plenum Press, New York (1997), pp. 285-293. 5. Kirkconnell, C.S., “Experimental Investigation of a Unique Pulse Tube Expander Design,” Cryocoolers 10, Plenum Press, New York (1999), pp. 239-247. 6. Kirkconnell, C.S., “Experiments on the Thermodynamic Performance of a ‘U-Tube’ Pulse Tube Expander,” Advances in Cryogenic Engineering, vol. 43 (1998), pp. 1973-1980. 7. Kirkconnell, C.S. and Price, K.D., “Thermodynamic Optimization of Multi-Stage Cryocoolers,” Cryocoolers 11, Plenum Press, New York (2001).

Numerical Study of a New Type of 4 K GM/PT Hybrid Refrigerator Liqiang Liu, Linghui Gong, Jingtao Liang and Liang Zhang

Cryogenic Laboratory Chinese Academy of Sciences Beijing 100080, China

ABSTRACT

A new type of hybrid refrigerator operating in the liquid helium temperature region has been proposed. The warm stage of this refrigerator is a typical G-M refrigeration cycle, on which the

cold stage of a pulse tube cooler is coupled thermodynamically. The phase shift structures for the pulse tube are supported at the temperature of the G-M refrigeration stage. There are several methods that can be applied as the phase shifter for the pulse tube; for example, a cold piston that is

connected to the displacer of the G-M refrigerator. Physical and mathematical models have been established to describe the unique thermophysical aspects of this type of refrigerator, and numerical methods have been employed to solve the theoretical models. Through the calculations, some helpful results have been obtained, and the structure of the refrigerator has been optimized. Experimental evaluation is now underway. INTRODUCTION

In recent years, owing to the introduction of magnetic regenerator materials, the low temperature performance of G-M refrigerators has been improved, particularly at 4.2K, and other significant performance enhancements have been achieved.1 However, due to the seal rings having to operate in the low temperature environment in a G-M refrigerator, the life and the stability of the refrigerator are seriously compromised; furthermore, the structure of its cold stage is complex. On the other hand, substantial performance improvements have been recently achieved in pulse tube coolers through the introduction of an orifice and double inlet as well as other phase shifters.2 Whereas the cooling power of a pulse tube cooler may be less than that of a comparable G-M refrigerator, its cooling temperature with two stages can reach below 4.2 K. Also, the structure of its cold stage is very simple, as the seal rings are eliminated. Based on theoretical considerations, a new type of 4K refrigerator has been proposed in which the warm stage is a typical G-M refrigeration cycle, onto which the cold stage of the pulse tube cooler is coupled. This new type of refrigerator can overcome the disadvantages of G-M refrigerators, and has promise for future applications that require somewhat lower cooling power (e.g., W at 4.2 K). At present, a prototype of this new type of refrigerator has been manufactured, and experimental measurements are under way. In this paper, a numerical analysis and optimization

calculations are presented. Cryocoolers 11, edited by R.G. Ross. Jr.

Kluwer Academic/Plenum Publishers, 2001

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Figure 1. Brief structure of the new type of 4K refrigerator.

STRUCTURE OF THE NEW TYPE OF REFRIGERATOR As shown in Fig. 1, the new type of refrigerator has two stages. The warm stage of the refrigerator is a typical G-M refrigeration cycle; it is coupled by a pulse tube to the cold stage. With the goal of enhancing the ability to control the pulse tube phase, and building upon the phase shifting technique of the G-M refrigerator, a cold piston connected to the G-M displacer has been introduced as the phase shifter for the pulse tube. Besides this mode of shifting the pulse tube phase by a cold piston, other styles of phase shifters, such as an orifice or double inlet, can be conveniently introduced to improve the pulse tube's performance. Through skillful design, these other phase shifting concepts can be used either independently, or in combination with the cold piston. As a result, some new methods of shifting phase are introduced, and opportunities for improving the performance of the refrigerator are increased. From Fig. 1 one can see that the new refrigerator is not just a pulse tube cooler pre-cooled by a G-M refrigerator. In this new concept, all gas passages are linked, and the pressure, temperature, and mass flow oscillations share the same frequency. Therefore, analytical models of either G-M refrigerators or pulse tube coolers by themselves are not capable of explaining the working mechanisms of the new refrigerator; a new theoretical model must be built. In the present study, a numerical model for simulating the dynamic performance and characteristics of the oscillating flow in the new refrigerator was developed to analyze its performance and optimize its structure. PHYSICAL MODEL AND GOVERNING EQUATIONS Physical Model

Figure 2 shows the physical model of the new refrigerator that is the subject of the study. The focus is on the low temperature stage regenerator and pulse tube. In the model, the junction between the cold chamber of the upper stage and the low temperature regenerator is taken as the left boundary, and the inner surface of the pulse tube phase-shifter piston is taken as the right boundary. Each method of shifting phase can be embodied by dealing with the right boundary properly. For example, when the cold piston and the orifice phase shifters are both used, the right boundary is movable, and the width of the grid to the left of it is varied accordingly; in addition, the gas velocity through this boundary u, is determined by the orifice and

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267

Figure 2. Schematic of the physical model for the new type of refrigerator.

reservoir. Otherwise, the length of the grid left to the right boundary is constant if the cold piston is not employed, and is zero if the orifice phase shifter is not applied. The basic assumptions in the model are as follows: 1. one-dimensional laminar compressible flow of 2. constant wall temperature of heat exchangers at cold and warm end of the pulse tube; 3. axial heat conduction neglected; 4. pressure drop in the regenerator also neglected Governing Equations

The governing equations for the numerical study are given as follows, where the gas velocity, u, is defined as positive if the gas flow is from left to right and as negative for the opposite flow. Continuity equation for the gas:

where

is the density of the gas. Energy equation for the gas:

where is the heat transfer area per unit volume of the matrix, h is the specific enthalpy of the gas, is the temperature of the gas, is the temperature of the matrix, a is the coefficient of heat transfer, which is taken from the reference 3. Energy equation for the regenerative materials:

Equation of state for the real gas of helium:

Boundary Conditions

The gas temperature at the right boundary is as follows:

where is the gas temperature in reservoir, which is taken to be the constant, is also taken as the wall temperature of the heat exchanger at warm end. Meanwhile the wall temperature of the heat exchanger at cold end is taken to be another constant, The process in the reservoir is regarded as isothermal and isobaric. The gas velocity through the right boundary (this velocity is zero if the orifice phase shifter is not employed) is determined by a formula for a nozzle with a correction factor as follows:

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where denotes the pressure in the reservoir and is the correction factor for the geometry of the nozzle. The motion of the piston for phase shift is given by

where is the crank angle, is the phase difference between the pressure wave and the movement of the piston, is the stroke, is the dead gap, which generally is 0.5 mm. The gas temperature at the left boundary is as follows:

The gas inlet temperature is maintained constant and the outlet temperature depends on the efficiency of the regenerator. For the left boundary, the pressure oscillation is taken as input data. The pressure varies with as where

and

denote the maximum and the minimum of the pressure oscillation respectively.

NUMERICAL METHODS AND PROCEDURE

To carry out the numerical simulation, discretization procedures were employed. The governing equations were discretized using a control-volume approach with the internal node method (the length of the last spatial grid, from left to right, was variable if the cold piston was used as the phase shifter). To reduce the number of spatial grids, a nonuniform grid was used. An implicit approach for time items in the governing equations and the upwind one-order scheme were applied to achieve

calculational stability. The staggered grid approach, which uses control volumes for the velocities that are staggered with respect to those for temperature and pressure, was also adopted to eliminate erroneous pressure profiles. The governing equations were solved with the under-relaxation iteration method. In particular, it should be noted that during the solution, close attention was paid to a special case of zero gas velocity. This case sometimes occurred along the spatial grid when the direction of gas velocity reverses at a certain place; when this special case occurs, the iteration procedure is lightly divergent. An efficient method to resolve the problem is to employ the upwind scheme strictly at every spatial grid, especially at the grid with zero gas velocity. COMPUTATIONAL RESULTS AND DISCUSSION The numerical analysis was first applied to the new refrigerator with the phase shifting accom-

plished merely by the cold piston (abbreviated ‘Pis.’); next it was applied assuming both the cold piston and the orifice (abbr. Pis.-Hol.). For comparison, the performance of the orifice-only version of the pulse tube cooler (abbr. Hol.) is also presented using the above numerical analysis. Finally, the optimized configuration of the new refrigerator is presented. If no special explanation is given, the parameters used are as shown in Table 1. The major operating conditions are:

MPa, frequency was

(in the case of the Pis.-Hol. shifter), and the operating (that is, the rotary speed of the refrigerator was 60 rpm).

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269

Cooling Power The temperature dependence of cooling power obtained using the three versions of the phase shifter is shown in Fig. 3. It is obvious that the performance of the refrigerator with only a cold piston as its phase shifter is not as good as expected. Nevertheless, the performance can be improved (the cooling temperature can be decreased more than 4.5 K) in the case where the cold piston is combined with another version of the phase shifter, such as an orifice. A promising future candidate may be to combine the cold piston with a double-inlet phase shifter.

Mass Flow Rate Figures 4 and 5 show the transient mass flow rates at the warm and cold ends of the pulse tube with the three versions of the phase shifter, respectively. The pressure wave is also presented in

these figures to help discriminate the difference in the phases. It is found that the phase difference between the pressure wave and the mass flow rate of the Pis.-Hol. shifter is superior to that of the Hol. or Pis. shifters at the cold end of pulse tube. In contrast, the former is inferior to that of the Hol. shifter at the warm end. This phenomenon predicts that the performance of the refrigerator with the combined phase shifter will be better than that with either of the two individual phase

shifters.

Figure 3.

Cooling power of three versions of shifting phase.

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Figure 4. Mass flow rate at warm end.

Figure 5. Mass flow rate at cold end.

Temperature Oscillations The temperature oscillations at the warm and cold ends of the pulse tube, respectively, are shown in Figs. 6 and 7 for the three versions of the phase shifter. To make the results comparable, the cooling temperature of each curve is the no-load temperature where the cooling power approaches zero. It can be seen that the difference between the average temperature of the gas at the cold end of the pulse tube and the cooling temperature with the Pis.-Hol. or Hol. shifters is much higher than that with the Pis. shifter. The fact that the average temperature of the gas at the cold end of the pulse tube is much lower than the cooling temperature with the Pis.-Hol. or Hol. shifters at low cooling power suggests a serious loss in the regenerator with these two phase shifters. As for

the warm end of the pulse tube, the difference between the average temperature and

for the Pis.-

Hol. or Hol. shifters is also much higher than that with the Pis. shifter; this implies that the Pis.-Hol. and Hol. shifters have a higher capacity for pumping heat.

Optimum Structure From the above numerical analysis we can see that to obtain better performance, the cold piston should be combined with an orifice-type phase shifter. In the following, interest is focused on the new refrigerator with the Pis.-Hol. type phase shifter to determine its optimum structure. For the Pis.-Hol. type phase shifter, Figures 8-11 show the influence on the cooling power at 8.0 K of the original phase difference, the diameter of the orifice, the volume ratio of the piston chamber to pulse tube (keeping the stroke constant and altering the diameter), and the volume ratio of the

regenerator to the pulse tube (keeping the diameter constant and altering the length). In Fig. 11, represents the cooling power per unit gas mass at the warm end of the regenerator. From these

figures, we can determine the optimum shifter for the new refrigerator with the Pis.-Hol. phase shifter, that is: 1) the relation between the phases of the pressure wave and piston movement is

Figure 6. Temperature wave at warm end.

Figure 7. Temperature wave at cold end.

NUMERICAL STUDY OF A NEW 4 K GM/PT HYBRID COOLER

Figure 8. Cooling power as a function of the original difference of phase.

Figure 10.

Cooling power as a function of the

volume ratio of piston to pulse tube.

271

Figure 9. Cooling power as a function of the diameter of orifice.

Figure 11.

Cooling power as a function of the

volume ratio of regenerator to pulse tube.

reversed, 2) the diameter of the orifice is 0.25 mm, and 3) the volume ratio of the piston chamber to pulse tube is 0.3. It is important to note that increasing the volume of the regenerator always improves so there is no optimum volume ratio for the regenerator to pulse tube for but the optimum volume ratio for exists, that is 5.3. CONCLUSIONS

We have proposed a new type of hybrid refrigerator and developed a numerical simulation to model its performance. Some useful results have been achieved as follows: 1. The behavior of the new type of hybrid refrigerator with only a cold piston as its phase shifter is not as good as expected; the reason is that its phase shifting capacity is very limited. However, the performance is improved for the case where a cold piston is combined with another version of a phase shifter, such as an orifice. A future possibility is to combine the cold piston with a double-inlet type phase shifter. 2. The loss that most affects the performance of the hybrid refrigerator is the regenerator loss. 3. The optimum configuration for the new hybrid refrigerator with a cold piston and orifice as its

phase shifter is: 1) the relation between the phases of the pressure wave and piston movement is reversed, 2) the diameter of the orifice is 0.25 mm, 3) the volume ratio of the piston chamber to pulse tube is 0.3, and 4) the volume ratio of the regenerator to pulse tube is 5.3 (for cooling power efficiency). ACKNOWLEDGMENT

This activity was supported by the National Natural Science Foundation of China and Hui Guo Ren Yuan Foundation of CAS.

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REFERENCES 1. Li, R., “Great Process in Magnetic Regenerator Material and 4K G-M Cryocooler,” Proceedings of ICCR’98, International Academic Publishers, Beijing (1998), pp. 67-73. 2. Wang, C., Thummes, G., and Heiden, C., “A Two Stage Pulse Tube Cooler Operating below 4K,” Cryogenics, vol. 37, no. 3 (1997), pp.159.

3. Kays, W.M. and London, A.L., Compact Heat Exchanger, 2nd ed., McGraw-Hill, New York (1964).

Thermally Actuated

Pulse Tube Cooler

Y. Matsubara, H. Kobayashi and S. L. Zhou

Atomic Energy Research Institute, Nihon University Funabasbi, Chiba 274-8501 Japan

ABSTRACT

An isotope of Helium gas, is an attractive working gas for cryocoolers that strive to efficiently provide cooling temperatures below 4 K. The drawback of this gas is that it is extremely expensive. To minimize the total amount of gaseous a hybrid cycle has been proposed. A single-stage pulse tube cooler using as the working gas and either a Stirlingtype or GM-type compressor system can provide a cooling temperature around 40 K, starting

from room temperature. A secondary cycle using can then be thermally attached to this cold head to produce a secondary cooling temperature below 4 K. The operating frequency of the secondary cycle should be lower than 2 Hz to prevent degradation of the performance of the lower-stage regenerator. In this study, a thermally actuated pressure wave generator driven between the temperatures of 40 K and 300 K was selected; it has no difficulty in generating a pressure wave below a frequency of 2 Hz. Workflow analysis calculations indicate that the total amount of working gas required for this secondary cycle may be minimized by the use of a warm expander type of phase control. INTRODUCTION

Additional performance improvement is still required of cryocoolers designed to cool LTC devices such as low noise signal detectors. The long-term goal is a 4 K cryocooler that is more reliable, more compact, generates minmal disturbance of the cooled devices, and has low cost and

low required input power. To satisfy these requirements, a pulse tube cooler operating at a low cycle frequency by a non-mechanical compressor would be a good choice. There are many papers on 4 K pulse tube coolers. However, all of these studies have been based on mechanical compressor systems.1-22

Since there are no moving components in the low-temperature region of a pulse tube cooler, the most significant feature to classify it is the gas compressor system that is located at room temperature and generates the pressure oscillation. Thus, pulse tube coolers can be classified into

two different styles: GM-type, utilizing the valved compressor, and Stirling-type, utilizing the valveless compressor. A multi-staged GM-type pulse tube cooler is able to achieve lower temperatures, down to 4 K, because of its lower operating frequency. However the thermodynamic efficiency is reduced due to the work losses within the rotary valve.

Cryocoolers 11, edited by R.G. Ross, Jr.

Kluwer Academic/Plenum Publishers, 2001

273

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HYBRID CRYOCOOLERS USING PULSE TUBES

Figure 1. Basic configuration of thermally actuated pulse tube cooler. In the case of a Stirling-type pulse tube cooler, the required pressure wave is provided by a direct-coupled compressor piston without any valves. Therefore, the driving frequency can be increased up to 50 Hz. For required temperatures above about 40 K, the thermodynamic efficiency is also good because it has no irreversible components such as the valves. However, it is difficult to get lower temperatures below about 10 K because of its higher frequency operation. As an alternative to the Stirling-type pulse tube cooler, the Vuilleumier (VM-type) pulse tube cooler has been reported.23 The feasibility of applying the VM cycle to a 4 K pulse tube cooler has also been studied.24 Further modifications of the VM-type 4 K pulse tube cooler are discussed in this paper.

BASIC MODEL OF A THERMALLY-ACTUATED PULSE TUBE COOLER

A hybrid pulse tube cooler for cooling temperatures below 4 K is schematically illustrated in Fig. 1. It consists of two cycles thermally coupled at the cold heat exchanger (f); this allows for a minimum usage of gas in the secondary cycle, or a different operating pressure or frequency to be used. The operating frequency of the secondary cycle should be lower than 2 Hz to produce efficient cooling performance below 4 K. A single-stage pulse tube cooler (a ~ d) provides a cooling temperature around 40 K at the cold heat exchanger (f). This portion of the cooler is the upper stage or primary cycle, and can be driven by either a Stirling-type or a GM-type compressor system. The secondary cycle is pre-cooled by the primary cycle and is driven by the thermal compressor (e ~ h) similar to that of the VM cycle. Here the thermal compressor also rejects its heat to the primary cycle at the cold heat exchanger. Since pulse tube coolers for the primary cycle are now commercially available, this paper focuses on the secondary cycle. Figure 2 highlights the basic function of the thermal compressor. A warm expander was

selected as the phase shifter. When the warm heat exchanger is heated to a temperature,

and

the cold heat exchanger is cooled to a temperature, by the primary cycle, the movement of the displacer generates a pressure wave. When the displacer moves from the warm end to the cold end, the cold gas will be displaced to the warm end through the warm regenerator, so it is heated to the temperature as a result, the pressure within the closed volume increases. The reverse movement of the displacer decreases the pressure in the same way. Therefore, the back and force movement of the displacer generates an oscillating pressure. However, if there is no pulse tube section, no work is generated. If there is a work receiver such as a warm expander or an orifice, then work is done and flows in the direction of the arrow indicated in the figure.

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Figure 2. Energy flow map for thermally driven pulse tube cooler.

Figure 3. Basic model for workflow analysis of DPT model.

This situation can be explained by an energy flow map. The work is generated within the warm regenerator by consuming the heat flow of the opposite direction. The generated work is transferred back to the cold end of the warm regenerator through the displacer. Most of the work flows to the pulse tube cooler section; however a small fraction of workflow to the warm regenerator is required, because the work amplification within the warm regenerator is limited due to the limited temperature ratio of and according to the 2nd law of thermodynamics. In general, this warm regenerator has a limited heat transfer surface and flow friction. The limited heat transfer surface generates the enthalpy flow, which increases the required cooling capacity of the primary cycle at the cold heat exchanger. The flow friction in the warm regenerator generates a pressure drop and decreases the work amplification, which results in additional mechanical work being required to move the displacer. Therefore, the design of the warm regenerator is very important to make the system efficient. A simplified workflow analytical method for the double inlet pulse tube (DPT) model is given in Fig. 3. Sinusoidal movement of the displacer gives the variable volume, and The mass Sow at the orifice, bypass valve, warm regenerator, and cold regenerator are given as and respectively. It is assumed the flow friction within the regenerator is generated at the middle of the regenerator, exclusively. With this assumption, four different pressure equations are

276

obtained as

HYBRID CRYOCOOLERS USING PULSE TUBES

and

The equations used in this analysis are as follows:

and are the swept volume at the hot end and the cold end of the displacer. and are the void volume at the hot end and the cold end heat exchangers. is the angular velocity of the cycle. are the flow coefficients at the warm regenerator, cold regenerator, bypass valve and the orifice valve respectively. The equivalent volume at the both

end of the pulse tube, are obtained by assuming a gas piston having the variable volume within the pulse tube. The PV work and workflow at each interesting point is solved by,

where R is the gas constant as an ideal gas (2078 for He4 and 2757 for He3). jj is the

number of time division in a cycle used in this numerical analysis. The enthalpy flow through the warm regenerator is given by,

where

is the inefficiency of the regenerator and was given as a constant value.

CALCULATED RESULT OF DPT AND WEPT MODEL

The calculated result of the case of double inlet method is given in Fig. 4. Pulse tube inner diameter is 10 mm and the length is 350 mm. Cv value of orifice and bypass for double inlet are fixed near their optimum opening rates and were preliminary calculated as 0.008 for orifice and 0.048 for bypass. The reservoir volume is 1 liter. The cold regenerator located between 4 K and 40 K is 17 mm in diameter and 135 mm in length. The porosity and the hydraulic diameter are fixed to 0.4 and 0.04 mm respectively. The thermal compressor is operated between 40 K and the room temperature 300 K. The warm regenerator inner diameter is 25 mm and the length is 100 mm. The mean pressure is 1.3 MPa and driving frequency of the displacer is 1 Hz. The hydraulic diameter of the packing material of the warm regenerator is changed to find out the optimum condition under the constant porosity of 0.678. At each calculation, is changed until the cold end workflow at 4 K becomes 1 watt. The result indicates the workflow at the Va (Wa in Fig. 4 (a)) is almost constant but it

quickly increases at dense mesh region. The enthalpy flow through the warm regenerator r decreases with increasing mesh number. The sum of these two becomes the heat flow (Q) and is the same as the minimum required cooling capacity at 40 K to keep this temperature constant.

Here the enthalpy flow through the cold regenerator is neglected. We found the minimum heat

THERMALLY ACTUATED

PULSE TUBE COOLER

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Figure 4. Calculated result of DPT model.

flow is 22 watts when the hydraulic diameter of the warm regenerator is around 0.07 mm. The required displacer stroke volume is about The workflow Wa separates into two directions, to the cold and warm regenerator. The work flowing to the warm regenerator (Wrl in Fig. 4(b)) is amplified and flows out of the warm end of the warm regenerator, and then separates again into two directions, to the bypass and to the warm end of the displacer (Wh). Therefore the additional mechanical work, which is the difference of Wa and Wh, is required to move the displacer. In this example, the pressure ratio up to 1.9 was generated, and it decreases with the increase of the stroke volume. However the use of higher pressure ratio range should be avoided because the required cooling capacity at 40 K also increases for same cooling capacity at 4 K. Similar calculation has been done for the case of the warm expander pulse tube cooler (WEPT). A warm expander with of the swept volume and with a phase lead of 70

degrees over the displacer volume (Va) is used instead of the double inlet phase shifter system. All other sizes and driving conditions are the same with the previous double inlet method. The result is shown in Fig. 5 and it indicates that the most of the important parameters are not a strong function of the warm regenerator mesh size. The minimum required cooling capacity at 40 K reduced to 14 watts. And the stroke volume of the displacer also reduced to Additional mechanical work is negligibly small, because the work used for the double inlet line is no longer required in this case.

Figure 5. Calculated result of WEPT model.

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Figure 6. Variation of thermal compressor for 4 K pulse tube cooler.

Figure 7. Gas movement within the pulse tube and work transfer tube.

VARIATION OF THERMAL COMPRESSOR FOR 4 K PULSE TUBE COOLER

If the displacer is removed from the cold section to the room temperature, it becomes the configuration as shown in Fig. 6 (a). Here the tube without the solid displacer is called as the work transfer tube, because it mainly transfers the work to the lower temperature with minimum transfer of the heat flow. The calculated result indicates the required cooling capacity at 40 K is 18.7 watts, which is somewhat between the DPT and WEPT of the previous case. The volume of the work transfer tube is

and the stroke volume of the displacer is

which is

THERMALLY ACTUATED

PULSE TUBE COOLER

279

slightly larger than the previous result of DPT. However, this method has no moving parts in its cold section, and we can separate all of the moving parts from the cold section by means of flexible tube. Similar calculation for the double inlet model has been done. However it requires

the cooling capacity of 31 watts at 40 K, and is not comparable with the other methods.

Fig. 7 indicates the phase difference of the pressure and the gas pistons of the warm expander method. Both of the gas piston length of the pulse tube and the work transfer tube are almost constant within a cycle similar to the solid piston. This result is caused by the lower pressure ratio. The PV diagrams plotted from this gas trajectory at Va and Ve indicate that this

part of the cycle represents the Stirling cycle operating between 40 K and 4 K, although the required cooling capacity at 40 K is increased to 14 watts from 10 watts for the ideal Stirling cycle. Another variation of the thermal compressor is schematically given in Fig. 6 (b). It consists of a pressure wave generator operating between 900 K and 300 K, a phase shifter and the work transfer tube. In this particular case, the work transfer tube can be replaced with the regenerator and it becomes an ordinary thermally precooled single stage pulse tube cooler. Most of the example calculations in this study are based on 1 watt workflow at 4 K for He4, and the cooling capacity is reduced to about 20 % due to the non-ideal gas properties of He4. In the case of He3, this reduction is not clear so far, because the thermal transport properties of He3 are not available for us. From the viewpoint of minimizing the total amount of the He3, WEPT with cold displacer will be the best, however, the WEPT with warm displacer seems to be the best for the application adaptability. CONCLUSION

1. Detail of simplified workflow analysis of the thermally actuated pulse tube cooler was given. It indicates the warm expander method is the best choice from the both viewpoints of minimizing the amount of enclosed gas and the required cooling capacity at the precooling temperature. Therefore this method could be applied for He3 pulse tube cooler providing the cooling temperature below 4 K efficiently. 2. The work transfer tube has been introduced. The function of this tube is very similar to the pulse tube, which transfers the work from end to end with minimum heat pumping effect. Only the difference is the direction of the workflow. This tube transfers the work from high temperature to low temperature, which is the reverse of pulse tube. Minimizing the heat flow

through the work transfer tube is the most important requirement. REFERENCES 1. Y.Matsubara, J.L.Gao, K.Tanida, Y.Hiresaki and M.Kaneko, “An Experimental and Analytical Investigation of 4 K Pulse Tube Refrigerator”, Proceedings of 7th International Cryocooler Conference, (1993), pp. 166-186.

2.

J.L.Gao and Y.Matsubara, "4 K Pulse Tube Refrigeration", Proceedings of 4th JSJS on Cryocoolers and Concerned Topics, (1993), pp. 69-73.

3.

J.L.Gao and Y.Matsubara, "Experimental Investigation of 4 K Pulse Tube Refrigerator", Cryogenics, vol. 34, (1994), pp. 25-30.

4.

Y. Matsubara and J.L. Gao, “Novel Configuration of Three-stage Pulse Tube Refrigerator for Temperatures below 4 K”, Cryogenics, Vol.34 (1994) 259.

5. Y. Matsubara and J.L. Gao, “Multi-staged Pulse Tube Refrigerator for Superconducting Magnet Applications”, Proceedings of the Fifteenth International Cryogenic Engineering Conference, (1994) pp. 155.

6. Y.Matsubara and J.Gao, "Multi-Stage Pulse Tube Refrigerator for Temperatures below 4 K", Cryocoolers 8, Edited by R.G.Ross.Jr, Plenum Press. (1995), pp. 345-352.

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K.Tanaida, J.L. Gao, N.Yoshimura and Y. Matsubara, “Three-Staged Pulse Tube Refrigerator controlled by Four-Valve Method”, Advances in Cryogenic Engineering, Vol. 41, Plenum Press, (1996), pp. 1503-1509.

8.

J.L. Gao, Y. Hiresaki and Y. Matsubara”, “A Hybrid Two-Stage Refrigerator operated at Temperatures below 4 K”, Advances in Cryogenic Engineering, Vol. 41, Plenum Press, (1996), pp. 1495-1502.

9.

J. L. Gao and Y. Matsubara, “An Inter-Phasing Pulse Tube Refrigerator for High Refrigeration Efficiency”, Proc. of Sixteenth ICEC (1997) pp. 295-298.

10. K. Tanida, J. L. Guo Y. Hiresaki and Y. Matsubara, “Performance of the Hybrid Two-Stage Pulse Tube Refrigerator”, Proc. of Sixteenth ICEC (1997) pp. 303-306. 11. Y. Ohtani, G. R. Chandratilleke, H. Nakagome, N.Yoshimura, Y. Matsubara, M. Narita and H. Okuda, “Development of a Three Stage Pulse Tube Refrigerator”, Proc. of the Fifth Japanese-Sino Joint Seminar (1997) pp. 118-122. 12. Yoshimura, N., Zhou, S.L., Matsubara, Y., Chandratilleke, Y, Ohtani, Y, Nakagome.H., Okuda, H., and Shinohara, S., “Conceptual Design of Space Qualified 4 K Pulse Tube Cryocooler”, Cryocoolers 10, KA/Plenum Press (1999) pp. 221-226. 13. Yoshimura, N., Zhou, S.L., Matsubara, Y, Chandratilleke, Y, Ohtani, Y, Nakagome.H., Okuda, H., and Shinohara, S., “Performance Dependence of a 4 K Pulse Tube Cryocooler on Working Pressure”, Cryocoolers 10, KA/Plenum Press (1999) pp. 227-232.

14. C. Wang, G. Thummes, C. Heiden, “Effects of DC gas flow on performance of two-stage 4 K pulse tube coolers", Cryogenics, vol. 38, no. 6 (1998), pp. 689-695. 15. G. Thummes, C. Wang, C. Heiden, “Small scale liquefaction using a two-stage 4 K pulse tube cooler”, Cryogenics, vol. 38, no. 3 (1998), pp. 337-342.

16. C. Wang, G. Thummes, C. Heiden, “Experimental study of staging method for two-stage pulse tube refrigerators for liquid temperatures”, Cryogenics, vol. 37, no. 12 (1997), pp. 857-863. 17. G. Chen, L. Qiu, J. Zheng, P. Yan, Z. Gan, X. Bai, Z. Huang, “Experimental study on a double-orifice two-stage pulse tube refrigerator”, Cryogenics, vol. 37, no. 5 (1997), pp. 271-273. 18. M.Y. Xu, A.T.A.M. De Waele, Y.L. Ju, A pulse tube refrigerator below 2 K, Cryogenics, vol. 39, no. 10 (1999), pp. 865-869.

19. A. von Schneidemesser, G. Thummes, C. Heiden, “Generation of liquid helium temperatures using a lead regenerator in a GM precooled pulse tube stage”, Cryogenics, vol. 40, no. 1 (2000), pp. 67-70. 20. A. von Schneidemesser, G. Thummes, C. Heiden, “Performance of a single-stage 4 K pulse tube cooler with neodymium regenerator precooled with a single-stage GM refrigerator, Cryogenics, vol. 39, no. 9 (1999), pp. 783-789. 21. G. Thummes, S. Bender, C. Heiden, “Approaching the lambda line with a liquid nitrogen precooled two-stage pulse tube refrigerator”, Cryogenics, vol.36, no. 9 (1996), pp. 709-711. 22. S.L. Zhou, G. Thummes, and Y. Matsubara, “Experimental Investigation of Loss Mechanisms in a 4 K Pulse Tube”, Advances in Cryogenic Engineering, Vol. 45, Plenum Press, (to be published).

23. Kaneko, M. and Matsubara, Y, “Thermally Actuated Pulse Tube Refrigerator”, Cryocoolers-5, Naval Postgraduate School, USA (1988), pp. 103-112.

24. Y. Matsubara and S.L. Zhou, “Feasibility Study of Applying Thermal Compressor to 4 K Pulse Tube Cooler”, Proceedings of the eighteenth International Cryogenic Engineering Conference, (2000), (to be published).

Investigation of Helium and Nitrogen Mixtures in a Pulse Tube Refrigerator Z.H. Can, G.B. Chen, G.Thummes† and C.Heiden†

Cryogenics Lab. Zhejiang University, Hangzhou, 310027, P.R.China † Institute of Applied Physics, University of Giessen, 35392, Germany

ABSTRACT

Normally, the working fluid used in regenerative cryocoolers is not condensed at any point in the working cycle. Helium is the popular working medium for such systems, rather than a gas mixture because of its excellent thermodynamic and transport properties. However, our theoretical analyses indicate that pure helium is not the best working fluid for regenerative refrigerators near

80 K, either for the best cooling power, or for the best coefficient of performance (COP). Motivated by the advantages of pulse tube refrigerators, which have no moving parts in the cold end, this paper describes an experimental investigation on a two-component, multi-phase helium and nitrogen gas mixture in a single-stage pulse tube refrigerator. The experimental results show that the COP and the cooling power can be increased to some extent near 80 K with the use of two-component gas mixtures with less than 25% nitrogen. A relatively stable temperature platform at the triple point of nitrogen (63.15K), which is independent of the nitrogen fraction, is obtained when the cooling power is below 7W.

INTRODUCTION Investigations of pulse tube refrigerators have made great progress over the past few years since Mikulin proposed the orifice type configuration in 1985. With the application of various novel phase shifters and construction improvements, the refrigeration temperature has continuously dropped and the thermodynamic efficiency has continuously improved. So far, refrigeration temperatures as low as 20 K have been reached with a single-stage pulse tube refrigerator.1 In addition, cooling temperatures below 4.2 K have been obtained by some two-stage pulse tube refrigerators,2,3 and a three-stage pulse tube refrigerator with has reached a temperature lower than the point of 1.78K.4 The results confirm that refrigeration performance can be increased if the appropriate refrigerants or mixed fluids are used in the corresponding temperature regions. Due to their advantages of simple construction and no moving parts in the cold end, pulse tube refrigerators provide feasible conditions for using a multi-phase fluid as the refrigerant.

One may imagine that, if a mixture of helium and nitrogen were used as the refrigerant in a pulse tube refrigeration system, the nitrogen will liquefy when the refrigeration temperature reaches the liquefaction point of nitrogen. In such a system, refrigeration performance could be

higher than that using pure helium, since the latent heat of the phase change of liquid nitrogen could be used. A number of researchers are interested in utilizing mixtures as the refrigerant for Cryocoolers 11, edited by R.G. Ross, Jr.

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Figure 1. Schematic of a GM-type single-stage pulse tube refrigerator.

regenerative refrigerators.5,6 The possibility of using mixtures of helium and nitrogen to improve

the refrigeration performance of pulse tube refrigerators was also analyzed in our previous paper.7

However, few reports of theoretical studies or successful experimental results have been published so far. An experiment with helium and neon mixtures as the refrigerant in a two-stage GM cooler was described by P C Mcdonald.8 Although no benefit in cooling power was obtained from the mixture of helium and neon, compared with pure helium in the refrigerator, a relatively stable temperature was obtained for cooling superconducting devices at 26K, close to the triple point of neon. In order to increase the refrigeration performance of pulse tube refrigerators near 80K, the authors have been exploring the possibility of using mixture gases, both theoretically9 and experimentally. In this paper, we present some experimental results using helium and nitrogen mixtures in a pulse tube refrigerator. EXPERIMENTAL ARRANGEMENT

The experimental setup used for measuring refrigeration performance of mixture fluids consists of the refrigeration system, vacuum system, measuring system, and mixture preparation system. Figure 1 shows a sketch of the refrigeration system, which consists of a helium compressor (Leybold, RW2), a GM-type rotary valve, and a single stage pulse tube refrigerator. Dimensions of the regenerator and pulse tube are and respectively. A copper tube for water cooling is wrapped and soldered around the hot end of the pulse tube. The pulse tube refrigerator is equipped with a double-inlet configuration via two needle valves (Nupro, type M, 25div/turn) and a 0.5 liter reservoir. The temperature profile along the pulse tube wall is measured by Pt100 resistance thermometers in Figure 1). Piezoelectric pressure sensors (Siemens, type KPY 46R) are used for monitoring the dynamic pressures at the hot ends of the regenerator and pulse tube as well as in the reservoir The cooling power is measured by a calibrated resistive heater, PBH-100 attached to the cold-end heat exchanger. The input power of the compressor is measured by EKM 265. Pressure and temperature recordings are accomplished by means of a data acquisition system controlled by a personal computer. EXPERIMENT PROCEDURE The test pulse tube refrigerator was operated with a charge pressure of 1.7 MPa at room temperature (pressures quoted throughout this paper are absolute values), and at a frequency of 2 Hz.

Mixture Preparation. Dolton’s law of partial pressures may be used for preparing a mixture of helium and nitrogen. For example, a mixture of 80% helium and 20% nitrogen can be prepared by firstly filling the system with helium gas to 1.36 MPa, then filling nitrogen gas up to a total

pressure of 1.7 MPa. With this method: 97/3, 94/6, 92/8, 88/12, 83/17, 80/2 and 75/25 % helium/ % nitrogen gas mixtures were prepared for the experiment.

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Comparison Criterion. A proper comparison standard is particularly important when comparing the refrigeration performance of a mixed fluid with that of pure helium. Both the orifice valve and double-inlet valve settings for the refrigeration at 80K, and for the lowest cooling temperature (about 30K), are different when pure helium is used as the working fluid. In addition, it is difficult to adjust the valve setting to a temperature below the condensation point of nitrogen due to the multi-phase fluid that exists when a mixed fluid is used. Therefore, the optimized settings for each refrigerant in this paper were fixed for a condition of 9W of cooling at about 80K. Furthermore, since the optimized double-inlet valve setting was found to be relatively insensitive to the

helium/nitrogen mixture ratio, the same double-inlet valve setting was adopted for all cases including pure helium; this simplified the experimental procedure. In short, the experimental procedure was divided into two steps. The first was to optimize the orifice valve opening for a 9W cooling load at 80K. The second was to measure the refrigeration load-line (heat load vs. cooling temperature) with the optimized orifice setting. RESULTS AND DISCUSSION Cooling Power and Coefficient of Performance

Figure 2 shows the experimental results of cooling power (a) and COP (b) for the pulse tube refrigerator with mixtures of helium and nitrogen. The figure also gives the curves for pure helium as a comparison. We can see that both the cooling power and COP with a mixed fluid are higher than that of pure helium for various nitrogen fractions up to 25%. When the cooling temperature is

higher than 70K, the degree of performance improvement depends on the composition of the mixture. It can be seen from Figure 2 that when the cooling power is lower than 7W, all curves in this figure are concentrated at approximately the triple point of nitrogen (63.14K). An approximate isothermal line near 64K is obtained under 7W of cooling power, which is independent of the fraction of nitrogen Because of the advantage of no moving parts in the cold space of the pulse tube, the isothermal working platform creates a very constant temperature for devices to be cooled, and the system is resistant to degradation caused by particles of liquid or solid nitrogen that are carried into the pulse tube by the working gas flow. Figure 3 shows the relationship between specific cooling power and COP versus composition of the mixture at 80K. When the fraction of helium is greater than about 85%, the refrigeration performance using a mixture is better than that with pure helium. The maximum gains obtained in the experiments include a 3.85% increase in cooling power with a helium fraction of 88%, and a 4.05% increase in COP with a helium fraction of 97%.

Figure 2. Cooling capacity (a) and COP (b) versus temperature.

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Figure 3. Specific cooling power and COP versus fraction of mixture.

Figure 4. Phase change phenomena of 3% nitrogen gas mixture.

Two transition temperatures for helium/nitrogen mixtures can be observed in Figure 2. This

can be explained as follows. During the cooling-down process, nitrogen, as the secondary refrigerant in the mixture, will be condensed at its liquid point, which depends on the partial pressure of

nitrogen and the fraction nitrogen; this is the vapor-liquid transition temperature. Furthermore, the liquid-solid transition temperature will occur at the triple point of nitrogen, 63.14K, when the liquid nitrogen freezes. After that, the nitrogen is in a vapor-liquid-solid multi-phase state. If the nitrogen fraction is much less, all of the nitrogen will be solidified. Figure 4 shows the phase change phenomena of a 3% nitrogen mixture during the cooling process. It indicates that the liquid-solid phase change occurs around 63K. Then another tempera-

ture platform appears on the cooling curve near 58K. Finally, after the nitrogen fraction in the mixture is completely solidified and the refrigerant is almost pure helium, the refrigeration process

gradually approaches a minimum point of 33K or so. Figure 5 and Table 1 give experimental results of the vapor-liquid phase change temperatures which are clearly in accordance with the partial pressure of nitrogen in the mixture. However, the

liquid-solid change temperature for all tests with various mixture fractions (3-25%) is almost the same as 63-65K; this is because the liquid-solid phase change is caused by the triple point phenomenon of nitrogen.

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Figure 5. Vapor-liquid phase change temperature in helium-nitrogen mixture.

Temperature Profile

The locations of the eight temperature measuring points in the experiment were presented in Figure 1. Accordingly, the temperature at the hot end of the pulse tube is 290K, due to the water cooling heat exchanger, while the hot end of the regenerator is at room temperature (305K). It is useful to define a dimensionless relative temperature as

where T is the variable value to be measured, and is the cold end temperature of the pulse tube; is 290K for the pulse tube and 305K for the regenerator. Figure 6 shows the relative temperature profile during the refrigeration process with pure helium versus relative position along the pulse tube or regenerator. We can see from Figure 6(a) that the temperature distribution profile along the regenerator shifts to the outside with increasing heat

Figure 6. Temperature profile with position along regenerator (a) and pulse tube (b).

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Figure 7. Temperature profile versus heat load at solidification process.

Figure 8. Temperature profile versus heat load at complete solidification state of nitrogen.

load. Figure 6(b) gives the same curve shape in the opposite direction along the pulse tube. But the

latter seems closer together. This is in accordance with the theoretical result given by de Waele for explaining the non-ideal gas effect.11 In conclusion, the temperature curves along both the pulse tube and the regenerator are expanded in the direction of increasing heat load. Figure 7 and Figure 8 show different temperature profiles during the refrigeration process with a mixture fluid. It can be divided into three steps. The first step, when the cooling temperature drops to above the liquid-solid phase change point, the curve shape appears the same as that with pure helium. That is, both the temperature profile along the pulse tube and regenerator expand in the direction of increasing heat load, as shown in Figure 7(a). The second, when the refrigeration tem-

perature is at the solidification point of nitrogen, the curves along both the pulse tube and regenerator appear opposite to their state with pure helium, and shift to the inside with increasing heat load,

as shown in Figure 7(b). The third, when the refrigeration temperature is lower than the phase change point (63.15K), the curves again show a tendency to the same shape as with pure helium, as

shown in Figure 8. This is because the working fluid becomes pure helium again, as most nitrogen in the mixture has been completely solidified.

Effect of Orifice and Double-inlet Valves Figure 9 shows the relationship between optimized orifice valve setting and helium fraction. The optimized valve opening becomes larger with an increase in nitrogen fraction in the mix-

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Figure 9. Optimized orifice valve setting versus fraction of gas mixture.

Figure 10. Refrigeration temperature versus orifice valve setting.

ture. For example, the refrigeration temperature in Figure 10 gradually drops down with an increase of the orifice opening when 12% nitrogen is used in the mixture. The minimum refrigeration temperature, 79.7K, was obtained with a 8.94W heat load and an orifice opening of 70 div. In contrast, with a pure helium refrigerant, the optimized orifice opening is 57 div (in Figure 9) with almost the same heat load. Figure 11 shows coldend temperature versus double-inlet valve opening for a pure helium refrigerant; the best opening of the double–inlet valve is 95 div. Similar data for sensitivity to double-inlet valve opening for a 25% nitrogen mixture are shown in Figure 12. These data indicate that the optimized opening is also 95 div with almost the same heat load. It seems that the same double-inlet valve setting can be adopted for both cases. This fact simplifies the experimental procedure. CONCLUSION Experiments using a number of helium and nitrogen mixtures in a GM-type single-stage pulse tube refrigerator have been carried out. The experimental results show that both the coefficient of performance (COP) and the cooling power can be increased to some extent when a mixed helium refrigerant is used with a nitrogen fraction of up to 25%. A cooling temperature platform near 63.14K has been observed for the tested fractions of nitrogen. The vapor-liquid and liquid-solid

transition temperatures of nitrogen are also observed during the pulse tube cooling process. The

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Figure 11. Refrigeration temperature versus double-inlet valve setting for pure helium.

Figure 12. Double-inlet valve setting for 25% nitrogen mixture.

optimized double-inlet valve settings are the same for both pure helium and the helium/nitrogen

mixed fluid, while the optimized orifice valve setting depends on the composition of the mixture. ACKNOWLEDGMENT

The project is financially supported by the National Natural Sciences Foundation of China and the Deutscher Akademischer Austauschdienst (DAAD). REFERENCES 1. Ishizaki, Y., et al., “Experimental Performance of Modified Pulse Tube Refrigerator Below 80K. Down 23K,” 7th Intl. Cryocooler Conf. Proceeding, Part 1 (1993), p. 140. 2. Wang, C., Thummes, G., and Heiden, C., “A Two-stage Pulse Tube Cooler Operating Below 4K,” Cryogenics, vol.37 (1997), p. 159.

3. Chen, G.B., Qiu, L.M., Zheng, J.Y., Yan, P.D., Gan, Z.H., Bai, X., and Huang, Z.X., “Experimental Study on a Double-orifice Two-stage Pulse Tube Refrigerator,” Cryogenics, vol. 37 (1997), p. 271. 4. Xu, M.Y., De Waele, ATAM., and Ju, Y.L., “A Pulse Tube Refrigerator below 2K,” Cryogenics, vol. 39, (1999), p. 865.

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5. Walker, G., “Stirling-Cycle Cooling Engine with Two-Phases, Two-Components Working Fluid,” Cryogenics, vol. 8 (1974), p. 459. 6. Patwardhan K.P. and Bapat S.L., “Cyclic Simulation of Stirling Cycle Cryogenerator Using Two

Component Two Phase Working Fluid Combinations,” Proceedings of ICCR’98, Hangzhou, International Academic Publishers (1998), p. 397. 7. Zhao L., Chen, G.B., and Yu, J.P., “The possibility use with Gas Mixtures in Pulse Tube Refrigeration”, Cryogenics Engineering, 5 (1996), p. 15. 8. McDonald, P.C., “Self-regulating Temperature Control of a Gifford-McMahon Refrigerator for Potential Use with Neon in High-Tc Power Applications,” Superconductor Science & Technology, 11 (1998), p. 817.

9. Chen, G.B., Gan, Z.H., Thummes, G., and Heiden, C., “Thermodynamic Performance Prediction of Pulse Tube Refrigeration with Mixture Fluids,” Cryogenics, to be published. 10. De Waele ATAM., Xu, M.Y., and Ju, Y.L., “Nonideal-gas Effect in Regenerators,” Cryogenics, vol.39 (1999), p. 847.

Pulse Tube Refrigeration with a Combined Cooling and Freezing Cycle for HTSC Devices Guobang Chen, Zhihua Gan, Limin Qiu, and Jianping Yu Cryogenics Laboratory Zhejiang University Hangzhou 310027, China

ABSTRACT Based on the analysis of a modified Brayton cycle, a combined cooling and freezing cycle

is introduced for pulse tube refrigeration with a binary refrigerant. A novel configuration of a pulse tube refrigerator with a frozen cryogen accumulator is proposed as an example for possible applications. Some experimental phenomena explaining the working principles are presented. INTRODUCTION In recent years, substantial progress has been made in the study of high-

SQUID-based heart scanners cooled by small Stirling cryocoolers that are cryogen-free.1,2 When the refrigerators are running, a magnetic noise level as high as from 10 Hz to 100 Hz has been measured in SQUID magnetometer tests. When the refrigerators are switched off, the noise level drops to about (26 times lower), which is very close to the white-noise level of a SQUID cooled by liquid nitrogen. Additionally, recent investigations3-5 have been conducted on the use of mixed refrigerants to improve the refrigeration performance of pulse tube refrigerators in the 80 K region. Since pulse tube refrigerators bear the promise of high reliability and economic operation, the authors propose a novel configuration of a pulse tube refrigerator with a freezing cryogen accumulator for some particular applications. One may imagine that the freezing cryogen accumulator may provide the necessary refrigeration for a certain period of time after the pulse tube refrigerator with two-phase binary mixture is switched off. To explore the possibility of providing such a pulse tube refrigerator, a modified Brayton cycle with a binary mixture of nitrogen and helium was conceived and is described in detail. Then, a combined cooling and freezing two-phase system was added to the pulse tube refrigerator via a freezing cryogen accumulator. Finally, some experimental phenomena explaining the working principles of the novel pulse tube refrigerator are presented.

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Figure 1. T-S diagram for the pulse tube refrigeration cycle with two-isentropic and two-isobaric processes.

MODIFIED BRAYTON REFRIGERATION CYCLE

In order to predict the refrigeration performance of pulse tube refrigerators with a binary mixture, the authors recommend a practical model of the refrigeration cycle which can meet the requirements of the thermodynamic analysis. In the case of pulse tube refrigeration, the following assumptions must be made:

1. The processes of compression and expansion of the working refrigerant that result in refrigeration are adiabatic processes rather than isothermal ones 2. The rejection of the heat of compression into the coolant in the after-cooler is an isobaric process

3. The heat transfer process occurring at the cold end of the pulse tube is also isobaric; thus, the refrigeration effect of the pulse tube refrigerator is absorbed over a temperature range rather than at a constant temperature 4. Heat transfer processes in the regenerator are carried out under constant pressure conditions. The enthalpy imbalance between the compressed warm stream and the expanded cold stream in

the regenerator must also be taken into consideration. Figure 1 shows the T-S diagram of this assumed pulse tube refrigeration cycle with two isentropic and two isobaric processes. The working refrigerant leaving from the cold head heat exchanger in state 1 enters the regenerator, and then is warmed up at a constant pressure to state 2 ambient temperature). It is compressed adiabatically in the compressor from state 2 to 3 and then passes through the after-cooler to reach state 4 The working fluid then enters the regenerator and is cooled at a constant pressure to state 5 It is then expanded adiabatically at the cold end of the pulse tube to its lowest refrigerating temperature in state 6 which is lower than Tc. Finally, the refrigerant absorbs the cooling load at constant pressure P1 in the cold head heat exchanger (temperature and returns to state 1, finishing a complete cycle. The cycle shown in Fig. 1 is essentially a Brayton cycle. However, a conventional Brayton cycle utilizes a recuperative heat exchanger, and turbine machines are generally used as the compressor and expander in the cycle. In contrast, in the described pulse tube cycle, the recuperative heat exchanger of slotted plates is replaced by a regenerator in the pulse tube refrigerator, and the turbine expander is replaced by an expansion space within the pulse tube. Thus, the thermodynamic cycle for a pulse tube refrigerator can be considered as a variation of the Brayton cycle, and will be referred to as a modified Brayton cycle in this paper.

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Coefficient of Performance

Coefficient of performance (COP) of the modified Brayton cycle, which consists of two isobaric and two adiabatic processes, can be expressed as follows. The heat rejected to ambient is

The heat removed from the cold end heat exchanger at the constant pressure refrigeration effect of the system, is

namely the

The coefficient of performance of the system then can be expressed as follows

The above analysis is based on enthalpy-balanced condition in the regenerator. It must be stressed that the phenomenon of unbalance enthalpy flow in the regenerator cannot be neglected in the computation process. The heat rejected by compressed working refrigerant in the regenerator is

The heat absorbed by the expanded working refrigerant in the regenerator is

In an ideal case, For example, it is true when pure helium is used as the working refrigerant at around 80K. However, for mixture fluids, enthalpy difference of the warmer refrigerant flux could be greater or less than that of the cooler one in the regenerator. In this case,

the enthalpy deficit occurs. The enthalpy deficit may be positive or negative depending on the fluids and operating parameters, and can be expressed as

The enthalpy difference is an additional heat load (if it is a positive value) or refrigeration power (if it is a negative value) of the cycle, then the Eq. (2) is rewritten as

and Eq.(3) is turned into

Calculations have shown that the enthalpy deficit cannot be eliminated if mixture fluids are used as the working refrigerants4.

Computed Results

In the following calculations, the ambient temperature the filling pressure of working refrigerant 1.7 MPa and the pressure ratio of the refrigeration system are assumed, so as to make a comparison with experimental results. The refrigeration effect and COP with varied fraction of nitrogen in the mixture are computed according to Eqs.(7) and (8). The results are shown in Figs.2 and 3. We can find that by using a mixture of

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Figure 2. Theoretical cooling power versus helium fraction in the mixed fluid.

Figure 3. Calculated COP versus helium fraction in the mixed fluid.

10% nitrogen and 90% helium, the refrigeration effect and COP can be improved by 6.7% and 9.5%, respectively, compared with that using pure helium. Obviously, the nitrogen-helium mixture is a promising fluid for pulse tube refrigeration at around 80 K. VAPOR-LIQUID TWO-PHASE REFRIGERATION CYCLE The modified Brayton refrigeration cycle (Fig. 1) cannot provide two isothermal processes, thus the heat transfer efficiency will be lower. In addition, the cooling power of a cold gas refrigeration system is only contributed by the sensible heat of the refrigerant, which is much smaller than its latent heat. These drawbacks may be eliminated by utilizing a vapor-liquid two-phase refrigeration cycle. In such a cycle, both condensation and vaporization processes occur at a constant pressure and temperature, respectively. Thus more refrigeration effect can

be expected from the phase change process. The working fluid gives its latent heat of evaporation to the cold heat exchanger, while the refrigeration temperature does not change. The thermodynamic process describing the vapor-liquid two-phase refrigeration cycle is shown in Fig. 4 The working refrigerant leaving from the warm end of the regenerator at state is compressed isentropically to state 3 rejecting heat isobarically to a coolant in process 3-4. Then the compressed refrigerant enters the regenerator and is cooled to the saturated vapor line at state 5 Process 5-6 is a condensation process in the twophase region. The saturated liquid then expands isentropically from 6 to 7. In the process 7-9,

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Figure 4. Vapor-liquid two-phase refrigeration cycle

the liquefied refrigerant absorbs heat from the cold heat exchanger and evaporates. Gaseous refrigerant further absorbs heat from state 9 to state 1 then enters the regenerator, being heated to state 2 completing one cycle. The vapor-liquid two-phase refrigeration cycle is widely used in general refrigeration processes in which the critical temperature of the refrigerant is required to be higher than room temperature so as to reject the condensation heat to the ambient. Otherwise, the heat of condensation will be an interior heat source in the system. In the present case associated with obtaining refrigeration in the 80 K temperature region, the only refrigerant that may be used is

nitrogen. It is well known that the critical temperature of nitrogen (126.2 K) is much lower than room temperature, thus the heat transferred in process 5-6 in Fig. 4 becomes a heat load within the refrigerator itself. It offsets the effective cooling. In addition, the enthalpy deficit in the regenerator is possibly so high that the efficiency and benefits of the two-phase cycle become unfavorable. A combined cooling and freezing cycle with a binary refrigerant might be more preferable for cryogenic refrigeration, as explained in the following.

COOLING AND FREEZING COMBINED CYCLE

The T-S diagram of the proposed cooling and freezing combined cycle is shown in Fig. 5. The diagram consists of a sensible heat refrigeration cycle called the cooling cycle (1-2-3-4-5-6-1) and a complementary latent heat refrigeration cycle called the freezing cycle (1-2-3-4-5-7-8-10-1).

Figure 5. Cooling and freezing combined cycle.

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The working refrigerant is assumed to be a mixture of helium and another component, such as nitrogen, which has a suitable boiling-point temperature for the required application. Helium in the mixture enables sensible heat refrigeration to temperatures as low as 30 K using a singlestage pulse tube refrigerator; it is referred to as the cooling medium. Nitrogen has a higher boiling point and can be condensed into its liquid phase at the cold head temperature; it is referred to as the freezing medium, and contributes latent heat refrigeration. The working process of the cooling and freezing combined cycle includes the following three steps: 1. The system firstly works in accordance with the modified Brayton refrigeration cycle when the refrigeration temperature is higher than the boiling point of nitrogen. In this case, both helium and nitrogen play the role of a cooling media. In fact, this is a pre-cooling process of the system, called the sensible heat refrigeration process. 2. When the cooling temperature reaches the condensation temperature of nitrogen at a corresponding partial pressure, the cycle turns into a vapor-liquid two-phase refrigeration cycle. The nitrogen then plays the role of a freezing medium, making latent heat refrigeration. In contrast, the helium continues as a cooling medium in the sensible heat refrigeration state. When the cooling temperature is dropped down to the liquefaction point of nitrogen, the twophase state appears, and nitrogen is gradually condensed at the cold head. With continued decrease of the refrigeration temperature, more and more liquid nitrogen will be obtained. This is the cooling and freezing combined process.

3. The total pressure of the system may slightly decrease during step 2, since some fraction of the

nitrogen has been liquefied. As a result, the gaseous component of the system gradually becomes nearly pure helium gas. At this point, the system transitions to work again as a sensible heat refrigeration cycle based on pure helium as the refrigerant. Hence, we can forget the influence of the transport properties of the freezing medium nitrogen in the binary mixture, which are normally poorer than those of pure helium. It is found that the cooling cycle can cool the system down to a temperature considerably below the triple point of nitrogen, where the freezing medium nitrogen is in a state of supercooled liquid or even solid.

PULSE TUBE REFRIGERATOR WITH A FREEZING CRYOGEN ACCUMULATOR Based on the above analysis, the authors propose a novel pulse tube refrigerator with a liquid or solid cryogen accumulator as shown in Fig. 6. In the pre-cooling operation, the

system is firstly in a sensible heat refrigeration process with a mixed refrigerant of helium and nitrogen. Then, the nitrogen in the mixture begins condensing as the cooling temperature reaches its liquefaction point. The accumulator may accept the condensed fraction of nitrogen by means

Figure 6. Pulse tube refrigerator with a freezing cryogen accumulator.

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Figure 7. Cooling power versus cooling temperature of pulse tube refrigerator of the phase separator under the cold end of the pulse tube while the system operates in a

cooling and freezing combined process. The liquefied nitrogen in the accumulator can further be solidified when the cooling temperature drops below its triple point. The liquid or solid

freezing medium contained in the accumulator provides additional cooling capacity associated with the latent heat of evaporation or sublimation to meet the requirements of the application, even if the machine is switched off. The time duration of the provided refrigeration depends on the volume of the accumulator and the cooling load of the device to be cooled. Of course,

the volume of the accumulator must reflect the possible fraction of nitrogen in the mixture of the system. EXPERIMENTAL OBSERVATIONS

In experiments with a G-M type pulse tube refrigerator, cooling power versus composition of the mixture was measured for mixtures of helium and nitrogen in which the mole fraction of nitrogen varied from The test machine was a single-stage U-type pulse tube refrigerator with 2.0 kW input power. Figure 7 shows an example curve of cooling power versus temperature for a mixture of 97% He and 3%

compared with that for pure helium. The cooling power of the mixture is higher than that of pure helium when the cooling temperature is higher than the condensing temperature (71.8 K) at the corresponding partial pressure of nitrogen in the mixture. It is interesting that the liquid-solid phase change at 63.14 K is visible. The experimental results indicate that the machine can provide a cooling power at an approximate temperature platform around 63.14 K with nitrogen fractions up to 25%. We can also see in Fig. 7 that solid nitrogen begins to appear when the cooling temperature drops below 63.14 K. The solidification process reaches completion when the cooling temperature drops to around 58 K. Finally, the refrigeration process, when the refrigerant is reduced to almost pure helium, gradually approaches a minimum point of 35 K or so. Obviously, the possibility of producing solid nitrogen depends on the refrigeration ability of the cryocooler. Figure 8 shows a comparison of temperature distributions along the pulse tube and regenerator for pure helium and a mixture of 3% nitrogen and 97% helium, respectively. For the mixed fluid case, we can see from Fig. 8 that temperatures along about 30% of the length of the pulse tube are almost the same as those at the cold end of the tube. This means that some nitrogen in the mixture has been liquefied. Figure 9 shows temperature variations at the cold end of the pulse tube during re-heating after the machine was switched off. The pulse tube refrigerator was operated with a mixture of 12% nitrogen and 88% helium. In this example, the machine can keep the cooling temperature close to 64 K for about 4.3 minutes with a cooling power of 200 mW. The capacity of the

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Figure 8. Temperature distribution along pulse tube and regenerator.

Figure 9. Temperature variation of pulse tube refrigerator without accumulator in re-heating process.

accumulator, which is installed under the cold end of the pulse tube and can contain of liquid or solid nitrogen, can be calculated. The results indicate that 2.0 minutes of dwell at 63.16 K can be achieved for a 200 mW cooling power input for each of solid nitrogen. This means that 20 minutes of dwell can be achieved via the solid-liquid phase change at 200 mW cooling power for a accumulator. In conclusion, the pulse tube with the additional accumulator may provide a rather steady temperature near 64 K for about 24 minutes with an applied cooling load of 200 mW. CONCLUSIONS

Based on the above analysis, we can draw conclusions as follows: 1. A G-M type single-stage pulse tube refrigerator of 2.0 kW input power with a mixture of 3% N, and 97% He can provide a cooling capacity of 7 to 10.5 W at 63.5 K to 90 K. This is about 7% greater than would be obtained with pure helium. The liquid-solid phase change at 63.16 K and

below has been observed. 2. In our experiment, a pulse tube refrigerator without a frozen cryogen accumulator was able to

keep a temperature of about 64 K for about 4 minutes with a 200 mW load when the machine was switched off. Therefore, if a frozen cryogen accumulator is added, it may be able to meet the measurement requirements of SQUIDs by eliminating the magnetic noise caused by the running machine.

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3. Calculation results indicate that a frozen nitrogen accumulator, if installed in the cold head of the pulse tube, can provide approximately 200 mW of cooling power at 64 K for about

20 minutes for a SQUID application.

ACKNOWLEDGMENT The project was financially supported by the National Natural Sciences Foundation of China. The authors would appreciate the opportunity to work with Professor C. Heiden and G. Thummes in their laboratories at University of Giessen, Germany.

REFERENCES 1. ter Brake, H.J., et al., “Construction and Test of a High-T, SQUID-based Heart Scanner Cooled by Small Stirling Cryocoolers,” ICEC-17, Institute of Physics Publishing (1998), pp. 341-344. 2. van den Bosch, P.J., et al., “Cryogenic Design of a High-T, SQUID-based Heart Scanner Cooled by Small Stirling Cryocoolers,” Cryogenics, vol.37, no.3 (1997), pp.139-151.

3. Chen, G.B., et al., “Study on Two-component Gas Mixture in Regenerative Refrigerators,” ICEC-17, Institute of Physics Publishing (1998), pp. 197-200.

4. Chen, G.B., Gan, Z.H., Thummes, G., and Heiden, C., “Thermodynamic Performance Prediction of

Pulse Tube Refrigerators with Mixture Fluids,” to be published in Cryogenics (2000). 5. Gan, Z.H., Chen, G.B., Thummes, G. and Heiden, C., “Experimental Study on Pulse Tube Refrigerator with Helium-nitrogen Gas Mixture,” Cryocoolers 11, Kluwer Academic/Plenum Publishers, New York (2001). 6. J. A. McCormick, et al., “Design and Test of Low Capacity Reverse Brayton Cryocooler for Refrigeration at 35K and 60K,” Cryocoolers 10, Kluwer Academic/Plenum Publishers, New York (1999), pp. 421429.

Experimental Investigation of a Pulse Tube Refrigerator Driven by a Thermoacoustic

Prime Mover L.M. Qiu, G.B. Chen, N. Jiang, Y.L. Jiang and J.P. Yu

Cryogenics Lab, Zhejiang University

Hangzhou 310027, P.R.China

ABSTRACT

A breadboard pulse tube refrigerator driven by a standing-wave thermoacoustic prime mover has been set up to study the relationship among stack, regenerator, and working fluids.

The stack of the thermoacoustic prime mover is packed with dense-mesh wire screens because

of their low cost and ease of construction. The effect of packing factor in the stack on onset temperature, refrigeration temperature, and input power has been explored. An optimum packing factor of 1.15 pieces per millimeter has been found; this is an empirical value that provides a compromise between enhancing the thermoacoustic effect, and decreasing the heat conduction and fluid-friction losses along the stack. The pulse tube cooler driven by the thermoacoustic prime mover is able to obtain refrigeration temperatures as low as 138 K and 196 K with helium and nitrogen, respectively.

INTRODUCTION

Thermoacoustic engines can be an attractive alternative for specialized applications because of their simplicity, and their absence of lubrication, seals, and environmentally harmful working fluids.1-3 On the other hand, research and development of pulse tube refrigerators has reached such a stage that commercial products have been already developed for cooling temperatures around 77 K and 4.2 K, respectively.4,5 One may imagine that the reliability of a pulse tube refrigerator will be greatly improved if it is driven by a thermoacoustic prime mover. In the past decades, researchers have developed several kinds of stacks to enhance the thermoacoustic effect. Among these, stacks made of wire screen meshes have the merit of low heat conduction and reasonable efficiency. S.L. Zhou, Y. Matsubara and G.B. Chen have reported on measurements of thermoacoustic prime movers with stacks made of copper wire

mesh.6-8 They found that the overall performance of a thermoacoustic prime mover is mainly dependent on where is the thermal penetration depth, and is the hydraulic radius. Also, an optimized value of has been experimentally obtained.6 Additionally, the authors have recognized that the packing factor of the stack plays an important role in influencing the thermoacoustic effect of the prime mover and the refrigeration temperature of the pulse tube. Attempts to find an optimum packing factor will be of benefit to the design of stacks made of wire mesh. Therefore, experiments have been conCryocoolers 11, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2001

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Figure 1. Schematic diagram of the experimental apparatus.

ducted to determine the optimum packing factor. Meanwhile, the pressure characteristics of the acoustic prime mover have also been extensively studied. Finally, a refrigeration temperature as low as 138 K has been obtained using a thermoacoustic-driven pulse tube. EXPERIMENTAL APPARATUS

The experimental apparatus consists of a 1/2-wavelength, standing-wave thermoacoustic prime mover and a pulse tube unit as shown in Fig. 1. The thermoacoustic prime mover includes a resonator tube, hot buffers, hot heat exchangers, stacks, and cold heat exchangers. They are symmetrically arranged on both sides of the resonator tube as shown. The hot heat exchanger consists of a copper block with two internal heaters (about 200 W each), and an outside heater of about 400 W. A transformer adjusts the heating power delivered to the hot heat exchanger. The structure of the stack is similar to the regenerator used in pulse tube coolers. No. 6 and No. 10 copper mesh disks are packed in turn with a proportion of Various packing factors were tested to determine the optimum packing factor. The same copper mesh disks were also packed in the cold heat exchanger with a proportion of 2:1. The hot section of the thermoacoustic prime mover is insulated with ceramic fiber. The resonant tube is a stainless steel tube, 4 m in length. Dimensions of the thermoacoustic prime mover are presented in Table 1. The pulse tube unit, which is connected to the cold heat exchanger via a copper pipe as shown in Figure 1, has a coaxial configuration and works with a double-inlet. Some parameters and data of the pulse tube refrigerator are shown in Table 2. In the control and measuring system, a temperature controller and a voltage adjuster were used to control the heating temperature. The heating power is calculated from the resistance of

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the heaters and the applied voltage. The maximum heating temperature is limited to around 500°C because of the structure and silver-alloy brazing of the hot heat exchanger. The pressure oscillations were measured by a strain pressure sensor that is installed at the inlet of the pulse tube unit. The heating temperature of the prime mover and cooling temperature at the cold end of the pulse tube are measured by thermocouples and a Rh-Fe resistance sensor, respectively. The accuracy of the resistance sensor is approximately 0.1 K.

EXPERIMENTAL RESULTS Pressure Characteristics

The pressure characteristics of the prime mover connected to the pulse tube cooler were measured in the present work. Figure 2 shows the dependence of refrigeration temperature and heating temperature on helium charge pressure. We can see that, at a certain charge pressure, the heating temperature of the mover increases and the refrigeration temperature of the pulse tube decreases with increasing heating power. For a fixed input power, the higher the charge pressure of the system, the lower the refrigeration temperature that can be obtained. This can be explained from the characteristics of pressure ratio and pressure amplitude at the inlet of the pulse tube as shown in Figure 3. Note that the pressure amplitude increases with increased heating power, which leads to more mass flow taking part in the refrigeration process. Meanwhile, an increase in pressure ratio

Figure 2. Dependence of charge pressure on minimum temperature and heating temperature.

Figure 3. Pressure amplitude and pressure ratio vs. heating power.

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Figure 4. Dependence of charge pressure on onset temperature.

also improves the refrigeration performance of the pulse tube. Therefore, the prime mover has the advantage of being able to adjust both the pressure amplitude and the pressure ratio by adjusting the

heating power. By comparison, it can be difficult to increase the pressure ratio or amplitude of a system driven by some mechanical compressors.

Dependence of onset temperature on charge pressure is presented in Fig. 4. We can see that the onset temperature at a heating power of 294.5 W increases with the charge pressure. In summary,

our inhouse-fabricated thermoacoustic driver can generate a pressure ratio of about 1.1 for helium or nitrogen, which is adequate for a Stirling type pulse tube refrigerator.

Packing Factor of Stack Zhou and Matsubara found that the overall performance of the thermoacoustic prime mover is mainly dependent on They reported that a thermoacoustic prime mover with a wire mesh stack works well with near This result was also verified in our experiments. According to our measurements, the resonant frequency is about 70 Hz and 25 Hz for helium and nitrogen, respectively. Dimensions of the meshes of the stack used in our experiment are listed in Table 3; when the average temperature of the stack is 573 K, equals 0.37. Thermal penetration depths of helium and nitrogen are calculated and shown in Fig. 5. The packing factor of the stack is expressed as pieces per unit length for convenience. Experiments for finding the optimum packing factor were done using a constant charge pressure of 1.8 MPa and the same open ratios for the orifice and second inlet. The heating temperature was kept at by adjusting the input power; the working fluid was nitrogen. The temperature of the cooling water was about 300 K. Figure 6 shows the dependence of the minimum temperature and input power on packing factor. An optimum packing factor of 1.15 was obtained, corresponding to the minimum refrigeration temperature. If the packing factor is greater than 1.15, the minimum temperature will increase sharply because of the significant heat conduction and fluid friction losses along the stack. The temperature at the cold end of the stack was also measured to estimate the heat conduction along the stack (see Fig. 1). This temperature could be as high as 80-110°C for a heating temperature of 500°C. If the packing factor is smaller than 1.15, the minimum temperature changes smoothly.

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Figure 5. Calculated thermal penetration depth of helium and nitrogen.

Figure 6. Dependence of packing factor on minimum temperature and input power.

Figure 6 also shows that heating power increases with a decrease of the packing factor. It means that the smaller the packing factor, the larger the heating power that can be absorbed. The dependence on packing factor of onset temperature and total thermal energy consumed for onset was also measured for a heating power of 1347.2 W (see Fig. 7). We find that onset temperature decreases

Figure 7. Dependence of packing factor on onset temperature and minimum input power.

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Figure 8. Cool-down behavior of the pulse tube cooler.

with increase of packing factor. The trend of the total thermal energy consumed for onset is in accordance with the onset temperature.

Refrigeration Characteristics of the Pulse Tube Typical cool-down tests were carried out to study the refrigeration characteristics of the pulse tube driven by both the thermoacoustic prime mover and a mechanical compressor. Based on the above experiments, the packing factor of the stack was set at 1.15 pieces/mm. The operational parameters of the pulse tube cooler are listed in Table 4. Figure 8 shows the experimental results of the pulse tube unit driven by both the thermoacoustic prime mover and the compressor. We can see in Fig. 8 that the temperature at the cold end increases slightly in the first 19 minutes until the onset temperature of 390°C is reached. Then, the temperature decreases sharply and reaches the minimum temperature of 138 K within 150 minutes. In comparison, with the same pulse tube driven by the mechanical compressor with a swept volume of a lowest temperature of 74 K (64 K lower than the former) was obtained in 140 minutes. Clearly, the efficiency of the pulse tube refrigerator driven by the acoustic driver is still lower and must be improved. CONCLUSIONS 1. The experimental results show that our thermoacoustic prime mover can generate a pressure ratio of about 1.1 for nitrogen and helium, which is adequate for a Stirling type pulse tube cryocooler.

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2. An optimum packing factor of about 1.15 pieces/millimeter of the stack has been found experimentally; this corresponds to the No.6 and No. 10 copper mesh disks being packed in turn with a proportion of 1:2. 3. The pulse tube cooler driven by our thermoacoustic prime mover is able to obtain refrigeration temperatures as low as 138 K. ACKNOWLEDGMENTS

This work was financially supported by the National Natural Science Foundation of China. Z.H. Gan, K. Tang, W. Zhang and J. Yan are appreciated for their contributions to the experimental work. REFERENCES 1. Rott, N., “Thermoacoustics,” Advances in Applied Mechanics, vol.20 (1980), pp. 135-175. 2. Swift, G.W., “Thermoacoustic Engine,” Acoust. Soc. Am., vol. 84 (1988), p. 1145. 3. Swift, G.W., Radebaugh, R. and Matin, R.A., “Acoustic Cryocooler,” U. S. Patent, No. 4 953 366 (1990).

4. Radebaugh, R., “Recent Development in Cryocoolers,” Proceedings of 19th International Congress on Refrigeration, vol.3b (1995), pp. 973-988.

5. Chen, G.B., Qiu, L.M., Zheng, J.Y., Yan, P.D., Gan, Z.H., Bai, X., and Huang, Z.X., “Experimental Study on a Double-orifice Two-stage Pulse Tube Refrigerator,” Cryogenics, vol.37 (1997), pp. 271-

273. 6. Zhou, S.L., Matsubara, Y., “Experimental Research of Thermoacoustic Prime Mover,” Cryogenics,

vol.38, no.8 (1998), pp. 813-822. 7. Chen, G.B., Jin, T., et al., “Experimental Study on a Thermoacoustic Engine with Brass Screen Stack Matrix,” A.C.E., vol.43b (1998),pp. 713-718. 8. Bai, X., Jin, T. and Chen, G.B., “Experimental Study on a Thermoacoustic Prime Mover,” Proceed-

ings of ICCR ’98, (1998), pp. 522-525.

Design, Development, and Operation of a Thermo-Acoustic Refrigerator Cooling to below –60 °C M.E.H.Tijani, J. Zeegers, A.T.A.M. de Waele

Eindhoven University of Technology Low Temperature Group Eindhoven, The Netherlands

ABSTRACT

A thermoacoustic cooler, using a resonant standing acoustic wave, has been built. It employs a loudspeaker to sustain a standing wave in a resonance tube. In the helium-filled tube, a layered parallel-plate structure, called stack, and two heat exchangers are installed. The interaction of the compressed and expanded gas in the channels of the stack with the surface generates heat transport. A description of the cooler is presented, together with the first performance measurements using two different stacks, two different average pressures, and different dynamic pressures. INTRODUCTION

Over the past two decades, thermoacoustic cooling has been investigated as a new cooling technology.1-8 Thermoacoustic coolers can reach temperatures of –70°C and can have a coefficient of performance of 20 % of Carnot. Instead of CFKs, inert gases are used.

Typically, a thermoacoustic cooler consists of an acoustic resonator (e.g. tube) filled with an inert gas at some average pressure in which a structure with channels, called a stack, is placed. The stack is the heart of the cooler; it is where the heat transfer takes place. At both ends of the stack, heat exchangers are installed. The temperature of the hot heat exchanger is fixed at room temperature; at the cold heat exchanger, cooling power is generated. A modified loudspeaker generates sound in the form of a standing resonant wave. This wave causes the gas particles to oscillate while compressing and expanding. The thermoacoustic cooling cycle can be illustrated by considering a parcel of gas oscillating along the stack surface as a response to the standing wave, as illustrated in Fig. 1. During one period of the acoustic cycle, the parcel of gas undergoes two adiabatic steps (1 and 3), and two constant pressure heat transfer steps (2 and 4). In step 1, the parcel of gas moves forward, in the direction of lower pressure, expands, and cools. At this time, the parcel of gas is colder than the local stack surface, and heat transfer from the stack to the parcel takes place (step 2). In step 3, the parcel of gas moves back to its initial position, is compressed, and warms up. Now, in step 4, the parcel of gas is warmer than the local stack surface and heat flows from the parcel to the stack.

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Figure 1. (a) A typical gas parcel in a thermoacoustic cooler experiences a four-step cycle with two adiabatic steps (1 and 3) and two constant-pressure heat transfers steps (2 and 4); (b) An amount of heat is shuttled along the plate from one parcel of gas to the next; as a result, heat Q is transported from one end of the plate to the other using acoustic energy W to compress the gas.

At this stage the parcel of gas is returned to its initial position and the cycle starts again. Although the parcel excursion is much smaller than the length of the stack, the net effect of many parcels along the stack is that heat is transported from one end of the stack to the other (c.f. Fig.1b). If the hot end is fixed at room temperature, the other end cools down. Only the gas layer approximately within a distance of one thermal penetration depth from the stack's surface contributes to the thermoacoustic effect. The quantity is the distance across which heat can diffuse through the gas in a time where f is the acoustic frequency, is defined in terms of the thermal conductivity of the gas k, the gas density and its isobaric specific heat

In standing wave coolers, an important geometrical requirement is the transverse channel dimension in the stack, which amounts 1 to 4 times and depends on the used geometry.3,4 A detailed discussion of the theory of the thermoacoustic effect can be found in the literature.1-4

DESCRIPTION OF THE THERMOACOUSTIC COOLER

In Fig. 2, a schematic diagram of our thermoacoustic cooler is shown. It is a Hofler type cooler.8 Thermoacoustic theory3 was applied to the design this cooler. The cooler consists mainly of five parts: a loudspeaker, a helium-filled resonator, a stack, and two heat exchangers. From the loudspeaker, the fabric dome was cut off near the voice coil and replaced by a thin-walled light aluminium cone glued onto the voice coil. A rolling diaphragm is used to seal the resonator from the loudspeaker housing. Commercially available loudspeakers have an efficiency of 3-5%. The performance can be improved when a loudspeaker is coupled to an acoustic resonator, provided an appropriate selection is made of the parameters of the system. From a model simulating the coupling between the resonator

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Figure 2. Cross-sectional illustration of the thermoacoustic cooler, showing the different parts.

and the loudspeaker, it follows that the electroacoustic efficiency of the loudspeaker is maximum when the mechanical resonance of the loudspeaker is equal to the fundamental acoustic resonance of the resonator.6 The electroacoustic efficiency of the loudspeaker is defined as the ratio of the output acoustic power to the input electric power delivered to the voice coil. Because, in general,

the two resonant frequencies are different, a new concept is used successfully in our cooler to control the mechanical resonant frequency of the loudspeaker. The volume of gas behind the cone is used as an adjustable extra spring that adds to the spring of the loudspeaker. As a consequence, the resonant frequency of the loudspeaker can be tuned to the frequency of the resonator to obtain

high performance. We have used a system consisting of a cylindrical volume and a piston that can be positioned at different heights, and thereby changes the volume behind the cone. This is indicated in Fig. 2 as the cylinder and piston system. A detailed description of this tuning concept will be published elsewhere. The resonator was optimised for minimum viscous losses leading to a fundamental operating frequency of nearly 430 Hz. It consists of many parts. First is a copper flange, which contains the

hot heat exchanger. This flange is used to connect the resonator to the driver housing, and ensures a good thermal contact with the bottom of the loudspeaker housing through which cooling water

circulates. The part of the resonator which contains the stack has a low thermal conductivity to minimize the heat flow from the hot end to the cold end of the stack. Next, a copper contraction is used to connect the stack holder to the smaller copper resonator part and to reduce turbulence as the cross section changes. The cold heat exchanger is soldered to the neck. Finally, the small resonator part terminates in the stainless-steel buffer volume to complete the acoustically resonant system. The whole system is designed to maintain a static pressure of 12 bar. Copper tubes, in which cooling water circulates, are soldered in the bottom of the loudspeaker housing to remove the

thermoacoustic heat of the stack and of the loudspeaker. As discussed above, the transverse dimension of the channels in the stack is determined by the

thermal penetration depth

This is a function of the parameters of the type of gas and the acoustic

frequency. The cooler is designed to meet the requirements of a relatively large temperature span of 90°C over the stack and a cooling power of 4 watts. The parameters concerning the stack position,

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stack length, and layer spacing, were all taken into account for the design. The hot and cold heat exchangers are made of copper fins, 0.1 mm thick. They consist of sine channels, and have different sizes in the longitudinal direction. Two different stacks have been made. One stack is made by

winding spirally a long sheet of 0.06 mm thick plastic. Fishing lines, 0.3 mm thick, are used as spacers at 5-mm intervals, glued laterally onto the sheet. The second stack consists of Mylar parallel plates, 0.1 mm thick. The plates are also separated by 0.3-mm thick monofilament fishing line spacers, at an interval of 7 mm. The two stacks have a diameter of 38 mm and a length of 84 mm. The second stack was more difficult to construct, but it has a more uniform channel structure. An electrical heater (Fig. 2) was used to apply heat to the cooler in order to determine the coefficient of performance of the cooler at different temperatures. Thermometers are placed at the cold heat exchanger, at the hot heat exchanger, on the loudspeaker, and on the buffer volume. A more detailed description of the cooler will be published later. MEASUREMENTS AND RESULTS

The input acoustic power into the resonator is measured using a dynamic pressure transducer placed near the pusher cone and an accelerometer on the pusher cone7 (c.f. Fig. 2). By using two lock-in amplifiers we determined the dynamic pressure the velocity of the cone (u), and phase difference between them. The input acoustic power is given by

The performance of the cooler is described by the coefficient of performance,3 which is given by the ratio of the cooling load (including heat leak), Q, at the cold heat exchanger and acoustic power, W, delivered by the loudspeaker to the resonator

It is convenient to characterise the performance of the cooler by the coefficient of performance relative to Carnot’s coefficient of performance COPR, defined as 3

where the Carnot coefficient of performance COPC is given by 3

and are the temperatures of the hot and cold heat exchangers, respectively. The measurement procedure is as follows: the cooler is first evacuated and filled with helium up to the desired average pressure. Then, the vacuum vessel is evacuated until a good vacuum Pa) is reached. Finally a power amplifier, controlling the loudspeaker, is set at the desired amplitude of the dynamic pressure by amplifying the signal from a function generator. A series of measurements of the electroacoustic efficiency for different piston heights are shown in Fig. 3. Two peaks can be noticed in all plots: one peak at 430 Hz is due to the fundamental acoustic resonance and the other is due to the mechanical resonance of the loudspeaker. The peak due to the loudspeaker shifts as the height of the piston changes. The height can be varied from 31 mm to 123 mm. As the height becomes lower the included volume becomes smaller and the corresponding spring constant becomes larger. This results in a shift of the mechanical resonance of the loudspeaker to higher frequencies. A maximum efficiency is reached when the two peaks nearly match, at 72.5 mm. When the two frequencies are equal, the efficiency is constant over a wider frequency range, which is desirable, as the resonance frequency of the resonator decreases when the temperature decreases during the cooldown.

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Figure 3. Measured electroacoustic efficiency h as function of the frequency for nine different piston positions. A maximum efficiency is reached when the resonance frequency of the loudspeaker and that of the resonator are nearly equal at a height of 72.5 mm.

The loudspeaker used for these measurements had a large electro-mechanical damping. A higher efficiency of 30-40% is possible if an appropriate loudspeaker is used in combination with our

tuning concept. All performance measurements were made using helium as the working gas. The temperature of the cooler decreases until steady state conditions are established (c.f. Fig. 5). Then, stepwise, a heat load is applied to the cold end. Each time the steady state temperature is reached, a set of parameters is recorded, consisting of f, Q, P1, u, etc. After this set of performance measurements, parameters like static pressure and dynamic pressure can be changed for a new series.

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Figure 4. Measured performance of the thermoacoustic cooler with the first (coiled) stack. Figure 4 shows the performance measurements and the temperature difference, over the stack for the first stack (coiled sheet) for two static pressures, 5 and 10 bar, using three different dynamic pressures in each case. We have chosen to use the ratio of the dynamic pressure to the static pressure instead of the dynamic pressure as the parameter. In all cases, the temperature difference over the stack is a linear function of the heat load Q, and it increases as the dynamic pressure increases. The coefficient of performance relative to Carnot shows a maximum, which shifts to higher heat loads when the dynamic pressure increases. For the second stack, the behaviour of the temperatures of the hot heat exchanger

cold heat

exchanger and the buffer as function of time, are also plotted in Fig. 5a. The temperature of the hot heat exchanger is held constant by the cooling water. The whole part of the cooler below the cold heat exchanger cools down. As the temperature of the buffer lags that of the cold heat exchanger, we believe that the kink in at the time of nearly 200 minutes, can be attributed to convection in the resonator, which is set up as a consequence of the temperature gradient over the cold side. The long cooling time constant is due to the large mass of the cold side (1300 g). Only one measurement of the temperature difference and coefficient of performance relative to Carnot is available for the second stack at this time. The same behaviour can be concluded as for the first stack. The maximum coefficient of performance of the second stack for a mean pressure of 10 bar helium and a pressure ratio of 2.1% is 11.5% at a heat load of watts. CONCLUSIONS

A thermoacoustic cooler has been designed and built. The first measurements were successful. It cools down to below – 67°C and has a COPR of 11.5% for a static pressure of 10 bar helium and 2.1% pressure ratio. From the measurements, one can conclude that the parallel-plate stack has a higher performance than the coiled one. A number of improvements to the cooler are underway; a lighter aluminium resonator has recently been constructed. It weighs only 350 g instead of the present 1300g, which will decrease the thermal time constant and decrease the temperature gradient over the cold end to a low level.

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Figure 5. Some measurements of the thermoacoustic cooler with the second stack. a) Temperatures of the hot heat exchanger, cold heat exchanger and the buffer volume as function of time. b) Measured temperatures and performance as function of heat load.

This will also accelerate the measurements that last normally a day per heat load for low dynamic pressures. The use of other parallel plate stacks with different spacing and other gases is also planned for the near future. ACKNOWLEDGEMENT

We like to acknowledge the following persons: Lock Penders (electronics), Leo van Hout (engineering), the assistance of the department and central workshops of our university. We like to thank Greg Swift and Chris Espinoza of Los Alamos National Laboratories for their advise in the engineering of the parallel plates stack. We are much indebted to Guido d’Hoogh of Philips speaker systems for the development of the loudspeakers.

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REFERENCES 1. J. Wheatley, T. Hofler, G.W. Swift, and A. Migliori, “Understanding some simple phenomena in thermoacoustics with applications to acoustical heat engines,” Am. J. Phys, Vol. 53, no.2 (1985), pp. 147-162.

2. J. Wheatley, T. Hofler, G.W. Swift, and A. Migliori, “An intrinsically irreversible thermoacoustic heat engine,” J.Acoust.Soc.Am., Vol. 74, no.1 (1983), pp. 153-170. 3. G.W. Swift, “Thermoacoustic engines,” J. Acoust. Soc. Am. Vol. 84, no.4 (1988), pp.1146-1180.

4. G.W. Swift, “Thermoacoustic engines and refrigerators,” Encyclopaedia of Applied physics, Vol. 21, (1997), pp. 245-264. 5. G.W. Swift, “Thermoacoustic engines and refrigerators,” Physics Today, (1995), pp. 22-28.

6. S.L. Garrett, “ThermoAcoustic Life Science Refrigerator,” NASA Report, no. LS-10114 (1991). 7. T.J. Hofler, “accurate acoustic power measurements with a high-intensity driver,” J. Acoust. Soc. Am.

Vol. 83, no.2 (1988), pp. 777-786. 8. T.J. Hofler, “Thermoacoustic Refrigeration Design and Performance,” Ph.D. Thesis, Physics

Department, University of California, San Diego (1986).

Design of a Miniature Pulse Tube Refrigerator A. Halouane*, J-C Marechal and Y. Simon Ecole Normale Superieure, 24 rue Lhomond 75005 Paris, France * French Institute of Petroleum 92852 Rueil-Malmaison cedex, France

ABSTRACT Using a miniature pulse tube refrigerator and a RICOR compressor of 1cc swept volume, we have achieved a lowest temperature of 80 K at which the cooling power is 1W. The main purpose of this paper is to explain the rules used to design each component of a miniature PTR: regenerator, tube, reservoir, heat exchangers, and valves. The regenerator will be investigated in the following two ways: Firstly, a complete hydrodynamic study (based on mass and momentum conservation). Secondly, a thermal study of the system based on energy conservation in the gas and the matrix. These allowed the computation of the gas temperature profile in the regenerator and in the tube. The losses of the system were deduced. INTRODUCTION In an OPTR / DIPTR Pulse Tube Refrigerator1 helium gas is periodically compressed and expanded in a closed system; the main part is a simple tube ended by two heat exchangers: cold and hot. During compression, the gas, before entering the tube, has passed through a regenerator filled with wire screens, whose role is to bring the gas from room temperature to the cold temperature During expansion, the same regenerator brings the gas from to The gas is pulsed at the entrance of the regenerator by a oscillating piston compressor.2 The system (tube, regenerator, and heat exchangers) is in a vacuum chamber. CHARACTERISTICS OF THE MINIATURE OPTR The piston of the compressor, driven in a cyclic manner, sweeps (under a process somewhere between isothermal and adiabatic) a volume at a frequency The swept volume amplitude is around An aftercooler (heat exchanger with a dead volume of 0.4 is inserted between the compressor and the regenerator to cool the gas down to room temperature The bronze wire screens are packed along the regenerator with a cross-section of The open gas section is A = n S, where n = 0.61 represents the porosity. Another parameter called hydraulic radius, , defined as the ratio (open volume / exchange section) gives an idea about the dimensions of the pores. In the present case Cryocoolers 11, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2001

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Figure 1. Description of the PTR.

The thickness of the stainless steel tube is about 0.2 mm, its length is 72 mm, and its diameter is 4 mm Both heat exchangers are made with the same copper wire screens soldered at the periphery of the tube over a length of 6 mm, (n=05, each has a dead volume of The orifice is a needle valve, and the reservoir volume is bigger than the whole system volume as shown in Fig. 1. In Fig. 2 we represent the temperature profile of the system obtained using the theoretical model and we will compare it to the experimental results. The monatomic gas used is helium; given its unity mass one can write:

and the specific heats:

Figure 2. Temperature profile of the PTR.

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Figure 3. Electrical analogy of the PTR.

The amplitude of the mass flow rate in a cycle is an important parameter to enhance the refrigeration power. Working at high pressure, this will increase automatically the mass flow

rate especially in miniature PTR. In our case the main pressure is : The uniform pressure in the tube is the reference, where At the entrance of the regenerator the pressure amplitude is bigger and the phase angle has a lead over all others To all these measured parameters we can also add the measure of the reservoir pressure amplitude which allows us to compute the mass flow rate value through the orifice:

is the equivalent capacity of the reservoir in an electrical analogy pressure/voltage and mass flow rate/ current( 3) , (6 ) (see Fig. 3). Through the orifice and after linearizing the relation one can write :

From the measure of the pressure amplitude of the reservoir , relation :

we get

using the

Typically and The mass flow rate at the entrance of the hot heat exchanger is a bit different from because of the dead volume. Since the transformation in the compressor could be adiabatic, isothermal or just between them, the mass flow rate flowing out of the compressor will be expressed by:

takes into account the mean value of the compressor and all dead volumes located between the compressor and the regenerator entrance.

Hysteresis effect in the tube At the hot end of the tube (respectively the cold end): at the instant t’ the gas enters in and goes up to a maximum distance The mass

the tube with a temperature

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conservation of gas in the tube allows the computation of this ratio of maximum penetration depth

where :

is the capacity of the tube.

For any edge x(t), we can write the relation between the position and the pressure :

In this approximation «sinusoidal waves», the mass flow rate at the cold end is :

We find:

with: At the cold end the « ratio » of maximum penetration depth is given by

At the moment t the gas will leave the hot end with a temperature T(t) linked to the temperature T(t’) by the isentropic relation :

According to the first law of thermodynamics the heat evacuated exchanger is expressed as :

at the hot heat

In the case of small sinusoidal amplitudes :

When the system reaches the coldest temperature, and since the system is in a vacuum chamber, the « energy flux » is constant along the regenerator and the tube

is the term representing the regenerator losses takes into account the cryocooler losses and is the conduction loss of the tube Since is small compared to then the regenerator losses are equivalent to the energy flux along the system

DESIGN OF A MINIATURE PULSE TUBE REFRIGERATOR

321

Figure 4. Flow through a porous media.

STUDY OF THE REGENERATOR

Applying a small perturbation around the equilibrium point with for a cyclic steady operation, the thermodynamic parameters P(x,t) and T(x,t) should be decomposed as

with

Starting with the hydrodynamic analysis we can write the following equations of conservation in the regenerator: Mass conservation In the porous media. Figure 4, the mass flow rate is related to the mean velocity u and

the gas density ρ by the relation

The mass conservation equation is then :

Momentum conservation

We can write, after simplification, the momentum conservation :

where R is the resistance per unit length of the regenerator expressed by(4), (5) :

a and b depends on the geometry of the wire matrix and it porosity (in our case a = 44 and the Reynolds number.

b = 0.4) . µ is the dynamic viscosity of the gas and

In order to study the thermal behaviour of the regenerator ; we should know firstly the heat exchange between the gas and the matrix (matrix = regenerator wire screens). This work was done by Tanaka and all:

and

is the thermal conductivity of the gas.

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PULSE TUBE ANALYSES AND EXPERIMENTAL MEASUREMENTS

HEAT EQUATION

For an ideal gas

the heat equation is expressed as :

is the matrix temperature, the term sources can be written :

where ( div q ) is the term

and the viscous dissipation

is represented by

is a numerical parameter 1-Analysis of steady terms in the Heat equation

Steady terms obtained from the heat equation can be written as follows:

let us compare different terms of this equation at the hot end of the regenerator; the same work

was done at the cold part and gave same conclusions. At the hot end the term and

represent exactly the same expression:

and calculation shows that

A simple

is:

from this result we can assume that the mean temperature of the gas and the matrix are the same in each point x along the regenerator. 2-Analysis of unsteady terms of the Heat equation Unsteady terms obtained from heat equation allow us to write the following equation

let us compare different terms of this equation at the hot end of the regenerator (respectively at

the cold side).

and All these terms have the same order so we will not neglect any

one of them.

DESIGN OF A MINIATURE PULSE TUBE REFRIGERATOR

323

ENERGY EQUATION OF REGENERATOR MATRIX Between sections x and x + dx of the regenerator, the matrix occupies a volume A dx (1-n)/n. Its heat capacity per unit volume is matrix density and mass heat capacity of the matrix); The longitudinal conduction of the matrix and the tube is represented by: taking into account the stainless steal tube conduction and the matrix conduction is a coefficient depending on diameter

and

thickness of the regenerator tube. As in the heat equation, we deduce the energy conservation equation of the regenerator matrix.

Global Equation (gas - matrix) Combining the equation of heat conservation (24) (unsteady terms) with equation (25)

(of matrix conservation of energy) one obtains:

To this equation we can add the conservation of energy along the regenerator :

Let us introduce the complex expression of the following parameters at the first order development : and

and the time constant

with an effective specific heat

of the matrix :

Combining equations (27) and (28) we obtain the relation between the local temperature gradient

where,

and the energy flux

:

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PULSE TUBE ANALYSES AND EXPERIMENTAL MEASUREMENTS

Figure 5. Temperature profile in the regenerator.

COMPUTER MODELLING8 The pressure in the tube

and the pressure in the reservoir

with a common phase angle will be our reference zero. The couple

is completely

defined at the cold end of the regenerator. Applying the electrical analogy we start our obtained for example when we suppose that the temperature profile is constant along the regenerator. From a cell (i=1) to the next one (i+1), starting at the cold end, we arrive at the entrance of the regenerator where we compare with the experimental values of (pressure amplitude), (angle phase), (compressor swept volume) and If these parameters correspond to the experimental data we stop the program and we print at the same time the temperature profile and the regenerator losses ; (see Figure 5). program with an

The ratio between experimental and theoretical regenerator heat loses is :

STUDY OF THE TUBE7

The heat exchanged over a cycle between the gas and the tube-wall has a very large affect on the refrigeration power of the system. In order to study the tube losses we can follow the same

work done with the regenerator. Considering the motion of the gas in the tube, Equations (8) and (12) allow us to compute point by point according to x, for different time instants (where i= 1 to 10) regularly spaced in a period the instantaneous temperature profile of the gas in the tube. Figure 6 shows the instantaneous temperature profile at the hot and cold ends of the tube. Since the buffer gas never leaves the tube, the mean temperature profile can be estimated to

Figure 6. Instantaneous temperature profile at the hot and cold end.

DESIGN OF A MINIATURE PULSE TUBE REFRIGERATOR

325

Figure 7. Mean temperature profile of the gas in the tube.

Figure 8. Mean temperature profile of the tube and the gas.

Figure 9. Mean temperature profile of the tube (gas + tube).

be linear. The mean temperature profile of the gas at a moment The thermal conduction along the stainless steel tube is:

in the tube is shown in Fig. 7.

Figure 8 shows both the mean temperature profile of the tube alone, due to thermal conduction, and the gas temperature profile. In Fig. 9, the temperature profile of the tube (gas + tube) is compared to the experimental results and shows that the theoretically predicted curves fit well with the experimental results. DISCUSSION

In this work we have shown that knowing the real losses of the regenerater requires knowledge of the real temperature profile. The optimization is, then, easier. We changed the geometry and nature of the matrix and selected the best one. The regenerator tube volume was twice the com-

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PULSE TUBE ANALYSES AND EXPERIMENTAL MEASUREMENTS

pressor swept volume. For the tube, the most important thing was that its volume is near the volume of the compressor; thus the penetration depth of the gas into the tube at both ends never exceeded 20 to 30%. If the penetrartion was much greater, the gas coming from the cold side could travel all the way to the hot heat exchanger, which would seriously decrease the efficiency. The mixed gas chamber in the valve system must be as small as possible so as to reduce the dead volume of the needle valves. The volume of the reservoir was about 100 times greater than the compressor volume. For miniature pulse tube refrigerators, we have to focus our work on reducing every dead volume in the system.

REFERENCES 1. Radebaugh, R., Zimerman, J., Smith, D. R. and Louie, B., “A Comparison of Three Types of Pulse Tube Refrigerators: New Methods for Reaching 60 K,” Advances in Cryogenic Engineering, vol. 31, Plenum Press, New York (1986), p. 779. 2. Roach, P., Kashani, A. and Lee J. M., “Theoretical Analysis of a Pulse Tube Refrigerator,” Advances in Cryogenic Engineering, 41 B (1995), p. 1357. 3. Huang, B.J and Chuang, M.D., “System Design of Orifice Pulse Tube Refrigerator Using Linear

Network Analysis,” Cryogenics, vol. 36, no. 11 (1996), pp.1357. 4. Tanaka, M., Yamashita, I. and Chisaka, F., “Flow and Heat Transfer Characteristics of Stirling Engine Regenerator in an Oscillating Flow,” JSME Int Series II (1990). 5. Kays, W., M. and London, A. L., “Compact Heat Exchangers,” Mc Graw-Hill Book Company. 6. Halouane, A., Marechal, J.C. and David, M., “Study by Electrical Analogy of Different Pulse Tubes Refrigerators: Model and Experiments,” The European Physical Journal Applied Physics, vol. 4, (1998), pp. 31-35. 7. Halouane, A., “Hydrodynamic and Thermal Study of the Pulse Tube Refrigerator: Model and Experiments,” PhD Paris Vl University, April 1999, (in French ). 8. Halouane, A., Marechal, J.C., David, M. and Simon, Y., “Study of Regenerator Temperature Gradient in a Pulse Tube Refrigerator,” Proceedings ICEC 17, IOP (1998), pp. 227-235.

Investigation of a Single Stage Four-Valve Pulse Tube Refrigerator for High Cooling Power T. Schmauder, A. Waldauf, M. Thürk, R. Wagner and P. Seidel Institut für Festkörperphysik, Friedrich-Schiller-Universität Jena D-07743 Jena, Germany

ABSTRACT

We discuss the optimization of a pulse tube refrigerator for high cooling power. Our approach is to increase the system efficiency by analyzing and reducing the various loss mechanisms. Because stationary losses (such as radiation and thermal conduction in the system) as well as design principles for the regenerator are well understood, our main effort is focused on controlling the flow behaviour of the working gas at the various tube connections between the components. For time resolved measurements of the gas velocity and gas temperature we use hot wire anemometry and thermocouples respectively. The results of this analysis are used to improve the design especially of the cold head heat exchanger and the hot end setup of the pulse tube. Despite the consequent separation of the in- and outlet gas at the hot end of the pulse tube we find a strong hot end loss caused even by very simple flow parallelizing devices at the hot end of the pulse. A cooling power of 67 W at 70 K has been achieved. The aim of this project is a cooling power of 100 W at 80 K for thermal shielding for magnets and for cryopumps.

INTRODUCTION

Today there are various applications for cryorefrigerators such as sensors, electronics and superconductors cooling and cryopumping, thermal shields cooling or even cryogenic fluids liquification for the cooling of large high-field magnets. Low cost and high reliability as well as low interference make the pulse tube a suitable cooler for many applications. Amongst the various types of pulse tube refrigerators, the four-valve pulse tube refrigerator (FVPTR) has great potential for being a highly efficient pulse tube cooler.1 Compared with the well known explanation of the basic phenomena inside of an ideal pulse tube refrigerator,2-6 a clear explanation of the real effects has not been given even though a few experimental studies have been reported,7-10 in which the intrinsic behaviour of the orifice pulse tube refrigerators was investigated. Nevertheless, open questions about the special intrinsic effects inside the FVPTR remain. The idealized working principle of the FVPTR is illustrated in Figure 1. The working fluid (usually Helium) is compressed by an external compressor and enters the pulse tube at the cold end after being cooled to the cold head temperature Tcold while passing the regeneraCryocoolers 11, edited by R.G. Ross, Jr.

Kluwer Academic/Plenum Publishers, 2001

327

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PULSE TUBE ANALYSES AND EXPERIMENTAL MEASUREMENTS

Figure 1. Working cycle of the four-valve pulse tube refrigerator (FVPTR): A) a) to e):

temperature distribution along regenerator and pulse tube for different phases of the cycle, B) pressure wave during the cycle.

tor (Figure 1a). The gas in the pulse tube is adiabatically compressed from low pressure to high pressure and heats up by a factor . In the next phase (see Figure 1b) the part of the working gas with a higher than ambient temperature is removed from the system at the hot end of the pulse tube. Next, the gas in the pulse tube is

expanded towards the cold end through the regenerator, adiabatically cooling below before passing the cold end heat exchanger (HX). Finally, the amount of gas, which was removed in the second phase of the working cycle is replaced. However, the replaced gas has only temperature while the removed gas had a higher temperature, thus removing an amount of heat from the system and generating a cooling power. THEORETICAL CONSIDERATIONS

Generated Cooling Power The heat removed during one cycle from a machine running a gas cycle process as described above for the FVPTR is:

Eq. (1) can be analysed in different ways: Separation in phases of the working cycle. In this case the cycle is broken up into easily described steps. For the FVPTR cycle as described above, Eq. (1) reduces to:

INVESTIGATION OF 1 -STAGE 4-VALVE PT REFRIGERATOR

specific heat of working gas,

(M: molar mass of working gas, exchange ratio) and temperatures:

329

amount of heat exchanged) with exchanged mass

gas constant,

pulse tube volume,

Eq. (2) to (4) yield

Obviously, large values of pressure ratio, pulse tube volume and exchange ratio are desirable. Thermoaccoustic model. A different approach to analyse eq. (1) is to treat volume flow and pressure wave in a harmonic approximation. The work flow at the cold end of the pulse tube is then11: phase angle between pressure and volume wave, amplitudes of pressure and volume waves, respectively). The phase angle is crucial for the cooling power. It is not only influenced by the construction details of regenerator and heat exchanger, but also by the timing of the gas exchange at the hot end. Detailed investigations based on phasor diagrams are given in Ref. 4.

Loss Mechanisms There are a number of obvious loss mechanisms in the pulse tube cryocooler. Some of them are very well understood and routinely considered when designing a cryocooler setup.5 These include losses due to heat conduction and radiation. Losses due to the finite thermal capacity of the regenerator are also well understood and we considered them when designing our regenerators6. There remain additional losses which are not quantitatively understood yet and on which our experimental efforts will be focussed: Turbulence Generation at Tube Connections between control valves, regenerator, heat exchanger and pulse tube. As described above, the working principle of the cooler relies on a stable temperature separation of hot and cold gas in the pulse tube. Turbulences in the gas streams entering or leaving the tube disturb this temperature separation and thus reduce the net cooling power. On the other hand a certain degree of turbulence helps to improve heat

transfer in regenerator and the heat exchanger. Dead Volume and Hot End Exchange Volume. The mass of working gas exchanged at the hot end of the pulse tube requires work from the compressor for its displacement. In contrast to orifice/buffer pulse tube refrigerators, this work is lost and can not be recovered when the displacement is reversed later during the cycle. For a large exchange volume in Eq. (3)) the additional load for the compressor may overweight the gain in cooling power as expected from Equation (5). A similar problem arises for the regenerator construction: Low pressure drop over the regenerator requires a large effective diameter, but this increases the dead volume in which gas is compressed but does not contribute to cooling power generation.

330

PULSE TUBE ANALYSES AND EXPERIMENTAL MEASUREMENTS

Hot End Loss. Gas leaving the pulse tube at the hot end will heat up the flow channel it passes. When the replaced gas enters later through the same flow channels, it will collect this heat and bring it back into the pulse tube. This is not desired since it reduces the amount of heat in Eqs. (1) to (6)) removed from the system during the work cycle. Thus, the design of the hot end of the pulse tube is crucial for the operation of the FVPTR. EXPERIMENTAL SETUP Refrigerator Setup

A FVPTR testbed has been set up which allows one to quickly exchange the main components including the pulse tube, cold heat exchanger, and regenerator, to vary the timing of the regenerator and hot end inlet and outlet valves, and to monitor the temperatures and pressures at various points of the cryocooler. A commercial Helium compressor (CTI, model 9600) with 5.5kW input power was used. Figure 2 A) shows a photograph of the testbed.

A schematic overview of the setup is given in Figure 2 B). We note that the inlet and outlet channels at the hot end of the pulse tube are strictly separated to avoid hot end losses due to regenerative effects (hot gas leaving the pulse tube heats the outlet channel walls which would return the excess heat to the entering replacement gas later during the work cycle). This construction enables us to avoid the use of a hot end heat exchanger. Thus, the refrigera-

tor design remains very simple since only the helium supply lines and the control valve power line, but no active cooling, are necessary on the refrigerator unit. Hot Wire Anemometry To understand and minimize losses due to turbulence at the junctions of the various components, we conducted hot wire anemometry (HWA) studies. The setup for these measurements is shown schematically in Figure 3a). A continuous helium flow at 8 bar was used as the working gas. Measurements of the gas velocity have been carried out within the pulse

tube at 6 points along the longitudinal axis using hot wire anemometers which were able to scan over the tube diameter. The response time of the HWA-sensor is about 500µs (2kHz). As indicated in Figure 3b) various designs of flow straighteners for the hot end, the

Figure 2. A) photograph of our FVPTR test bed, and B) schematic of the setup indicating probe positions for temperature measurements (Si diodes and thermocouples) and pressure measurements (piezoelectric sensors).

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331

Figure 3. Experimental setup for hot wire anemometry measurements: a) schematic of the entire apparatus, b) the inlet head allows for various configurations by exchanging insets representing regenerator and pulse tube junctions.

cold end, and the regenerator intake manifold were tested by introducing the constructions in the changeable inlet head on top of the apparatus. RESULTS AND DISCUSSION Pneumatic Flow Optimization

Typical velocity profiles of the working gas right behind the intake design of the hot end of the pulse tube, the cold end of the pulse tube and the regenerator are documented in Figure 4a, b and c, respectively. An intake design without any flow straightener is leading to a jet stream in the centre of the tube ( ´s in Figure 4a). The gas penetrates more deeply into the tube increasing the turbulence. This results in a high convective heat transfer between the gas elements and high axial heat loss, respectively. A noticeable compensation of the mismatched flow is reached by using a stack made from a couple of screens (+ and curves in Figure 4a). On the other hand we have to reduce thermal mass at the hot end of the pulse tube to prevent hot end loss9. As pointed out above, however, the losses due to flow instabilities increase for an unfavourable designed intake device. A compromise is to be sought. An efficient flow straightener design with low thermal mass could be a thin plate perforated with fine holes ( in Figure 4a). At the cold end of the pulse tube the working gas stream is perpendicular to the tube axis. This results in asymmetrical distribution of the flow as it is displayed in Figure 4b. As the cold heat exchanger made of a stack of more than 20 screens also acts as flow straightener, the velocity profile of the gas entering the tube from the regenerator is flat and the incident flow is laminar. In order to minimize regenerator losses a uniform fluid distribution over the cross section of the regenerator is necessary. The -curve in Figure 4c indicates an asymmetrical pattern of the flow distribution over the cross section of the regenerator due to the lack of any plenum chamber at the end of the regenerator. On the other hand a too large volume of the plenum chamber means a large compression and expansion of the remaining gas. This does not contribute to the cooling power, but the necessary compression power reduces the COP; again a compromise is to be sought. A sufficient balance of the flow was

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PULSE TUBE ANALYSES AND EXPERIMENTAL MEASUREMENTS

Figure 4. Anemometry measurements of the flow velocity a) at the hot end,and b) at the cold end of the pulse tube, c) at the regenerator inlet. Solid lines are to guide the eye.

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333

Figure 5. Pressure and temperature waves during the working cycle at different positions along

the pulse tube for optimized (maximum cooling power) exchange volume and dc-flow.

achieved by using a plenum spacer with a height of 3 mm. The high velocities near the wall indicate reduced mesh density due to a small gap between the stack and the regenerator housing which can be prevented by using slightly overdimensioned discs. Refrigerator Operation As discussed above, the exchanged volume has significant influence on the cooling power: according to Eq.(5), a large exchange volume is desired, while due to dead volume loss this volume needs to be limited. Similarly, the balance between the removed and replaced amount of gas (referred to as dc-flow) will influence the regenerator efficiency and the coolers performance. Thus we studied the performance of the FVPTR in dependence of these two parameters and dc-flow which we control using needle valves at the hot end in- and outlet. For studies of the heat transfer within the pulse tube we use thermocouple probes for time resolved measurements of the gas temperature at various positions along the pulse tube. Figure 5 shows these temperature curves together synchronised to the respective pressure waves. It is obvious that the steepest temperature gradient is in the lower half of the tube and a huge temperature amplitude of more than 100K (peak-peak) is observed. Surprisingly, no indication of the temperature drop due to the replacement gas intake is seen even very close to the hot end flow straightener. For cryocoolers of high cooling power, the cold heat exchanger becomes a critical component. In the pulse tube refrigerator this heat exchanger also serves as flow conditioner at the cold end. We performed measurements of the temperature drop as well between the cold head housing and the wires of the copper gauze serving as exchange surface as of the drop between those wires and the working gas. The results of these measurements are given in Figure 6A. To reduce the temperature drop per transferred heat

power we improved the contact between the cold head housing and the wire gauze disks

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PULSE TUBE ANALYSES AND EXPERIMENTAL MEASUREMENTS

Figure 6. Transferrable power vs. temperature drop over the cold heat exchanger: a) for the

original cold head heat exchanger, b) for the improved heat exchanger with soldered gauze discs.

and further increased the exchange surface. The new heat exchanger (Figure 6B) has a significant reduced temperature drop per transferred power of 0.11 K/W. Hot End Design

We were not satisfied with the available cooling power of the test FVPTR even with pneumatically optimized component junctions, improved heat exchanger and optimal exchange volume and dc-flow. From the available data we assumed a problem in the hot end flow conditioner: No sign of temperature change is recorded even very close to the flow conditioner disk inside the pulse tube during the phase of refilling the (according to our view of the working cycle relatively cold) exchange gas from the hot end. In addition we see a strong overheating of the hot end of the pulse tube during cooldown time (Figure 7). The heat is not efficiently removed from the hot end. To overcome this problem, a variety of flow conditioning insets were tested at the hot end of the pulse tube. Figure 8 summarizes the results. The originally used punched disc (Figure 8 a and c) causes a phase shift between the temperature wave above and below the disc. The result is also a shift in the volume wave due to

Figure 7. Temperature vs. time development during the cooldown phase of the refrigerator at

various points of regenerator and pulse tube. Note the strong overheating of the hot end during initial cooldown and during operation under heat load at the cold head!

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335

Figure 8. Temperature waves at the hot end inside the pulse tube and behind the respective

flow parallelizer: a) and b): experimental setup, c):punched disk as flow straightener, d): 4, and e):12 wire gauze disks. Note the strong and very setup dependent phase shift between pressure and temperature wave caused by the flow conditioners!

(ideal gas, assuming p=const.), and finally a change in the phase shift

in Eq. (6).

Alternative flow conditioners at the hot end are stacks of wire gauze dics (Figure 8b, d, e, compare also Figure 4a). The phase shift of the temperature wave over this stack depends strongly on the number of discs used. Almost no shift is seen for 4 discs (Figure 8D) and

almost 180° shift for 12 disks (Figure 8E). The regenerator performance behaves similar: both, the ultimate cold head temperature (35 K and 29 K respectively) and the cooling

power at 80K (30 W and 70 W respectively) improve strongly when changing from 4 to 12 discs. Further increase of the number of discs in the hot end flow straightener does not improve refrigerator performance.

Figure 9. Power characteristic of the FVPTR with different flow straighteners at the hot end of the pulse tube.

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PULSE TUBE ANALYSES AND EXPERIMENTAL MEASUREMENTS

Figure 9 contains the cooling power characteristics of the cooler with the punched disc

(dashed line) and the stacked gauze hot end flow straightener (solid line). While the available cooling power is significantly increased, we are still concerned about the

regenerative effect of the new flow conditioner: the temperature waves of the gas inside the disc stack in Figure 8d) and e) show a very small amplitude, indicating strong hot end loss. CONCLUSIONS

A single stage FVPTR was designed and constructed. As the preliminary experiment the performance of the pulse tube was investigated. The essential results can be summarized as follows: z The minimum no-load temperature achieved is 28 K. The observed net cooling power at 70 K is 67 W. z The influence of several parameters on the FVPTR performance has been studied to test their sensitivity for potential modifications in future improved designs. z

The hot-end loss due to induced regenerative effects is incompatible with the adiabaticity

required for the maximum enthalpy flow, and thus is a main loss source. A successful design of the hot end of the pulse tube has to match the requirements of low mixing and low turbulence of the gas by using as little thermal mass as possible for the insets in order to prevent hot end losses. The current work has mainly focused on the reduction of these two types of losses. ACKNOWLEDGMENT

The authors wish to acknowledge the financial support provided for this project by the German BMBF under grant number FKZ 13 N 7395. REFERENCES 1. Blaurock.J., Hackenberger, R. Seidel, P. and Thürk, M., “Compact Four-Valve Pulse Tube Refrigerator in Coaxial Configuration,” Cryocoolers 8, Plenum Press, New York (1995), pp. 395-401. 2. Storch, P.J. and Radebaugh, R., “development and experimental test of an analytical model of the orifice pulse tube,” Adv. Cryog. Eng., vol. 33 (1988), p 851. 3. Radebaugh, R., “A review of pulse tube refrigeration,” Adv. Cryog. Eng., vol. 35 (1990), p. 1191. 4.

5. 6. 7.

8.

9.

Hoffmann, A., Wild, S., “A Model for Analyzing Ideal Double Inlet Pulse Tube Refrigerators,” Cryocoolers 8, Plenum Press, New York (1995), pp. 371-381. Walker, G., Cryocoolers, Part I:Fundamentals, Plenum Press, New York (1983). Ackermann, R.A., Cryogenic regenerative heat exchangers, Plenum Press, New York (1997). Lee, J.M., Kittel, P., Timmerhaus, K.D., Radebaugh, R. “Flow patterns intrinsic to the pulse tube refrigerator,” Proceedings of the 7th Int. Cryocooler Conf., PL-CP-93-1001, Part 1, Kirtland AFB (1993), pp. 125-139. David, M., Merechal, J.-C. and Encrenaz, P., “Measurements of instantaneous gas velocity and temperature in a pulse tube refrigerator,” Adv. Cryog. Eng., vol. 37 (1992), pp. 939-943. Gerster, J., Thürk, M., Reißig, L., Seidel, P., “Hot end loss at pulse tube refrigerators,” Cryogenics, vol. 38 (1998), pp. 679-682.

10. Thürk. M., Brehm, H., Wagner, R., Gerster, J., Seidel, P., “Intrinsic behaviour of a four valve pulse tube refrigerator,” Proc. ICEC 16/ ICMC, Elsevier Science, Amsterdam (1996), pp. 259-262. 11. Xiao, J.H., “Thermoacoustic heat transportation and energy transformation; part 1: formulation of the problem,” Cryogenics, vol. 35, no. 1 (1995), pp. 15-19.

Analysis and Experimental Research of a Multi-Bypass Version Pulse Tube Refrigerator L.W. Yang, J.T. Liang and Y. Zhou

Cryogenic Laboratory Chinese Academy of Sciences

Beijing 100080, China

ABSTRACT

One important configuration of a pulse tube refrigerator is the multi-bypass pulse tube refrigerator (MPTR). This paper analyzes the working processes of this PT configuration and presents some experimental results taken using operating frequencies below 5 Hz. The analysis shows that a MPTR is different from a general multi-stage cooler, but works somewhat like a multi-stage

cooler to a certain extent. Experimental results are presented and compared to former results. The experiments demonstrate that the multi-bypass channel leads to an improvement in performance, and that the function of the multi-bypass more resembles a double-inlet than the first-stage of a multi-stage cooler. INTRODUCTION

The multi-bypass pulse tube refrigerator (MPTR) was invented by Y. Zhou in 1993.1 As shown in Fig. 1, the MPTR contains not only an orifice valve and double-inlet valve at the pulse tube hot end, but also a connection between the middle of the regenerator and the middle of the pulse tube; this forms a new structure. In the figure, both the pulse tube and the regenerator have two parts, just like a two-stage pulse tube cooler. In the past, experimental investigations of MPTRs were mainly carried out using a valveless compressor and an operating frequency higher than

These tests showed that the MPTR

structure is effective at lowering the refrigeration temperature. However, experiments at lower

Figure 1. Multi-bypass version pulse tube refrigerator. Cryocoolers 11, edited by R.G. Ross, Jr.

Kluwer Academic/Plenum Publishers, 2001

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PULSE TUBE ANALYSES AND EXPERIMENTAL MEASUREMENTS

Figure 2. Difference between gas piston and actual piston.

frequencies are not available. The objective of this paper is to analyze and do experiments using operating frequencies generally lower than 5 Hz to lay the foundation for further research. ANALYSIS Actual Piston and Gas Piston

There are different viewpoints as to the working mechanism of a multi-bypass pulse tube refrigerator. Some investigators treat it as a special version of a multi-stage cooler,4 and some treat

it just like another double-inlet PT, but at a lower temperature.3 The function of the double-inlet is easy to understand from Figure 1, where the multi-bypass and double-inlet are similar with just different positions. When a MPTR is compared with a cold head of a two-stage G-M refrigerator, its functional similarity to a multi-stage cooler is also easy to understand, as shown in Figure 2. In Figure 2 (A), with an actual moving piston, three moving volumes are formed: at the hot end, at the middle part, and at the cold end. By controlling the phase between the pressure wave and the piston motion, refrigeration is developed in the two lower-temperature volumes. A similar process take place in the MPTR, as shown in Figure 2 (B). With the assumption of a gas piston existing in the pulse tube, three moving volumes are formed, just like with a mechanical piston. However, considering the compressibility and fluidity of the gas piston, the middle chamber in the pulse tube may exhibit other effects that make the actual process more complicated than a simple mechanical piston. Figure 2 (C) shows this feature. The gas piston in the pulse tube is actually divided into two parts by a multi-bypass flow chamber. This assumption makes the multi-bypass resemble a double-inlet. For a G-M refrigerator, as shown in Figure 2 (A), the middle chamber volume change is predetermined by the piston area and its stroke. And generally, the flow channel from the regenerator to the middle chamber is made large enough to minimize flow resistance; thus the gas flow rate into the middle chamber is determined by the piston position and the pressure wave. However, a MPTR is different. In Figure 2 (B), the cold-chamber movement is determined by the hot-end gas movement, the gas piston change, and the middle chamber change. In contrast, the middle chamber volume change is generally not determined by the hot-end gas movement. This makes the phase relationship required for refrigeration a problem. In fact, the flow channel connecting the pulse tube and the regenerator must be specially designed to limit or control the middle chamber amplitude. This makes the phase relation of the middle chamber in a MPTR quite different from that of a G-M refrigerator. Flow resistance becomes very important in a MPTR. Flow Resistance and Mass Flow

For a MPTR, a small opening of the multi-bypass valve will form a small middle chamber, and a large opening will result in a large one. The middle chamber volume is controlled by the

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339

multi-bypass tube flow area, and in this process, the pressure wave before and after this flow area becomes very important. As a comparison, at the cold end of pulse tube, the connection tube is generally large enough to minimize flow resistance. In the pulse tube, flow resistance is small, and pressure wave is of little relation to position. Assuming the pressure wave at the multi-bypass position of the regenerator is the

mass flow into and out of the pulse tube multi-bypass is determined by.

where c is the flow coefficient, A is the flow area, and

Consider at time gas begins to flow into the pulse tube, volume change will be: When me pressure in the pulse tube changes, tube:

then the partial

will change correspondingly in the pulse

where n is component to consider gas change, and for an adiabatic process The bypass position volume change will thus be:

changes from zero to a maximum and then to zero, and generally has a special form. Refrigeration Effect Eq. (4) only gives the potential middle chamber volume change. The function of this flow and the extent to which it contributes to refrigeration needs further consideration. Assuming the pulse tube pressure is when some gas with temperature flows into pulse tube through the multi-bypass, and when this gas with temperature Tmo flows out of the regenerator, the pressure in pulse tube is Similar to Eq. (3), the relationship is as follows:

If and there will be a temperature drop effect when gas stays in the pulse tube. Considering the whole process, gas flowing into the pulse tube should have an average higher inlet pressure. In general, such a condition is obtainable in experiments with a particular orifice opening, and then refrigeration will happen. But the degree of phase between pressures is mainly determined by flow resistance, and the refrigeration effect is generally small. This has been explained by others.5 REFRIGERATION SYSTEM

Design of the system is based on the viewpoint of multi-stage cooler.

As shown in Figure 1, the regenerator includes two parts: a high temperature part and a low temperature part. The high temperature part was 120 mm long, with 28 mm outside diameter, and 0.45 mm thick stainless steel tube; it was filled with 250-mesh stainless steel screen. The low temperature part was 110 mm long, with 20 mm outside diameter and 0.45 mm thick stainless steel tube; it was filled with 250-mesh stainless steel screens for the first 30 mm length, and 0.20.3 mm lead spheres for the rest. The pulse tube also includes two parts. The high-temperature

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part is 150 mm long, with 18.4 mm outside diameter, and 0.2 mm thick stainless steel tube. The low-temperature part is 100 mm long, with 14.4 mm outside diameter, and 0.2 mm thick stainless steel tube. The joint between the two parts is filled with 20 pieces of 60-mesh copper screen. The connection of the multi-bypass is made with a copper tube with an inner diameter of 1.86 mm. By changing the minimum flow area of the multi-bypass tube, different experiments were carried out. The pulse tube refrigerator was driven by a 750-W air-conditioning compressor. The frequency of the rotary valve was easily changeable from 1 Hz to 10 Hz. All tests were done under a pressure of about 10 bar and 700 W input power to the compressor. In the experiments, the main parameters measured were: a. Temperatures. There are three as shown in Figure 1: at cold tip, at the multi-bypass position, and near the pulse tube hot end. b. Pressure waves. One is before the regenerator, the other in the pulse tube. c. Cooling capacity at the cold tip. d. Multi-bypass minimum flow area. EXPERIMENT RESULT AND ANALYSIS Effect of Multi-bypass Area

The most important aspect of an MPTR is the multi-bypass effect. In experiments, for a fixed multi-bypass flow area, the orifice and double-inlet were optimized to reach the lowest temperature. Six groups of experiments were done and the main parameters measured are listed in Table 1. In the table, different forms of flow channels for multi-bypass have been converted to flow area. The most important effect of multi-bypass area is temperature. There is an optimized flow area to achieve the minimum cold-tip temperature this is the same as former experiments.2,3 Temperature at the multi-bypass position drops continuously with the opening of the multibypass. This is easy to understand. Larger opening of the multi-bypass represents larger flow rate and larger refrigeration at the bypass position, and thus the temperature will drop. This can also explain the temperature of As to the relation between and it is possibly due to the mass flow distribution. With pulse tubes, increase of the bypass flow will decrease the cold-tip flow. An optimized lowest temperature is thus gained with a decrease of gas flow at the cold tip in combination with a relatively high regenerator efficiency, and a relatively low middle temperature. The pressure wave is evidently affected by the multi-bypass. Through bypass flow, flow resistance is much smaller than that without multi-bypass. From no multi-bypass to largest multibypass, the pressure drop changes from 0.742 to 0.923, and this makes the pressure amplitude in the pulse tube larger.

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Figure 3. Cool-down curves for three different openings of the multi-bypass valve including: no maximum multi-bypass and optimum multi-bypass

multi-bypass

Orifice and double-inlet opening are important parameters for a MPTR. Both use standard needle valves. From Table 1, optimization of the opening of the orifice and double-inlet valve is seen to have little relation to the flow area of the multi-bypass. This is different from that for twostage pulse tube refrigerators, where there are two pulse tubes, and each one has a best value. In contrast, this feature is more like the function of a double-inlet, whose effect on orifice opening is small. And this is the rationale for a high- and low-temperature double-inlet version.5 Cooling Down Curve and Cooling Capacity

In past experiments with a 16-Hz MPTR, there is a slow cool-down rate in comparison to an OPTR or DPTR.2 However, this feature was not evident in this experiment, as shown in Figure 3. The three curves in Fig. 3 are for, respectively: 1) largest multi-bypass, 2) no multi-bypass, and 3) optimized multi-bypass. Although the lowest temperatures are different, the cooling rates are similar. Refrigeration capacity is very important for a cryocooler. Some experiments for MPTRs have shown a temperature rise per unit cooling capacity that is higher than for a DPTR.2 It seems that in this experiments, such a tendency is not very evident, as shown in Figure 4. The results for all five groups of multi-bypass designs (groups 1 to 5 as defined in Table 1) are almost parallel and slightly lower in slope than the DPTR results (group 6).

Figure 4. Cooling capacity of each of the six experiment groups.

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Figure 5. Effect of double-inlet opening on refrigeration temperature at different orifice opening.

Relation of Orifice and Double-inlet

With the optimized multi-bypass flow area of orifice and double-inlet are adjusted to reveal their effect on refrigeration temperature. The typical result is shown in Figure 5. There are four groups of orifice openings. For every orifice opening, there is a best corresponding double-inlet opening. It seems that larger opening of the orifice will need a larger opening of double-inlet in order to reach the lowest temperature. Also, the best orifice opening for the OPTR to gain the lowest temperature doesn' t correspond to the best orifice opening for the DPTR to gain its lowest temperature. However, the temperature difference is only 2-4 K. Frequency Effect In the experiment, frequency is easy to adjust. Tests were conducted to optimize the frequency. As shown in Table 2, the optimized frequency is about 2.2 Hz. And throughout the test, this frequency led to the best results in comparison to other frequencies. Also, as with previous experiments, frequency strongly affects the optimum opening of the orifice, while its effect on the opening of the double-inlet is quite small. DISCUSSION

Other results on the parameters influencing the MPTR are available. However, they are based on a pulse tube cooler with a relatively high refrigeration temperature. In fact, the high temperature part of the regenerator and the high temperature part of the pulse tube once formed one pulse tube refrigerator. That pulse tube refrigerator could easily reach 40 K. And with the second part of the regenerator added, the refrigeration temperature is higher. The whole regenerator in this paper came from one two-stage pulse tube refrigerator, and this cooler reached a temperature of 11 K at the second stage coldtip when the high-temperature stage was cooled to 50 K. Thus, a multi-bypass pulse tube refrigerator should not be treated as general multi-stage cooler.

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In this experiment, one thing that needs to be improved is to further lower the temperature of high temperature part. However, the design of this part is not easily isolated, as it is also related to the low-temperature part. Another future consideration is how to make the MPTR act like a twostage cooler. CONCLUSION

The analysis shows that a MPTR is different from a general multi-stage pulse tube refrigerator. This makes it difficult to design. In this investigation, several groups of experiments were conducted to reveal the features of the MPTR at low frequency. Throughout the work, it could be concluded that the design of the MPTR should not follow the idea of a general multi-stage unit. Also, the experiments have shown that the MPTR resembles a double-inlet pulse tube more than a multi-stage unit. REFERENCES 1. Zhou, Y., Han, Y.J., “Pulse tube refrigerator research,” 7th International Cryocooler Conference Proceedings, Air Force Phillips Laboratory Report PL-CP--93-1001, Kirtland Air Force Base, NM, April 1993, p. 147. 2. Wang, C., Wang, S.Q., Cai, J.H., Zhou, Y., “Experimental study of multi-bypass pulse tube refrigerator,” Cryogenics, vol. 36 (1996), pp. 605-609. 3. Cai, J.H., Wang, J.J., Zhu, W. X., and Zhou, Y., “Experimental analysis of the multi-bypass principle in pulse tube refrigerator,” Cryogenics, vol. 34 (1994), pp. 713-715.

4. Olson, J.R., Kotsubo, V., Champagnes, P.J., Nast, T.C., “Performance of a two-stage pulse tube cryocooler for space application,” Cryocoolers 10, Kluwer Academic/Plenum Publishers (1999), pp. 163-170.

5. Yang, L.W., Zhou, Y., Liang, J.T., “Research of pulse tube refrigerator with high and low temperature double-inlet,” Cryogenics, vol.39 (1999), pp. 417-423.

Experimental Study of the Heat Transfer in Pulse Tubes S. Jeong, K. Nam, M. G. Kim, H.-M. Chang* and E. S. Jeong* Korea Advanced Institute of Science and Technology Department of Mechanical Engineering Taejon, 305-701, Korea *Hong Ik University, Dept. of Mech. Engineering Seoul, 121-791, Korea

ABSTRACT The present study has been conducted to observe the details of heat transfer under pulsating pressure and oscillating flow in a pulse tube. An experimental apparatus was fabricated to measure the gas temperature, wall temperature, pressure, and the instantaneous heat flux inside a pulse tube. The measured gas temperature and heat flux must be corrected to compensate for their finite time constant under oscillating flow conditions. In experiments performed from 1 Hz to 3 Hz, the phase difference between the instantaneous heat flux and the gas-wall temperature difference was clearly observed. The experimental heat fluxes were compared to theoretical correlations such as the Complex Nusselt Number Model (CNNM) and the Variable Coefficient Model (VCM).

In general, the absolute value of the heat flux predicted by the CNNM was greater than that of the VCM. The experiment confirmed the validity of the VCM for the instantaneous heat flux under the

pulsating pressure and oscillating flow in the warm end of the basic pulse tube.

INTRODUCTION

Oscillating flow under pulsating pressure is a common phenomenon in an engineering system such as a pulse tube cryocooler, Stirling cryocooler, or G-M cryocooler. Due to the complex physics and lack of experimental data, the heat exchangers in these systems are usually designed by conventional steady-state heat transfer relations that can not predict the oscillating heat transfer phenomena properly. It is known that a phase shift exists between the instantaneous heat flux and the gas-wall temperature difference under oscillating flow and pulsating pressure conditions. The conventional Newton’s law of cooling does not contain a term that explains this phase shift phenomenon. Kurzweg1 attempted to apply the previous

oscillating heat transfer data to a Stirling cycle heat exchanger. Gedeon2 introduced a complex Nusselt number using the results of Kurzweg’s work. He obtained a Nusselt number for incompressible oscillating flow and showed the existence of the phase shift between the heat flux at the wall and the gas-wall temperature difference when the oscillatory frequency was high. Kornhauser3 showed that heat transfer analyses using the complex Nusselt number could predict the experimental data well for the Stirling engine. Jeong et al.4 obtained two-dimensional Cryocoolers 11, edited by R.G. Ross, Jr.

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velocity and temperature profiles for the oscillating flow caused by pulsating pressure that was induced by piston motion. He suggested a new heat transfer relation, so called the ‘Variable Coefficient Model’. Lee et al.5 studied analytically the heat transfer of the Stirling cycle heat exchanger. He obtained the gas and wall temperature profiles when the axial gradient of the wall temperature was not constant and the oscillating flow only existed. He also examined the effect of the frequency and the maximum displacement. Jeong et al.6 installed the heat flux sensor on

the outer surface of the heat exchanger for the basic pulse tube and calculated the instantaneous heat flux at the inner wall using the measured data. In their experiment, however, the calculated

heat flux at the interior wall had an uncertainty due to the capacitance effect and the complex geometry of the heat exchanger. This paper describes the experiment of the instantaneous measurement of the heat flux and the temperature at the heat exchangers of the pulse tube under oscillating flow and pulsating pressure. The measured heat flux data were compared with the theoretical predictions that had

been previously developed. EXPERIMENTAL CONFIGURATIONS

An experimental apparatus was fabricated as shown in Fig. 1. At the inlet of the pulse tube, the stainless steel mesh (#200) was stacked to make one-dimensional flow between 1/4 inch tube

and 1 inch tube. Its thickness was 4 mm. The cold-end heat exchanger was made of 1 mm thick copper tube and nicely fitted to the pulse tube. To install the heat flux gauge on the inside wall of the heat exchanger, the flange with 7 mm thickness was brazed to the heat exchanger and assembled with bolts. The warm-end heat exchanger had the same shape as that of the cold-end

heat exchanger, but it had the water jacket to cool the warm-end by the cooling water as shown in Fig. 2.

Figure 1. Schematic diagram of the experimental apparatus. (CHX: cold-end heat exchanger,

WHX: warm-end heat exchanger).

EXPERIMENTAL STUDY OF HEAT TRANSFER IN PULSE TUBES

Figure 2. Warm-end heat exchanger.

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Figure 3. Sensor Installation (T/C: thermocouple).

Fig. 3 illustrates the sensor installation at the heat exchanger. A fast response thermocouple (OMEGA Model EMQSS-010E, type-E) was inserted into the heat exchanger by reducer fitting. The wire diameter of this thermocouple was 0.038 mm and its junction was exposed to the flow field for the fast response characteristics. The heat flux gauge was attached to the inner wall of the heat exchanger. It was a thick micro-foil type sensor and its time constant for the step input was 20 ms according to the manufacturer. Thermal resistance of the heat flux gauge was very small compared to that of the convection with the helium gas. The difference of the heat flux between the case with the sensor and without the sensor was about 5 % of the heat flux

value. The heat flux gauge had also the T-type thermocouple that could measure the wall

temperature of the heat exchanger. The pulse tube was made of 1 mm thick stainless steel tube with outer diameter of 25.4 mm and the length of 200 mm. Helium compressor (CTI-cryogenics Model 8200) was used to supply helium gas as working fluid. The rotary valve system provided the pulse tube with pulsating gas pressure and flow. The frequency of the pulsating pressure was adjusted by the rotational speed of the rotary valve, which was controlled by the stepper motor and the function generator. The strain gauge type pressure transducer (Sensym Model ST2000) was installed at the inlet

of the pulse tube to monitor the pulsating pressure. The signal of the heat flux gauge was so small (the order of and so sensitive to the environmental noise that this signal was preamplified by isolation pre-amplifier (YOKOGAWA Model 313100-61E). All the experimental data were acquired by data acquisition board (Keithley Model DAS 1600) and stored in the personal computer when the cyclic steady-state was reached. EXPERIMENTAL RESULTS AND DISCUSSION

As mentioned earlier, utilizing complex Nusselt number concept, Kornhauser3 proposed a heat transfer correlation as follows:

where

k f

= thermal conductivity of the gas = hydraulic diameter of the pulse tube = oscillating frequency of pressure wave = real part of the complex Nusselt number

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= imaginary part of the complex Nusselt number The complex Nusselt number in Eq. (1) was determined by using the least square method from the experimental data.3 He also presented the complex Nusselt number for various Peclet number On the other hand, Jeong et al.4 has derived the following relation from the two-dimensional energy equation for the case of the fluid flow between two flat plates. The heat flux at the wall was expressed as follows:

where

2H = distance between two flat plates R = ideal gas constant P = pressure of the gas The thermal conductivity of the gas and the wall temperature were assumed to be constant when Eq. (2) was derived. When Eq. (2) was applied to the pulse tube, 2H was assumed to be the

diameter of the pulse tube.7 The fundamental difference of Eq. (2) from Eq. (1) is that the heat flux term associated with the oscillation is derived from the direct instantaneous pressure change. Although the temperature gradient term in Eq. (1) is replaced by the pressure gradient term in Eq. (2), two equations describe virtually the same physical phenomenon because the temperature is varied by the pressure change. Kornhauser’s relation contained the complex Nusselt number that must be determined from the extensive experiment, but Jeong’s relation included no empirical constant. Kornhauser’s relation was also known as the Complex Nusselt Number Model (CNNM) and Jeong’s relation as the Variable Coefficient Model (VCM). Under the pulsating pressure and oscillating flow, the heat flux and the gas-wall temperature difference have been measured as shown in Fig. 4. In this paper, the experimental data are shown only for the case of the closed orifice (basic pulse tube type) and no load condition at the coldend heat exchanger. There was an apparent phase shift between the heat flux and the temperature difference at 2 Hz, but the phase shift was not clear at 1 Hz. The heat flux was also calculated by Eq. (1) and Eq. (2) using the measured temperature and the pressure. The predicted flux was compared to the measured heat flux at the cold-end and the warm-end heat exchanger as shown in Fig. 5 and Fig. 6. The negative heat flux means the heat flow from the gas to the wall. At the warm-end heat exchanger, both the CNNM and the VCM can predict the heat flux reasonably well. On the other hand, Fig. 6 shows that both the results from the two theoretical relations failed to predict the experimental data at the cold-end heat exchanger. The CNNM predicted the heat flux variation pretty well, but with large amount of discrepancy. This result indicated that the complex Nusselt number suggested by Kornhauser could not be applicable to the cold-end heat exchanger. Since Kornhauser correlated the complex Nusselt number only by Peclet number his empirical correlation could not represent the general case. Therefore, it is suggested that another non-dimensional parameter, such as maximum Reynolds number

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Figure 4. Heat flux and gas-wall temperature difference at the cold-end heat exchanger: (a) 1 Hz (b) 2 Hz

Figure 5. Heat flux at the warm-end heat exchanger: (a) 1 Hz (b) 2 Hz

Figure 6. Heat flux at the cold-end heat exchanger: (a) 1 Hz (b) 2 Hz.

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should be included in the complex Nusselt number. Here, is the maximum mass flow rate of the oscillating flow, is the viscosity of the gas and A is the cross-sectional area. The VCM did not match the variation pattern of the heat flux in one cycle, but the cyclic integrated value was more close to the measured value than that of the CNNM. Table 1 shows the comparison of these values. The CNNM had another difficulty in its application at higher frequency. Since it contains a

time derivative term of the measured temperature as shown in Eq. (1), the prediction of the heat flux by the CNNM may require one more calculation step. As shown in Fig. 7(a), a small variation in gas temperature was amplified by the time derivative operation and the heat flux from this gas temperature had a large oscillation as shown in Fig. 7(b). This inevitable error amplification could be virtually reduced by approximating the measured temperature to a smooth function like sinusoidal wave (continuous line in Fig. 7). Fig. 8 shows, thus, the predicted heat flux of the CNNM by this error reducing method. CONCLUDING REMARKS

The experiment was performed to the heat exchangers of the basic pulse tube for the frequency of 1 ~ 3 Hz. From this experiment, we recognized the following results. (1) The VCM (Variable Coefficient Model) could predict the heat flux in the warm-end heat exchanger of the pulse tube better than the CNNM (Complex Nusselt Number Model).

Figure 7. Heat flux from the CNNM using the curve fitted gas temperature (3 Hz, warm-end).

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Figure 8. Comparison of the heat flux using the calculated complex Nusselt number. (3 Hz, warm-end).

(2) Neither the VCM nor the CNNM predicted the heat flux at the cold-end heat exchanger where the oscillating flow effect was noticeable. The better model is necessary for the

oscillating flow with large amplitude. ACKNOWLEDGMENT

This research was supported by the Korea Research Foundation under the contract no. 1998018-E00164 and partially supported by the Brain Korea 21 project. The authors would also like to acknowledge the support of the Ministry of Science and Technology through the Dual Use

Technology project (grant no. 99-DU-04-A-02). REFERENCES 1. Kurzweg, U. H., "Enhanced heat conduction in fluids subjected to sinusoidal oscillations", J. Heat Transfer, Vol. 107 (1985), pp. 459-462.

2. Gedeon, D., "Mean-parameter modeling of oscillating flow", J. Heat Transfer, Vol. 108 (1986), pp. 513-518. 3. Kornhauser, A. A., Gas-wall heat transfer during compression and expansion, Sc.D. Thesis, Dept. of Mech. Eng., Massachusetts Institute of Technology, Cambridge, MA (1989). 4. Jeong, E. S. and Smith, Jr., J. L., "An analytic model of heat transfer with oscillating pressure", General papers in Heat Transfer, ASME, Vol. 204 (1992), pp. 97~104.

5. Lee, D. Y., Park, S. J. and Ro, S. T., "Heat transfer by oscillating flow in a circular pipe with a sinusoidal wall temperature distribution", Int. J. Heat Mass transfer, Vol. 38, No. 14 (1995), pp. 2529-2537.

6. Jeong, S. and Nam, K., "An experimental study on the heat transfer characteristics of the heat exchangers in basic pulse tube refrigerator", Crycoolers 10, Plenum Publishers, New York (1999), pp.

249-256. 7. Jeong, E. S., "Secondary flow in basic pulse rube refrigerators", Cryogenics, Vol. 36, No. 5 (1996), pp. 317-323.

Shuttle Loss in Pulse Tubes L.W. Yang

Institute of Applied Physics Justus-Liebig-University Giessen D-35392 Giessen, Germany

ABSTRACT

Pulse tube refrigerators have the unique feature of a hollow tube replacing the cold piston. This paper analyzes some losses resulting from this hollow tube due to the oscillating flow; the losses are quite similar to the shuttle loss and pumping loss in a Stirling or G-M refrigerator. The analysis starts with surface heat pumping in a basic-type pulse tube cooler. Then, flow features and the origin of losses are further explained. As an important parameter, boundary layer thickness is predicted. The analysis shows that, due to the large temperature difference, the distribution of boundary layer thickness along the pulse tube changes greatly. For different boundary layer thicknesses, shuttle losses and pumping losses differ significantly. With the predicted boundary layer thickness, the amplitude of the shuttle loss and pumping loss are just the opposite of what they are in a traditional Stirling or G-M cooler. Pumping gas loss becomes an important loss in the pulse tube refrigerator. INTRODUCTION

In a cryocooler, evaluation of the losses is as important as the refrigeration mechanism.1 Generally, the various kinds of losses occupy a major part of the gross refrigeration. For example, in a Stirling refrigerator, there are regenerator heat transfer losses and flow resistance losses, displacer pumping gas losses and shuttle losses, conductance losses, cold tip radiation losses, and heat-exchanger losses. These losses make up about 80-90% of the gross refrigeration at low temperatures. Pulse tube refrigerators have developed very fast since 1984, with the orifice and reservoir being added at the hot end of the pulse tube.2 These improvements gave rise to a revolutionary improvement in the refrigeration mechanism. Many theoretical methods have been used to explain pulse tube operation based on different physical models. However, as to loss evaluation, especially as to the loss in the pulse tube, no detailed introduction could be found in the literature. The pulse tube working process has been treated in many references now. Some researches consider the flow or working process in the pulse tube itself, including flow pattern research and numerical simulation.3, 4 However, these results are not directly related to the evaluation of losses and the design of pulse tubes. Recently, the unique gas movement and potential heat transfer in the pulse tube was explained with a change of temperature ratio.5 Other researchers have also pointed out the problem of shuttle losses in a pulse tube.6 Here shuttle loss and pumping loss are analyzed and calculated in detail for the first time. Cryocoolers 11, edited by R.G. Ross, Jr.

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Figure 1. Structure of pulse tube refrigerator and gas movement.

Figure 2. Gas movement characteristics of BPTR. ORIGIN OF LOSSES

Limit of Surface Heat Pumping

The pulse tube refrigerator was invented in 1963 by Gifford and Longthworth.7 This first pulse tube is referred to as a basic-type pulse tube refrigerator (BPTR), and its working mechanism is

attributed to surface heat pumping as shown in Fig. 1 (BPTR) and Fig. 2. In Figure 1, accompanying the movement of the compressor piston, there is another gas movement at the jointing position between the pulse tube and the regenerator. At this jointing position the temperature is lowest and a temperature gradient is formed along the pulse tube wall. For such a process, the working mechanism and its limitations are shown in Figure 2. Assuming an adiabatic process in the pulse tube, with a pressure increase, every part of the gas in the pulse tube will move further into the tube, with the change in position and temperature described by Eqs. (1) and (2):

where is the adiabatic component and y is the normalized pulse tube position from the hot end. Figure 2 illustrates the gas moving process and corresponding temperature at four initial pulse tube positions, y = 0.2, 0.5, 0.8 and 1.0, where 1.0 represents cold end. First look at the curve with the hot end temperature and cold end temperature Evidently, at any position the temperature of moving gas is higher than the temperature of that based on a linear distribution. Then, during the high pressure process, moved gas may transfer heat to the corresponding pulse tube wall. Thus, the whole process results in heat being pumped to a higher-temperature position from an initial lower-temperature position. Such a process is called surface heat pumping. This shows why there is temperature drop in the pulse tube refrigerator.

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However such a process has its limit or shortcoming with low temperature decreasing. Let’s look at in Figure 2. At the position of 1.0, gas temperature change almost coincides with assumed linear temperature distribution. Evidently, surface heat pumping will have problem. With even lower temperature, the inverse process happens compared to y = 0.8 and 1.0 of situation. Though the gas temperature increase upon compression, this temperature is much lower than the temperature based on a linear distribution. The mechanism of surface heat pumping at certain temperature ratio will terminate, and the process turns to be a loss. Through controlling gas temperature distribution along pulse tube, the lowest temperature reached in theory may be low. For example, assuming that the temperature distribution has the following form: n= 1 means linear distribution of temperature along the wall, n>1 means there is a slow temperature change at the hot end, n <1 means a slow temperature change at the cold end. Based on, such a hypothesis, a group of ideal lowest temperature can be calculated as shown in Table 1. Table 1 shows that with smaller value of n, a low temperature of about 70K is attainable. But if temperature distribution at n= 0.2 is plotted, we will find that this is a non-realistic curve, because the low temperature part is long and flat. Gas Movement with Reservoir and Double-inlet As shown in Figure 1 (OPTR DPTR), after the invention of orifice version pulse tube refrigerator (OPTR)2 and further double-inlet version pulse tube refrigerator (DPTR),8 the above mentioned low temperature limit has been overcome. Different to the BPTR, the gas in the pulse tube consists of three parts for the OPTR. Besides the gas section (here called gas piston) staying in the pulse tube and the gas section flowing into pulse tube through the cold end, now a gas section flowing through the hot end of

pulse tube is added, as shown in Figure 1. Neglecting mechanism, complex moving process of different gas section is considered. For basic version, cold end flow position change has the following form according to adiabatic process:

where is the pulse tube length, pressure wave in the pulse tube, lowest pressure in the pulse tube, are the gas piston length at average pressure For orifice version, assuming that hot end gas section movement is and the reservoir has a large volume, generally there is very small pressure change in the reservoir, and position change at hot end has the following relation:

where represents pressure in reservoir. With double-inlet, another gas section pressure difference of the following relation:

will be added at the hot end according to

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where is pressure wave at inlet of regenerator. And then hot end flow with orifice and double inlet change to be: where are coefficients to determine their amplitude. With hot end flow predetermined, cold end gas flow is easily to attain. Assuming that the gas piston change is adiabatic, a simple formula to reveal their relation is that:

All three gas groups move in the pulse tube continuously. Because of the hot end flow, the refrigeration process is different to Figure 2, and at the same time, loss may occur. Loss due to Gas Piston Flow

A typical result of the actual flow feature of OPTR is as shown in Figure 3. It’s easy to notice the refrigeration mechanism. Different to Figure 2, cold tip gas flow position forms one circle in clockwise direction. The corresponding inlet pressure at is larger than the lowest pressure The circled area represents the theoretical refrigeration capacity or work done to gas piston near it. Comparing to Figure 2, the circle in the

cold end is due to the circle in the hot end. Let’s look at the gas piston movement. In the Figure 3, six groups of length and position are depicted by double-arrow line. The length of gas piston changes only with pressure. But we should notice that the biggest length is not at cold tip, nor the shortest length at the hot tip. Its change with position is complex, while the process is from cold tip to hot tip and then back. If we depict the flow curve of any point of gas piston, a circle will form and contact with tube wall and heat transfer is inevitable. What we are really interested in is that whether this heat transfer contributes to refrigeration or a kind of loss. First look at cold tip. Different to basic version pulse tube, at the highest pressure, which corresponding to the highest temperature of the gas piston, the gas piston is not corresponding to the largest position of but at the middle of the pulse tube, and at the largest position of the pressure drops by a certain extent. Compared to Figure 2, gas moving distance is much larger with hot end flow adding. Also important is temperature ratio. If the lowest temperature is

lower than that determined by Table. 1, surface heat pumping will not work for cold tip, and loss will be inevitable. For basic version pulse tube, hot end of pulse tube always contributes to surface heat pumping. From Fig. 3, with pressure increasing, increases first while basic version decreases: the highest pressure is not at the shortest position, the lowest pressure is not at the lowest position. Heat transfer is inevitable, but not transferring heat to higher temperature position. The loss is inevitable. In fact, the loss resulting from the movement of gas piston is just like the actual cold piston or displacer in Stirling refrigerator or G-M refrigerator. The most evident one is shuttle loss. When considering this loss, we could assume that gas piston length itself could keep unchanged but the position is constantly changing. Then heat is transferred from low temperature to high temperature gradually according to wall and gas piston

contact. This loss in Stirling refrigerator is large and important. Another is pumping gas loss. This loss is to consider the effect of gas gap or gas clearance

between gas piston and wall. Due to viscosity, this gas gap adheres to tube wall and can not move like gas piston. Then some more gas will go into this gas gap from cold end or gas piston during compression and be released during expansion, and the corresponding heat transfer forms the loss. This loss is generally small in Stirling or G-M refrigerator.

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Figure 3. Typical gas piston movement in pulse tube.

METHOD TO CONSIDER SHUTTLE LOSS Shuttle loss is considered similarly to traditional G-M type or Stirling refrigerator.1 Total heat transfer amount could be expressed as: where Q represents heat transfer amount in one cycle,

is average moving distance of the gas

piston and f is frequency. According to the mechanism of shuttle loss, one cycle heat transfer amount:

where is gas conduction coefficient, A is heat transfer area, is temperature difference between wall and gas piston, is heat transfer time, is distance between tube wall and gas piston. In theory, A, are different to traditional ones, and they are changing with pressure wave and flow of orifice and double-inlet. These parameters should be considered in

priority. Length of Moving Distance

in Eq.(10) represents moving distance. In ST or GM cooler, it is just related to cold chamber length minus potential dead volume length. But for pulse tube, its value is not so easy to determine because hot end part and cold end part generally have different moving length and the two are combined together. Here based on Eq. (4-9), the result is deduced. For basic version pulse tube, the largest moving distance is determined by: For orifice version, the largest value at the hot end should be determined first, which depends on orifice valve. Then we could roughly predict the largest value at the cold end: For double-inlet version, we should have the largest movement by double-inlet valve, and then hot end largest movement is:

at the hot end resulted

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while the cold end movement becomes: With the upper largest values, average moving distance will be:

Length of Gas Piston Though in Eq. (9) gas piston length at average pressure is given, its actual value is complex to determine. The problem is that have phase relation. From Eq. (9), at average pressure, we could get: For basic version, the value is easy to get according to Eq. (4):

For orifice version, we could assume pressure ratio is not large and reservoir is large, then: For double-inlet version, if

generally, we have:

Besides average length of gas piston, we should notice that the actual gas piston length changes with pressure: For high pressure ratio such as 1.8, for low pressure ratio such as 1.2, The gas piston of low pressure ratio more resembles actual piston. Heat Transfer Temperature Difference

Along gas piston, a simple and reasonable assumption is linear distribution of temperature. And in actual, wall temperature of pulse tube is in linear distribution for orifice version, but this

may be destroyed by double-inlet. Here the nonlinear effect is neglected. When the gas piston is located at certain position difference is:

the heat transfer temperature

means certain middle position, where gas piston temperature distribution is the same to wall temperature distribution. It should be noticed that is joint position, which separates heat conducting to tube wall and from tube wall, or where heat transfer temperature difference changes its direction. Gas Gap Between Wall and Gas Piston

The most important and difficult consideration is about gas gap . If neglecting viscous of gas, there should be no so called gap. But this will result very large loss according to Eq. (11). The simplest consideration is to use steady state boundary layer method. The following

formula is the simplest method to calculate the boundary layer thickness (BLT) of laminar flow through a large plate:

where

is gas viscosity, x distance form plate inlet,

gas density, U main flow velocity.

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359

Accurately, flow characteristics in pulse tube does not fit upper equation. There are two questions. One is distance x. In theory, every part of the gas forming gas piston performs cyclic flow around certain point, length itself will also expand and shrink with pressure changing, and actual velocity U also has such a feature. Thus in theory, every time, every position has a different BLT based on definition of distance x, because any position will be start point of some gas and at the same time middle position of some gas. Also, viscosity will effect BLT because it changes with temperature. For the convenience of calculation, here based on some main parameters in pulse tube, a simple boundary layer thickness is determined. Considering actual movement, assume the length parameter in Eq. (22) to be:

Considering frequency f , the average velocity will be: Neglecting the complex of actual distance and velocity, upper two characteristic parameters could determine the boundary layer thickness. Then Eq. (22) changes to be:

Eq. (25) means that the boundary layer is determined by gas itself, temperature, pressure and frequency. Viscosity and density should also be used. Viscosity is changing evidently with temperature, while changing little with pressure at 0.1 MPa—5MPa. A best fit line is:

with temperature from 14K to 400K and unit (Accuracy will be 98%). Density could be calculated use ideal gas state equation with accuracy of about 99%:

for helium4, R =2118.4 J/(kg K). Calculation has been done to reveal the effects of temperature and pressure on It shows that with the increasing of temperature and decreasing of pressure, will increase evidently. The difference will be more than 100 times with same pressure at 15K and at 300K. This means, according to equation (25), the thickness of boundary layer at high temperature will be 10 times thicker than that at low temperature. Using helium gas, a group of calculated results is shown on Table 2. In Table 2, temperature has a very large effect on BLT. Then along the pulse tube, velocity distribution form will change greatly from high temperature to low temperature. Also, pressure and frequency also have evident effect on boundary layer thickness. Higher pressure and higher frequency corresponds to smaller BLT.

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Heat Conduction Coefficient

Gas heat conduction coefficient changes greatly with temperature and slowly with pressure. Neglecting pressure effect, a simple formula is given for helium gas:

Heat Transfer Area

Considering the boundary layer thickness, the average heat transfer area will be: where

is pulse tube diameter.

Calculation of Shuttle Loss

With upper parameters, Eq. (11) could be integrated to get the result. Because along the gas piston, parameters are changing, twice integration is needed. First, according to energy conservation, integration of whole area and whole time will be:

From upper integration, the middle point could be gained.

Then for half cycle amount from

that separates Q from positive to negative

we have:

A reasonable and simple method will be:

Upper integration process is possible. For many parameters of actual condition are coupled together, a computer is needed to get the result easily. In calculation, for different working conditions, shuttle loss should be calculated differently. For example, heat transfer time is not easy to determine. There are two simple methods. One is to consider linear distribution of time with temperature distribution. This is similar to Stirling type pulse tube refrigerator. At very part, gas move is similar to sinusoidal wave. The other is to divide a period into three parts: one is for gas piston staying at hot end, one is for gas piston staying at cold end, and the third is for gas piston staying at middle part where no heat transfer takes place. This method is suitable for G-M type pulse tube refrigerator. METHOD TO EVALUATE PUMPING GAS LOSS

Pumping gas loss is different from shuttle loss. Generally, when shuttle loss is large,

pumping gas loss will be small. And it also relates to boundary layer thickness. Using upper parameters, pumping loss similar to traditional method is deduced:1 The average boundary thickness is defined to be:

Mass flow rate in and out of this average thickness due to pressure wave is:

Adopting following heat transfer equation:1

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361

where is heat transfer coefficient, is equivalent diameter, is specific heat of gas which is nearly 21 KJ / Kmol · K for helium gas at 50-400K. Then non-efficiency of heat transfer in the boundary layer is determined according to heat transfer unit (NTU) method:

where

is heat transfer area.

Pumping gas loss is:

LOSS ANALYSIS---COMPARISON WITH THEORETICAL REFRIGERATION

Based on thermodynamic analysis of the hot end flow of pulse tube and considering sinusoidal pressure wave, following formula is reduced to calculate theoretical refrigeration:9

where

is pressure wave amplitude and

a special parameter determined by:

Combining refrigeration capacity with upper two losses, one typical calculated result according to a typical group of parameters is given, as shown in Figure 4, where relative proportion of shuttle loss, pumping loss and sum of two to refrigeration are given. For different

average boundary layer thickness, there will be different results of shuttle loss and pumping loss. With the increasing of shuttle loss decreases evidently while pumping loss increases quickly. There is a smallest value of the sum of two, which shows the best boundary layer thickness for pulse tube. Such a thickness is easy to execute for actual piston and not easy for gas piston. In Figure 4, the predicted point of actual boundary layer thickness is 0.57mm, which corresponds to the smallest shuttle loss and largest pumping loss. This is not strange. Shuttle loss becomes small with thick film preventing heat transfer, but pumping loss becomes large. In Stirling or G-M refrigerator, pumping gas loss is much small comparing to shuttle loss, because BLT or gas gap could be as thin as 0.1mm.

Figure 4. A group of loss evaluation result. Condition: pressure 2.5MPa, pressure ratio 1.1, frequency 50Hz, pulse tube diameter 9mm, length 70mm.

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CONCLUSION

In this paper, losses due to oscillating flow in a pulse tube have been analyzed in detail. Through analysis, some unique features have been found in the pulse tube flow. One is that surface heat pumping has a limit. The other is that boundary layer thickness in the pulse tube may be very thick compared to the actual piston. The above features will result in a small shuttle loss and a large pumping loss. It is important that, independent of a thick or thin boundary layer thickness, the sum of the shuttle loss and pumping loss will be an important loss in a pulse tube refrigerator. This may be helpful for future pulse tube refrigerator design. ACKNOWLEDGEMENTS

The author thanks Dr. G. Thummes (Giessen) for his helpful discussion and support. REFERENCE 1. 2.

3.

Bian, S.X., Small-type Cryocooler, China Mechanical Industry Press, Beijing (1985).

Mikulin, E.I., Tarasov, A.A., Shkrebyonock, M.P., “Low-temperature expansion pulse tubes,” Adv. in Cryo. Eng., vol.29, Plenum Press, New York (1984), p. 629. Shiraishi, M., Nakamura, N., Seo, K., Murakami, M., “Visualization study of velocity profiles and displacements of working gas inside a pulse tube refrigerator,” Cryocoolers 9, Plenum Press, New

4.

York (1997), pp. 355-364. Lee, J.M., Kittel, P., Timmerhause, K.D., Radebaugh, R., “Higher Order Pulse Tube Modeling,”

Cryocoolers 9, Plenum Press, New York (1997), pp. 345-353. 5. 6. 7. 8. 9.

Yang, L.W., Zhou Y., Liang, J.T., “Analysis of theory refrigeration and losses in pulse tube,” Adv. in Cryo. Eng., vol. 45, Plenum Press, New York (2000). Hiratsuka, Y., Kang, Y.M., Matsubara, Y., “Development of a 1 to 5W at 80K Stirling pulse tube cryocooler”, Cryocoolers 10, Kluwer Academic/Plenum Publishers (1999), pp.149-155. Gifford, W. E., Longsworth, R.C., “Pulse tube refrigeration”, Trans. ASME, J. Eng. Ind., vol. 63 (1964), p.264. Zhu, S.W., Wu, P.Y., Chen, Z.Q., “Double-inlet pulse tube refrigerators: an important improvement,” Cryogenics, vol. 30 (1990), p.514. Yang, L.W., Research of pulse tube refrigeration mechanism and practical development, Post-doctor Report, Institute of Mechanics, Beijing (1998), China.

Numerical Study of Gas Dynamics Inside of a Pulse Tube Refrigerator Yoshikazu Hozumi, Masahide Murakami1, Masao Shiraishi2 Chiyoda Corporation, 2-12-1, Tsurumichuo, Tsurumi-ku

Yokohama, 230-8601, Japan 1 Institute of Engineering Mechanics, University of Tsukuba, Tennodai 1-1-1, Tsukuba 305-8573, Japan 2 Mechanical Engineering Laboratory, MITI Namiki 1-2, Tsukuba, Ibaraki 305-8564, Japan

ABSTRACT A simulation program for viscous compressible flow has been developed to study the details of fluid motion and gas dynamics inside a pulse tube refrigerator. Axisymmetric two-dimensional Navier-Stokes equations are solved numerically. Simulation results inside of a basic-type pulse tube refrigerator have been reported in past research, and showed that secondary mass flow and enthalpy flux, going from the cold end to the hot end along the tube wall, contribute to the heat transfer. They also suggested that the boundary layer on the tube wall might play an important role in transferring enthalpy from the cold end to the hot end of the pulse tube. In order to investigate the heat transfer mechanisms within the entire pulse tube refrigerator,

the present research also includes the simulation of the flow and gas dynamics inside of the regenerator and after-cooler. The simulation results are compared with experimental data. The simulation results of the pulse tube temperature profile, when compared to the experiment, are seen to be in good agreement. The simulation results appear to well describe the gas dynamics and refrigeration mechanisms of the pulse tube refrigerator.

INTRODUCTION

Pulse tube refrigerators are categorized as thermoacoustic refrigerators. One approach to explaining the refrigeration phenomenon in a pulse tube refrigerator is to note that a pulse tube refrigerator works the same as a Stirling cycle refrigerator; the Stirling displacer is just replaced by the pulse tube in the pulse tube refrigerator. Thus, the pulse tube works as a gas piston similar to the Stirling displacer. A key feature of thermoacoustic refrigerators is their regenerator, which strongly determines the refrigerator's performance. Heat conduction in the regenerator and phase shift between the pressure change and the volumetric change are important phenomena in all Stirling-type refrigerators. Phase shift devices are installed at the hot end of the pulse tube to improve refrigeration performance. Examples include an orifice, an inertance tube, and a double inlet bypass connection. Cryocoolers 11, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2001

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A relevant question is: what happens in a heat-transfer-only pulse tube without a regenerator. Inside of the tube, when driven with an adequate pressure pulse, there is a differential temperature established between the closed end and the open end; the closed end gets hotter and the open end gets colder. In order to explain this heat transfer, a surface heat pumping mechanism was suggested by early investigators Gifford and Longsworth.1 Although the gas elements do not move directly from the cold end to the hot end in the pulse tube, the gas transfers heat from the open end to the closed end by successive heat-pumping cycles. Although this concept of heat pumping was suggested in the 1960s, it has not been verified by experimental research. This is because the mechanism is very local with a short cycle time, and experimental probes that measure temperature, velocity, and pressure can not respond to the working gas status. However, the phenomenon had been investigated by previous analytical researchers.2,3 The heat transfer mechanism has been described by the Lagrange frame approach (surface heat pumping mechanism), and also by using the pressure versus volume diagram.

In the pulse tube, a convective driving mechanism occurs in the oscillatory boundary layer at the solid wall of the pulse tube. During the cyclic expansion and compression process, the gas

elements close to the wall experience different viscous drag associated with the alternating gas temperature, and experience a net drift from the cold end to the hot end. This means that gas elements around the boundary layer move from the cold end to the hot end; this is referred to as the secondary mass flux. Since the net mass flux along the tube must be zero, so-called streaming is introduced in the center part of the pulse tube as the secondary-flux gas returns to the cold end. In order to investigate the heat transfer, one must evaluate the particle path of the working gas element and the temperature difference between the gas elements and the tube side-wall. The secondary mass flux and the streaming are observed by tracing the particle path of gas elements as shown in Figure 1. The phenomenon can also be explained by the approach using the pressure versus volume diagram. Figure 2 shows pressure versus specific volume (1/ Density) diagrams for three small gas parcels placed at different radial positions in the pulse tube. The PV diagram for the parcel closest to the tube side wall shows large work. The diagram does not show only the refrigeration work, but also the viscous dissipation of the gas parcel. The swept-out area of the diagram suggests the magnitude of the conversion of pressure to work. In contrast, the gas at the center of the pulse tube does not contribute refrigeration work. Thus, the gas on the boundary layer is what contributes the

heat transfer from the cold end to the hot end of the pulse tube. The purpose of the present research is to evaluate the whole pulse tube refrigerator performance. The gas dynamics in the regenerator must be investigated because the heat conduction in the regenerator is very important to the refrigeration performance. A simulation program incorporating a regenerator has been developed, and simulation results are presented in this paper.

Figure 1. Particle path showing secondary mass flux and streaming.

NUMERICAL STUDY OF GAS DYNAMICS INSIDE PULSE TUBE

365

Figure 2. Pressure versus volume diagram on the moving parcel inside pulse tube.

SIMULATION MODEL The state of compressible viscous flow is expressed by the Navier-Stokes equations, including the equations of mass conservation, momentum conservation, and energy conservation. Since the solution of discretized full three-dimensional Navier-Stokes equations places a heavy computational burden on computers, several simplified models of the governing equations have been investigated.4 Fortunately, the phenomena inside of a pulse tube can be treated as an axisymmetric flow. This flow feature is used and axisymmetric two-dimensional Navier-Stokes equations are directly solved for the pulse tube flow. The simulated refrigerator model matches the experimental apparatus at MITI's Mechanical Engineering Laboratory.5,6,7 The pulse tube dimensions are 250 mm x 15.6 mm (length x radius), and the regenerator length is 250 mm. Figure 3 shows the simulation model of the pulse tube. Although metal screens are installed in the actual regenerator, a pipe-type regenerator is used for the simulation modeling. The pipe-type regenerator has the advantage that

Figure 3. Schematic diagram of the pulse tube refrigerator used for simulating the axisynmetric twodimensional flow region: (1) Pulse tube, (2) Regenerator.

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PULSE TUBE ANALYSES AND EXPERIMENTAL MEASUREMENTS

the flow inside of the pipes can be treated using the same simulation model that is applied for the pulse tube simulation. The modeling concept is also shown in Figure 3. In reality, parallel plates and pipes are not well suited for the real regenerator because of their longitudinal configuration. In order to obtain the equivalent surface for heat conduction, the regenerator length in the simulation had to be made approximately three times the length of the actual experimental apparatus. Practical evaluation of the refrigeration performance of the entire pulse tube is available with the incorporated regenerator.

SIMULATION METHOD The phenomena inside of the pulse tube can be treated as axisymmetric flow. Thus, axisymmetric two dimensional viscous compressible Navier-Stokes equations were used as the governing equations. The finite volume method was applied in the simulation. An implicit flux split difference approximation was applied to discretize the Navier-Stokes equations. The flux split

difference approximation has an advantage in describing the compressible flow. Gauss-Seidel line relaxation was used to solve the discretized Navier-Stokes equations. This numerical approach is a popular method in the field of treating highly compressible flow.8 In order to close the equations, the assumption of a perfect gas was made; helium is used as the working gas in this study. BOUNDARY CONDITIONS A numerical simulation requires boundary conditions for solving the governing equations.

Because of its large influence on the simulation results, a practical velocity profile must be specified for the tube inlet boundary condition. The analytical solution of the Navier-Stokes equations for pulsating flow in the inlet of the regenerator is obtained in terms of Bessel functions.9

At the pulse tube wall, a heat conduction wall model is applied, and the wall temperature is determined by the heat conduction from the gas temperature. The gas status inside of the boundary layer near the tube wall is important for heat transfer, and heat conduction is also important for the heat pumping mechanism. Mesh points are concentrated toward the wall surface and approximately eight points are imposed within the viscous boundary layer for better resolution of the

viscous shear stress on the wall. Boundary conditions at the pulse tube hot end assume a heat exchanger is installed here, and the amount of heat transferred is computed based on the temperature difference between the gas and the wall. Orifice type pulse tube refrigerators can also be analyzed by the program by changing the pulse tube hot-end boundary condition to allow flow in and out of the pulse tube though the orifice hole, driven by the pressure difference between the tube and the reservoir. SIMULATION RESULTS AND DISCUSSION The first step in the numerical simulation was verification that the simulation method and model appropriately describes the gas dynamics phenomena inside of the pulse tube. Figure 4 shows the time variation of velocity profiles throughout the entire pulse tube simulation in three locations: the cold end, the middle of tube, and the hot end. The phase difference in oscillating flow between the tube center and near the tube wall is clearly visible. The time variation of the temperature profile is also shown in Figure 4. To verify the simulation method and model, the simulation results were compared with experimental data obtained on the pulse tube at the Mechanical Engineering Laboratory MITI.5,6,7 The operational frequency was 2 Hz, which is the operational frequency for best performance of this pulse tube, and the pressure ratio was set to 1.8 to match the experiment. Figures 5 to 7 present a

comparison of the simulated and experimental gas temperature profiles versus time. Note that the simulated gas temperature oscillation has a much larger amplitude than the experimental data; this is because the experimental temperature probes have significant heat capacity and cannot respond to the rapidly alternating working gas temperature. The experiments tend to show the average value. A similar level of gas temperature profile is observed.

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367

Figure 4. The time variation of velocity profiles and temperature profiles (velocity profiles are observed at the cold end, middle, and hot end of the pulse tube).

A comparison of the pulse tube wall temperature obtained from the experiment and simulation

is shown in Figure 7. The temperature profile from the simulation agrees relatively well with the experiment and validates the simulation. Thus, the simulation appears to appropriately describe the gas dynamics inside the pulse tube. The refrigerator performance dependence on operational frequency was also investigated. The optimum operational frequency for the actual pulse tube refrigerator to achieve its minimum cold-

end temperature is approximately 2 Hz. The simulation results, presented in Figure 8, agree in that they show the tube temperature profile decreases (improves) with lower operational frequency.

Figure 5. Pulse Tube Pressure (Basic Type, Frequency : 10Hz).

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Figure 6 Pulse tube gas temperature versus time.

Figure 7. Comparison of experimental data and simulation results; pulse tube gas temperature

(basic type pulse tube, frequency = 10 Hz, pressure ratio = 1.8).

Figure 8. Pulse tube gas temperature profiles at various operational frequencies (basic type pulse tube, pressure ratio = 1.8)

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369

CONCLUDING REMARKS A simulation program for viscous compressible flow has been developed to study the details of fluid motion and gas dynamics in a pulse tube refrigerator. The gas dynamics and the heat conduction of the working gas in the regenerator are appropriately simulated by the pipe type regenerator modeling. The simulated temperature profile of the pulse tube agrees with the experimental data. The program has the capability to study the entire pulse tube refrigerator including the detailed gas dynamics inside the boundary layer near the tube wall. Future work will address the optimization of orifice and double-inlet type pulse tube refrigerators.

REFERENCES 1. W.E. Gifford and R.C. Longsworth, “Surface Heat Pumping,” Advances in Cryogenic Engineering, Vol.11, Plenum Press, NY (1965), p. 171.

2. Y. Hozumi and M. Murakami et al., “Numerical Study of Pulse Tube Flow,” Cryocoolers 9, Plenum Press, NY (1998), p. 321.

3. Y. Hozumi and M. Murakami, “Numerical Study of Gas dynamics inside of a Pulse Tube Refrigerator,” Advances in Cryogenic Engineering, Vol.45, Plenum Press, NY (1999), p. 167. 4. J.M. Lee and P. Kittel “Higher Order Pulse Tube Modeling,” Cryocoolers 9, Plenum Press, NY (1997), p. 345. 5. M. Shiraishi and N. Nakamura and M. Murakami, “Visualization Study of Velocity Profiles and Displacements of Working Gas Inside a Pulse Tube Refrigerator,” Cryocoolers 9, Plenum Press, NY (1997), p. 355. 6. M. Shiraishi et al., “Visualization Study of Flow Fields in a Pulse Tube Refrigerator,” Advances in Cryogenic Engineering, Vol.43, Plenum Press, NY (1998). 7. M. Shiraishi et al., “Start-up Behavior of Pulse Tube Refrigerator,” Advances in Cryogenic Engineering, Vol.41, Plenum Press, NY (1996), p. 1455. 8. R.W. MacCormack and G.V.Candler, “The Solution of The Navier-Stokes Equations using GaussSeidel Line Relaxation,” Computer & Fluid, Vol. 17, No.1 (1989), p. 135. 9. A.F. D’Souza and Oldenburger, Tran. ASME, D, 86-4 (1964-9), p. 589.

Visualization of DC Gas Flows in a Double-Inlet Pulse Tube Refrigerator with a Second Orifice Valve M.Shiraishi, K.Takamatsu1, M.Murakami1, A.Nakano, T.Iida2, and Y.Hozumi3

Mechanical Engineering Laboratory, MITI Tsukuba, Ibaraki 305-8564 Japan 1 University of Tsukuba Tsukuba, Ibaraki 305-8573 Japan 2 National Space Development Agency of Japan Tsukuba, Ibaraki 305-8505 Japan 3 Chiyoda Corporation Yokohama 230-9601 Japan

ABSTRACT We have observed the secondary flow induced in a double-inlet pulse tube refrigerator by using a smoke-wire flow visualization method and investigated the effects of the opening of the bypass valve on the flow behavior of the secondary flow, especially of the DC flow. Also, the effect of a second orifice valve on the secondary flow has been visually investigated. Based on the observations, the relationship between the cooling performance and the dynamic behavior of the secondary flow has been determined. It has been found that for the double-inlet pulse tube refrigerator, a DC flow is induced by opening the bypass valve, and that the DC flow is strengthened with additional opening of the valve. Further, the behavior of secondary flow in the pulse tube is well

modeled as a superposition of the DC flow and the convection of acoustic streaming driven by the oscillating main flow. The valve opening for optimum cooling performance is found to be that that balances the DC flow and the acoustic streaming to reduce the net velocity of the secondary flow to zero in the core region of the pulse tube. Similarly, it is found for a double-inlet pulse tube refrigerator with a second orifice valve, that the optimum cooling performance again corresponds to near

zero secondary flow in the core region. INTRODUCTION

In a pulse tube refrigerator, several kinds of secondary flow are induced; these include acoustic streaming associated with the oscillating main flow, natural convection caused by the difference in gas temperature between the hot and cold ends, and DC flows resulting from closed-loop flow

paths as in a double-inlet configuration.1,2,3,4 Here, we define a secondary flow as any kind of

induced flow except the main oscillating flow in the pulse tube. Acoustic streaming and natural Cryocoolers 11, edited by R.G. Ross, Jr.

Kluwer Academic/Plenum Publishers, 2001

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convection lead to a deterioration in cooling performance, whereas DC flow does not always degrade the cooling performance. Recently, it has been empirically determined that DC flows can have a serious effect on the cooling performance of multistage pulse tube refrigerators with several closed-loop flow paths, especially in the case of refrigerators reaching temperatures below 4 K.5,6 How to adjust the opening of valves, i.e. tune the DC flow, is an important aspect of optimizing the performance of such refrigerators.7,8 There is evidence that a suitable DC flow improves the cooling performance, but a DC flow can also lead to unstable oscillation of the cold end temperature of a pulse tube refrigerator.8,9 Recently, it was reported that a second (or double) orifice valve that is installed between the reservoir and a supply or return line of the compressor can significantly improve the cooling performance when optimally installed and adjusted.10,11 However, some valve installations bring no advantage to the cooling performance, and the reason is not well understood. The second orifice valve is thought to work as a control means for the secondary flow. It is of importance to understand the role of the second orifice valve and its effect on the secondary flow. The behavior of DC flows has been investigated only by measuring the change of the wall temperature distribution. The relation between the flow behavior and cooling performance, the temperature distribution, and the optimum condition for bypass and second orifice valves has not

yet been clarified. Visual observation is surely one of the best ways to understand the flow behavior, but there are very few reports concerning visualization experiments. The objective of this study was to observe the secondary flow in a double-inlet pulse tube refrigerator and to clarify the relationship between the flow behavior and the cooling performance. We also investigated the effects

of DC flow and second orifice valve opening on the secondary flow.

EXPERIMENT The experimental apparatus for the visualization study is schematically shown in Fig. 1. The pulse tube, which is 16 mm in diameter and 320 mm long, is made of a transparent plastic tube. The regenerator is 18 mm in diameter, 170 mm long, and composed of a pile of #100 stainless-steel screens with a wire diameter of 0.1 mm. The bypass line connects the inlet of the regenerator and the hot end of pulse tube through the bypass valve installed near the hot end. The reservoir, which is composed of a plastic vessel with a volume of about 10 times as large as the pulse tube, is connected to the high or low pressure lines through the second orifice valve. Needle valves are used as the bypass valve (Nupro, type BM), the orifice valve (type BM), and the second orifice valve (type M). Generally, needle valves have an asymmetrical flow resistance with respect to the flow direction; that is to say, gas flow in the direction of the arrow indication on the valve is not the same as that in the opposite direction. As a preliminary experiment, the flow resistance was investigated

by introducing a steady flow of air. In this experiment, air was used as the working gas for the pulse

Figure 1. A schematic drawing of experimental apparatus.

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373

tube refrigerator. The relationship between pressure drop and mass flow rate was obtained, and the data were used to estimate the mass flow rate through the bypass valve from the measured pressure difference. The resistance to air flow in the direction of the arrow is larger than that for the opposite

flow direction, but the difference is less than 12% under experimental conditions. The bypass valve was set so that the arrow direction is toward the hot end as shown in the figure. The pressure oscillation was generated via a rotary valve that periodically introduced pressurized air of about 0.2 MPa, and then released it to the atmosphere. The smoke-wire is a thin tungsten wire with 0.1 mm diameter. Both ends of the wire were soldered to copper supports that act as

electrodes and structural supports to keep the wire taut. The wire was stretched across a pulse tube cross section through a pair of tube fittings installed at the middle point of the pulse tube. Pressure transducers of the strain gauge type were installed near the inlet of the regenerator, at the hot end of the pulse tube, and in the reservoir as shown in the figure. Mineral insulated fine thermocouples (CA) of 0.15 mm in diameter were inserted into the gas space at the cold and hot end to evaluate the

cooling performance. The experimental conditions of frequency and amplitude of pressure oscillation were determined by balancing competing needs. Higher frequency is desirable because it yields a larger number of cycles in a limited time of smoke-line life. On the other hand, at higher frequency the gas moves rapidly, and the smoke-line fades out more quickly. In the end, 6 Hz was selected as the best frequency. As the gas velocity increases with an increase of the amplitude of the pressure oscillation, the observations also become more difficult. We found an amplitude of about 1.2, which is defined as the ratio between the high and low pressures, as a good compromise. The opening of the orifice valve, not the second orifice valve, was adjusted so as to maximize the difference in the gas temperatures between the cold and hot ends under the condition of a closed bypass valve. After that, the opening of the orifice valve was fixed throughout the experiment. During the conduct of the experiment, the smoke-wire surface is coated with paraffin, it is stretched in the tube, and the lead wires are connected to the high voltage pulse generator. The pulse tube is then operated under the prescribed condition. After confirming a steady state is reached by monitoring the gas temperature, a smoke-line is emitted. The movement of the smoke-line is viewed and recorded using a high speed video camera with a frame rate of 400 frames/sec over 10 cycles. The pressure and the gas temperatures are also measured. RESULTS AND DISCUSSION

Observation of Oscillating Main Flow and Secondary Flow

Figure 2 shows a typical oscillatory movement of the smoke-line. The smoke-line is emitted at the moment the gas starts to move toward the hot end, so in the figure the initial motion is upward. The shape of the smoke-line in the core region is gradually deformed from its original shape of a

Figure 2. A typical oscillatory movement of a smoke-line in the pulse tube. Vb = 0. Each picture is selected at intervals of 20 msec from the original sequential pictures. A reversal black-and-white gradient is

used to emphasize the smoke-line.

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Figure 3. A comparison of the smoke-line deformations for three different valve openings; near the optimum opening for maximum performance from the double-inlet configuration.

is

straight line into a concave parabolic line over time as the result of the secondary flow. With increased time its dent becomes steeper, and the length of the smoke-line between the tips near both walls and the vertex in the core region increases. To distinguish the secondary flow component from the oscillating main flow, we examine smoke-lines at the moment of flow reversal in each cycle. Typical results are shown in Fig. 3 for every other smoke-line in the first 5 cycles, for three different openings of the bypass valve Here, (and which appears in a later section) is an arbitrary unit of opening measured by a scale on the valve. The case corresponds to a closed bypass valve, and thus to an orifice pulse tube configuration. For this case, the length between the tips near the walls and the vertex in the core region is elongated toward the cold (left in the figure) and the hot end (right) with the passage of cycles. This result indicates the existence of an axisymmetric convection where the fluid

goes toward the cold end in the core region, and the associated return flow goes toward the hot end along the peripheral wall. This convection is one of the typical secondary flows regarded as acoustic streaming driven by the oscillating main flow.1 In the case of which is in a region of the optimum opening, only the smoke-line near the wall is elongated toward the hot end. The smoke-line in the core region remains around the initial position of the first cycle. This indicates that the secondary flow is diminished to almost zero in the core region and it only exists near the wall. For further increases in the valve opening,

50, the smoke-line drifts toward the hot end as a whole. The smoke-line deforms into an entirely different shape from those in the former cases. The tips near the wall disappear and the concave parabolic line in the core region changes to a convex one. This result indicates that the secondary

flow goes toward the hot end at every radial position and the velocity is much faster in the core region than near the wall.

Velocity Estimation of Secondary Flow The observed smoke-lines were converted into the computer-identified smoke-lines by using an image-processing system. The results are shown in Fig. 4 as the smoke-lines at every other cycle in the first 5 cycles, that is, 1st, 3rd and 5th cycles. This figure shows that the opening of the bypass valve

has serious effects on the flow behavior of the secondary flow in the double-inlet pulse

tube configuration. In the case of smaller than 20, the smoke-line in the core region shifts toward the cold end with an increase of oscillation cycles, but the overall displacement across the cross section decreases with increasing valve opening. In the case of larger than 20, the direction of drift of the smoke-line in the core region changes toward the hot end, but the smoke-line near the wall still keeps shifting toward the hot end. This result is well explained by considering the generation of DC flow in the direction from the cold to the hot end, which is superposed on the convection of acoustic streaming driven by the oscillating main flow. The magnitude of the DC flow depends on the valve opening and becomes larger with an increase in the valve opening; in contrast, the valve opening has little affect on the convection, as long as the pressure wave attributes such as the frequency and compression ratio are kept the same. The net secondary flow is determined as the superposition of both flows. Consequently the direction of drift in the core region changes. It is

VISUALIZATION OF DC GAS FLOWS IN A DOUBLE-INLET PT

375

Figure 4. A change of the displacement of smoke-lines after 1, 3 and 5 cycles in the double-inlet pulse

tube configuration with the increase of bypass valve opening corresponds to the orifice pulse tube configuration and to the optimum valve opening for the cooling performance.

worth while noting that the acoustic streaming component is not affected by the valve opening and always exists irrespective of the valve opening. The radial velocity profile of the secondary flow is calculated based on the drift distance of the smoke-line during two cycles, from the 2nd to the 4th cycles. Here, a one-dimensional flow along

the axis of the pulse tube is assumed. Figure 5 shows the measured variation of the profiles with valve opening. Negative velocity corresponds to flow toward the cold end, and positive toward the hot end. The solid curves are the fittings of fourth-order polynomials through each profile. In the

case of

the net mass flow rate passing across a cross section was calculated from the velocity

Figure 5. Change of the radial velocity profiles of secondary flow in the double-inlet pulse tube configuration with an increase of the bypass valve opening. Solid curves are the fitting of 4th-order polynomials through each profile. Negative velocity corresponds to the flow toward the cold end.

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Figure 6. A change of the radial velocity profiles of DC flow in the double-inlet pulse tube configuration with increasing the bypass valve. Solid curves are the fitting curves of quadratic polynomial. A positive

velocity corresponds to the flow toward the hot end.

profile in the figure to confirm that the net mass flow rate was balanced as it should be for acoustic streaming convection in a closed tube of the orifice pulse tube configuration. It was found that the

mass flow rate toward the cold end in the core region agrees with that toward the hot end near the wall within 10%. By considering the accuracy of the experiment, this result leads to the conclusion that the net mass flow rate is well balanced and that the secondary flow is dominated by the convection of acoustic streaming.3 With increasing the core region of negative velocity disappears, and the velocity becomes positive over the cross section, as explained in Fig. 4.

By using the data presented in Fig. 5, we extracted the component of the DC flow by subtracting the component of the convection of acoustic streaming for from the measured convection with other openings. Here, we assumed that the secondary flow for the double-inlet configuration is composed of a superposition of the DC flow on the convection for The velocity profiles of the DC flow are shown in Fig. 6. Solid curves are again the fitting curves of quadratic polynomials. The data are well correlated with the quadratic polynomial approximations, considering the accuracy of the visualization, except for small openings of 10 and 20. These parabolic velocity profiles are similar to those of the Poiseuille flow in a tube. The velocity increases with an increase in valve opening as expected from the results in Fig. 4. The mass flow rate of the DC flow was calculated using two methods: one based on the velocity profiles expressed by the solid curves in Fig. 6, and the other based on the measured pressure difference through the bypass valve combined with the pressure drop characteristics that were obtained from our preliminary experiment that used a steady flow (as discussed in the earlier 'Experiment' section). In the case of the latter method, the mass flow passing through the bypass valve changes the direction of flow with respect to the pressure difference in a cycle. However, the calculation shows that, irrespective of the valve opening, the mass flow rate of the outflow at the hot end through the bypass valve is larger than that of the inflow. This result indicates that the DC flow in the pulse tube is induced toward the hot end, and this agrees with the observations of Fig. 6 relative to flow direction. The mass flow rates, calculated using both methods, are compared in Fig. 7. The mass flow rate calculated from the pressure drop is twice as large as that converted from the velocity profile. This discrepancy is probably due to the measured pressure drop characteristics of the valve, and most likely because the applied pressure drop characteristics were obtained for steady flow, not for unsteady flow. The actual pressure difference essentially originates from the unsteady oscillating flow. Therefore, we consider the mass flow rates converted from the velocity profiles to be closer to reality than those calculated from the pressure drop.

VISUALIZATION OF DC GAS FLOWS IN A DOUBLE-INLET PT

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Figure 7. A comparison of estimated mass flow rates calculated from the velocity profiles of DC flow and from the measured pressure difference through the bypass valve.

Figure 8. A change of the velocity of secondary flow and the cooling performance for the double-inlet

pulse tube configuration as a function of Core corresponds to the velocity of the secondary flow in the core region and near wall to near the wall of pulse tube.

Figure 8 shows the relation between the cooling performance and the velocity of the secondary flow in the core region. The larger temperature difference means better performance. The results in Fig. 8 indicate that the region of optimum performance corresponds to the region where the absolute value of the velocity decreases to almost zero, i.e., where the convection is cancelled by the DC flow in the core region. Therefore, it is concluded that the bypass valve can work as the means of adjusting the DC flow to a suitable level in the double-inlet pulse tube configuration. Secondary Flow in Double-inlet Configuration with a Second Orifice Valve

In the same manner as in the above experiment, the secondary flow in a double-inlet configuration with a second orifice valve was visually observed, and the role of the second orifice valve was investigated. Figure 9 shows a comparison of the smoke-line movement for two cases of connections of the second orifice valve. The 'high pressure line' case is for a gas line connection between the reservoir and the high pressure supply line through a second orifice valve as shown in Fig. 1; the 'low pressure line' data is for a connection to the low pressure return line. The doubleinlet valve was kept at a constant opening of throughout the experiment; that is the same condition of in Fig. 4. As shown in Fig. 4, the secondary flow already exists in the pulse tube even if the opening of the second orifice valve is zero. In the case of the secondary flow in the core region flows toward the hot end due to a larger DC flow than convection. For the high pressure line connection, an extra flow is induced in the direction from the hot to

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Figure 9. Variation in the displacement of the smoke-lines for the double-inlet pulse tube configuration for different openings of the secondary orifice valve Vso. The 'high pressure line' corresponds to a connection between the reservoir and the high pressure supply line through the secondary orifice valve,

while the 'low pressure line' is for connection to the low pressure return line. The top figure corresponds to

the double-inlet pulse tube configuration with

(see for

in Fig. 4).

the cold end with increasing opening of the second orifice valve. This extra flow goes counter to the

existing DC flow and reduces the DC flow. If the opening is adjusted so that the DC flow is reduced to a suitable level, the DC flow and the convection cancel each other in the core region as in the case of in Fig 9. Further increase in the valve opening makes the DC flow decrease to a small level so that the secondary flow in the core region turns toward the cold end. With a low pressure line connection, the extra flow is induced in the same direction as the

existing DC flow, from the cold to the hot end. Both flow components flow in the same direction so that the displacement of the smoke-lines only increases with increasing valve opening. The velocity profiles were determined in the same manner as in the previous discussion, and

the effect of the velocity in the core region on the cooling performance was investigated. Figure 10 shows a comparison of the flow velocity and the cooling performance between both connections.

Figure 10. Measured dependence of the velocity of secondary flow and cooling performance on secondary orifice value opening for a double-inlet pulse tube configuration. The velocity corresponds to that in the core.

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379

In the case of the high pressure line connection, the optimum operating region, where the temperature difference is maximized, coincides with the region where the absolute value of the

velocity is reduced to roughly zero. On the other hand, in the case of the low pressure line connection, the velocity increases with increasing opening and the improvement in the cooling performance is not clearly distinguishable compared to the former case. This fact indicates that the cooling performance should be improved if the secondary flow in the core region is reduced to roughly zero. This conclusion is found to hold for the double-inlet configuration without the second orifice valve. We may, therefore, reasonably conclude that the cooling performance of the double-inlet pulse tube refrigerator should be improved by reducing the secondary flow in the core region.

SUMMARY We have observed the secondary flow induced in the double-inlet pulse tube refrigerator by using a smoke-wire flow visualization method. Some results are as follows: 1. The secondary flow in a double-inlet pulse tube refrigerator is well explained by considering

the superposition of the DC flow on the convection of acoustic streaming driven by the oscillating main flow. 2. The magnitude of the DC flow increases with increasing bypass valve opening. The cooling performance is improved by adjusting the valve opening to reduce the velocity of the secondary flow to almost zero in the core region. 3. For a double-inlet pulse tube refrigerator with a second orifice valve, an extra flow is introduced by opening the second orifice valve and the cooling performance can be improved by adjusting the extra flow to reduce the velocity of the secondary flow in the core region to roughly zero. REFERENCES 1. Olson, J.R. and Swift, G.W., “Suppression of Acoustic Streaming in Tapered Pulse Tube,” Cryocoolers 10, Kluwer Academic/Plenum Publishers, New York (1999), pp. 307-313.

2. Thummes, G. et al., “Convective Heat Losses in Pulse Tube Coolers:Effect of Pulse Tube Inclination,” Cryocoolers 9, Plenum Press, New York (1997), pp. 393-402.

3. Shiraishi, M. et al., “Visualization Study of Secondary Flow in an Inclined Pulse Tube Refrigerator,”

Advances in Cryogenic Engineering, Vol. 45 (2000), pp. 119-126. 4. Gedeon, D., “DC Gas Flows in Stirling and Pulse Tube Cryocoolers,” Cryocoolers 9, Plenum Press, New York (1997), pp. 385-392.

5. Wang,C., Thummes, G. and Heiden.C., “Control of DC Gas Flow in a Single-Stage Double-inlet Pulse Tube Cooler,” Cryogenics, Vol. 38 (1998), pp. 843-847.

6. Wang,C., Thummes, G. and Heiden.C., “Effects of DC Gas Flow on Performance of Two-stage 4 K Pulse Tube Coolers,” Cryogenics, Vol. 38 (1998), pp. 689-695. 7. Kotsubo, V., Huang, P. and Nast, T. C., “Observation of DC Flows in a Double Inlet Pulse Tube,” Cryocoolers 10, Kluwer Academic/Plenum Publishers, New York (1999), pp. 299-305. 8. Charles, I., Duband, L. and Ravex, A., “Permanent Flow in Low and High Frequency Pulse Tube Coolers-Experimental Results,” Cryogenics, Vol. 39 (1999), pp. 777-782. 9. Duband, L., et al., “Experimental Results on Inertance and Permanent Flow in Pulse Tube Coolers,”

Cryocoolers 10, Kluwer Academic/Plenum Publishers, New York (1999), pp. 281-290. 10. Chen, G. et al., “Experimental Study on a Double-orifice Two-stage Pulse Tube Refrigerator,” Cryogenics, Vol. 37 (1997), pp. 271-273. 11. Yang, L., Zhou, Y. and Liang, J.,”DC Flow Analysis and Second Orifice Version Pulse Tube Refrigerator,” Cryogenics, Vol. 39 (1999), pp. 187-192.

A Gifford-McMahon Cycle Cryocooler below 2 K T. Satoh, A. Onishi*, I. Umehara**, Y. Adachi**, K. Sato** and E.J. Minehara*** R&D Center, Sumitomo Heavy Industries, Ltd. Hiratsuka, Kanagawa 254-0806, Japan *Precision Products Division, Sumitomo Heavy Industries, Ltd. Tanashi, Tokyo 188-0001, Japan ** Yokohama National University Yokohama, Kanagawa 240-0067, Japan ***FEL Lab., Japan Atomic Energy Research Institute Naka, Ibaraki 319-1195, Japan

ABSTRACT According to theory, a Gifford-McMahon (GM) cycle cryocooler with 4He cannot cool below 2 K because of the 4He superfluid transition near this temperature. However replacing 4 He by 3He removes this temperature limitation. The cooling performance of a GM cryocooler with a magnetic regenerator material is investigated using 3He. The minimum temperature of 2.3 K with 4He goes down to 1.65 K when the 4He working fluid is replaced by 3He. The maximum cooling capacity at 2 K is 53.9 mW with a compressor power of about 2.5 kW, and the cooling capacity at 4.2 K is enhanced by more than 20%. The effect of a new regenerator material on the cooling performance was also investigated. The minimum temperature decreased to 1.64 K and the cooling capacity at 2 K improved to 57.1 mW with the use of this material in the bottom 40% of the regenerator. INTRODUCTION We can now reach the 2 K temperature region very easily using a regenerative cryocooler such as a GM cryocooler or a pulse tube cryocooler with magnetic regenerator materials. Nagao et al.1 reached 2.09 K with a three-stage GM cryocooler, while the Giessen group2 obtained 2.07 K with a liquid nitrogen precooled pulse tube cryocooler; this is the lowest temperature reached to date with a regenerative cryocooler. These cryocoolers used 4He as the working fluid, and their lowest temperature is limited by the superfluid transition of 4He near 2 K. The only possibility to reach below 2 K is using 3 He instead of 4He as the working fluid in a regenerative cryocooler. According to the theoretical analysis of Xu et al.,3 temperature does not change by adiabatic expansion or compression if the thermal expansion coefficient is zero. Since the thermal expansion coefficient is zero for 4 He at the superfluid transition temperature, and even negative at lower temperatures, we canCryocoolers 11, edited by R.G. ROM, Jr. Kluwer Academic/Plenum Publishers, 2001

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not expect to cool below 2 K using 4He. For 3He, on the other hand, the expansion coefficient is still positive below 2 K, and we can expect to cool below 2 K.

Xu et al.3 investigated the performance of a three-stage pulse tube cryocooler. It was found that the minimum temperature could be reduced from 2.19 K to 1.87 K by replacing the 4He working fluid by 3He. This indicates the possibility that a GM cryocooler with 3He could also achieve

temperatures below 2K. Following this lead, we have investigated the cooling performance of a two-stage GM cryocooler with 3He working fluid and a magnetic regenerator material. EXPERIMENTAL SETUP

Figure 1 shows a schematic diagram of our two-stage GM cryocooler. The compressor is a model CKW-21 from Sumitomo Heavy Industries, Ltd. The rated input power of the standard unit is 2.6 kW at its driving frequency of 60 Hz. However, in this experiment, to save 3He, shorter flex

hoses were used and the adsorber of the compressor was changed to a smaller one. The 3He volume was limited to about 230 liters at standard conditions and the charge pressure was about 15 bar; the purity of the 3He was 99.95%. The cylinder of the cold head is made of thin stainless steel tube, the dimensions of which are as follows: the first-stage cylinder diameter is 52 mm with an inner length of 191.5 mm, and the second-stage cylinder diameter is 25 mm with an inner length of 165 mm.4 A copper block is silver brazed to the stainless steel tube at the end of the 2nd-stage cylinder to improve heat exchange. Each cylinder contains a displacer, in which regenerator material is stuffed. The displacer stroke is 20 mm and the displacers are driven by an AC synchronous motor that can vary the cycle speed by changing the supply frequency. The first-stage regenerator is composed of

Figure 1. Schematic of our two-stage GM cooler.

A GIFFORD-McMAHON CYCLE CRYOCOOLER BELOW 2 K

383

Figure 2. Specific heat of magnetic regenerator materials. #180 copper screens in the higher-temperature region, and lead spheres in the lower-tempera-

ture region. The second-stage regenerator is composed of lead spheres (150g) in the highertemperature side, and magnetic regenerator material in the lower-temperature side, spheres of diameter 0.2 mm to 0.5 mm and a new material crushed powder) of size 0.2 mm to 0.5 mm were used. Figure 2 shows the temperature dependence of specific heat for and below 20 K. has a specific heat peak due to its magnetic phase transition to its antiferromagnetic state at 7 K and 10 also has a similar type peak of specific heat at 2 K due to its magnetic phase transition to its anti-ferromagnetic state.6 Since it has a larger heat capacity at 2 K than as shown in Fig. 2, is a good candidate to improve the cooling performance in the 2K region. As shown in Table 1, four types of second-stage regenerators were prepared to investigate the effect of the material on the refrigerator's cooling performance below 4 K. The intake/exhaust valve timing was optimized for each case. RESULTS AND DISCUSSION

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 that was used to measure the first-stage temperature (see Fig. 1). An electric heater of manganin

wire was installed on each stage to allow known heat loads to be applied independently. To reduce the radiation heat load from room temperature to the second stage, a thermal shield was attached to the first stage.

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Figure 3. Dependence of no-load temperature on cycle speed.

Figure 4. Relationship between heat load and the second-stage temperature.

No-load Temperature Figure 3 shows the cycle speed dependence of the first and second-stage temperature with no heat load. The No. 1 regenerator was used for this experiment. The compressor power was about 2.5 kW for both cases. As shown in this figure, both the first and the second-stage

temperatures are lower with 3He than with 4He for all the investigated cycle speeds. The second-stage temperature goes down when the cycle speed decreases. The lowest temperature is 2.27 K at 54 rpm with 4He, and 1.65 K at 54 rpm with 3He. Temperatures below 2 K are obtained at cycle speeds lower than 90 rpm. The first-stage temperature, on the contrary, goes up when the cycle speed decreases. The relationship between the second-stage temperature and the heat load is shown in Fig. 4; the cycle speed is 60 rpm. The cooling capacity of the cooler with 3He is improved compared to that of the cooler with 4He not only in the 2 K region, but also in the higher temperature region. The cooling capacity at 4.2 K is 0.63 W with 4He and 0.82 W with 3He; this is improved more than 20%. The cooling capacity at 2 K with 3He is 53.9 mW. Effect of on Cooling Performance The specific heat of the regenerator material is a very important factor effecting the performance of regenerative cryocoolers. has a larger specific heat below 2 K than as shown in Fig. 2. Thus, is expected to be an effective regenerator material to improve the cooling capacity below 2 K. Four types of the second-stage regenerators were prepared to investigate the effect of on the cooling performance. Fifty percent 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 various regenerators, either or a combination of the two was loaded into the lower temperature region. The lowest temperature 1.64 K was obtained at 54 rpm and 48 rpm with the No.3 regenerator configuration installed and no heat load on the first stage. This is the lowest temperature achieved by any regenerative cryocooler. The cooling capacity at 2 K is plotted against regenerator makeup in Fig. 5 for cycle speeds of 48 rpm and 60 rpm. The cooling capacity at 60 rpm was improved 8% when 40% of the was replaced by Figure 6 shows the cycle speed dependence of the cooling capacity at 2 K. The largest cooling capacity, 57.1 mW, was obtained at 54 rpm with the No.3 regenerator.

A GIFFORD-McMAHON CYCLE CRYOCOOLER BELOW 2 K

Figure 5. Cooling capacity at 2K versus regenerator makeup.

385

Figure 6. Cycle speed dependence of cooling capacity with regenerator No. 3.

Figure 7. The cooling capacity below 4.2 K. The effect of on the cooling capacity is shown in this figure.

Figure 7 shows the relationship between the heat load and the second-stage temperature with regenerators No.1 and No.3. The heat load versus the second-stage temperature curves cross at about 2.3 K. This figure shows that the cooling capacity is enhanced below 2.3 K by replacing a part of the with However, the performance is poorer above 2.3 K; this is because the specific heat of is much lower in this higher temperature region above

2.3 K. The cooling capacity at 2 K does not fall at the first stage temperature up to about 50 K. Effect of Orientation on Cooling Performance The cooling performance around the 4K region of a GM cooler is much influenced by the orientation. Figure 8 shows the orientation dependence of the first and the second-stage temperature with no heat load. An orientation of 0° means that the cold head is set vertical with the room temperature end up. The cycle speed is 60 rpm and the No.4 regenerator is used in this experiment. The second-stage no-load temperature is nearly the same at 0° and 180°, and even lower at 180° compared to 0°. The maximum temperature is obtained at 135°, but it is still much lower than 2 K. On the other hand, the first-stage temperature goes up when the orientation angle increases.

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Figure 8. Orientation dependence of no-load temperature.

Figure 9. Orientation dependence of cooling capacity at 2K

The cooling capacity at 2 K has the orientation dependence shown in Fig. 9. As shown in

this figure, the cooling capacity at 90° is the lowest, and about 17% lower than that at 0°. The largest cooling capacity is achieved at 180°. This orientation dependence is similar to that of

the cooling capacity at 4.2 K with 4He working fluid.7 CONCLUSIONS

The cooling performance of a two-stage GM cryocooler with has been investigated using 3He as the working fluid. The lowest temperature goes down from 2.27 K to 1.65 K when the 4He is replaced by 3He, and the cooling capacity at 2 K is 53.9 mW at 60 rpm. The cooling capacity at 4.2 K is improved more than 20% with 3He. The effect of the new magnetic regenerator material on the cooling performance has also been investigated. The low-

est temperature goes down to 1.64 K and the cooling capacity at 2 K is improved to 57.1 mW when 40% of the in the lower temperature end of the regenerator is replaced by NdInCu2. A GM cooler with magnetic regenerator can be a reliable and practical 2 K cooler when the 4 He working fluid is replaced by 3He. REFERENCES 1. M. Nagao, T. Inaguchi, H. Yoshimura, S. Nakamura, T. Yamada and M. Iwamoto, “Generation of Superfluid Helium by a Gifford–McMahon Cycle Cryocooler,” Proceedings of the 6th International Cryocooler Conference, Plymouth, MA, DTRC-91/002, David Taylor Research Center (1991) pp. 37-47. 2. G. Thummes, C. Wang, S. Bender and C. Heiden, “Pulsrohrenkuhler zur erzeugung von temperaturen im bereich des flussingen heliums”, DKV-Tagungsbericht, 23 (1996) (Jahrgang Band), pp. 147-159.

3. M.Y. Xu, A.T.A.M. de Waele and Y.L. Ju, “A Pulse Tube Refrigerator below 2K,” Cryogenics, vol. 39

(1999), pp. 865-869. 4. T. Satoh, R. Li, H. Asami, Y. Kanazawa and A. Onishi, “Development of High Efficiency 0.5W Class 4K GM Cryocooler,” Cryocoolers 10, Plenum Press, New York (1999), pp. 575-580.

5. J. Bischof, M. Divis, P. Svoboda and Z. Smetana, “Specific Heat of in Magnetic Fields,” Phys. Stat. Sol. (a) 114 (1989), p. 229. 6. K. Sato, Y. Ishikawa and K. Mori, “Magnetic specific heat of light rare-earth Heusler compounds ( Ce, Pr, Nd and Sm),” J. Magn. Mater., 104-107 (1991), pp. 1435-1436. 7. Sumitomo Heavy Industries, Ltd. SRDK-408D catalog.

High Efficiency, Single-Stage GM Cryorefrigerators Optimized for 20 to 40 K C. Wang and P. E. Gifford

Cryomech, Inc. Syracuse, NY 13211, USA

ABSTRACT

Cryomech, Inc. has developed (2) single-stage GM cryorefrigerators optimized for the 2040K temperature range. The Models AL230 and AL330 were designed for high efficiency and high power cooling of HTc Superconducting Devices. The structural parameters of the AL230

and the AL330 were optimized using a numerical simulation program. The regenerative materials used are the industry standard lead spheres and phosphor bronze screens. Both the AL230 and the AL330 obtain the minimum temperature of <11 K. The AL230 provides 30 W at 20 K, or 65 W at 30 K when operating with a CP950 Helium Compressor Package, with an input power of 5.4 kW (60 Hz). The AL330 provides 45 W at 20 K, or 101 W at 30 K with CP970 Compressor Package, with an input power of 7.0 kW (60 Hz). INTRODUCTION

For the next generation of HTc Superconduting Devices; such as, Fault Current Limiters, SMES, SC Motors and Magnets etc., to become viable products, higher capacity and lower cost cryorefrigerators are necessary. The HTc applications require that the prototype devices increase in size and current densities and to be cooled to < 30K. The present options of single and 2-stage GMs, which supply 10-30 watts, are not large enough. A single stage GM cryorefrigerator has a bottom temperature between 20 and 25 K. Because they cannot reach temperatures below 20 K, they cannot provide the high cooling capacities at 30 K. The small temperature difference between the bottom temperature of 25 K and 30 K will not transfer and carry away the required heat. The conventional 2-stage GM has a bottom temperature around 7 K with the second stage, but is not as efficient in the temperature range of 20 to 40 K, as the single stage GM. Also, the cost of manufacture and maintenance for a 2-stage GM is higher. Therefore, the conventional single-stage and 2-stage GMs have not been able to satisfy the requirements of the new HTc superconductor applications. Fielder et al. presented a low temperature single-stage GM1, Leybold Model RGS120-T, which has 25 W at 20 K for 6 kW power input. This single-stage GM is simple and efficient at the temperatures from 20 to 30 K.

Cryomech, Inc. has recently developed lower temperature single-stage GM cryorefrigerators, the Models AL230 and AL330. A simulation program was used to optimize the structural Cryocoolers 11, edited by R.G. Ross, Jr.

Kluwer Academic/Plenum Publishers, 2001

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Figure 1. Model of single-stage GM cold head for numerical simulation

parameters of AL230 and AL330 Cold Heads. The AL230 and AL330 have bottom temperatures of less than 11 K and are highly efficient in the temperature range of 20 K to 40 K. DESIGN OF COLD HEADS

A conventional single-stage GM Cold Head, like Cryomech AL200, is designed for efficient cooling in the temperature range of 50 to 80 K. The AL200 utilizes phosphor bronze screens as the regenerative materials. The specific heat of phosphor bronze will be very low when the temperature in the regenerator is below 50 K. Lead is a good candidate for regenerative material in the temperature range of 20 to 50 K. For the new low temperature refrigerators, Cryomech Model AL230 and AL330, utilizes a hybrid regenerator with phosphor bronze screens at warm end and lead spheres at cold end. Theoretical Optimization

A numerical simulation program was developed to optimize the structure parameters of the refrigerator. Figure 1 shows a physical model of a single stage GM for the simulation. The regenerator inlet is taken as the left boundary. The right boundary is the surface of the displacer. The basic assumptions in the simulation model are as follows: 1. ideal gas; 2. one-dimensional flow of helium; 3. constant wall temperature of cold heat exchanger; 4. adiabatic expansion process in expansion chamber; 5. axial heat conduction and shuttle loss neglected. The governing equations and method for carrying out the numerical simulations used in this paper are the same as that in Reference 2. The left boundary conditions are: pressure wave at the inlet of regenerator, and constant gas temperature flowing into the regenerator. The right boundary condition is the motion of the displacer. The normal operating parameters for the optimization program are: warm end temperature 300K, cold end temperature 30K, and operating frequency 2.4Hz. When designing the new AL230, we used many of the components from the AL200 changing only the displacer, the displacer tube and 30 K heat exchanger. The same valve motor assembly, valve plate, rotary valve, and displacer diameters were used. The AL330, and the 77 K AL300 were completely new products. For the AL330, the phasing of rotary valve, size of regenerator, and heat exchanger were all optimized by the program. The procedures for the optimization and the design are illustrated in Figure 2.

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Figure 2. Diagram of optimization and design.

Figure 3. Cutaway of the AL230 and AL330.

Physical design Figure 3 shows the cutaway of the AL230 and the AL330 Cold Heads, as well as their size. They have rotary valves for gas distribution and the displacer is driven pneumatically. The regenerator consists of two layers of regenerative materials: phosphor bronze screens in the upper part, lead spheres in the lower part. A Cryomech patented cold heat exchanger is used in the refrigerators to enhance heat

transfer. Figure 4 is a photo of the AL330, AL230 and AL200 cold heads. The AL230 and the AL330 have the same length and different diameters.

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Figure 4. Photo of AL330, AL230 and AL200. (a) AL330, (b) AL230, (c) AL200.

PERFORMANCE OF THE SINGLE-STAGE GM CRYOREFRIGERATORS

Performance of AL230 For the proof of principal our first attempt at a low temperature high capacity single stage GM was the AL230. The AL230 was first tested with an in-house CP640 Helium Compressor Package with input power around 5.5kW. Figure 5 shows cooling capacities of the AL230 with displacer strokes from 15.2mm to 25.4mm. The shorter the displacer stroke, the lower the bottom temperature of the cold head and the less cooling capacity in higher temperature range. The lowest temperature, 10.5 K, was obtained with 15.2 mm stroke. Further reducing the stroke

could not lower the temperature of the cold head. The lowest temperature may be limited by the regenerator efficiency and shuttle loss. We decided on a stroke of 17.8mm for the production of the AL230, since it maximizes the cooling capacities in the temperature range from 20K to 50K.

Figure 5. AL230 Cooling capacities with different displacer strokes

HIGH EFFICIENCY, SINGLE-STAGE GM FOR 20 TO 40 K

Figure 6. Performance graph comparing AL200 and AL230.

Figure 7. Cooling capacities of AL330 and AL300

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A comparison of cooling capacities of the AL200 and AL230 are displayed in Figure 6. The curve for AL230 in Figure 6 is a typical cooling capacity, which was obtained with 17.8mm displacer stroke and the new CP950 Compressor Package. The CP950 utilizes a helium scroll compressor and has an efficient absorber with lifetime >20,000 hours2. The AL230 has a bottom temperature of 11.5 K, 30 W at 20 K and 65 W at 30 K. Compared to the AL200, the AL230 has a higher cooling power below 60 K. Our prediction for the AL230 performance is also shown in Figure 6. We believe the discrepancy is most likely due to the fact that there was no factor for shuttle losses included in the simulation model. Performance of the AL330 The AL330 and AL300 operate with the CP970 Compressor Package. The CP970 also utilizes a helium scroll compressor module with input power approximately 7kW. The AL300 is designed for 80 K operation. Figure 7 shows the cooling capacities of the AL300 and AL330 as well as the performance predicted by our program for the AL330. The AL300 and AL330 are both working at the theoretically optimized stroke of 25.4 mm. The AL330 has a bottom temperature of 10.5 K and 101 W at 30 K. It has better performance than the AL300 when the cooling temperature is below 35 K. Table 1 lists the cooling capacity, input power and efficiency of the AL230 and AL330. The AL230 has 65 W at 30 K for 5.44 kW of input power, for a carnot efficiency of 10.7%. The AL330 has 101 W at 30 K for 7.04 kW of input power, for a carnot efficiency of 12.9%. The highest percent of carnot for the AL230 is 11.6% and for the AL330 is 14.0%. Both cryorefrigerators have their highest efficiency near 40 K which is close to the optimization

temperature of 30 K. The AL300 and AL330 have the highest efficiency and cooling capacity to date of any available GM cryorefrigerators. CONCLUSION

Single-stage GM cryorefrigerators can supply the high capacity, high efficiency, 30 K cooling requirements of many of the new HTc applications. Predictions from our simulation program can be used to design even larger systems. The Cryomech Models AL230 and AL330 are built with a proven technology. Both cryorefrigerators have the minimum temperatures of

approximately 11 K. The AL230 provides 30 W at 20 K or 65 W at 30 K for an input power of 5.44 kW. The AL330 provides 45 W at 20 K and 101 W at 30 K for an input power of 7.04 kW. ACKNOWLEDGMENT

We would like to thank R. Dausman and B. Zerkle for useful discussions and J. Cosco for testing the cryorefrigerators. REFERENCES 1. Fiedler, A., Gerban, J. and Haefner, H.U., “Efficient Single Stage Gifford-MacMahon Refrigerator Operating at 20 K”, in: Advances in Cryogenic Engineering, Vol. 43B, Plenum Press, New York (1997), pp.l823-1830. 2. Wang, C., Wu, P.E. and Chen, Z.Q. “Numerical Modelling of an Orifice Pulse Tube Refrigerator”, Cryogenics, vol.32, no.9 (1992), pp.785-790.

Remote Cooling with a G-M Cryocooler by Use of Cold Electromagnetic Valves Driving an External Flow Loop K. M. Ceridon and J. L. Smith Jr.

Massachusetts Institute of Technology Cambridge, Massachusetts, USA 02139

ABSTRACT

A major limitation of Gifford-McMahon, G-M, cryocoolers is the requirement to conduction cool the refrigeration load. A number of earlier attempts to use check valves to rectify the oscillating pressure from the G-M expansion in an external flow loop were unsuccessful because of valve leakage. A G-M cryocooler is being modified with the addition of Boreas style cold electromagnetic valves to utilize the oscillating pressure to drive the unidirectional flow in an external cooling loop. The cold valve of the Boreas cooler demonstrated minimum leakage and acceptable electrical dissipation even at 4.5 K. Preliminary analysis indicated that the G-M blow down process could be used to power the external flow without seriously degrading the G-M cooler capacity. INTRODUCTION

Cold electromechanical EM valves offer significant potential for improving cryocoolers. When fitted with two cold EM valves, a G-M cryocooler can drive an external, unidirectional flow loop

without the need for a high efficiency recuperative heat exchanger and a circulator at room temperature. As shown in Figure 1, the EM valves allow a blow down into a cold surge volume and a return flow through a load heat exchanger without any increase in the gas circulation in and out of the warm end through the charge and exhaust valves. In contrast, an oscillating flow into and out of a cold surge volume will not cool the surge volume. This is because of the regenerative heat transfer associated with the oscillating flow which blocks of energy flow from the surge volume to the G-M displacement volume. Earlier attempts to rectify the oscillating pressure of a G-M cycle with cold check valves have

generally been unsuccessful. With spring loaded check valves, the gas must enter the surge volume near the system-high-pressure and exit the surge volume near the system-low-pressure. As will be

shown, this process does not maximize the cooling capacity of the external flow loop. In addition, the closing force is typically not sufficient enough to avoid valve leakage. Electro-mechanical valves can overcome these problems. The cold EM valves are spring loaded in the closed position. Clearly, commercial solenoid valves are not suitable for application as cold EM valves. In a commercial solenoid valve, the electrical dissipation is high and the mechanical reliability is low. On the other hand, a cold EM valve can have low dissipation. The use of high purity copper reduces the coil resistance, which will also reduce dissipation. The high current used to open the valve is Cryocoolers 11, edited by R.G. Ross, Jr.

Kluwer Academic/Plenum Publishers, 2001

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GM REFRIGERATOR DEVELOPMENTS

Figure 1. Modified G-M cycle including external helium flow loop.

reduced to a low hold-open value to limit the dissipation and to prevent the impact with the full-open stop. On closing, the energy stored in the magnetic field and the energy in the valve spring is significantly dissipated in the control circuit at room temperature. This further reduces the cold dissipation and reduces the impact with the valve seat. The cold EM valve is guided by elastic flexures rather than by sliding guides. Designs of the electrical circuits that actuate the valves are critical to low loss and long life operation. The work on valves for the G-M cooler draws heavily on the technology developed for the cold one-watt Boreas cryocooler [1,2]. The Boreas valve operated at 4.5 K had a cold dissipation of only a few tens of milliwatts and proved to be quite reliable. Cold EM valves offer significant potential for improving other cryocooler cycles, especially for operation below 10 K where thermal regenerators have very limited heat capacity. An associated project in cooperation with Advanced Mechanical Technology, Inc. is applying cold EM valves to a miniature Collins cycle cryocooler for operation at 10K. The valves are to be applied to the floating piston expanders by Jones and Smith [3]. SYSTEM DESIGN

First Order Thermodynamic Analysis of System and Processes

In order to examine how the external flow loop effects the performance of the refrigerator, a simplified first order thermodynamic analysis of a modified G-M cycle was compared to that of an unmodified G-M cycle. As shown in Figure 1, the principal thermodynamic components of an unmodified G-M cycle are a regenerator, R, contained within a displacer, D, that moves between a cold volume, C, and a warm volume, W. The compressor maintains a constant high pressure, upstream of valve V1 and a constant low pressure, down stream of valve V2. In addition to these components, the modified G-M cycle has two cold EM valves, V3 and V4, a large surge volume, S, and a heat exchanger, HE. Valve V3 opens for flow from cold volume C to surge volume S. Valve V4 opens for flow from surge volume S through heat exchanger HE to cold volume C. For the purposes of this analysis, the unmodified G-M cycle is taken to consist of the following four discreet processes and four equilibrium states. State 1 W, R and C, all at D at top dead center (TDC) Process 1-2 Close valve V2 and open V1, charge W, R and C to State 2 W, R and C at D at TDC Process 2-3 Move displacer from TDC to BDC State 3 W, R and C at D at BDC Process 3-4 Close valve V1 and open Valve V2 to blow down to State 4 W, R and C at D at BDC

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Process 4-1 Move displacer from BDC to TDC In the modified cycle, valves V3 and V4, surge volume S, and heat exchanger HE are added to the cycle. The mass of gas stored in the heat exchanger HE is assumed to be small. The modified cycle consists of the following discreet processes and equilibrium states State 1 W, R, C and S all at D at top dead center (TDC) Process 1-2 Close valve V2 and open V1, charge W, R and C to

State 2 Process 2-3 State 3 Process 3-4 State 4 Process 4-5 State 5 Process 5-1

W, R and Cat S at D at TDC Move displacer from TDC to BDC W, R and Cat S at D at BDC Close valve V1 and open Valve V2 until W,R and C reach W, R and C at Sat D at BDC With V2 closed, open V3 to equalize W,R and C with S W, R, C and S at D at BDC Close V3, open V4 and V2 to equalize W,R,C and S to then

move displacer from BDC to TDC The following analysis presents the method used to perform a first order analysis on these two cycles. The performance of the two cycles is then compared to determine how the external flow loop degrades the performance of the G-M cycle. The simplified analysis of the two cycles is based on a mass and energy balance for each component and each process. As shown in Figure 1, the interfaces between components are: RWE between R and W, RCE, between C and R, SI between S and C, HEI between S and HE, and HEO between HE and C.

The G-M Cryocooler Cycle. The system starts at state 1. Process 1-2 opens valve V1 and allowing high-pressure helium to charge the regenerator R and the warm volume W to state 2. Process 2-3 fills the cold volume C with gas that is cooled as it flows through the regenerator to State

3. Work, flows up the displacer because of its motion at Process 3-4 expands the gas in volume C and the gas in the regenerator to by allowing flow out through valve V2. As the displacer moves during process 4-1 the cool gas flows through the regenerator and out valve V2. During processes 3-4 and 4-1, heat flows from the walls of volume C to the cold gas. From an energy and mass balance for volumes C and R, the cooling is given by

where is the displacement volume for the displacer. The first term is the net work up the displacer from cold to warm and the second term is the net enthalpy, h, flowing from warm to cold

along the regenerator due to its ineffectiveness for the temperature span of the regenerator. The net enthalpy flow for the commercial cooler is evaluated from the measured cooling and the displacer work term. Modified G-M Cycle with External Flow Loop. States 1, 2 and 3 as well as processes 1-2 and 2-3 are the same for the modified cycle. Thus the mass entering the regenerator R across section

RWE is the same for both cycles. The regenerator is assumed to have the same effectiveness and temperature span for both cycles. In process 3-4, gas flows out through valve V2 until the pressure in volume C and regenerator R decrease to The value for is selected as a design parameter for the cycle. Then valve V2 closes. In process 4-5, gas flows out through valve V3 into volume S until the pressure in volume C and regenerator R decreases to Then valve V3 closes and valve V2 opens to start process 5-1. The displacer moves to BDC and gas exits to through valve V2. The regenerator R, volume C and volume S go to to start the next cycle. The operating states of the modified cycle are selected so that the regenerator has the same warm-end net enthalpy flow and the same temperature span as in the simple cycle. From an energy and mass balance for volumes C, R and S, the cooling QR is again given by Eq. (1) when the electrical losses in valves V3 and V4 are neglected and there is no additional external heat leak to volume S. The addition of the external flow loop does not influence the value of the when

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GM REFRIGERATOR DEVELOPMENTS

Figure 2. Refrigeration temperature as a function of intermediate pressure,

processes 1 -2 and 2-3 match the simple cycle. Also, the value selected or

does not influence the

refrigeration The remaining task is to relate the cooling temperature of the heat exchanger HE to the low temperature defined by the span of the regenerator. The required calculations are the mass and gas

temperature or energy for R, C, and S at each state 1, 2, 3, 4, 5. The mass and energy balances together with the boundary fluxes of both energy and mass at sections RWE, RCE, SI, HEI, and HE2 allow the calculation of the energy flux across HEI during process 5-1. This is accomplished by starting with process 1-2 and following each process in sequence. Except for process 4-5, the end states of each process is determined by the initial states. Process 4-5 requires numerical integration from its initial state to find the final state at pressure By calculating the required state 1 at the end of process 5-1 for the regenerator and volume C, the energy flux across section HEO from HE to volume C for process 5-1 is determined. If gas crossing section HEO is assumed to be at the load temperature, then the value of the cooling load temperature is fixed by the energy fluxes across section HEO and HEI.

Since the cooling temperature for the modified cycle is higher than for the simple cycle, the analysis shows the penalty associated with the modification as a temperature increase at the cooling

load rather than as a reduction of the cooling

The analysis also shows how the selected value for

influences the increase in cooling temperature. Figure 2 shows the cooling temperature as a

function of the intermediate pressure

for the regenerator span that gave a cooling temperature of

77 K for the simple cycle of the commercial G-M system. Figure 2 indicates that the maximum available mass should be circulated into the surge volume and should equal Thus, process 3-4 is not necessary since it does not improve the refrigerator performance. The mass circulation in the surge volume is also maximized when the volume S is large compared with volume C. This initial analysis, also indicates that, at the same cooling load, the refrigeration temperature increases with the modifications to the cycle. When is equal to the refrigeration temperature

is 82 K. Although this higher temperature could be a result of approximations in the model, it is likely that second-order affects cause this increase in temperature. Process 4-5 puts the cooling load, into the mass circulating in the surge volume. As this mass decreases, a larger temperature rise is required of the circulating mass to absorb the cooling load. The result is a larger refrigeration

temperature at

REMOTE COOLING WITH G-M COOLER BY USE OF FLOW LOOP

397

Figure 3. Cross section of cold EM valves as placed in the cold cylinder. The valve housing with

the valve assembly appears on the left and the valve housing appears on the right.

Overall, this analysis examines the cycle only on a first-order basis. Second order effects such as losses at the valves, regenerator temperature distribution and effectiveness and external radiation

heat leak have been neglected. A more thorough analysis and experimental procedure is required to evaluate the impact of these second order affects on the cycle performance. Design of Cold EM Valves

The design of the EM valves, Figure 2, was selected to meet the following objectives: • Implement an external circulation loop with minimal modification of the existing

commercial cryocooler; • Minimize electrical and mechanical dissipation.

• Maximize operating life; • Demonstrate an electronic valve controller that provides a programmed valve opening force; • Demonstrate a removable valve assembly inserted in a valve housing; • Demonstrate a soft seat design for effective valve sealing. The existing cooler was modified by boring two holes in the cold cylinder head for attachment of the valve enclosures. The gas connection to the valve is through the end flange of the enclosure which is bolted in place to compress an indium seal. Since the gas flows through the valve assembly, all valve dissipation is effectively transferred to the helium stream. The valve assembly is composed of the following parts: Valve seat and valve head and stem; A Kel-F insert in the valve head to form a soft sealing surface; Ferromagnetic core attached to the valve stem; Solenoidal, high purity, copper-wire winding; Ferromagnetic back iron to close the magnetic circuit; Two multi-leaf flexures to guide the valve and core and prevent friction; Helical valve spring to provide valve-sealing force in addition to the pressure force.

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GM REFRIGERATOR DEVELOPMENTS

Both the valves open into the cold volume of the cylinder so that the pressure difference across the valve is in a direction that increases the valve sealing force. The cold blow down valve exits the helium directly into the cold surge volume as shown in Figure 1. The surge volume exits the gas into the load heat exchanger. For the experiment, an electric heater in the exit line from the surge volume supplies the cooling load. The load heat exchanger exits the helium back into the cold volume of the cylinder via the return valve. Manufacturing and Geometric Design Considerations. The two cold EM valves V3 and V4 are attached to the cylinder head to control the flow through the external flow loop. The attachment of the valves to the cold cylinder head is shown in Figure 3 at a 1:2 scale. The cylinder of the cryocooler has a 3-inch internal diameter and a 0.035-inch wall thickness. The stroke of the displacer is from 0 to 0.625-inch from the cylinder head. All components of the commercial cryocooler are to be maintained and modified only for the valve attachment. This geometrically limits the size and position of the valves relative to the cold cylinder. The valve assemblies are contained within the vacuum tight housings that are welded to the ports drilled through the cold cylinder head. The valve assemblies are assembled externally and then placed into the housing. The valve housing is closed by end flange that is bolted to the housing and sealed with indium for the helium-vacuum seal. The end flange has the 0.375-inch port for the tube that connects to the external components of the flow loop. The valves are at the maximum possible distance apart, consistent with a valve port diameter of 0.375-inch for the gas flow. The valve center lines are 1.875 inches apart and the valve housing has an internal diameter of 1.430-inches. The valve housing is assembled by first welding an attaching tube to the cylinder head over the valve port. Then the other parts of the housing are sequentially welded on. The internal components are then manufactured, welded together and fully assembled as a separate valve assembly. Finally, the valve assembly is fit into the housing. The valve assembly is sealed against the lower flange of the housing by the force of cup-washer springs that are compressed by the housing flange bolts. Material and Dynamic Considerations. With the size and position of the valves set by the geometry of the cold cylinder, the valve assembly components are designed as follows. Three criteria are addressed in this design concept: 1) material limitations, 2) low-temperature valve leakage and 3) dynamic constraints that arise from the operation of the G-M cycle. Since these valves operate at low temperatures, material wear is critical. To maximize the life of the valves, this design virtually eliminates sliding components and stress concentrations. This is accomplished by using multi-leaf flexures to guide the motion of the single moving core. This single moving core includes the valve stem, valve head and moving electromagnetic core. Unlike a typical solenoid valve design, this valve is not allowed to hit the valve stops at full force. This is accomplished by the dynamic control of the valve motion through prescribed electrical input to the coil so that impact forces are minimized.

Electromagnetic Design and Analysis. To open the valve, enough magnetic force must be created to overcome the pressure force on the valve head and the initial force in the valve spring. However, once the valve begins to move, the required force to continue valve motion is substantially reduced. Only enough magnetic force to hold the valve in the open position against the valve spring is required. The magnetic force should then be further reduced to allow the valve spring to gently reseat the valve in the closed position. Therefore, it is desired to first build up enough electromagnetic energy in the coils to overcome the opening force. Once motion starts, the magnetic energy in the coils should be reduced twice to hold the valve open to return the valve to the closed position. Controlling the amount of electrical input to the coils will accomplish this. In a typical solenoid valve, enough electromagnetic energy is delivered to coils to open the valve. However, this energy is not reduced and the valve is allowed to hit the stops with full force at a high velocity. To return the valve to its closed position, a solenoid design allows the electromagnetic energy to be dissipated in the coils. This seats the valve at high speed using the full force of the spring. Conventional solenoid-valve designs result in large impact forces and stresses, valve bounce and energy dissipation that reduces cooling.

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399

In the cold EM valve design, the electromagnetic energy in the coils is controlled by the electrical input to the valves. Enough electrical power is supplied to the magnetic coils to open the

valves. The valve will initially open at a high velocity. The energy in the coils is then reduced through a room temperature resistor to provide enough force to maintain the valves in the open position. The valve spring now acts to slow down the valve motion and the valve will reach its full open position at low velocity. Reducing the energy in the coils again to allows the valve spring to

begin closing the valve. This electromagnetic force retards the closing motion of the valve and the valve reseats with minimum velocity. The velocity profile of the cold EM valve will be similar to the velocity profiles produced in cam driven designs. However, unlike cam driven designs, the cold EM valve design has no frictional components. From initial calculations, the force required to open the valve is 53 lbf and the force required to hold the valve open is 13 lbf. From this information, a differential analysis of the electromagnetic design was performed using the Runge-Kutta method. It was found that the magnitude of the power

dissipation in the valves for the opening and closing motions is on the order of 10 millliwatts. These calculations were performed using tabulated room temperature properties of pure copper. As temperature decreases, the resistance in the windings also decreases; thus less electrical power is required to produce a magnetic field in the ferromagnetic core. Copper wire of 99.95% purity will be used for this design. At very low temperatures, the resistance of this copper wire will decrease to a low value. This resistance will be experimentally determined at a later date.

Flexure Design. In order to achieve a design with no frictional components, the valve stem must remain centered in the hole in the back iron throughout the travel of the valve. Additionally, the valve guides should not resist the axial motion of the valve. Thin multi-leaf flexure stacks are

used to accomplish this. These multi-leaf flexures are designed to have a high radial stiffness with a low axial stiffness.

Axial stiffness and stress concentrations are minimized by making each flexure 0.010-inch thick. A single flexure is designed with as a 300° circular beam spanning between the moving valve stem and the stationary housing. The circular beam is subject to torsional and bending stresses.

Each flexure is manufactured of a single sheet of spring steel and the curved beam is formed by the milling of the required pattern into the steel sheet. Stress concentrations are minimized by

maintaining a corner radius 3/64-inch. Each three-leaf flexure consists of three single flexures stacked with a 120° relative rotation and with 0.035-inch spacers between the single flexures. Three-leaf flexure stacks are located above and below the moving ferromagnetic core. The required radial stiffness of the flexure must be sufficient to overcome the instability from the radial magnetic force exerted on the stem at its maximum off-center position. The valve stem

has a standard sliding fit clearance in the hole in the back core of 0.003-in on the radius. An electromagnetic flux exists in this gap. The flexures stiffness was selected so that at a 0.003-inch radial deflection, the net force will return the valve to a centered position. At a 0.003-inch radial deflection the off-center magnetic force is 9.4 lbf. Each multi-leaf flexure stack has a stiffness that exceeds this value.

Valve Seat Design. The head of the cold EM valve is designed with a Kel-F insert to form the valve seat. Miller and Brisson investigated the behavior of Kel-F valve seats at room temperature

and at low temperatures. Their data shows that the required seating force necessary so that the creep of the Kel-F at room temperature shapes the valve to the valve seat and provides a tight low temperature seal. At room temperature with no electrical input to the electromagnetic coils, the cold EM valves are closed with sufficient force so that the Kel-F seat deforms to the shape of the seating surface. At low temperature, this deformation provides a tight seal for the EM valves with leak rates well below tolerable limits [4]. CONCLUSIONS

From an initial, first-order, thermodynamic analysis, it has been determined that the addition of an external flow loop to a G-M cycle will not significantly degrade the performance of the cryocooler. With the cooling load the same as that for an unmodified G-M cycle, the external flow

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loop increases the refrigeration temperature by 5 K. This temperature increase may be attributed to the second-order affects that arise with the introduction of the external flow loop.

The cold EM valves are designed with good sealing properties and low frictional and impact wear. The valve seat is made of Kel-F to reduce leakage. The valves are guided by multi-leaf

flexures that maintain the valve stem in a centered position. Since the valve is designed to hit the stops and reseat with a negligible amount of force and momentum, the impact and valve bounce are reduced. This is accomplished by controlling the velocity profile of the valve through controlled electrical input to the magnetic coils. Preliminary calculations indicate that the dissipation of these valves is on the order of a 10 milliwatts. ACKNOWLEDGMENTS

We gratefully acknowledge the National Science Foundation and Massachusetts Institute of Technology for the financial support of this work.

We would also like to thank Leybold Cryogenics for the donation of the cryogenic equipment used for this analysis and experimentation. REFERENCES

1.

J.A. Crunkleton, “A new Configuration for Small-Capacity Liquid-Helium-Temperature Cryocoolers,” 7th International Cryocooler Conference Proceedings, Vols. 1-4, Santa Fe, NM, November 17-19, Air Force Phillips Laboratory Report PL-CP-93-1001, Kirtland Air Force Base, NM, 1993, pp. 187-196.

2. G.R. Gallagher and J.A. Crunkleton, “Thermodynamic Analysis of the Boreas Cryocooler,” Advances in Cryogenic Engineering, vol. 39, Plenum Press, New York (1994), pp 1543-1550. 3. R.E. Jones and J.L. Smith Jr., “Design and Testing of Experimental Free-Piston Cryogenic Expander,” presented at the 1999 Cryogenic Engineering Conference in Montreal, Canada (July 1999), to be published in Advances in Cryogenic Engineering, vol. 45.

4. F.K. Miller and J.G. Brisson, “Development of a low-dissipation valve for use in a cold-cycle dilution refrigerator.” Cryogenics, vol. 39, Elsevier Science, Oxford, UK (1999), pp. 859-863.

Optimum Intermediate Temperatures of Two-Stage Gifford-McMahon Type Coolers T. C. Chuang*, S. Yoshida** and T. H. K. Frederking

Cryogenics Laboratory, Chemical Eng. Dept. SEAS, UCLA Los Angeles CA 90095 * Raytheon Corp., Philadelphia, PA 19101

** Taiyo Toyo Sanso Co., Ltd., Kawasaki, Kanagawa, 210 Japan

ABSTRACT

Optimum intermediate temperatures are evaluated for two-stage Gifford-McMahon (GM) coolers including shield coolers protecting cryostats with low-boiling-point liquids, e.g. liquid Helium-4 (He I) baths near 4 K. Recent cooler progress has raised refrigeration loads to the 1-watt order of magnitude at 4 K. Thus, present three-stage systems with two-stage GM and single-stage Joule-Thomson (JT) expanders are potentially replaceable as two-stage GMs reach the refrigeration power of the three-stage GM/JT systems. We have extended earlier studies with the objective of evaluating realistic optimum intermediate temperatures, e.g. of two-stage GMs, in order to replace previous ideal limits. Our present extended model includes previous components such as the cascade-shunt model. Its optimum temperatures are found to be lower than ideal values, consistent with cooler data, in the range of parameters covered. This feature relates in part to higher loss parameters in the low temperature stage exceeding those of the elevated temperature stage. INTRODUCTION

Three-stage GM/JT systems have been known since the 1960’s. The two-stage GM cooler, combined with the low-temperature single JT-stage, has reached high reliability, e.g. in NASA’s deep space network.1,2 In more recent time, modifications of the three-stage systems have become known for the cooling of SQUIDS.3,4 In the 1990’s cooler progress has been remarkable permitting two-stage GM systems (and similar coolers) for refrigeration load on the order of magnitude of 1 watt at 4 K. In this context, the objective of the present studies is to determine the optimum intermediate temperature between two stages of the two-stage cooler. The objective function is minimization of the total power input to the cryocooler. The calculations aim at a generic approach, to the extent possible, using non-dimensional variables. We outline simplified early cooler models which have led to ideal optimum temperatures for restrictive assumptions. We relax these idealized constraints using an extended model based on our preceding work of Chen et al.3,6 In lieu of constant properties, realistic parameters are introduced; for instance, different Carnot fractions (CF) for the two stages. Mostly, the low temperature stage has smaller CF-values than the moderate temperature stage. Finally, we discuss model results, compare them with cooler data and present conclusions. Cryocoolers 11, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2001

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Figure 1. Chen et al. cooler model, shown schematically in temperature-entropy diagram.

SIMPLIFIED SMALL CRYOCOOLER MODELS

Aside from extensive simulation codes, there have been relatively few closed-form models for small cryocoolers. We outline a few examples of simplified cooler modeling. Chen et al. model. W.E.W. Chen et al.5,6 have considered a “dynamic insulation” model for multi-stage coolers. Figure 1 is a schematic diagram of the two-stage case in the temperature (T)-

entropy (S) plane. The main assumptions of the model have been introduced for the following scenario. Cooler losses are represented by parasitic heat inputs to each stage. For a stage “j”, the parasitic input is intercepted at the lowest temperature of the stage and “pumped” up by power input The total power input is minimized varying the intermediate temperature be-

tween stages for a specified set of the high temperature and the low temperature (cold box temperature For constant parameters, the model for two stages has an optimum ideal temperature equal to the geometric mean of and : The “high” temperature usually is the environs temperature

i.e.

The optimum temperature ratio of a multi-stage

system with “n” stages is given by

In Fig. 1 contact with a working fluid is indicated, e.g. operation along an isobar. This case leads to simple constraints for constant fluid properties, e.g. constant specific heat. The assumption of Carnot cooler operation is relaxed readily assuming constant Carnot fractions for each stage. The optimum result, equation (1) is not altered. For a stage, we have a power input of is the Carnot power input to stage j.

Jeong-Smith cascade model.7 Figure 2 is a schematic diagram of the model’s basic unit. A stage in the cascade system7 absorbs its refrigeration load at its low temperature “bus”. Upon power input into the cooler stage, the Q-rate is “amplified”. The amplified rate is rejected to the next “high-T” stage. Counting from the bottom up (instead of from the top down), the heat leaving a

stage is For constant parameters and constant Carnot fractions, the total entropy generation rate of the multi-stage system is minimized readily. This leads to the optimum T-ratio result, equation (1). A specific example referred to in reference 7 is the thermo-electric cooler. For the simplified system, e.g. two-stage system, the postulate of minimum entropy generation rate is equivalent to the postulate of total power input used in the Chen et al. model.5,6

OPTIMUM INTERMEDIATE TEMPERATURES OF 2-STAGE G-M

Figure 2. Jeong-Smith

cascade model, schematically.

403

Figure 3. Bejan shunt model,

schematically.

Bejan shunt model.8,9 Figure 3 shows the shunt model8,9 schematically. For small single-stage coolers, the sum of the thermodynamic losses in the cooler is a multiple of the refrigeration load The Bejan shunt model represents this aspect in a simplified manner. The cooler losses are lumped together by the “shunt” carrying an entropy-generating heat flow rate. In other words, all losses are considered to be represented by the parasitic rate of the shunt. The total refrigeration load is the sum The resulting Carnot fraction is

When is zero, the ideal Carnot process limit is reached. In equation (2) the losses are represented by the ratio Losses are found to be high for small cryocoolers reaching ratios of up to six and more. Bejan8,9 has used a convenient non-dimensional form of the shunt’s thermal

conductance. He normalized it using a reference entropy generation rate at the warm end Small cooler performance results suggest that a more convenient normalization is based on the entropy generation rate at the cold end.10,11 Using the latter, we describe the shunt-related losses by one single loss parameter C of the cooler.19-12 We express the power input to the entire cooler in terms of its Carnot value

is the Carnot coefficient of performance. The Carnot power input is

The “cooler map”12, CF versus of various cooler types.

provides a first order of magnitude assessment of the regimes

For illustration, a recent very small cooler is selected. Kuo et al.13 describe a miniaturized Stirling cooler. The data set [78 K at 300 K at 0.3/14 watt/watt] leads to i.e. C significantly above unity. This C-constraint for the mini-cooler supports the Chen et al. model5,6 (omitting ) in so far as the equivalent parasitic of the shunt is In the present extended model, we leave the minicooler range, relax previous conditions and incorporate the (modified) Bejan shunt model, [Eq. (3)], into an extended Chen et al. model.

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GM REFRIGERATOR DEVELOPMENTS

Figure 4. Extended model with ingredients of the Chen et al., shunt-cascade models, schematically.

EXTENDED MODEL

Extended model characteristics. Figure 4 is a schematic drawing of the present system. Ingredients of the three preceding models are utilized relaxing the ideal nature of the Chen et al. model.5,6 Carnot fractions of the stages are not equal, noting that the loss parameters are different

as well. This is expected to shift the optimum temperature at the intermediate “bus” away from the ideal system temperature In comparison to the Chen et al. model,5,6 we have more efficient GM stages than in the small coolers considered initially in the early modeling.5,6 Stage loss parameters are close to the order of

magnitude unity or below. Further, the cold box load is incorporated readily. The scenario chosen for the calculations is zero externally applied refrigeration load on the intermediate “bus” at (Fig. 4). The total power input to the two stages is

The total refrigeration load at the low-T stage (stage 1) is the sum of the cold box refrigeration load and the parasitic load or

The related power input into the first stage is

or

The refrigeration load imposed on the upper cooler stage 2 is expressed by the cascade condition. “Amplification” of via -input provides the load imposed on stage 2.

405

OPTIMUM INTERMEDIATE TEMPERATURES OF 2-STAGE G-M

The entire refrigeration load on stage 2 is expressed using the stage loss parameter This ad-hoc condition is formulated as an analog of equation (6), e.g. In this context we note that there is a lack of loss parameter information in the literature on GM - stages. Accordingly,

the power input to stage 2 is written as

Because of coupling condition (8) we arrive at two power input contributions to stage 2, with

and

and

Orders of magnitude. We assess orders of magnitude for the low-T asymptotic conditions and A simplified (truncated) set of power input contributions is

The second term on the right hand side is independent of

i.e.

We vary

in

order to obtain a minimum of the power input. This order of magnitude estimate leads to Thus, our qualitative result implies a lowering of below the ideal value as the loss parameter becomes larger than The related total power input turns out to be on the order of magnitude Therefore, in the subsequent numerical calculations we use a non-dimensional total power input as follows: The y-values are in the order of magnitude range from 0.1 to unity. In the model calculations the following set is selected: In the numerical parameter range selected for the subsequent modeling, we have increasing roughly proportional to

However, the input

dition implies that the power ratio

is a relatively weak function of

is a distinctly decreasing function of

This con-

Because of

other two-stage cooler constraints, it is necessary to have a power ratio on the order of magnitude of ten, e.g. values of 20 or higher.22 Sample calculation. For illustration, the following cooler stage loss parameters are selected: at a temperature ratio The low-T stage has a (normalized) power input of The power contributions to the upper stage are , in normalized form: where The second term is Thus, the upper stage receives The final nondimensional total power input is The stage power ratio turns out to be for the present numerical set. This is consistent with the preceding order of magnitude discussion. Values of leading to low power ratios are physically not meaningful. Numerical results. Results for the set [4.2 K at 300 K] are presented in Figs. 5 and 6. The nondimensional total power (y) is displayed as a function of Figure 5 is based on the condition of equal loss parameters in both stages, The curves for constant C-values show a distinct minimum. For the smallest C of 0.45, the location of the minimum total power input is fairly close to the ideal value Trends shown for larger C-values beyond 0.45 however are at variance with the asymptotic order of magnitude estimate discussed above. Further, the related stage power ratio tends to drop toward values below 10. Therefore, Fig. 5 suggests that the assumption is not realistic for high C-values. Figure 6 presents curves for loss parameter sets The moderate-T stage is characterized by a constant loss parameter kept at The parameter of the low-T stage is increased. For a small rise in the minimum intermediate temperature is reduced but little. For a larger mismatch in C-values however the minimum moves toward low As is raised beyond unity, e.g. 1.2, the minimum is too low. It cannot be reached with the capability of the upper stage. For the range of loss parameters of small coolers, there is no chance to reach 4 K with

406

GM REFRIGERATOR DEVELOPMENTS

Figure 5. Total power input (normalized) versus for equal loss parameters

Figure 6. Total power input (normalized) versus for stage loss parameters

two stages. Instead, multi-stage cooler systems are needed. The latter are contrary to the usual rationale applied to small cryocoolers. From Fig. 6 we conclude that there is a relatively narrow range of loss parameter ratios 1, not too far above unity, for favorable two-stage system operation. Figure 6 suggests that running costs require C1-values not exceeding unity. The related optimum values are lower than the ideal value All of these constraints pertain to the set of our present assumptions. COMPARISON WITH COOLER DATA

Table 1 presents GM-cooler parameters which reflect the two-stage system progress of the 1990’s. The data are selected for the limit of zero externally applied refrigeration load at the intermediate “bus” (Fig. 4). The cooler system’s low temperatures are chosen toward the low-T end of the versus plot for various loads. The choice does not necessarily imply that the -value coincides with the highest CF location. We select these figures in order to see whether our model is consistent with GM cooler data. Mostly the cooler tests display rather steep characteristics as a function of (for specific load conditions): is changed but little for the specified intermediate external load. Because of the steep characteristics the entries in Table I are considered representative of for the load conditions presumed. The cold box T L-values have decreased with time in the 1990’s, and accordingly also and the -values. All the -values, i. e. -values for our load conditions, are below the ideal optimum temperatures of constant property models. This feature agrees with our extended model. Further, it supports Carnot fraction behavior: the -values of the low-T stage are smaller than the of the moderate-T stage. In this context we note further CF-values of reference 12, and recent single-stage GM data of reference 19. For background information we refer to references 4 and 20.

Additional cooler information may be found in the manufacturer compilation of reference 21.

OPTIMUM INTERMEDIATE TEMPERATURES OF 2-STAGE G-M

407

CONCLUSIONS For the cold box near 4 K, both the present extended model and the cooler data support stage loss parameters given by the inequality (for the parameter range considered). The low temperature stage has a larger relative thermodynamic loss than the moderate temperature stage. This implies a Carnot fraction of the low-T stage smaller than the of the moderate-T stage. Consequently, the optimum intermediate temperature is below the ideal value Less stringent requirements are imposed when the cold box temperature is raised to 15 K or 20 K, e.g. for high-temperature superconductor magnets. The minimum power input function is lowered, and Carnot fractions of the stages are raised.

ACKNOWLEDGMENT The senior author (thkf) acknowledges with appreciation the hospitality of Professor K. Andres, Director, Walther-Meissner-Institut (WMI), Garching FRG during a brief stay (Summer 1998 / Fall Quarter 1998). Dr. Uhlig and Mr. Hehn of WMI provided useful input in several interesting discussions on coolers. Dr. Sidney Yuan’s help has been essential. It is appreciated very much.

NOMENCLATURE C CF COP

loss parameter of cooler, C of cooler stage “j” Carnot fraction = for stage “j”

coefficient of performance Carnot coefficient of performance

s T W

y

refrigeration load, stage “j” refrigeration load imposed on cold box parasitic heat flow rate through shunt representing stage losses entropy temperature; temperature at stage cold box temperature intermediate temperature of two-stage cooler power input; for stage Carnot value of power input non-dimensional power input temperature ratio

Subscripts 1 stage 1 2 stage 2 Carnot value c ideal ideal value, e.g. E environs, e.g. high temperature value, mostly H L low temperature value m intermediate stage tot total * optimum value at minimum power input REFERENCES 1. 2.

Higa, W. H. and Wiebe, E., “One Million Hours at 4.5 K,” Proc. 1977 International Cryocooler Conf., Boulder, CO., NBS Special Publication 508 (1978), pp. 99-107. Brithcliffe, M., “A 2-Watt 5-Kelvin Closed-Cycle Refrigerator System for Micro Wave Low-Noise Amplifiers,” Proc. 7th Intersoc. Cryog. Sympos., 1989, ASME, New York, pp. 131-134.

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GM REFRIGERATOR DEVELOPMENTS

3. Buchanan, D. S., Paulson, D. N., Klemic, D. A. and Williamson, S. J., “Development of a Hybrid Gifford-McMahon Joule-Thomson-Based Neuromagnetometer Cryosquid,” Proc. International Cryocooler Confernece, Monterey CA, 1989, pp. 35-46. 4. Yoshida S., Umeno, T. and Kamioka, Y., “Development of a Flexibly Separated 4 K Head Refrigerator For a SQUID System,” Advances in Cryogenic. Engineering, vol. 39, Plenum Press, New York (1994), pp. 1263-1270.

5. Chen, W. E. W., Turner, J. R. and Frederking, T. H. K., “Thermodynmamic Performance Measure 6. 7.

8. 9. 10.

11. 12.

13.

14. 15.

16. 17. 18.

for Cryogenic Vessel Insulation,” Proc. 6th Intersoc. Cryog. Sympos., AIChE Symposium Ser.. 251, vol. 82, New York (1986), pp. 101-103. Chen, W.E.W., M.S. thesis, University of California, Los Angeles, 1987. Jeong S. and Smith J.L., Jr., “Optimum Temperature Staging of Cryogenic Refrigeration System,” Cryogenics 34 (1994), pp. 929-933. Bejan, A., Advanced Engineering Thermodynamics, Wiley, New York, 1985. Bejan, A., Thermal Science Seminar, UCLA, Academic Year 1989-1990. Ravikumar, K. V., Yoshida S. and Frederking, T.H.K., “Comparison of Performance Measures for Cryocoolers and Refrigerators,” Intersoc. Cryog. Sympos., Houston, TX, March 1995, paper 33b, unpublished. Pinsky, C. et al., “Evaluation of the Thermodynamic Performance of Pulse Tubes,” Advances in Cryogenic Engineering, vol. 41 (1996), pp. 1365-1372. Rohlin, L. et al., “Comparison Studies of Thermodynamic Losses in Pulse Tube Cooler Components,” Proc. ICEC 17, 17* Intern. Cryog. Eng. Conf. Bournemouth UK, IoP Institute of Physics Publishing, Bristol, Philadelphia (1998), pp. 89-92. Kuo, D. T., Loc, A. S. and Yuan, S. W. K., “Qualification of the BEI B512 Cooler, Part 1 Environmental Tests,” Cryocoolers 10, ed. R. G. Ross, Jr., Kluwer Academic/Plenum Publishers, New York (1999), pp. 105-115. Dr. Peter Kerney (1995), private communication on Cryodyne Model 1020. Chafe, J., Green, G. and Riedy, R. C., “Neodymium Regenerator Test Results in a Standard GiffordMcMahon Refrigerator,” 7th International Cryocooler Conference Proceedings, Air Force Phillips Laboratory Report PL-CP—93-1001, Kirtland Air Force Base, NM, April 1993, pp. 1157-1164. Uhlig, K. and Hehn, W., “3He/4He Dilution Refrigerator Precooled by Gifford-McMahon Refrigerator,” Cryogenics, 37 (1997), pp. 279-282.. Uhlig, K. and Hehn, W., private communication (1998). Yoshida, S., private communication on the Sumitomo GM [1 watt at 4 K] characteristics (2000).

19. Wang, C. and Gifford, P. E., “High Efficient Single-Stage GM Cryorefrigerator for Temperatures around 20-40 K,” Cryocoolers 11, ed. R. G. Ross, Jr., Kluwer Academic/Plenum Publishers, New

York (2001). 20. Chuang, C., Ph.D. thesis, University of California, Los Angeles, 1981.

21. Walker, G. amd Bingham, E. R., Low-Capacity Cryogenic Refrigeration, Oxford, Clarendon (1994), pp. 283-287. 22. Dr. T. Kuriyama, private communication at poster session of ICC 11.

Regenerator Behavior with Heat Input or Removal at Intermediate Temperatures Ray Radebaugh, E. D. Marquardt, J. Gary, and A. O’Gallagher National Institute of Standards and Technology Boulder, CO 80303

ABSTRACT Regenerators with finite losses are capable of absorbing a limited amount of heat at intermediate temperatures along their length. This paper discusses a simple analytical model and a rigorous numerical model of regenerator behavior under the influence of heat input or heat removal at intermediate temperatures as well as the influence of a steady mass flow superimposed on the oscillating mass flow within the regenerator. The finite time-averaged enthalpy transport through the regenerator undergoes a discontinuity at the location of the heat input to satisfy the First Law of Thermodynamics. The discontinuous enthalpy flow leads to a discontinuous temperature gradient in the axial direction and to an increase in the regenerator loss that must be absorbed at the cold end. However, the increased loss is less than the heat input at the intermediate temperature, which allows the regenerator to provide a certain amount of cooling without the need for a separate expansion stage. This phenomenon is particularly useful for shield cooling and for precooling a gas continuously or at discrete regenerator locations prior to liquefaction at the cold end. For continuous precooling the total heat load can be reduced by as much as 23%. A comparison is made of the system performance with and without intermediate heat input under various conditions. The paper presents design guidelines to determine the amount of heat a regenerator is capable of absorbing at various temperatures. Methods for optimizing the location of discrete heat inputs are presented. The analytical and numerical models are in very good agreement with each other and are consistent with very limited experimental data. INTRODUCTION The function of a regenerator in cryocoolers is to transfer heat from an incoming, highpressure stream to an outgoing, low-pressure stream, just as in a recuperative heat exchanger. The only difference is that in regenerators the heat extracted from the incoming stream is stored temporarily in the heat capacity of the matrix before transferring it to the outgoing stream a short time later. Even though there are discontinuities in the flow in regenerators, there are no fundamental differences between regenerative and recuperative heat exchangers when only time-averaged behavior is considered. The finite heat capacity of a regenerator results in a degradation of its performance, but there is no difference between the two heat Cryocoolers 11, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2001

409

410

REGENERATOR ANALYSIS AND MATERIALS DEVELOPMENTS

exchangers in regard to their behavior under the influence of steady external factors. For example, the effect of a steady heat input along the length of a heat exchanger or a superimposed steady flow of fluid in one direction is the same for both recuperative and regenerative

heat exchangers. Therefore, the arguments being presented here will apply to both types of heat exchangers, but the calculations performed here were carried out only with regenerators. The second law of thermodynamics shows that it is always desirable to remove heat in a refrigerator at the highest possible temperature since the increase in entropy flow carried by the refrigerant to the warm end is given by where is the change in entropy flow, is the heat input, and T is the temperature from which the heat is being input. In the case of liquefying a gas it is desirable to have many stages of cooling to remove the sensible heat from the gas and to cool it to the liquefaction temperature. The last and coldest stage only removes the heat of vaporization. In practice the additional

stages may lead to a system that is too complex and costly. The simplest case is a one-stage refrigerator used to liquefy a gas like nitrogen or oxygen. For nitrogen at atmospheric pressure

the specific enthalpy change from 300 K to the saturated vapor phase at 77 K is 234.0 J/g and the enthalpy change during liquefaction at 77 K is 199.2 J/g. Of the total heat that must be removed from the nitrogen, 54% should be removed at temperatures between 300 and 77 K. If this

sensible heat is removed only at the single stage operating at 77 K, then system efficiency is decreased. But, with a single-stage refrigerator other heat sinks to remove some of the sensible heat at a higher temperature and improve the system efficiency are not normally available. We propose here that the heat exchanger (either recuperative or regenerative) can be used to remove a portion of the sensible heat. This concept works only with non-ideal heat exchangers as will be shown in the next section. A perfect heat exchanger cannot absorb heat along its length. In the next sections we analyze the behavior of heat exchangers in three different cases that involve heat input (or removal) to the heat exchanger along its length. In case 1 a fixed amount of heat is input at a specific location along the length of the heat exchanger. In case 2 the heat input is proportional to the temperature change, such as with the precooling of a gas. In this case a portion of the total heat input is at some fixed location along the length of the heat exchanger and the remainder is at the cold end. Case 3 is like the previous case except that the heat is continuously removed all along the length of the heat exchanger. Case 3 also applies to the situation of a superimposed steady flow of refrigerant through the heat exchanger. In the case of regenerators this applies to the DC flow superimposed on the oscillating flow. The analysis

given here applies to either sign of heat flow or to either direction of steady flow. For instance

the analysis applies to the cold finger heat interceptor discussed by Johnson and Ross1 where heat was removed at some location along the regenerator to increase the refrigeration power at the cold end. However, the emphasis in this paper is for using the regenerator or recuperator to absorb heat along its length in a manner to increase the system efficiency. CASE 1, FIXED HEAT INPUT AT ONE LOCATION

Simple Analytical Model For cryocooler operation above about 20 K the temperature profile in the regenerator is very near linear and the regenerator loss or energy flow (time-averaged enthalpy flow plus conduction) is approximately proportional to the temperature gradient. The contribution to the regenerator loss due to the compression and expansion of the gas in the void space is small for this temperature range. Therefore, for a simple model we assume the regenerator loss or energy flow is proportional to the temperature gradient, as given by

REGENERATOR BEHAVIOR WITH HEAT INPUT

411

where

is the time-averaged enthalpy flow in the regenerator (ignoring real gas effects), is the conduction in the regenerator, is the regenerator loss or energy flow with no heat applied to the regenerator, L is the regenerator length, is the temperature of the hot end, and Tc is the temperature at the cold end. Heat input along the regenerator length increases the temperature at that location so that the temperature gradient in the regenerator before and after

this location is altered as shown in Fig. 1. The increased gradient at the cold end causes an increase in the regenerator loss to the cold end. The main question is whether this increased loss is less than the heat input or whether the heat input simply ends up in the cold end with no attenuation. Applying the first law of thermodynamics along the length of the regenerator shows that the new regenerator energy flow to the cold end (region 2 in Fig. 1) is given by

where is the heat input at some intermediate location xi. To solve Eq. (3) for the new regenerator energy flow to the cold end, we begin by introducing the following dimensionless variables:

Figure 1 shows the use of dimensionless position and temperature. By using Eq. (2) and these dimensionless variables, the energy balance equation Eq. (3) can be rewritten as

where the last term represents the dimensionless temperature gradient in region 1. From Fig. 1

we see that the two dimensionless temperature gradients are given by

Figure 1. Diagram showing location of heat input and its effect on regenerator temperature and energy flow.

412

where

REGENERATOR ANALYSIS AND MATERIALS DEVELOPMENTS

is the dimensionless location for the heat input and is the dimensionless By substituting Eq. (7) into Eq. (6) and

temperature of the regenerator at that location. combining that with Eq. (5) yields

Solving Eq. (8) for qreg gives the result

This equation shows that the heat input at xi is attenuated by

the cold end unless

and not all of the heat reaches

The ratio of the total heat load on the cold end to the heat load if the

heat were input at the cold end (heat load ratio) is given by

The dimensionless temperature at the location Eqs. (7) and (9) to give

of the heat input qi is found by combining

Figure 2 shows the locus of points for several values of qi and Fig. 3 shows the variation of qr with for various qi from Eq. (10). We note from Fig. 2 that qi values much above 1 give rise to significant heating of the regenerator mid section. Thus, as a general statement we can say that the regenerator can be used to beneficially absorb heat along its length

Figure 2. Dimensionless regenerator temperature at location of discrete heat input.

Figure 3. Heat load ratio at cold end for

various heat inputs at discrete locations.

REGENERATOR BEHAVIOR WITH HEAT INPUT

413

only for heat inputs not much larger than the original regenerator loss. An ideal regenerator with zero loss cannot absorb any heat along its length and the temperature at would approach infinity since would be infinity in Eq. (11). In simple terms, the regenerator loss in an ideal regenerator remains zero even as the temperature gradient approaches infinity. A comparison of

the calculated behavior with experiment requires that

be known in order to find

from a

known Typically, for 80 K cryocoolers, is comparable to the net refrigeration power, whereas for a 60 K cryocooler may be 50% larger than the net refrigeration power.

Numerical Model The numerical model used here for more accurate calculations is known as REGEN3.22. It is an update of REGEN3.13, 4, which is a finite difference program using the conservation of energy, mass, and momentum equations to describe the behavior of regenerators. One of the additions in this new program is the ability to add or subtract heat at any location along the regenerator and to allow for a DC flow in either direction. The baseline case used for the 5 calculations here was an optimized design similar to that for a pulse tube oxygen liquefier. The hot and cold temperatures were 300 and 90 K. The length of the regenerator was 40 mm and it was divided into 40 cells (41 mesh points) for these calculations. The baseline regenerator loss (RGLOSS + HTFLUX in REGEN3.2) was 8.19 W. The RGLOSS term in REGEN3.2 is the total enthalpy flux minus the enthalpy flux caused by pressure changes (real gas effects). The HTFLUX term is the matrix conduction. Figure 4 compares the temperature profiles calculated from REGEN3.2 with those from the analytical model. For zero heat input there is only a slight deviation from linearity in the profile calculated by REGEN3.2. The midpoint dimensionless temperature is 0.533 compared with 0.500 for a linear profile. Heat inputs or removal were at cell midpoints and occurred at and 0.49. The results in Fig. 4 show that adding has a greater effect on in the numerical model than it does in the analytical model. For heat removal the two models agree very well. Figure 5 compares the dimensionless regenerator loss calculated from the numerical model with that from the analytical model. This figure shows that the effect of on is nearly the same from the two models for heat input. The largest difference in the values for is 0.07. For heat removal the largest difference is 0.27, which occurs with at the regenerator midpoint. The temperature gradient at the cold end under those conditions is nearly zero, and the regenerator loss would be that caused by the compression and expansion in the void space. The temperature profile for heat input calculated by the analytical model would be in better

Figure 4. Dimensionless temperature profile in regenerator for various heat inputs at

two different locations.

Figure 5. Dimensionless regenerator loss as a function of heat input at two different locations.

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REGENERATOR ANALYSIS AND MATERIALS DEVELOPMENTS

agreement with the numerical model if the compression and expansion in the void space of the regenerator were accounted for in the analytical model. That contribution is independent of the temperature gradient, which would then indicate that a constant should be added to the expressions for the regenerator loss in Eqs. (2) and (4). The dimensionless regenerator loss at any location then would be expressed as

where a + b = 1. Equation 5 would then be changed to

By using the modified approximation to the regenerator loss given by Eq. (4*) we find that Eqs. (9) and (10) do not change, but Eq. (11) becomes

The numerical results at xi = 0.49 show that

when qi = 1.50. Thus, a = 0.77 and b = 0.23 give perfect agreement with the numerical model for at that value of qi and deviates only by 0.03 at qi = 0. At qi = -2.0 for the midpoint we found that qreg = 0.27 when the temperature gradient at the cold end went to zero. That value agrees well with b = 0.23. However, the results for from the numerical model for heat removal agree very well with the analytical model when b = 0. To maintain simplicity in the remaining sections, we will consider only the case of

a = 1.0 and b = 0. Such an approximation is reasonably good for a first stage regenerator, but a second stage regenerator will have a rather large b and require the use of the modified analytical model. For temperatures below about 20 K the regenerator temperature profile with zero heat input to the regenerator begins to deviate considerably from linearity with a dimensionless temperature at the center point being much less than 0.5. Thus, we do not expect that the analytical model would be useful for temperature below about 20 K. Figures 4 and 5 show that when heat is removed at to a heat sink at about 180 K the regenerator loss can be reduced by 36%. For the case of regenerator loss being

comparable to the net refrigeration power, the heat intercept leads to an increase in net refrigeration of 36% for the same power input or a reduction in input power by 36% for the same net refrigeration power. These calculated improvements are consistent with that found experimentally by Johnson and Ross1 with a Stirling cryocooler when a heat interceptor at 150 K increased the net cooling power at 60 K by 30%. CASE 2, TEMPERATURE DEPENDENT HEAT INPUT

We now consider the case where the regenerator is used to precool a gas. The amount of heat flow at a particular location along the regenerator depends on the temperature difference between the hot end temperature and the temperature at the specified location on the regenerator. The gas is cooled the rest of the way to the cold temperature by the refrigeration power available at the cold end. In this case we still consider only one heat sink located along the regenerator. The following section deals with a continuous heat transfer along the entire length. The total heat that must be removed by the regenerator and the cold end for this case is given by where

of

for the case of fluid flow at a mass flow rate of

The heat flow into the regenerator at the temperature

and the heat flow into the cold end becomes

is

with a constant specific heat

REGENERATOR BEHAVIOR WITH HEAT INPUT

415

The sum of the heat flows into the cold end when the regenerator loss is considered is

which, when normalized by the baseline regenerator loss

becomes

Because Ti is not known at this time, the only heat flow known at this time is from Eq. (12). When it is normalized by it becomes qt . The other two heat flows are related to qt in the following manner:

We now solve for

in terms of the known quantities

into Eq. (11) allows us to write

The solution for

at any

and qt . Substituting qi from Eq. (17)

as

for various qt becomes

The sum of heat flows to the cold end is found by substituting Eqs. (17) and (9) into Eq. (16), which yields The ratio of this heat flow to the heat flow at the cold end if all the heat was delivered to the cold end is

Figures 6 and 7 show how

and vary with for different values of qt from the simple analytical model. For all qt the optimum location for the heat intercept is at the midpoint of the regenerator. The minimum qr is 0.889, which occurs with qt = 2.0. The value of at the midpoint with this heat input is 0.667, which means that 1/3 of the heat is transferred to the regenerator and 2/3 is transferred to the cold end. Thus, the use of the regenerator to precool a gas can reduce the total heat load on the cold end by up to 11%. It should be pointed out that the baseline heat load considered here includes the baseline regenerator loss as well as the net heat load of cooling the gas from the warm temperature to the cold temperature. If the regenerator loss were equal to the net refrigeration power, then precooling at the midpoint would reduce the required net refrigeration by up to 22% at the optimum condition. CASE 3, CONTINUOUS HEAT TRANSFER (GAS PRECOOLING)

Analytical Model The most efficient use of the regenerator to precool a gas or a conduction member is with continuous heat transfer all along the length of the regenerator. The first law of thermodynamics

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REGENERATOR ANALYSIS AND MATERIALS DEVELOPMENTS

Figure 6. Dimensionless temperature at location of single heat sink for various total

heat inputs.

Figure 7. Heat load ratio at cold end versus heat sink location for various total heat inputs.

applied to an infinitesimally small element shows that the regenerator energy flow at any location, given by Eq. (2), undergoes a change given by

where is the heat flow caused by changing the temperature of a gas. In accordance with Eq. (12) this heat flow is represented by By using the linear assumption in Eq. (2) for the regenerator, Eq. (22) becomes

Equation (23) can be written in dimensionless quantities as

The solution to Eq. (24) gives the dimensionless temperature profile along the regenerator as

Figure 8 shows the dimensionless temperature profile for various qt. The dimensionless regenerator loss or energy flow at any location with this additional heat input to the regenerator according to Eq. (4) is given by

Figure 9 shows this regenerator energy flow along the regenerator for variousqt. The ratio of the

total heat flow to the cold end when using the regenerator for precooling to that without the precooling is given by

REGENERATOR BEHAVIOR WITH HEAT INPUT

Figure 8. Dimensionless temperature

profile in regenerator for continuous heat input or steady mass flow toward cold end.

417

Figure 9. Dimensionless regenerator energy flow versus position in regenerator for continuous heat input or steady mass flow.

A graph of qr is shown in Fig. 10 and is discussed in the following section where it is compared

with the results from the numerical model. The minimumqr of 0.770 occurs at qt = 1.80. Thus, the continuous heat transfer can reduce the total heat load (including baseline regenerator loss) on the cold end by a maximum of 23%. Numerical Model (with steady mass flow) The REGEN3.2 numerical model can be used to simulate continuous heat transfer to the regenerator by superimposing a steady or DC mass flow on the oscillating flow. When simulating the precooling of any gas other than helium, the steady mass flow rate of helium in REGEN3.2 needs to be adjusted to give the same enthalpy change between the hot and cold temperatures as for the gas to be precooled. We are assuming that the specific heats of both gases are independent of temperature. Starting with the same baseline case discussed for Case 1, we ran REGEN3.2 for three different steady mass flows which simulated total dimensionless heat flows from the hot to the cold end of qt = 0.5, 2.0, and 5.0. The calculated temperature profile for the case of qt = 2.0 is shown in Fig. 8. The numerical model shows slightly higher temperature sensitivity than does the analytical model, Eq. (25). As discussed previously, the difference can be explained by the neglect in the analytical model of a contribution due to

Figure 10. Heat flow ratio to cold end with heat transfer at the optimum discrete

location and with continuous heat transfer or steady mass flow.

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compression and expansion in the void space. The regenerator energy flow calculated with the numerical model for the case of qt = 2.0 is compared with that from the analytical model in Fig. 9. Results from the two models agree very well, particularly at the two ends of the regenerator. Values of much greater than about 2 or 3 would tend to reduce the net refrigeration power to zero for cryocoolers operating around 80 K. Figure 10 compares the heat flow ratio qr of Eqs. (10) and (27) for the analytical model with that obtained from REGEN3.2. This figure shows that the continuous heat transfer significantly reduces the cold end heat load compared with the case of heat transfer only at the midpoint. The numerical and analytical models agree well for both discrete and continuous heat inputs. COMPARISON WITH EXPERIMENT Gilman6 shows that with a heat intercept strap at a location of with the cooling power increased from 1.65 W to 2.40 W, or a 45% gain at 60 W input power. From Eq. (11) we find that qt = -1.56 and from Eq. (9) we find that or a 39% reduction in the regenerator loss. The radiator load from Gilman’s results was 3.1 W, which for a qi = -1.56 gives a baseline regenerator loss of 1.99W. A 39% reduction of this loss would cause a 0.78 W increase in refrigeration power compared with Gilman’s measured result of 0.75 W. Our numerical model would predict a gain of about 0.62 W and agrees to within about 17% of the experimental result. Our experimental results with liquefaction of nitrogen7 showed an increased liquefaction rate of 17% when the incoming gas was continuously precooled by the outer surface of the regenerator. The ratio of sensible heat of nitrogen to the heat of vaporization is 1.18. In this example the refrigeration power is comparable to the regenerator loss, and the calculated results from Fig. 10 show a predicted increased liquefaction rate of about 22%. CONCLUSIONS A simple analytical model has been developed that can be used to calculate the effect of heat input or removal along the length of a regenerator. The model assumes a linear temperature profile in the regenerator, which is a good approximation for temperatures down to about 20 K. The model is also extended to the case of continuous heat transfer all along the length of the regenerator and to a steady mass flow superimposed on the oscillating flow. The analytical model is in good agreement with our most recent numerical model, REGEN3.2. For continuous precooling in gas liquefaction, the heat load can be reduced by up to 23%. The calculated results are in reasonable agreement with experiments using heat interceptor straps and with those obtained in the liquefaction of nitrogen with a pulse tube refrigerator.

REFERENCES 1. Johnson, D. L., and Ross, R. G., Jr., “Cryocooler Coldfinger Heat Interceptor”, Cryocoolers 8, Plenum Press, New York (1995), pp. 709-717.

2. Gary, J., O’Gallagher, A., Radebaugh, R., and Marquardt, E., “REGEN3.2 Regenerator Model: User Manual”, NIST Technical Note, to be published. 3. Gary, J., Daney, D. E., and Radebaugh, R., “A Computational Model for a Regenerator”, Proc. Third Cryocooler Conf., NIST Special Publication 698 (1985) pp. 199-211. 4. Gary, J., and Radebaugh, R., “An Improved Model for Calculation of Regenerator Performance (REGEN3.1)”, Proc. Fourth Interagency Meeting on Cryocoolers, David Taylor Research Center Technical Report DTRC-91/003 January 1991, pp. 165-176. 5. Marquardt, E. D., and Radebaugh, R., “Pulse Tube Oxygen Liquefier”, Adv. Cryogenic Engineering, vol. 45, Plenum Press, New York (2000), in press. 6. Gilman, D. C., “Cryocooler Heat Interceptor Test for the SMTS Program”, Cryocoolers 9, Plenum Press, New York (1997), pp. 783 -793. 7. Marquardt, E. D., Radebaugh, R., and Peskin, A. P., Vapor Precooling in a Pulse Tube Liquefier,” Cryocoolers 11, Plenum Press, New York (2001).

Measurement of Heat Conduction through Metal Spheres M. A. Lewis and R. Radebaugh National Institute of Standards and Technology Boulder, Colorado, USA 80303

ABSTRACT This paper describes the results of the measurements of heat conduction through a bed of packed metal spheres. Spheres were packed in a fiberglass-epoxy cylinder, 24.4 mm in diameter and 55 mm in length. The cold end of the packed bed was cooled by a Gifford-McMahon (GM) cryocooler to cryogenic temperatures, while the hot end was maintained at room temperature. Heat conduction through the spheres was determined from the temperature gradient in a calibrated heat flow sensor mounted between the cold end of the packed bed and the GM cryocooler. The samples used for these experiments consisted of stainless steel spheres, lead spheres, and copper spheres. The spheres were screened to obtain a uniform diameter of 80 to 120 µm. Porosities of the packed

beds varied between 0.371 and 0.398. The measurements to determine the thermal conductance were carried out with various pressures of helium gas in the void space. The results indicated, as

expected, that the helium gas between each sphere enhances the heat conduction across the contacts between the individual spheres by several orders of magnitude compared with vacuum in the void

space. The conduction degradation factor, defined as the ratio of actual heat conduction to the heat conduction if the metal were in the form of a solid rod of the same metal cross-sectional area, was

about 0.11 for stainless steel, 0.08 for lead, and 0.02 for copper. The conduction degradation factor of 0.11 for stainless steel spheres agrees very well with the factor of 0.10 for stainless steel screen measured previously in the same apparatus. INTRODUCTION Beds of packed spheres are commonly used as a regenerator for cryocoolers operating at temperatures below about 80 K.1 Because of the large temperature gradient in the regenerator, heat conduction through the packed spheres can be a significant loss. Previous studies of heat flow through columns of packed spherical material have considered conduction in the fluid due to a temperature gradient, conduction within solid particles, conduction from one particle to the next through a separating film, heat transfer between particle and main body of the fluid, enthalpy carried along by the moving fluid, and possible heat generation through a chemical reaction.2 However, in cryocooler regenerators, the high thermal conductivity of the helium working fluid can transport a large fraction of the heat between the solid contacts.3 Schumann and Voss investigated experimentally the thermal conductivity of packed beds in static gas.4 Recent work by Slavin et al.5 considers the thermal conduction in packed beds of alumina spheres in static helium gas, but the temperature range is 100 to 500 degrees Celsius. Experimental and analytical research data for heat Cryocoolers 11, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2001

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Figure 1. Experimental test apparatus.

Figure 2. Experimental test section.

conduction loss are not available for practical use in the design of cryocooler regenerators. The

purpose of this study is to directly measure the heat conduction loss from room temperature to cryogenic temperatures through packed spheres in static helium gas. EXPERIMENTAL APPARATUS AND PROCEDURE

Figure 1 shows the experimental apparatus used for the present 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 and were described previously.3 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.6 Stainless steel, copper, and lead spheres were used in the study. The sphere materials were carefully screened to obtain diameters between 80 and 110 µm. The spheres were packed in the regenerator cylinder to a height of approximately 45 mm. 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 were connected to a cold plate, which was cooled by the GM cryocooler with the heat flow sensor between them. The hot ends of both regenerators were

capped by piston-shaped water jackets. Flowing water maintained the hot end temperature at room temperature. A bellows was attached to the lower water jacket. Changing the filling pressure of the helium in the bellows moves the lower water-cooled piston to apply any desired load. For these measurements, an applied load of 5.1 MPa on the packed sphere bed was used. The cold plate and the two regenerators were free to move with respect to the water jackets, so the force exerted by the bellows was applied equally to the two regenerator columns of packed spheres. The heat flow sensor was 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 oxygen-free copper with a cross-sectional area of 72 mm2 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.

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The experimental procedure was as follows. After pumping the vacuum vessel, the two-stage GM cryocooler was turned on. Both the cold plate and the heat flow sensor were cooled by the first stage of the GM cryocooler. The cold plate temperature was 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 could be varied over a wide temperature range, but for these tests was maintained at 80 ±0.05 K. Once the cold plate temperature was set and the temperature difference at the heat flow sensor was measured, an additional heat load was supplied to a heater mounted on the cold plate, and its effect on the temperature difference at the copper bar was 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 sphere beds, the two fiberglassepoxy 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

packed spheres was measured to provide information needed to determine the heat flow through the columns of sphere beds only. Characteristics of the spheres tested in this study are given in Table 1. The porosity for each

type sphere shown in this table is calculated using the density of the material and the actual mass of spherical material used to fill each regenerator column volume to capacity. To achieve the desired results, the regenerators along with material to be tested had to be assembled to minimize the amount of void space that was inherent to the G-10 cylinders and sphere material. The test sections needed to be packed efficiently to obtain low porosity values. The spherical material was poured into the regenerators and placed on a vibration apparatus that allowed the spheres to tightly pack together in a uniform distribution. As the material settled over the desired vibration time, the regenerators were filled to maximum capacity. This enabled us to reduce the void volumes in the G10 cylinders as much as possible. Once the regenerators were packed to the necessary uniform and solid packing, the G-10 regenerators were installed into the apparatus. Using this procedure, we

were able to obtain porosity values that were near the theoretical value expected for packed beds of spheres for a known volume and a specific material density.

EXPERIMENTAL RESULTS AND DISCUSSION The first measurement was performed to determine the system heat leak through the two regenerators without any spherical material packed inside and with vacuum pressure applied to the regenerator tubes. The cold plate temperature was 80 K and the hot plate temperature was 285 K. The total measured heat leak was 1.06 W, with 0.42 W calculated to be the heat conduction through the cylinder walls of the two regenerators. The 1.06 W of heat flow was then subtracted from

subsequent measurements with spheres in the regenerators and various helium pressures applied to obtain the heat conduction through the packed sphere beds.

The effect of helium gas pressure inside the regenerator on the total measured heat leak is shown in Figure 3. These measurements were made with the cold end of the apparatus maintained

at a constant temperature of 80 K. As this figure shows, the heat leak increases rapidly with increasing helium pressure until there is little significant change for pressures above 0.5 MPa. Fig-

ure 4 shows the pressure dependence of the mean free path and thermal conductivity of bulk helium gas at 80 K, 200 K, and 300 K. Although the thermal conductivity of bulk helium at pressures to

5 MPa is almost independent of pressure, the heat leak varies with pressure conditions at pressures

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REGENERATOR ANALYSIS AND MATERIALS DEVELOPMENTS

Figure 3. Heat leak vs. helium gas pressure.

Figure 4. Mean free path and thermal

conductivity vs. pressure dependency over temperature ranges.

below 0.5 MPa. In order to understand the test results, we discuss the heat transfer mechanism in

the regions of molecular flow, viscous flow, and intermediate flow. Generally gas flow can be treated as molecular flow when the mean free path, is larger than 10 times the distance, d, between the two solid plates which transfer the heat. On the other hand, when is less than 0.01 d, the gas flow can be treated as viscous flow. In the viscous flow region, heat transfer between two solid plates is proportional to the thermal conductivity and is independent of pressure for an ideal gas. In the molecular flow region, the heat transfer is proportional to the gas pressure. Figure 3 shows that at approximately 0.5 MPa the thermal conduction through the helium gas has reached the viscous flow region where the mean free path of the helium atoms has become less than 0.01 d. Because the temperature varies from 80 to 285 K along the length of the regenerator, the mean free path of helium at 0.5 MPa according to Figure 4 varies from 0.010 to 0.039 µm. Therefore, the effective distance between each spherical ball contributing most to the heat flow is about 1 to 4 µm, or 100 times the mean free path. Since the sphere is in the range of 80 to 110 µm, most of the heat is transported by the helium gas in a region very close to the individual contacts between spheres. According to Fig. 3, the heat leak at very low helium gas pressure approaches a value very near the background value. Such behavior indicates that there is very little electronic heat conduction associated with direct metallic contact between the spheres, in agreement with electrical resistance measurements.7 The electrical resistance measurements7 indicate a heat conduction of about 4 mW for our system. Because of the many spheres in a packed bed, the thermal conduction through the bed is reduced compared to a solid bar of the same material and same cross-sectional area as the metal in the sphere bed. Therefore, to estimate this heat leak through the packed spherical columns, a conduction degradation factor is applied. The actual conduction through a packed sphere bed is then given as a proportional reduction to the bulk material conduction as

where Heat flow through packed sphere bed Conductivity degradation factor Total cross-sectional area of regenerator Length of packed beds Porosity of packed beds k:

Temperature at cold end of regenerator Temperature at hot end of regenerator Thermal conductivity of regenerator matrix material

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Figure 5. Pressure (MPa) vs. CDF for packed bed spheres.

Figure 5 shows how the conductivity degradation factor (CDF) varies with increasing pressure within the regenerator for the stainless steel, lead, and copper sphere material. This value begins to level out at around a pressure of 0.5 MPa, which is consistent with the previous explanation of mean free path and thermal conductivity. The values of 0.11, 0.077, and 0.019 for the degradation factors of stainless steel, lead, and copper material, respectively, represent our data well over the pressure range above 0.5 MPa at a temperature of 80 K. Earlier experiments3 were performed in

this test apparatus using various screen materials as the regenerator matrix. Figure 6 shows the conductivity degradation factor for stainless steel 400-mesh, 25.4 µm wire; 325-mesh 27.9 µm wire; and 325-mesh, 22.9 µm wire over a wide porosity range. These porosities were obtained using a static load of 5.1 MPa in addition to no load conditions. These data indicate that a value of

0.11 is a good representation of the conductivity degradation factor for the stainless steel screen material and agrees with the value of 0.11 in the present investigation for stainless steel spheres. Our previous measurements3 of phosphor bronze screen gave a degradation factor of about 0.025 compared with the value of 0.02 for the copper spheres. The slightly lower value for copper is consistent with a lower for higher conductivity material. Table 2 gives the effective thermal conductivity integration between 80 and 300 K of various sphere and screen materials using Equation 1. Shown for comparison is the thermal conductivity of bulk helium gas.

Figure 6. Porosity vs. CDF for stainless steel screens.

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CONCLUSIONS

The heat conduction through packed spheres from room temperature to cryogenic temperature was measured experimentally. The experimental apparatus allows for a change in the regenerator material, the cold end temperature, and the helium gas pressure in the regenerator. The measurements were performed using stainless steel, lead, and copper spheres between 80 and 120 µm in size with a porosity of about 0.38. The experimental results showed that the helium gas between each sphere contact point plays an important role in transporting the heat. The heat conduction through the packed beds was enhanced by at least two orders of magnitude using helium gas compared to vacuum conditions inside the regenerator. The heat conduction reached a constant maximum value for helium pressures above 0.5 MPa. A short helium mean free path of 0.010 to 0.039 µm, indicates that most of the heat is transported a distance of the order of 3 µm from one sphere to the next. The conduction degradation factor, which is the ratio of actual heat conduction to the heat conduction where the regenerator material is assumed to be bulk, was about 0.11 for the stainless steel sphere materials, 0.077 for the lead sphere material, and 0.021 for the copper sphere material. This factor was relatively constant for the 80 K temperature at the cold end and for pressures over 0.5 MPa. For stainless steel and lead spheres the drop off below 0.5 MPa was much more significant than the copper spheres. This more constant value for the conductivity degradation factor for copper spheres could be attributed to possible deformation due to pressure applied during the experiments as well as a better heat transfer at the contacts This test apparatus provided NIST with valuable information using the packed sphere columns as well as the stacked screens. The conductivity degradation factors that were obtained for the stainless steel materials of 0.11 were very consistent for both screen and spheres. These new conductivity degradation factors for calculating the thermal conduction through packed sphere columns gives valuable information for regenerator optimization. With the NIST regenerator optimization software REGEN3.1,10,11 an improved coefficient of performance for regenerators can be achieved with proper optimization of regenerator geometry. REFERENCES

1. Walker, G., Crycoolers, Plenum Press, New York (1983). 2. Singer, E. and Wilhelm, R.H., Chemical Engineering Progress, Vol. 46 (1950), pp. 343-357. 3. Lewis, M.A., Kuriyama, T., Kuriyama, F., Radebaugh, R., “Measurement of Heat Conduction through Stacked Screens,” Advances in Cryogenic Engineering, Vol.43 (1998), pp. 1611-1618.

4.

Schumann, T.E.W., and Voss, V., “Heat Flow Through Granulated Material,” Fuel, Vol. 13 (1934),

5.

pp. 249-256. Slavin, A.J., Londry, F.A., Harrison, J.H., “A new model for the effective thermal conductivity of packed beds of solid spheroids: alumina in helium between 100 and 500°C,” International Journal of Heat and Mass Transfer, Vol. 43 (1999), pp. 2059-2073.

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6.

425

Takeno, “Thermal and Mechanical Properties of Advanced Cryogenic Materials at Low Temperatures,” Cryogenic Engineering, Journal of the Cryogenic Society of Japan, Vol. 21, No. 3 (1986), pp. 182-187.

7.

Lee, A.C., “Contact resistance of 500 mesh regenerator screens,” Cryogenics, Vol. 34, No. 5 (1994),

8.

pp. 451-456. Simon, N.J., Drexler, E.S., Reed, R.P., Properties of Copper and Copper Alloys at Cryogenic

Temperatures, and Below, NIST Monograph 177, Material Reliability Division, Boulder, CO, Feb. 1992. Childs, G.E., Bricks, L.J., Powell, R.I., Thermal Conductivity of Solids At Room Temperature and Below, NBS Monograph 131, Cryogenics Division, Boulder, CO, Sept. 1973. 10. Gary, J., Daney,D.E., and Radebaugh, R., “A Computational Model for a Regenerator,” Proc. Third Cryocooler Conference, NBS Special Publication 698, (1985), p. 199. 11. Gary, J. and Radebaugh, R., “An Improved Numerical Model for Calculation of Regenerator Performance (REGEN3.1),” Proc. Fourth Interagency Meeting on Cryocoolers, Report DTRC-91/003, David Taylor Research Center, 1991, p. 165.

9.

Innovative Technology for Low Temperature Regenerators L. Tuchinskiy, R. Loutfy and B. J. Tomlinson* MER Corp., Tucson, AZ, USA

*AFRL, Albuquerque, NM, USA

ABSTRACT

A novel approach for fabrication of cryocooler regenerators is proposed. The regenerators consist of a matrix with a system of through parallel microchannels of pre-assigned shapes and sizes. Diameters of the channels can be strictly controlled and set at any value from a few microns to a few millimeters. The volume fraction occupied by the microchannels can be also precisely controlled and set at any value from 1% to 95%. Regenerators of various customized shapes, geometrical and structural characteristics to ensure desirable end properties may be produced. The patented technique offers a possibility to fabricate cryocooler regenerators with controlled surface area and low fluid flow resistance from any powder materials including brittle magnetic intermetallics. Mechanical stability of the regenerators is expected to be much better than that of beds packed with spheres. It is anticipated that the cost of manufacture using the MCS technology will be significantly lower because of much higher yields compared to the traditional sphere production and spherical regenerator beds packing. INTRODUCTION The large demand for compact, energy efficient and powerful cryocoolers is driven by recent developments in infrared sensor detection technology, conductive cooling of conventional superconducting magnets for MRI systems, and wireless communications, where further increases of tower capacity and transmission quality require superconducting-electronics-based filters and preamplifiers. Current applications of cryogenic regenerative materials and regenerative heat exchangers include small cryogenic refrigerators (cryocoolers), such as Stirling, Gifford-McMahon, and pulse tubes to reach and maintain temperatures between 4.2 and 100 K at cooling powers ranging from 100 mW to 100 W. Regenerators are exceptionally suitable for conditions that occur in heat engines and cryogenic coolers. The three most important requirements to regenerator matrix are: 1) a maximum thermal storage capacity; 2) a maximum surface heat transfer area; and 3) a minimal hydraulic resistance to the fluid flow. Thermal storage capacity of a cryogenic regenerator matrix is proportional to the heat capacity of a regenerator material and, therefore, successful candidate materials must have large heat capacity. The heat capacities of conventional solid regenerator materials such as stainless steel or bronze or lead rapidly decrease with decreasing temperature below ~70 K. This limits practical application of stainless steel and bronze to temperatures above ~50 K and lead to temperatures above ~10 K. Cryocoolers 11, edited by R.G. Ross, Jr.

Kluwer Academic/Plenum Publishers, 2001

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Figure 1. Volumetric specific heat of two rare earth intermetallic compounds and lead:

Magnetic solids have received much attention because of their large heat capacities near the magnetic ordering temperatures (Curie temperature, or Nèel temperature, The unique feature of these materials is that they exhibit sharp increase in heat capacity at temperatures near and below 10 K. This phenomenon is shown in Fig. 1, where the volumetric specific heat of and are compared with the volumetric specific heat of Pb.1 One of the most successful regenerator materials is

that has been used in Gifford-McMahon

cryocoolers to reach 4 K with a cooling power of 1 W. A number of other rare-earth intermetallic

compounds, including ErNi, ErNi2, DyNi2, ErCo2, Er3Co, GdRh, ErRh, Nd3Ni, HoNi2, HoCu2, have demonstrated the greatest potential for enhancing the heat capacity at temperatures below 20K. However, effective use of these compounds in cryocoolers is limited. Because of their brittleness they are used today only in the form of spherical particles. The brittleness of the best intermetallic regenerator materials does not allow the use of more effective regenerators. The critical parameter defining the effectiveness of a regenerator packing is the ratio of the heat transfer surface area to the fluid pressure drop. In theory, the best design for regenerators is a matrix with multiple microchannels. It has a highest ratio of heat transfer area to pressure drop. An attempt to use this approach has been utilized in ribbon regenerators, where the channels are formed by

either dimpling or embossing the ribbon and then winding it on a mandrel. Unfortunately, the advantages of higher thermal efficiency never have been practically achieved with the ribbon regenerator because of channel tortuosity and associated maldistribution of the flow through the channels.1 If the core passages are uniform and the flow is laminar, the regenerator compactness can be expressed as2 where p = core porosity (dimensionless)

Nu = h D / k, the appropriate Nusselt number for laminar flow (dimensionless) the hydraulic diameter of a core passage (m) free-flow area of the passage (m) wetted perimeter of the passage (m) Equation (1) shows an inverse-squared dependence of core compactness on the passage hydraulic diameter. This means that if everything else is held constant, the regenerator compactness will increase dramatically as the diameter of core passages is decreased. If the regenerator is made as a solid block with straight microchannels or regularly stacked series of perforated plates, its effectiveness may be significantly increased. The problem is that there are no techniques for fabricating such kind of structures from brittle magnetic intermetallics. Should the technique be developed, it may result in the creation of higher efficiency low temperature (near and below 10 K) regenerators.

INNOVATIVE TECH FOR LOW TEMPERATURE REGENERATORS

429

The objective of this work is to demonstrate a novel approach for fabrication of high-efficiency microchannel cryocooler regenerators from rare-earth magnetic intermetallics and other materials.

TECHNOLOGY Our approach is based on a new, proprietary MER Corp. fabrication technique for microchannel structures (MCS).3 MCS represent a new type of engineered powder materials consisting of a matrix with a system of through parallel microchannels of pre-assigned shapes and sizes. Diameters of the channels can be strictly controlled and set at any value from a few microns to a few millimeters. The regenerator porosity (volume fraction occupied by the microchannels) can be also precisely controlled and set at any value from 1% to 95%. The MCS technology allows the fabrication of products of various customized shapes, geometrical and structural characteristics to ensure desirable end properties. This technique offers a possibility to fabricate cryocooler regenerators with controlled surface area and low fluid flow resistance from any powder materials including brittle magnetic intermetallics. Mechanical stability of MCS regenerators is expected to be much better than that of beds packed with spheres. It is anticipated that the cost of manufacture using the MCS technology will be significantly lower because of much higher yields compared to the traditional sphere production and spherical regenerator beds packing. The typical technological process for the fabrication of a MCS, as illustrated in Fig. 2, consists of the following steps. In the step 1 bi-material rods consisting of a shell and a core are produced. The shell comprises a mixture of a matrix powder (e.g. Er3Ni) and a thermoplastic binder. The core

comprises a mixture of the binder with a channel-forming filler, which can be removed afterwards by evaporation, melting, dissolution, etc.

The rods 1 produced in the step 1 are cut and assembled into a bundle 2 (step 2). The bundle is re-extruded through a die of the prescribed diameter (step 3). As a result of the step 3, green

composite rods that comprise the matrix (Er3Ni powder + binder) and fibers (filler + binder) are obtained. Plasticity of these green rods can be controlled by temperature and mixture composition; they can be subjected to any plastic deformation. If the rods produced by the step 3 still have filler fibers of larger diameter than required, they can be collected into a bundle again and step 3 is repeated for further reduction in scale. The rods of any cross-section shape can be obtained in this step. The produced green rods are debound by heating at temperature, which provides evaporation of the binder and filler, and sintered (step 4). The filler fibers burn out during or after debinding with the

resulting formation of the channels in the matrix. A variety of MCS with a prescribed density and anisotropy that are tailored to desired end use can be produced by this technique. The extrusion ratio and number of extrusion steps control the channel diameter. The porosity of the interchannel walls is controlled by the size of the intermetallic powder and sintering conditions (temperature and time); thus, these walls may be both solid and porous. The layers produced in the step 3 can be stacked layer-by-layer and pressed so that the future channels in the adjacent layers are oriented in parallel or at any pre-assigned angles. The proposed approach offers several important advantages, as listed below:

Figure 2. Main fabrication steps for MCS.

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REGENERATOR ANALYSIS AND MATERIALS DEVELOPMENTS

• Multichannel regenerators can be made of any materials • Controlled porosity may range from 1 to 90-95%

• Channel diameter in the range from a few microns to a few millimeters can be precisely controlled • Near-net-shape products can be produced. The technology for fabrication of microchannel Er3Ni structures is described below. FABRICATION OF MICROCHANNEL Er3Ni

Er3Ni is a very brittle intermetallic compound that can be relatively easily crushed and milled. Melt Er3Ni ingots produced by Iowa State University were crushed to pieces about 3-5 mm in average diameter and milled in a steel jar using hard-alloy ball media. After milling, the powder was sieved through a 325-mesh sieve. The percent output of powder as a function of milling time is shown in Fig. 3. After 5.5 to 6 hrs of milling, ~ 40 % of the source material is converted into 325mesh powder. The mean particle size is d50=17.5 µm, and 10% of the particles have a diameter less than 2.5 µm, and 90% less than 47µm. Specific surface area is about 0.7 m2/cm3.

Next, a mixture of the Er3Ni powder and a binder, as well as a mixture of a filler powder with the same binder, was prepared. These mixtures were co-extruded to produce green samples. Extrusion included the following steps (Fig. 4). 1. Extrusion of the bimaterial rods (1) with the (filler + binder) core and (Er3Ni + binder) shell. 2. Packing 19 rods (1) into the bundle (2) and re-extrusion of this bundle through the hexagonal die for fabrication of the hexagonal rod (3). The rod (3) consists of an (Er3Ni + binder) matrix and 19 (filler + binder) fibers. 3. Packing 19 rods (3) in the bundle (4) and re-extrusion for fabrication of the rod (5), which comprises the (Er3Ni + binder) matrix and 361 (filler + binder) fibers.

Figure 3. Output of 325-mesh Er3Ni power versus milling time.

Figure 4. Extrusion steps for fabrication of Er3Ni microchannel regenerators.

INNOVATIVE TECH FOR LOW TEMPERATURE REGENERATORS

431

Figure 5. Temperature versus time for the thermal debinding and filler removal.

Figure 6. SEM micrographs of multichannel Er3Ni.

The produced extrudates were subjected to heat treatment to decompose the organic binder and filler and remove them from the final product. Heating was performed in argon to prevent oxidation of the Er3Ni; the heating schedule in shown in Fig. 5. According to the phase diagram, pure Er and Ni melt at 1529°C and 1455°C, respectively. The binary compound Er3Ni, however, melts peritectically at considerably lower temperature (the peritectic temperature is 845°C, and the liquidus temperature is ~910°C). That enables sintering Er3Ni powders at temperatures below 845°C. In our experiments sintering was performed at 820°C during 2 hours. After sintering, the Er3Ni samples demonstrated uniform shrinkage without flaws and cracks, they had no visible defects, except a micron-size black film formed on the external surface. The average diameter of the microchannels was 60-80 µm as shown in Fig. 6.

CONCLUSIONS The proposed approach enables a significant mass and volume reduction in parallel with an increase in efficiency of cryocooler regenerators. The microchannel regenerators will be applicable

to pulse tube and Stirling coolers for all temperature ranges. If the regenerators are made out of stainless steel, then they will improve the performance of coolers above 50 K due to the improved compactness factor. Multichannel lead regenerators will improve the cooling capacity at 35 K. At 10 K, the regenerators, either of lead or a rare-earth material, will be applicable. The developed

technique may also be used for fabrication of high performance micro-heat exchangers, large and small prime movers in Stirling engines, catalyst carriers, etc. Due to the much lighter weight, better vibration damping and sound absorption of the microchannel structures the application possibilities will extend to general transportation. ACKNOWLEDGMENT The authors acknowledge the support of the Department of Air Force on this work.

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REFERENCES

1. R. Ackerman, Cryogenic regenerative heat exchangers, Plenum Press, N.Y., 1997. 2. D.S.Beck, D.G.Wilson, Gas-Turbine Regenerators, Chapman & Hall, International Thomson Publishing, 1996. 3. L. Tuchinskiy, “Multi-Channel Structures and Processes for Making Such Structures,” US Patent 5,774,779 (1998).

Ductile, High Heat Capacity, Magnetic Regenerator Alloys for the 10 to 80 K Temperature Range K. A. Gschneidner, Jr.,1,2 A. O. Pecharsky1 and V. K. Pecharsky1,2 Ames Laboratory1 and Dept. of Materials Science and Engineering2

Iowa State University Ames, IA 50011-3020, USA

ABSTRACT New erbium-based regenerator materials have been developed as a replacement for lead in

low temperature cryocoolers. These alloys have volumetric heat capacities which are 20 to 185% larger than that of lead from 10 to 80 K. These magnetic, ductile, oxidation resistant

erbium alloys are more than 10 times stronger than lead and have a thermal conductivity ~10 times lower than lead. The alloys can easily be fabricated into spheres, foils, ribbons, and wires, and are environmentally friendly, non-toxic materials. A layered regenerator composed of three different erbium alloy compositions is recommended as the most efficient system to improve the

cryocoolers’ performance to “get the lead out.” INTRODUCTION

The use of lanthanide intermetallic compounds, which exhibit low magnetic ordering temperatures (<10 K), as cryogenic magnetic regenerator materials was pointed out by Buschow

et al.1 25 years ago. The practical use of lanthanide regenerators was not realized until 15 years

later, when Sahashi et al.2 and Kuriyama et al.3 used Er3Ni as the low temperature stage regenerator material in a two-stage Gifford-McMahon (GM) cryocooler. In these articles the Japanese scientists proposed that low temperature regenerator material in common use, Pb, be partially replaced by Er3Ni. This replacement allowed one to reduce the low temperature limit of

GM crycoolers from ~10 K to ~4 K. The achieved improvement is due to the higher volumetric heat capacity of Er3Ni relative to Pb below 25 K. Above 10-15 K lead is the material of choice because of its high heat capacity, which is due to its low Debye temperature of 102 K.4 Above 50 K stainless steel or bronze have much larger heat capacities than Pb and the former are used as the upper stage regenerator materials of a two-stage GM and other cryocoolers. Since the efficiency of cryocoolers is proportional to the heat capacity of the regenerator material, the higher the volumetric heat capacity the greater the amount of heat that can be transferred from the low temperature heat exchanger (or the material to be cooled) to the hot heat exchanger (exhaust) per cycle of fluid flow through the regenerator bed. Cryocoolers 11, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2001

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The utilization of the magnetic contribution to the total heat capacity offers a promising avenue to follow to improve the efficiency of a cryocooler. The magnetic contribution to the heat capacity arises from the magnetic ordering process itself and this can give rise to either a very narrow heat capacity peak for a first order magnetic transformation or a modestly broader lambda-like peak for a second order magnetic transition. There is also a third type of magnetic heat capacity peak, which is known as a Schottky anomaly and is quite broad. It is due to the thermal excitation of electrons from lower energy levels to higher energy levels due to the crystalline electric field (CEF) splitting of the 4f levels of a magnetic lanthanide metal. In the case of Er3Ni the 7 K heat capacity peak is due to a second order paramagnetic to antiferromagnetic (on cooling) transition, but because the CEF levels are probably fairly low,5 the Schottky heat capacity contribution above 7 K prevents the total heat capacity from falling off as rapidly as is normally observed for a second order magnetic transition. There are many lanthanide compounds which order magnetically below 20 K, but because most intermetallic compounds are quite brittle, this can lead to particle fragmentation when they are used as regenerator materials in GM and other types of cryocoolers. This has been discussed in some detail by Merida and Barclay.6 Thus we have turned our attention to ductile, magnetic materials, primarily solid solution alloys based on pure lanthanide metals as a replacement for lead in cryocooler regenerators. The main results obtained to date are discussed below. HEAT CAPACITY OF MAGNETIC LANTHANIDE METALS

The magnetic ordering processes in the magnetic lanthanide metals is generally quite complicated, and for most of them at least two ordering processes take place as the samples are cooled below room temperature.7,8 The volumetric heat capacities for high purity, electrotransport purified ( solid state electrolysis) heavy lanthanides (Gd, Tb, Dy, Ho and Er) are shown in Fig. 1 for the 4 to 100 K temperature interval. Above 100 K the heat capacities for Gd, Tb, Dy and Ho are comparable to those of bronze and stainless steel at least in the vicinity of their respective ordering temperatures. Below about 70 K Pb has a much higher heat capacity than either bronze or stainless steel, see Fig. 1, and is used as the low temperature regenerator material for cooling down to ~10 K. However, Er metal has a significantly higher volumetric heat capacity than lead from ~20 K to ~90 K, and is about the same as lead from 90 to 350 K and from 5 to 20 K. Thus Er or Er-based alloys might be an excellent substitute for Pb as a regenerator material. However, the sharp peak at 19 K is not particularly useful because it occurs over a narrow temperature range, 2 K. Thus if the entropy associated with this peak were distributed over a wider temperature range, or better yet, if the peak temperature was lowered it might have a higher heat capacity than lead below 20 K. This might be accomplished by adding various alloying agents to Er, and this is discussed in the next section. ERBIUM-BASED SOLID SOLUTION ALLOYS

Interstitial Alloys

The interstitial Er-based alloys containing O and C can readily be prepared by arc-melting high purity Er with Er2O3 or graphite, respectively. The alloys containing H or N are prepared by reacting Er metal with H2 at moderate temperatures (~500°C [~775 K]) or with N2 at elevated temperatures (>1000°C [>1275 K]). However, most of the commercially available “pure Er” metal already contain a substantial amount of interstitial impurities, and the interstitial alloy (noted below) was purchased from a commercial vendor without any additional alloying or processing. The presence of 2.7 at.% O, plus smaller amounts of N and C, 0.3 and 0.2 at.%, respectively, in pure Er, reduces the height of the sharp peak at 19 K and shifts it upward by about 3 K. Furthermore, it probably combines with the smaller 25 K peak in pure Er to give a small, but broader, heat capacity peak at 22 K (see Fig. 2). Furthermore, the 52 K peak is nearly eliminated while the 88 K peak is shifted downward to 84 K. As a whole the volumetric heat

MAGNETIC REGENERATOR ALLOYS FOR 10 TO 80 K RANGE

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Figure 1. The volumetric heat capacity from 0 to 100 K of high purity Gd, Tb, Dy, Ho and Er metals,

along with those of bronze, stainless steel and lead. The letters SSE stand for solid state electrolysis, which was used to purify these metals.

capacity of pure Er between the various peaks is not changed significantly by the addition of the interstitial elements and is still better than lead over 20 to 80 K region. Throughout the rest of the paper this alloy will be used as the prototype Er alloy against which the heat capacities of the other Er-based alloys will be compared.

Figure 2. The volumetric heat capacity of an Er-based interstitial alloy

100 K. Also shown are the heat capacities of high purity Er and lead.

from 3.5 to

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REGENERATOR ANALYSIS AND MATERIALS DEVELOPMENTS

Non-lanthanide Substitutional Alloys A five atomic percent addition of Sc or Y or Zr or Hf wiped-out the 19, 25 and 52 K peaks and lowered the 88 K one. Below 18 K the volumetric heat capacity is slightly larger by a few percent over that of the prototype alloys and Pb. But as a whole these alloys are not much of an improvement, if at all, over the Er-based interstitial alloy as a regenerator alloy. Lanthanide Substitutional Alloys

The heavy lanthanide elements (Gd, Tb, Dy and Ho) tend to raise the magnetic transitions temperature of Er, and thus these metals are not useful in increasing the heat capacity below 20 K. Lutetium, the last member of the lanthanide elements, has no unpaired 4f electrons, and its

behavior is similar to that observed for the non-lanthanide alloying agents Sc, Y, Zr and Hf, see above.

The light lanthanides (La, Ce, Pr and Nd), however, have an unusual influence on the magnetic ordering phenomena in Er. With respect to the upper (or Néel temperature) transition of Er, all four of these elements lower it, in proportion to their size: the larger La atom lowers 88 K peak the most (to ~70 K for 5 at.%), followed by Ce (about the same as La), which is

followed by Pr (to ~75K for 5 at.%) and Nd (to ~78 K at 5 at.%). For the lower temperature transformations, La and Ce behave quite similarly, and significantly different from that of Pr and

Nd, which have similar affects. For La and Ce small additions (5 at.%) wipe-out the 52 K transition while rapidly raising

and merging the 19 and 25 K peaks into one. The 19/25 K peak eventually merges with the rapidly dropping Néel temperature (88 K) at 20 at.% (La or Ce). This results in large somewhat broad (11 K) heat capacity peak at ~40 K, but the heat capacity below 34 K and above 45 K lies below that of the alloy. The maximum heat capacity is 1.8 J/cm3K at 40 K for the and and the alloys would be useful if one needed a high heat capacity material for the 35 to 45 K range. The addition of 5 at.% Pr or Nd destroys the 25 K heat capacity anomaly shifting the entropy toward the 19 K Curie temperature peak while reusing its temperature to about 23 K.

The 52 K peak is rapidly lowered by the initial additions of Pr reaching a minimum at ~7 at.% Pr before rising slightly and leveling-off when more than 10 at.% Pr is added. As more

Pr (or Nd) are added to the Er (~15 at.%), a double peak structure develops, which eventually merges into one peak at ~27 at.% Pr, due to rapidly dropping of the 88 K transition temperature. These behaviors are shown in Fig. 3a which shows the change of the transition temperatures and Fig. 3b which shows the associated heat capacity peak values, both as a function of Pr content. The heat capacities of a series of Er-Pr alloys are shown in Fig. 4. It is noted that for alloys

containing more than 30 at.% Pr the heat capacity is larger than that of either Pb below 20 K. This is shown in more detail in Fig. 5, where it is seen that highest heat capacity of any of the Er-Pr alloys and the Er interstitial alloy

Pb or Er3Ni between 10 and 20 K. At 10 K the

or has the

alloy’s heat capacity is 185% larger than

that of Pb and the same as that of Er3Ni.

The Nd additions follow a similar trend as shown by the Pr addition, but since the upper transition is lowered at a slower rate, more than 30 at.% Nd is required before the two peaks merge into one. Based on these results we suggest the following combination of Er-based alloys as a replacement of lead in the cryocooler regenerator: at the high temperature end the interstitial

alloy; the

alloy for the intermediate temperature region; and the

alloy for the low temperature end of the regenerator. The heat capacities of these three

Er-based alloys are shown in Fig. 6 along with that of lead. It is evident from Fig. 6 that the alloy would be the best regenerator material from 40 to 85 K having a heat capacity 20 to 40% larger than that of Pb; while which has a heat capacity 20 to 30% larger than that of Pb, would cover the 24 to 40 K range, and would be the most efficient below 24 K, since its heat capacity is 20 to 185% larger than lead. Of course if cooling down

MAGNETIC REGENERATOR ALLOYS FOR 10 TO 80 K RANGE

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Figure 3. The variation of the magnetic ordering temperature (a) and the peak value of the heat capacity at the respective ordering temperature (b) as a function of the Pr concentration.

much below 10 K (i.e. down to 4 K) is required, a lower temperature stage regenerator composed of Er3Ni or HoCu2 or Nd would be needed, as is the standard practice today, when Pb is used as the intermediate temperature stage regenerator material.

Figure 4. The volumetric heat capacity of a series of Er-Pr alloys from 3.5 to 100 K and the prototype interstitial alloy and Pb.

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REGENERATOR ANALYSIS AND MATERIALS DEVELOPMENTS

Figure 5. The volumetric heat capacity of some Er-based alloys from 3 to 22 K. The heat capacities of Er3Ni and Pb are shown for comparison.

Other Properties In addition to the significantly higher heat capacity, these three Er alloys have other distinct advantages over lead with a lower thermal conductivity and a higher strength; and they are not as toxic as Pb, which is a federally regulated poison. The thermal conductivity which is shown in Fig. 7, is nearly one order of magnitude lower than that of Pb, comparable to that of stainless steel. The reduced thermal conductivity would lead to lower longitudinal heat losses in the

Figure 6. The heat capacities of

and

along with that of Pb.

MAGNETIC REGENERATOR ALLOYS FOR 10 TO 80 K RANGE

439

Figure 7. The thermal conductivities for interstitial Er, Pb and stainless steel.

regenerator. The thermal conductivity of the slightly lower than that of the Pr atoms in the Er matrix.

Er73Pr27 and Er50Pr50 would be expected to be because of additional phonon scattering from the

The ultimate tensile strength of interstitital Er, Pb, stainless steel and bronze are listed in

Table 1. It is evident that the tensile strength of is an order of magnitude larger than that of Pb, and about the same as bronze. This is important if the Er-based alloy are used in the form of spheres, since the materials are strong enough to prevent the loss of sphericity as can happen with Pb (Sb-hardened) alloys. The strength of the Pr-Er alloys would be expected to be about the same as that of the interstitial Er alloy. This level of strength is important when these Er-based alloys are used as wires, screens, flat sheets and jellyrolls in cryocooler regenerators, which one cannot do with Pb since it is so weak.

There are two other important features which need to be mentioned. One is that the Er alloys do not oxidize like the light lanthanides, e.g. Nd, which has been used in cryocooler regenerators. We have held and Er60Pr40 at 396±5 K for over 15 months and there has been no measurable weight gain or loss (within ±0.1 mg) for samples weighing 3.6, 2.5 and 4.1 g, respectively. This is in contrast to Nd metal which oxidizes to form Nd2O3 within a few hours after being exposed to ambient air at room temperature. Thus these alloys can be easily handled and stored without any special precautions. The second point is that since the Er-based alloys are solid solution alloys, and not intermetallic compounds, they can be readily fabricated into spheres, wires and foils (ribbons), see Fig. 8. Since they are ductile alloys with reasonable strength they will not decrepitate or crumble, which can easily occur when using brittle intermetallic compounds in regenerators due

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REGENERATOR ANALYSIS AND MATERIALS DEVELOPMENTS

Figure 8. Fabricated forms of spherical powders – 0.30mm (12 mil) diameter; ribbon – 0.05mm (2 mil) thick; and wire – 0.30mm (12 mil) diameter.

to the high pressure pulses of the gases that are rapidly recycled through the regenerator. Details on the fabrication of the and alloys into spherical powders are being presented in another paper at this conference.12

Initial Test Results An initial test using in place of Pb as the regenerator material in a single stage pulse tube cryocooler indicated that at low frequencies the Er interstitial alloy performed better than Pb. Additional results using conference.13

are presented in another paper at this

CONCLUSIONS AND SUMMARY

We have shown that Er-based alloys in comparison to Pb have significantly (1) higher volumetric heat capacities from 10 to 80 K, (2) lower thermal conductivities, and (3) improved tensile strength. The Er alloys are oxidation resistant below 396 K (123°C), and can readily be formed into wires, ribbons (foils) and spheres. We have suggested that the most efficient replacement for lead in a regenerator is to use a combination of three alloys at the hot end, at the intermediate temperatures, and at the cold end. Finally we wish to note that these Er-based alloys are not a replacement for Er3Ni, Nd or HoCu2, which are required to reach temperatures below 10 K, but are to be used in conjunction with them if temperatures below 8 K are required. Our theme is not only to improve cryocooler performance

but also to “get the lead out.” ACKNOWLEDGEMENTS This work was supported in part by Atlas Scientific, Sunnyvale, California via a SBIR, and in part by the Materials Sciences Division, Office of Basic Energy Sciences, U.S. Department of

Energy under contract W-7405-ENG-82.

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441

REFERENCES

1. Buschow, K.H.J., Olijhoek, J.F. and Miedema, A.R., “Extremely Large Heat Capacities Between 4 and 10 K”, Cryogenics, vol. 15 (1975), pp. 261-264. 2. Sahashi, M., Tokai, Y., Kuriyama, T., Nakagome, H., Li, R., Ogawa, M. and Hashimoto, T., “New Magnetic Material R3T System with Extremely Large Heat Capacities Used as Heat Regenerators”, Adv. Cryogen. Eng., vol. 35 (1990), pp. 1175-1182.

3. Kuriyama, T., Hakamada, R., Nakagome, H., Tokai, Y., Sahashi, M., and Li, R., Yoshida, O., Matsumoto, K. and Hashimoto T., “High Efficient Two-Stage GM Refrigerator with Magnetic Materials in Liquid Helium Temperature Region”, Adv. Cryogen. Eng., vol. 35 (1990), pp. 12611269. 4. Gschneidner, K.A., Jr., “Physical Properties and Interrelationships of Metallic and Semimetallic Elements”, Solid State Phys., vol. 16 (1964), pp. 275-426. 5. Although the crystalline electric field splitting of the Er 4f electrons in the ground state multiple 4I15/2

in Er3Ni are not known, those of the isostructural Er3Co compound have been determined. The calculated and measured Schottky heat capacity of Er3Co shows a broad peak at ~12 K which slowly falls off with increasing temperature reaching about half the peak value at 30 K. We would expect that the Er3Ni heat capacity would show a very similar behavior. This work was reported by: Takahashi Saito, A., Tutai, A., Sahashi and M., Hasimoto, T., “Crystal Field Effects on Thermal and Magnetic Properties of Er3Co”, Jpn. J. Appl. Phys., vol. 34 (1995), pp. L171-L173. 6. Merida, W.R. and Barclay, J.A., “Monolithic Regenerator Technology for Low Temperature (4 K) Gifford-McMahon Cryocoolers”, Adv. Cryogen. Eng., vol. 43 (1998), pp. 1597-1604. 7. Sinha, S.K., “Magnetic Structures and Inelastic Neutron Scattering: Metals, Alloys and Compounds” in Handbook on the Physics and Chemistry of Rare Earth, Gschneidner, K.A., Jr. and Eyring, L., eds., North-Holland Publishing Co., Amsterdam (1978), vol. 1, chap. 7, pp. 489-589. 8. McEwen, K.A., “Magnetic and Transport Properties of the Rare Earths” in Handbook on the Physics

and Chemistry of Rare Earth, Gschneidner, K.A., Jr. and Eyring, L., eds., North-Holland Publishing Co., Amsterdam (1978), vol. 1, chap. 6, pp. 411-488. 9. Scott, T.E., “Elastic and Mechanical Properties” in Handbook on Physics and Chemistry of Rare

Earths, Gschneidner, K.A, Jr. and Eyring, L., eds., North-Holland Publishing Co., Amsterdam (1978), vol. 1, chap. 8, pp. 591-705. 10. Anonymous in Properties and Selection: Nonferrous Alloys and Special-Purpose Materials, Metals Handbook, 10th ed., ASM Intern., Materials Park, OH (1990), vol. 2, p. 217 (Bronze) and p. 550 (Pb). 11. Anonymous in Properties and Selection: Irons, Steels and High-Performance Alloys, Metals

Handbook, 10th ed., ASM Intern., Materials Park, OH (1990), vol. 1, p. 855.

12. Miller, S.A., Nicholson, J. D., Gschneidner, K.A., Jr., Pecharsky, A.O. and Pecharsky, V.K., “Manufacturing Considerations for Rare Earth Powders Used in Cryocooler and Magnetic

Refrigerator Applications”, paper presented at 11th International Cryocooler Conference, June 20-22, 2000, Keystone, Colorado.

13. Kashani, A., Helvensteijn, B.P.M., Gschneidner, K.A., Jr., Pecharsky, A.O. and Pecharsky, V.K., “New Regenerator Materials for Use in Pulse Tube Coolers”, paper presented at 11th International Cryocooler Conference, June 20-22, 2000, Keystone, Colorado.

Low Temperature Properties of HoSb, DySb, and GdSb H. Nakane, S. Yamazaki, H. Fujishiro*, T. Yamaguchi**,

S. Yoshizawa**, T. Numazawa*** and M. Okamura****

Kogakuin University, Shinjuku-ku, Tokyo, 163-8677, Japan * Iwate University, Ueda, Morioka, 020-8551, Japan ** Meisei University, Hino-shi, Tokyo, 191-8506, Japan ***

National Research Institute for Metal, Tsukuba-shi, 305-0003, Japan

****

Toshiba Corporation, Yokohama-shi, 235-8522, Japan

ABSTRACT

Materials including rare-earth and Sb compounds were developed as regenerator materials for a 4K GM refrigerator. Sb compounds have a remarkably large peak value of specific heat in the low temperature range. First, the heat capacities, thermal conductivity, and thermal expansion of HoSb, DySb and GdSb, which are anti-ferromagnetic with a large spin value in the low temperature range, were measured. Then, the results of these measurements were used to analyze the problems encountered when Sb compounds were used as the regenerative materials in a GM refrigerator. As for the thermal expansion, the measured values of Sb compounds were compared with that of fabric-impregnated phenol-formaldehyde resin, which is usually used as the wall material of a regenerator. As regards the thermal conductivity, the thermal diffusivity was obtained by using the

measured values of thermal conductivity and specific heat. The heat penetration depth was evaluated from the thermal diffusivity. Discussion of the heat penetration depth is based on the figures obtained in a previous experiment using a GM refrigerator with HoSb packed at the cold end of the second stage of a multi-layer regenerator. INTRODUCTION

In a small 4 K refrigerator, the refrigeration capacity depends on the response speed of heat exchange, i.e., the exhaustion and absorption of heat between the working gas and the materials packed into the regenerator. In order to obtain higher effectiveness in the regenerator, the heat capacity of the regenerator materials must be larger than that of 4He used as the working gas. Only a magnetic material that has a large specific heat based on magnetic phase transition is effective below 15 K. However, the specific heat of pressurized 4He gas is very large below 15 K. Since the heat exchange region of 4He gas is wide, a single magnetic material cannot cover the specific heat of4He gas. Multi-layer regenerators composed of several rare-earth compounds that have different specific heat peak temperatures are usually used. HoSb and DySb compounds were discovered to have a remarkably large specific heat peak. HoSb compound has the peak value of 2.7 J/K cm3 at 5.3 K, and DySb, 2.0 J/Kcm 3 at 9.5 K. The measured results of the specific heat1 of HoSb, DySb and GdSb are shown in Fig. 1. HoSb was Cryocoolers 11, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2001

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Figure 1. Temperature dependence of specific heat of Sb compound.

packed into a multi-layer regenerator and the refrigeration capacities were measured and are included in the report.2

In regards to the measured values of thermal expansion, the relation between the fabric-impregnated phenol-formaldehyde resin (fabric-impregnated Bakelite), which is usually used as the wall material of regenerators, and Sb materials, which have a remarkably large peak, was analyzed in this experiment. Then, the thermal conductivities of HoSb, DySb and GdSb were measured, and the thermal diffusivity was obtained from the measured values of thermal conductivity and the specific heat. The heat penetration depth of HoSb was compared with that of4He used as the working gas. Whether the driving cycle of the displacer was optimal was verified for the former experiment2 in which HoSb was packed at the cold end of the second stage multi-layer regenerator of a GM refrigerator.

EXPERIMENTS The samples used were intermetallic compounds of HoSb, DySb and GdSb made by arc welding at the melting points of 2433 K (2160°C), 2443 K (2170°C) and 2403 K (2130°C), respectively. Two Sb compounds were found to have remarkably large specific heat peak values in the low temperature range. One of these compounds is HoSb, with the peak value of 2.7 J/K cm3 at 5.3 K, and the other one is DySb, with 2.0 J/K cm3 at 9.5 K. GdSb has a broad specific heat peak around 24 K. The measuring methods of thermal conductivity and thermal expansion of such magnetic materials are explained and the measuring results are shown as follows:

Measurement and Discussion of Thermal Expansion The thermal expansions of HoSb, DySb and GdSb were measured by a clip type dilatometer as shown in Fig. 2. A resistance bridge was formed by four strain gauge elements (Kyowa Electronic Instruments, KFL-1-120-C1-11) bound to a phosphor-bronze clip. The balance of the resistance bridge was directly measured to an accuracy of 71/2 figures with an automatic thermometer bridge (TINSLEY 5840D) by the four-wire method. At a certain temperature, the balance of the resistance bridge changes almost linearly to the length of the specimen. To obtain the calibration values of for the expansion, the of two copper reference specimens of different lengths was measured, as the thermal expansion rate of copper is well known.

Figure 2. Schematic diagram of the clip type dilatometer using four strain gage elements.

LOW TEMPERATURE PROPERTIES OF HoSb, DySb AND GdSb

445

Figure 3. Thermal expansions of GdSb, DySb and HoSb.

Figure 4. Temperature dependence of Sb compounds and Bakelite.

The lengths of HoSb, DySb and GdSb samples were 10 mm, and the cross sections about 6.3 mm x 2.3 mm. The accuracy of the measurement increases with the size of the cross section. As for the fabric-impregnated phenol-formaldehyde resin (fabric-impregnated Bakelite), which is usually

used as the wall material of regenerators, a sample of 10 mm in diameter and 10 mm in length was cut out from a circular rod. The linear expansions of HoSb, DySb and GdSb are shown in Fig. 3. As shown in Fig. 1, HoSb had the highest specific heat peak, DySb second, and GdSb third. However, the sharpness of change in the linear expansion around the Néel temperature was DySb, first; HoSb,

second; and GdSb, third (not sharp). In Fig. 4, the linear expansions normalized at 300 K are expressed as percentages. The linear expansions of the samples at 4 K were about – 0.2 % for Sb compounds, – 0.5 % for Bakelite against the diameter and – 0.2 % against the length. The contraction of Bakelite in diameter was larger than those of HoSb and DySb. In a former experiment,2 in which HoSb was packed at the cold end of the second stage of a multi-layer regenerator of a GM refrigerator, the effect of contraction of the wall materials was compensated by using felt as the partition material.

Measurement and Discussion of Thermal Conductivity The thermal conductivity was measured by the steady state heat flow method. Fig. 5 shows a schematic view of the sample set on top of the cold head of a GM refrigerator used as a cryostat3. Both samples were glued to the cold-finger of the refrigerator and the metal film resistance heater (10 kW) to the sample with GE7031 varnish. AuFe(0.07 at.%)-Chromel thermocouples with a di-

ameter of 73 µm were used as thermometers. The temperature range for the measurement was from 4 to 300 K. Air from the whole chamber of the sample was removed to 1 x 10–6 Torr with an oil diffusion pump. The samples were cut to the dimensions of 3 mm x 3 mm x 10 mm as the accuracy

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Figure 5. Schematic diagram of the thermal diffusivity measurement system.

Figure 6. Temperature dependence of thermal conductivity.

of measurement inversely increases with the size of the cross section in this method. The experimental results of the thermal conductivity of HoSb, DySb and GdSb are shown in Fig. 6. The thermal conductivity of lead (Pb), stainless steel, ErNi2, ErCo2 and DyNi2 which are used as conventional regenerator materials, is shown in Fig. 6 for comparison. Heat is periodically absorbed and exhausted by the materials used in a regenerator. Thermal diffusivity is often a more convenient parameter than thermal conductivity in discussing such a non-steady heat transfer subject. As is well known, thermal diffusivity D is given by: where is the thermal conductivity and Cp the volumetric specific heat. By substituting the temperature properties of the specific heat and thermal conductivity of each material into Eq. (1), the thermal diffusivities D are obtained as shown in Fig. 7. As shown in Fig. 7, the thermal diffusivities

of HoSb and DySb at the Néel temperature are very low in comparison with that of Pb. It is, therefore, necessary to discuss whether the materials are effective under the conditions of non-steady heat transfer that occurs in regenerators. Obviously the